INTERNATIONAL EXERGY, ENERGY AND ENVIRONMENT SYMPOSIUM

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1 8 th INTERNATIONAL EXERGY, ENERGY AND ENVIRONMENT SYMPOSIUM May 1-4, 2016 Antalya, Turkey Proceedings of the 8 th International Exergy, Energy and Environment Symposium ISBN:

2 Proceedings of the 8 th International Exergy, Energy and Environment Symposium ISBN: th INTERNATIONAL EXERGY, ENERGY AND ENVIRONMENT SYMPOSIUM May 1-4, 2016 Antalya, Turkey i

3 ORGANIZING COMMITTEE CONFERENCE CHAIR Onder KIZILKAN Suleyman Demirel University, Turkey HONORARY CHAIR T. Nejat VEZIROGLU International Association for Hydrogen Energy, USA FOUNDING CHAIR Ibrahim DINCER University of Ontario Institue of Technology, Canada ORGANIZING COMMITTEE MEMBERS Ibrahim DINCER University of Ontario Institue of Technology, Canada Ali Kemal YAKUT Suleyman Demirel University, Turkey Reşat SELBAS Suleyman Demirel University, Turkey İskender AKKURT Suleyman Demirel University, Turkey Onder KIZILKAN Suleyman Demirel University, Turkey Ahmet KABUL Suleyman Demirel University, Turkey Can Ozgur COLPAN Dokuz Eylül University, Turkey Sandro NIZETIC University of Split, Croatia Mehmet Akif EZAN Dokuz Eylül University, Turkey Melik Ziya YAKUT Suleyman Demirel University, Turkey Fatih YIĞIT Suleyman Demirel University, Turkey LOCAL ORGANIZING COMMITTEE MEMBERS Adnan OZDEN Dokuz Eylul University, Turkey Anil ERDOGAN Dokuz Eylul University, Turkey Ceren YUKSEL Dokuz Eylul University, Turkey Omer Faruk ATACAN Dokuz Eylul University, Turkey David OUELLETTE Dokuz Eylul University, Turkey Gamze YAKUT Suleyman Demirel University, Turkey Mustafa ERCELIK Dokuz Eylul University, Turkey Ozum CALLI Izmir University of Economics, Turkey Ugur GENCALP Dokuz Eylul University, Turkey ii

4 INTERNATIONAL ADVISORY COMMITTEE (In Alphabetic Order) A. Bejan, USA A. Beyene, USA A. Hepbasli, Turkey A. Kabul, Turkey A. Midilli, Turkey A. E. Ozgur, Turkey A. F. Miguel, Portugal A. H. Reis, Portugal A. K. Yakut, Turkey B. V. Reddy, Canada C. Koroneos, Greece D. Queiros-Conde, France E. Michaelides, USA E. Sciubba, Italy E. Shirani, Iran F. Aloui, France F. Hamdullahpur, Canada G. Lebon, Belgium G. Tsatsaronis, Germany G. F. Naterer, Canada H. El-Qarnia, Morocco H. Kwak, Korea H. Peerhossaini, France H. Yamaguchi, Japan H. C. Bayrakci, Turkey H. M. Sahin, Turkey I. Benko, Hungary I. Dincer, Canada I. Yildiz, Canada J. Yan, Sweden M. Feidt, France M. A. Gadalla, UAE M. A. Rosen, Canada M. B. Pate, USA O. Altuntas, Turkey O. Arnas, USA O. Kizilkan, Turkey P. Grammelis, Greece P. Lund, Finland R. El Emam, Egypt R. Selbas, Turkey S. Lorente, France S. Obara, Japan S. Nizetic, Croatia S. A. Sherif, USA T. Akiyama, Japan T. H. Karakoc, Turkey T. N. Veziroglu, USA X. Li, Canada X. R. Zhang, China V. I. Ugursal, Canada Y. Demirel, USA Y. A. Cengel, Turkey Y. Iwamoto, Japan Z. Sen, Turkey iii

5 PREFACE During the past few decades, the world has faced various critical challenges in various dimensions, ranging from energy to environment and from economy to sustainability. In all these dimensions, energy plays the most critical roles since the way that we produce, transfer, transport, convert, and use energy affects all other dimensions significantly. There is a strong need for clean energy solutions to overcome environmental, resource, efficiency, cost, energy security, and sustainability issues. When the Exergy, Energy and Environment Symposium was first launched 13 years ago in 2003 by the founding chair Dr. Ibrahim Dincer in Izmir, Turkey, this was a kind of ultimate goal.. The symposium, since then, has been running successfully under the title of International Exergy, Energy and Environment Symposium (IEEES). The conference was held in the following cities: Kos, Greece (2005); Evora, Portugal (2007); Sharjah, United Arab Emirates (2009); Luxor, Egypt (2011); Rize, Turkey (2013) and Valenciennes, France (2015). All symposium chairs, in this regard, deserve a clear recognition. Our warm thanks go to all the organizers who have contributed in the success of this conference series. This year, the 8 th International Exergy, Energy and Environment Symposium (IEEES-8) is bringing all disciplines together. IEEES-8 is a multi disciplinary international symposium, covering three main areas of exergy, energy and the environment and aims to provide a forum for researchers, scientists, engineers and practitioners from all over the world to exchange information, to present high-quality research results and new developments in the wide domain covered by exergy, energy, and the environment, and discuss the future direction and priorities in the field. The primary theme of the conference is exergy, energy and environment, not only in engineering and science but also in all other disciplines (e.g., ecology, education, social sciences, economics, management, political sciences, and information technology). Therefore, papers on related topics were solicited from all relevant disciplinary areas, ranging from current problems, projections, new concepts, modeling, experiments and measurements, to simulations. The IEEES-8 has received extraordinary international attention from every corner of the world. Here are some summary figures to share with you: Number of abstracts received: 319 Number of papers received: 169 Number of presentation scheduled for the program: 267 (with 139 oral and 128 poster presentations) IEEES-8 includes plenary sessions, keynote lectures, and several oral and poster sessions on different topics related to exergy, energy and environment. The conference aims to offer a distinctive platform for an effective and fruitful communication between the research, government and industrial communities. We iv

6 hope that everyone will find IEEES-8 both enjoyable and technically enlightening. We are sure everyone will also enjoy Antalya, which is one of the most beautiful cities in Turkey. As we are all aware, the efforts required in organizing and holding this kind of symposium are extensive. First, I would like to take this opportunity to express my sincere appreciation to Dr. T. Nejat Veziroglu, who is the Honorary Chair of the symposium. Second, I would like to express my warmest thanks to Dr. Ibrahim Dincer as the Founding Chair of the symposium. Third, my special thanks go to all the organizing committee members for their exemplary efforts. Last, but not least, I acknowledge my gratitude to the IEEES-8 keynote speakers, authors, session chairpersons and attendees, whose contributions and efforts have made us to come up with this stellar program. Dr. Onder Kizilkan IEEES-8 Symposium Chair v

7 TABLE OF CONTENTS ORGANIZING COMMITTEE... ii INTERNATIONAL ADVISORY COMMITTEE... iii PREFACE... iv TABLE OF CONTENTS... vi ENERGY, ENTROPY AND EXERGY ANALYSIS AND MANAGEMENT... 1 A Study on Exergetic Performance of Afsin Lignite Stoichiometric Combustion Process, Sefa Yalcin, Alp Er S. Konukman, Adnan Midilli... 2 Page Exergy Analysis of Nitrogen Liquefaction Process, Arif Karabuga, Resat Selbas, Ahmet Kabul Investigation of Irreversibility with CO 2 Emission Measurement in Industrial Enamel Furnace, Sedat Vatandas, Atakan Avci, M. Ziya Sogut Advanced Exergy Analysis of an Application of Waste Heat Powered Ejector Refrigeration System to Rotary Kiln, Abid Ustaoglu, Mustafa Alptekin, Mehmet Emin Akay, Resat Selbas Thermodynamic Evaluation of Absorption-Compression Cascade Refrigeration Cycles for Advanced Exergy Analysis, Mustafa Alptekin, Abid Ustaoglu, Mehmet Emin Akay, Resat Selbas Exergy Optimization of the Hybrid Compression-Absorption Industrial Refrigeration Systems, Mahmoud Afshar, Hamid Rad Energy and Exergy Analysis of a Steam Power Plant Considering Effect of Varying Plant Loads, Mehmet Bilgili, Mehmet Tontu, Besir Sahin Long Term Energy Demand and Supply Projections and Evaluations for Turkey, Esra Ozdemir, Muhsin Kilic Evaluating Exergetic Sustainability Indicators for an Electrolyte Supported SOFC Stack, Adnan Midilli, Ugur Akbulut Life Cycle Assessment of Nuclear Based Ammonia Production Options: A Comparative Study, Yusuf Bicer, Ibrahim Dincer Energy and Exergy Efficiency Evaluations of R134a Clathrates with Additives for Cooling Applications, Sayem Zafar, Ibrahim Dincer, Mohamed Gadalla Thermodynamic Performance Analysis of a Raw Mill System in Cement Plant, Mehmet Altinkaynak, Murat Ozturk, Ali Kemal Yakut THERMAL SYSTEMS AND APPLICATIONS Exergetic Assessment of PTSC Integrated Power-Refrigeration System Working with CO 2, Ahmet Kabul, Onder Kizilkan Cooling of Concentrated Photovoltaic System Using Microchannel Heat Sink, Ali Radwan, Mahmoud Ahmed, Shinichi Ookawara Thermodynamic Analysis of Parabolic Solar Collector Driven Double-Effect Absorption Cooling System, Fatih Yigit, Ahmet Kabul, Onder Kizilkan Energy and Exergy Analyses of a Biomass Fired Regenerative ORC System, Ozum Calli, Can Ozgur Colpan, Huseyin Gunerhan vi

8 Transient Analysis of an Absorption Solar Refrigerator with External and Internal Irreversibilities, Yasmina Boukhchana, Ali Fellah, Ammar Ben Brahim A Study on Adsorption Characteristics of Activated Carbon-R134a and Activated Carbon- R404a Pairs, Muhsin Kilic, Ersan Gonul Performance Investigation of a Geothermal Powered Organic Rankine Cycle for Natural Working Fluids, Mustafa Alptekin, Onder Kizilkan, Ahmet Kabul, Resat Selbas An Experimental Investigation on Exergy Analysis of an Ejector Expansion Refrigeration System, Nagihan Bilir Sag, Halil Kursad Ersoy, Arif Hepbasli Thermodynamic Assessment of Ozone Friendly Cascade Refrigeration System Using Natural Refrigerants, H. Cenk Bayrakci, Onder Kizilkan, Ahmet Kabul, Selin Cekin Thermodynamic Analysis of an Integrated System with A Concentrating Collector for Multi-Generation Purposes, Yunus Emre Yuksel, Murat Ozturk Heat Recovery Analysis of a Rotary Kiln in Cement Industry, Ahmet Yakup Cumbul, Mehmet Akif Ezan, Ismail Hakki Tavman, Arif Hepbasli, M. Ziya Sogut SOLAR ENERGY Experimental Analysis of Latent Thermal Energy Storage for Solar Heating Applications: Preliminary Results, Onder Kizilkan, Ahmet Kabul, Sefika Yildirim, Gamze Yildirim A review of Solar Energy Status in Iraq and Current Status, Ahmed Emad, Ahmet Kabul, Onder Kizilkan Effect of Solar - Geothermal Heat Exchanger Design and Fluid Type on the Thermodynamic Performance of a Power Plant, Anil Erdogan, Duygu Melek Cakici, Can Ozgur Colpan Thermal Regulation Enhancement of Concentrated Photovoltaic Systems Using Phase- Change Materials, Mohamed Emam, Mahmoud Ahmed, Shinichi Ookawara Solar Radiation Exergy and Enviroeconomic Analysis for the West Black Sea Region in Turkey, Yusuf Kurtgoz, Emrah Deniz Comparison of Regression Analysis, ANN and ANFIS Methods in the Prediction of Monthly Mean Global Solar Radiation: A Case Study, Yusuf Kurtgoz, Emrah Deniz Reduction of Entropy Production by Using of Solar Cooling Systems Based on SOLITERM Parabolic Trough Collectors Combined with Double Effect Absorption Chillers, Ahmet Lokurlu Optimal Off-Design Conditions for Solar-Driven Organic Rankine Cycles, Caglan Sevinc, Eray Uzgoren One-Dimensional Transient Thermal Model for Parabolic Trough Collectors Using Closed- Form Solution of Fluid Flow, Eray Uzgoren Design and Performance Analysis of Linear Fresnel Reflector, Melik Ziya Yakut, Arif Karabuga, Ahmet Kabul, Resat Selbas Exergy Analysis of a Solar Photovoltaic Panel within Karabük Climate Conditions, Mutlucan Bayat, Mehmet Ozalp A Comparative and Experimental Study on Different Exergetic Efficiency Methods of a Solar Panel, Mehmet Ozalp, Mutlucan Bayat Solar Assisted Multi-Generation System Using Nanofluids: A Comparative Analyzes, Muhammad Abid, Tahir A. H. Ratlamwala, Ugur Atikol New Climate Zone Definitions of Turkey by Using Typical Meteorological Year Data, Serpil Yilmaz, İsmail Ekmekci vii

9 Exergetic and Energetic Performance Evaluation of a Flat Plate Solar Collector in Dynamic Behavior, Hamed Mouna, Ben Brahim Ammar Optimization of Tilt Angles of PV Arrays for Different Seasons, Ahmet Senpınar Key Factors for the Operation of a Solar Air Collector: A Parametric Study, Ahmet Caglar, Mustafa Burak Bahadir Experimental Analysis of Solar Space Heater Performance, Guvenc Umur Alpaydin, Serhan Kucuka SUSTAINABLE AND RENEWABLE ENERGY DEVELOPMENT Load Side Management in Smart Grids using a Global Optimizer, Abdelmadjid Recioui, Mossaab Djehaiche, Abderrahim Boumezrag Performance Assessment of Various Greenhouse Heating Systems; a Case Study in Antalya, M. Tolga Balta, Fatih Yilmaz, Resat Selbas Cost Risk Modeling for Hybrid Power Generation from Geothermal, Biomass Resources and CSP in Turkey - Southeastern Anatolia and Eastern Anatolia Region, Yildirim Ismail Tosun Sustainable Re Use of Dairy Cow Manure as Bedding and Compost: Nutrients, Pathogens and Self-Heating Potential from Increased Residence Time in a Tumbling Composter, Joe Ackerman, Ehsan Khafipour, Nazim Cicek Cost Modeling for Thermal Energy Storage in Hybrid Power Generation from CSP and Biomass Resources in Turkey - Southeastern Anatolia and Eastern Anatolia Region, Yildirim Ismail Tosun Vacuum Stripping Membrane Desalination for Marmara Sea-Water, Filiz Ugur Nigiz, Nilufer Durmaz Architecture in the Net Zero Houses of the Future, Okay Gonulol, Ayca Tokuc Control System for a Novel Photobioreactor in the Building Envelope, Gulden Kokturk, Ayca Tokuc, Anil Unal Energy and Exergy Analysis of a Solar Air Heater Having Transverse Wedge Shaped Rib Roughness, Cihan Yildirim, Ismail Solmus Performance Analysis of Three Soft Computing Methods for Predicting the Heat Load of Buildings, Cihan Turhan, Tugce Kazanasmaz, Gulden Gokcen Akkurt Ventilation Strategies for the Preventive Conservation of Manuscripts in Necip Paşa Library, İzmir-Turkey, Turgay Coskun, Cem Dogan Sahin, Ozcan Gulhan, Zeynep Durmus Arsan, Gulden Gokcen Akkurt Investigation of Thermodynamic and Environmental Performance Based on Subcooling of Refrigerants in Direct Expansion System for Supermarket Applications, Onder Altuntas, M. Ziya Sogut, Enver Yalcin, T. Hikmet Karakoc Techno-Economic Assessment of Solar-Geothermal Based Multigeneration System for a Community, Farrukh Khalid, Ibrahim Dincer, Marc A. Rosen PV Array Based Smart Home Automation System, Ahmet Senpinar Energy and Exergy Analyses of a Solar Energy Driven Multigeneration System for Green Buildings, Yunus Emre Yuksel, Murat Ozturk, Ibrahim Dincer Achieving Sustainable Buildings via Energy Efficiency Retrofit: A Case Study of an Industrial Building, Muhsin Kilic, Ayse Fidan Altun Passive Thermal Management of a Photovoltaic Panel: Influence of Fin Arrangements, Ceren Yuksel, Cem Kalkan, Mustafa Aydin, Güven Nergiz, Mehmet Akif Ezan viii

10 Multi-Criteria Selection Factors for Evaluation of Intelligent Buildings; A Novel Approach for Energy Management, Elnaz Asadian, Katayoun Taghizadeh Azari, Ali Vakili Ardebili, Samira Mahmoodkelayeh HYDROGEN GENERATION, STORAGE AND TECHNOLOGY Effect of Nitrogen Doping on Hydrogen Storage of Graphene-TiO 2 Nanocomposites, Zahra Gohari Bajestani, Alp Yurum, Omid Akhlaghi, Yuda Yurum Performance Assessment of Solar-Based Hydrogen Production via H 2SO 4 Cycle, Fatih Yilmaz, M. Tolga Balta, Resat Selbas Exergoeconomic and Optimization of a Solar Based Integrated Energy System for Hydrogen Production, Shoaib Khanmohammadi, Parisa Heidarnejad, Nader Javani, Hadi Ganjehsarabi Exergy Analysis and Optimization of a Solid Oxide Electrolysis Cell for Hydrogen Production, Abdullah A. AlZahrani, Ibrahim Dincer Energy and Exergy Analyses of a Solar Based Hydrogen Production and Compression System, Hasan Ozcan, Ibrahim Dincer Energy and Exergy Analysis of a Novel Combined System Producing Power, Water and Hydrogen, Kiyan Parham, Hamed Alimoradiyan, Mohsen Assadi Energy and Exergy Analyses of a Solar, Wind and Geothermal Based Integrated System for Hydrogen Production, Abbas Alpaslan Kocer, Murat Ozturk Exergy Based Environmental Effect of PEM Electrolyser Integrated Hydrogen Gas Storage System, Mert Ozsaban, Selcuk Inac, Adnan Midilli FUEL CELLS Studying the Effect of Electrolyte Thickness on Exergetic Performance for an Electrolyte Supported SOFC Stack, Ugur Akbulut, Adnan Midilli, Ibrahim Dincer Multiphase Non-Isothermal Modeling of a Flowing Electrolyte - Direct Methanol Fuel Cell, Faruk Atacan, David Ouellette, Can Ozgur Colpan Multi-Inlet- Multi-Outlet Anode Flow Field Design for Micro Direct Methanol Fuel Cells, Radwan M. El-Zoheiry, Mahmoud Ahmed, Shinichi Ookawara Synthesizing and Testing of Nafion/SiO 2 and Nafion/TiO 2 Composite Membranes for the DMFC Applications, Mustafa Ercelik, Adnan Ozden, Yilser Devrim, Can Ozgur Colpan Evaluation of Design and Performance of Two Different Power Systems for a Small UAV, Mohamed Gadalla, Sayem Zafar Effect of Cathode Flow Field Configuration on the Performance of Flowing Electrolyte- Direct Methanol Fuel Cell, Ugur Gencalp, David Ouellette, Can Ozgur Colpan The Effects of Three Different Coating Techniques on the Performance of DMFCs, Adnan Ozden, Mustafa Ercelik, Yagmur Nalbant, Hasan Kiyik, Can Ozgur Colpan The Effects of Bio-Inspired Flow Field Design on the Performance of DMFCs, David Ouellette, Adnan Ozden, Mustafa Ercelik, Can Ozgur Colpan FLUID MECHANICS, HEAT AND MASS TRANSFER Diesel-Like Fuel from Waste Engine Oil by Thermo-Catalytic Pyrolysis, Tarabet Lyes, Maamouri Mohamed, Zouad Youcef, Khiari Karim, Mahmoud Rachid, Mohand Tazerout 558 MHD Natural Convection and Entropy Generation in a Nanofluid Filled Cavity with a Conductive Partition, Fatih Selimefendigil, Hakan F. Oztop ix

11 CFD Simulations to Optimize Flow Distribution in a FGD Wet Scrubber, Osman Gozutok, Murat Baranak, Goktug N. Ozyonum, Asli I. Kaya Determination of Flow Characteristics of Multiple Slot-Jets Impingement Cooling, Nuri Kayansanayan, Ersin Alptekin, Caner Erdogan Experimental and Numerical Investigations of Heat Transfer in Multi-Port Tubes, Kemal Ermis, H. Ibrahim Coban, Mehmet Coban Effect of Length of the Wavy Shaped Splitter Plate on Flow around a Circular Cylinder, Mustafa Sarioglu, Mehmet Seyhan, Yahya Erkan Akansu Investigation of the Effect of the Plasma Actuators Vertically Placed On the Spanwise of NACA0015 Airfoil, Hurrem Akbiyik, Hakan Yavuz, Yahya Erkan Akansu Second Law Analysis of Coupled Heat and Mass Transfer through Combined Non Gray Gas Radiation within a Cylindrical Annulus, Sakly Abir, Mazgar Akram, Ben Nejma Faycal A Numerical Study on Phase Change inside a Spherical Capsule, Ersin Alptekin, Muhammet Ozer, Murat Top, Nuriye Bozkurt, Muruvvet Zenginoglu, Fazil Erinc Yavuz, Mehmet Akif Ezan Phase Change Materials in Textile Fabrics: A Numerical Survey, Mehmet Akif Ezan, Berkant Murat Gul, Hüseyin Kurt, Atif Canberk Ezan, Ersin Alptekin Experimental Investigation of a Panel Radiator with Latent Heat Storage, Guvenc Umur Alpaydin, Serhan Kucuka Energy and Exergy Analyses of a Hybrid Solar-Geothermal Power Plant, Duygu Melek Cakici, Anil Erdogan, Can Ozgur Colpan FUELS AND COMBUSTION TECHNOLOGY MW Hybrid Power Generation in ORC Unit from Co-Incineration of Agricultural, Forestry Biomass Waste and Biogas in Stoker and Through Parabolic Solar Panel (CSP), Yildirim Ismail Tosun Optimization Methods of Radiative Transfer Calculation Applied to a Cylindrical Sodium Vapor Plasma, Soumaya Hadj Salah Waste to Energy with a Combine Membrane Technology: Biobutanol Production and Purification, Filiz Ugur Nigiz, Nilufer Durmaz Hilmioglu Biodiesel Production from High Acid Value Sunflower Oil By Using Zirconium Sulfate as a Heterogeneous Acid Catalyst, Melike Imge Senoymak, Oguzhan Ilgen Biodiesel Production over CaMgAl Hydrotalcite like Compounds from Waste Cooking Oil, Emine Emel Cakirca, Ayse Nilgün Akın Production of Upgraded Bio-Oils by Biomass Catalytic Pyrolysis Using Low Cost Food Industry Waste, Nurgul Ozbay, Adife Seyda Yargic, Rahmiye Zerrin Yarbay Sahin, Elif Yaman Numerical Investigation of Fixed Bed Downdraft Woody Biomass Gasification, Ebubekir Siddik Aydin, Ozgun Yucel, Hasan Sadikoglu Influence of Boron Loading Sequence on HDS Catalyst Activity, Yesim Dusova-Teke, Esra Yonel-Gumruk, Orhan Ozcan, M. Efgan Kibar, A. Nilgun Akin Production of a Low-Sulfur Oil from Scrap Tires Pyrolysis Using a Two-Stage Pyrolysis Process and Additives, Gyung-Goo Choi, Young-Kon Choi, Joo-Sik Kim Utilization of Kayseri-Menteş Iron Ore as Oxygen Carrier in Chemical Looping Combustion of Syngas: Deconvolution of the Gas Analysis Data, Nesibe Dilmac, Omer Faruk Dilmac x

12 Pyrolysis of Waste Polyethylene Plastics and Investigation of the Fuel Potential of Pyrolysis Products, Merve Sogancioglu, Esra Yel, Gulnare Ahmetli Pyrolysis of Washed Waste HDPE Plastics and Production of Epoxy Composite from the Pyrolysis Char, Merve Sogancioglu, Esra Yel, Gulnare Ahmetli Production of Hazelnut Oil Biodiesel through Investigating KOH-Catalyzed Transesterification Reaction Parameters, Mert Gulum, Atilla Bilgin Second Law Analysis of a CI Engine Fueled With Biodiesel-Diesel Blends, Abdulvahap Cakmak, Atilla Bilgin Coupled Diesel Engine Hazardous Emissions Fixation, and Microalgae Biomass Production Enhancement, D. O. Correa, B. Santos, J.V.C. Vargas, A.B. Mariano, W. Balmant, M.P. Rosa, D.C. Savi, V. Kava, J.C. Ordonez Regression Models for Predicting Some Important Fuel Properties of Corn Oil Biodiesel- Diesel Fuel Blends, Atilla Bilgin, Mert Gulum Rational and Hyperbolic Models to Estimate Kinematic Viscosities of Hazelnut Oil Biodiesel-Diesel Fuel Blends, Mert Gulum, Atilla Bilgin Experimental Investigation of the Effects of Water Adding to the Intake Air on Diesel Engine Performance and Heat Release Analysis, Zehra Sahin, Orhan Durgun, Mustafa Tuti Emission due to Pollution from Ships Main Engine and Auxiliary Machinery, Munir Suner POSTER PRESENTATIONS Heat and Mass Transfer in a Composite Fluid-Porous Layer, Noureddine Hadidi, Ziane Farouk, Rachid Bennacer, Yacine Ould-amer Fuzzy Control of the Compression System by the Throttle and Coupled Valves in Petroleum Companies, Razika Zamoum Boushaki, Tarik Boushaki, Farida Kessal Experimental Study and Energy Optimization of a Solar Domestic Refrigerator Incorporating a Phase Change Materials, Tetbirt Ali, Mokrane Mehdi, Abbas Mohammed, Berdja Mohand, Ferhat Yahi Removal of Ions Pb 2+ and Cd 2+ from Aqueous Solution by Containment Geomaterials, Souhila Ait Hamoudi, Boualem Hamdi, Jocelyne Brendle The Dissolution Behavior of Lead Oxide in Aqueous Organic Acid Solutions, S. Bendebane, S. Djerad, L. Tifouti Effects of Parameters on the Extraction Yield of Acid Orange10 by Elm from an Aqueous Solution: Application of Plackett-Burman Design, Farida Bendebane, Lynda Bahloul, Hazem Meradi, Mohammed Saddek Lachgar, Abbes Boukhari, Fadhel Ismail Effect of Biopolymer on the Properties of Oil-In-Water Microemulsions, Nedjhioui Mohammed, Moulai Mostefa Nadji, Tir Mohamed Taguchi Optimization Approach for Methyl Orange Removal from Aqueous Solution Using Electrochemical Process, Mohamed Tir, Mohamed Nedjhioui Multilayer Perceptron Model for Predicting Acute Toxicity of Fungicides on Rats: Validation and Domain of Application, Hamadache Mabrouk, Benkortbi Othmane, Hanini Salah, Amrane Abdeltif Thermoeconomic and Enviroeconomic analysis of ISCCS in Algeria, Tarik Boushaki, Pr. Kacem Mansouri Effect of Pb Content and Heat Treatment on Thermoelectric Properties of AgPb 18+xSbTe 20 alloys, Sheng-Long Lee, Jo-Kuang Nieh, Yu-Chih Tzeng xi

13 Thermodynamic and Kinetic Studies for the Adsorption of Amoxicillin onto Modified Wheat Grains, Othmane Benkortbi, Asmaa Boukhelkhal, Mabrouk Hamadache, Salah Hanini A Novel Technique for the Production of Fuel Bioadditive Ethyl Levulinate: Green Process by the Catalytic Membrane, Derya Unlu, Nilufer Hilmioglu Biodiesel Synthesis by Using the Smart Catalytic Membrane, Derya Unlu, Aynur Hacioglu, Nilufer Hilmioglu Characterization of Bio-Oil Obtained from a Food Industry Waste Pyrolysis, Nurgul Ozbay, Elif Yaman, Adife Seyda Yargic, Rahmiye Zerrin Yarbay Sahin Synthesis Gas Production from Tri-Reforming and Partial Oxidation of Simulated Biogas over Ni/ZrO 2-MgO-Al 2O 3, Merve Dogan, Emel Engintepe, Orhan Ozcan, Murat Efgan Kibar, Ayse Nilgun Akin Partial Oxidation of Biogas for Hydrogen Production over Ce-Promoted Ni/Mgal Hydrotalcite-Like Catalyst, Emel Engintepe, Merve Dogan, Orhan Ozcan, Murat Efgan Kibar, Ayse Nilgun Akin Desalination in Algeria, Case of Skikda Seawater Desalination Plant, Mounira Rouainia, Karima Mehri Performance Enhancement of Ni-based Oxygen Carrier by Adding Other Oxygen Carrier, Ho-Jung Ryu, Dong-Ho Lee, Chang-Keun Yi, Sung-Ho Jo, Jeom-In Baek Continuous Operation Results of 263 kwth Chemical Looping Combustor, Ho-Jung Ryu, Dong-Ho Lee, Gyoung-Tae Jin, Seung-Yong Lee, Jeom-In Baek Rehabilitation Alternatives for Flue Gas Desulfurization Units, Asli Isler Kaya, Fatih Aydin, Mustafa Malkoc, Savas Altinisik, Omer Orcun Er Effect of Zeolite Supported Iron Catalyst on Upgrading of Pyrolysis Bio-Oil, Elif Saracoglu, Esin Apaydin-Varol, Basak Burcu Uzun Air Gasification of Dried Sewage Sludge Using a Multi-Stage Gasifier: Effects of the Equivalence Ratio and Activated Carbon on Tar Removal, Young-Kon Choi, Gyung-Goo Choi, Joo-Sik Kim Life Cycle Assessment of a Maintenance Process for a Training Aircraft, Yasin Sohret, Selcuk Ekici, Onder Altuntas, T. Hikmet Karakoc Development and Multi Objective Exergy Based Optimization of a Solar Micro CHP System Based on Organic Rankine Cycle for domestic applications, Alireza Noorpoor, Parisa Heidarnejad, Shoaib Khanmohammadi, Nader Javani Evaluation of Bio oils produced from Pomegranate Pulp Catalytic Pyrolysis, Eylem Pehlivan, Nurgul Ozbay Effect of Air Exchange Rate on the Economic Outputs of Aircraft Environmental Control Systems, Ramazan Atilgan, M. Ziya Sogut, Onder Turan Interactions between Polysaccharide and Anionic Surfactant and Their Effects on the Interfacial and Rheological Behaviours, Nedjhioui Mohammed, Moulai Mostefa Nadji, Tir Mohamed, Skender Abdelhak Numerical Study of Latent Heat Thermal Energy Storage inside a Porous Matrix, Mehdi Fetiti, Amel Alidrous CD 4 effective Hamiltonian in order 6 for the Pentad (2ν 4, ν 2+ν 4, ν 1, 2ν 2, ν 3). A simultaneous Line Position analysis of GS-GS, Dyad (ν 2, ν 4)-Dyad, Dyad-GS and Pentad- GS, Ouardi Okkacha, Kaarour Abdelkrim Study of the ν 3 Fundamental Band of 12 CD 4, Kaarour Abdelkrim, Ouardi Okkacha xii

14 Exergy Analysis of Benzene Production Cycle, Masoud Taghavi, Gholamreza Salehi, Rasoul Hajibabaei xiii

15 ENERGY, ENTROPY AND EXERGY ANALYSIS AND MANAGEMENT 1

16 A Study on Exergetic Performance of Afşin Lignite Stoichiometric Combustion Process Sefa Yalcin 1*, Alp Er Ş. Konukman 2, Adnan Midilli 3 1 Energy Institute, TUBITAK Marmara Research Center, Gebze 41470, Kocaeli, Turkey 2 Gebze Technical University, Department of Mechanical Engineering, Gebze 41400, Kocaeli, Turkey 3 Recep Tayyip Erdoğan University, Department of Mechanical Engineering, Rize, Turkey * Abstract This study aims to investigate parametrically the energetic and exergetic performance of Afşin lignite combustion process. In this regard, in terms of the First law and the Second law of thermodynamics, the energy and exergy analyses have been achieved by using the complete combustion (air fuel ratio, λ=1) reaction and the proximate and ultimate analyses results of Afşin lignite samples taken from Afşin basin that is the largest lignite basin of Turkey. For these purposes, a schematic model system to produce process steam (151 o C, 5 bar) and hot water (90 o C) for practical applications has been developed, which consists of combustion chamber, steam production heat exchanger, hot utilization water production heat exchanger, recuperator for combustion air heater and chimney. The combustion chamber exit gas temperatures are taken to be 1000, 1200 and1300 K, respectively. Meanwhile, Afşin lignite adiabatic flame temperature has been calculated to be 1327 K. However, maximum reaction temperature is taken to be 1300 K for the analyses. Accordingly, it is determined that, in order to increase the energetic and exergetic performances of Afşin lignite combustion process, the heat losses resulting from the convectional, conductional and radiational irreversibilities should be minimized by adducting the practical flame temperature to the adiabatic flame temperature of Afşin lignite and increasing the useful energy from the process. Keywords: Afşin lignite, energy, exergy, efficiency, combustion. I. Introduction Afşin-Elbistan lignite basin constitutes 38% of the total lignite reserves with 4,8 billion tonnes reserves in Turkey (EÜAŞ Yıllık Rapor, 2014). Afşin lignite basin that is the largest lignite basin in Turkey has coals whose caloric values are between 900 and 1250 kcal/kg (World Energy Council Energy Report 2013). Therefore, it is inevitable to do performance improvements for lignite combustion system. In order to determine the performance of the lignite combustion system thermodynamic analysis should be done. The first law of thermodynamics does not give any information about irreversibilites and degredations, occurring in the system. Therefore, energy analysis of industrial systems is needed to be performed as well as exergy analysis (Ohijeagbon IO, 2013). Exergy analysis is a result of the second law of thermodynamics and it is a method used to determine the useful work potential of given amount of energy at some special cases. Exergy analysis is commonly used in thermal and thermo-chemical system design, simulation and performance assessment (Saidur, 2010). efficient energy system design and energy resources use on environment. In Turkey, steam boilers are used in many sectors of industry. Textiles, paper, sugar, tea and many other industries, for use in production processes, boilers for steam generation is used. Most of steam boilers used in industry utilize fossil fuels. According to the percentage use of fossil fuels for steam generation sector; food procesesing (57%), paper processing (%81), chemical production (42%), oil refining (23%) and primary metal production (10%) is realized (Saidur R, 2010). The largest share in the food processing sector, sugar and tea processing plants are located. Tea processing plants need a lot of process steam during wet tea processing. ÇAYKUR produced 65% of the dry tea in Turkey, also private sector produces 35%. ÇAYKUR has 46 fresh tea processing factory also private sector has 230. Production capacity of Çaykur is 6760 ton/day, also private sector s production capacity is 8746 ton/day (Turkey Black Tea Sector Report, 2009). According to Dincer et al. (2003) exergy analysis is an important tool to determine the impact of more 2

17 II. Main Considerations II.I. System Design and Operating Principle In this work, a facility whose daily fresh tea production capacity is 280 t/h is considered, which has five steam boilers. The desired properties of steam to be obtained from the system are given in Table 1 (Korkmaz, 2012). Tab. 1: Properties of steam needed tea processing plant Steam pressure 5 bar Steam Temperature C Steam amount 11 t/h A schematic model system to produce process steam (151 o C, 5 bar) and hot water (90 o C) for practical applications has been developed, which consists of combustion chamber, steam production heat exchanger (HeX-1), hot utilization water production heat exchanger (HeX-2), recuperator for combustion air heater (R-1) and chimney (Fig. 1). Coal properties used in the study were obtained by doing proximate and elemental analysis of samples taken from the Afşin basin. Proximate and elemental analysis of coal are conducted in accordance with ASTM D 1372 and ASTM D 3176 standards, respectively. Coal properties are presented in Table 2. Tab. 2: Proximate and elemental analysis of Afşin lignite Sym Value Unit c 15,05 %w h 1,48 %w o 4,80 %w n 0,37 %w s 1,98 %w w 46,43 %w a 29,88 %w LHV 919 kcal/kg HHV 1250 kcal/kg (c; carbon content, h; hydrogen content, o; oxygen content, n; nitrogen content, s; sulphur content, w; moisture content, a; ash content) Fuel W F1 Q 1 Q Fresh Air 25 C, 1 atm Air Preheater Hot Air 300 C, 1 atm Fuel 25 C Combustion Chamber CC Flue gas T stack, 1 atm Flue gas Tstack, 1atm Ash Stack W F3 Flue gas 150 C, 1 atm R - 1 Hot Air 300 C, 1 atm WF2 Flue gas T HeX 2,out Utilization Water 90 C HeX - 2 Feed Water-2 20 C HeX - 1 Flue gas T stack,hex-1,out Feed Water-1 20 C Process Steam 151 C Fresh Air 25 C, 1atm Fig. 1: A schematic representation of the coal-fired steam generating plan 3

18 Closed formula for Afşin lignite is determined at equation (6) by using values obtained from the results of the analysis (El-Wakil,1984, Bejan et al., 1996). n C = %c, n M H = %h n C M O = %o H M O n S = %s n M N = %n (1) S M N n top = n C + n H + n O + n N + n S (2) n C = n C n H = n H n n O = n O top n (3) top n top n N = n N n n S = n S top n top C nc H nh O no N nn S NS + γo min (O 2 + 3,762N 2 ) aco 2 + bh 2 O + cso 2 + dn 2 (4) where, γ; excess air coefficient (for stoichiometric combustion is taken 1). III.1. Energy Analysis In this work, energy analyses of each equipment individually and the whole system are performed according to the 1 st law of thermodynamics. a) Combustion Chamber Energy Analysis Energy production in coal fired steam and hot water producing system is provided by the transformation of chemical energy of coal burned with atmospheric air into heat energy. Combustion air with energy m m E air and fuel with energy E fuel enter into the combustion chamber and they perform combustion reaction and then they leave the system with the m energy E flue gas. Heat transfer occurs by the amount Q of E radiation and E ash m from the combustion chamber to the outside (Fig. 2.). E Q Radiation O min = n C + n H 4 + n S n O 2 (5) E m fuel C 0,4025 H 0,4728 O 0,0964 N 0,0085 S 0, ,4924(O 2 + 3,762N 2 ) 0,4025CO 2 + 0,2364H 2 O + 0,01985SO 2 + 1,857N 2 (6) E m air Control Volume CV E m fg The fuel fed into the system is burned with stoichiometric conditions in the combustion chamber. Process steam is obtained by transferring the the energy of the resulting combustion gas to the feed water at at 20 C in the HeX-1-exchanger. The flue gas releasing the some fraction of its energy in the HeX-1-exchanger produces the proses water at 90 C with its a portion of energy by entering the HeX- 2 that is used for production of the hot water used in the system. Then the flue gas releasing the some fraction of its energy in the HeX-2-exchanger is fed to R1 recuperators where the combustion air temperature is increased necessary for the combustion of the fuel. Finally, the flue gas that increases the entering fresh air temperature from 25 C to 300 C in the recuperator is discharged to the atmosphere from the chimney at 150 C due to the fact that condensation occurs at below temperatures. III.Analysis Assumptions The assumptions made in this energy and exergy analyses of the system; 1. The system is constantly open continuous flow 2. Heat transfers from exchanger to the environment are neglected. 3. The fans used in the system are chosen to be 1 kw. 4. Subsystems and processes are considered steady state. 5. Combustion air consist of 21 % O2 and 79 % N2 and it is assumed as an ideal gas. E m ash Fig. 2: Combustion chamber energy flow diagram Fuel and air enter into the combustion chamber as the mass. Outlet mass of the combustion chamber consists of the flue gas occurring after combustion and the unburned ash. In this way, For a general steady-state process, mass and energy balances, respectively, can be written as: m in = m out (7) m in = m coal + m air (8) m out = m fluegas + m ash (9) E in E out = 0 (10) E in = E in m + E in W Q + E in (11) E in m = m fuel LHV fuel + m air h air,573 K (12) E in W = 0 (13) E in Q = 0 (14) Mass energy input stem from only air and fuel into the combustion chamber. Any work or heat transfer is not an issue. Similarly, the energy outputs from the system are to be written as follows. 4

19 E out = E out m E out m + E out W Q + E out (15) flue = E gas ash out + E out (16) flue E out gas flue = m flue gas (h gas T CC ) (17) E out ash = ignore (18) E out W = 0 (19) E out Q Q = E Radiation (20) Energy efficiency of the combustion chamber is defined as the ratio of the output useful energy value to input energy value at equation 21. ƞ 1,CC = E flue gas out E in (21) b) HeX 1 Heat Exhanger Energy Analysis The process steam, the main useful output, is obtained by transferring the flue gas energy to the feed water in the HeX-1 heat exchanger. Hot flue gas in with E fg(t) energy entering into the heat exchanger leaves its energy to the feed water entering into the in heat exchanger with E fw(293 K) energy and then out leaves the system with E fg(t HeX1 out) energy. Cold feed water enters into the system leaves the system out as process steam with the energy E ps(424 K) by changing phase (Fig.3). Energy input to the HeX-1 heat exchanger takes place with the hot flue gas from the combustion chamber. Apart from this, there is not any work or interaction. Energy output from the HeX-1 is provided with the flue gas that have transferred a portion of its energy to the feed water for steam production and the generated process steam. The energy balance equation for the HeX-1; E in E out = 0 (24) m fg h fg(t) m fg h fg(t HeX1 out) = m ps h ps(424 K) m fw h fw(293 K) (25) Q HeX1 = m fg (h fg,1000k h fghex1 out ) (26) h fghex1 out = (Q HeX2 m fg) + h fghex2 out (27) m steam = Q HeX1 (h ps(424 K) h fw(363 K) ) (28) Efficiency expression for HeX-1 is defined in the equation 29 as the ratio of the total useful energy obtained from the heat exchanger and the total energy entering the system. ƞ 1,HeX1 = Q HeX1 E in (29) The process steam mass flow rate as defined in equation 28, is calculated as the main useful output obtained from the system. c) HeX 2 Heat Exhanger Energy Analysis E in fg (T) HeX - 1 E out fg (T-HeX1-out) CV E in fw (20 C) E out ps (151 C) Hot water to be used at the factory is also obtained from the system which is designed for steam generation. The flue gas that has leaved a portion of its energy for steam production in the HeX-1 enters to the HeX-2 heat exchanger for hot water production in with the energy E fg(t HeX1 out) and leaves the heat out exchanger with the energy E fg(t HeX2 out) by leaving a portion of its energy to the feed water that in enters to the system with the energy E fw(293 K). The cold feed water that has entered to the system leaves out the system with the energy E uw(363 K), which takes from the flue gas (Fig. 4). Fig. 3: HeX - 1 heat exhanger energy flow diagram The output hot flue gas from the combustion chamber and the feed water fed for steam production enter to the HeX-1 heat exchanger. The total mass from the system consists of prosess steam, which is obtained in the HeX-1 heat exchanger, and the flue gas. Thereby; The mass balance equation for the HeX-1; E out fg (T-HeX2-out) E out uw (90 C) HeX - 2 CV E in fw (20 C) E in fg (T-HeX1-out) m in = m out (22) m fw + m fg,in = m ps + m fg,out (23) Fig. 4: HeX-2 heat exchanger energy flow diagram 5

20 The output hot flue gas from the HeX-1 and the feed water fed for hot use water production enter to the HeX-2 heat exchanger. The total mass from the system consists of hot use water and the flue gas. Thereby; The mass balance equation for the HeX-2; m in = m out (30) m fw + m fg,(t HeX1 out) = m uw + m fg,(t HeX2 out)(31) The energy balance equation for the HeX-2; E in E out = 0 (32) Efficiency expression for HeX-2 is defined in the equation 37 as the ratio of the total useful heat obtained from the heat exchanger and the total energy entering the system. ƞ 1,HeX2 = Q hex2 E in d) R1 Recuperator Energy Analysis (37) Combustion air is fed into the combustion chamber at 573 K for the purpose of increasing the combustion efficiency in the coal combustion system. The flue gas exits from the HeX-2 which is produced hot out utilization water with the energy E fg(t HeX2 out) and enters into the R1 in order to increase the fresh air temperature to 573 K. The flue gas leaving a portion of its energy leaves from the system with the energy out E fg(t 423 K) and is thrown from the chimney into the atmosphere. In order to prevent the flue gas from condensation in the chimney, the flue gas exit temperature is kept constant at 423 K (Fig.5). E out air,300 C the system consist of hot combustion air quantity derived in the recuperator and the flue gas. Thereby; The mass balance equation for the R-1; m in = m out (38) m air,298 K + m fg,(t HeX2 out) = m air,573 K + m fg,(t 423 K) (39) The energy balance equation for the R-1; E in E out = 0 (40) m air,573 K h air,573 K m air,298 K h air,298 K = m fg h fg(t HeX2 out) m fg h fg(t 423 K) (41) Q R1 = m air (h air,573 K h air,298 K ) (42) Q R1 = m fg (h fghex2 out h fg423 K ) (43) h fghex2 out = Q R1 + h (44) m fg 423 K fg Efficiency expression for R1 is defined in the equation 45 as the ratio of the total useful heat obtained from the recuperator and the total energy entering the system ƞ 1,R1 = Q R1 E in e) Coal-Fired Steam Production System Energy Analysis (45) According to the first law of thermodynamics, energy analysis of the common system was carried out. The energy efficiency of the system is defined as the ratio of total produced useful output to the total entering energy value and is presented in the equation 46. ƞ I,Sys = Q HeX1+Q HeX2+Q R1 E in (46) III.2 Exergy Analysis E out fg (T-150 C) R1 E in air,25 C CV E in fg (T-HeX2-out) E in w,f1 Fig. 5: R1 recuperator energy flow diagram The output hot flue gas from the HeX-2 and the atmospheric air fed for combustion air production at 573 K enter to the R-1 recuperator. The total mass of Second law of thermodynamics, which describes the exergy term, defines the quality of energy as well as the quantity of energy. The final state of the system is called as dead state. This is the state that kinetic and potential energy exchanges are zero (Taner T, 2015). As the thermodynamics point of view, exergy is described as the maximum work amount that can be produced by the system at the reference ambient conditions. Exergy is not protected like energy. it is consumed due to irreversibilities or destroyed in the real prosess (Dincer I, 2003). A general exergy equation consists of different form of exergy like kinetic and potential exergy, physical and chemical exergy and exergy of radiation (Amelio A, 2016). General exergy equation consists of physical and chemical exergy at steady-state condition with negligible potential and kinetic exergy 6

21 changes for combustion process (Ohijeagbon O, 2013). Physical exergy is defined as the maximum theoretical useful work that can be obtained from a flowing stream as it is brought to the environmental state (Riverio R, 2006; Dincer I, 2003). According to Dincer and Rosen (2007), the physical exergy can be calculated as follows; E x ph = m [(h h 0 ) T 0 (s s 0 )] (47) The chemical exergy is defined as the maximum theoretical useful work that can be obtained from a flowing stream as it is brought from the environmental state to the dead state (Ohijeagbon Ol, 2013; Bilgen S, 2008; Dincer I, 2003). individually and the whole system are performed according to the 2nd law of thermodynamics. a) Combustion Chamber Exergy Analysis Combustion air enter into the combustion chamber with exergies E x ph ch air and E x air and fuel enter into ch the combustion chamber with exergy E x fuel and they perform combustion reaction and then they leave the system with the exergy E x ph fg. ph Heat transfer occurs by the amount of E x Rad from the combustion chamber to the outside as well as exergy destruction by the amount of E x D due to internal irreversibilities (Fig. 6). Chemical exergy value of pure substances is calculated via standard chemical exergy values taken from tables. Chemical exergy of gas mixtures is to be calculated by standard chemical exergy of pure substances generating the mixture as following; E x m fuel E x Q Radiation ch e x gas mix = x k e x ch k k + R T 0 k x k lnx k (48) where; xk, mole fraction of gas mixture at gas phase; e x k ch, standard chemical exergy values of gases; R, universal gas constant; T0, ambient temperature. E x m air Control Volume CV E x,d E x m fg Industrial fuels occur tens of chemical compounds that are difficult to calculate their chemical exergy. Standard chemical exergy of fuels that their chemical formulations are known can easily be found from the tables. However, some fuels are composed of multiple components and can not be read directly from the tables. Szargut and Styrylska developed a statistical method that represent the chemical exergy of industrial fuels (Midilli A, Chemical Exergy Calculations, 2014; Szargut J and Styrylska T, 1964). According to this method, φ dry colerations are expressed below according to the cases for the o/c ratio of a fuel composed of C, H, O and N: o/c 0.667; φ dry = h c o c n c (49) o/c 2.67; φ dry = h c ( h c ) n c o c (50) Standard chemical exergy value of a moist fuel is expressed as: ch e x fuel = ((NCV) fuel + w + h fg ) φ dry + (e x ch s (NCV s )) s (51) Here; w is the mass fraction of water; s is the mass fraction of sulphur and hfg is the enthalpy of vaporization of water. In this work, energy analyses of each equipment 7 E x m ash Fig. 6: Combustion chamber exergy flow diagram Exergy balance for the combustion chamber; E x in E x out E x D = E x sys (52) For steady-state condition, E x in E x out = E x D (53) Ex in = E x,in W + E Q x,in + E x,in m + E x,in,ke + E x,in,pe (54) Any work or heat transfer does not enter into the combustion chamber. Kinetic and potential energy changes are neglected. Therefore, E x,in W = 0 (55) Q E x,in = 0 (56) as written. Mass and exergy transfer into the combustion chamber takes place by fuel and combustion air. The fuel fed into the combustion chamber has chemical exergy because of being exposured to chemical reactions. Due to the fact that the fuel fed into combustion chamber at dead state, there is no physical exergy of the fuel. Combustion air fed into the combustion chamber at 573 K, so it transports physical exergy into the combustion chamber. Also combustion air reacts with fuel, so it has chemical exergy. Exergy transfer equations with mass are written as follows:

22 E x,in m = E fuel x,in + E air (57) x,in fuel = E E x,in ph E x,fuel ph E xair ch x,fuel ph + E x,fuel (58) = 0 (59) = (hair,573 K h air,298 K ) (s air,573 K s air,298 K )T 0 (60) Ex PH in,fg (T) HeX - 1 CV Ex PH in,fw1 (20 C) e x ch air = [(x O2 e x ch O2 ) + (x N2 e x ch N2 )] + R T 0 [(x O2 lnx O2 ) + (x N2 lnx N2 )] (61) Chemical exergy of combustion air is calculated with the help of values at Table 3. Tab. 3: Standard chemical exergy of air compounds at reference conditions (Bejan vd., 1996) Compounds of Air Molar Flow Rate (kmol/s) Molar Fraction (xk) Standard Molar Chemical Exergy (kj/kmol) O2 0,0153 0, N2 0,0577 0, Total 0, Exergy output from the combustion chamber takes place with the physical exergy of the flue gas and the exergy of the heat transferred by radiation and convection from the combustion chamber surfaces and discharged ash from the system. In this case, the physical exergy value carried by the flue gas which is the useful output obtained from the combustion chamber was calculated according to the equation 63. ph E x,out = E x,flue gas ph E x,flue gas Q rad + E + E x,out Q conv x,out +E x,out ash (62) = m flue gas (h T fg h 0 fg ) T 0 (s T fg s 0 fg ) (63) Exergy efficiency expression for the combustion chamber is defined as the ratio of the physical exergy value of the flue gas, which is the output useful obtained from the combustion chamber, to the total exergy value enters to the combustion chamber and shown in equation 64. ƞ II,cc = E ph x,flue gas fuel E x,in +E x,in air (64) b) HeX-1 Heat Exchanger Exergy Analysis The flue gas from the combustion chamber with the ph exergy E x in,fg(t) and the feed water with the ph E x in,fw1 enter to the heat exchanger for the production of process steam. While flue gas leaves ph the system at the exergy E x out,fg(t HeX1 out), the proses steam leaves the heat exchanger at the ph exergy E x out,ps (Fig. 8). Ex PH out,fg (T-HeX1-out) Ex PH out,ps (151 C) Fig. 8: HeX-1 heat exhanger exergy flow diagram Exergy balance for the HeX-1 heat exchanger; E x in E x out = E x D (65) ph ph ph [E x fg(t) + E x fw1(293 K) ] [E x fg(t HeX1 out) + ph E x ps(424 K) ] = E x D (66) written as that. If instead of the values in the equation are to be written; m fg [(h fg(t) h 0,fg ) T 0 (s fg(t) s 0,fg )] + m fw1 [(h fw1 h 0,fw1 ) T 0 (s fw1 s 0,fw1 )] m fg [(h fg(t HeX1 out) h 0,fg ) T 0 (s fg(t HeX1 out s 0,fg )] m ps [(h ps h 0,ps ) T 0 (s ps s 0,ps )] = E x D (67) The useful physical exergy obtained from the HeX-1 heat exchanger was calculated according to the equation 68. Q E x usefull,hex1 = Q HeX1 (1 T 0 T ps ) (68) The main useful output of the system is equal to the physical exergy value of the steam obtained from the heat exchanger. Exergy efficiency of HeX-1 heat exchanger is defined as the ratio of the useful exergy to the total exergy entering the system and shown in the equation 69. ƞ II,HeX 1 = Q HeX1 (1 T0 Tps ) ph E x fg(t) +E x fw1(293 K) ph (69) c) HeX-2 Heat Exchanger Exergy Analysis The flue gas that gets out from the HeX-1 enters into ph the HeX-2 with the exergye x out,fg(t HeX1 out), and increases the temperature of the water that enters ph into the heat exchanger with the exergy E x in,fw2, and finally provides to obtain the use water with the 8

23 ph exergy E x out,uw. The flue gas that transfers a portion of its own exergy to the feed water at HeX-2, leaves ph the system at the exergy E x out,fg(t HeX2 out) (Fig. 9). gas that transfers a portion of its own exergy to the combustion air at R-1, leaves the system at the ph exergy E x out,fg(t 423 K) (Fig. 10). Exergy balance for the R-1 recuperator; Ėx PH out,fg (T-HeX2-out) HeX - 2 CV Ėx PH out,fg (T-HeX1-out) E x in E x out = E x D (75) ph [E x fg(t HeX2 out) ph ph + E x air(298 K) ] [E x fg(t 423 K) + Ėx PH out,uw (363 K) Ėx PH in,fw2 (293 K) ph E x air(573 K) ] = E x D (76) The useful physical exergy obtained from the R-1 recuperator was calculated according to the equation 77. Fig. 9: HeX-2 heat exchanger exergy flow diagram Exergy balance for the HeX-2 exchanger; E x in E x out = E x D (70) ph [E x fg(t HeX1 out) ph [E x fg(t HeX2 out) ph + E x fw2(293 K) ] ph + E x uw(363 K) ] = E x D (71) If instead of the values in the equation are to be written; m fg [(h fg(t HeX1 out) h 0,fg ) T 0 (s fg(t HeX1 out) s 0,fg )] + m fw1 [(h fw1 h 0,fw1 ) T 0 (s fw1 s 0,fw1 )] m fg [(h fg(t HeX2 out) h 0,fg ) T 0 (s fg(t HeX2 out s 0,fg )] m uw [(h uw h 0,uw ) T 0 (s ps s 0,uw )] = E x D (72) The useful physical exergy obtained from the HeX-2 exchanger was calculated according to the equation 73. Q E x useful,hex2 = Q HeX2 (1 T 0 T uw ) (73) The main useful output of the system is equal to the physical exergy value of the hot utilization water obtained from the heat exchanger. Exergy efficiency of HeX-2 heat exchanger is defined as the ratio of the useful exergy to the total exergy entering the system and shown in the equation 74. ƞ II,HeX 2 = Q HeX2 (1 T 0 Tuw ) ph E x fg(t HeX1 out) +E x fw(293 K) ph (74) d) R-1 Recuperator Exergy Analysis The flue gas that gets out from the HeX-2 enters into ph the R-1 with the exergy E x out,fg(t HeX2 out), and increases the temperature of the fresh water that enters into the recuperator with the exergy ph E x air (298 K), and finally provides to obtain the ph combustion air with the exergy E x air (573 K). The flue 9 Q E x useful,r1 Ėx PH out,fg (T-423 K) = Q R1 (1 T 0 T air,573 K ) (77) Ėx PH air (298 K) Ėx PH air (573 K) R1 CV Ėx PH out,fg (T-HeX2-out) Ėx W f1 Fig. 10: R-1 recuperator exergy flow diagram The main useful output of the system is equal to the physical exergy value of the combustion air obtained from the recuperator. Exergy efficiency of R-1 recuperator is defined as the ratio of the useful exergy to the total exergy entering the system and shown in the equation 78. ƞ II,R1 = Q T0 R1 (1 ) T air,573 K ph E x fg(t HeX2 out) +E x air(298 K) ph (78) e) Coal-Fired Steam Production System Exergy Analysis Coal-fired steam generation system has three main useful output. The first of these is the process steam at 151 C, 5 bar pressure obtained from the HeX-1, the second is the use hot water with 2,5 t/h 90 C obtained from the HeX-2 and the third is combustion air at 1 atm to 300 C fed into the combustion chamber. In order to obtain the beneficial outcomes, coal and combustion air are supplied as the input to the system.

24 Exergy efficiency of the system is defined as the ratio of the useful exergy output to the total exergy entering the system. Thus, exergetic efficiency definition of the coal fired steam production system is expressed in the equation 79. ƞ II,sys = E Q x HeX1 E x,in Q +E x HeX2 air +E fuel +E x,in Q +E x R1 fan (79) x,in IV. Results and discussions In this study, the coal-fired system that are to be produced prosess steam and hot utilization water needed for the tea factory. As an energy resource Afşin lignite is used and analyses has been realized with assumption of stoichiometric combustion. Firstly adiabatic flame temperature is calculated according to ultimate and proximate analysis results of Afşin lignite. Because maximum flue gas temperature obtained with burn of coal related to adiabatic flame temperature. Adiabatic flame temperature of Afşin lignite has been estimated 1327 K. Hence, flue gas exit temperature has been limited to max K. Energy and exergy analysis results for the combustion chamber are reported in Tables 4 and 5. Tab. 4: Combustion chamber energy analysis results T [K] E in [kw] E out [kw] E fg [kw] E rad+conv+ash [kw] ƞ 1,CC % Tab. 5: Combustion chamber exergy analysis results A [K] E x in [kw] E x out [kw] E x D [kw] fuel [kw] E x,in air [kw] ch [kw] E x,in E x,fuel ph E xair ch E xair ph E x,flue gas [kw] [kw] [kw] Ƞ II,CC % According to combustion chamber energy and exergy analysis results has shown that as flue gas exit temperature close to adiabatic flame temperature combustion chamber efficiency increases. Maximum energy and exergy efficiency have been calculated 93% and 43%, respectively. Owing to the fact that Afşin lignite has high ash content even if heat losses stem from the system is prevented exergy efficiency is to be obtained as max. 43%. According to HeX-1 energy and exergy analysis results, is to be obtained maximum steam amount and HeX-1 heat exchanger efficiency for different flue gas exit temperature and using 1 kg/s Afşin lignite are presented in Table 6 and 7. Tab. 6: HeX-1 heat exchanger energy analysis results T [K] in [kw] E fg(t) out E fg(t HeX1 out) [kw] Q HeX1 [kw] m steam [kg/h] ƞ I,HeX1 % Because of the fact that HeX-1 exhanger is gas-fluid exchanger its energy and exergy efficiency is lower than gas-gas or fluid-fluid heat exchanger. Maximum steam production takes place approximately 3,5 t/h with 54% energy and 31% exergy efficiency at flue gas exit temperature with 1300 K in the HeX-1 heat exchanger. Accordingly, system capacity ought to be tripled by using 1 kg/s Afşin lignite for providing steam needs of the tea factory. Tab. 7: HeX-1 heat exchanger exergy analysis results T [K] E x in [kw] E x out [kw] E x D [kw] ph [kw] E x,fg(t) ph [kw] E x,fw1(20 C) ph [kw] E x,fg(t HeX 1 out) ph [kw] E x,ps(151 C) Q E x useful,hex1 [kw] ƞ II,HeX 1 % HeX-2 heat exchanger provides produce hot utilization water at 90 C for internal needs at tea factory. Mean hot water need of tea factory is assumed as 2,5 t/h. This requirement is provided by HeX-2 heat exchanger by using flue gas energy. Hot water production takes place with decrease of flue gas temperature at 23 C and with 9,4% energy and 7,36% exergy efficiency. The main reason of energy and exergy efficiency is so much low by using very high quality exergy resource is obtained to output that has low exergy value. Another reason for low exergy efficiency of HeX-2 heat exchanger is its gasfluid exchanger like HeX-1 heat exchanger. HeX-2 heat exchanger energy and exergy analysis results are presented in tables 8 and 9. 10

25 Tab. 8: HeX-2 heat exchanger energy analysis results T [K] in [kw] E fg(t HeX1 out) out E fg(t HeX2 out) [kw] Q HeX2 [kw] m water [kg/h] ƞ 1,HeX2 % Tab. 9: HeX-2 heat exchanger exergy analysis results T [K] E x in [kw] E x out [kw] E x D [kw] ph [kw] E x,fg(t HeX 1 out) ph [kw] E x,fw2(20 C) ph [kw] E x,fg(t HeX 2 out) ph [kw] E x,uw(90 C) Q E x useful,hex1 [kw] ƞ II,HeX 2 % R-1 recuperator used for the purpose that raising the temperature of the combustion air fed to the combustion chamber in a coal-fired steam generation system. Especially, when combustion air is given the high temperature plays a role in enhancing the combustion efficiency. So, the combustion air fed into the combustion chamber at 573 K in the system. Combustion air entering into R-1 recuperator at 298 K its temperature has been reached 573 K by using flue gas energy with 30% energy efficiency and 63,3% exergy efficiency. R-1 recuperator energy and exergy analysis results are presented in tables 10 and 11. Tab. 10: R1 recuperator energy analysis results T [K] E in [kw] in [kw] E fg(t HeX2 out) out E fg(t 150 C) [kw] Q R1 [kw] in E w,f1 [kw] ƞ 1,R1 % Tab. 11: R1 recuperator exergy analysis results T [K] E x in [kw] E x out [kw] E x D [kw] ph [kw] E x,fg(t HeX 2 out) ph [kw] E x,air (25 C) w [kw] E x,f1 ph [kw] E x,fg(t 150 C) ph [kw] E x,air (300 C) Q E x usefull,r1 [kw] ƞ II,R1 % As a consequence, while coal-fired steam production system energy efficiencies were calculated as 44%, 60% and 66%, respectively, its exergy efficiencies were calculated 13%, 17% and 19%, respectively at combustion chamber exit gas temperatures at 1000 K, 1200 K and 1300 K. Coal-fired steam production system energy and exergy analysis results are presented in tables 12 and 13. Tab. 12: Coal-fired steam production system energy analysis results T [K] E in [kw] Q HeX1 [kw] Q HeX2 [kw] Q R1 [kw] ƞ I,Sys % Tab. 13: Coal-fired steam production system exergy analysis results T [K] E x in [kw] fuel [kw] E x,in air [kw] E x,in fan [kw] E x,in Q E x HeX1 [kw] Q E x HeX2 [kw] Q E x R1 [kw] Q E x useful [kw] ƞ II,sys % V. Conclusion In this study is a parametric study on the energetic and exergetic performance of Afşin lignite stoichiometric combustion process. In this regard, in terms of the First law and the Second law of thermodynamics, the energy and exergy analyses have been achieved by using the stoichiometric combustion (air fuel ratio, λ=1) reaction and the proximate and ultimate analyses results of Afşin lignite samples taken from Afşin basin that is the largest lignite basin of Turkey. The following concluding remarks are made; The overall system exergy efficiency increases with an increase adiabatic flame temperature of the Afşin lignite but in that Afşin lignite has high ash and moisture content and low heating value its adiabatic temperature is to be constrictedly increase. Increasing of quality of the Afşin lignite with enrichment process like ash-free and drying process or using high quality coal requires for increase exergy efficiency of the system. 11

26 Acknowledgements The authors gratefully acknowledge the financial support from The Scientific and Technological Research Council of Turkey (TUBITAK). Nomenclature LHV HHV M Omin T E x ph E x ch E x D s T0 : low heating value (kcal/kg) : high heating value (kcal/kg) : molecular weight (kg/kmol) : minimum oxygen requirements (kcal/kg) : temperature ( C) : physical exergy (kw) : chemical exergy (kw) : exergy destruction (kw) : entropy (kj/kgk) : reference temperature (K) : enthalpy (kj/kg) h Greek letters λ : excess air coefficient (-) : efficiency (%) Superscripts m : mass Q : heat W : work rad : radiation Subscripts max : Maximum min : Minimum in : input out : output fw : output fg : flue gas ps : process steam uw : utilization water sys : system cc :combustion chamber Electricity Generation Company., Annual Report, Turkey, (2014). El-Wakil M.M., Power Plant Technology, 1st edition, Chapter 4, McGraw Hill, (1984). Korkmaz F., Current situation of Turkish tea sector and energy efficiency analysis of a tea factory, MsC Thesis, İstanbul Technical Universty, (2012). Midilli A., Chemical exergy calculations, Summer Course on Exergy and Its Applications, Ankara, June (2014). Ohijeagbon I.O., Waheed M.A., Jekayinfa S.O., Methodology for the physical and chemicial exergetic analysis of steam boilers, Energy 53, (2013). Rivero R., Garfias M., Standard chemical exergy of elements updated, Energy 31, (2006). Saidur R., Ahamed J.U., Masjuki H.H., Energy, exergy and economic analysis of industrial boilers, Energy Policy 38, (2010). Szargut J., Styrylska T., Approximate evaluation of the exergy of fuels, Breennstoff Waerme Kraft 16 (12), (1964). Turkey Black Tea Sector Report., Enterprise Europe network, Turkey (BlackSea), (2009). Taner T., Sivrioglu M., Energy-exergy analysis and optimisation of a model sugar factory in Turkey, Energy 93, (2015). World Energy Council., Energy Report, ISSN: , Ankara (2014). References Amelio A., de Voorde T.V., Creemers C., Degreve J., et. al., Comparison between exergy and energy analysis for biodiesel production, Energy 98, (2016). Bejan A., Tsatsaronis G., Moran M.J., Thermal Design and Optimization, John Wiley, , ISBN : (1996). Bilgen S., Kaygusuz K., The calculation of the chemical exergies of coal-based fuels by using the higher heating values, Applied Energy 85, (2008). Dincer I., Bejan A., Exergy : Energy, Environment and Sustainable Development, 1st Edition, Elsevier Science, (2007). Dincer I., Hussain M.M., Al-Zaharnah I., Energy and exergy use in the industrial sector of Saudi Arabia, Proc. Instn. Mech. Engrs, Vol. 217 Part A : J. Power and Energy (2003). 12

27 Exergy Analysis of Nitrogen Liquefaction Process Arif Karabuga 1 *, Resat Selbas 2, Ahmet Kabul 2 1 Suleyman Demirel University, Keciborlu Vocational School, Electrical Energy Generation, Transmission and Distribution, Isparta, 32260, Turkey 2 Süleyman Demirel Üniversitesi, Faculty of Technology, Energy Systems Engineering, West Campus, Isparta, 32260, Turkey * Abstract Component of air that nitrogen, oxygen and argon is separated and liquefied by cryogenic method. Cryogenics is the science of very low temperature. Conventionally, the field of cryogenics is taken to start at temperatures below 120 K. Cryogenic air separation is the main method to separate air into its components. The nitrogen is used in chemical industry, food freezing, medical purpose, particle accelerators, colliders, synchrotrons, metal processing technology etc. In this study; a real nitrogen liquefaction unit has been examined. This nitrogen liquefaction unit is integrated to an air separation unit. Nitrogen provided by air separation creates the source of liquefaction unit. Energy and exergy analysis of the studied nitrogen liquefaction unit has been done. In numerical calculations and graphics EES (Equation Engineering Solver) software has been used. In results of thermodynamic calculations; exergy efficiency %36, COPactual and COPreversible 0.77 has been calculated. Furthermore, heat-exchanger block in nitrogen liquefaction unit is formed from HE-71, HE-72, HE-73. For each heat exchanger; exergy efficiency has been calculated. Exergy efficiency values are 0.55, 0.81 and 0.89 respectively. Keywords: Air separation unit, cryogenic, nitgrogen, energy and exergy. I. Introduction The earth is surrounded by air. The components of air are nitrogen, oxygen and argon. There is a virtually unlimited supply of nitrogen, oxygen and argon because of their natural occurrence within the atmosphere. Currently several methods are known in air separation. Two processes for air separation exist; cryogenic distillation and non-cryogenics distillation. Cryogenics is the science and techonolgy of very low temperature, usually below 120 K by Weisend II (1998). Non-cryogenic method to include that pressure swing adsorpsition (PSA) and membrane separation. The choice of the process to be used is based on the desired products. Cryogenic air separation is used when product high purity is needed. It is also advantageous when products are required in liquid form by Rizk et al.(2012). The cryogenic systems have the capability to deliver the largest capacities for products and for very high purities. Non-crygenics systems are employed at the lower end of production scale and generally for lower product purities by Castle (2002). In table 1 different air separation processes compared. Tab.1: Compare the process of air separation method by KLM Technology Group (2013) Process Advantages Disadvantages Cryogenic Low amount of electricity per unit nitrogen PSA Produces very high purity nitrogen Can generate liquid nitrogen for storge on site Cost-effective nitrogen production of relatively high purities Quick installation and start-up Membrane Low capital cost Large site space and utility requirements High capital cost Limited scaleability in production Long start-up and shutdown Low to moderate capital cost High maintenance equipment Production output is very flexible Quick installation and start-up Easy to vary purity and flow rate Noisy operation Limited scalability Uneconomical for high purity requirements Uneconomical for large outputs Requires relatively large amount of electricity per unit nitrogen Tab. 2: According to compare the purity values of the air separation unit by Campestrini (2014) Process Purity (%) Cryogenic N: PSA O 2: N: 99,9 up Membrane O 2: 85 99,7 N: 10 ppb * *ppb: Parts per billion 13

28 The largest markets for oxygen are in primary metals production, chemicals and gasification, clay, glass and concrete products, petroleum refineries and welding. The use of medical oxygen is an increasing market. Gaseous nitrogen is used in the chemical and petroleum industries and it is also used extensively by the electronic and metal industries for its inert properties. Liquid nitrogen is used in applications ranging from cryogenic grinding of plastics to food freezing. Argon, the third major component of air, finds uses as an inert material primarily in welding steelmaking, heat treating and in manufacturing processes for electronics by Vinson (2006). Otherwise liquid nitrogen is used in physic applications, particle accelerators, colliders, synchrotrons by Thomas et al. (2011). An exergy analysis is carried out to analyze the possibilities of fuel saving in the cryogenic distillation process. It is shownthat more than half of the exergy loss takes place in the liquefaction unit and almost one-third in the air compression unit. Exergy loss in the compressor is reduced by improving by Cornelissen and Hırs (1998). Amin et al. (2014) made simulation of nitrogen separation from air. In this study, liquefaction processes are predicated on Linde-Hampson method as a thermodynamic cycle. Liquefaction degree is taken -200 C under maintanence and affective parameters. In nitrogen from air simulation, HYSYS programme used and purity rate of nitrogen found as % as a result of simulation. Rizk et al. (2012) Made simulation of three types of cryogenic process column and calculated exergy losses of different columns. For each column accurate analyses has been defined. Exergy analyses between distillations columns have been compared double diabetic column s exergy efficiency is 23 % more efficient than traditional adyabatic double columns. Van der Ham and Kjelstrup (2010) made two different air separation units exergy analyses one of the studied units is three columns other is two columns. Three columns design has 12 % less exergy losses than two columns design. II. Experimental facility In this study an integred system is examined. Nitrogen liquefaction unit has been integrated to air separation system. Nitrogen cryogenic is used as air enters the separation unit under atmosphere pressure and ambiente temperature. Air liquid passed from air filter, passes through three staged air compressor, then enters air purity units. In this stage, separated from particles and damp air enters to cold box. Cold box consists of main heat exchanger blocks and distillation column is formed two different column and argon column. These two different columns are high and low pressure column. After that dry air enters heat exchanger and leaves it in a temperature close to liquafection degree and that dry air enters high pressure column of to cold box. There are 50 separation trays in high pressure column and 78 separation trays in low pressure column. 14 Fig. 1: Distillation column in cold box Air is separated as oxygen, nitrogen and argon through temperature differences. In distillation column, nitrogen is separated from other components and transferred into liquefaction unit. Nitrogen liquefaction unit consists of nitrogen recycle compresson CP-77, booster compressor CE-77, booster compressor last cooler HE-771, nitrogen chiller R-60 and three exchangers. While entering nitrogen liquefaction unit with a nearly 5 bar pressur, nitrogen has been risen to nearly 32 bar pressure in recycle compresson. Nitrogen leaves booster compressor, turbine with nearly 45 bar pressure. Compressor enables this pressure work from turbine through booster compressor last cooler and enters to first heat exchanger blocks HE-1. It leaves from HE-1 with a temperature of 251 K. Entering nitrogen chiller cooler leaves its heat here. After that nitrogen enters to HE-2 heat exchanger and leaves it with a temperature of 182 K. Nitrogen s 3/4 liquid mass leaving HE-2 heat exchanger is send to booster compressure turbine and enables the necessary work for pressure in compressor. Liquid from turbine combines with average pressure nitrogen and passes though heat exchanger blocks in liquefaction unit them of combines with average nitrogen from main heat exchanger of cold box and finally enters nitrogen recycle compressor ¼ of nitrogen from HE-2 heat exchangerenters to HE-3 heat exchanger and leaves HE-3 with a temperature of 112 K and 45 bar pressure in liquid phases.

29 COP actual = q L,gas w in (6) For COP vaule of liquid per mass in liquefaction unit, equation 2 is calculated as reversible work by Dinçer and Rosen (2007). COP rev = Q L,liquid w rev (7) Reversible work in equation 7 is defined in equation 8 by Dinçer and Rosen (2007). w rev = h 17 h 2 T 0 (s 17 s 2 ) (8) Fig. 2: Air separature unit by Linde Group (2009) III. Exergy analysis For making exergy analysis of nitrogen liquefaction unit, it s necessary to define two cooling effect in liquid and gas phases. The refrigeration effect per unit mass of the liquefied gas is give by Dinçer and Rosen (2007). q L,gas = h 4 h 2 (1) h 4 value defines the enthalpy leaving from compressure, h 2 value defines the enthalpy entering the compressor. From energy balance on the cycle, the refrigerant effect per unit mass of the liquefied gas is given by Dinçer and Rosen (2007). q L,liquid = h 2 h liquid (2) hliquid defines the enthalpy value of liquid nitrogen leaving cycle. If energy balance in compressor is written for gases per mass according to compression by Dinçer and Rosen (2007). w in = RT 0 ln(p 2 P 1 ) (3) Here R is nitrogen gases constant, T0 is ambient temperature, P values are entering and exit pressures. (T0=298.15) For finding exergy efficiency in cycle, COPactual value rate of COPrev values have been calculated. η ex = COP actual COP rev (9) To find exergy efficiency in heat exchanger of nitrogen liquefaction cycle by Thomas et al. (2011) η ex HE = m HP (ex heat out ex heat in ) m LP (ex cold in ex cold out ) (10) Here m is the rate of mass flow through heat exchanger, ex is defined as hot and cold exergy. Tab. 3: Enthalpy and entropy values in nitrogen liquefaction cycle Referance point values Enthalpy (kj/kg) Entropy (kj/kg) Referance point values Enthalpy Entropy (kj/kg) (kj/kg) For finding liquefaction rate of gases in cycle, fraction of equation 2 to equation 1 has been calculated by Dinçer and Rosen (2007). y = q L,gas q L,liquid (4) If actual work in cycle is written fornitrogen per mass by Dinçer and Rosen (2007). w actual = w in y (5) Fig. 3: T-s diagram of nitrogen liquefaction unit If actual coefficient of performance (COP) in cycle is written for gases per mass by Dinçer and Rosen (2007). 15

30 In table 6 ıt is observed analyses decreases if T2 temperature entering the system increases in circumstances of T0 value is 20 C and 25 C. But if ambiatte temperature is 20 C COPactual value is a little much comparing to 25 C however if ambiante temperature is 25 C exergy efficiency inreases comparing to 20 C. Fig.4: P-h diagram of nitrogen liquefaction unit IV. Results and discussions With the values given in table 3, equations solved and values in table 4 obtained. Tab. 4: Calculated result Calculated Result Calculated Result Values Values q L,gas 52.3 kj/kg COP actual q L,liquid kj/kg COP rev y η ex 0.36 w actual,gas kj/kg η ex HE w actual,liquid kj/kg η ex HE w rev kj/kg η ex HE It ambiente temperature increases COPactual and COPrev values decrease in pic 5. Fig. 7: COPactual and exergy efficiency change graphic depending on h4 value In figure 7 ıt s observed that exergy efficiency increases if enthalpy value rises leaving compressor. V. Conclusions In this study a real air separation and nitrogen liquefaction units have been examined. Energy and exergy analyses has been made and exergy efficiency has been calculated as 0.36, COPactual value and COPrev as Moreover exergy efficiencies of HE-1, HE-2 and HE-3 heat exchangers have been calculated as 0.55, 0.81 and In result of this study, compressor efficiency is low but ıf enthalpy value increses as showed in figure 7 exergy efficiency of cycle will increase. Nomenclature Fig. 5: COP values change graphics depending on T0 temperature m ex COP w h s : mass : Exergy : Coefficient of Performance : Work : Enthalpy : Entropy Fig. 6: Exergy effiency change graphics depending on T2 temperature 16 Greek letters η : Effiency Subscripts ηex : Exergy efficiency References Weisend II, J.G., Hanbook of Cryogenic Engineering. 504, Taylor & Francis, USA. Rizk, J., Nemer, M., Clodic, D., A Real Column Design Exergy Optimization of a Cryogenic Air

31 Separation Unit. Energy, 37, Castle, W. F., Air Separation and Liquefaction: Recent Developments and Prospects for the Beginning of the New Millennium. International Journal of Refrigeration, 25, Campestrini M., Thermodynamic study of solid-liquid-vapor equilibrium: application to cryogenizs and air separation unit, doctora thesis, 147. Vinson, D. R., Air separation control technology, Computers and chemical engineering, 30, Thomas, R. J., Ghosh, P., Chowdhury, K., Exergy analysis of helium liquefaction systems based on modifield Claude cycle with two-expanders. Cryogenics, 51, Amin R., Islam A., Islam R., Islam S., Simulation of N2 Gas Separation Process From Air, IOSR Journal of Applied Chemistry, Volume: 6, Issue: 5, van der Ham, L. V., Kjelstrup, S., Exergy Analysis of Two Cryogenic Air Separation Processes, Energy, 35, Dinçer, İ., Rose, M.R., Exergy: Energy, Environment and Sustainable Development, 454. Elsevier, Canada. KLM Technology group, Air Separation Units. The Linde Grup, About Air Separation Units. 17

32 Investigation of Irreversibility with CO2 Emission Measurement in Industrial Enamel Furnace Sedat Vatandas 1*, Atakan Avci 2, M. Ziya Sogut 3 1 Energy Efficiency and management Department, Enervis, Bursa Turkey 2 Mechanical Engineering Department, Engineering Faculty, Uludağ University, Bursa, Turkey 3 Mechanical Engineering Department, Engineering Faculty, Bursa Orhangazi University, Bursa, Turkey. * Abstract The increasing use of energy in Turkey, especially in industry, the need for energy efficiency raises to the forefront every day. External dependence on sources of energy, energy costs and competitive factors, addressed the energy efficiency to take into account as a new energy source. At the same time, increasing competitiveness, reducing energy costs and decreasing environmental impacts can be achieved only through energy efficiency. In this study, primarily energy and exergy analysis of enamel oven is made which has a significant energy consumption in the facility then energy efficiency improvements are evaluated. Due to the analysis, irreversibility found approximately 88.71%. Finally, economic savings provided by emission reductions are evaluated in accordance with improvements.irreversibility due to changes in their CO2 emissions are discussed separately in this study. Exergy and environmental effects of improvements are assessed. Keywords: Enamel ovens, energy analysis, exergy, analysis, CO2 emissions, irreversibility. I. Introduction The threat posed by an increasingly global warming today, many studies are performed for solutions. Considering the results and cost of the solutions, the importance of energy efficiency increases even more. Reducing energy costs and carbon emissions can be achieved only through energy efficiency. As half the energy used in industry, ensuring energy efficiency is extremely important in terms of competitiveness and reduction of environmental impact. It is impossible to continue to compete without reducing specific energy consumption with increasing energy needs and costs. One of the significant energy users in facilities are industrial furnaces. In processes where the industrial furnace is, energy cost is the highest after the raw material cost. Significant amount of energy in high-temperature furnaces cannot be converted to useful energy. Industrial furnaces carry significant potential for energy efficiency targets desired to be achieved in the industry. Minimizing losses and energy recovery will be returned as a profit by ensuring energy efficiency in industrial plants. In this study, in order to determine the current state of modeled enamel furnace, energy and exergy analysis was performed after calculating the energy, transferred into the furnace and turned into useful energy after processing. Afterwards studies were described that was made to upgrading the furnace. And energy and exergy analyses re performed with energy efficiency improvement projects. According to the results of analyses energy recovery potential was evaluated and results are discussed. II. Industrial furnaces Industrial furnaces are used in metallurgical production to smelt at high temperatures, in heat treatment, tempering, in some areas such as the food industry, drying and for fermentation at low temperatures (Hazi et al., 2009). Metal industry which has a significant share in the total industrial energy consumption, examples of furnaces working in the high temperature zones; Enamel cooking 600 C C Heat treatment of metals 1100 C Rolling, extrusion, cooking ceramic materials, such as heat treatment and pressing 1350 C Melting and smelting of metals 1700 C According to the type of heat generation, furnaces can be divided into two main groups. Electrical furnaces and furnaces that use fuel. According to the type of fuel that combustion furnaces use, furnaces classified as solid, liquid and gaseous furnaces (Trink et al., 2004). Electric ovens, works as an arc furnace or an induction furnace. Some advantages of electric ovens are easy to operate and easy to be managed. The absence of any loss due to the flue gas is another advantage of electric ovens. On the other hand main disadvantage is the price of electricity. In addition heating can be provided by plasma arc, laser, radio frequency or a combination of heating by electromagnetic. 18

33 Combustion furnaces can be classified according to the type of heat transfer of heat, operating state and by providing the shape of the heat recovery. Industrial furnaces shall design to provide maximum amount of the product with homogeneously heat diffusion. The specific energy consumption should be kept constantly under control as well as quality. Furnaces should be operated with minimum fuel and minimum maintenance while furnace design should allow maximum heat transfer to maximum material in defined time. In order to ensure these conditions, criteria's should be considered that ordered below; How much heat will be transferred to the material Determination of the heat necessary for heating the mass and losses. Once these issues are identified, the studies such as, how much of the losses will be minimized, what will be used as a refracter material and how will the temperature of the furnace body stabilize, should be made. The model industrial furnace is used for the curing of the applied enamel of the tank and the boiler to be used in the heating sector. After the processes of metal forming, welding, degreasing, cleaning and enameling of the tank and boiler curing process is taken in model furnace. According to the sizes, adequate number of tank and boiler is processed in the furnace at 860 C. The flow line of the furnace is given in Figure 1. III. Theoretical analyses Industrial furnaces are treated as thermal processes and analyzed as a continuous flow system under the first and second laws of thermodynamics. Such systems, in order to be performed energy and exergy analysis firstly temperature, environmental conditions, specific heat capacity and mass flow of the input and output materials must be defined. In the evaluation of industrial furnaces mass balance of input and output material defined as below. (Sogut ve Oktay, 2006); m. i. m o (1) Energy is a protected property under the first law(cheng et al,. In this case, the mass flow rate for the furnace, general energy balance caused by work and heat is;. E. Q i. E o. mi h W i. m o h o (2) (3) (Sogut ve Oktay, 2006; Balkan et al., 2005). In equation (2) E. i refers input total energy amount, Ė o refers output total energy amount. In equation (3) Q is ( Q Q and W refers (. W.. W net net. i. Q Q ) total heat. W o o. W i ) amount of work. h is the enthalpy value of input and output materials. The first law of thermodynamic describes a quantitative measure of energy as independent for the direction of energy. Practically systems lose energy, with the irreversibility and changings in environmental conditions. As this situation is evaluated as production of entropy, in the second law of thermodynamic it is expressed in the concept of exergy. The exergy of a system is the maximum useful work possible according to the environmental conditions of system. The exergy balance of the prose's expressed (Szargut vd., 1998 and Utlu e al.2011); E x = Ex kin + Ex pot + Ex phy+ Ex che (4) Fig. 1: The flow line of the furnace In the equation Ekin is kinetic exergy, Epot is potential exergy, Ephy is physical exergy, and Eche is chemical exergy. When the furnaces are considered as continuous flow system, exergy balance is;. Ex i. Ex o. l (5) 19

34 (Sogut, 2010).. Ex o Ex. i refers to output exergy. refers to input exergy and,. l defines the exergetic destruction. In these systems, the potential, kinetic and chemical exergy can be neglected. Hence general exergy balance can be written as below; T..... ( 1 0 ) Qk W mi i mo o l (6) T In equation (6) 1 Q. k, Ṫ k refers heat transfer rate, W. is work amount, is flow exergy, s entropy and 0 index represents (P 0 and T 0) in terms of dead state of environment. Hence mass and material flow exergy is; ( h h0 ) T0 ( s s0) (7) Performance evaluation in the system is defined by efficiency. Efficiency is defined as the ratio of input to output of the system. Depending on the analysis conducted under the first law energy efficiency of process is found by the equation below; Here process and (8) Ė refers to the total output energy of o Ė i refers to the total input energy of process. Exergy efficiency of process is calculated with the equation of (9) (Cornelissen 1997).. Ex o ıı (9). Ex i CO2 emission in thermal systems, caused by energy losses, is dependent on the type of fuel used and the waste energy potential. It can be expressed as CO 2 where CO Q 2 W CO 2 I (10) is the unit energy CO2 emission Q coefficient and W is the amount of total waste energy. In works where such waste energy is exploited, depending on the recycled energy potential, total CO2 should be calculated taking CO2 emissions caused by furnaces. Hence, total CO2 emission (CO ) may be expressed as follows: 2 CO. E. o ı E i Q Q (1 ) (11) CO CO 2 2 BW i Wi j j 2 CO2i CO2 j Ri I i I j Q where W is the total energy of furnace, is rational exergy efficiency. is the function among environment temperatures, waste heat, and inlet and outlet temperatures of furnace(kılkış, 2004). Van Gool (1997) has noted that maximum improvement in exergy efficiency for a process or system is obviously achieved when the exergy loss or irreversibility E x E x ) is minimized. ( i o Consequently, he suggested that it is useful to employ the concept of an exergetic improvement potential when analyzing different processes or sectors of the economy. Hammond and Stapleton (2001) give this improvement potential in a rate form, denoted as IP below. IP 1 )( ) (12) ( g IV. Energy and exergy analyses Providing mass balance of enamel furnace which is taken as a model, energy and exergy analyses are conducted according to the reference environmental conditions. The energy analyses of the furnace is made and results is given in Table 1. Material Tab. 1: Energy analyses of the furnace Through energy analyses, energy efficiency of furnace was found %17.3 according to the equation (8). mo. ho ,1 1 = 0,173 mi. hi In real furnace processes environmental conditions directly affect the efficiency. In this aspect the importance of exergy analyses become important which is connected to the second law of thermodynamics. However, firstly environmental ç Input Materials M T1 Cp Δh kg/h K kj/kgk kj/h Boiler 279, , ,31 N.gas 40, , ,20 Air 35, , ,17 Enamel 2, ,35 212,21 Electricity 4,30 Total 357, ,2 Material Output Materials M T1 Cp Δh kg/h C kj/kgk kj/h Boiler 279, , ,21 Flue gas 77, , ,87 Total 356, ,1 20

35 parameters need to be defined for exergy analyses. The environmental reference pressure was taken (P0) 1 atm and temperature 25 C for conditions where the furnace is located. For exergy analyses some assumptions are made. According to this; effect of pressure on the enthalpy and entropy characteristics of input and output materials were neglected. Gases are considered as an ideal gas mixture for input and output material flow. As the furnace is continuous flow, kinetic, potential and physical exergy value of input and output materials are neglected. According to the assumptions exergy analyses were conducted and represented in Table 2. Material Tab. 2:Exergy analyses of the furnace Furnace exergy efficiency is defined as the ratio of the total exergy output to total exergy input. Accordingly exergy efficiency of the furnace is;. Ex o ıı =. 0, 1128 Ex M T1 Cp Δh Δs ψi kg/h C kj/kgk kj/h kj/k kj/h Boiler 279, , ,31 8, ,8771 N.gas 40, , ,2 5, ,72 Air 35, , ,17 2, ,75538 Enamel 2, ,35 212,21 0, , Electricity 4,3 0 4,3 Total 356, ,2 16, ,89 Material i , ,89 Input Materials Output Materials M T1 Cp Δh Δs ψi i kg/h C kj/kgk kj/h kj/k Boiler 279, , ,21 172, ,5888 Flue gas 77, , ,87 146, ,2978 Total 356, ,08 319, ,89 According to the results of analyses, efficiency of the furnace was determined too low and causes were investigated. Causes were found as heat losses in the furnace surface (Figure 2 and 3), inefficient combustion system and optimum design is not provided. Also waste heat potential was determined according to the results of flue gas analyses shown in Figure 4. Fig. 3: Heat losses in the furnace surface (side) Fig. 4: Flue Gas Analyses Based on this information by changing the furnace combustion systems instead of conventional eight burners, six recuperative burners are used. Furnace panel is renewed in accordance with the modified combustion system in order to provide control. Furnace refreacter material has been renewed in order to avoid the leakage losses. Although waste heat was used by recuperative burner (Figure 5), still waste heat potential is determined for the bath process of the boilers. In bath process hot water is needed for the degreasing process. Economizer is used for heating water of bath process(figure 6). Nearly kcal was obtained with the economizer. Fig. 5: Recuperative burner section Fig. 2: Heat losses in the furnace surface (Front) 21

36 V. Measurement of CO 2 emissions As a result of studies exergy efficiencies were respectively 11.28% and 22.73%. With the equation of 12 improvement potential (IP) of furnace was found After improvement projects IP was calculated Table 5 is prepared according to the equation 10. Tab. 5 CO2 Emissions saving after improvement CO2 CO2 Exergy Improvement CO2 Emission Efficiency Potential (kg CO2/h) Emission Factor Before Improvement 11,28% , After Improvement 22,73% , Fig. 6: Recuperative burner in furnace Exergy analyses and energy analyses have been renovated and after efficiency improvement project and the results of the energy analysis are given Table 3. Tab. 3: The balance of energy analyses Energy efficiency of the furnace according to the equation 8 was found %35. Material Material M T1 Cp Δh kg/h C kj/kgk kj/h Boiler 449, , ,89 N.gas 26, , ,80 Air 58, , ,37 Enamel 3, ,35 318,31 Electricity 4,30 Total 538, Material Input Materials Output Materials M T1 Cp Δh kg/h C kj/kgk kj/h Boiler 449, , ,79 Flue gas 88, , ,90 Total 538, Tab. 4. The balance of exergy analysis M T1 Cp Δh Δs ψi kg/h C kj/kgk kj/h kj/k kj/h Boiler 449, , ,888 13, ,80196 N.gas 26, , ,8 Air 58, , ,369 4, ,4766 Enamel 3, ,35 318, , , Electricity 4,3 0 4,3 Total 538, ,7 18, ,72 Material Input Materials Output Materials M T1 Cp Δh Δs ψi i kg/h C kj/kgk kj/h kj/k Boiler 449, , ,79 277, ,01 Flue gas 88, , ,90 166, ,58 Total 538, ,69 444, ,59 Exergy efficiency of the furnace according to the equation 9 was found % 22, With the equation (CO2b / CO2a)/ CO2b, %74 CO2 saving potential has been identified with the projects that are explained in part IV. VI. Conclusions Studies show that energy efficiency projects pay their selves in a short time when compared to energy generation projects. Therefore, facilities should think constantly energy efficiency and perform systematically energy saving projects. In order to maximize the energy efficiency, facilities should focus on large energy consumer as furnaces. In this study, enamel furnace was taken as a model which is defined as significant energy user in energy management system. Firstly looses and leakages defined and fixed, afterwards efficiency was increased in the view of technological developments. To provide the maximum amount of production per unit time, furnace height was increased. And 3 pieces of boiler has started processed which was 2 before improvement. In addition, due to technological advances, burner systems were upgraded. After improvements, energy and exergy efficiency of the enamel furnace was doubled. Likewise enamel furnace improvement potential is reduced by half. CO2 emission depending on exergy efficiency was decreased four times. Also with the design changing's in furnace, not only more production was provided per unit time but also CO2 emission per unit of production was reduced. This situation also shows that the energy efficient design is very important in the case of energy efficiency. References Balkan, F., Colak N., Hepbasli, A. (2005) Performance evaluation of a triple-effect evaporator with forward feed using exergy analysis, Int. J. Energy Res

37 Chen X., Zhang Y., Zhang S., Chen Y., Liu S.(2007) Exergy analysis of iron and steel eco-industrial systems The third international Exergy, energy and Envıronment symposıum, 1 5 July, Evora, Portugal Cornelissen R.L.(1997) Thermodynamics and sustainable development: The use of exergy analysis and the reduction of irreversibility, Ph.D thesis, University of Twente, The Netherlands. Hammond G.P., Stapleton A.J. (2001) Exergy analysis of the United Kingdom energy system, Proceedings of the Institute of Mechanical Engineers, 215 (2) Hazi A., Badea A., Hazi Gh., Necula H., Grigore R. (2009) Exergy Evaluation of Renewable Use in the Pulp and Paper Industry, IEEE Bucharest Power Tech Conference, June 28th-July 2nd, Bucharest,Romania Kılkıs Bir. (2004) An Exergy a ware optimization and control algoritma for sustainable buildings, İnternational Journey of Green Exergy, 01/2004;No 1:65-77 Sogut Z., Oktay Z.,(2006) Energy And Exergy Analyses In Thermal Process Of Production Line Of Cement Factory And Application, Igec-2 International Green Energy Conference, Ontario Institute of Technology (UOIT), Canada, June 2006 Sogut Z., Oktay Z. Karakoç H.(2010) Mathematical modeling of heat recovery from a rotary kiln Applied Thermal Engineering 30 (2010) Szargut, J., Morris, D.R., Steward, F.R.(1988) Exergy Analysis of Thermal and Metallurgical Processes, Hemisphere Publishing Corporation. Utlu Z., Hepbasli A., Turan M.(2011) Thermodynamic analyses of a Industrial dryer mill, X National Sanitary Engineering congress, 13/16 April 2011, İzmir/Turkey Trinks W., Mawhinne M. H.,(2004) Shannon R. A., Reed R. J., Garvey J. R. Industrial Furnaces Sixth Edition, ISBN: , Copyright 2004 John Wiley & Sons, Inc, New Jersey, USA W. Van Gool, (1997) Energy policy: fairy tales and factualities, in: O.D.D. Soares, A. Martins da Cruz, G. Costa Pereira, I.M.R.T. Soares, A.J.P.S. Reis (Eds.), Innovation and Technology Strategies and Policies, Kluwer, Dordrecht, pp

38 Advanced Exergy Analysis of an Application of Waste Heat Powered Ejector Refrigeration System to Rotary Kiln Abid Ustaoglu 1*, Mustafa Alptekin 2, Mehmet Emin Akay 3, Resat Selbas 2 1 Bartin University, Faculty of Engineering, Department of Mechanical Engineering, Bartin, 74100, Turkey 2 Suleyman Demirel University, Faculty of Engineering, Department of Energy Systems Engineering, Isparta, 32260, Turkey 3 Karabuk University, Mechanical Engineering Department, Karabuk, 78050, Turkey. * ; Abstract The rotary kiln consumes the biggest share of energy in a cement factory and has great heat loss which causes significant reduction in efficiency. This reduction becomes excessive for a cement factory using the wet method. In the progress of clinker production, about 33% of total energy is exhausted from the chimney of rotary kiln. In this study, conventional and advanced exergy analyses of a waste heat powered ejector refrigeration system were performed by using the exhausted heat from the chimney of the rotary kiln. Conventional and advanced exegy analyses are carried out to the system. The exergy destructions, thermal efficiency and exergy efficiency were calculated. By means of the advanced exergy analysis, avoidable and unavoidable exergy destruction rates were found in order to determine the improvement potentials of both the components and the overall system. Moreover, the effects of condenser temperature, generator temperature and evaporator temperature on system performance were also investigated. The largest exergy destruction occurs in generator, accounting 48.9% of total exergy destruction and followed by ejector with ratio of 39.13%. The largest share of total avoidable exergy destruction occurs in ejector and that is about half of total avoidable part of overall system (82.94%). Therefore, it is important to concentrate on this component to improve overall system performance. About 31 % of the total destruction is falling into the part of unavoidable exergy destruction. Namely, the system has potential to improve by reducing the avoidable part of 40% Keywords: Advanced exergy analysis, waste heat, rotary kiln, ejector refrigeration I. Introduction Renewable energy is one of the most important solutions for a clean energy future. Another alternative to solve these problems is utilization of the waste heat. Ejector refrigeration systems are promising technologies since they can utilize renewable energy and harvest low-grade waste heat from industrial processes for cooling demand so that the problems related with greenhouse gas emission and the energy cost can be reduced. Cement industry is one of the most energy consuming industries in the world. The energy consumption rate reaches 12-15% of total energy consumption in industry. The rotary kiln has biggest share in terms of energy consumption in a cement factory and has great heat loss which causes significant reduction in efficiency. This reduction becomes excessive for a cement factory using the wet method. In the progress of clinker production, about 33% of total energy is exhausted from the chimney of rotary kiln apart from the heat loss through the wall of rotary kiln Ustaoglu et al. (2016). Therefore, reutilization of the exhausted gas becomes substantial due to the great waste heat. In order to utilize this heat, one of the preferable options is to use ejector refrigeration cycle. In the ejector refrigeration cycle, ejector is an important component. An ejector can increase the pressure without using mechanical energy directly. Therefore, it is better to use an ejector than applying mechanical devices to increase the pressure such as compressor, pump, etc. since it may be safer and simpler technology. Apart from that, the system combined with ejectors and other components are also simple. Therefore, researchers have been made endeavor to have more knowledge about the ejector behavior and improve the system performance with various methods. The ejector theory was proposed by Keenan et al (1950) and that underlies many other ejector model and designs. Since proposition of the ejector theory, many studies have been carried out to improve the performance of the refrigeration cycles by improving design or combining with other cycle, and to evaluate the first and second low efficiency of thermodynamic for more realistic and detail approach. Huang et al. (1999) predicted the ejector performance for critical conditions and validated the results with experimental data. Zhu et al. (2007) developed a two-dimensional ejector model by considering of a shock circle at the entrance of the constant area section. The second law efficiency evaluation is important tool to determine the location, magnitude and source of the exergy destruction. Pridasawas and Lundqvist (2004) carried out exergy analysis for a solar-driven 24

39 ejector refrigeration system using butane as a working fluid and obtained that the most substantial exergy destruction in the ejector refrigeration cycle occurred in the ejector. Dahmani et al. (2011) said that more than half of the total exergy destruction in the ejector refrigeration cycle by using working fluid, R134a was due to the ejector. The exergy analysis for ejector enhanced refrigeration systems has also been largely carried out to evaluate the improvement of the system performance by adapting the ejector. Yan et al. (2015) proposed a new ejector enhanced auto-cascade refrigeration cycle using R134a/R23 and compared it with a conventional auto-cascade refrigeration cycle. Their cycle achieved % better efficiency than a basic cycle at the same operation conditions as the ejector achieved pressure ratio lifts of They found that the highest exergy destruction occurs in the compressor, the condenser, cascade condenser, expansion valve, ejector and evaporator, respectively. Yang et al. (2016) proposed a novel combined power and ejector-refrigeration cycle using zeotropic mixture, and evaluated the cycle performance with different fluid composition and compared the novel cycle with conventional combined cycle. The results showed that the cycle has better performance in lower condenser temperature. Although the refrigeration cycle achieves lower evaporating temperature in higher generating temperature, the power output decreased. Zhao et al. (2015) conducted a numerical study to analyze the performance of the ejectorexpansion refrigeration cycle (EERC) using zeotropic mixtures (R134a/R143a). The results showed that the compressor and ejector have the most exergy destruction, and the cycle exergy efficiency achieves a maximum value with the mass fraction of 0.7 for R134a. Apart from the conventional analysis method, the exergy method is an important tool that can show the useful work that can be generated through the process. However, it cannot explain the interaction among the component or estimate the actual improvement potential. Conventional approach for optimizing of the system may be wrong without taking into account of the interaction between the components particularly for complicated systems in which many components having interaction with each other. In the case of the ejector refrigeration cycles, the operation parameters of the ejector depend on itself and other components Chen et al. (2014). A recent developed technique, the advanced exergy analysis by splitting the exergy destruction into avoidable-unavoidable and endogenous-exogenous part can enable us to evaluate the system in more detail and investigate the capacity of the improvement Tsatsaronis (1999). The advanced exergy analysis of different refrigeration systems, including vapor compression refrigeration systems Morosuk and Tsatsaronis (2009), Morosuk et al. 25 (2012), absorption refrigeration systems Gong and Goni (2014), Morosuk T, Tsatsaronis (2014) ejector refrigeration system Chen et al. (2015) and heat pump Erbay and Hepbasli (2014) have been carried out. In a previous study, energy and exergy analysis of a wet type rotary kiln were carried and recovery capacity of waste heat was evaluated by using an organic Rankine cycle Ustaoglu et al. (2016). The results showed that a great amount of heat energy of 30.5 MW is exhausted from the chimney of rotary kiln. In this study, the waste heat from the rotary kiln was evaluated for an ejector refrigeration system and advanced exergy analysis was carried out to determine avoidable and unavoidable exergy destruction rates in order to determine the improvement potentials of both the components and the overall system. II. Material and Method II.1. System Description Figure 1 shows a waste heat powered ejector refrigeration system.. Each state of the working fluids is represented in the point as seen in figure. The working principle of this system can be expressed as follows: the working fluid in saturated liquid phases leaving from the condenser is separated as two parts; one goes to pump where that fluid is compressed to generator pressure, and the other one goes to the expansion valve. The properties of the working fluid in points 1 and 4 are same as point 8. The compressed fluid is pumped to generator to be vaporized by using the exhausted gas from the rotary kiln chimney and leaves the generator as saturated vapor. The other fluid s pressure and temperature decreases to the evaporator level in expansion valve. The fluid is vaporized in the evaporator by using the heat of the cooling ambient. Either saturated vapors from the upper and below cycle enter to ejector and mix. When the condenser pressure below the critical value, ejector can entrain same amount of secondary fluid Huang (1985). Thus, the cooling capacity and COP are kept constant. The mixture leaves from the ejector to be superheated vapor at condenser pressure. The forking fluid entering water-cooled condenser release its heat to the water and leaves the condenser from point 8 to be saturated fluid, and then again separates. Thus, the cycle is completed. R142b is used as working fluid. II.2. Thermodynamic Evaluation The ejector refrigeration system is considered as steady flow open-system and modeled based on the first and second laws of thermodynamics, and these laws are applied to each component in the system. In the steady flow open systems, the mass and energy of control volume are stationary. General equations of mass, energy and exergy balances by ignoring the kinetic and potential energy variations can be expressed.

40 E F,tot = E F,GE + W PU (6) E P,tot = E P,EV (7) E D,tot = E D,k (8) The coefficient of the performance (COP) can be described as a ratio of the refrigerating capacity generated by evaporator to the heat supplied to generator and electricity consumption of the pump COP = Q EV (Q GE + W PU ) = m EV(h 6 h 5 ) m GE(h 3 h 1 ) = μ(h 6 h 5 )/(h 3 h 1 ) (9) where the ratio of the mass flow rate of secondary fluid coming from evaporator to the primary flow in the reversible ejector coming from the generator indicates ideal entrainment ratio McGovern et al. (2012). Table 2 shows the equations of the entropy calculation for each component in the ejector refrigeration system. Fig. 1: Waste heat powered ejector refrigeration system m in = m out (1) Q + W = m out h out m outh out (2) E heat + W + E D = E out E in (3) S F S P + S GEN = 0 (4) E D = T 0 S GEN (5) where ṁ is mass flow rate, h is enthalpy, Q is heat, and Ẇ is net work. Ėheat, ĖD, Ėin and Ėout are the exergy input through heat, exergy destruction, exergy input and output, respectively. ṠF, ṠP and ṠGEN are the entropy of fuel, product and generation, respectively. Basic thermodynamic evaluation and conventional exergy analysis can be carried out by using these equations. Table 1 shows the equations of the exergy calculation for each component in the ejector refrigeration system. By using these equations, fuel exergy, product exergy and exergy destruction for overall system can be determined by the following equations, respectively Tab. 1: Evaluation of the exergy for each component Component Exergy of fuel (Ė F,XX) Exergy of product (Ė P, XX) Exergy destruction Generator (GE) ṁ 9(e 9- e 10) ṁ 1(e 2- e 3) Ė F,GE- Ė P,GE Condenser (CO) ṁ 7(e 7- e 8) ṁ 11(e 12- e 11) Ė F,CO- Ė P,CO Evaporator (EV) ṁ 4(e 6- e 5) ṁ 13(e 13- e 14) Ė F,EV- Ė P,EV Ejector (EJ) ṁ 3(e 3- e 7) ṁ 6(e 7- e 6) Ė F,EJ- Ė P,EJ Pump (PU) W PU ṁ 1(e 2- e 1) Ė F,PU- Ė P,PU Expansion valve (EXV) ṁ 4e 4 ṁ 4 e 5 Ė F, e 4, EXV - Ė P, e 4, EXV 26 Tab. 2: Evaluation of the entropy for each component Component Entropy of fuel (Ṡ F,XX) Entropy of product (Ṡ P,XX) Exergy destruction Generator (GE) ṁ 3s 3+ṁ 10s 10 ṁ 2s 2+ṁ 9s 9 (Ṡ P,GE - Ṡ F,GE)T 0 Condenser (CO) ṁ 8s 8+ṁ 12s 12 ṁ 7s 7+ṁ 11s 11 (Ṡ P,CO - Ṡ F,CO)T 0 Evaporator (EV) ṁ 6s 6+ṁ 14s 14 ṁ 5s 5+ṁ 13s 13 (Ṡ P,EV - Ṡ F,EV)T 0 Ejector (EJ) ṁ 7s 7 ṁ 3s 3+ṁ 6s 6 (Ṡ P,EJ - Ṡ F,EJ)T 0 Pump (PU) ṁ 2s 2 ṁ 1s 1 (Ṡ P,PU - Ṡ F,PU)T 0 Expansion valve (EXV) II.3. Ejector Model ṁ 5s 5 ṁ 4s 4 (Ṡ P,EXV - Ṡ F,EXV)T 0 An ejector can increase the pressure without using mechanical energy directly. Therefore, it is better to use an ejector than applying mechanical devices to increase the pressure such as compressor, pump, etc. since it may be safer and simpler technology. An ejector system is composed of four section including nozzle, mixing chamber, throat and supersonic diffuser. The high pressure steam known to be primary fluid coming from the generator expands and accelerates along with primary nozzle. The steam blows out with supersonic speed to achieve very low pressure at the end of the nozzle and in mixing chamber. Due to the pressure different, a lower-pressure vapor, secondary fluid, can be entrained within the mixing chamber.they are mixed in this stage. In the diffuser stage, the pressure is recovered to the condenser pressure. In order to facilitate evaluating of ejector model, the following assumption were made; 1. The ejector flow was assumed to be one dimensional and steady state. 2. The velocity variation of the fluid at the ejector inlet and outlet was neglected. 3. The pressure drop of the working fluid and the heat loss were neglected. 4. The pressure in mixing process in the ejector was considered to be constant with mass,

41 momentum and energy conservation. 5. The losses of the flow inside the ejector are all considered using isentropic efficiencies in the motive nozzle (n), in the mixing chamber (m) and in the diffuser (d). The isentropic efficiency of motive nozzle can be determined by following equation n = (h P,3 h P,3a ) (h P,3 h P,3s ) (10) where hp,3, hp,3a and hp,3s are the enthalpy of inlet flow, enthalpy of the exit flow, and enthalpy of the exit flow with isentropic expansion process. For a given nozzle efficiency, hp,3a can be determined. By applying the energy equations to the motive nozzle, the speed of nozzle at the exit can be calculated by using following equations 2 u p,3a 2 = h P,3 h P,3a (11) When Eqs 9 and 10 are solved, the following equation can be obtained u p,3a = 2 n (h P,3 h P,3s ) (12) In a similar way, the velocity of the exit flow can be determined for the secondary fluid u p,3a = 2 n (h P,6 h P,3s ) (13) However, this value can be neglected since it is very small compare to the primary flow velocity. After mixing the fluid, the mass conservation can be defined by m p + m s = m t (14) where mp, ms and mt represents the mass flow rates of primary fluid, secondary fluid and total mass flow rate of primary and secondary mixture. Although the actual results deviate slightly in ejector component, in order to facilitate the evaluation, the mass flow rate of the primary fluid is assumed to be 0.15 kg/s as a first approximation and that is very close the value of the primary fluids in literature. However, it should be defined by calculating of entrainment ratio. According to the momentum conservation, the following equation can be obtained m p u p,3a + m s u s,3a = m t u t,3m (15) The efficiency of the fluid mixture can be calculated by 2 m = u t,3m 2 u t,3s (16) when the energy conservation equation is applied to the mixture of the fluid. 2 m p (h P,3a + u p,3a 2 2) + m s (h s,3a + u s,3a 2) 2 = m t (h t,3m + u t,3m 2) (17) 27 Thus, the value of the enthalpy ht,3m and velocity ut,3m of the mixture can be determined. The first term can be neglected due to its comparatively low value. After the throat section, the mixture comes to supersonic diffuser section. The isentropic efficiency of the diffuser section can be expressed by d = (h 7s h t,3m ) (h 7 h t,3m ) (18) where h7 and h7s are the enthalpy of the exit flow and the enthalpy of the exit flow for isentropic compression process. II.4. Advanced Exergy Analysis All actual operations are irreversible due chemical reactions, heat transfer at finite temperature difference, mixture of matters, infinite expansion and frictions Petrakopoulou (2011). Conventional exergy analysis can describe components having high exergy destruction and its reasons, which component has irreversibility and its magnitude. However, it cannot explain the interaction among the component or estimate the actual improvement potential. Conventional approach to optimize the system may be wrong without taking into account of the interaction between the components particularly for complicated systems in which many components having interaction with each other. Therefore, advanced exergy analysis has been carried out. A detail exergy analysis, where the exergy destruction is split into several parts, is called to be advance exergy analysis. These parts are avoidable-unavoidable and endogenous-exogenous exergy destructions Morosuk and Tsatsaronis (2009). The value of the total exergy destruction can be calculated through an exergy balance for this component as follows E D,k = E F,k E P,k = T 0 S P,k = T 0 m ks P,k (19) where T 0, E F,k and E P,k shows reference temperature, exergetic fuel and exergetic product, respectively. Thus, the exergy efficiency can be found by following equations for kth component ε k = E P,k E F,k = 1 E D,k E F,k (20) where the exergy destruction E D,k can be split into AV UN avoidable E D,k and unavoidable E D,k parts, Tsatsaronis and Park (2002), Cziesla et al. (2006), Tsatsaronis and Morosuk (2008a, 2008b); Petrakopoulou et al. (2012). E D,k = E D,k UN AV + E (21) D,k This can provide more realistic approach to measure the potential improvement of the thermodynamic efficiency of the components. The unavoidable part of the exergy destruction cannot be shrunk due to the

42 technological limitations including material and manufacturing cost and availability. The other part of indicates avoidable destruction. EN The other approach is split into endogenous E D,k EX and exogenous E D,k parts as follows, Tsatsaronis (1999), Morosuk and Tsatsaronis (2006) E D,k = E D,k EN EX + E (22) D,k The endogenous and exogenous part of the exergy destruction provides the reason of the exergy destruction in the component caused by the EN component itself or by the other components. E D,k part can be obtained when the all irreversibilities occurs in the kth component while the other component is assumed to be ideal and has no irreversibility with having its current efficiency. On the EX other hand, E D,k part occurs within the kth component due to the irreversibilities in the other components. For each case, the power output of the overall system is kept constant and equal to actual case. Thus, the exogenous destruction is the remaining part of the total exergy destruction in kth component and can be calculated by subtracting the endogenous exergy destruction from the total exergy destruction Petrakopoulou (2011). In order to calculate the unavoidable exergy destruction, each component should be considered isolated and separated from the system. The exergy UN destruction rate per unit product exergy (E D E P ) k can be calculated by assuming the system operating with high efficiency and low losses. Thus, the unavoidable exergy destruction for kth compoentn by using the real case product exergy rate can be expressed by Petrakopoulou (2011) UN = E E D,k P,k real (E D,k E P,k ) UN (23) when the unavoidable exergy destruction is known, the avoidable part can be calculated by Eq 20. The unavoidable endogenous and exogenous, and avoidable endogenous and exogenous part of the exergy destructions are expressed respectively, by Tsatsaronis and Morosuk (2008a, 2008b) UN,EN = E E D,k E D,k P,k EN (E D,k E P,k UN,EX = E UN E E D,k D,k AV,EN = E EN E E D,k D,k AV,EX = E AV E D,k ) UN (24) UN,EN D,k UN,EN D,k AV,EN D,k (25) (26) (27) In order to evaluate the advanced exergy analysis, modified exergy efficiency can be described by Tsatsaronis and Morosuk (2008a, 2008b) In this study, the exergy destruction rate split into avoidable-unavoidable parts were evaluated. The endogenous- exogenous exergy destructions and combined two splitting approaches were not considered. In a further study, these parts will be evaluated. In order to facilitate the analyses some assumptions are made; 1. Each process in the system is assumed to be steady state. 2. Potential and kinetic energy variations are neglected. 3. The heat transfer to/from ambient and pressure drops in the pipes are neglected. 4. The working fluid at the inlet of pump is assumed as saturated liquid. 5. The dead state pressure P0 and temperature T0 are considered to be kpa and 20 C, respectively. Tab. 3: Input values to the system Parameters Values Pump isentropic efficiency 85% Evaporator temperature 4 C Condenser temperature 35 C Generator temperature 95 C Cooling capacity 20 kw Mass flow rate of exhausted gas 44.5 kg/s Mass flow rate of refrigerant at the inlet of 0.15 kg/s pump Inlet temperature to generator of exhausted 277 C gas Inlet temperature to condenser of cooling 27 C water Outlet temperature from condenser of 32 C cooling water Inlet temperature to evaporator of water 15 C Outlet temperature from evaporator of water 10 C Ambient pressure kpa Ambient temperature 25 C III. Results and Discussions The calculation program was written in Engineering Equation Solver (EES). Table 4 expresses the parameters of the components in the ejector system for the real, ideal and unavoidable conditions. The calculated thermodynamic data of the ejector refrigeration cycle at real operation case, at ideal operation case and at the unavoidable operation conditions for the working fluid R142b are shown in Tables 5-7, respectively. Tab. 4: The parameters used for real, ideal cycles, and the cycle for the unavoidable exergy destruction Component Parameter Real Ideal Unavoidable Pump (PU) ɳ P Generator (GE) ΔT GE 7.5 C C Ejector (EJ) ɳ n Ejector (EJ) ɳ m Ejector (EJ) ɳ d Condenser (CO) ΔT CO 3 C C Expansion valve (EXV) - Isenthalpic Isentropic Isenthalpic Evaporator (EV) ΔT EV 2 C C ε modified = E P,k (E F,k E D,k UN E AV,EX ) (28) D,k 28

43 Tab. 5: Data of the ejector refrigeration cycle at real operation case for the working fluid R142b m P T h s e Loc. Subs. (kg/s) (kpa) ( C) (kj/kg) (kj/kg.k) (kj/kg) 1 R142b R142b R142b R142b R142b R142b R142b R142b R142b Air Air Air Water Water Water Water Water Tab. 6: Data of the ejector refrigeration cycle at ideal operation case for the working fluid R142b m P T h s e Loc. (kg/s) (kpa) ( C) (kj/kg) (kj/kg.k) (kj/kg) Tab. 7: Data of the ejector refrigeration at the unavoidable operation conditionsfor the working fluid R142b Loc. Loc. m (kg/s) P (kpa) T ( C) h (kj/kg) s (kj/kg.k) e (kj/kg) 1UN UN UN UN UN UN UN UN III.1. Evaluation of Convensional Exergy Analysis In recognition of the fuel exergy of the overall system, it was provided by generator and pump. The fuel exergy of generator is provided by the exhausted gas from the rotary kiln and that of pump is provided from the electrical work. 29 Table 8 indicates the results of the conventional exergy analysis of ejector refrigeration system in terms of entropy balance and destruction. The major part of the exergy destruction is arisen from the generator system (48.9%) and followed by the ejector system (39.1%). The pump, expansion valve and evaporator systems have the lowest exergy destruction. The total exergy destruction rate of these components generates about 4.73% of the overall system exergy destruction. In terms of the exergetic efficiency, the expansion valve and pump show the best performance to be about 91 and 90.4 %, respectively. On the other hand, the lowest efficiency was obtained in the generator (34.2%) and followed by the condenser with an exergy efficiency of 38.3%. The remaining components show relatively better efficiency. Although the exergy destruction rate of the ejector forms the major part, its exergetic efficiency is quite preferable. It is observed that the main attention to improve the overall system performance should be paid to the generator system since the exergy destruction is mostly resulted from this system. The overall system shows a low exergetic efficiency of 5.58%. This is arisen from that most of the fuel exergy is destroyed due to limited use of the ejector refrigeration system. Tab. 8: The conventional exergy analysis results S F Component (kw) (kw) (kw) ε (%) Pump (PU) Generator (GE) Ejector (EJ) Condenser (CO) Expansion valve (EXV) Evaporator (EVA) Overall System III.2. Evaluation of Advanced Exergy Analysis Table 8 shows the advanced exergy analysis results where the exegy destruction rates are split into unavoidable-avoidable parts for each component. The second column indicates the exergy destruction for each component. The third and fourth columns show the unavoidable and avoidable parts which can provide the potential improvement of the system components. Tab. 9: The advanced exergy analysis results Splitting the exergy destruction Component Ė D (kw) Ė UN (kw) Ė AV (kw) Pump (PU) Generator (GE) Ejector (EJ) Condenser (CO) Expansion valve (EXV) Evaporator (EVA) Overall System Figure 2 shows these exergy destructions as percentages in the total destruction rate for each S P Ė D

44 component. Although the generator has the largest value of the exergy destruction; most of the destruction is composed of unavoidable part. Namely, the improvement potential for itself and for overall system is very weak. The expansion valve and evaporator show very similar characteristic and have unavoidable exergy destruction rates of about 70%. Moreover, they have very low exergy destruction. On the other hand, the ejector is the most promising component to improve the overall system performance since it has the largest avoidable exergy destruction. Moreover, about 65% of the exergy destruction can be avoided. About 31% of the total exergy destruction is in the part of avoidable exergy destruction for overall system and the ejector has the largest share of overall system. Although the conventional exergy analysis results indicate that the largest exergy destruction is in the generator and may be considered for the system improvement, the advanced exergy analysis results show that the main attention to improve the overall system performance should be paid to the ejector system since the avoidable part of the exergy destruction is mostly resulted from this system. Fig. 2: Exergy destruction rate for avoidable and unavoidable parts IV. Conclusions Conventional and advanced exergy analyses are carried out for ejector refrigeration cycles. Moreover, the exergy destruction was split into avoidable-unavoidable parts in order to determine the improvement potential of each component in the system. The following part summarizes the study. 1. The largest exergy destruction occurs in generator, accounting 48.9% of total exergy destruction and followed by ejector with ratio of 39.13%. 2. The exergetic efficiency of the overall system is about 5.58%. 3. The largest share of total avoidable exergy destruction occurs in ejector and that is about half of total avoidable part of overall system (82.94%). Therefore, it is important to concentrate on this component to improve overall system performance. 4. About 69 % of the total destruction is falling into the part of unavoidable exergy destruction. Namely, the system has potential to improve by reducing the avoidable part of 31%. Nomenclature COP : The coefficient of the performance (-) 30 E : Exergy (kw) e : Specific exergy (kj.kg-1) h : Specific enthalpy (kj.kg-1) ṁ : Mass flow rate (kg.s-1) P : Pressure (kpa) Q : Heat load (kw) s : Specific entropy (kj.kg-1 K-1) T : Temperature (C) W : Work (kw) Greek letters ε : Exergetic efficiency Superscripts AV : Avoidable EN : Endogenous EX : Exogenous UN : Unavoidable Subscripts a : Actual CO : Condenser D : Destruction EJ : Ejector EXV : Expansion valve EV : Evaporator F : Fuel GE : Generator In : Inlet k : The k-th component n : nozzle out : Outlet P : Product PU : Pump s : Isentropic tot : Total XX : Component representative References Cai D.H., He G.G., Tian Q.Q., Tang W.E., Thermodynamic analysis of a novel aircooled non-adiabatic absorption refrigeration cycle driven by low grade energy, Energy Convers. Manage., 86, (2014). Chen J., Havtun H., Palm B., Conventional and advanced exergy analysis of an ejector refrigeration system, Applied Energy, 144, (2015). Chen J., Havtun H., Palm B., Parametric analysis of ejector working characteristics in the refrigeration system, Appl. Therm. Eng., 69, (2014). Cimsit C., Ozturk I.T., Analysis of compression-absorption cascade refrigraiton cycles, Applied Thermal Engineering, 40, (2012). Cziesla F., Tsatsaronis G., Gao Z., Avoidable thermodynamic inefficiencies and costs in an externally fired combined cycle power plant, Energy, 31, (2006). Dahmani A., Aidoun Z., Galanis N., Optimum design of ejector refrigeration systems with environmentally benign fluids., Int. J. Therm. Sci., 50, (2011).

45 Erbay Z., Hepbasli A., Application of conventional and advanced exergy analyses to evaluate the performance of a ground-source heat pump (GSHP) dryer used in food drying, Energy Conversion and Management, 78, (2014). Gong S., Goni B.K., Parametric study of an absorption refrigeration machine using advanced exergy analysis, Energy, 76, (2014). Huang B.J., Chang J.M., Wang C.P., Petrenko V.A., A 1-D analysis of ejector performance, Int. J. Refrig.,22, (1999). Huang B.J., Jiang C.B., Hu F.L., Ejector performance characteristics and design analysis of jet refrigeration system, Trans ASME, 107, (1985). Kairouani L., Nehdi E., Cooling performance and energy saving of compressioneabsorption refrigeration system assisted by geothermal energy, Applied Thermal Engineering, 26, (2006). Kaska O., Energy and exergy analysis of an organic Rankine for power generation from waste heat recovery in steel industry, Energy Conversion and Management, 77, (2014). Keenan J.H., Neumann E.P., Lustwerk F., An investigation of ejector design by analysis and experiment, J. Appl. Mech. Trans. ASME, 72, (1950). McGovern R.K., Narayan G.P., Lienhard V.J.H., Analysis of reversible ejectors and definition of an ejector efficiency, Int. J. Therm. Sci., 54, (2012). Morosuk T., Tsatsaronis G., A new approach to the exergy analysis of absorption refrigeration machines., Energy, 33, (2008). Morosuk T., Tsatsaronis G., Advanced exergetic evaluation of refrigenration machines using different working fluids, Energy, 34, (2009). Morosuk T., Tsatsaronis G., Splitting the exergy destruction into endogenous and exogenous parts application to refrigeration machines. In: Frangopoulos C, Rakopoulos C, Tsatsaronis G, editors. Proceedings of the 19th international conference on efficiency, cost, optimization, simulation and environmental impact of energy systems, July 12 14; 2006, National Technical University of Athens, Greece, 1, (2006). Morosuk T., Tsatsaronis G., Zhang C.. Conventional thermodynamic and advanced exergetic analysis of a refrigeration machine using a Voorhees compression process, Energy Convers. Manage., 60, (2012). Petrakopoulou F., Tsatsaronis G., Morosuk T., Carassai A., Conventional and advanced exergetic 31 analyses applied to a combined cycle power plant, Energy, 41, (2012). Pridasawas W., Lundqvist P., An exergy analysis of a solar-driven ejector refrigeration system, Sol. Energy, 76, (2004). Tarique S., Siddiqui M.A., Performance and economic study of the combined absorption/compression heat pump, Energy Conversion Management, 40, (1999). Tsatsaronis G., Recent developments in exergy analysis and exergoeconomics, Int. J. Exergy, 5, (2014). Tsatsaronis G., Strengths and limitations of exergy analysis, In: Bejan A, Mamut E, editors. Thermodynamic optimization of complex energy systems. Dordrecht: Kluwer Academic Publishers, (1999). Tsatsaronis G., Morosuk T., A general exergy-based method for combining a cost analysis with an environmental impact analysis, Part I. Theoretical development, in: Proceedings of the ASME IMECE, Boston, Massachusetts, USA (2008a). Tsatsaronis G., Morosuk T., A general exergy-based method for combining a cost analysis with an environmental impact analysis. Part II. Application to a cogeneration system, in: Proceedings of the ASME IMECE, Boston, Massachusetts, USA (2008b). Tsatsaronis G., Park M.H., On avoidable and unavoidable exergy destructions and investment costs in thermal systems, Energy Convers. Manage., 43, (2002). Ustaoglu A., Alptekin M., Akay M.E., Energy and exergy analysis of a wet type rotary kiln and utilization of waste heat powered ORC, Applied Thermal Engineering, Submitted (2016). Yan G., Chen J., Yu J., Energy and exergy analysis of a new ejector enhanced auto-cascade refrigeration cycle, Energy Conversion and Management, 105, (2015). Yang X., Zheng N., Zhao L., Deng S., Li H., Yu Z., Analysis of a novel combined power and ejector-refrigeration cycle, Energy Conversion and Management, 108, (2016). Zhao L., Yang X., Deng S., Li H., Yu Z., Performance analysis of the ejector-expansion refrigeration cycle using zeotropic mixtures, International journal of refrigeration, 57, (2015). Zhu Y., Cai W., Wen C., Li Y., Shock circle model for ejector performance evaluation., Energy Convers. Manage., 48, (2014).

46 Thermodynamic Evaluation of Absorption-Compression Cascade Refrigeration Cycles for Advanced Exergy Analysis Mustafa Alptekin 1*, Abid Ustaoglu 2, Mehmet Emin Akay 3, Resat Selbas 1 1 Suleyman Demirel University, Faculty of Engineering, Department of Energy Systems Engineering, Isparta, 32260, Turkey 2 Bartin University, Faculty of Engineering, Mechanical Engineering Department, Bartin, 74100, Turkey 3 Karabuk University, Mechanical Engineering Department, Karabuk, 78050, Turkey. * Abstract In this paper, thermodynamic assessment of waste heat powered absorption-compression cascade refrigeration system is performed using advanced exergy method by using the exhausted heat from the chimney of rotary kiln. NH3-H2O fluid pairs and R134a are selected as working fluids in the absorption-compression cascade refrigeration cycles. The heat energy of generator is provided from waste heat of rotary kiln. Conventional and advanced exegy analyses are carried out to the system. The exergy destruction was split into endogenous-exogenous and avoidable-unavoidable parts in order to reveal interdependency within the components and determine the improvement potential of each component in the system. The largest exergy destruction occurs in the generator, accounting 63% of total exergy destruction. Therefore, it is important to concentrate on this component to improve overall system performance. In respect of overall system destruction, about 53.1% of the total destruction is caused by the system components themselves. Furthermore, about 60% of the total destruction is falling into the part of unavoidable exergy destruction. Namely, the system has potential to improve by reducing the avoidable part of 40%. Keywords: Advanced exergy analysis, absorption-compression cascade refrigeration system, refrigerant pairs I. Introduction Cement industry is one of the most energy consuming industries in the world. The energy consumption rate reaches 12-15% of total energy consumption in industry. The rotary kiln has biggest share in terms of energy consumption in a cement factory and has great heat loss which causes significant reduction in efficiency. This reduction becomes excessive for a cement factory using the wet method. In the progress of clinker production, about 33% of total energy is exhausted from the chimney of rotary kiln apart from the heat loss through the wall of rotary kiln Ustaoglu et al. (2016). Therefore, reutilization of the exhausted gas becomes substantial due to the great waste heat. Research and development studies for cooling systems have been gradually increased recently. The conventional refrigeration systems consume a great amount of energy through the compressor. Absorption systems are promising technologies in the refrigeration applications they can be operated by providing of a low level of thermal energy, including solar energy, biomass, geothermal and waste heat energy in the industrial process, compared to compression refrigeration systems Cai et al. (2014). Thus, the utilization of waste heat through production process and renewable energy even if low level can be provided in a useful form of energy thereby reducing amount of carbon emission and environmental pollutions. The compression-absorptions refrigeration cycles are another alternative among the refrigeration systems. Even tough more complex structure, these cycles can provide electrical energy save compering to vapor compression cycles and utilize advantages of absorption and vapor compression cycles by providing the use of electricity and heat for refrigeration. These type cycles also can be categorized into two group including combined refrigeration cycles and cascade refrigeration cycles. In the case of combined cycles, the compression rates of vapour compression and absorption sections are same with each other and that of the combined cycles. However, absorption and vapour compression cycles are connected in serial form for cascade systems Cimsit and Ozturk. (2014). There have been done many studies about the compression-absorption refrigeration cycles. Tarique and Siddiqui (1999) compared a conventional vapor compression cycle using ammonia with a combined refrigeration cycle using NH3 and NaSCN in the identical conditions in terms of performance and economical aspect. İt is found that the capital and operation costs were reduced significantly for the case of combined cycle. Kairouani et al. (2011) carried out a performance evaluation of compression-absorption refrigeration cascade cycle for NH3-H2O fluid pair for absorption section as the R717, R22 and R134a were used for the vapour compression section. The results showed that performance coefficient of the cycle have improvement of %37-54 comparing to a conventional 32

47 vapour compression cycle in which the same working fluids were used in the same operation conditions. Cimsit and Ozturk (2012) used LiBr-H2O and NH3-H2O pairs in the absorption section as R134a, R-410A and NH3 in the vapour compression section of a cascade refrigeration cycle. They showed that the cycle using LiBr-H2O fluid pair could show better performance coefficient compare to the cycle using NH3-H2O fluid pair for all working fluid of the vapour compression section. The second law efficiency evaluation is important tool to determine the location, magnitude and source of the exergy destruction. Rezayan and Behbahaninia (2011) carried out a thermoeconomic optimization of the system and applied exergy analysis to a CO2/NH3 cascade refrigeration cycle. The results showed that regarding to the exergy analysis on the optimized system, the highest exergy destruction occurred in the condenser (33.5%) as the lowest exergy destruction became in the expansion valve of the carbon dioxide circuit (5.2%). Yan et al. (2015) proposed a new ejector enhanced auto-cascade refrigeration cycle using R134a/R23 and compared it with a conventional auto-cascade refrigeration cycle. Their cycle achieved % better efficiency than a basic cycle at the same operation conditions as the ejector achieved pressure ratio lifts of They found that the highest exergy destruction occurs in the compressor, the condenser, cascade condenser, expansion valve, ejector and evaporator, respectively. Cimsit et al. (2015) carried out thermoeconomic optimization of LiBr/H2O-R134a compression-absorption cascade refrigeration cycle. The analysis results showed that the evaporator equipage and solution heat exchanger should be designed carefully according to the exergoeconomic factor values. Apart from the conventional analysis method, the exergy method is an important tool that can show the useful work that can be generated through the process. However, it cannot explain the interaction among the component or estimate the actual improvement potential. Conventional approach for optimizing of the system may be wrong without taking into account of the interaction between the components particularly for complicated systems in which many components having interaction with each other. In the case of the ejector refrigeration cycles, the operation parameters of the ejector depend on itself and other components Chen et al. (2014). A recent developed technique, the advanced exergy analysis by splitting the exergy destruction into avoidable-unavoidable and endogenous-exogenous part can enable us to evaluate the system in more detail and investigate the capacity of the improvement Tsatsaronis (1999). The advanced exergy analysis of different refrigeration systems, including vapor compression refrigeration systems Morosuk and Tsatsaronis (2009), Morosuk et al. (2012), absorption refrigeration systems Gong and Goni (2014), Morosuk T, Tsatsaronis (2008) ejector refrigeration system Chen et al. (2015) and heat 33 pump Erbay and Hepbasli (2014) have been carried out. To the best of author s knowledge, there is no study about evaluation of an absorption-compression cascade refrigeration cycle in terms of advanced exergy analysis. In a previous study, energy and exergy analysis of a wet type rotary kiln were carried and recovery capacity of waste heat was evaluated by using an organic Rankine cycle Ustaoglu et al. (2016). The results showed that a great amount of heat energy of 30.5 MW is exhausted from the chimney of rotary kiln. In this study, the waste heat from the rotary kiln was evaluated for an absorption-compression cascade refrigeration cycle and advanced exergy analysis was carried out to determine endogenous, exogenous, avoidable and unavoidable exergy destruction rates in order to determine the improvement potentials of both the components and the overall system along with the interactions within components. II. Material and method II.1. System description Figure 1 shows the schematic view of the absorption compression cascade refrigeration cycle. The below and upper cycles are vapor compression and absorption refrigeration cycles, respectively. R134a is used as working fluid in the case of vapor compression cycle. LiBr-H2O and NH3-H2O pairs are selected for absorption refrigeration cycle. Each state of the working fluids is represented in the point as seen in figure. The working principle of this system can be expressed as follows: Vaporized working fluid in evaporator enters to the compressor to be compressed to point 2. The compressed and slightly heated fluid releases its heat to the other working fluid having lower temperature in the upper cycle, and leaves the heat exchanger. In the expansion valve, the temperature and pressure of the working fluid decreases to evaporator level. The working fluid entering evaporator absorbs the heat to be vaporized and again enter to the compressor. Thus, the below cycle is completed. As the refrigerant is circulated in points 11, 12, 13 and 14, fluid pair is circulated in points 5, 6, 7, 8, 9 and 10. The vaporized refrigerant by receiving heat from the below cycle enters from point 14 as weak solution enters from point 10 to the absorber where they dissolves and reacts with each other to form strong solution, fluid pairs. This strong solution is pumped to the solution heat exchanger and receives some of the heat of the fluid pairs coming from the generator and enters to generator where the heat of the exhausted gas is transfer to the fluid pair. The fluid having low boiling point in the strong solution is vaporized and enters to the condenser. However, all of that fluid cannot be vaporized. The weak solution leaving from the generator releases some of its heat in the heat exchanger and enters to the expansion

48 valve where its heat and pressure decreases to the absorber level. The fluid leaving generator from point 11 is condensed in condenser and enters to the expansion valve where its pressure and temperature decreases then the fluid enters to the heat exchanger to be vaporized by heat from the below cycle. Thus, all cycle is completed. frictions Petrakopoulou (2011). Conventional exergy analysis can describe components having high exergy destruction and its reasons, which component has irreversibility and its magnitude. However, it cannot explain the interaction among the component or estimate the actual improvement potential. Conventional approach to optimize the system may be wrong without taking into account of the interaction between the components particularly for complicated systems in which many components having interaction with each other. Therefore, advanced exergy analysis has been carried out. A detail exergy analysis, where the exergy destruction is split into several parts, is called to be advance exergy analysis. These parts are avoidable and unavoidable, or endogenous and exogenous exergy destructions Morosuk and Tsatsaronis (2009). The value of the total exergy destruction can be calculated through an exergy balance for this component as follows E D,k = E F,k E P,k = T 0 S P,k (4) where T 0, E F,k and E P,k shows reference temperature, exergetic fuel and exergetic product, respectively. Thus, the exergy efficiency can be found by following equations for kth component ε k = E P,k E F,k = 1 E D,k E F,k (5) Fig. 1: Schematic view of absorption-compression cascade refrigeration cycle II.2. Thermodynamic Evaluation The absorption- compression cascade refrigeration cycle is considered as continuous flow open-system and modeled based on the first and second laws of thermodynamics, and these laws are applied to each component in the system. In the continuous flow open systems, the mass and energy of control volume are stable. General equations of mass, energy and exergy balances by ignoring the kinetic and potential energy variations can be expressed m in = m out (1) Q + W = m out h out m outh out (2) E heat + W + E D = E out E in (3) where ṁ is mass flow rate, h is enthalpy, Q is heat, and Ẇ is net work. Basic thermodynamic evaluation and conventional exergy analysis can be carried out by using these equations. All actual operations are irreversible due chemical reactions, heat transfer at finite temperature difference, mixture of matters, infinite expansion and Exergy efficiency of the overall system can be determined by ε overall = COP COP carnot (6) where the coefficient of performance COP can be calculated by COP = Q EV/(Q GE + W CO + W PU) (7) where Q GE = m 8h 8 + m 11h 11 m 7h 7 (8) W CO = m R(h 2 h 1 ) (9) W PU = m 5(h 6 h 5 ) (10) ṁ R indicates the mass flow rate of the R134a. The coefficient of performance for Carnot can be obtained by COP carnot = ((T GE T AB )T EV ) ((T CO T EV )T GE ) (11) where T indicates the temperature, the subscripts GE, AB and EV represent generator, absorber and evaporator. In the evaluation, the temperature should be in Kelvin. The exergy destruction E D,k can be split into 34

49 AV UN avoidable E D,k and unavoidable E D,k parts, Tsatsaronis and Park (2002), Cziesla et al. (2006), Tsatsaronis and Morosuk (2008a, 2008b), Petrakopoulou et al. (2012). E D,k = E D,k UN AV + E (12) D,k This can provide more realistic approach to measure the potential improvement of the thermodynamic efficiency of the components. The unavoidable part of the exergy destruction cannot be shrunk due to the technological limitations including material and manufacturing cost and availability. The other part of indicates avoidable destruction. EN The other approach is split into endogenous E D,k EX and exogenous E D,k parts as follows, Tsatsaronis (1999), Morosuk and Tsatsaronis (2006) E D,k = E D,k EN EX + E (13) D,k The endogenous and exogenous parts of the exergy destruction provide the reason of the exergy destruction in the component caused by the EN component itself or by the other components. E D,k part can be obtained when the all irreversibilities occurs in the kth component while the other component is assumed to be ideal and has no irreversibility with having its current efficiency. On the EX other hand, E D,k part occurs within the kth component due to the irreversibilities in the other components. For each case, the power output of the overall system is kept constant and equal to actual case. Thus, the exogenous destruction is the remaining part of the total exergy destruction in kth component and can be calculated by subtracting the endogenous exergy destruction from the total exergy destruction Petrakopoulou (2011). In order to calculate the unavoidable exergy destruction, each component should be considered isolated and separated from the system. The exergy UN destruction rate per unit product exergy (E D E P ) k can be calculated by assuming the system operating with high efficiency and low losses. Thus, the unavoidable exergy destruction for kth component by using the real case product exergy rate can be expressed by Petrakopoulou (2011) UN = E E D,k P,k real (E D,k E P,k ) UN (14) When the unavoidable exergy destruction is known, the avoidable part can be calculated by Eq 12. The unavoidable endogenous and exogenous, and avoidable endogenous and exogenous part of the exergy destructions are expressed respectively, by Tsatsaronis and Morosuk (2008a, 2008b) AV,EN = E EN E E D,k E D,k D,k AV,EX = E AV E D,k UN,EN D,k AV,EN D,k (17) (18) In order to evaluate the advanced exergy analysis, modified exergy efficiency can be described by Tsatsaronis and Morosuk (2008a, 2008b) ε modified = E P,k (E F,k E D,k UN E AV,EX ) (19) In order to facilitate the analyses some assumptions are made; 1. Each process in the system is assumed to be steady state. 2. Potential and kinetic energy variations are neglected. 3. The heat transfer to/from ambient and pressure drops in the pipes are neglected. 4. The fluid pairs in absorption refrigeration system at the inlet of pump is assumed as saturated liquid. III. Results and Discussions The calculation program was written in Engineering Equation Solver (EES). Table 1 shows the input parameters in order to carry out the energetic and exergetic analysis of the cycle. Table 2 expresses the parameters of the components in the absorption-compression cascade refrigeration system for the real, ideal and unavoidable conditions. The calculated thermodynamic data of the absorption-compression cascade refrigeration at real operation case, at ideal operation case and at the unavoidable operation conditions for the working fluid R142b are shown in Tables 3-5, respectively. Tab. 1: Input parameters to the system Parameters Values Pump isentropic efficiency 85% Compressor isentropic efficiency 85% Evaporator temperature -20 C Condenser temperature 45 C Generator temperature 100 C Absorber temperature 43 C Heat exchanger temperature for bottom cycle 5 C Heat exchanger temperature for upper cycle 20 C Inlet temperature to generator of exhausted gas 277 C Inlet temperature to condenser of cooling water 25 C Outlet temperature from condenser of cooling 30 C water Inlet temperature to evaporator of water 30 C Outlet temperature from evaporator of water 25 C Inlet temperature to absorber of cooling water 30 C Outlet temperature from absorber of cooling water 35 C Subcooling and superheating temperature 5 C Mass flow rate of exhausted gas 44.5 kg/s Effectiveness of solution heat exchanger 50% Cooling capacity 20 kw Ambient pressure kPa Ambient temperature 25 C D,k UN,EN = E E D,k P,k EN (E D,k E P,k ) UN (15) UN,EX = E UN E E D,k D,k UN,EN D,k (16) 35

50 Tab. 2: The parameters used for real, ideal cycles, and the cycle for the unavoidable exergy destruction Component Parameter Real Ideal Unavoidable Evaporator ΔT EV 10 C C Compressor η C Heat exchanger ΔT HE 5 C C Expansion valves - Isenthalpic Isentropic Isenthalpic Condenser ΔT CO 10 C C Absorber ΔT AB 3 C C Pump Solution heat Exchanger β Generator ΔT GE 7 C C Tab. 3: Data of the ejector refrigeration cycle at real operation case Loc. Subs. m P h s e T ( C) (kg/s) (kpa) (kj/kg) (kj/kg.k) (kj/kg) 1 R134a R134a R134a R134a R134a NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH NH NH NH Water Air Air Air Water Water Water Water Water Water Water Tab. 4: Data of the ejector refrigeration cycle at ideal operation case Loc. Subs. m P h s e T ( C) (kg/s) (kpa) (kj/kg) (kj/kg.k) (kj/kg) 1 R134a R134a R134a R134a NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH NH NH NH Tab. 5: Data of the ejector refrigeration at the unavoidable operation conditions Loc. Substance m P T ( C) h s e (kg/s) (kpa) (kj/kg) (kj/kg.k) (kj/kg) 1 R134a R134a R134a R134a NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH 3 H 2O NH NH NH NH III.1. Evaluation of Conventional Exergy Analysis Table 6 shows the results of the conventional exergy analysis of absorption-compression cascade refrigeration system. The generator shows the largest exergy destruction on the cycle which is more than one half of the overall exergy destruction rate of system (%63.86), and followed by evaporator (10.8%) and solution heat exchanger (6%). The lowest exergy destructions are observed in pump, and followed by the expansion valve 2, 1 and 3, respectively. The exergy efficiency of pump is the highest in the system and followed by Expansion valve 3, 1 and 2. These are arisen from the relatively low exergy destructions of these systems compare to their fuel and product exergy. The lowest exergy efficiency is obtained in evaporator with about 23% and followed by condenser, generator and heat exchanger with 28%, 32% and 42%, respectively. The other components show quite preferable performances. The system COP and COPcarnot were calculated to be and , respectively. As for the overall system, the efficiency, which can be calculated by Eq 6, is about 46%. The fuel exergy of overall system is composed of the fuel exergy of generator, compressor and pump as the exergy product of overall system indicates that of evaporator. Tab. 6: The conventional exergy analysis results Component Ė F (kw) Ė P (kw) Ė D (kw) ε (%) Compressor (COMP) Heat exchanger (HE) Expansion valve Evaporator (EV) Condenser (CO) Expansion valve Absorber (AB) Pump (PU) E Solution heat exchanger(she) Generator (GE) Expansion valve Overall System III.2. Evaluation of Advanced Exergy Analysis Table 7 shows the advanced exergy analysis results where the exegy destruction rates are split into endogenous-exogenous and unavailable-available 36

51 parts for each component. The second column indicates the exergy destruction for each component. The third and fourth columns show the endogenous EN and exogenous parts, E D,k ande D,k EX, those state how much exergy destruction caused by components itself and how much exergy destruction caused by the other components in the cycle. Tab. 7: The advanced exergy analysis results Component Compressor (COMP) Heat exchanger (HE) Expansion valve 1 Evaporator (EV) Condenser (CO) Expansion valve 2 Ė D (kw Splitting the exergy destruction Ė EN (kw) Ė EX (kw) Ė UN (kw) ĖAV (kw) Absorber (AB) Pump (PU) Solution heat exchanger (SHE) Generator (GE) Expansion valve 3 Overall System 1.02E E E E E In Figure 2, the endogenous exogenous parts were shown in terms of percentages in the total exergy destruction. In the case of evaporator, the exergy destruction is mostly caused by the component itself. The endogenous exergy destruction is more dominant for the compressor and generator compare to the exogenous. The remaining components exergy destructions are mostly caused by the rest of the components. For the case of expansion valve 3, pump and absorber, the exogenous exergy destruction rates are over the 90% and followed by compressor and solution heat exchanger with about 82% and 77%, respectively. The heat exchanger, expansion valve 1 and expansion valve 2 show similar behaviors and have exogenous exergy destruction rates of about 65%. As regards to the overall system, the endogenous part of the exergy destruction is comparable with the exogenous exergy destruction. Figure 3 shows the avoidable and unavoidable parts of the total exergy destruction for each component. These results state the improvement potential of each system with more realistic approach. Although 42.3% of total exergy destruction is avoidable part in generator, it shows the largest avoidable part of the total exergy destruction and that is more than two third of avoidable part of overall system as seen in Table 7. Therefore, it is important to concentrate on this component to improve overall system performance. The second biggest share of the avoidable part is observed in the solution heat exchanger and 75% of the exergy destruction can be avoided. The third largest share is the absorber and 81% of the total exergy destruction is avoidable part for that component. Although the second largest exergy destruction occurs in the evaporator, all exergy destruction is unavoidable part. In the case of condenser, it is 93.2%. The highest rates of avoidable parts are observed expansion valve 3, absorber, and pump and solution heat exchanger, respectively. Together, they contribute 25% of the avoidable part of overall system. However, some components including expansion valves and pump have very small share. As for overall system, about 60% of the total exergy destruction is unavoidable part while the 40% of the exergy destruction can be eliminated. Fig. 2: Exergy destruction rate for exogenous and endogenous parts Fig. 3: Exergy destruction rate for avoidable and unavoidable parts 37

52 IV. Conclusions Conventional and advanced exegy analyses are carried out for absorption-compression cascade refrigeration cycles. Moreover, the exergy destruction was split into endogenous-exogenous and avoidable-unavoidable parts in order to reveal interdependency within the components and determine the improvement potential of each component in the system. The following part summarizes the study. 1. The largest exergy destruction occurs in generator, accounting 63% of total exergy destruction and followed by evaporator and solution heat exchanger with ratios of 10.8% and 6.6%, respectively. 2. The exergetic efficiency of the overall system is about 46%. 3. The endogenous exergy destruction is more dominant for the compressor and generator compare to the exogenous part of the exergy destruction. 4. In the case of the expansion valve 3, pump and absorber, the exergy destruction can be largely reduced by improvement of other remaining components. 5. The largest share of total avoidable exergy destruction occurs in generator and that is more than two third of avoidable part of overall system (68.23%). Therefore, it is important to concentrate on this component to improve overall system performance. 6. Although the second largest exergy destruction occurs in the evaporator, the 100% of exergy destruction is caused by unavoidable exergy destruction part. 7. In respect of overall system destruction, about 53.1% of the total destruction is caused by the system components themselves. Furthermore, about 60% of the total destruction is falling into the part of unavoidable exergy destruction. Namely, the system has potential to improve by reducing the avoidable part of 40%. Nomenclature E : Exergy (m.s-1) h : Specific enthalpy (kj.kg-1) ṁ : Mass flow rate (kg.s-1) P : Pressure (kpa) Q : Heat load (kw) s : Specific entropy (kj.kg-1 K-1) T : Temperature (C) W : Work (kw) Greek letters ε : Exergetic efficiency Superscripts AV : Avoidable EN : Endogenous EX : Exogenous UN : Unavoidable Subscripts D : Destruction F : Fuel 38 k : The k-th component P : Product Abreviations AB : Absorber EV : Evaporator COMP : Compressor HE : Heat exchanger EXV : Expansion valve EXP V : Expansion valve PU : Pump SHE : Solution heat exchanger GE : Generator References Cai D.H., He G.G., Tian Q.Q., Tang W.E., Thermodynamic analysis of a novel aircooled non-adiabatic absorption refrigeration cycle driven by low grade energy, Energy Convers. Manage., 86, (2014). Cimsit C., Ozturk I.T., Analysis of compression-absorption cascade refrigraiton cycles, Applied Thermal Engineering, 40, (2012). Cimsit C., Ozturk I.T., Hosoz M., Second law based thermodynamic analysis of compression-absorption cascade refrigeration cycles, J. of Thermal Science and Technology, 34, 9-18 (2014). Chen J., Havtun H., Palm B., Parametric analysis of ejector working characteristics in the refrigeration system, Appl. Therm. Eng., 69, (2014). Cziesla F., Tsatsaronis G., Gao Z., Avoidable thermodynamic inefficiencies and costs in an externally fired combined cycle power plant, Energy, 31, (2006). Erbay Z., Hepbasli A., Application of conventional and advanced exergy analyses to evaluate the performance of a ground-source heat pump (GSHP) dryer used in food drying, Energy Conversion and Management, 78, (2014). Gong S., Goni B.K., Parametric study of an absorption refrigeration machine using advanced exergy analysis, Energy, 76, (2014). Kairouani L., Nehdi E., Cooling performance and energy saving of compressioneabsorption refrigeration system assisted by geothermal energy, Applied Thermal Engineering, 26, (2006). Kaska O., Energy and exergy analysis of an organic Rankine for power generation from waste heat recovery in steel industry, Energy Conversion and Management, 77, (2014). Morosuk T., Tsatsaronis G., Advanced exergetic evaluation of refrigenration machines using different working fluids, Energy, 34, (2009). Morosuk T., Tsatsaronis G., Splitting the exergy

53 destruction into endogenous and exogenous parts application to refrigeration machines. In: Frangopoulos C, Rakopoulos C, Tsatsaronis G, editors. Proceedings of the 19th international conference on efficiency, cost, optimization, simulation and environmental impact of energy systems, July 12 14; 2006, National Technical University of Athens, Greece, 1, (2006). cycle, Energy Conversion and Management, 105, (2015). Morosuk T., Tsatsaronis G., A new approach to the exergy analysis of absorption refrigeration machines., Energy, 33, (2008). Morosuk T., Tsatsaronis G., Zhang C.. Conventional thermodynamic and advanced exergetic analysis of a refrigeration machine using a Voorhees compression process, Energy Convers. Manage., 60, (2012). Petrakopoulou F., Tsatsaronis G., Morosuk T., Carassai A., Conventional and advanced exergetic analyses applied to a combined cycle power plant, Energy, 41, (2012). Rezayan O., Behbahaninia A., Thermoeconomic optimization and exergy analysis of CO2/NH3 cascade refrigeration systems, Energy, 36, (2011). Tarique S., Siddiqui M.A., Performance and economic study of the combined absorption/compression heat pump, Energy Conversion Management, 40, (1999). Tsatsaronis G., Strengths and limitations of exergy analysis, In: Bejan A, Mamut E, editors. Thermodynamic optimization of complex energy systems. Dordrecht: Kluwer Academic Publishers, (1999). Tsatsaronis G., Morosuk T., A general exergy-based method for combining a cost analysis with an environmental impact analysis, Part I. Theoretical development, in: Proceedings of the ASME IMECE, Boston, Massachusetts, USA (2008a). Tsatsaronis G., Morosuk T., A general exergy-based method for combining a cost analysis with an environmental impact analysis. Part II. Application to a cogeneration system, in: Proceedings of the ASME IMECE, Boston, Massachusetts, USA (2008b). Tsatsaronis G., Park M.H., On avoidable and unavoidable exergy destructions and investment costs in thermal systems, Energy Convers. Manage., 43, (2002). Ustaoglu A., Alptekin M., Akay M.E., Energy and exergy analysis of a wet type rotary kiln and utilization of waste heat powered ORC, Applied Thermal Engineering, Submitted (2016). Yan G., Chen J., Yu J., Energy and exergy analysis of a new ejector enhanced auto-cascade refrigeration 39

54 Exergy Optimization of the Hybrid Compression-Absorption Industrial Refrigeration Systems Mahmoud Afshar*, Hamid Rad, Petroleum University of Technology, MahmoudAbad, Energy, , Iran * Abstract To prevent the creation of condensate liquid in gas trunk lines, the heavy molecules of the supply natural gas have to be removed. Proper natural gas dew point adjustment process is crucial in the separation of its heavy molecules: It is typically performed by the compression refrigeration system with a vast power demand for gas compressors, usually provided by gas turbines. In this paper a Hybrid Compression-Absorption Refrigeration System (HCARS) is presented and optimized in terms of exergy losses. The main idea in the proposed hybrid system is to utilize the recovered compression refrigeration system gas turbines hot exhaust gases (with a temperature of 540 ) as the heat source required for the Absorption Refrigeration System (ARS) absorber regeneration process which would reduce energy consumption and greenhouse gases generation. The proposed system is worked out for the dew point unit of a gas refinery which results in 63% additional cooling capacity of the hybrid system (12550 KW) in comparison o the current compression refrigeration system (7670 KW) for the same fuel gas consumption. The coefficient of performance of the hybrid system is , and totally about SCMD of fuel gas is saved by the possibility of shutting off one of three running turbo-compressors of the unit. Keywords: Gas Refinery Dew Point Unit, Hybrid Refrigeration System, Exergy Destruction Optimization I. Introduction Energy efficiency has received heightened interest with the increasing attention paid to climate change and environmental pollution. It has been a major topic of discussions on natural resources preservation and costs reduction. Environment preservation must also be considered through energy optimization studies. Oil and gas industry consumes a large amount of energy and gas turbines are typically used to drive mechanical equipment and to generate electricity. High temperature exhaust gas from the gas turbine is discharged to the surroundings. Energy efficiency can be improved by using an absorption refrigeration system (ARS) to convert waste heat into useful cooling energy. The application of ARSs reduces the energy consumption of conventional vapor compression refrigerators. Hydrocarbon liquid dropout can cause a number of problems in gas transmission lines, including increased pressure drop, reduced line capacity, and equipment problems such as compressor damage. To avoid liquid dropout, most current operating specifications for gas transmission lines require that the lines be operated above the hydrocarbon dew point. As noted, natural gas dew point temperature is regulated by refrigeration units and needs a lot of cooling load. Benjamin Brant et al (1998) has Installed the first kind Waste Heat Absorption Refrigeration Plant. The refrigeration unit was designed to provide refrigeration for two process units at the refinery while utilizing waste heat as the energy source. The ARS enables to cool the gas to C, thereby recovering 45% of high molecular weight hydrocarbons by kalinowski et al (2009). Priedeman and Christensen (1999) presented a general ammonia-water absorption heat pump cycle that was modeled and tested. The experimental findings were found to be in close agreement with the simulation results by Jianbo et al (2013). Studies on the absorption and compression refrigeration systems in recent years were continued. A.K. Pratihar et al. (2011) simulated an ammoniawater compression-absorption system by incorporating detailed heat and mass transfer calculations in the absorber and desorber of the system. They studied the effect of relative solution heat exchanger area on the COP. With increase in the relative solution heat exchanger area from 17% to 50%, COP increased initially, got maximum value at 39% and then decreased. N.A. Darwish et al. (2013) analyzed the absorption refrigeration water ammonia system using Aspen Plus flow sheet simulator. A very good agreement between the simulator s results and the experimental measurements was found. Based on the above Description, in addition to simulation of absorption refrigeration cycles, analysis of them requires thermodynamic equations to 40

55 knowledge of simulation accuracy and compliance the design assumptions with the physics laws. Morethrefore simulations that have been associated with experimental models, have been matched with simulation software, especially sequential software. In the dew point units of natural gas, compression refrigeration systems are mainly used to achieve the desired temperature. Absorption refrigeration cycles can be used alongside conventional compression cycle to achieve the desired temperature. In other words, the hybrid refrigeration system used. A hybrid absorption compression refrigeration system consists of a heat-driven compression refrigeration subsystem and an absorption refrigeration subsystem, which are integrated by an absorber and a rectifier. High-temperature flue gas successively goes into a heat recovery generator to heat the working fluids of the compression refrigeration subsystem. Exist of restrictions in the unit and keeping constant defines the basic structure of the propane cycles cause going away from the original definition of hybrid refrigeration cycle and intersections of absorptioncompression refrigeration limited to nature of the system does not change. II. Energy Analysis of Absorption System Absorption systems have been modeled in the past in a variety programs, such as the one by Lazzarin et al (1996). Modern modeling is usually done by one of two software, developed by Oak Ridge National Laboratory (2014) and Engineering Equation Solver (EES), developed at the University of Wisconsin allows the user to compute thermophysical properties of working fluids, providing results with very good accuracy when compared to experimental results by Liao et al (2004). Selecting the correct property method is crucial for getting meaningful results. As in the previous paper written on modeling ammonia/water systems in ASPEN HYSYS program, it was found that Peng- Robinson was the best available method by Darwish (2013). This section explores the possibility of using absorption chillers to utilize waste heat. To accomplish this, models are created in ASPEN HYSYS and a variety of cycle options are considered. The waste heat source is the exhaust stream from a gas turbine, and since gas turbine models are already available, the bulk of the modeling work reported here focuses on developing absorption cycle models by Rad et al (2015). In fact, many calculations of thermodynamics are based on the assumption of ideal conditions such as reversible processes, real processes are nevertheless amenable to thermodynamic analysis. The object of such an analysis is to determine the overall efficiency of the use of energy and to calculate 41 the inefficiencies of the various steps of a certain process. The principle of mass conservation and the First and Second Laws of Thermodynamics were applied to each component of the system for the analysis. Every component was considered as a control volume, taking into account the heat transfer, work interaction and inlet and outlet streams. The gas turbine was not modeled at the same level of detail as the absorption chiller, since it is not the focal point of this work by Rad and Afshar (2015). The governing equations for mass conservation are: Mass balance: m i m o = 0 (1) Mass balance for ammonia: m ix i m ox o = 0 (2) Where x i and x o correspond to the inlet and outlet ammonia mass fractions. Energy balance: The First Law of Thermodynamics yields the energy balance of each component of the whole system in following form Q W = m oh o m ih i (3) Entropy balance: S gen,k = m es e m is i Q K T K (4) The known data of the model are the pump s flow rate, the condition of some streams, maximum and minimum pressure of cycle, ammonia concentration in all streams and the values of other points must therefore be calculated. II.1. Exergy Analysis In the absence of nuclear, magnetic, electrical and surface tension effect, the total exergy of the system E can be divided into four components: physical exergy E PH, kinetic exergy E K, potential exergy E PT and chemical exergy E CH. If the kinetic and potential exergy are neglected: E = E PH + E CH (5) E PH = (h h 0 ) T 0. (s s 0 ) (6) E CH = x n M a. e a ch + 1 x n M w. e w ch (7) Exergy destruction at the individual component level could also be calculated using the following equation: E D,k = in m (h T 0 s) m (h T 0 s) out (8)

56 Where T 0 is a reference temperature maintained at K here. It is noted that the definition of the exergy destruction rate in (Eq. 8) accounts for both the physical and chemical exergy of the fluid streams, as suggested by Palacios Bereche et al. and is also consistent with the approach of Morosuk et al (2005). The calculation procedure of the chemical exergy of various substances based on standard chemical exergy values of respective. Exergy balance at the steady state for a control volume reads: i m i. e i + (1 T 0 j ). Q T j = e m j e. e e + W CV + E D (9) A detailed exergy analysis includes calculation of exergy destruction, exergy loss, exergetic efficiency, two exergy destruction ratios in each component of the system along with the overall system and second law efficiency for the cycle. Mathematically these are expressed as follows: E D = E F E P E L (11) ε = E P E F = 1 [ E D+E L ] (12) E F II.2. COP and exergetic efficiency The COP is defined as the useful heat rate from the Q E divided by the required heat rate to the generator Q G by Lostec et al (2013). COP Abs = Q E = [m (h e h i )] ammonia (12) Q G [m (h i h e )] steam The COP of compression refrigeration system is defined as: COP Comp = Q Evap W Comp (13) Which can be defined COP for compression refrigeration operating units that are using gas turbines to run compressors of compression refrigeration system as: COP Comp = Q Evap Input Heating Energy of NG (14) Therefore, with the help of the equations 12 and 14, the COP of Hybrid refrigeration System defined by Rad (2015) as: COP Hybrid = (Q Evap) Abs +(Q Evap) Com Input Heating Energy of NG (15) The purpose of this section is to compare of the absorption cycle COP with compression refrigeration cycle COP and achieve to definition of hybrid refrigeration system COP that uses gas turbine as power source. III. Hybrid Refrigeration System The Dew Point Unit is already working utilizing standard mechanical compression refrigeration systems, and the proposed Hybrid Refrigeration System design is explained based on this refinery design specifications as well as real operational working conditions. Gas turbine exhaust operational specifications are shown in Table 1. Tab.1: Exhaust flue gas operational specifications Exhaust flue gas temperature 540 Exhaust flue gas mass flow rate 75.5 kg s Exhaust flue gas pressure kpa The cooling of the main cycle evaporator, i.e MW, is utilized for cooling of natural gas parallel to the available compression refrigeration cycle as shown in Figure 1. Considering the 7.26 MW cooling of natural gas in operational compression refrigeration cycle, the hybrid system natural gas cooling capacity will be =11.75 MW, which is 1.62 times the original cooling capacity. The original design is for cooling of 19.95MMSCMD natural gas, therefore the proposed hybrid system can cool MMSCMD of natural gas. Table.2 Shows the COP of Compression System, Absorption Refrigeration System that added to available system and Hybrid Absorption-Compression Refrigeration System COP in design condition and real condition. Tab.2: COP of System in Design and Real Condition Design COP Real COP Compression system Absorption system Hybrid system The results presented indicate the feasibility of the proposed hybrid refrigeration system for natural gas cooling under current real working condition of the compression refrigeration cycle. IV. Results of Advanced Analysis Exergy analysis of the process is performed and the process is optimized to minimize the exergy destruction of the process. The effect of partial load operation of the available compression refrigeration system on the performance of ARS is shown in Figure 3. This is done by extrapolation of the real field data collected in compression refrigeration cycle in Fajr-e-jam refinery by Ashar and Rad (2015). The exhaust temperature and mass flow rate of the flue gas is obtained with second order regression. As shown in Figure 3. The absorption cycle cooling load changes are within 5% of full load operation. Figure 4 also shows COP of Hybrid Refrigeration System in partial cooling load of compression refrigeration system. 42

57 Fig. 1: Proposed Hybrid Refrigeration System Layout Fig. 2: Schematic View of Suggested Cascade Absorption Cycle for Dew Point Unit 43

58 In the field data collected from propane compression refrigeration system by Afshar and Rad (2015), reduced cooling load of compression refrigeration system results in the exhaust flue gas temperature and mass flow rate changes: mass flow rate reduces in direct ratio with cooling load of compression system, while the temperature increases. Therefore, the increase in flue gas temperature compensates to some degree the drop in its mass flow rate. Indeed, despite reduction in compression refrigeration cooling load, the absorption refrigeration cycle cooling load remains almost unchanged and works near full load conditions. Tab. 3: Stream Constraints used in optimization of the proposed hybrid refrigeration system Stream Constrains Strea Constrains m 1A Quality=0 1B Quality=0 3A 4A Quality=0 Quality=0 X 3 X 5 3B X 3 X 5 T 5 T T 5 T X NH X NH T B T 4 40 P 5 P 4 P 5 P 4 5A Quality = 0 5B Quality = 0 P 5 P 3 &P 4 P 5 P 3 &P 4 6A T 6 T 2 6B T 6 T 2 8A Quality = 0 P 8 P 4 P 8 P 4, 8B T 8 T 4 T 8 T 4 P A P B T A T B Quality A Quality B P 12 P 1 12A P 12 P 1 12B X 14 = X 1 14A X 14 = X 1 T 14 = T 1 T 14 = T 1 14B P 14 = P 1 P 14 = P 1 P 8 P 4 Fig. 3: Cooling Load of Absorption Cycle Based on Partial Cooling Load of Compression Cycle Fig. 4: COP of Hybrid Refrigeration System in Partial Cooling Load of Compression Cycle The proposed hybrid refrigeration cycle is optimized by using optimization methods (such as Sequential Linear Programming Method) in HYSYS (version7.2). The objective function is: The total optimized Exergy destruction rate is 4005 KW, and the refrigeration rate increased to 4.68 MW from previous 4.49 MW, and also the COP of the cascade absorption refrigeration cycle increased to from its previous value of The exergetic efficiencies of all components are calculated. The rate of fuel exergy ( E Fuel ), rate of product exergy (E P), rate of destruction exergy (E D), rate of loss exergy ( E L ) for all components are calculated first. Four other parameters that are first exergy destruction ratio ( Y D ), second exergy destruction ratio (Y D ), exergy loss ratio (Y L ) and finally the exegetic efficiency are calculated based on the above four values. All of theses parameters are defined in section 3.2 of chapter three. The values of these parameters for all components of hybrid refrigeration cycle are shown in Table 4. Only % of the input exergy is converted to cooling as the system useful output, while the remaining exergy is either lost to the environment or destructed due to irreversibilities in the various components of the system. The total exergy supplied to the system is kw, out of which % is converted to useful product which is equivalent to kw, % is destroyed which is equivalent to kw, and the remaining 31.1 % is lost to environment which is equivalent to kw. n Minimize k=1 (E D,k) (k is number of cascade cycle components) Subject to constrains shown in table3. (16) 44

59 Tab. 4: Exergetic Destruction, Loss and Efficiency Using EDM Component E Fuel E P E D E L Y D Y L Y D ε kw kw kw kw % % % % Pump(A) Generator(A) Rectifier(A) SHX(A) Valve1(A) Cond.Evap.assly(A) L/V HX(A) Valve2(A) Absorber(A) Pump(B) Generator(B) Rectifier(B) SHX(B) Valve1(B) Cond.Evap.assly(B) Valve2(B) L/V HX(B) Absorber(B) Overall System The highest total exergy destruction is found in SHX- A which is equal to 9.08 % of the total exergy destruction of the system. The second highest exergy destruction is found to be in the SHX-B which amounts to kw, equivalent to 7.89 % of the total exergy destruction. The exergetic efficiency for condenser-evaporator assembly and L/V HX of main and cascade cycles are almost low. The reason for low exergetic efficiency of condenser-evaporator assembly may be attributed to the large temperature difference between the working fluid and brine in the evaporator and the working fluid and cooling air in the condenser. Similarly the high temperature difference between both the fluids and relatively very less mass flow rate of gaseous ammonia from the evaporator compared to the condensed ammonia from the condenser may be the reason for the poor exergetic efficiency of L/V HX. V. Conclusion Energy efficiency can be improved by using an absorption refrigeration system to convert waste heat into useful cooling energy and one example is the dew point unit of gas refineries. Since the source of the heat is the flue gas of compression refrigeration systems, a hybrid compression-absorption refrigeration system is proposed in this paper for full achievement of possible energy savings. By conducting exergy destruction optimization, the total minimum exergy destruction rate is 4005 KW, while the ARS cooling rate and COP increased to 4.68 MW and from their its previous values of 4.49 MW and , respectively. The exergetic efficiencies of the rectifiers and absorbers are the highest, and the exergetic 45 efficiencies of condenser evaporator assemblies and liquid/vapor heat exchangers are the least. The highest total exergy destruction is found in SHX-A which is equal to 9.08 % of the total exergy destruction of the system. The second highest exergy destruction is found to be in the SHX-B, equivalent to 7.89 % of the total exergy destruction. Low exergetic efficiency of condenser evaporator assembly is attributed to the large temperature difference between the working fluid and NG in the evaporator and the working fluid and cooling air in the condenser. Similarly the high temperature difference between both the fluids and relatively very low mass flow rate of ammonia from the evaporator compared to the condensed ammonia from the condenser may be the reason for the poor exergetic efficiency of L/V HX. References Rad, H Feasibility Study of Hybrid Cooling Systems in Separation of High Molecular Weight Hydrocarbons in Natural Gas. Thesis Submitted to the University of Petroleum University of Technology In Energy Systems Engineering. Afshar, M. Rad, H Simulation and Exergy Analysis of Cascade Absorption Refrigeration System with Heat Recovery of Dew Point Units Waste Heat. 4 th Conference of Emerging Trends in Conservation Energy. Wei, H., Liuli, S., Danxing, Z., Hongguang, J., Sijun, M., and Xuye, J., New hybrid absorption compression refrigeration system based on cascade use of mid-temperature waste heat. International Journal of Applied Energy, Vol.106, pp Andre, M., Sergio, M., Luben, G., and Ricardo, S.,

60 2010. Using engine exhaust gas as energy source for an absorption refrigeration system. International Journal of Applied Energy, Vol.87, pp Kalinowski, P., Yunho, H., Reinhard, R., Hashimi, S., and Rodgers, P., Application of waste heat powered absorption refrigeration system to the LNG recovery process. International Journal of Refrigeration, Vol. 32, pp Manzela, A., Hanriot, S., Cabezas, G., and Sodre, J., Using engine exhaust gas as energy source for an absorption refrigeration system. International Journal of Applied Energy, Vol.87, pp Erickson, D., Anand, G., and Kyung, I., 2004, Heatactivated dual-function absorption cycle. International Journal of ASHRAE Trans, Vol.110, Part1. Fan, Y., Luo, L., and Souyri, B., Review of solar sorption refrigeration technologies: development and applications. Journal of Renew Sustain Energy Rev, Vol.11, pp Srikhirin, P., Aphornratana, S., and Chungpai, S., A review of absorption refrigeration technologies. Journal of Renew Sustain Energy Rev, Vol.5, pp Gas technology. Technical information about natural gas cleaning and treatment. See also URL Priscilla, B., Machado, J., Monterio, J., Medeiros, H., and Epsom, O., Supersonic separation in onshore natural gas dew point plant. International Journal of Natural Gas and Engineering, Vol.6, pp Pongsid, S., Satha, A., and Supachart, C., A review of absorption refrigeration technologies. International Journal of Renewable and Sustainable Energy Reviews, Vol.5, pp Jianbo, L., and Shiming X., The performance of absorption-compression hybrid refrigeration driven by waste heat and power from coach engine. International Journal of Applied Thermal Engineering, Vol.61, pp Brant, B., Brueske, S., Erickson, D., and Papar, R., Refinery Waste Heat Ammonia Absorption Refrigeration Plant. Journal of ESL-IE Pratihar, A., Kaushik, S., and Agarwal, R., Performance evaluation of a small capacity compression-absorption refrigeration system. International Journal of Applied Thermal Engineering, Vol. 42, pp Darwish, N., Hashimi, S., and Mansoori, A., A Performance Analysis and Evaluation of a 46 Commercial Absorption-Refrigeration Water- Ammonia (ARWA) System. International Journal of Refrigeration, Vol. 31, No. 7, pp Kececiler, A., Acar. H., and Dogan A, Thermodynamic analysis of absorption refrigeration system with geothermal energy: an experimental study. Journal of Energy Conversion and Management, Vol. 41, pp Lazzarin, R.M., Gasparella, A., and Longo, G.A., 1996, Ammonia-Water Absorption Machines for Refrigeration: Theoretical and Real Performances, International Journal of Refrigeration, Vol. 19, No. 4, pp Ruiz, E., Ferro, V.R., Riva, J., Moreno, D., and Palomar, J., 2014, Evaluation of ionic liquids as absorbents for ammonia absorption refrigeration cycles using COSMO-based process simulations. International Journal of Applied Energy, Vol. 123, pp Grossman, G., Zaltash, A., Modular Simulation of Absorption Systems. International Journal of Refrigeration, Vol. 24, No. 6, pp Liao, X., 2004, The Integration of Air-Cooled Absorption Chiller in CHP Systems. Ph.D. Thesis, University of Maryland, College Park, MD, USA. See also URL Darwish, N., Al-Hashimi, S., Al-Mansoori, A., Performance Analysis and Evaluation of a Commercial Absorption-Refrigeration Water- Ammonia (ARWA) System. International Journal of Refrigeration, Vol. 31, No. 7, pp Peng, D., Robinson, D., A New Two-Constant Equation of State. Industrial & Engineering Chemistry Fundamentals. International Journal of Chemistry Fundumental, Vol.15, pp Janilson, A.R., Edson, B., Thermodynamic Modeling of an Ammonia-Water Absorption System Associated with a Micro-turbine. International Journal of Thermodynamics, Vol. 12, pp Garousi, L.F., Infante F., Mahmoudi, S.M.S., and Rosen, M.A., First and second law analysis of ammonia/salt absorption refrigeration systems. International Journal of Refrigeration, vol. 40, pp Lostec, B.L., Galanis, N., and Millette, J., Simulation of an ammonia-water absorption chiller. International Journal of Renewable Energy, Vol.60, pp Aghniaea, S., Mahmoudi, S.M, Exergy Analysis of a Novel Absorption Refrigeration Cycle with Expander and Compressor. Indian Journal of Sci.Res, Vol.1, pp

61 Zare, V., Mahmoudi, S.M., Yari, M., and Amidpour, M., Thermoeconomic analysis and optimization of an ammonia-water power/cooling cogeneration cycle. International Journal of Energy, vol.47, pp Takeshita, K., Amano, Y., and Hashizume, T., Experimental study of advanced cogeneration system with ammonia-water mixture cycles at bottoming. International Journal of Energy, Vol. 30, pp

62 Energy and Exergy Analysis of a Steam Power Plant Considering Effect of Varying Plant Loads Mehmet Bilgili 1, Mehmet Tontu 2, Besir Sahin 2* 1 Cukurova University, Ceyhan Engineering Faculty, Mechanical Engineering Department, Adana, 01950, Turkey 2 Cukurova University, Engineering and Architecture Faculty, Mechanical Engineering Department, Adana, 01330, Turkey * Abstract In this study, the energy and exergy analyses of the existing steam power plant have been performed with three different operating loads (100%, 70% and 40%). This steam power plant is designed to operate at a subcritical steam conditions and has a power capacity of 660 MW at full load. The primary objectives of this study are to analyze the system components separately and to identify and quantify the sites having largest energy and exergy losses. Influences of three different loads on the useful power, reversible power and irreversibility have been investigated for each component. In addition, the second law efficiency of system components and the overall efficiency of the power plant have been computed. Energy losses mainly occur in the condenser and that exergy losses mainly take place in the boiler. According to these results, the exergy analysis is more significant compared to the energy analysis. It is found that if the exergy losses are reduced, the power plant efficiencies are positively affected. It is concluded that notable amount of input exergy can t be used as a useful work due to friction, mixing and irreversibility in the cycle. Keywords: Efficiency, energy analysis, exergy analysis, operating load I. Introduction Energy is one of the basic necessities in modern life. There is not any field of activity in everyday life that is the energy not being used. Nowadays, energy consumption of societies is considered as an indicator of development. Especially, electrical and heat energies play an important role in human lives. These energies which are produced from limited natural resources have become more valuable due to an increase of demand. Energy is considered to be a vitally important tool in economic, social and industrial developments (Yazıcı and Selbaş, 2011). Erkin (2006) reported that Turkey is one of the fastest growing energy markets in the world for nearly two decades because of a young and growing population, an increase of electricity consumption per person, rapid urbanization and strong economic growth. Unfortunately, the growth in electricity generation in recent years has been lower comparing to the growth of electricity demand. According to the report of IEA, (2013), the population growth, industrialization, expansion of technology and rising of welfare are proportional to the increase of energy consumption. In order to reach the level of developed countries, developing countries must involve in research and development and produce more advanced technological products. Although the development rate of renewable energy technologies increases rapidly, the world energy needs are still heavily depended on fossil fuels for electricity production. That is to say, the majority of the world s power generation is met by fossil fuels, particularly coal and natural gas. Despite the growth of renewable energy installations like wind and solar power, a heavy dependence on fossil fuels is expected to continue for decades as stated by Regulagadda et al. (2010). They also reported that despite the depletion of fossil fuel reserves and environmental concerns such as climate changes, the growth in the fossil oil demand is expected to be 47.5% (For example, 91.6% from the natural gas and 94.7% from the coal, until 2030). On the other hand, cleaner renewable energy sources are being rapidly developed. But, their relative cost and current state of renewable energy technologies have not advanced to a stage where they can significantly reduce our dependence on the fossil fuel. Environmental concern, climate changes, and continued reliance on the fossil fuel forces the fossil fuel plants to reduce their environmental impact by operating thermal power stations more efficiently and improving their technologies. In this study, the energy and exergy analyses of the existing steam power plant were performed with three different operating loads (100%, 70% and 40%). Influences of three different loads on the useful power, reversible power and irreversibility were investigated for each component. II. Power Plant Description Steam power plant considered in the present study is the coal fired plant located in Turkey. It operates with Rankine cycle supplying competitive power to the national grid under acceptable environmental protection. Power plant generally consists of pumps, 48

63 turbines, boiler, condensers and heaters also includes flue gas desulfurization unit and electrostatic filters. The present power plant has one HP turbine, one IP turbine and two LP turbines. HP turbine is a single flow type and has 13 stages. IP is double flow and has three uncontrolled extraction points and 14 stages. LP turbines are double flow and each LP turbines have three uncontrolled extraction points and 14 stages. All turbines directly coupled with the generator which rotates at 3000 rpm (50Hz). Gross outlet power is 660 MW and net power is 605 MW at a full load for each unit. In coal fired thermal power plant, steam is obtained in very high pressure inside the steam boiler by burning the pulverized coal. This steam is then super-heated in the super heater to extreme high temperature. This super-heated steam is then allowed to enter into the turbine, as the turbine blades are rotated by the pressure of the steam. After entering into the turbine, the steam pressure suddenly falls leading to corresponding increase in the steam volume. After having imparted energy into the turbine rotors, the steam is made to pass out of the turbine blades into the steam condenser of turbine. In the condenser, cold water at ambient temperature is circulated with the help of pump which leads to the condensation of the low pressure wet steam. Then this condensed water is further supplied to low pressure water heater where the low pressure steam increases the temperature of this feed water, it is again heated in high pressure. This outlines the basic working methodology of a thermal power plant. The water-steam cycle of power plant is shown in Fig. 1. III. Analysis III.1. Energy Analysis In a steady state control volume: m in = m out (1) The total energy content remains constant as the mass balance during steady state operation: E in = E out (2) Q W = m out h out m in h in (3) Here, subscripts in and out show inlet and outlet states, Q is the heat transfer rate, W is the work rate, m is the mass flow rate and h is the specific enthalpy. Thermal efficiency (η ı ), may also be defined as the 1st law efficiency, which can be expressed as the ratio of the work rate to the fuel energy input rate: η I = W net Q in (4) III.2. Exergy analysis Regulagadda et al. (2010) indicated that exergy analysis will characterize the work potential of a system. In the exergy analysis of this study, the properties at the dead state were denoted by subscript zero. For instance P0 and T0 refer to the dead-state pressure and temperature, respectively. Here, T0 was assumed to be 25 o C (298 K) and P0 was assumed to be 1 bar. The method of exergy analysis overcomes the limitations of the first law of thermodynamics. The concept of exergy is based on both the FLT and the SLT. Exergy analysis clearly indicates the locations of energy waste in a process and can therefore lead to improved operation or technology. Exergy analysis can also quantify the quality of heat in a waste stream. A main aim of exergy analysis is to identify exergy efficiencies and true magnitudes of exergy losses. Exergy analysis provides those tools and it helps in locating weak spots in the process. (Dinçer and Rosen, 2011). The exergy analysis for present system is carried out for each component in the subsystems, to evaluate the exergy losses in the individual component and then the analysis is performed on the overall individual subsystems to find out the exergy losses in each subsystem. Finally the exergy analysis for the overall plant is carried out and the second law efficiency is computed. Quantitative measure of disorder is called entropy of the system at the microscopic level. Entropy generation, Sgen of the system is presented as follows: Ql S ms ms gen (5) out in Exergy balance can be obtained by using following equations: X heat X X mass, out T T in 1 0 k m work X Ql out X loss W mass, in X m X T 0 out loss X 0 in (6) (7) Where Tk is the surrounding temperature and the change in the flow exergy is; out in h out h in T ( s 2 2 Vout V in (8) g( zout z in) 2 X m( in ) (9) out 0 out s in ) 49

64 Fig.1: The water-steam cycle of power plant 50

65 Reversible work is obtained with following equations. out V W 2 out rev 2 in V h 2 in out h g( z in out T ( s 0 z in ) out m( in ) out s in ) (10) (11) Irreversibility can be calculated as: I T0S gen or, I W rev W in (12) (13) Considering no kinetic and potential energy, the expression for exergy becomes; ( h h0 ) T0 ( s s0) (14) Where h, s are specific enthalpy and entropy values at a given state, h0, s0 are specific enthalpy and entropy values at a dead state. III.3. Fuel and Combustion Analysis Boilers need a heat source which has an enough temperature to generate steam with a high pressure and temperature using fossil fuel in the combustion chamber. It is known that the main constituent of coal is carbon. Coal also contains varying amounts of oxygen, hydrogen, nitrogen, sulfur, moisture, and ash as shown in Table.1. In the present thermal power plant, imported coal with high calorific value is used and coal components are indicated below. Lower calorific value of coal is defined as 25,800 kj/kg. (Unal, 2009). Pure oxygen is used only in special application but in this study air is used for combustion. Air is considered to be % 21 oxygen and % 79 nitrogen on a molar basis. With this idealization the molar ratio of the nitrogen to the oxygen is 0.79/0.21 = 3.76, when air supplies is accompanied by 3.76 moles of nitrogen (Moran and Shapiro, 1995). Tab. 1: Chemical compositions of presently used Coal Coal Components (% ratio) C H N S O Ash Water Combustion process in this study is: (cc+hh+nn+ss+oo+wh2o+)+1.5a(o2+3,76n2) (x CO2+yN2+zH2O+tSO2) Required minimum amount of oxygen for completing combustion is calculated using this equation 17 (Geredelioğlu, 2011); Amount of oxygen = %C %H %S %O2 (15) Enthalpy of flue gases can be calculated at desired temperature using these equations: h h fg xco. hco xn. h N xso. h SO xh O. h H O (16) fg h M fg fg Entropy of components of flue gases, s s CO N s s ln 2 K CO R x, 2 CO2 (17) (18) (19) s ln 2 K R x SO, N 2 N 2 (20) s ln 2 K R x, SO2 SO2 s s R x 2 H O 2 K, H2O ln H O (21) fg xco. sco xn. sn xso. sso xh O s H O (22) s s fg s fg (23) M fg Chemical exergy of flue gas and combustion air formula are represented below respectively. ch fg x RT ( x CO2. ln x CO2 x N2 x. N2 ln x x SO2 x. SO2 ln x H2O x. H2O 0 CO2 CO2 N2 N2 SO2 SO2 H2O H2O ln (24) ) ch ch fg fg (25) M fg X ch ch fg m (26) fg ch fg air x x. RT ( x ln x x ln x ) (27) O. 2 O2 N2 N2 0 O2 O2 N2 N2 ch ch air air (28) M air X ch air ch mair air (29) Coal chemical exergy, h o n (30) c c c w f h f (31) 4.18 X f m f f (32) 51

66 Standard chemical exergy of components is given Table 2 (Kotas,1995). Tab. 2: Standard chemical exergy of components Components ψ ch kj/kmol O CO N2 720 SO H2O 9500 IV. Results and Discussions Energy losses of thermal power plant equipment at different loads are shown in Fig. 2. As it can be seen from the figure energy loss of condenser is always higher than the other equipment. For example, 48.1%, 48.8% and 49.6% of total energy are rejected to the environment via condenser at 100%, 70% and 40% loads respectively. Because enthalpy of exhaust steam which comes from LP turbines is higher compared to the condense water. Reason of condenser losses is the latent heat of exhaust steam is transferred to the cooling water. Also 5.5 %, 6.3% and 6.6% of total energy rejected to the environment via boiler s outer surface and flue gases at 100%, 70% and 40% loads, respectively. Fig. 3 shows exergy losses of all thermal power plant equipment. As it is observed from the figure, exergy losses of the boiler is much higher than the other equipment. Despite the fact that major energy loss occurs in the condenser, but, major exergy loss is taken place in the boiler. That is to say, 51.9 %, 53.6% and 56.1% of input exergy are destroyed through exhaust gases at 100%, 70% and 40% loads respectively. The high level of exergy loss in the boiler is due to finite differences between flue gas and working fluid (water and steam). A combustion process usually occurs simultaneously with heat transfer. Both chemical reaction and heat transfer are irreversible processes. The losses in the boiler are due to increase in the entropy generation of flue gases. Additionally substantial amount of heat loss is conveyed by the flue gas to the environment with high temperature. The exergy losses take place in pumps due to compression ratio and flow control method. Pump losses are usually very small compared with the other equipment. Exergy loss in the condenser is due to heat transfer between exhaust steam and cooling water but it s thermodynamically insignificant because temperature differences between them is about 5 K. So that the rate of exergy loss is small. Fig. 2: Energy losses of power plant equipment Fig. 3: Exergy losses of thermal power plant equipment 52

67 Fig. 4: Irreversibilities of turbines Fig. 5: Irreversibility of heaters Fig. 4 shows irreversibility of turbines at three different loads. In turbines, 5.34 %, 5.43% and 5.55% of total exergy are destructed at loading capacities of 100%, 70% and 40% respectively. The exergy losses in turbines are due to the pressure drop and expansion process. Fluid friction happens when the fluid expands through the steam turbine blades. These friction losses result in the dissipation of part of its energy into heat itself at the expense of useful work. The fluid then does less work and leaves through the exhausts with a higher temperature. The more irreversible process, the higher turbine exit temperature and the less the work. As the load increases, amount of irreversibility increases in turbines but percentage of irreversibility decreases. The exergy loss in LP turbines is higher than IP and HP turbines because LP turbines work at a vacuum pressure so that condensation process begins and steam leaves the LP turbines as a water-steam mixture therefore the rate of entropy and irreversibility increases substantially with condensation process. But the exergy loss in HP turbines is lower, because pressure ratio between inlet and outlet is small compared to the IP and LP turbines. The value of exergy losses in the LP 1 and LP 2 turbines are close to each other, because the operating conditions are almost the same. Irreversibility in the HP and LP heaters at different load are shown in Fig. 5. It can be seen from this figure that irreversibility in the HP heaters is always higher than LP heaters due to the extraction steam of HP heaters at a high pressure and high temperature. Besides temperature differences between extraction steam and working fluid is higher and also as the load increases, although irreversibility in the HP heaters raises, irreversibility in the LP heaters drops down because while the load increases, temperature differences between extraction steam and working fluid diminishes in LP heaters but in HP heaters temperature differences between them are increases. In addition aim of using HP and LP heaters is to recover latent heat of exhaust steam just before conveying to the condenser. The second law efficiency which is belonged to power plant equipment is shown in Fig. 6. At three different loads, boiler feed pump is very efficient but boiler is least efficient. And also according to the figure the second law efficiency of equipment enhances with increasing load. Besides boiler feed pump efficiency is higher compared to the condenser extraction pump because flow rate of boiler feed pump is controlled by frequency converter but flow rate of condenser extraction pump is controlled by a valve (throttling). Pump exergy loss is related with the 53

68 compression ratios which are 21.4, 21.6 and 22.2 in the boiler feed pump and 595.2, and 913,4 in the condenser extraction pump at different loading capacities such as100%, 70% and 40% respectively. Overall efficiency of thermal power plant is shown in Fig. 7. According to energy analysis based on the first law analysis the major energy losses are due to heat rejection in the condenser and heat losses with flue gases but according to exergy analysis based on the second law analysis the major losses are due to steam generation where the fuel exergy is destroyed. It is seen that the first and second law efficiencies enhance with increasing load capacity. Overall power plant efficiencies are affected by its component efficiencies and its raises with increasing load. Although the amount of energy loss and exergy loss of power plant is increased, percentage of these losses is decreased while load is raised so that overall efficiencies are increased. As a consequence, the power plant is not suitable for partial load, it is suitable for full load. Net output powers of thermal power plant are measured 605 MW, 420 MW and 235 MW at a loading capacity of 100%, 70% and 40% respectively. Second law efficiency BFP Turbines CEP Boiler Fig. 6: The second law efficiency of equipments 100% 70% 40% First Law Second Law Efficiency V. Conclusions Plant Load % 40 Fig. 7: Overall efficiencies of power plant thermal power plant. In this study, energy and exergy analyses of the active thermal power plant were performed using thermodynamic principles for three different loads. Energy and exergy losses and also second law efficiencies of each equipment of the thermal power plant were calculated. From these analyses, the vulnerable spots can be seen in power production processes clearly. In actual coal fired thermal power plant operation, abnormal exergy parameters of the corresponding equipment s can be detected and used to define the faulty locations. So, the exergy efficiency and energy loss analysis of the power plant are helpful to malfunctions identification and diagnosis of From the present results, the following conclusions have been drawn: In the present coal fired thermal power plant, the maximum energy loss was found in the condenser at three different loads, but, exergy losses are lower because these losses thermodynamically are insignificant which are at low quality. Heat losses can be used in heating greenhouses if it is possible to set it near the thermal power plant. In terms of the exergy loss or irreversibility, maximum losses were found in the boiler at three 54

69 different loads. Because flue gas temperature is very high according to the ambient temperature. These losses may be used in the absorption cooling system or it may be used for coal drying purposes. Also, the losses may be used for heating sites. But it is considered that in the case of flue gas temperature which is lower than 120 C, acid formation can occur with condensation of flue gas, this situation cause defect the channel and any other equipment. As the load increases, irreversibility or the exergy loss increases at each equipment but percentage of irreversibility decreases. Irreversibility in pumps is directly proportional to the compression ratio. And the type of flow control method can also affect the irreversibility. As the load increases, compression ratio falls down but second law efficiency increases. Irreversibility in the HP heaters is always higher than LP heaters at three different operating loads. Irreversibility in heaters is directly proportional to the temperature differences between heating medium and working fluid. If the differences between them are increased, heat transfer rate is enhanced as a result, entropy and irreversibility rise. The causes of exergy losses in turbines are mainly related to their design, frictional losses and pressure ratio between inlet and outlet. Also condensation process should be considered in the last blades of LP turbines. It may cause wear and tear in the last blades of LP turbines because of water droplets and hence the entropy and irreversibility increase rapidly. The first law efficiency of thermal power plant was found to be 41.5%, 39.7% and 36.4% % at the loading capacities of 100%, 70% and 40% respectively. The second law efficiency of thermal power plant was found to be 39.1%, 37.4% and 34.3% at loading capacities of 100%, 70% and 40% respectively. Finally, it can be concluded that the overall efficiency of the thermal power plant is best at a full operating load. analysis of a thermal power plant with measured boiler and turbine losses, Applied Thermal Engineering, 30: (2010). Yazıcı H., Selbaş R., Energy and exergy analysis of steam power plant, Selçuk University, Journal of Technical Online, 10(1): (2011). Dinçer, I., Rosen, M.A., Thermal energy storage; systems and applications, John Wiley and Sons, 599p (2011). Geredelioğlu, Ç., Energy and exergy analysis of Çayırhan thermal power plant: Msc Thesis, Süleyman Demirel University, The Institute of Science, Isparta, 122p (2011). Unal, F., Exergy analysis of a thermal power plant: Msc Thesis, Yıldız Teknik University, The Institute of Science, İstanbul, 91p. (2009) References Erkin T., Energy and exergy analysis of thermal power plant, MSc Thesis, Denizli: Pamukkale University (2006). IEA, Energy Efficiency Policy Recommendations, (Date of Access: ). Kotas. T.J., The exergy method of thermal plant analysis, Krieger Publishing Company, Malabar, FL (1995). Moran, M.J., Shapiro, H.N., Fundamental of Engineering thermodynamics, Third Edition. John Wiley & Sons, New York (2000). Regulagadda P., Dincer, I., Naterer G.F., Exergy 55

70 Long Term Energy Demand and Supply Projections and Evaluations for Turkey Esra Ozdemir 1, Muhsin Kilic 2* 1 Uludag University, Engineering Faculty, Department of Mechanical Engineering, Bursa, Turkey 2 Uludag University, Yenisehir Ibrahim Orhan of Vocational School, Department of Machine, Bursa, Turkey * Abstract Energy demand of countries changes depending on many socia-economic factors like their population, the level of social and economic development, industrialization, urbanization and technological development. In the past decade, economic growth and social development in our country has been led to increase in energy consumption. Accordingly, the amount of energy needed must be provided that so as to realize the economic growth and social development in time, satisfactory, uninterrupted and taking into account the environmental impact. For this reason, it is necessary to determine and prioritize the alternative energy strategies. In this study, used energy datas from Turkey s Energy-Balance Tables and developed a long-term energy demand projections for Turkey in Long Range Energy Alternatives Planning System Program. Two scenarios for Turkey's energy demand are created and their results are evaluated. In addition, scenarios are created that the renewable energy resources' ratios which are solar, wind, hydro and geothermal energy etc. is increased in total energy supply and their results are evaluated for energy demand and for electricity generation. Keywords: Energy, energy demand, energy supply, environmantal effects, long term projection. I. Introduction Energy has become one of the most significant factors in ensuring that countries provide a competitive advantage since the beginning of the 20th century. In the 21st century, the technological innovations, increasing the permeability of international borders, capital mobility and development of communication cause increasing the amount and speed of energy use by Kavak (2005). For this reason, generating policies for the future and energy management should be done by planning energy from today. The Republic of Turkey with population 78 million and area km 2, forms a natural bridge between Europe and Asia by TUIK (2016). Turkey is a rapidly growing economy and over the past decade, its Gross Domestic Product (GDP) has increased at an significant rate compared to other OECD countries. Turkey is the 17th largest economy of the world by IEA (2001). Turkey has experienced considerable changes in its electricity market in the past decades. Rapid growth in the electricity demand has led to considerable transformation in the electricity sector with large increases in the generation capacity to accompany it. According to the energy balance sheets of Turkey, which are published by World Energy Council-Turkish National Committee (2016), between the years 2000 and 2013, the electricity demand of Turkey almost doubled and reached GWh. Energy has a strategic importance for developing Turkey. According to Turkey Statistical Institute's datas (2016), the energy imports of Turkey increased to 60.1 billion dolars in However this value decreased in the last two years and that was billion dolars and 54.91billion dolars in 2013 and 2014 respectively. Accounting for these values, it is considered that Turkey doesn t have substantial reserves of conventional fuels. Hence, Turkey should make energy planning and energy policy. Due to the lack and poor quality of primary resources, Turkey is highly dependent on imported energy. According to the Ministry of Energy, import dependency was above 72% in This is underpinned by the dependency on natural gas imports which account for nearly 43% of total electricity production by Republic of Turkey Ministry of Energy (2016). There are several challenges that the Turkish energy market faces. The high level of dependence on imported energy sources and the negative externalities caused by the utilization of fossil fuels stand as the main problems the policy makers will strive to solve for the immediate future. The primary aim of Turkey is to realize its own energy security. To this end, Turkey has for objective to -diversify its energy supply routes and source countries, -increase the share of renewables and include the nuclear in its energy mix, -take significant steps to increase energy efficiency, -contribute to Europe s energy security by Republic of Turkey Ministry of Foreign Affairs (2016). As a potential candidate country for Europe Union accession, Turkey is under a pressure to reduce its CO2 emissions. Thus Turkey should focus on building up its capacity for mitigation of greenhouse gas 56

71 (GHG) emissions and adaptation to climate change. One of the common approaches in the world for mitigation of GHGs is the sector-based emission mitigation policy. Accordingly, Turkey has a national policy on increasing share of renewable energy sources in the electricity generation by Ozer at al. (2013). Energy demand and supply projections working which lead to the energy planning and energy policy, has gained importance in recent years. Long term energy supply and demand projections constitute the basis of long term energy planning and investment. A combination of an optimal system to meet the energy demand can be determined in the lowest cost with energy projections. Thus, it is possible that both energy and financial resources are used more efficient and possible scenarios can be tested and revised. In the light of this working, decision-makers are able to have an idea about needed policies and practises for future from today. Some energy projections in the literature are as follows; Zhang et al. (2007) calculated the external costs of electricity generation in China under different energy scenarios by using Long Range Energy Alternatives Planning (LEAP) system. They estimated the energy demand in electricity generation of China from 2003 to Lin and Ouyang (2014) evaluated how demand of fossil fuels and carbon dioxide and sulphur dioxide in China are affected by the reforms in the country and macro-economic developments with general equilibrium models they created. Cai et al. (2013) rated between by creating different scenarios of electricity production models for transition to clean and more efficient use of energy in China by using LEAP software. Ozer et al. (2012) predicted electricity demand until 2030 in Turkey by using the LEAP program and have planned electricity production improvement scenarios for reduction of greenhouse gas emissions. Hotunoglu and Karakaya (2011) evaluated the final results by scenarizing how energy demand will develop in case of economic growth is stabilized, energy densities decrease in the future and economic growth changes in each five years by using artificial neural Networks technique. Dilaver and Hunt (2011) estimated Turkish aggregate electricity demand depending on GDP, electricity price and Underlying Energy Demand Trend (UEDT). Shin et al. (2005) have made planning electricity production by creating projection of being used waste gas (landfill gas) to produce electricity more environmentally friendly and economically in Korea by using LEAP model, within the context of the Kyoto Protocol. Egelioglu et al. (2001) investigated the influence of economic variables on the annual electricity consumption in Northern Cyprus and they found that a model using number of customers, number of tourists and electricity prices has a strong predictive ability. Song et al. (2007) accomplished environmental and economic assessents of chemical absorption processes in Korea using the LEAP model. They analyzed the scenario based on the data of a pilot plant (2 ton/day) that is installed in the Seoul coal steam power plant, and an alternative scenario 57 is set according to energy policy chang by climate change egreement and development of CO2 mitigation technology. Ozturk et al. (2005) used the genetic algorithm approach to investigate the relationship between total electricity consumption, gross national product (GNP), population, imports and exports for the period in Turkey with annual data. Total electricity demand of Turkeywas estimated as 220 TWh and 300 TWhin 2020 with exponential and quadratic forms of the genetic algorithm electricity demand models respectively. However, Ozturk and Ceylan (2005) concluded that aggregate electricity demand for Turkey would be between about 462 TWh and 500 TWh in 2020 for the low and high growth scenario respectively as the results of genetic algorithm electricity demand (GAED) quadratic model. The goal of this paper is to evaluate the future energy demand, energy supply and CO2 emission potential of Turkey s energy sector. For this purpose, firstly, the total energy demand was estimated depending on the six sectors: industrial, residential and services, transport, agriculture and non energy use. The estimation was based on the population, gross domestic product (GDP) and the proportion of each demand sector in total consumption with the annual growth rates. Correspondingly electricity generation scenarios were built. In this paper, two scenarios for energy demand and four scenarios for electricity generation based on Long-range Energy Alternatives Planning system (LEAP) model were employed to simulate the current energy situation and to develop forecasts under certain assumptions. The demand scenarios include Business As Usual (BAU) and Mitigation Scenario options. The electricity generation scenarios are created that the renewable energy resources' ratios which are solar, wind, hydro and geothermal energy etc. is increased in total energy supply and their results are evaluated for energy demand and for electricity generation. II. Energy Consumption and Supply in Turkey II.1. The structure of the energy sector in Turkey Importance of Turkey increases as a regional energy transit hub and growing consumer in the energy market. According to Energy Information Administration (2013), Turkey's energy demand has increased in recent years in very quickly and it is predicted that this increase will continue in the next years. According to Energy Report (2013), overall distribution of energy resources in Turkey is that, Turkey has become one of the fastest growing energy markets in the world and has been experiencing rapid demand growth in all segments of the energy sector for decades. Turkey comes in possession of the most dynamic energy economies of the world in terms of increase in energy demand.

72 Having a substantial potential for the renewable energy resources, Turkey ranks seventh in the world and first in Europe in terms of geothermal energy. Turkey aims at further increasing its use of hydro, wind and solar energy resources and Turkey has potential producing %30 of its electricity need from the renewable by Turkey is geographically located in close proximity to more than %70 of the world s oil and gas reserves Annual electricity generation is approximately 179, 5 billion kwh in Turkey. Renewable energy and energy efficiency projects are assisting to reduce CO2 emissions in Turkey by more than 3 million tons annually. Turkey has different kinds of energy sources which Turkish energy sector is becoming more active, competitive and attracting the attention of investors. According to Energy Balance Sheets of Turkey (EBST), thousand tons of oil equilavent (TOE) primary energy supply occurred with thousand TOE domestic production and thousand TOE the value of imported energy in Fig. 1 shows the distribution of sources in total primary energy supply. The highest energy resource was natural gas with the rate of 32%. This value was followed by coal and oil sources with 29% rates. According to Fig.1, Turkey's energy supply comprises of fossil fuels by 90%. Wood 2% Hydro 4% Oil 29% Geo., wind, solar 2% Others 2% Coal 29% Naturel gas 32% Fig. 1: The distribution of sources in Turkey's primary energy supply On a sectoral basis, thousand TOE of the energy-supply were consumed by the sector of cycle and energy, thousand TOE were consumed by the sectors of industrial, residential and services, transport, agriculture and non energy use. The highest energy demand took place in the industrial sector and the residential and service sector (Tab. 1). Generally, the industrial sector in the energy consumption have increased since 2002, although in 2008 and 2009 the production decreased due to the global economic crisis in Turkey. However, the value of energy consumption of industrial sector has continued to rise since 2010 (WEC-TNC). 58 Tab. 1: Sectoral energy consumption of Turkey in 2013 (EBST 2013) Sectors Energy consumption Rate (Thousand TOE) (%) Industry 30, Transportation 22, Housing (Residence) 31, and Services Agriculture Non-Energy Usage Cycle and Energy 30, II.2. The structure of the electricity sector in Turkey According to Energy Balance Sheets of Turkey (EBST), electricity generation and consumption increased more than threefold since The gross electricity demand of Turkey increased by 6.8% annually from GWh (Gigawatt-hours) in 1990 to GWh in The total installed capacity of the power industry is approximately 49.5 GW (Gigawatt) at the end of 2013, while it was 16.3 GW in The annual growth rate is about 5.8% (Fig. 2) Electricity Energy Generation (GWh) Installed Capacity (MW) Fig. 2: Turkey's installed capacity and electricity energy generation Production and consumption in 2013, compared to 2012, increased by 0.3% and 1.6% respectively. As seen in Tab. 2., generation and consumption increased by 0.3% and 1.6% respectively. When we look at the past five years, the change in consumption and peak demand was the average annual level of 5.6%. Compared with the previous five years, the decline of increase rate is realized significantly. However, growth in the installed capacity has continued and reached MW with an increase 12%. Electricity Generation Company (EGC) with its subsidiaries in the production has a share by 34%. As of 2013, the total share of public sector in the market was 60% and the share of free market production was 40%. The weight of natural gas continues in electricity generation. Consumption of natural gas increased significantly from 1990 to 2013 while the share of natural gas in electricity generation increased from

73 about 18% in 1990 to 43.8% in Hydro, lignite and imported coal power plants had 25%, 13% and 12% respectively. Tab. 2: General electricity generation and consumption of Turkey (EBST 2013) Unit change (%) change (%) Installed MW capacity Peak Demand MW Generation GWh İmportation GWh Exportation GWh Consumption GWh According to Energy Balance Sheets of Turkey (EBST), electricity consumption per capita increased more than threefold since 1990, as shown in Fig. 3. Net electricity consumption per capita was 2577 kwh/person and 2568 kwh/person in 2012 and Gross electricity consumption per capita was 3205 kwh/person and 3132 kwh/person in 2012 and 2013 respectively. LEAP based on a comprehensive accounting such as, energy production, conversion and consumption in a particular region or economy under conditions of alternative assumptions based on population, economic development, technology, price and so on. With the program, possible future problems are identified; perspective is created for the energy supply and demand in the future years by evaluating possible effects of energy policy and it allows energy planning and policy from today. In addition, an environmental assessment of greenhouse gas emissions arising from energy use can be done. LEAP contains technology and environment database. It allows extensive information indicated the impacts of the environment, cost, technical characteristics of energy technology and also it allows to do projections of energy supply and demand for long-term planning. Fig. 4: The structure of LEAP (Song et al. 2007). Fig. 3: Electricity consumption per capita by the years When the final electricity demand by sector of Turkey for 2013 is analyzed, Turkey s energy end-use is dominated by the industrial sector, which takes up about 45% of total end-use electricity consumption while it was 62.4% in 1990 and has an average annual growth rate of about 5%. The residential sector accounts for about 25%, the commercial and services sectors follow by a cumulative 27% in III. Methodology This study uses an accounting and scenario-based modeling platform called LEAP to assess the impact of energy consumption and energy supply processes in Turkey. Long Range Energy Alternatives Planning (LEAP) system is an energy environment modeling tool developed at Stockholm Environment Institute (SEI), Boston, to assess the effects physical, economic and environmental of alternative energy programs, technologies and other energy initiatives by Song et al. (2007). 59 The structure of LEAP is presented in Fig. 4. In this approach, the LEAP software tool is used to analyze the current energy patterns and to simulate alternative energy futures along with environmental emissions under a range of user-defined assumptions. LEAP consists of four modules: energy scenarios, aggregation, environmental data base and fuel chain. Each module makes it possible to analyze extending impact by technology and policy change through a description of natural resources structured as energy sector, conversion course, final energy form and final energy demand. Moreover, it is possible to analyze for resource, conversion, demand and environmental emission through scenario analysis by (Song et al. 2007). Gross domestic product (GDP), which expresses the income level of countries is an entry data in the creation of projections. In LEAP model, the GDP PPP and GDP MER datas for the value of GDP which are gotten from the World Bank, were used. Turkey Statistical Institute population value for population, which is another factor in the incresing of energy demand is used in LEAP model. For this data,

74 scenarios was created using the growth rate of population in the newsletter which is "the Demographic Structure and Future of Turkey, " published by TUIK. All of the datas were gotten form Energy Balance Sheets of Turkey (EBST), and for which were entered for years in LEAP model. Fig. 6: The fuel ratios of enery demand According to the BAU scenario, the energy consumption of housing and services sector will be million GJ at the end of 2023, while it was million GJ in Compared with the 2011 value, the total growth rate is about 45% in The energy consumption of industry sector will reach million GJ at the end of 2023, while it was million GJ in Compared with the 2011 value, the total growth rate is about 40.5% in In the agricultural sector, energy demand will be million GJ with a 117% increase at the end of In the transportation sector, energy demand will be million GJ with a 28.5% increase at the end of In the non energy use sector, energy demand wil be million GJ with a 103% increase at the end of 2023 (Tab. 3). Fig. 5: Energy system diagram of LEAP model IV. Results and discussions IV.1. Energy Demand Scenarios IV.1.1. Business As Usual (BAU) Scenario The Business As Usual represents the energy pathway that is implied if current energy policies, supply and demand trends in Turkey persist. This includes basically economic growth and energy conversion. Current trends in the Turkish economy and the power sector continue in the BAU Scenario by Ozer et al. (2013). For the growth rate of Turkey's population and the growth rate of GDP MER, 1.25% and 2.9% are used in scenario. Accordingly, it is estimated that Turkey's population will reached to 86.7 million people and GDP MER will reached to billion $ in Scenarios were developed in based the historical development of fuels from 1985 until 2012 for energy demand sectors. According to the scenario, Turkey's total energy demand in 2023 will be million GJ (Gigajoule). Electricty, natural gas and oil energy demand will reached to million GJ, million GJ and million GJ respectively in 2023 (Fig. 6). Tab. 3: Sectoral energy consumption of Turkey in next years Sectors (MGJ) Housing (Residence) and Services 1254, , , , ,93 Agriculture 240,97 292,76 355,68 432,13 525,00 Industry 1348, , , , ,61 Transportation 667,80 708,98 754,07 803,58 858,12 Non-Energy Usage 185,99 222,14 265,32 316,89 378,49 Total 3698, , , , ,15 According to Tab.4, it was predicted that the electricity demand of housing and services sector will reached to TWh with about 54% growth, the electricity demand of industry sector will reached to TWh with about 49.5% growth in The electricity demand of agriculture sector and transportation sector will reached to TWh and 0.42 TWh respectively at the and of Tab. 4: Sectoral electricity consumption of Turkey in next years Sectors (TWh) Housing (Residence) and Services 88,07 97,06 107,35 120,75 136,17 Agriculture 4,36 6,10 8,51 11,86 16,49 Industry 93,90 104,26 114,52 126,56 140,24 Transportation 0,53 0,34 0,37 0,39 0,42 Total 186,86 207,76 230,75 259,56 293,32 60

75 In Turkey's energy demand, the most destructive gas is CO2 emissions. According to BAU scenario,as a result of energy consumption, it is estimated that greenhouse gas affect will reach to million metric tons of CO2 at the end of Tab. 5: Greenhouse gas affect IV.1.2. Mitigation Scenario Economic and demographic situation are jumping the shark in the mitigation scenario. It is assumed that the develepment of Turkey's economy slowed towards Hence, the growth rate of Turkey's population has been 1%. The growth rate of GDP MER has been as "Growth (2,9%; 2015; 2,5%; 2018; 1,5%; 2020; 0,5%; 2023; -0,5%)" in projection. The energy demand of housing and services sector is decreased 1% annually, the energy demand of industry sector remains constant and the enerdy demand of non-energy use sector is increased by 0.5% growth. Accordingly, it is estimated that Turkey's population will reached to 84.2 million people and GDP MER will reached to billion $ in According to the mitigation scenario, the energy consumption of housing and services sector will be million GJ at the end of 2023, while it was million GJ in Compared with the 2011 value, the total decrease rate is about 4.5% in The energy consumption of industry sector will be million GJ at the end of 2023, while it was million GJ in Compared with the 2011 value, the total decrease rate is about 0.3% in In the agricultural sector, energy demand wil be million GJ with a 29.6% increase at the end of In the transportation sector, energy demand wil be million GJ with a 11.3% increase at the end of In the non energy use sector, energy demand wil be million GJ with a 29.6% increase at the end of 2023 (Tab. 6). According to Tab.7, it is estimated that the electricity demand of housing and services sector will reached to TWh, the electricity demand of industry sector will reached to TWh in The electricity demand of agriculture sector and transportation sector will reached to 9.81 TWh and 0.37 TWh respectively at the end of Tab. 7: Sectoral electricity consumption of Turkey in next years Sectors (TWh) Housing (Residence) and Services 88,07 87,10 86,45 87,26 88,31 Agriculture 4,36 5,55 6,97 8,40 9,81 Industry 93,90 95,69 96,47 97,85 99,52 Transportation 0,53 0,34 0,36 0,37 0,37 Total 186,86 188,68 190,25 193,88 198,01 According to mitigation scenario, it is estimated that greenhouse gas affect will reach to million metric tons of CO2 at the end of Tab. 8: Greenhouse gas affect IV.2. Electricity Generation Scenarios IV.2.1. Hydro Scenario In this scenario, it is aimed that the installed capacity of hydro-based electricity generation will be MW in Accordingly, it can be seen from the Fig.7 that the electricity generation with hydro-power growth will be TWh at the end of 2023, while it was TWh in Tab. 6: Sectoral energy consumption of Turkey in next years Sectors (MGJ) Housing (Residence) and Services 1254, , , , ,55 Agriculture 240,97 266,51 291,33 306,18 312,34 Industry 1348, , , , ,42 Transportation 667,80 708,98 745,31 754,75 743,42 Non-Energy Usage 185,99 205,70 224,87 236,32 241,08 Total 3698, , , , ,31 61 Fig. 7: The fuel ratios of hydro scenario IV.2.2. Nuclear Scenario In addition to the increase of hydro power generation capacity, in the nuclear scenario, it is aimed that Akkuyu Nuclear Power Plant with 4800 MW capacity will be activated in 2019 and Sinop Nuclear Power

76 Plant with 1200 MW capacity will be activated in For these data, step function was used in modelling. was used in modelling. Results were shown at the Fig.10. According to nuclear scenario as seen in Fig.8, it is predicted that electricity generation will reach to TWh and the genaration from nuclear energy will be TWh. Fig. 10: The fuel ratios of total scenario Fig. 8: The fuel ratios of nuclear scenario According to total scenario, it is estimated as given in the Tab.8 that electricity generation will reach to TWh. The genaration from hydro, geothermal, wind, natural gas, solar and nuclear energy will be ; 3.68; 87.6; 12.67; 9.2; and TWh, respectively. Tab. 8: A comparison of the value of electricity generation scenarios Scenario (TWh) Hydro 240, , , ,625 Geo and wind 246, , , ,243 Nuclear 240, , , ,465 Total 247, , , ,340 Fig. 9: The fuel ratios of geo and wind scenario IV.2.3. Geo and Wind Scenario In addition to the increase of hydro power generation capacity, in the geo and wind scenario, it is targetted that the value of geothermal energy in electricity generation in 2016 and 2023 will reached to 300 MW and 600 MW respectively. In addition that, the value of wind energy in electricity generation in 2016 and 2023 will reach to MW and MW respectively. According to the scenarios, the highest energy generation values were obtained with the total scenario which combined hydro, geothermal, wind, solar, natural gas and nuclear energy. When looking at the environmental emission values, the most environmentally friendly scenario was expected to take place with total scenario as shown in the Fig.11. According to this scenario, it is calculated as shown in the Fig.9 that electricity generation will reach to TWh. The generation from geothermal energy will be 3.68 TWh and the generation from wind energy will be 87.6 TWh. IV.2.4. Total Scenario Total scenario consists of increase of hydro-power, nuclear, geo and wind generation capacity. Also, it is aimed that the installed capacity of solar energy will be 3000 MW and natural gas installed capacity will be MW in For these data, step function 62 Fig. 11: The greenhouse effect of scenarios V. Conclusions Energy supply and demand forecast from today is significant for investors and energy planners to

77 evaluate energy policies and environmental policies. The importance of energy demand and electricity generation is dealt with in this study. The energy demand and electricity generation of Turkey were estimated according to population and GDP increasing rate up to Affect the energy consumption are analyzed. The modelling for demand of energy and electricity generation has been done in Long Range Energy Alternatives Planning system program. Two energy demand scenarios and three electricity generation scenarios were built up for energy consumption, power technology and environmental policies. In the scenarios generated, the results of energy demand were obtained primarily by making BAU scenario in which everything continues in conditions of today and mitigation scenario. In the electricity generation scenarios, hydro, nuclear, geo and wind and total, energy consumption and emission values were calculated by creating possible future plans with various policies. In this line of study, generating policies for the future and energy management will be possible by planning energy today. References Başkan, Ö., Haldenbilen, S., Ceylan, H,. The modeling of energy demand in transportation sector and sustainable policies (in Turkish). World Energy Council-Turkish National Committee, 10. Energy Congress of Turkey (2005). Cai, L., Guo, J., Zhu, L., China's future power structure analysis based on LEAP. Energy Sources, Part A: Recovery, Utilization and Environmental Effects, 35:22, (2013). Canyurt, O.E., Öztürk, H.K., Hepbaşlı, A., Utlu, Z., Genetic Algorithm (GA) approaches for the transport energy demand estimation: model development and application. Energy Sources, Part A: Recovery, Utilization and Environmental Effects, 28: (2006). Dilaver Z., Hunt L.C., Turkish aggregate electricity demand. An outlook to 2020, Energy, 36: (2011). Egelioglu F., Mohamad A.A., Güven H. Economic variables and electricity consumption in Northern Cyprus. Energy, 26: (2001). EIA (Energy Information Administration), Online at: U, [accessed: ] IEA (International Energy Agency), Online at: [accessed: ] Kavak K., Energy Efficiency in the world and in Turkey and the analysis of energy efficiency in the Turkish industrial, Dissertation Thesis, State Planning Organisation (in Turkish), publication number: SPO:2689, Ankara (2005). 63 Lin B., Ouyang X., A revisit of fossil-fuel subsidies in China: challenges and opportunities for energy price reform, Energy Conversion and Management, 82, (2014). MEANR, Republic of Turkey Ministry of Energy and Natural Resources, Ankara. Online at:http://www.enerji.gov.tr/en- US/Mainpage, [accessed: ] Ozturk H.K., Ceylan H., Canyurt O.E., Hepbaslı A., Electricity estimation using genetic algorithm approach: a case study of Turkey, Energy, 30: (2005). Ozturk H.K., Ceylan H., Forecasting total and industrial sector electricity demand based on genetic algorithm approach: Turkey case study, International Journal of Energy Research, 29: (2005). SEI. LEAP long-range energy alternatives planning system; user guide for leap version Online at: 08UserGuideEnglish.pdf; [accessed: ]. Shin H., Park J., Kim H. ve Shin E., Environmental and economic assessment of landfill gas electricity generation in Korea using LEAP model, Energy Policy, 33, (2005). Song H.J., Lee S., Maken S., Ahn S.W., Park J.W., Min B., Koh W., Enviranmental and economic assessment of the hemical absorption process in Korea using the LEAP model, Energy Policy, 35: (2007). Zhang Q., Weili T., Yumei W., Yingxu C., External costs from electricity generation of China up to 2030 in energy and abatement scenarios, Energy Policy, 35: (2007). TUIK (Turkish Statistical Institute), TUIK Datas, Ankara. Online at: [accessed: ] WB (The World Bank), Datas and statistics for Turkey. Online at: [accessed: ] WEC-TNC, The Energy Balance Sheets of Turkey, Ankara. Online at: [accessed: ] WEC-TNC (World Energy Council-Turkish National Committee), Energy Report 2013, Ankara. Online at: ]

78 Evaluating Exergetic Sustainability Indicators for an Electrolyte Supported SOFC Stack Adnan Midilli*, Ugur Akbulut Department of Mechanical Engineering, Faculty of Engineering, Recep Tayyip Erdogan University, Rize, Turkey * Abstract In terms of exergy, this paper introduces and evaluates exergetic sustainability indicators for an Yttria Stabilized Zirconia (YSZ) electrolyte supported SOFC stack. For this purpose, the perfectly insulated SOFC stack is considered, having 300 μm-thickness-ysz electrolyte, 50 μm-thickness Ni-YSZ anode, 50 μm-thickness LSM-YSZ cathode, and consuming fuel as 97% H2+3% H2O and oxygen of the air, and operating at temperatures ranging from 1073 to 1273 K. In order to introduce exergetic sustainability indicators of such a stack, the following parameters are taken into account, which are total exergy input, desired exergy output, exergy destruction, exergy outputs by unused hydrogen and air, and exergy output by water. For better understanding and evaluating the critical points of the exergetic sustainability indicators as a function of current density, the critical exergetic efficiency is assumed to be 0.30 Accordingly, for making an evaluation of such a SOFC stack in terms of exergy based environmental and sustainability assessment, the following indicators should be taken into consideration, which are exergetic efficiency, waste exergy ratio, exergy destruction ratio, environmental effect factor and exergetic sustainability index. Under the selected operating conditions, it is noticed that waste exergy ratio increases with the rise of temperature. However, exergy destruction ratio and environmental effect factor decrease with temperature elevation at a constant current density while increasing with the rise of current density at a constant temperature. Moreover, exergetic sustainability index decreases with the increment of current density at a constant temperature while increasing with the rise of temperature at a constant current density. Considering the critical value of exergetic efficiency (=0.30), the critical environmental effect factor and the critical exergetic sustainability index are respectively determined to be and while the critical current density is A/cm 2 for 1073 K, A/cm 2 for 1173 K and A/cm 2 for 1273 K. Thus, it is suggested that, in order to exergetically operate the YSZ electrolyte supported SOFC stack under the selected design and operating conditions in accordance with the required electricity generation and fuel consumption, i) environmental effect factor should not be higher than 2.320, ii) exergetic sustainability index should not be lower than 0.428, iii) current density should be selected lower than A/cm 2 at 1073 K, A/cm 2 at 1173 K and A/cm 2 at 1273 K. Keywords: YSZ electrolyte supported SOFC, exergetic efficiency, waste exergy ratio, exergy destruction ratio, environmental effect factor, exergetic sustainability index I. Introduction Solid oxide fuel cells, commonly manufactured to be tubular and planar SOFCs, has recently attracted considerable interest for domestic and industrial applications. The most important features of these cells can be compiled as higher energy efficiency, lower pollutant emissions, fuel flexibility, and high operating temperature which allows a variety of cogeneration possibilities (Singhal, 2002; Khaleel et al., 2004; Xue et al., 2005; Ni et al., 2007; Hussain et al., 2009; Ahn et al., 2009; Xu et al., 2014). In terms of the energy conversion management, it is generally said that there are three types of the SOFCs which are anode supported SOFCs, cathode supported SOFCs and electrolyte supported SOFCs. In electrolyte-supported SOFCs, the thickness of YSZ electrolyte is generally between 150 ~ 300 μm, which operates during 800~1000 C (Singhal and Kendall, 2004; Han et al., 2010). In anode-supported SOFCs, the thickness of YSZ electrolyte is generally 15~30μm, which operates during 600~800 C (Singhal and Kendall, 2004; Han et al., 2010). In an electrolyte-supported SOFCs, the ohmic effect of the electrolyte layer should be also taken into account for energetic or exergetic performance improvement because electrolyte is the thickest component. Moreover, exergy based environmental and sustainability assessments of these fuel cells should be investigated. In this regard, it can be said that in order to ensure the exergetic sustainability, a clean and abundant fuel such as hydrogen should be used in the SOFCs. Of the most promising energy carriers for the future, hydrogen is more versatile, an energy-efficient, low polluting, environmentally benign fuel that will meet most of our energy needs in near future (Dincer, 2002). If so, it can be said that use of renewable hydrogen in the SOFCs can help reduce environmental effect of the SOFCs and achieve exergetic sustainability. Under these important considerations, in order to introduce and evaluate the exergy based environmental and sustainability assessment of an YSZ electrolyte supported SOFC stack, the following exergetic sustainability indicators, which are derived based on the operating principle of the YSZ electrolyte supported SOFC stack, should be taken into account; i) exergetic efficiency, ii) waste exergy ratio, iii) exergy recoverability ratio, iv) exergy 64

79 destruction ratio, v) environmental effect factor, vi) exergetic sustainability. Moreover, some effective parameters are considered, which are i) operating temperature (ranging from 1073 K to 1273 K), ii) operating pressure (=1 atm), iii) anode and cathode thicknesses (= 50 μm each) and iv) current density (ranging from 0 to 3 A/cm 2 ), v) electrolyte thickness (ranging from 100 to 500 μm). For this purpose, a detailed literature review has been carried out by considering the evaluation of exergetic sustainability indicators for an YSZ electrolyte supported SOFC stack. However, it is noticed that no studies are conducted on this subject. Actually, this lack of information indicating the originality of this paper is the motivation behind this work. However, some works have been found on exergetic sustainability indicators (e.g. Midilli and Dincer, 2009; 2010; Midilli et al., 2010; Kucuk and Midilli, 2015; Ozsaban and Midilli, 2016). Considering such important facts, as the scientific and industrial benefits, this study, which includes all details on exergy-based sustainability of a high pressure hydrogen production and storage system, aims to contribute to find out new exergetic dimensions and the exergy based-environmental and sustainability aspects of an YSZ electrolyte supported SOFC stack. II. Main Considerations In order to introduce the exergetic sustainability indicators of an YSZ electrolyte supported SOFC stack, the following required assumptions and parameters have been taken into consideration. II.1. Assumptions The following assumptions have been considered for this evaluation: An electrolyte supported SOFC stack is considered. Hydrogen is used as a fuel and only hydrogen is electrochemically reacted. Fuel consists of 97% H2 and 3% H2O and air as oxidant consists of 79% N2, 21% O2 (Costamagna et al., 2004; Ni et al., 2007; 2009; Yonekura et al., 2011; Ranjbar et al., 2014). The chemical exergy of nitrogen is not taken into account because it is inert gas and not a function of temperature. The fuel cell is insulated perfectly, so there is no heat interaction with environment. Temperatures at channel inlets and exits are the same (Colpan et al., 2007; Ranjbar et al., 2014). Radiation heat transfer between gas channels and solid structures is neglected (Ranjbar et al., 2014). Contact resistances are ignored. Radiation transfer between solid structure and gas channels is ignored. The electrodes and electrolyte materials of the SOFC are taken to be Ni-YSZ/YSZ/LSM-YSZ (Chan et al., 2002; Costamagna et al., 2004; Singhal and Kendall, 2004; Colpan et al., 2007; Han et al., 2010; Zheng et al., 2014). SOFC stack operates under steady-state conditions (Hussain et al., 2006; Colpan et al., 2007). Kinetic and potential exergies are neglected. SOFC operating pressure is taken to be 1 atm (Costamagna et al., 2004, Colpan et al., 2007; Mirahmadi et al., 2011; Verma et al., 2013). Dead state pressure is 1 atm and dead state temperature is 298 K. SOFC operating temperatures are taken between 1073 and 1273 K, and membrane thickness is selected between 100 and 500 μm considering the values in the literature (Singhal and Kendall, 2004; Han et al., 2010). Each reactant in SOFC is an ideal gas (Chan et al., 2002; Larminie and Dicks, 2003; Hussain et al., 2006; Colpan et al., 2007; Tanim et al., 2014). Flow of reactants is steady, incompressible and laminar. Pressure drops along the fuel cell are neglected (Hussain et al., 2006; Colpan et al., 2007; Tanim et al., 2014). The product water is in vapor phase. Current density is taken from 0 to 3 A/cm 2. The utilization ratios of hydrogen and oxygen are taken to be 80% and 50%, respectively (Ishak et al., 2012; Tanim et al., 2014) Active area of a single cell is selected to be 100 cm 2 and a stack is composed of 100 cells. All activation, ohmic, and concentration polarizations are considered. All values for hydrogen, oxygen, nitrogen and water vapor are taken from the NIST (Website) Tab. 1: The parameters taken for the calculations Parameter Value Ref. Operating temperature K (Singhal and Kendall, 2004; (T) Han et al., 2010). Operating pressure (P) 1 atm (Mirahmadi and Valefi, 2013) Current density (J) 0-3 A/cm 2 Assumed. Faraday constant (F) A.s/mol (Verma et al.,2013) Number of electrons per 2 (Costamagna et al., 2004) mol (ne) Anode thickness (La) 50 μm (Ni et al., 2007; Verma et al, 2013) Electrolyte thickness μm (Singhal and Kendall, 2004; (Le) Han et al., 2010). Cathode thickness (Lc) 50 μm (Ni et al., 2007) Porosity (ε ) 30% Tortuosity (ξ) 6 (Ishak et al., 2012) Pore radius (r p ) 0.5 μm Diffusion volume of 6.12x10-6 m 3 /mol hydrogen (v H2 ) Diffusion volume of 13.1x10-6 m 3 /mol (Zheng et al., 2014) water (v H2O ) Diffusion volume of 16.3x10-6 m 3 /mol oxygen (v O2 ) Diffusion volume of 18.5x10-6 m 3 /mol nitrogen (v N2 ) Molar fraction of %97 hydrogen (Yonekura et al., 2011) Molar fraction of water %3 Molar fraction of oxygen %21 II.2. Main Calculations In order to perform this work, the required equations for the specific calculations related to the electrolyte supported SOFC stack are taken from the literature, which are not presented in this study. 65

80 1. The values of ohmic, activation and concentration overpotentials, total cell overpotential (irreversible cell voltage) and net cell voltage can be calculated from the literature (Colpan et al. 2008; Kazempoor et al., 2010; Costamagna et al, 2004; Tanim et al., 2014). 2. Assuming that a single cell has 100 cm 2 active area and a stack is composed of 100 cells, total active area of the SOFC stack can be calculated from the literature (Trendewicz and Braun, 2013). 3. In order to estimate the molar flow rate of hydrogen, the fuel utilization factor (UF) can be calculated from the literature (Campanari, 2001; Ishak et al., 2012). 4. The molar flow rates of hydrogen consumed in the SOFC, and the molar flow rates of water output, oxygen and nitrogen at the air channel inlet and exit can be calculated from the literature (Colpan et al., 2007). 5. The heat generated from the SOFC stack can be calculated by considering the general energy balance equation for a SOFC stack (Chan et al., 2001). II.3. System Description The control volume of an YSZ electrolyte supported SOFC stack can be illustrated as below. H assumptions), v) E x 2 O out includes chemical and W physical exergies of water vapor, vi) E x d,out includes exergy of electricity generated by the stack. Fig. 2: Exergy balance diagram of an YSZ electrolyte supported SOFC stack II.4. Exergy Balance of the Stack Considering Fig. 2, the general exergy balance equation can be written as below in terms of the Second Law of Thermodynamics. Fig. 1: Control volume of an YSZ electrolyte supported SOFC stack In this figure, fuel is hydrogen; desired output is electricity generated by the stack; air includes mostly oxygen and nitrogen; and H2O is in vapor phase while crossing the boundary of control volume; heat loss will be equal to zero because it is assumed that the stack is perfectly insulated. However, the heat loss should be taken into account if it occurs in case of practical applications. Considering the general control volume of an YSZ electrolyte supported SOFC stack, general exergy balance diagram is shown in Fig. 2. In Fig. 2, i) E x in fuel includes chemical and physical fuel exergies of hydrogen gas, ii) E x uu includes chemical and physical exergies of unused hydrogen air gas, iii) E x in includes chemical and physical air exergies of air (see assumptions), iv) E x uu includes chemical and physical exergies of unused air (see 66 E x in - E x out = E x loss (1) In this equation, total exergy output ( E x out ) includes desired output (E x W d,out ) from the stack while total exergy loss ( E x loss ) can sometimes include waste exergy and exergy destruction based on the operating principle of the system (Midilli and Dincer, 2009). In this regard, Eq. (1) can be written as E x in - E x d,out = E x w + E x D (2) In Eq. (2), E x w can consist of recoverable waste exergy ( E x rw ) and unrecoverable waste exergy (E x urw ) (Midilli and Dincer, 2009; 2010). If so, Eq. (2) can be written as E x in - E x d,out = E x rw + E x urw + E x D (3) Considering Eq. (3) and assumptions, general exergy balance of the stack can be derived as ch E x in,h2 ch (Ėx uu,h2 ph + E x in,h2 ph + E x uu,h2 ch + E x in,o2 ph E x out,h2o ) + (E x uu,o2 ph + E x in,o2 ph + E x in,n2 W = ph ph + E x uu,o2 + E x out,n2 + E x out,h2o ch ) + E x D (4) ch + Under these considerations, exergetic sustainability parameters depending on the exergy parameters in Eq. (4) can be derived as below. III. Exergetic Sustainability Parameters How to derive the exergetic sustainability indicators

81 for a system or process? In order to derive the exergetic sustainability indicators for a system or a process, the steps summarized below should be basically followed, i) detailed operating principle is determined, ii) detailed control volume including all inputs and outputs is drawn, iii) operating parameters and assumptions are determined, iv) mass balance equation is written, v) energy analysis is achieved in terms of the First Law of thermodynamics, vi) exergy analysis is achieved in terms of the Second Law of thermodynamics, vii) Exergy efficiency is defined as the ratio of exergy of desired output to total exergy input, viii) Waste exergy ratio is defined as the ratio of total waste exergy output to total exergy input. Here, total waste exergy output should contain all exergy outputs from the system or process to reference environment, which should not include exergy destruction, ix) Exergetic recoverability ratio is defined as the ratio of the recoverable exergy from total waste exergies to total exergy input. Here, recoverable exergy from total waste exergy output should cover the exergies that are possible to be reused for the same system or same process, and/or for any system or any process, x) Exergy destruction ratio is defined as the ratio of exergy destruction in the system or process to total exergy input. Here, exergy destruction is determined through the exergy analysis of the system or process, xi) Environmental effect factor is defined as the ratio of total waste exergy output to exergetic efficiency, xii) Exergetic sustainability index is defined as the reverse of environmental effect factor. More details on derivation, conceptual and physical meanings and mathematical formulations of these indicators were presented in the literature (Midilli and Dincer, 2009; 2010; Midilli et al., 2012). Therefore, mathematical derivation procedure of these parameters will not be presented in this study. However, main definitions of these parameters will be introduced here as below. i) Exergetic efficiency (ee) Total exergy of useful output ee (5a) Total exergy input In this regard, exergetic efficiency of the SOFC is mainly based desired output exergy, total exergies of hydrogen and air. E x (ee) = W = E x in,h2 +E x in,air W E x ch ph in,h2 +E x in,h2 +E x ch ph in,o2 +E x in,o2 +E x in,n2 ph (5b) where E x W indicates exergy of SOFC stack power ch ph ch ph output while E x in,h2, E x in,h2, E x in,o2, E x in,o2 and ph E x in,n2 represents chemical and physical exergy of hydrogen, oxygen and nitrogen gas entering the SOFC stack, respectively. ii) Waste exergy ratio (wer) Total waste exergy output wer (6a) Total exergy input Waste exergy consists of physical and chemical exergies of unused hydrogen, unused oxygen, nitrogen, water vapor leaving the SOFC is taken into consideration. In this regard, waste exergy ratio (ranging from 0 to 1) can be written as wer = ch ph ch ph ph ch ph E x uu,h2 +E x uu,h2 +E x uu,o2 +E x uu,o2 +E x out,n2 +E x out,h2o +E x out,h2o E x ch ph in,h2 +E x in,h2 +E x ch ph ph in,o2 +E x in,o2 +E x out,n2 (6b) ph ch where E x out,h2o and E x out,h2o indicates physical and chemical exergies of water vapor, respectively. iii) Exergy recoverability ratio (err) Total re coverable exergy err (7a) Total exergy input Exergy recoverability ratio (ranging from 0 to 1) indicates the exergy potential that is possible to be recovered in the system. If so, it includes i) physical exergies of unused hydrogen, unused oxygen, nitrogen, water vapor, and ii) the chemical exergies of unused hydrogen and water vapor leaving the SOFC. In this regard, exergy recoverability ratio can be written (err) = E ch x uu,h2 (7b) ch +Ėx uu,h2o E x ch in,h2 ph ph ph ph +E x uu,o +E x 2 out,n2 +E x uu,h2 +E x out,h2o ph +E x in,h2 +E x ch ph ph in,o2 +E x in,o2 +E x out,n2 ch ch where E x uu,h2 and E x uu,o2, indicates chemical exergy of unused hydrogen and oxygen output; ph ph ph E x uu,h2, E x uu,o2, E x out,h2o, represents the physical exergy of unused hydrogen, oxygen and water outputs respectively. iv) Exergy destruction ratio (edr) Total exergy destruction edr (8a) Total exergy input Exergy destruction ratio (ranging from 0 to 1) is a function of exergy destruction and total exergy input. (edr) = E x d E x ch ph in,h2 +E x in,h2 +E x ch ph in,o2 +E x in,o2 +E x out,n2 v) Environmental effect factor (eef) eef ph (8b) Waste exergy ratio Exergy destruction ratio (9a) Exergetic efficiency Environmental effect factor (ranging from 0 to + ) indicates whether or not SOFC has some damage potential on the environment due to waste exergy output and exergy destruction. 67

82 (eef) = wer+edr = ch E x uu,h2 ee ph ch ph ph ch ph +E x uu,h2 +E x uu,o2 +E x uu,o2 +E x out,n2 +E x out,h2o +E x out,h2o +E x d W (9b) vi) Exergetic sustainability index (esi) Exergy of useful output esi (10a) Waste exergy output Exergy destruction Exergetic sustainability index (ranging from 0 to + ), which is defined as a function of environmental effect factor, reveals exergy-based sustainability of system or process in terms of the second-law of thermodynamics. (esi) = 1 eef = E x ch uu,h2 (10b) W ph ch ph ph ch ph +E x uu,h2 +E x uu,o2 +E x uu,o2 +E x out,n2 +E x out,h2o +E x out,h2o +E x d IV. Result and Discussion In this study, the main aim is to introduce and evaluate the exergetic sustainability indicators for an Yttria Stabilized Zirconia (YSZ) electrolyte supported SOFC stack. In this regard, the following investigations have been performed; i) variation of exergetic efficiency as a function of current density under various operating temperatures (see Fig. 3), ii) variation of environmental effect factor as a function of current density under various operating temperatures (see Fig. 4), iii) variation of exergetic sustainability index as a function of current density under various operating temperatures (see Fig. 5). Figure 3 introduces the variation of exergetic efficiency as a function of current density under various operating temperatures which are taken to be 1073, 1173 and 1273 K. It is assumed that anode, cathode and electrolyte thicknesses are taken to be 50, 50 and 300 μm, respectively while operating pressure is equal to 1 atm. As noticed from this figure, exergetic efficiency (ranging from 0 to 0.497) decreases with the rise of current density from 0.1 to 3 A/cm 2. At a constant current density, for example, for 0.3 A/cm 2, exergetic efficiency goes up from to with the increase of operating temperature from 1073 to 1273 K. Accordingly, in order to increase the exergetic efficiency of the YSZ electrolyte supported SOFC stack, lower current densities are targeted under the given design and operating conditions. The exergetic efficiency can be commercially assumed to be higher than at least 0.3. Considering this assumption and the critical point of current densities (0.271 A/cm 2 at 1073 K, 0.61 A/cm 2 at 1173 K, and 1.17 A/cm 2 at 1273 K) resulting from the operating temperatures, the graph can be separated in two region which are efficient operating region (exergetic efficiency is higher than 0.3) and inefficient operating region (exergetic efficiency is lower than 0.3). On the other word, it can be said that, considering the assumed reference line, commercially available operating region is upper side of the reference line. 68 Figure 3. Variation of exergetic efficiency as a function of current density under various operating temperatures. Figure 4 display the variation of environmental effect factor as a function of current density under various operating temperatures. It is assumed that anode, cathode and electrolyte thicknesses are taken to be 50, 50 and 300 μm, respectively while operating pressure is equal to 1 atm. Moreover, operating temperatures are taken to be 1073, 1173 and 1273 K. Environmental effect factor (eef) La=50 μm, Lc=50 μm, Le=300 μm, P=1 atm T=1073 K T=1173 K T=1273 K Current density (A/cm 2 ) efficient operating region critical environmental effect factor line Figure 4. Variation of environmental effect factor as a function of current density under various operating temperatures. As shown in this figure, environmental effect factor (ranging from to ) goes up with the rise of current density from 0.1 to 2.2 A/cm 2. At a constant current density, for example, for 0.3 A/cm 2, rising operating temperature (from 1073 to 1273 K) decreases environmental effect factor (from to 1.201). Considering the reference line of exergetic efficiency (as in Fig. 3) and the critical points of the current densities at the selected operating temperatures, the critical values of environmental effect factor are determined to be at 1073 K, at 1173 K, at 1273 K in case La=Lc=50 μm and Le=300 μm. Thus, under the selected design and operating conditions, in order to minimize exergy based environmental effect of the YSZ electrolyte supported SOFC stack, it should be operated at the lower values than almost of environmental effect factor.

83 Figure 5 displays the variation of exergetic sustainability index as a function of current density under various operating temperatures. It is assumed that anode, cathode and electrolyte thicknesses are taken to be 50, 50 and 300 μm, respectively while operating pressure is equal to 1 atm. Moreover, operating temperatures are taken to be 1073, 1173 and 1273 K. As shown in this figure, exergetic sustainability index (ranging from 0.04 to 0.988) decreases with the increase of current density (from 0.1 to 2.2 A/cm 2 ). At a constant current density, for example, for 0.3 A/cm 2, exergetic sustainability index increases (from to 0.833) with the rise of operating temperature from 1073 to 1273 K. In order to increase exergetic sustainability index the current density should be decreased while operating temperature increases. Considering the reference line of exergetic efficiency (as in Fig. 3) and the critical points of the current densities at the selected operating temperatures, the critical values of exergetic sustainability index are estimated to be at 1073 K, at 1173 K, at 1273 K in case La=Lc=50 μm and Le=300 μm. Thus, under the selected design and operating conditions, in order to maximize exergetic sustainability index of the YSZ electrolyte supported SOFC stack, it should be operated at the higher values than almost of the exergetic sustainability index. Exergetic sustainability index (esi) La=50 μm, Lc=50 μm, Le=300 μm, P=1 atm Efficient operating region Current density (A/cm 2 ) citical exergetic sustainability index line T=1073 K T=1173 K T=1273 K Figure 5. Variation of exergetic sustainability index as a function of current density under various operating temperatures. V. Conclusion In this paper, exergetic sustainability indicators were introduced and evaluated for an Yttria Stabilized Zirconia (YSZ) electrolyte supported SOFC stack. In this regard, the following concluding remarks can be drawn: Exergetic sustainability index decreases with the rise of current density at a constant temperature while going up with the increase of operating temperature at a constant current density. Considering the critical value of exergetic efficiency (=0.30), the critical current densities are determined to be A/cm 2 for 1073 K, A/cm 2 for 1173 K and A/cm 2 for 1273 K. 69 Considering the reference line of exergetic efficiency, the critical values of environmental effect factor are determined to be at 1073 K, at 1173 K, at 1273 K in case La=Lc=50 μm and Le=300 μm. Considering the reference line of exergetic efficiency, the critical values of exergetic sustainability index are estimated to be at 1073 K, at 1173 K, at 1273 K in case La=Lc=50 μm and Le=300 μm. Consequently, it can be said that, under the selected design and operating conditions, in order to minimize environmental effect factor of the YSZ electrolyte supported SOFC stack, it should be operated at the lower values than of the critical environmental effect factor. Moreover, under the selected design and operating conditions, in order to maximize exergetic sustainability index of the YSZ electrolyte supported SOFC stack, it should be operated at the higher values than of the critical exergetic sustainability index. Thus, it is suggested that, in order to exergetically operate the YSZ electrolyte supported SOFC stack under the selected design and operating conditions in accordance with the required electricity generation and fuel consumption, i) environmental effect factor should not be higher than 2.320, ii) exergetic sustainability index should not be lower than 0.428, iii) current density should be selected lower than A/cm 2 at 1073 K, A/cm 2 at 1173 K and A/cm 2 at 1273 K. Nomenclature Symbols E x : Exergy rate (kw) W : Power (kw) Subscripts D : Destruction d,out : desired output in : Input w : waste References Ahn J.S., Pergolesi D., Camaratta M.A., Yoon H., Lee B.W., Lee K.T., Jung D.W., Traversa E., Wachsman E.D., High-performance bilayered electrolyte intermediate temperature solid oxide fuel cells, Electrochemistry Communications 11, (2009). Campanari S., Thermodynamic model and parametric analysis of a tubular SOFC module, Journal of Power Sources, 92, (2001). Chan S.H., Khor K.A., Xia Z.T., A complete polarization model of a solid oxide fuel cell and its sensitivity to the change of cell component thickness, Journal of Power Sources, 93, (2001). Chan S.H., Low C.F., Ding O.L., Energy and exergy analysis of simple solid-oxide fuel cell, power

84 systems, Journal of Power Sources, 103, (2002). Colpan C.O., Dincer I., Hamdullahpur F., Thermodynamic modeling of direct internal reforming solid oxide fuel cells operating with syngas, International Journal of Hydrogen Energy, 32, (2007). Colpan C.O., Dincer I., Hamdullahpur F., A review on macro-level modeling of planar solid oxide fuel cells. International Journal of Energy Research, 32, (2008). Costamagna P., Selimovic A., Borghi M.D., Agnew G., Electrochemical model of the integrated planar solid oxide fuel cell (IP-SOFC), Chemical Engineering Journal, 102, (2004). Dincer I., Technical, environmental and exergetic aspects of hydrogen energy systems, International Journal of Hydrogen Energy, 27, (2002). Han M.F., Liu Z., Zheng Z., Liu M., High performance solid oxide fuel cells based on tri-layer yttria-stabilized zirconia by low temperature sintering process, Journal of Power Sources, 195, , (2010). Hussain M.M, Li X., Dincer I., Mathematical modeling of planar solid oxide fuel cells, Journal of Power Sources, 161, , (2006). Hussain M.M, Li X., Dincer I., A general electrolyte electrode-assembly model for the performance characteristics of planar anode-supported solid oxide fuel cells, Journal of Power Sources, 189, (2009). Ishak F., Dincer I., Zamfirescu C., Energy and exergy analyses of direct ammonia solid oxide fuel cell integrated with gas turbine power cycle, Journal of Power Sources, 212, (2012). Kazempoor P., Dorer V., Ommi F., Modelling and Performance Evaluation of Solid Oxide Fuel Cell for Building Integrated Co- and Polygeneration, Fuel Cells, 10(6), (2010). Khaleel M. A., Lin Z., Singh P., Surdoval W., Collin D., A finite element analysis modeling tool for solid oxide fuel cell development: coupled electrochemistry, thermal and flow analysis in MARC, Journal of Power Sources, 130, (2004). Kucuk H., Midilli A., Assessment of exergetic sustainability indicators for a single layer solar drying system, International Journal of Exergy, 16(3), (2015). Larminie J.E., Dicks A., Fuel Cell Systems Explained, John Wiley and Sons, West Sussex (2003). Midilli A., Dincer I., Development of some exergetic parameters for PEM fuel cells for measuring 70 environmental impact and sustainability, International Journal of Hydrogen Energy, 34(9), (2009). Midilli A., Dincer I., Effects of some micro-level exergetic parameters of a PEMFC on the environment and sustainability, International Journal of global Warming, 2(1), (2010). Midilli A., Kucuk H., Dincer I., Environmental and sustainability aspects of a recirculating aquaculture system, Environ Prog Sustain, 31, (2012). Mirahmadi A., Valefi K., Study of thermal effects on the performance of micro-tubular solid-oxide fuel cells, Ionics, 17, (2011). Ni M., Leung M.K.H., Leung D.J.C., Parametric study of solid oxide fuel cell performance, Energy Conversion and Management, 48, (2007). Ni M., Leung D.J.C., Leung M.K.H., Electrochemical modeling and parametric study of methane fed solid oxide fuel cells Energy Conversion and Management, 50, (2009). Ozsaban M., Midilli A., A parametric study on exergetic sustainability aspects of high-pressure hydrogen gas compression, International Journal of Hydrogen Energy, 41(11), (2016). Ranjbar F., Chitsaz A., Mahmoudi S.M.S., Khalilarya S., Rosen M., Energy and exergy assessments of a novel trigeneration system based on a solid oxide fuel cell, Energy Conversion and Management, 87, (2014). Singhal S. C., Solid oxide fuel cells for stationary, mobile, and military applications, Solid State Ionics, , (2002). Singhal S.C., Kendall K., High Temperature Solid Oxide Fuel Cells: Fundamentals, Design and Applications, Elsevier Ltd, Oxford, (2004). Tanim T., Bayless D.J., Trembley J. P., Modeling a 5 kwe planar solid oxide fuel cell based system operating on JP-8 fuel and a comparison with tubular cell based system for auxiliary and mobile power applications, Journal of Power Sources, 245, (2014). Trendewicz A.A., Braun R.J., Techno-economic analysis of solid oxide fuel cell-based combined heat and power systems for biogas utilization at wastewater treatment facilities, Journal of Power Sources, 233, (2013). Verma J.K., Verma A., Ghoshal A.K., Performance analysis of solid oxide fuel cell using reformed fuel, International Journal of Hydrogen Energy, 38, (2013)

85 Website, accessed at March 20, Xue X., Tang J., Sammes N., Du Y., Dynamic modeling of single tubular SOFC combining heat/mass transfer and electrochemical reaction effects, Journal of Power Sources, 142, (2005). Xu H., Dang Z., Bai B.F., Electrochemical performance study of solid oxide fuel cell using lattice Boltzmann method, Energy, 67, (2014). Yonekura T., Tachikawa Y., Yoshizumi T., Shiratori Y., Ito K., Sasaki K., Exchange current density of solid oxide fuel cell electrodes, ECS Transactions, 35(1), (2011). Zheng K., Li L., Ni M., Investigation of the electrochemical active thickness of solid oxide fuel cell anode, International Journal of Hydrogen Energy, 39, (2014). 71

86 Life Cycle Assessment of Nuclear Based Ammonia Production Options: A Comparative Study Yusuf Bicer*, Ibrahim Dincer Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, Ontario L1H 7K4, Canada * Abstract In this study, nuclear power based ammonia production options ranging from thermochemical cycles to high temperature electrolysis are comparatively evaluated using life cycle assessment (LCA) tool. Ammonia is produced by extracting nitrogen from air and hydrogen from water and combining them with the help of nuclear energy. Since production of ammonia contributes about 1% of global greenhouse gas (GHG) emissions, new methods with less environmental impact are under close investigation. Nuclear, as a sustainable energy source compared to conventional fossil fuels, emerges as an alternative option for ammonia synthesis. Within the current study, the selected ammonia production systems are (i) three step nuclear Cu-Cl thermochemical cycle, (ii) four step nuclear Cu-Cl thermochemical cycle, (iii) five step nuclear Cu-Cl thermochemical cycle, (iv) nuclear power based electrolysis integrated to Haber-Bosch process and (v) nuclear high temperature electrolysis integrated to Haber-Bosch process. Electrolysis units for hydrogen production and a Haber-Bosch process for ammonia synthesis are utilized for the electrolysis based options while hydrogen is produced thermochemically using excess heat in nuclear power plants for thermochemical based systems. The waste heat in nuclear power plant is also utilized for high temperature electrolysis in order to decrease the amount of required power for electrolysis process. Using the LCA methodology, the environmental impacts of selected ammonia synthesis methods are comparatively identified and quantified from cradle to gate. The life cycle assessment includes fuel elements, chemicals, and diesel requirements as well as the relevant transport requirements. Cryogenic air separation is mostly used method for massive amount of nitrogen production. In the life cycle assessment of nitrogen production, electricity for process, cooling water, waste heat and infrastructure for air separation plant are included. The LCA results for the selected ammonia production methods show that nuclear electrolysis based ammonia production method yields lower global warming and climate change impacts while thermochemical based options yield higher abiotic depletion and acidification values. Keywords: Ammonia production, nuclear, life cycle assessment, environmental impact. I. Introduction Ammonia is potentially treated as a significant hydrogen carrier with a much higher hydrogen content. In recent years, expectations are rising for hydrogen and hydrogen carriers as a medium for storage and transportation of energy and use of renewable energy. Transportation and storage issues of hydrogen are important as hydrogen is in gas form at ambient temperature and pressure. Ammonia is one of the major synthesized industrial chemicals in the world. Ammonia synthesis consumes almost 1.2% of total primary energy and contributes about 1% of global GHGs emissions (Gilbert et. al., 2010). Approximately 1.5 tonnes of CO2 is released to the environment during the production of 1 tonne of ammonia with the current technology (Anderson et. al., 2008). Natural gas is the primary feedstock used for producing ammonia worldwide via steam methane reforming. The delivery and storage infrastructure of ammonia is similar to liquefied petroleum gas (LPG) process. Under medium pressures, both of the substances are in liquid form which brings significant benefit because of storage options. Today, vehicles running with propane are mostly accepted and used by the public since their on-board storage is possible and it is a good example for ammonia fueled vehicle opportunities since the storage and risk characteristics of both substances are similar to each other. Zamfirescu and Dincer (2009) examined the use of ammonia as a clean fuel in evaluation with further conventional fuels. They defined the possible technical benefits of ammonia usage as a sustainable fuel aimed at power production on vehicles based on some efficiency indicators containing the system efficiency, the driving distance, fuel tank compactness and the price of driving. Verma and Kumar (2015) offered a model to evaluate life cycle GHG emissions in hydrogen production from underground coal gasification with and without carbon capturing. Utilization of carbon capturing technology permits a substantial reduction in total life cycle emissions in hydrogen production from underground coal gasification. Kalinci et al. (2012) performed a life cycle assessment of hydrogen production from CFBG/DG biomass production in 72

87 order to use the generated hydrogen in PEM fuel cell vehicles by investigating the costs of GHG emissions reduction. The extreme energy consumption rates were observed in the compression and transportation of hydrogen steps for the CFBG based system. Zamfirescu and Dincer (2008) reported a few possible occasions and benefits of using ammonia as a sustainable fuel in transportation vehicles. They have compared ammonia with other conventional fuels in different aspects. Moreover, using ammonia both as a refrigerant and a fuel, they calculated refrigeration effect with respect to refrigeration power vs engine s power ending up with that ammonia is the cheapest fuel on $/GJ basis. Cryogenic air separation is usually used method for massive amount of nitrogen production which is used in this study. In the life cycle assessment of nitrogen production, electricity for process, cooling water, waste heat and infrastructure for air separation plant are taken into account. The allocation factors were obtained from the heat of vaporization and the specific heat capacity multiplied with the temperature difference from 20 C to the boiling point. The utilized software, SimaPro, has the values of nitrogen production from cryogenic air separation process in the database. Makhlouf et al. (2015) presented the results of a life cycle assessment of 1 tonne of ammonia produced in Algeria considering anhydrous liquid ammonia. They specified that Algerian ammonia plant consumes more energy than world average. Reformer processes are the main reasons of overconsumption of energy and GHG emissions. This was because of the low effectiveness of the catalytic reaction in which the catalysts were used more than 10 years. A few of the available ammonia utilization pathways can be listed as follows (Dincer et. al., 2011): Direct feed of ammonia into an internal combustion engine Ammonia thermal cracking and feed of the products (H2 and N2) all together in the internal combustion engine cylinder for combustion Separation of N2 and H2 streams simultaneously with the decomposition such that only pure H2 is combusted; and the nitrogen is expanded for work production Direct ammonia high-temperature fuel cell systems, Ammonia thermal cracking and separation and further using the hydrogen into high temperature fuel cells Ammonia electrolysis and hydrogen used in proton exchange fuel-cells with additional exploitation of ammonia s refrigeration effect Fig. 1. Selected nuclear based ammonia production options In nuclear-based high-temperature ammonia production, the system consists of a nuclear power plant, high temperature electrolyzer, cryogenic air separation unit and a Haber-Bosch synthesis plant as shown in Fig. 2. The required electricity is utilized from nuclear power plant and the required heat for high temperature electrolysis is supplied from nuclear waste heat. Nuclear power plant electricity is assumed as a mixture of 66.5% from pressure water reactor (PWR) and 33.5% from boiling water reactor (BWR) type reactors since SimaPro software database does not include CANDU type reactors. Note that ammonia is also a suitable fuel for spark-ignition engines because of its high opposition to auto ignition. On the other hand, it is of great interest to utilize ammonia in compression-ignition engines due to the popularity of compression ignition engine-driven electricity generators. For internal combustion engines, service network is already available and ready in addition to mature manufacturing technology. II. Systems description In the present study, five different ammonia production methods are selected for comparative assessment purposes as illustrated in Fig. 1 where Haber-Bosch process is utilized for ammonia synthesis. 73 Fig. 2. Nuclear high temperature electrolysis and Haber-Bosch process for ammonia production

88 Nuclear based electricity yields lower cost and reliable supply. Combining nuclear power plant with ammonia production plant is an encouraging method. In high temperature electrolysis, the excess heat in the nuclear power plant is utilized to decrease the required amount of electricity for electrolysis as seen in Fig. 2. Nitrogen (N2) Water Uranium Nuclear Electrolysis & Haber-Bosch Ammonia Synthesis Ammonia (NH3) Nuclear Power Plant Electricity Fig. 3. Energy and material nuclear electrolysis based ammonia production On the other hand, in nuclear electrolysis based option, electricity is produced in nuclear power plant and directly utilized in electrolysis coupled with Haber-Bosch ammonia synthesis loop. There is no heat assisting in this method. Hence, more electrical energy is required to split water into hydrogen and oxygen. The schematic diagram of nuclear electrolysis based ammonia production option can be seen in Fig. 3. The copper-chlorine (CuCl) cycle is a multiple step thermochemical cycle for the production of hydrogen. The CuCl cycle is a combined process that employs both thermochemical and electrolysis steps. The CuCl cycle involves four chemical reactions for water splitting, whose net reaction decomposes water into hydrogen and oxygen. Both heat and electricity are provided at the same time for hydrogen generation and then hydrogen reacts with nitrogen to produce ammonia. Input of water and energy for the production of steam are included but other infrastructure is not included, as the heating infrastructure is already a part of the respective heating modules used in the plant. Nuclear power plant electricity is assumed as a mixture of 66.5% from PWR and 33.5% from BWR type reactors for this method, too. The life cycle assessment includes fuel elements, chemicals, and diesel requirements as well as the relevant transport necessities. Water use for cooling is also taken into account. Considered radioactive waste streams are: spent fuel to reprocessing and conditioning; operational low active waste for conditioning in the intermediate repository; and, contaminated waste from dismantling. Non-radioactive wastes are taken into account. The average burnup relates to an average enrichment of 3.8% U235 for fresh uranium fuel elements in BWR type reactor. The average burnup corresponds to an average enrichment of 4.2% U235 for fresh uranium 74 fuel elements in PWR type reactor. The diesel requirements for the yearly test of diesel emergency generators are accounted for. The transport requirements are calculated with the standard distances for chemical and diesel requirements and specific distances for fuel recharge and radioactive waste. Schematic diagram of energy and material flows of nuclear thermochemical CuCl cycle based ammonia production options are shown in Fig. 4. Nitrogen (N2) Water Uranium Nuclear CuCl Thermochemical Cycle & Haber-Bosch Ammonia Synthesis Ammonia (NH3) Nuclear Power Plant Heat Electricity Fig. 4. Energy and material flows of nuclear CuCl cycle based ammonia production III. Life Cycle Assessment (LCA) LCA is a methodology from cradle to grave. This tool helps to make effective decision by analyzing the system systematically. LCA analyses the environmental impact of a product or process over the length of its entire life, beginning from raw material extraction to final disposal. LCA deliberates all the life periods of product or process to assess the overall environmental impact. There are a number of assessment methods progressed over the time to categorize and characterize the environmental flows of system. In this study, LCA is performed using CML 2001 method which was proposed by a set of scientists under the principal of CML (Center of Environmental Science of Leiden University) including a group of impact classes and characterization procedures for the impact assessment phase in The environmental impact categories considered in this study are explained as follows: III.1 Depletion of abiotic resources The key concern of this category is the human and ecosystem health that is affected by the extraction of minerals and fossil as inputs to the system. For each extraction of minerals and fossil fuels, the Abiotic Depletion Factor (ADF) is defined. This indicator has globe scale where it is related with concentration reserves and rate of de-accumulation. III.2. Human toxicity Toxic substances on the human environment are the

89 core concerns for this category. In the working environment, the health risks are not included in this category. Characterization factors, Human Toxicity Potentials (HTP), are determined with (The Uniform System for the Evaluation of Substances) USES-LCA, describing fate, exposure and effects of toxic substances for an infinite time horizon. 1,4-dichlorobenzene equivalents/kg emissions is used to express each toxic substance. Depending on the substance, the geographical scale differs between local and global indicator [9]. III.3. Fresh water aquatic eco-toxicity This indicator considers the effect of the emissions of toxic substances to air, water, and soil on fresh water and ecosystems. USES-LCA is used to calculate the Eco-toxicity Potential by describing fate, exposure and effects of toxic substances. 1,4-dichlorobenzene equivalents/kg emissions is used to express infinite Characterization factors which is the time horizon. The scale of this indicator can be applied to global/continental/regional and local scale. Marine eco-toxicity is related to effects of toxic substances on marine ecosystems [9]. III.4. Ozone depletion Due to stratospheric ozone depletion, a bigger portion of UV-B radiation spreads the world surface. It may have damaging properties upon human health, animal health, terrestrial and aquatic ecosystems, biochemical cycles and on materials. The category is output related and it is at global scale. The model of characterization is advanced by the World Meteorological Organization (WMO) and describes ozone depletion potential of various gasses in unit of kg CFC-11 equivalent/kg emission. The geographic scope of this indicator is at global scale and the span of time is infinity [9]. III.5. Acidification potential Acidifying substances causes a wide range of impacts on soil, groundwater, surface water, organisms, ecosystems and materials. RAINS 10 model is used to calculate the Acidification Potential (AP) for emissions to air, describing the fate and deposition of acidifying substances. The Regional Air Pollution Information and Simulation (RAINS) model is a European-scale integrated assessment model dealing with air quality and associated effects. SO2 equivalents/kg emission is utilized to state the AP. This category has a different geographical scale that can be local and global. Depending on the availability the Characterization aspects containing fate were used. But, when not available, the aspects used without fate (In the CML baseline version only factors including fate were used). The method was stretched for nitric acid, water, soil, and air; sulphuric acid, water; sulphur trioxide, air; hydrogen chloride, water, soil; hydrogen fluoride, water, soil; phosphoric acid, water, soil; hydrogen sulphide, soil, all not including 75 fate. Nitric oxide, air (is nitrogen monoxide) was added containing fate [9]. III.6. Global warming The greenhouse gases to air are associated with the climate change. Adversative effects upon ecosystem health, human health and material welfare can result from climate change. The Intergovernmental Panel on Climate Change (IPCC) developed the characterization model which is selected for the development of characterization factors. A kg carbon dioxide/kg emission is used to express the Global Warming Potential for time horizon 500 years (GWP500). This indicator has a global scale [9]. III.7. Eutrophication This category reflects the impacts of to excessive levels of macro-nutrients in the environment caused by emissions of nutrients to air, water and soil. The stoichiometric procedure of Heijungs is the base of the Nutrification potential (NP) which is expressed as kg PO4 equivalents per kg emission and the geographical scale varies between local and continental scale, time span is infinity, and fate and exposure are not involved [9]. Overall environmental impact of any process is not complete if only operation is considered, all the life steps from resource extraction to disposal during the lifetime of a product or process should be considered. Mass and energy flows and environmental impacts related to plant construction, utilization, and dismantling stages are taken into account in LCA analysis [10, 11]. Using SimaPro software for life cycle analysis, cradle to grave considerations of various nuclear based ammonia production methods are investigated and comparatively assessed. IV. Results and discussion Various nuclear resources based ammonia production pathways are determined, and the energy and material requirement for each route are identified and calculated. The values are used in SimaPro software for the calculations of life cycle assessment. The calculations are based on one kg of ammonia end product. The environmetanl impact results are presented herein. The impact on human health due to human toxicity is maximum for the ammonia production from nuclear electrolysis method where it corresponds to 1.41 kg 1,4-DB-eq per kg of ammonia. Ammonia from nuclear thermochemical based methods yield lower human toxicity values where the lowest is 5 step cycle with a value of 0.80 kg 1,4-DB-eq as seen in Fig. 5.

90 Fig. 5. Human toxicity values of nuclear based ammonia production methods Fig. 7. Acidification values of nuclear based ammonia production methods Fig. 8. Eutrophication values of nuclear based ammonia production methods Fig. 6. Abiotic depletion values of nuclear based ammonia production methods The abiotic resources are natural resources including energy resources, such as iron ore and crude oil, which are considered as non-living. The abiotic depletion is highest for nuclear 5 step CuCl cycle method with a value of kg Sb eq. as it is illustrated in Fig. 6. Besides, nuclear electrolysis based option has the lowest abiotic depletion corresponding to kg Sb eq. In terms of global warming potential, nuclear based electrolysis option yields the lowest environmental impact with a value of 0.48 kg CO2 eq. GHG emission. However, it is very high (3.70 kg CO2 eq.) for the thermochemical based ammonia production as shown in Fig. 9. The nuclear-based high temperature electrolysis has higher global warming potential (0.84 kg CO2 eq.) than nuclear electrolysis. Nuclear electrolysis options have approximately same values with many renewable based ammonia production options found in the literature. The acidification values are lowest for nuclear electrolysis based option (0.002 kg SO2 eq.) followed by nuclear high temperature electrolysis method (0.003 kg SO2 eq.) as shown in Fig. 7. It is higher in thermochemical cycles because of used chemical substances. Eutrophication values are in parallel with acidification values in which nuclear 5 step CuCl cycle has about kg PO4 eq. as illustrated in Fig. 8. Fig. 9. Global warming values of nuclear based ammonia production methods 76

91 The ozone layer depletion is currently an important issue which needs to be decreased. In terms of ozone depletion, nuclear electrolysis yields highest environmental impact corresponding to 7.26E-07 kg CFC-11 eq. In terms of human toxicity, nuclear 5 step CuCl cycle has the lowest environmental impact. Nuclear electrolysis based options yield lower acidification, eutrophication and Abiotic depletion values among all methods. Global warming potentials of nuclear electrolysis and nuclear high temperature electrolysis are relatively lower compared to thermochemical cycles. Nuclear based ammonia production is a promising option especially when utilizing excess heat and electricity. Acknowledgement The authors acknowledge the support by the Natural Sciences and Engineering Research Council of Canada and the Mitacs. Fig. 10. Ozone layer depletion values of nuclear based ammonia production methods The other methods have approximately same values where nuclear 5 step CuCl cycle represents a value of 5.12E-07 kg CFC-11 eq. Nomenclature BWR Boiling water reactor CCS Carbon capture storage CFBG Circulating fluidized bed gasifier DG Downdraft gasifier GHG Greenhouse gas HHV Higher heating value IPCC Intergovernmental panel on climate change LCA Life cycle analysis LPG Liquefied petroleum gas PV Photovoltaic PWR Pressurized water reactor SMR Steam methane reforming UCG Underground coal gasification USES The Uniform System for the Evaluation of Substances Fig. 11. Marine sediment ecotoxicity values of nuclear based ammonia production methods The marine eco-toxicity refers to impacts of toxic substances on marine ecosystems. Highest environmental impact is observed in 3 step CuCl thermochemical cycle which corresponds to 9.11 kg 1,4-DB eq. On the other hand, nuclear high temperature electrolysis occurs to be most environmentally benign method (7.28 kg 1,4-DB eq.) in terms of marine sediment ecotoxicity. V. Conclusions A life cycle assessment of the nuclear ammonia production methods is conducted and environmental impacts are comparatively assessed. The following concluding remarks can be written for this study: 77 References Gilbert P, Thornley P. Energy and carbon balance of ammonia production from biomass gasification. Poster at Bio-Ten Conference, Birmingham Anderson K, Bows A, Mander S. From long-term targets to cumulative emission pathways: Reframing UK climate policy. Energy Policy. 2008;36(10): Zamfirescu C, Dincer I. Ammonia as a green fuel and hydrogen source for vehicular applications. Fuel Processing Technology. 2009;90(5): Verma A, Kumar A. Life cycle assessment of hydrogen production from underground coal gasification. Applied Energy. 2015;147(0): Kalinci Y, Hepbasli A, Dincer I. Life cycle assessment of hydrogen production from biomass gasification systems. International Journal of Hydrogen Energy. 2012;37(19): Zamfirescu C, Dincer I. Using ammonia as a sustainable fuel. Journal of Power Sources. 2008;185(1): ) Makhlouf A, Serradj T, Cheniti H. Life cycle impact

92 assessment of ammonia production in Algeria: A comparison with previous studies. Environmental Impact Assessment Review. 2015;50: Dincer I, Zamfirescu C. Apparatus for using ammonia as a sustainable fuel, refrigerant and NOx reduction agent. Google Patents; Consultants P. SimaPro Life Cycle Analysis Database version 7.3 (software). International Organization for Standardization (ISO) ISO Environmental Management - Life Cycle Assessment e Requirements and Guidelines, Dincer I, Rosen MA, Exergy: Energy, Environment and Sustainable Development, 2nd ed., Elsevier, NY

93 Energy and Exergy Efficiency Evaluations of R134a Clathrates with Additives for Cooling Applications Sayem Zafar* 1, Ibrahim Dincer 1, Mohamed Gadalla 2 1 Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, Ontario L1H 7K4, Canada 2 Department of Mechanical Engineering, American University of Sharjah, Sharjah, UAE * s: Abstract An experimental investigation is performed to evaluate the energetic and exergetic efficiencies of R134a with additives, being treated as phase change materials (PCMs). PCMs charging and discharging characteristics are analysed and evaluated for cooling applications. PCMs are formed using R134a clathrate and distilled water with five different additives. The used additives are sodium chloride, magnesium nitrate hexahydrate, aluminum, copper and ethanol. The refrigerant mass percentage is maintained at 35% while the additive mass percentage at 1% is selected for all the PCMs. The discharge tests are also conducted in which the PCMs are used to cool the air. The main objective of using additives is, then, to study their potential for enhancing the R134a clathrate and exergy contents. The tests are also conducted to determine the energetic and exergetic efficiencies. A comparative assessment study is implemented to compare both energy and exergy efficiencies of different PCMs made up of suggested additives. The exergy destruction evaluations and thermoeconomic analyses are also performed. The present results indicate that the sodium chloride based PCM has the highest and magnesium nitrate hexahydrate based PCM has the lowest exergy destruction. The thermoeconomic analyses include the evaluation of thermoeconomic factor and cost-benefit analyses for the PCMs and the related parameters are studied for each PCM which included its energy, containment and PCM components costs. The ethanol additive is found to have the best overall efficiency when compared with other PCMs. It can be safely concluded that liquid additives are more feasible than other tested additives as they dissolve homogeneously and improve the phase change heat absorption. Finally, the thermoeconomic results show that magnesium nitrate hexahydrate based PCM has the highest thermoeconomic factor while sodium chloride based PCM has the lowest thermoeconomic factor. Keywords: R134a clathrate, phase change materials, cooling, energy, exergy, efficiency, thermoeconomics. I. Introduction The energy management is a challenge that needs to be dealt with in order to achieve the goal of sustainable development. In order to improve the performance of energy systems, more effective tools need to be utilized (Rosen et al. 1997). Exergy is defined as the maximum obtainable work producing ability by a system or a flow of matter as it comes to equilibrium. The exergy of an energy form or a substance is a measure of its usefulness or quality or potential to cause change (Rosen and Dincer, 1997). Exergy analysis is an effective thermodynamic method for design and analysis of thermal systems while it is an efficient technique for revealing the improvement capacity of thermal systems. With energy and exergy precisely known, it becomes easier to evaluate their efficiency for any given system. In thermodynamics, efficiency specifies the effectiveness of the energy conversion process. Efficiency is sometimes misunderstood and defined incorrectly. This is due to the fact that efficiency is often used without being properly defined first (Cengel and Boles, 2015). Basically, efficiency can be described as the ratio of output against the input. This definition of efficiency holds true for all thermodynamic systems and is clearly understood. However the output and input parameters are specific to a system or a component which needs to be specified clearly for efficiency assessment (Zafar et. al. 2014). Research has shown that refrigerant clathrates can be used for cooling applications where phase change is desired above freezing (Mori et al. 1991). Clathrates tend to form when gas molecules get trapped in the water molecule cage under low temperature and high pressure (George. 1989, Sloan. 1990). Refrigerant clathrates can be used for active as well as passive cooling applications hence are considered more effective compared with other type of PCMs (Bi et al. 2004, Inba. 2000). Refrigerant clathrates have high heat of fusion and high density which allows them to store more energy per unit volume. Refrigerant clathrates are no more toxic than the base refrigerant so the existing systems can be utilized to contain them. Many refrigerants form clathrates but only handful are in commercial use. Several chlorofluorocarbons 79

94 (CFCs), hydro- chlorofluorocarbons (HCFCs), and hydrofluorocarbons (HFCs) can form clathrates of refrigerant (Eslamimanesh. 2011). For use in cold thermal energy storage system, the refrigerant clathrate should form between temperature range from 278 K to 285 K (Guo et al. 1996). CFCs are forbidden due to stratospheric ozone layer depletion concerns which leave the hydro-chlorofluorocarbon and hydrofluorocarbons to be used for PCM. Refrigerant clathrates of R-134a show they can be an effective in their role as cold thermal energy storage through phase change (Guo et al. 1996). PCMs based on refrigerant clathrates have poor thermal transport properties. To make refrigerant clathrates as effective PCMs, additives of different materials have been studied. For instance, adding calcium hypochlorite or benzenesulfonic acid sodium salt improved the cold energy storage capacity and the cold energy transfer rate of R141b based clathrate (Bi et al. 2006). Adding alcohol in R134a based clathrate has also been studied which shows it accelerates the cool storage rate and eliminates the floating clathrate during the hydration process (Wua et al. 2012). Adding ethanol as an additive in R134a clathrate has also shown to improve the charging and discharging perforamce of the PCM (Zafar et al. 2015). This paper studies the energy and exergy of the R134a clathrates with and without additives. The paper also studies the thermoeconomic analyses of R134a clathrate with and without additives. The refrigerant clathrate and additives were studied as phase change materials (PCMs) for cooling applications. The PCMs are studied for their charging and discharging energy and exergy values and efficiencies. The cost-benefit analyses are also presented in this paper for each PCM. II. Experimental Setup For the experiments, a cold constant temperature bath from The Clifton Range is used as a constant temperature source (Clifton, NE7). The refrigerant clathrate with additive, named PCM, are formed in glass tubes from ACE Glass Incorporated (ACE Glass Inc.). The tubes are submerged in the constant temperature water bath for which the temperature is set at 276 K and 278 K. The constant temperature bath works by providing cold energy and heat simultaneously to the distilled water in the bath to maintain its temperature at a set value. The graphic illustration of the experimental system is shown in Fig. 1. the amount of energy. A stirrer is also used which circulated the water in the bath. Without the stirrer, the water near the hot or cold source would change its temperature while the water away from the source would see its effect later. Fig. 1: A schematic diagram of the proposed PCM testing system The PCM is formed in the glass tubes. First the glass tube is filled with distilled water and the desired additive. The exact mass of the tube with its constituents is measured using a high accuracy digital weighing scale. The tube is sealed and then vacuumed to get rid of excess air. The last step is to fill the desired refrigerant using a needle valve that allows one way flow. The glass tube is then submerged into the cold temperature water bath for charging. The tubes are visually observed after regular interval to observe the onset and end set of freezing. The freezing times, PCM temperatures and pressures are recorded for each test. Onset of freezing is usually east to detect as the top layer starts freezing. The end set is challenging to pin point so it is important to continue observing the PCM until after the last observed changes in the PCM structure. PCM usually rises as it freezes so height is observed for the end set. The K-type thermocouples are attached to a reader to read the temperatures. For initial charging test, only one temperature reading is taken. For thermal property tests, temperatures are taken at two different locations. The tube is comprehensively tested for leaks and provisions are made to make sure there are no leaks. It is important to use a glass tube since the onset of phase change needs to be observed visually. The illustrative Fig. of the glass tube, its connections and used systems are shown in Fig. 2. A refrigeration system with cooling coils around the water bath pumps out the heat. A controller constantly monitors the water temperature in the bath while continues to provide the desired heat to maintain the desired temperature. The bath is converted into constant energy bath for thermal properties experiments. A constant cold and hot energy is provided to the water in the bath to maintain 80

95 Q The thermal exergy absorbed by the PCM Ex in,pcm is described as: Q (m PCM ex PCM )f - (m PCM ex PCM )i = Ex in,pcm (7) The thermal exergy released by the discharging fluid Q Ex out,c or the thermal exergy released by the stationary solid is described as: Q (mcexc)i - (mcexc)f =Ex out,c = E x Q supply Δt (8) The overall system efficiencies can now be described since useful output and required inputs have been established. The required input is the energy/exergy released by the charging material to change the phase or charge the PCM. The useful output is the energy/exergy absorbed by the discharging material which in turn is absorbed by the PCM. The overall system s energy efficiency can be described as ηoa = Q in Q out (9) The overall system s exergy efficiency can be described as Fig. 2: Instruments for experimental measurements III. Analysis In order to determine the efficiencies, it is first important to describe the useful input and required output of the system. For the charging process, the heat absorbed by the charging fluid Qin,c is described as: [(m chc)δt]out - [(m chc)δt]in = Qin,c (1) While heat given out by the PCM Qout,PCM is (m PCM h PCM )i - (m PCM h PCM )f = Qout,PCM (2) The thermal exergy absorbed by the charging fluid Q is described as: Ex in,c Q [(m cexc)δt]out - [(m cexc)δt]in = Ex in,c (3) Q While thermal exergy given out by the PCM Ex out,pcm is defined as: Q (m PCM ex PCM )i - (m PCM ex PCM )f = Ex out,pcm (4) For the discharging process, the heat absorbed by the PCM Qin,PCM is described as: (m PCM h PCM )f - (m PCM h PCM )i = Qin,PCM (5) The heat released by the discharging fluid Qout,c or the heat emitted by the stationary solid is described as (mchc)i - (mchc)f = Qout,c = Q Δt (6) 81 Ψoa = Ex in Ex out (10) The thermoeconomic analysis on the PCMs is also conducted with the following equation: f TE = Z k Z k +ξ Ex dst (11) where Z k is the total cost of the items used in the PCM in dollars, ξ is the energy cost in $/J and and f TE is the thermoeconomic factor. IV. Results and discussion Fig. 3 shows the average energy utilizations for charging R134a clathrate using five tested additives. The graph shows that magnesium nitrate hexahydrate has the lowest overall energy utilization followed by copper, ethanol, aluminum and then sodium chloride. For 0.01 additive mass fraction, the energy decreased by 55% for magnesium nitrate hexahydrate and 22% for copper. Ethanol and aluminum maintained the energy utilization relatively the same as required by the base R134a clathrate. Sodium chloride increased the energy utilization by 13%. At high additive concentrations, the energy utilization decreased by 27% for magnesium nitrate hexahydrate. Copper, ethanol, aluminum and sodium chloride increased the energy utilization by 27%, 26%, 23% and 60% respectively. Fig. 4 presents the average exergy utilizations during charging process of R134a clathrate for the five tested additives. Similar to the energy utilization trend, magnesium nitrate hexahydrate has the lowest overall energy utilization followed by copper, ethanol, aluminum and then sodium chloride. For 0.01 additive mass fraction, the exergy decreased by 55%

96 for magnesium nitrate hexahydrate and 33% for copper. Ethanol and aluminum maintained the energy utilization relatively the same as required by the base R134a clathrate. Sodium chloride increased the energy utilization by 13%. At high additive concentrations, the energy utilization decreased by 27% for magnesium nitrate hexahydrate. Copper, ethanol, aluminum and sodium chloride increased the energy utilization by 27%, 26%, 23% and 60% respectively. Fig. 5: Energy comparison between PCMs during discharging phase Another factor that hampers the energy absorption is non-homogenous mixing of additives. The metal additives would absorb more energy if they mix well in the clathrate. Ethanol on the other hand, makes hard solid clathrate which allows it to provide cool energy longer than the others. Fig. 3: Energy comparison between PCMs during charging phase Fig. 6 presents the average cool exergy released by each PCM during the discharge phase. Exergy follows the same trend as energy. Magnesium nitrate hexahydrate, again, has the lowest exergy while ethanol additive has the highest. The reason for magnesium nitrate hexahydrate additive s low energy release is its soft clathrate structure. Ethanol, on the other hand, makes hard solid clathrate which allows it to provide cool energy longer than the others. Fig. 4: Exergy comparison between PCMs during charging phase Fig. 5 shows the average cool energy released by each PCM during the discharge phase. The graph shows that ethanol additive has the highest overall energy release followed by aluminum, copper, base clathrate and magnesium nitrate hexahydrate. The reason for magnesium nitrate hexahydrate additive s low energy release is its soft clathrate structure. Additives that make soft small clathrate structures tend to absorb low energy. 82 Fig. 6: Exergy comparison between PCMs during discharging phase Fig. 7 presents the overall energy and exergy efficiencies of each PCM. The solid bar represents the exergy values while bars with pattern fill represent the energy. The input and output values are utilized to evaluate these efficiencies. The efficacy graph shows the true picture of which additive results in most gain. Ethanol additive shows the highest efficiency while sodium chloride shows the least. Ethanol, in spite taking long to freeze, yields the most during discharge hence proved to be the most useful. Sodium chloride additive talks long

97 to charge while does not last very long during discharge which makes it the least efficient additive. Fig. 8 shows the exergy destruction for each PCM used in the experiments. Sodium chloride based PCM has the highest exergy destruction of 12 kj. Magnesium nitrate hexahydrate based PCM has the lowest exergy destruction of 4 kj. Sodium chloride based PCM has high charging time and a comparative low discharge time which results in high exergy destruction. Magnesium nitrate hexahydrate based PCM has low charging time which results in relatively low exergy destruction. It is to be noted that even though magnesium nitrate hexahydrate based PCM has the lowest exergy destruction, it may not necessarily be the most useful PCM, overall. IV.2. Results of Thermoeconomic Analysis Using the equations described in the Analyses section, thermoeconomic analyses are conducted on the PCMs used in the experiments. Thermoeconomic analyses include the evaluation of thermoeconomic factor and cost-benefit analyses for the PCMs. Fig. 9 shows the variation of thermoeconomic variable, f TE as it changes with respect to each PCM. Thermoeconomic variable, f TE, is studied for each PCM including its energy, containment and PCM components costs. For thermoeconomic factor, higher the value, more feasible it is. The results show that magnesium nitrate hexahydrate based PCM has the highest thermoeconomic factor while sodium chloride based PCM has the lowest thermoeconomic factor. The low thermoeconomic factor for magnesium nitrate hexahydrate based PCM is due to its low exergy destruction. Similarly, high thermoeconomic factor for sodium chloride based PCM is due to its high exergy destruction. Fig. 7: Overall energy and exergy efficiencies of PCMs Fig. 9: Thermoeconomic variable values of each PCM Fig. 8: Exergy destruction values of each PCM using different additives 83 Fig. 10: Energy costs of producing and using PCMs

98 Fig. 10 shows the energy cost of producing the PCM and amount saved using the PCM. The energy cost is calculated using the electricity unit rate of $0.32. Ethanol, having the highest efficiency, gives the highest return in terms of discharge energy. Sodium chloride has the greatest difference between charging and discharging price due to its low efficiency. As it can be seen from the graph that the energy cost of charging 100 units is not very high and it remains below $5. Magnesium nitrate hexahydrate shows to have the lowest energy cost but it also has the lowest return. aluminum and then sodium chloride. For discharging process, PCM with ethanol additive has the highest overall energy and exergy release followed by aluminum, copper, base clathrate and magnesium nitrate hexahydrate. The efficacy graph shows the true picture of which additive yields most gain. Ethanol additive shows the highest efficiency while sodium chloride shows the least. Sodium chloride based PCM has the highest exergy destruction of 12 kj. Magnesium nitrate hexahydrate based PCM has the lowest exergy destruction of 4 kj. The results show that magnesium nitrate hexahydrate based PCM has the highest thermoeconomic factor while sodium chloride based PCM has the lowest thermoeconomic factor. Ethanol, having the highest efficiency, gives the highest return in terms of discharge energy. Ethanol additive makes the most economical PCM since it costs relatively low to produce it yet it gives the highest efficiency. Fig. 11: Cost of producing 100 units of each PCMs Fig. 11 shows the costs of producing 100 units of each PCM with additives compared to the base PCM without any additive. This Fig. includes the price for 80 gram additive and the costs of energy to charge it. The energy cost is calculated using the electricity unit rate of $0.32. Copper additive proves to be the most expensive primarily due to the price of copper particles. Magnesium nitrate hexahydrate has low cost since it does not take long to get charged. It should be noted that magnesium nitrate hexahydrate does not produce a lot either so its low price is somewhat deceiving. Ethanol has the second lowest cost due to its low price. Ethanol additive makes the most economical PCM since it costs relatively low to produce it yet it gives the highest efficiency. V. Conclusions Experimental studies are undertaken on refrigerant clathrates for use in cooling applications. Refrigerant R134a is used to form the clathrate. Sodium chloride, magnesium nitrate hexahydrate, aluminum particles, copper particles and ethanol is used as an additive to determine their impact on the refrigerant clathrate. Energy, exergy, efficiencies and thermoeconomics are evaluated for R134a with and without additives. Based on the obtained results, the following findings may be drawn: For charging process, PCM with magnesium nitrate hexahydrate has the lowest overall energy and exergy utilization followed by copper, ethanol, 84 Nomenclature Ex : Exergy (J) ex : Specific exergy (J/kg) fte : Thermoeconomic factor h : Specific enthalpy (J/kg) m : Mass (kg) m : Mass flow rate (kg/s) PCM : Phase change material Q : Heat (J) Q : Heat flow rate (W) t : Time (sec) Zk : Total cost ($) Greek letters ξ : Energy cost ($/J) η : Energy efficiency ψ : Exergy efficiency Subscripts c : Charging d : Discharging i : initial f : Final Superscript Q : Thermal References ACE Glass Incorporated. Pressure Tube 185 ml. Item Number Bi YH, Guo TW, Zhu TY, Zhang L, Chen L., Influences of additives on the gas hydrate cool storage process in a new gas hydrate cool storage system. Energy Conversion and Management 47: (2006) Bi YH, Guo TW, Zhu TY, Fan SS, Liang DQ, Zhang L., Influence of volumetric-flow rate in the crystallizer on the gas-hydrate cool-storage process in a new gas-hydrate cool-storage system. Applied Energy 78:

99 (2004) Cengel Y A, M. A. Boles Thermodynamics: An Engineering Approach, 8th ed. McGraw-Hill: New York. Eslamimanesh A, Mohammadi AH, Richon D., Thermodynamic model for predicting phase equilibria of simple clathrate hydrates of refrigerants. Chemical Engineering Science 66: (2011) George A., Hand book of thermal design. Guyer C, editor. Phase change thermal storage materials. McGraw Hill Book Co.; chapter 1 (1989) Guo KH, Shu BF, Zhang Y., Transient behavior of energy charge discharge and solid liquid phase change in mixed gas-hydrate formation. In: Wang, B.X. (Ed.), Heat Transfer Science and Technology. Higher Education Press, Beijing (China): (1996) Inaba H., New challenge in advanced thermal energy transportation using functionally thermal fluids. International Journal of Thermal Science 39: (2000) Mori YH, Isobe F., A model for gas hydrate formation accompanying direct-contact evaporation of refrigerant drops in water. International Communications in Heat Mass Transfer 18: (1991) Rosen MA., Dincer I., On exergy and environmental impact, International Journal of Energy Research (1997) Rosen M.A. and I. Dincer On exergy and environmental impact. International Journal of Energy Research 21: Sloan ED., Clathrate hydrates of natural gases. New York: Marcel. (1990) The Clifton Range. Chillo Baths NE7 Series 8 igerated-water-baths/ Wua J, Wangb S., Research on cool storage and release characteristics of R134a gas hydrate with additive. Energy and Buildings 45: (2012) Zafar S, Dincer I., Efficiency Assessment of Crude Oil Distillation System. In Progress in Exergy, Energy and Environment, edited by Ibrahim Dincer, Adnan Midilli and Haydar Kucuk, Chapter 19, pp United States, ISBN : Springer International Publishing Switzerland, (2014). Zafar S, Dincer I, Gadalla. M., Experimental testing and analysis of R134a clathrates based PCMs for cooling applications. International Journal of Heat and Mass Transfer 91: (2015). 85

100 Thermodynamic Performance Analysis of a Raw Mill System in Cement Plant Mehmet Altinkaynak 1*, Murat Ozturk 2, Ali Kemal Yakut 1, 1* Süleyman Demirel University, Faculty of Technology, Department of Energy Systems Engineering, 32260, Isparta, Turkey 2 Süleyman Demirel University, Faculty of Technology, Department of Mechatronic Engineering, 32260, Isparta, Turkey * Abstract The cement production process is one of the most power-intensive and higher harmful gas emitting process in the world. The energy policies of many development and under developing countries focus on increasing energy efficiency in the industry, which in turn, causes to decrease harmful gas emissions. In this paper, the energy and exergy analyses of a raw mill in cement plant are investigated for better understanding of the system design dynamics. Exergy destruction rate and exergy efficiency are obtained using by the thermodynamic analysis. The system design parameters, which effect the process performance, such as the ambient temperature, the mass flow rate and component temperature are analyzed. Keywords: Thermodynamic analysis, cement plant, raw mill. I. Introduction The cement production facility is the energy extensive plant which makes every effort related to power consumption, performance and generation. The cement production process contains many indicators, such as i-) grinding and blending raw materials (limestone, shells or chalk, and shale, clay, sand, or iron ore), ii-) heating those materials to very high temperatures in a kiln, iii-) cooling and mixing those materials with gypsum, and iv-) finally, grinding down the mixture to form cement powder. The exergy analysis presents the system design as closely as allowable to the maximum theoretical limit. The development of design technique for the raw mill component with increasing performance is a vital task. The transfer of heat between inlet layer of the raw mill and environment is the most generally encountered operation in component design process. Numerous theoretical and experimental analyses to investigate the raw mill system in cement industry have reported in the literature. Sogut et al. (2010), have analyzed the heat recovery modelling from a rotary kiln process in cement industry to the environment using by the energy and exergy analysis viewpoints. Also, the authors have investigated the energetic and exergetic efficiencies, and exergy destruction rates of the rotary kiln for cement production process. Madlool et al. (2012), have given the energy and exergy analyses equations, the exergy balance equations, and also energy and exergy efficiencies for the components of a cement industry for investigating of the cement generation processes. In addition, Ahamed et al. (2012), have investigated for increasing the energetic efficiency, the exergetic efficiency and the recovery performance of a cooling process through the optimization of its working indicators, such as i-) the mass flow rate of working fluid and clinker, ii-) the cooling working fluid temperature, and iii-) the grate speed. Also, they have analyzed the thermodynamic performance analysis to investigate how the working indicators of grate clinger cooling process and the heat recovery from the hot exhaust gasses. Atmaca and Kanoglu (2012), have studied the energetic and exergetic analyses of a raw mill in cement plant and specific measures in order to decrease the quantity of energy consumption in grinding system. also, they have found the energetic and exergetic efficiency as 61.5% and 16.4%, respectively. Gutierrez et al. (2012), have investigated the power consumption and the exergy destruction rate of the calcination system in the vertical shaft kilns, in order to identify the indicators affecting energy consumption. They have given that the most exergy destruction rate, due to fuel combustion, internal heat and momentum transfer taking place in the kiln process. In this paper, the methods for determining the magnitudes and causes of exergy destruction in the raw mill system in cement industry is detailed investigated by using thermodynamic analysis. Furthermore, the impact of design parameters on the system performance are evaluated under different operating conditions. II. System description The raw mill system is an important component among other parts of the cement plant. Because, the raw mill system is used to grain the crude inputs into the farine output which is the semi-product of clinker output. The schematic diagram of the raw mill system in a cement plant is illustrated in Fig. 1. In this process, the raw materials, such as CaCO2, SiO2, Al2O3, Fe2O3, MgO, K2O, SO3 and Na2O at reference temperature and pressure enter the raw mill system to produce farine. The producing farine, which is consisting of CaO, CO2, SiO2, Al2O3, Fe2O3, MgO, 86

101 K2O, SO3 and Na2O enters to the farine silo at point 2. Also, the producing farine dust goes to the filter at point 3. The additional of stack gases for the farine production process has significant indicators for the heat needs during farine generation. Therefore, the heated gaseous, such as N2, O2, CO2, CO and SO2 input to the raw mill system at point 4 to give its heat to the raw materials. m i = m e (2) where m is the mass flow rate, subscript i and e are the inlet and outlet flows, respectively. III.2. Energy balance The energy analysis of the control volume deals with all energy parts of the chosen control volume. The energy balance equation, which is given as the first law of the thermodynamics, can be given as follows; Q + m inh in = W + m out h out (3) where Q, W and h are the heat transfer rate, power and specific enthalpy, respectively. III.3. Exergy balance The exergy can be described as the maximum work that should be provided from the process at a chosen state. To evaluate the exergy analysis, firstly the reversible work must be defined. The reversible work can be defined as the maximum useful work that can be provided as the system goes through a process between two given states. The general exergy balance rate can be written as follows; E x Q + m in ex in = E x W + m out ex out + E x D (4) Fig. 1. Schematic diagram of a raw mill system III. Thermodynamic analysis Thermodynamic assessments based on the energy and exergy analyses are used to examined the performance, energy loss rate and exergy destruction rate in order to increase the efficiency of the investigated process and its components. The mass, energy and exergy balance equations are used to investigate the exergy destruction rate, the energy and exergy content of any stream, the energy efficiency and exergy efficiency of the process for detailed information. Generally, based on the usual principle, the thermodynamic balance equation for a quantity in investigated system can be defined as (Dincer, 2012) Input + Generation Output Consumption = Accumulation (1) Also, in the steady state condition, the accumulated indicator in the Eq. (1) is equal to zero, because whole properties in the system are unchanging with time. III.1. Mass balance The conservation of mass is a fundamental procedure in investigating any thermodynamic process. The mass balance equation can be written as follows; where ex is the specific exergy and E x D is the rate of exergy destruction. E x Q and E x W are the rate of exergy transfer by heat, and work, respectively, and can be calculated as follows; E x Q = (1 T o T ) Q (5) E x W = W (6) where T_o is the reference temperature, and T is the temperature at which heat transfer takes place. The specific exergy can be given as follows; ex = ex ke + ex pe + ex ph + ex ch (7) where ex ke, ex pe, ex ph and ex ch are the kinetic, the potential, the physical and the chemical exergies, respectively. In this paper, kinetic and physical exergy are accepted negligible. The physical exergy or specific flow exergy can be written as ex ph = (h h o ) T o (s s o ) (8) where s is the specific entropy. The chemical exergy of gas mixture can be given as ex ch = x i ex o ch + RT o x i ln(x i ) (9) o where ex ch is the standard chemical exergy of an element in kj mol and x i is the mass fraction of in element i and subscript o stands for dead state. Total exergy rate can be given as follows; 87

102 E x = m ex (10) III.4. Thermodynamic analysis of raw mill In this section, the energy and exergy analyses of raw mill system in cement plant are investigated. To reach this aim, the mass, energy and exergy balance equations for input and output flows of the raw mill process are evaluated. The mass balance equation of the raw mill sub-system in the cement production process can be written as follows: m 1 + m 4 = m 2 + m 3 (11) Based on the general energy balance equation, which given in the Eq. (3), the energy balance equation for the raw mill can be defined as follows: m 1h 1 + m 4h 4 + Q loss RM = m 2h 2 + m 3h 3 (12) The exergy balance equation of the raw mill is given as m 1ex 1 + m 4ex 4 + E x RM Q,loss = m 2ex 2 + m 3ex 3 + E x RM D (13) The Raw materials enters the Raw mill at point 1. The chemical compositions of Raw materials are given in Table 1. Tab 1. Chemical compositions of Raw materials at point 1. Mass Molar Mass Mass Flow Raw Concentration (kg/kmol) Rate Materials (wt. %) M (kg/s) Y CaO 56,077 CaCO 2 CO 2 44, SiO 2 60, Al 2O 3 101, Fe 2O 3 159, MgO 40, K 2O 94, SO 3 80, Na 2O 61, Total The specific exergy of the flow at state1 is given as follows; ex 1 = Y CaO M CaO ex ch CaO ch + Y CO2 M CO2 ex ch CO2 + ch + Y SiO2 M SiO2 ex SiO2 + Y Al2 O 3 M Al2 O 3 ex Al2 O 3 ch ch Y Fe2 O 3 M Fe2 O 3 ex Fe2 O 3 + Y MgO M MgO ex MgO + ch Y K2 OM K2 Oex K2 O + Y SO3 M SO3 ex ch ch SO3 + Y Na2 OM Na2 Oex Na2 O (14) where chemical exergy of CaO, CO 2, SiO 2, Al 2 O 3, Fe 2 O 3, MgO, K 2 O, SO 3 and Na 2 O are given as follows, respectively; ch ex CaO 0 = h CaO (h Ca + 0.5hO2 ) T0 [s CaO (s 0 Ca + 0.5s 0 O2 )] + ex ch ch Ca + 0.5ex O2 ex ch CO2 0 = h CO2 ex ch ch C + ex O2 (15) (h C + ho2 ) T0 [s CO2 (s 0 C + s 0 O2 )] + (16) 88 ch ex SiO2 0 = h SiO2 s 0 O2 )] + ex ch ch Si + ex O2 ch ex Al2 O 3 0 (s Al2 ch ex Fe2 O 3 0 (s Fe2 ch ex MgO 0 (s Mg ex ch K2 O 0 = h Al2 O (h Si + ho2 ) T0 [s SiO2 (s 0 Si + 0 (h Al s 0 O2 )] + ex ch ch Al ex O2 0 = h Fe2 O 3 0 (h Fe s 0 O2 )] + ex ch ch Fe ex O2 0 = h MgO 0 (h Mg + 0.5s 0 O2 )] + ex ch ch Mg + 0.5ex O2 0 = h K2 O 0.5s 0 O2 )] + ex ch ch K ex O2 ex ch SO3 0 = h SO3 s 0 O3 )] + ex ch ch S + 1.5ex O2 ch ex Na2 O 0 (s Na2 0 = h Na2 O h O2 ) T0 [s Al2 O h O2 ) T0 [s Fe2 O h O2 ) T0 [s MgO (17) (18) (19) (20) (h K hO2 ) T0 [s K2 O (s 0 K (h S + 1.5hO2 ) T0 [s SO3 (s 0 S + 0 (h Na s 0 O2 )] + ex ch ch Na ex O h O2 ) T0 [s Na2 O (21) (22) (23) where Mi is the molar mass (kg/kmol) of ith substance, Yi is the mass concentartion (wt. %) of the ith substance. Also, Mi and Yi are given in Table 1. The stack gases enters the Raw mill at point 19. The chemical compositions of stack gases are written in Table 2. Tab 2. Chemical compositons of stack gases at point 4 Mass Molar Mass Mass Stack Concentration (kg/kmol) Flow Rate Gases (wt. %) M (kg/s) Y N 2 28, O 2 31, CO 2 44, CO 28, SO 2 64, Total The spesific exergy of the flow at state 4 can be calculated as follows; ex 4 = Y N2 M N2 ex ch N2 + Y O2 M O2 ex ch O2 + Y CO2 M CO2 ex ch CO2 + Y CO M CO ex ch ch CO + Y SO2 M SO2 ex SO2 (24) where the chemical exergy equations of CO and SO 2 are given as ex ch CO = h CO (hc + 0.5hO2 ) T0 [s 0 CO (s 0 C + 0.5s 0 O2 )] + ex ch ch C + 0.5ex O2 ex ch SO2 0 = h SO2 ex ch ch S + ex O2 (25) (h S + ho2 ) T0 [s SO2 (s 0 S + s 0 O2 )] + (26)

103 The Raw materials exit from the raw mill, is called as farine stored in the farine silo at point 2. The chemical compositions of farine at point 2 are given in Table 3. Tab 3. Chemical compositions of farine at point 2 Farine Molar Mass Mass Mass Flow (kg/kmol) Concentration Rate M (wt. %) (kg/s) Y CaO 56, CO 2 44, SiO 2 60, Al 2O 3 101, Fe 2O 3 159, MgO 40, K 2O 94, SO 3 80, Na 2O 61, Total The specific exergy of flow at point 2 is given as follows; ex 2 = Y CaO M CaO ex ch CaO ch + Y CO2 M CO2 ex ch CO2 + ch + Y SiO2 M SiO2 ex SiO2 + Y Al2 O 3 M Al2 O 3 ex Al2 O 3 ch ch Y Fe2 O 3 M Fe2 O 3 ex Fe2 O 3 + Y MgO M MgO ex MgO + ch Y K2 OM K2 Oex K2 O + Y SO3 M SO3 ex ch ch SO3 + Y Na2 OM Na2 Oex Na2 O (27) The stack gases plus farine dust exist from the raw mill, and goes to the electro filter at point 3. The chemicalcomposition at flowing materials at point 3 are given in Table 4. Tab 4.Chemical compositions of flowing materials at point 3. Molar Mass Mass Mass Flow Flowing (kg/kmol) Concentration Rate Materials M (wt. %) (kg/s) Y CaO 56, CO 2 44, SiO 2 60, Al 2O 3 101, Fe 2O 3 159, MgO 40, K 2O 94, SO 3 80,064 < < Na 2O 61, N 2 28, O 2 31, CO 28, SO 2 64, Total The specific exergy of flow at point 3 is given as follows; ex 3 = Y CaO M CaO ex ch CaO ch + Y CO2 M CO2 ex ch CO2 + ch + Y SiO2 M SiO2 ex SiO2 + Y Al2 O 3 M Al2 O 3 ex Al2 O 3 ch ch Y Fe2 O 3 M Fe2 O 3 ex Fe2 O 3 + Y MgO M MgO ex MgO + ch Y K2 OM K2 Oex K2 O + Y SO3 M SO3 ex ch SO3 + ch Y Na2 OM Na2 Oex Na2 O + Y N2 M N2 ex ch N2 + Y O2 M O2 ex ch O2 + Y CO M CO ex ch ch CO + Y SO2 M SO2 ex SO2 (28) The heat loss rate from the raw mill system to the environment, which given in Eq. (12), can be calculated as follows (Atmaca and Kanoglu, 2012); Q loss RM = T RM T o R tot (29) where T RM is the raw mill temperature R tot is the net thermal resistance of the raw mill, and can be evaluated as follows; R tot = R conv,1 + R cond + R conv,2xr rad R conv,2 +R rad (30) The convection 1 and 2, conduction and radiation thermal resistance can be calculated as follows; R conv,1 = 1 2πr 1 hl R conv,2 = 1 2πr 2 hl ) R cond = ln(r 2 r 1 2πkL R rad = 1 h rad A (31) (32) (33) (34) Where h and k are the convection coefficient and thermal conductivity, respectively, h rad the radiation heat transfer coefficient, and can be defined as follows; 2 h rad = εσ(t out,surf + T 2 out )(T out,surf + T out ) (35) where ε is the emissivity of the rwa mill system surface and σ is the Stefan-Boltzman constant as5.67x10 8 W m 2 K 4. III.5. Energy efficiency The efficiency of any system should be written in terms of useful outputs from the system boundary divided by the total inputs to the system. According to this description, for a general system, the energy efficiency can be defined as (Dincer, 2011) η = useful energy in outputs total energy inputs (36) The energy efficiency equation of the raw mill system in cement plant is defined as follows; η RM = m 2h 2 +m 3h 3 m 1h 1 +m 4h 4 (37) III.6. Exergy efficiency The exergetic efficiency analysis gives some very significant indicators about the process and its parts for increasing and efficiently use. For a general process, an exergy efficiency equation is given as follows (Dincer, 2011); ψ = exergy in outputs total exergy inputs (38) The exergy efficiency of the raw mill system in cement plant, which is presented in Fig. 1, is written as 89

104 ψ RM = m 2ex 2 +m 3ex 3 m 1ex 1 +m 4ex 4 (39) IV. Results and discussions In this paper, the reference temperature and pressure are taken as 25 C and kpa, respectively. The thermodynamic properties of the material flows in the raw mill system for the investigating cement production process are determined using by the EES software program (Klein, 2010). The EES code is also developed to investigate the performance of the raw mill system and its parts. The heat loss rate from the raw mill system to the ambient is calculated as 14, MJ/h. The exergy efficiency of the raw mill system is calculated as 34.45%. In addition, the drying part and grinding part have the maximum heat loss rate for the process. The heat loss rate for these components are calculated as and 64.74%, respectively. The design parameters of the raw mill system used in this study, such as the mass flow rate, temperature, pressure, specific exergy and exergy rate are given in Table 5. Tab 5. Thermodynamic properties of materials with each state point for raw mill system State point Mass flow rate (kg/s) Temperature ( C) Pressure (kpa) Exergy rate (MJ) , , , ,160 Generally, the exergy efficiency of any process is lower than energy efficiency, and that important development possibility exist. The cement producing sectors are the energy and exergy intensive system. There are significant opportunities to identify parts where energy and exergy savings measurements should be performed so that energy and exergy should be saved along with the increase of harmful gaseous. The effects of input material temperature on the exergy destruction rate and exergy efficiency of the raw mill system are illustrated in Fig. 2. As seen in this figure, increasing temperature of the input materials has positive effect on the system exergy efficiency. Also, the exergy destruction rate of the raw mill system is decreased with increasing input material temperature. The impacts of input materials mass flow rate on the exergy destruction rate and exergy efficiency of the raw mill system are given in Fig. 3. As illustrated in this figure, increasing input material mass flow rate increase the exergy destruction rate and exergy efficiency of the raw mill system. The heated gaseous temperature at point 4 is very important for efficiently system design. Therefore, the effects of heated gaseous temperature on the system energy efficiency and exergy efficiency is given in Fig. 4. The energy efficiency of the raw mill system is increased from 52.25% to 54.25% with increasing 90 heated gaseous temperature from 350 C to 750 C. Also, the exergy efficiency of the raw mill system is increased from 31.45% to % with increasing heated gaseous temperature. Ex D (MJ) Ex D (MJ) Temperature of input materials Fig. 2. Effect of input materials temperature on exergy destruction rate and exergy efficiency Fig. 3. Effect of mass flow rate on exergy destruction rate and exergy efficiency Energy efficiency Fig. 4. Effect of flow 4 temperature rate on energy and exergy efficiency V. Conclusions Ex D (MJ) y In this paper, the energy and exergy analyses of the raw mill system in cement plant are given to investigating cement generation process using with the actual facility data. Because energetic viewpoint cannot give adequate information about the energy losses, exergetic viewpoint is performed in order to investigated real efficiencies and destructions of the raw mill system in cement production process. Also, the parametric studies are conducted in order to find out how the input material temperature, mass flow rate of input materials and temperature of flow at Mass flow rate of input materials (kg/s) Ex D y h y Temperature of flow Exergy efficiency Exergy efficiency Exergy efficiency

105 point 4 affect the efficiency of the raw mill system in cement plant. References Sogut, Z., Oktay, Z., Karakoc, H. Mathematical modeling of heat recovery from a rotary kiln, Applied Thermal Engineering 30 (2010) N.A. Madlool, R. Saidura, N.A. Rahim, M.R. Islam, M.S. Hossian, An exergy analysis for cement industries: An overview, Renew Sust Energ Rev 16 (2012) U Ahamed, N A Madlool, R Saidur, M I Shahinuddin, A Kamyar, H H Masjuki, Assessment of energy and exergy efficiencies of a grate clinker cooling system through the optimization of its operational parameters, Energ 46 (2012) A Atmaca, MKanoglu, Reducing energy consumption of a raw mill in cement industry, Energ 42 (2012) A S Gutiérrez, J B Cogollos Martínez, C Vandecasteele, Energy and exergy assessments of a lime shaft kiln, Appl Therm Eng 51 (2013) Dincer, I., Rosen, M. A. Exergy: Energy, Environment and Sustainable Development, Elsevier, 225 Wyman Street, Waltham, MA 02451, USA, Second edition Dincer, C. Zamfirescu, Sustainable energy systems and applications, New York, NY: Springer, Klein, S.A., Engineering equation solver. Academic Professional, Version 8,

106 THERMAL SYSTEMS AND APPLICATIONS 92

107 Exergetic Assessment of PTSC Integrated Power-Refrigeration System Working with CO2 Ahmet Kabul 1, Onder Kizilkan 2* 1,2 Süleyman Demirel University, Faculty of Technology, Department of Energy Systems Engineering, 32260, Isparta, Turkey * Abstract This study deals with energy and exergy analysis of a solar driven combined power-refrigeration system. The system comprises a supercritical Brayton cycle, a transcritical organic Rankine cycle and a subcritical vapor compression refrigeration cycle. The three systems operates with carbondioxide (CO2) as working fluid because of its zero ozone depleting potential and with negligible global warming potential. Also it is a sustainable working fluid. The combined process includes parabolic trough solar collector system for providing the heat demand of the supercritical Brayton Cycle. The rejected heat from supercritical Brayton Cycle is used for heat energy demand of organic Rankine cycle. The refrigeration cycle is driven by the power generated from the organic Rankine cycle. With the results, all the irreversibility rates of the combined system are determined. Additionally, a parametric study is carried out to examine the variation of energy and exergy efficiency rates of the three systems. Keywords: Thermal energy storage, solar energy, phase change material, latent heat I. Introduction The demand for energy is continually increasing while conventional fossil fuel energy resources are being consumed at an alarming rate. It became very important that reliable and more sustainable energy resources are required to compensate for the uncertainty surrounding the supply of fossil fuels. Renewable energy sources, such as solar, biomass, geothermal, wind, and hydro, can be good alternatives to conventional fuel sources. These sustainable energy sources are available in sufficient quantities and have minimal impact on the environment (Al-Sulaiman and Atif 2015). Solar assisted power systems have the potential to generate electricity particularly in places with high insolation levels. In order to be competitive with conventional power plants, further development of this renewable technology is necessary, such as the use of more efficient power cycles and solar components, increasing deployment, reducing manufacturing cost and integration of thermal energy storage to improve dispatchable power on demand (Padilla et al., 2015). It is very important to develop a highly efficient, relatively low-cost power conversion system with the minimal environmental impacts by the development of the new concept and advanced power cycles as well as by the refinement of the existing conventional power cycles. Carbon dioxide, due to its non-toxicity, non-flammability, abundance and low cost, is an economic competitive and environmental favorable working fluid. Supercritical Carbon dioxide Brayton cycle (SCO2-BC), in view of its simple system layout, superior cycle efficiency, compact plant components, 93 and friendly environmental influence, have currently received significant attention and being studied for widely applications in nuclear, fossil, concentrating solar power (CSP), waste heat recovery and ship propulsion systems (Li et al., 2016) Research effort to find alternative methods of generating electricity is not new and the advantages of the supercritical Brayton cycle with carbon dioxide has been exposed and discussed during the 60's (Mecheri and Le Moullec, 2016). For the last decades, there is a pretty high interest on such power cycles those utilize low-moderate temperature heat sources. Akbari and Mahmoudi (2014) reported exergoeconomic analysis for a new combined supercritical CO2 recompression Brayton/organic Rankine cycle in which the waste heat from supercritical CO2 recompression Brayton cycle was utilized by an organic Rankine cycle for generating electricity. Al-Sulaiman and Atif (2015), conducted a thermodynamic comparison of five supercritical carbon dioxide Brayton cycles integrated with a solar power tower. Padilla et al. (2015) performed a detailed energy and exergy analysis of four different supercritical CO2 Brayton cycle configurations with and without reheat. Kim et al. (2016), investigated heat transfer performance and pressure drop of PCHE with CO2 as a working fluid with wide Reynolds number ranges using CFD analysis. Li et al. (2016), examined the modeling of the forced convection heat transfer of carbon dioxide at supercritical pressures within the PCHE for SCO2-BC. Mecheri and Moullec (2016), investigated the supercritical CO2 cycles performance from thermodynamic consideration for coal power plant application. Kouta et al. (2016), conducted the performance and cost analyses of a solar power tower integrated with supercritical CO2

108 Brayton cycles for power production and a multiple effect evaporation with a thermal vapor compression desalination system for water production. They performed analyses for two configurations based on two different supercritical cycles. Linares et al. (2016), presented an exploratory analysis of the suitability of supercritical CO2 Brayton power cycles as alternative energy conversion systems for a future fusion reactor based on a dual coolant lithium-lead blanket. Hu et al. (2015), studied the performance of a supercritical gas Brayton cycle using CO2-based binary mixtures as the working fluids have been studied. Liu et al. (2016), presented the theoretical analysis and on-site testing on the thermal performance of the waste heat recovery system for offshore oil production facilities. They used the ideal air standard Brayton cycle to analyze the thermal performance. Garg et al. (2015), introduced a quantitative methodology for load regulation of a CO2 based Brayton cycle power plant using the thermal efficiency and specific work output coordinate system. Padilla et al. (2016), proposed three S-CO2 Brayton cycle configurations without reheat by introducing an ejector prior the heater, which reduced the pressure at the solar receiver. They performed a comprehensive thermodynamic analysis and a multi-objective optimization. Serrano et al. (2014) proposed a new layout of the classical recompression supercritical CO2 Brayton cycle which replaces one of the recuperators by another which bypasses the low temperature blanket source. Manente and Lazzaretto (2014) analyzed the supercritical closed CO2 Brayton cycles. They made the analyses to explore the thermodynamic performance and the technical feasibility of such systems. Baroncia et al., (2015) investigated the integration of a molten carbonate fuel cell and a Brayton cycle that uses supercritical carbon dioxide as working fluid. Rovense (2015), performed a numerical analysis performed by SAM for the solar tower for the assessment of the performance of a solar closed air Brayton cycle. Rovira et al. (2015), proposed a configuration named Hybrid Rankine Brayton cycle with balanced recuperator, as well as its operating conditions and potential working fluids, for low to moderate temperature solar applications. Iverson et al. (2013), investigated the response of a prototype sco2 Brayton cycle under transient operating conditions similar to that experienced in a typical solar plant with a direct receiver. In this study, thermodynamic assessment of a solar driven combined power-refrigeration system is conducted. The system includes a supercritical CO2 Brayton cycle (sco2-bc), a transcritical CO2 Organic Rankine Cycle (tco2-orc) and a subcritical CO2 vapor compression refrigeration cycle (CO2-VCRC). A parabolic trough solar collector system (PTSC) provides the necessary heat demand of the supercritical CO2 Brayton Cycle. The ORC system uses the rejected heat from Brayton cycle. The CO2 refrigeration cycle is driven by the power generated from the ORC which are integrated together. II. Combined CO 2 System with PTSC The schematic representation of solar driven combined power-refrigeration system is shown in Fig. 1. The combined system consists of PTSC system, sco2-bc, and tco2-orc, for power generation and CO2-VCRC for refrigeration. Since the use of CO2 as a working-fluid of power and refrigeration cycles has been growing in recent years due to associated benefits (Singh et al. 2013), it has been selected as working fluid for all cycles. The P-h diagram of the three cycles which are being investigated are given in Fig. 2. Fig. 1: Schematic representation of proposed system 94

109 valve where it becomes wet vapor at low pressure. After expansion valve, the refrigerant passes through evaporator where it absorbs necessary heat energy to become saturated vapor while it refrigerates the cold room. Fig. 2: The P-h diagram of the three cycles Solar energy is collected using a PTSC system for supplying heat demand of the cycles. For PTSC system, Therminol-VP1 is selected as the heat transfer fluid (HTF) for its good heat transfer properties and good temperature control (Therminol 2014). Because of its good properties, it is being used in many high temperature applications driven by PTSC such as power plants (Kumar and Reddy 2009; Vogel et al. 2014; Cheng et al. 2012; Al-Sulaiman 2013; 2014). In sco2-bc, a compressor is used to increase the pressure of the gas and after compression process; the compressed gas enters to the boiler. In the boiler the gas is heated up to about 350 C by means of the absorbed solar energy using HTF. The high pressure CO2 then expands in the turbine and enters to the heat exchanger (HEX) where it gives the rest of its heat energy to the tco2-orc. In gas cooler, the gas is cooled to 32 C before the inlet of the compressor. The tco2-orc comprises of four compounds: a turbine, an evaporator, a condenser and a pump. The required heat energy for the evaporator of the tco2- ORC is supplied from the sco2-bc. The liquid CO2 from the condenser is pumped by means of liquid pump and fed to the HEX, where it is heated by the heat energy delivered from BC, and becomes superheated vapor. The superheated vapor then enters to the turbine and expands to a low pressure. At the exit of the tco2-orc turbine, the CO2 vapor enters to recuperator for preheating of the other fluid after pumping process. Subsequently, the turbine exhaust is intensified to liquid in the condenser by extracting heat to the environment by means of a cooling tower. The tco2-orc and the CO2-VCRC are coupled together by the turbine-compressor unit. They also use the same condenser and CO2 as working fluid. The compressor of the CO2-VCRC is driven by the turbine of tco2-orc system and the CO2 is compressed to the condenser as superheated vapor. The CO2-VCRC is subcritical cycle and after the condenser, the refrigerant enters to the expansion 95 For the refrigeration processes, the coolant is 23 % ethylene glycol water (EG-water) mixture with a freezing temperature of C. Also water is used in the cooling tower for absorbing heat energy from gas cooler and condenser. The general design parameters for modelling of the power-refrigeration system are given in Table 1. It must be noted that data for PTSC system is adapted from the reported data in the references Kalogirou (2009), Singh et al. (2013) and Al-Sulaiman (2014). III. Mathematical Modelling For the thermodynamic assessment of the proposed system, energy and exergy balance equations are employed. For the modelling of PTSC system, the equations in ref Kalogirou (2009) are used. Also the given assumptions are made for the analyses: All processes are of steady state and steady flow. The changes in potential and kinetic energies are negligible. The heat transfer to/from ambient and pressure drops in the pipes are neglected. The pump and compressor operations are adiabatic and isentropic. The dead state temperature and pressure are taken to be 25 C and kpa, respectively. Tab. 1: General design parameters of the combined system Pipe receiver inner diameter 0.08 m Pipe receiver outer diameter 0.09 m Glass cover diameter 0.15 m Total length of PTSC m Mass flow rate of HTF kg/s Receiver emissivity 0.92 Glass cover emissivity 0.87 Temperature of the sun 5739 K Absorbed solar radiation 850 W/m2 Wind velocity 5 m/s Turbine isentropic efficiency 0.93 Pump isentropic efficiency 0.92 Turbine inlet temperature 350 C Compressor inlet temperature 32 C Turbine inlet pressure kpa Turbine outlet pressure 8000 kpa Net power generation 120 kw Turbine isentropic efficiency 0.88 Pump isentropic efficiency 0.96 Turbine inlet temperature 85 C Condenser temperature 28 C Turbine inlet pressure 8000 kpa Turbine outlet pressure 6892 kpa Net power generation 120 kw Evaporator capacity kw Evaporator temperature -10 C Condenser temperature 40 C Entering EG-water temperature -4 C Exiting EG-water temperature -9 C PTSC sco2-bc tco2-orc VCRC

110 For steady-state and steady-flow processes a general mass balance can be expressed in rate form as (Dincer and Rosen, 2007) m in = m out (1) where m is the mass flow rate, and the subscripts in and out stand for inlet and outlet respectively. The general energy balance can be written as: E in = E out (2) The energy balance for a general steady-flow system can also be written more explicitly as Q + m inh in = W + m outh out (3) where E in is the rate of net energy transfer to the system, E out is the rate of net energy transfer from the system, Q is the rate of net heat, W is the rate of net work, and h is the specific enthalpy. The general exergy balance equation can be written as (Dincer and Rosen, 2007) Eẋ in = Eẋ out + Eẋ dest (4) where FR is the heat removal factor, S is the is the absorbed solar energy, Aa is the unshaded collector aperture, Ar is the receiver area, and UL is the overall heat loss coefficient of the solar collector. The useful collected energy can be also calculated from entering and exiting fluid properties: Q u = m c p (T out T in ) (11) The heat removal factor can be calculated from F R = m C p A r U L [1 exp ( A ru L F m C p )] (12) where F' is the collector efficiency factor and given by F = U 0 U L (13) Here, UL is the loss coefficient of the receiver and U0 is the overall heat transfer coefficient. Since the receiver is surrounded by glass cover and the inside space is evacuated, it is assumed that there is no heat transfer by convection. Therefore, based on the receiver area Ar and glass cover area Ag, the overall collector heat loss coefficient is given by For a fixed control volume the exergy balance in steady state can be expressed as U L = [ A r ] (h c,c a +h r,c a )A g h r,r c (14) Eẋ Q Eẋ W = m ine in m oute out + T 0 S gen (5) Here, Eẋ Q and Eẋ W terms are the exergies of heat and work, respectively, e is the specific exergy, T0 is the dead state temperature and S gen is the rate of entropy generation. In Eq. 5, the terms T 0 S gen, Eẋ Q and Eẋ W are given below: Eẋ dest = T 0 S gen (6) Eẋ Q = Q ( T T 0 T ) (7) Eẋ W = W (8) The specific exergy (thermomechanical exergy or flow exergy) is defined relative to the environment (T0, P0): e = (h h 0 ) T 0 (s s 0 ) (9) where h is enthalpy, s is entropy and the subscript 0 indicates properties at the reference (dead) state. For the thermodynamic modelling of the PTSC, mathematical equations given in reference Kalogirou (2009), is used. The useful collected energy rate is defined as Q u = F R [SA a A r U L (T in T 0 )] (10) where hc,c-a is the convection heat loss coefficient between ambient and the cover, hr,c-a is the radiation heat transfer coefficient for the glass cover to the ambient and hr,r-c is the radiation heat transfer coefficient between the receiver tube and the glass cover. These three heat transfer coefficients are defined below: h c,c a = Nu air k air D g (15) Where Nu air = Re 0.52 for 0.1 < Re < 1000 (16) Nu air = 0.3 Re 0.6 for 1000 < Re < (17) Here, k is the thermal conductivity of the air, Nu is the Nusselt number and Re is the Reynolds number. h r,c a = ε g σ (T g + T a )(T g 2 + T a 2 ) (18) where is Stefan Boltzmann constant and g is the emittance of the glass cover. h r,r c = σ(t r+t g )(T r 2 +T g 2 ) 1 εr +A r Ag ( 1 εg 1) (19) Here, the subscript g refers to glass cover and r is the emittance of the receiver. The glass cover 96

111 temperature Tg can be calculated using equation below: T g = A r h r,r c T r +A g (h r,c a +h w ) T a A r h r,r c + A g (h r,c a +h w ) (20) The overall heat transfer coefficient from the surroundings to the fluid in the tube is U 0 = [ 1 + D o + ( D o ln( U L h fi D i 2k Do ) D i )] 1 (21) where Di and Do are the inside and the outside tube diameters, hfi is the heat transfer coefficient inside the tube, and k is the thermal conductivity of the tube. The equation for hfi is given below: h fi = Nu fi k fi D i (22) where Nu fi = Re 0.8 Pr 04 for Re > 2300 (23) Nu fi = (constant) for Re < 2300 (24) The exergy from the solar radiation in terms of reference and sun s temperature given by given by Petela (2005) can be expressed as Eẋ solar = S A a ( ( T 0 T sun ) ( T 0 T sun )) (25) where Eẋ solar is the function of the outer surface temperature of sun where Tsun = 5739 K (Tiwari, 2002). Finally, the exergy efficiency can be expressed as the ratio of total exergy output to total exergy input (Dincer and Rosen, 2007): η ex = Eẋ out Eẋ in = 1 Eẋ dest Eẋ in (26) IV. Results and Discussion Solar assisted CO2 power-refrigeration system was analyzed based on the model and assumptions described previously. For determining the thermodynamic performance of the systems, energy and exergy analysis are applied to the PTSC integrated power-refrigeration system. Under the assumptions made and using the solar data of Isparta, Turkey, the calculated properties of the system are given in Table 2, according to reference points illustrated in Fig. 1. Tab. 2: Calculated properties of the PTSC integrated CO2 power-refrigeration system Reference point Fluid type T ( C) P (kpa) m (kg/s) h (kj/kg) s (kj/kgk) e (kj/kg) E x (kw) 1 CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO Water Water Water Water Water Water Water EG-water EG-water Therminol-VP Therminol-VP Therminol-VP

112 During the calculations, the net power generation of sco2-bc was taken as 865 kw, the net power generation of tco2-orc was taken as 120 kw and the refrigeration capacity of CO2-VCRC was taken as kw. According to the analyses, the energy efficiency the overall system was found to be while the overall exergy efficiency was found to be %. The exergy destruction rates of the all parts of the integrated system were determined according to the exergy analyses. The results showed that, the total irreversibility of the system was calculated as 4891 kw where the PTSC system leads with the exergy destruction of 2719 kw. In Table 3, the exergy destruction rates of the all elements are given with relative irreversibility rates. Tab. 3: Exergy destruction rate and relative irreversibility s of the system components System component E x dest RI (kw) (%) PTSC Brayton comp Boiler Brayton turbine HEX Gas cooler ORC pump Rectifier ORC Turbine Condenser VCRC Compressor VCRC Expansion valve Evaporator Cooling Tower Cooling Tower pump PTSC Pump Overall cooling system In order to identify the effect of different working conditions, some parametric studies were also carried out. For these analyses, the variable parameters were selected to be solar radiation intensity, PTSC length, and net power generated. Figure 3 shows the variation of solar radiation intensity with the total exergy destruction rate. As seen from the figure, while the solar irradiation intensity increases, the exergy destruction rate decreases. Also in Figure 4, the variation of exergy efficiency with solar radiation intensity is given. With the increase of solar radiation, the exergy efficiency increases as expected. This decrement is expected, This is because the exergetic efficiency is an inverse function of exergy destruction rate as long as the exergy loss is less than the exergy destructed. Ex dest, kw hex S, kw/m 2 Fig. 3: Variation of solar radiation intensity with exergy destruction rate S, kw/m 2 Fig. 4: Variation of solar radiation intensity with exergy efficiency In Figure 5, the variation of total power generation with PTDC length is given. With the increase of collector length, the power generation increases, as well. This is because, the more collector length results in increase of area thus the heat gain from solar energy increases. This means, the heat transfer absorbs more energy with higher temperature and transfers it to the sco2-bc. A detailed information about energy analyses can be found from the reference Kizilkan and Kabul (2015). W net, kw PTSC lenght, m Fig. 5: Variation of PTSC length with power generation 98

113 It can be considered the effect of varying the net power generation on the exergy destruction rate of the overall systems considered in Figure 6. As can be seen from the figure that, as the net power generation increases, the total exergy destruction rate increases linearly. This increase in exergy destruction rate is expected because of the increment in the area of the PTSC. Ex dest, kw W net, kw Fig. 6: Variation of power generation with exergy destruction rate V. Conclusions Exergetic assessment of PTSC integrated combined power-refrigeration system was investigated comparatively using CO2. The system was consisted of sco2-bc, tco2-orc and CO2-VCRC. For the design parameters of the cycles, the net power generation of the sco2-bc was taken as 865 kw, power generation of tco2-orc was taken as 120 kw and the refrigeration capacity of CO2-VCRC was kw. According to the second law analyses, the exergy efficiency of the overall system was found to be % while the exergy destruction rate of the whole system was found to be 4891 kw. The major contributor the exergy destruction rate was determined as PTSC system because of its huge area. Also, the effects of solar radiation intensity on the exergy efficiency and exergy destruction were investigated. It was found that with the increase of the solar irradiation intensity, the exergetic efficiency of the system increased. Additionally, it was observed that the main source of exergy destruction in the integrated cycle was the PTSC system, therefore it is very important illustrate carefully the exergy destruction of the collectors. References Akbari A.D., Mahmoudi S.M.S., Thermoeconomic analysis & optimization of the combined supercritical CO2 (carbon dioxide) recompression Brayton/organic Rankine cycle, Energy, 78, , Al-Sulaiman F.A., Atif M., Performance comparison of different supercritical carbon dioxide Brayton cycles integrated with a solar power tower, Energy, 82, 61 71, Al-Sulaiman F.A. (2013), Energy and sizing analyses of parabolic trough solar collector integrated with steam and binary vapor cycles, Energy, 58, Al-Sulaiman F.A. (2014), Exergy analysis of parabolic trough solar collectors integrated with combined steam and organic Rankine cycles, Energy Conversion and Management, 77, Baronci A., Messina G., McPhail S.J., Moreno A., Numerical investigation of a MCFC (Molten Carbonate Fuel Cell) system hybridized with a supercritical CO2 Brayton cycle and compared with a bottoming Organic Rankine Cycle, Energy, 93, , Cheng, Z.D., He, Y.L, Cui, F.Q., Xu, R.J. and Tao, Y.B. (2012), Numerical simulation of a parabolic trough solar collector with nonuniform solar flux conditions by coupling FVM and MCRT method, Solar Energy, 86, Dincer I, Rosen MA. Exergy: Energy, Environment and Sustainable Development. 1st ed. Oxford: Elsevier Science; Garg P., Kumar P., Srinivasan K., A trade-off between maxima in efficiency and specific work output of super- and trans-critical CO2 Brayton cycles, The Journal of Supercritical Fluids, 98, , Hu L., Chen D., Huang Y., Li L., Cao Y., Yuan D., Wang J., Pan L., Investigation on the performance of the supercritical Brayton cycle with CO2-based binary mixture as working fluid for an energy transportation system of a nuclear reactor, Energy, 89, , Iverson B.D., Conboy T.M., Pasch J.J., Kruizenga A.M, Supercritical CO2 Brayton cycles for solarthermal energy, Applied Energy, 111, , Kim S.G., Lee Y., Ahn Y., Lee J.I., CFD aided approach to design printed circuit heat exchangers for supercritical CO2 Brayton cycle application, Annals of Nuclear Energy, 92, , Li H., Zhang Y., Zhang L., Yao M., Kruizenga A., Anderson M., PDF-based modeling on the turbulent convection heat transfer of supercritical CO2 in the printed circuit heat exchangers for the supercritical CO2 Brayton cycle, International Journal of Heat and Mass Transfer, 98, , Kalogirou S.A., Solar Energy Engineering: Processes and Systems, Academic Press, Oxford, UK, Kizilkan O., Kabul A., Design and Energy Modelling of a Solar Driven Combined Power-Refrigeration System with Super-Trans-Sub Critical Cycles using CO2, The 2015 World Congress on Advances in Aeronautics, Nano, Bio, Robotics, and Energy (ANBRE15), Incheon, Korea, August 2015.

114 Kouta A., Al-Sulaiman F., Atif M., Marshad S.B., Entropy, exergy, and cost analyses of solar driven cogeneration systems using supercritical CO2 Brayton cycles and MEE-TVC desalination system, Energy Conversion and Management, 115, , Kumar K.R. and Reddy K.S. (2009). Thermal analysis of solar parabolic trough with porous disc receiver, Applied Energy, 86, Linares J.I., Cantizano A., Moratilla B.Y., Palacios V.M., Batet L., Supercritical CO2 Brayton power cycles for DEMO (demonstration power plant) fusion reactor based on dual coolant lithium lead blanket, Energy, 98, , Singh, R., Kearney, M.P. and Manzie, C. (2013), Extremum-seeking control of a supercritical carbondioxide closed Brayton cycle in a direct-heated solar thermal power plant, Energy, 60, Therminol (2014), Therminol VP-1, Heat transfer fluids by Eastman, accessed on Tiwari GN. Solar Energy: Fundamentals, Design, Modelling and Applications. Pangbourne: Alpha Science International Ltd.;2002. Vogel, T., Oeljeklaus, G., Görner, K., Dersch, J. and Polklas, T. (2014), Hybridization of parabolic trough power plants with natural gas, Energy Procedia, 49, Liu X., Gong G., Wu, Y., Li H., Thermal performance analysis of Brayton cycle with waste heat recovery boiler for diesel engines of offshore oil production facilities, Applied Thermal Engineering, In Press, Mecheri M., Moullec, Y.L., Supercritical CO2 Brayton cycles for coal-fired power plants, Energy, 103, , Manente G., Lazzaretto A., Innovative biomass to power conversion systems based on cascaded supercritical CO2 Brayton cycles, Biomass and Bioenergy, 69, , Padilla R.V., Too Y.C.S., Benito R., Stein W., Exergetic analysis of supercritical CO2 Brayton cycles integrated with solar central receivers, Applied Energy, 148, , Padilla R.V., Too Y.C.S., Benito R., McNaughton R., Stein W., Thermodynamic feasibility of alternative supercritical CO2Brayton cycles integrated with an ejector, Applied Energy, 169, 49 62, Petela R. Exergy analysis of the solar cylindricalparabolic cooker. Solar Energy 2005;79: Rovense F., A Case of Study of a Concentrating Solar Power Plant with Unfired Joule-Brayton Cycle, Energy Procedia, 82, , Rovira A., Muñoz M., Sánchez C., Val J.M.M., Proposal and study of a balanced hybrid Rankine Brayton cycle for low-to-moderate temperature solar power plants, Energy, 89, , Serrano I.P., Linares J.I., Cantizano A., Moratilla B.Y., Enhanced arrangement for recuperators in supercritical CO2Brayton power cycle for energy conversion in fusion reactors, Fusion Engineering and Design, 89, ,

115 Cooling of Concentrated Photovoltaic System Using Microchannel Heat Sink Ali Radwan 1*, Mahmoud Ahmed 1, Shinichi Ookawara 1,2 1 Department of Energy Recourses Engineering, Egypt-Japan University of Science and Technology (E-JUST), Egypt. 2 Tokyo Institute of Technology, Tokyo, Japan. * Abstract The high incident heat flux on the concentrated photovoltaic (CPV) system causes a significant increase in the cell temperature and thus reduces the system efficiency. Therefore, using an efficient cooling technique is a vital issue for those systems. In the present study, a new cooling method for CPV systems is introduced. A wide microchannel and a manifold microchannel are introduced as heat sinks for those of high thermal generated systems. A comprehensive thermo-fluid model is developed that includes the whole layers of photovoltaic cell integrated with the introduced microchannel domain. The developed model is numerically simulated and validated using various sets of the previous experimental, numerical and analytical results. In the manifold microchannel heat sink, the effect of the manifold pitch is investigated. Based on the simulation results, it is found that decreasing the manifold pitch distance enhances the solar cell temperature uniformity and avoids the hot spot formation. The manifold microchannel heat sink achieves a better solar cell temperature uniformity compared with the conventional wide one at lower mass flowrates lower than 8.3 g/s. In addition, the manifold microchannel heat sink consumes less pumping power than the wide microchannel heat sink. At CR=40, the wide microchannel heat sink is suitable to achieve the highest gained net electric power at cooling flow rate from 8.3g/s to 71.1 g/s. However, if the cooling mass flow rate islower than 8.3 g/s or greater than 71.1 g/s, the manifold microchannel heat sink is producing a higher net electrical power than the wide microchannel heat sink. The differences between the maximum and minimum solar cell temperature are about15 o C and 65 o C at CR=10 and 40 respectively for the wide MCHS, while these differences are about 0.3 o C and 0.8 o C at CR=10 and 40 respectively in the case of using manifold microchannel heat sink. Keywords: Heat transfer, concentrated photovoltaic, manifold-microchannel, solar cell efficiency I. Introduction In sunlight concentration onto photovoltaic (PV) cells, the replacement of expensive solar cells area with low-priced concentrating lenses or mirrors is one method to lower the cost of solar electricity. Because of the reduction in the solar cell area, higher PV cells efficiency may be used (Royne et al., 2005; Zelin Xu and Kleinstreuer, 2014). In the meantime, it is mentioned that PV cells efficiency is less than 20% for silicon solar cells and around 40% for multi-junction solar cells (Zelin Xu and Kleinstreuer, 2014). The remainder part of solar energy is converted into the heat causing temperature rise in PV cells. The generated thermal energy in the PV systems might cause junction damages and lead to a significant decrease in its electrical efficiency (Rejeb et al., 2015; Zelin Xu and Kleinstreuer, 2014; Zhao et al., 2011). Therefore, using an efficient cooling technique in concentrated photovoltaic (CPV) systems will achieve a high electrical efficiency and allow designing the high concentration ratio (CR) systems. In addition, the extracted thermal energy could be used for any domestic or industrial application in photovoltaic/ thermal (PV/T) system. The major design considerations for cooling of concentrated photovoltaic cells are the solar cell temperature, temperature uniformity, and the consumed pumping power (Rejeb et al., 2015; Zelin Xu and Kleinstreuer, 2014; Zhao et al., 2011). The rapid development of the Micro-Electro-Mechanical Systems (MEMS) makes it possible to construct very small scale cooling devices. These small-scale devices can dissipate a large amount of heat flux from hot surfaces(kalteh et al., 2011). So the use of microchannel heat sink (MCHS) is an effective technique to limit the temperature of those high generated heat flux areas such as CPV systems (Rosell et al., 2011). Long parallel microchannels and wide microchannel heat sinks (WMCHS) are described as the conventional microchannel heat sinks. Such heat sinks have been successfully examined for the use in electronic devices cooling applications. Numerous numerical and analytical models for predicting the heat transfer characteristics and pressure drop through such cooling systems have been investigated. Although the conventional microchannel heat sinks offer a significant heat transfer augmentation, they are associated with a dramatic pressure loss. Alternative configurations had been proposed to decrease the incurred pressure loss and simultaneously increase the heat transfer. One of those configurations is the manifold microchannel 101

116 heat sink (MMCHS) (Sarangi et al., 2014). The MMCHS consists of a manifold system which distributes the cooling fluid via multiple inlet-outlet ports. By reducing the flow length through the microchannels, a significant reduction in the pressure drop was attained. Additionally, a decrease in the thermal resistance was achieved by interrupting the thermal boundary layers growth. This design was originally suggested by (Harpole and Eninger, 1991), who confirmed a significant enhancement in the heat transfer coefficient relative to the conventional MCHS at a constant pumping power. A complete twodimensional thermo-fluid model of the MMCHS has been developed [8]. They concluded that distribution of manifold channel spacing should be 333µm, i.e. (30 channels/cm). Also, they found that by using the mentioned spacing, a small rise in the lateral surface temperature non-uniformities is achieved. Rahimi (Rahimi et al., 2013) experimentally studied the performance of the combination of micro-channels and a photovoltaic module as a hybrid PV/T system using water as a coolant. In their experiments, the microchannel hydraulic diameter is mm, and the Reynolds number (Re) varies up to 70. They reported that approximately 30% increased in the output power compared to uncooled conditions. Reddy et al. (Reddy et al., 2014) concluded, based on numerical simulation, that the optimum dimensions of the microchannel were 0.5 mm width and aspect ratio of 8. Moreover, the pressure drop was found to be low in straight flow channels. Bladimir et al. (Ramos- Alvarado et al., 2011) numerically calculated the pressure loss and temperature uniformity of the heated walls of different proposed microchannel configurations. They suggested a new design to achieve a smaller pressure drop and a better flow and temperature uniformity. They recommended using microchannel distributors for cooling the concentrated PV cells, fuel cells, and electronics. The main findings of the recent literature in the field of cooling CPV or PV systems can be summarized as follows: (1) until now the minimum studied channel depth was 500 µm, 5000µm in (Agrawal and Tiwari, 2011),1200µm in(ramos-alvarado et al., 2011),500 µm in (Rahimi et al., 2013), and 220 µm in (Yang and Zuo, 2015). These dimensions are not in agreement with the basis of microchannel dimensions which should be in the range from 10 to 200 µm (Kandlikar, 2014); (2) In the most recent theoretical modeling of the PV/T systems, one-dimensional model is the most popular one (3) most of the recent microchannel studies are investigated in cooling method of the integrated circuits and electronic devices, while a few studies was performed in the concentrated photovoltaic systems. Based on the recent summary, the objective of the current work is to compare the performance of CPV system integrated with two different designs of microchannels such as a manifold microchannel (MMCHS) and a wide microchannel (WMCHS) with channel height of 100 µm. A comprehensive 102 conjugate two-dimensional thermo-fluid model including the PV layers and the microchannel is developed. The developed model is more suitable for the new investigated MMCHS with unknown thermal performance. Also, the proposed numerical model is able to present the variation of temperature which can t be obtained by the conventional energy balance model. The developed model is numerically solved and validated with recent available numerical, analytical, and experimental data. II. Physical model In the present study, a concentrated photovoltaic cell with concentration ratio up to 40 is investigated. The concentration ratio is reached by using the Fresnel lens solar concentrator which has the minimum capital cost (Whitfield et al., 1999). The typical dimensions of the solar cell unit are 12.5 cm by 12.5 cm. In the current work, the effective area of the photovoltaic cell is divided into four equal areas of 6.25 cm by 6.25 cm. Consequently, four microchannel heat sinks of 6.25 x 6.25 cm 2 are used for a single solar cell unit to ensure better temperature uniformity and lower friction power. A schematic diagram of the WMCHS and MMCHS considered in the present work is shown in Fig.1. The manifold distribution system is placed on the top of the flat microchannels, in a direction transverse to the main flow direction. The coolant is pumped in through a common inlet header, which branches out into parallel manifold inlet channels. Upon entering the microchannel, the fluid undergoes a 90 o turn and passes through a distance of the microchannel midpitch distance P/2, removing the generated heat from the concentrated solar cell, and subsequently flows through another 90 o turn then exits upward through the outlet manifold channels. Another common outlet header is used to collect the outlet flow rate. In MMCHS, the pitch (P) is defined as the distance between two consecutive inlets or outlet ports. The inlet and the outlet ports of the manifold have the same cross-sectional area and the same number for the same pitch value. The pitch value is set to have different values of 2547, , 837.6, and µm to achieve a specific number of inlet ports of 25, 50, 75, and 100 respectively. The variation of the pitch is investigated to achieve the best effect on the solar cell temperature, reduce the friction power, and attain a temperature uniformity of the solar cell. In fig. 2-A and B, the PV layers rest on the designed microchannel heat sink. Water is flowing through the channel to maintain a high efficiency and avoid excessive cell temperatures. A constant sun radiation is assumed to be 1000W/m 2, and the incident solar heat flux increases according to the value of geometrical CR.

117 (A) Vw Ta Concentrated thermal radiation Glass cover Solar Cell Tedlar Cooling Channel configurations, the microchannel wall thickness, channel height, channel material, and cooling mass flow rate are selected to be the same. The thermophysical properties of the solar cell layers and microchannel wall materials are presented in Table 1. δg δsc δt δins Lsc H Insulation Table. 1: Thermophysical properties of the microchannel material and solar cell layers Material ρ (kg/m 3 ) Cp (J/kg.K) K (W/m.K) Glass Silicon Tedlar Aluminium (B) Concentrated thermal radiation II.1. The governing equations and numerical simulation δg δsc δt H (A) (B) Vw Ta p outlet port Lsc Glass cover Solar Cell Tedlar Cooling Channel Fig.1 a neat sketch of the proposed PV layers integrated with (A) WMCHS, (B) MMCHS Inlet y x q" w H q" w Solar Cell Tedlar Channel wall δ sc δ t δ w Outlet Two-dimensional solid fluid conjugate heat transfer model is developed to estimate the electrical and thermal performance of the CPV system. In the current study, the width of the MCHS (Wsc) is very large compared to the microchannel height (H). Hence, the two dimensions assumption is appropriate. The developed model adopts the following assumptions: 1. The flow in microchannel heat sink is laminar, incompressible, and steady. 2. The fluid properties are temperature dependent due to the effect of high concentrated solar radiation, while the properties of the solar cell, tedlar, and microchannel walls are assumed to be temperature independent. 3. The effect of viscous dissipation term in the energy equation is neglected. 4. For the investigated MCHS configurations, the back side is perfectly insulated. 5. The thermal contact resistances among each layer of the solar cell and microchannel heat sink are neglected. The comprehensive thermal model includes energy balance, and thermo-fluid equations. Applying energy balance equations on the PV module, the total absorbed energy by the PV cell can be written as: Solar Cell δ sc E g G(t) (1) sc sc sc y P Tedlar Channel wall H δ t δ w The total absorbed energy by the tedlar can be written as: 1 ( t) E g G (2) T sc T x Outlet Inlet Fig. 2 computational domain for (A) WMCHS-PV system, (B) MMCHS-PV system A comparison of the CPV system performance integrated with the proposed MMCHS and conventional WMCHS is presented. In the presented 103 Consequently, the total absorbed energy by the solar cell and its tedlar can be written as follows (G. N. Tiwari and Swapnil Dubey, 2010; Zelin Xu and Kleinstreuer, 2014): 1 G( t) G( t) E (3) sc T g sc sc The electrical energy generated by solar cell can be written as: g

118 E g G(t) (4) el sc sc Where sc is the solar cell efficiency (G. N. Tiwari and Swapnil Dubey, 2010; Zelin Xu and Kleinstreuer, 2014) and can be written as follows: 1 ( T T )) (5) sc ref ( ref sc ref Where: the ref and β ref are the solar cell efficiency and cell temperature coefficient at a reference temperature (T ref =25 o C) respectively. The reference solar radiation is G=1000W/m 2 (Hedayatizadeh et al., 2013). As reported earlier, part of the total absorbed solar energy (E) is converted into electricity (Eel) in the solar cell according to its efficiency. The second part of absorbed solar energy is lost from the top (Et) of the solar cell to the environment by the effect of the wind and radiation loss. The rest is conducting through the solar cell to the microchannel heat sink. This part of solar energy causes a temperature rise in the solar cell and reduces the electrical efficiency. Hence, the amount of thermal energy passes through the solar cell can be estimated according to the following relation: q' ' E E E (6) w el t Where the top loss (Et) of thermal energy due to the effect of the wind speed and radiation can be calculated by (Zelin Xu and Kleinstreuer, 2014). 1 1 Et Ut ( Tsc Ta) (7) g U t K h g conv (8) 2008). The fluid flow and heat transfer governing equations are described for the fluid and solid domains as follows (Lelea and Laza, 2014): Continuity equation: ( u) ( v ) 0.0 x y Momentum equations in x and y-direction u ( u) P u u u v ( ) ( ) x y x x x y y v u x ( v) P v v v ( ) ( ) y y x x y y Energy equation (Dehghan et al., 2015) CuT CvT T T ( Kl ) ( Kl ) x y x x y y (12) (13) (14) (15) Energy equation for solids T T ( K s ) ( K s ) 0 (16) x x y y Where: K l and K s are the liquid and solid material thermal conductivity. The solid energy equation is applied to the solar cell, tedlar, and the microchannel heat sink material. In the current study, the solar concentration ratio is varied from 10 to 40. While the cooling fluid mass flow rate ranged from 4.47g/s to g/s which is equivalent to Re=10 to 400 for the wide microchannel configuration respectively where Reynolds number based on twice of microchannel height as a hydraulic diameter(kandlikar, 2014). In addition, the wind speed and ambient temperatures are 1m/s and 30 o C respectively through the all simulation results. The top convection heat transfer coefficient from the glass surface to the ambient,(hconv), is depending on wind speed (Vw) (Agrawal and Tiwari, 2011) as follows: h (9) conv V w The useful thermal energy (Eu), friction power, and Reynolds number are calculated according to the following equations: E P u m. C ( T (10) friction f P. m f f T f, out f, in f ) (11) Fluid temperature significantly changes inside the microchannel, especially at high CR values. Hence, the variation in the fluid thermophysical properties is substantial. It is worth mentioning that in the present calculations, the variation of the thermophysical properties of cooling water is considered using the polynomial equations reported in (Jayakumar et al., 104 II.2. Boundary conditions At the inlet the fluid velocity is identified and assumed to be uniform and it is varied according to the value of mass flow rate. In addition, a uniform inlet temperature is assumed. In the meantime, the outlet flow boundary condition is identified at the microchannel outlets. Water molecular mean free path is about 0.25 nm so, Knudsen number falls in the no-slip regime (Kn< 0.001) (Dehghan et al., 2015). Consequently, no-slip and no temperature jump boundary conditions are taken at the interface between the solid fluid domains. The thermal boundary condition for the upper wall of the solar cell is a constant heat flux which is calculated by Eq. 6. Here, it is clear that the upper wall heat flux is a function of the solar cell efficiency which is dependent on the solar cell temperature. Therefore, an iterative technique is applied to calculate the actual incident heat flux. Finally, the lower and side walls of the computational domain are assumed to be adiabatic. In more details, for a flat microchannel domain, the applied boundary conditions are indicated as follows:-

119 1. At the channel inlet (x=0) 1.1. For fluid domain at 0 y H u = Uin, v =0, T = Tin=303 K 1.2. For solid domains (Microchannel walls, tedlar, and solar cell layer) T k 0 s x 2. At the channel outlet x= Lsc 2.1. For fluid domain 0 y H v u T 0, 0, and 0 x x x 2.2. For solid domains (Microchannel walls, tedlar, and solar cell layer) T k s 0 x 3. Upper wall y= (δw +H +δsc+δt), and 0 x Lsc T k '' sc q f ( T sc ) y w 4. Lower wall at y=0 and 0 x Lsc T k ch 0 y 5. For fluid-solid interface: u= v = 0, and k f Tf kchtch 6. For solid-solid interfaces: (a) Solar cell-tedlar interface K sc T sc = K T T T (b) the interface between the tedlar and the microchannel K ch T ch = K T T T Table 2: Optical and physical characteristics of PV cell and MCHS design parameters. Parameter Value, µm Parameter Value L sc δ g H 100 erf δ Sc β ref β sc τ g δ ch 200 α sc δ w 200 α T δ T ε g Pitch (A) P= 2547 µm, N= 4/cm (B) P=1260 µm, N= 8 /cm (C) P= 837 µm, N=12/cm (D) P= 627 µm, N=16/cm 1, (Hedayatizadeh et al., 2013); 2, (Zhou et al., 2015); 3, (Zelin Xu and Kleinstreuer, 2014); 4, (Z. Xu and Kleinstreuer, 2014) II.3. Solution methods and convergence criteria The governing equations along with the described boundary conditions are solved using the commercial CFD package ANSYS FLUENT 14.5 ( ANSYS FLUENT 14.5 Theory Guide) based on discretization using the finite volume method. Pressure-velocity coupling is addressed through the SIMPLE algorithm, along with an algebraic multigrid algorithm (AMG) for solving the linearized system of discretised equations. In addition, the average pressure at the inlet and the average temperature of the solar cell are also monitored to check for convergence of the flow and energy equations. A non-uniform grid arrangement with a large number of grid points near the channel walls is arranged to resolve fluid flow and heat transfer with a consideration of the effect of the boundary layer flow.the resulting system of algebraic equations is solved using the Gauss Seidel iterative technique. The computational domain dimensions, the solar cell optical properties, and layers dimensions are reported in Table 2. The numerical code is verified in some ways to ensure the validity. For every microchannel heat sink, a grid independence test is conducted using several different mesh sizes. II.4. Mesh independence test The mesh independence study is performed for each dimension of the computational domain. The first one is for a cooled CPV system using WMCHS. A different number of cells of , , , and are tested. It is found that there no significant change in the solar cell temperature and the outlet water temperature with further increase of cell numbers after cells. Accordingly, the cell numbers of are selected for the simulation of the WMCHS. The test is established four times for the MMCHS with the studied four-pitch value. It is found that the suitable number of cells are , , , and respectively II.5. Numerical results validation. The results of the present numerical simulation of the microchannel fluid domain are validated using different sets of the available experimental, numerical and theoretical data. The comparison between the predicted fully developed Nu number values with the analytical results (Rohsenow and Hartnett, 1998) is established at different applied Boundary conditions. Table 3 illustrates the comparison between predicted fully developed Nu number and the corresponding analytical values at different boundary conditions in the case of using water as a working fluid. In this comparison, the following three types of boundary conditions are applied: (i) both walls at constant heat flux; (ii) both walls at constant wall temperature; (iii) one wall is adiabatic and the other wall is heated at constant heat flux. Based on Table 3, the maximum error between numerically predicted values and analytical results of Nu number does not exceed 0.65%. Tab. 3. Fully developed Nu number validation Nu fully developed on the heated wall, water, Re=100 B.C CFD Analytical* %Error q w1,2=const % T w1,2=const % q w1=const., q w2= % *(Kandlikar, 2014) 105

120 Where: the numbers 1 and 2 refer to the upper and the lower wall of the wide microchannel. qw and Tw are the constant wall heat flux and constant wall temperature boundary condition respectively. Figure 3 presents the comparison between the current predicted Nusselt number and those measured [30] and predicted using Lattice Boltzmann method, (LBM) [28] at different values of Reynolds numbers. A good agreement between the predicted and measured Nusselt number is found. Also, the friction factor is validated with the analytical results of the fully developed laminar flow friction characteristics presented in (Kandlikar,2014; Rohsenow and Hartnett, 1998) in the case of wide microchannels. An excellent agreement between the predicted friction factor and the analytical results is obtained. inlet ports is 2547µm and the computational domain length is µm. So the total numbers of inlet ports are four per one cm. Similarly, in the case (B), (C), and (D), the pitch values are 1260, 837, and 624 µm which gives a 8, 12, and 16 inlet ports per cm respectively. The same total mass flow rate is used for each case, and CR is selected to be 40. The comparison between the four investigated pitches is implemented based on the average solar cell temperature, cell temperature uniformity, and the consumed pumping power. Figure 4 indicates the variation of the average solar cell temperature under the influence of the cooling fluid mass flow rate. Generally, it is found that increasing the cooling mass flow rate leads to reduce the solar cell temperature. This trend was observed by several researchers (Baloch et al., 2015) and (Du et al., 2012) using water and (Z. Xu and Kleinstreuer, 2014) using 5% vol. Al2O3-water nanofluid. There are different points of view to interpret the reason for this trend. It is reported that at a lower Re range, the heat transfer mechanism between the upper wall and cooling fluid is dominated by convection, while at a higher Re range, the heat transfer mechanism is dominated by conduction within the thin layer of the laminar wall region (Z. Xu and Kleinstreuer, 2014). Another point of view relates this trend to the reduction of contact time between the fluid and the upper wall due to higher velocity associated with higher flow rate (Baloch et al., 2015). The last interpretation is that at high Re, the heat extracted by the cooling water reaches the saturated level and therefore, the cell temperature slightly increases (Du et al., 2012). Fig. 3 Comparison between the predicted Nu number and those measured results of (Kalteh et al., 2012) and previous predicted of the author using twophase Lattice Boltzmann method (Ahmed and Eslamian, 2015). Also, the predicted friction factor and the analytical results (Rohsenow and Hartnett, 1998, Kandlikar, 2014) for flow between two parallel plates. III. Results and discussion In this section local and average solar cell temperature, electrical efficiency, pumping power and temperature uniformity will be discussed for the investigated configurations. In the first part, the effect of the manifold pitch on the average solar cell temperature, temperature uniformity, and the friction power will be discussed. Then the best manifold pitch design is compared with the conventional. The comparison is presented at a mass flow rate ranged from 4.47g/s to g/s and the solar concentration ratio (CR) from 10 to 40. III.1. Effect of the manifold pitch Different values of the manifold pitch are investigated. In case (A), the distance between the two consecutive 106 Fig.4. Variation of the average solar cell temperature with the cooling fluid mass flow rate at the various manifold pitch Regarding to the pitch effect, it is found that decreasing the pitch reduces the solar cell temperature especially at lower flow rates as it is clear in the figure. This is may be attributed to the reduction of the fluid pass under the heated CPV cell. Consequently, the flow might be in the entrance

121 region. The heat transfer in the entrance region is significantly higher than that of fully developed regime. This will cause a better reduction in the solar cell temperature as shown in the case in the case of the lowest pitch (D). The solar cell temperature reduces from o C in the case of pitch (A) to 86 o C for the case of using pitch (D) at the same cooling mass flow rate and solar concentration ratio of 4.47g/s and 40 respectively. In addition, further increase of the mass flow rate beyond 50g/s leads to no significant effect on the solar cell temperature. The reason might be due to the fully developed flow regime. Figure 5 shows the variation of the local solar cell temperature with the lateral distance for the studied pitch values. It is clear that at a lower mass flow rate and the highest solar concentration ratio, increasing the manifold pitch achieves non-uniformity in the solar cell. Moreover, the case (D) achieves better solar cell temperature uniformity. The temperature uniformity is measured by the difference between the minimum and maximum local solar cell temperature MMCHS associated with different number of inlet ports per cm. Results indicated that the optimum number of inlet and outlet ports per cm is found to be 30 per cm. which is in agreement with the results obtained in (Harpole and Eninger, 1991). Figures 7-a and b show the variation of the average solar cell temperature versus the cooling fluid mass flow rate for both WMCHS and MMCH. Generally, it is found that increasing the cooling fluid mass flow rate leads to enhance the average solar cell temperature until a certain limit. Further increasing of mass flow rate attains a slight enhancement in the solar cell temperature. Additionally, increasing the solar concentration ratio leads to the increase of the solar cell temperature due to the rise in the incident heat flux. Figure 6 presents the variation of the pumping power versus the cooling rate at different values of the manifold pitches. It s found that the configuration (A) consumed the highest pumping power. The reason for such trend is probably due to two factors. The first one is that increasing the pitch tends to increase the fluid path length and hence increase the pressure drop. The second one is that reducing the pitch decreases the flow velocity due to increasing the number of inlet ports at the same mass flow rate. To conclude, from the figs 4, 5 and 6, it is clear that increasing the manifold pitch number enhances the solar cell temperature and the temperature uniformity and reduces the consumed pumping power. Fig.6. Variation of the friction power per unit width with the cooling flow rate at various manifold pitches By comparing between 7-a, and 7-b, MMCHS achieve lower average solar cell temperature especially at lower mass flow rates, while increasing the mass flow rates the WMCHS achieve a relatively lower solar cell temperature. The reason for such trend is that at a lower mass flow rate, the heat transfer coefficient for MMCHS is higher than that of WMCHS. At a higher mass flow rate, the heat transfer coefficient is approximately close for both configurations. However, there are stagnation flow points that cause hot spots on the lower side of the microchannel walls due to the flow turning in the MMCHS. Accordingly, it leads to the increase of the cell temperature. Fig.5. Variation of the solar cell temperature with the solar cell lateral length at various manifold pitches III.2.Comparison between the MMCHS and WMCHS The conventional WMCHS is compared with the 107 Comparison of the solar cell temperature uniformity is presented in Fig.8 A and B at a mass flow rates of 4.74g/s and g/s respectively. In both figures, the local solar cell temperature increases with the increase of the WMCHS axial distance while for the MMCHS, the local solar cell temperature is nearly constant along the solar cell length. For instance, the difference between the maximum local solar cell temperature and the lower local solar cell temperature is about 15 o C and 65 o C at CR=10 and 40 respectively

122 for the WMCHS. However, in the case of MMCHS, the maximum local solar cell temperature difference is about 0.3 o C and 0.8 o C at CR=10 and 40 respectively. Regarding the gained electric power, it is found that the solar cell power integrated with the WMCHS is greater than that for MMCHS especially at a higher mass flow rate as shown in Fig.9. This is due to the fact that the solar cell power is directly proportional to its efficiency which is proportional to the solar cell temperature. Moreover, the solar cell temperature is lower in the case of WMCHS than MMCHS at a higher mass flow rate as discussed in Fig. 7. q"w Solar Cell δsc also until it reaches a nearly constant solar cell power after 50g/s. Further increase in the cooling fluid mass flow rate leads to a significant rise in the friction power while remaining the solar cell power as a constant value. So the net power which is defined as the difference between the solar cell electric power and the friction power will increase and then it decreases again. The same trend also observed in (Z. Xu and Kleinstreuer, 2014). Thus in the case of WMCHS, further increase in the cooling fluid mass flow rate beyond 20g/s is not effective at CR=10. Finally, there is an optimum value of mass flow rate which will attain a maximum net output power. This value is dependent on the solar concentration ratio. It is found that the mass flow rate of 25, 35, and 45 g/s are the best suitable for the CPV systems operating with a solar concentration ratio of 20, 30, and 40 respectively. Tedlar δt Inlet y x H Channel wall δw Outlet q"w Solar Cell δsc Tedlar δt Channel wall δw y P H x Inlet Outlet Fig. 7 Variation of the average solar cell temperature with the cooling fluid mass flow rate at various CR values for (A) WMCHS and (B) MMCHS with P=627.2µm It is well known that using microchannel heat sink will significantly increase the friction power loss due to its very small size. Figure 10 shows the variation of the solar cell electric power and both the consumed friction power and the net gained electric power versus the mass flow rate at CR=10. It is clear that for WMCHS, the friction power increase with an increase in the flow rates. While the solar cell power increase 108 Fig.8 the Variation of the local solar cell temperature with the axial distance at cooling fluid mass flow rate of (A) 4.74g/s and (b) 189.6g/s

123 The comparison of the net gained electric power for the CPV systems integrated with the MMCHS and WMCHS is presented in Figs.11 A and B for CR=10 and 40 respectively. It is found that for MMCHS, the net gained power increases with the increase of the cooling fluid mass flow rate and takes the expected trend of the solar cell efficiency. This is due to that the consumed friction power is the same with the gained power. However, in the case of WMCHS, the variation of the net gained electric power is varied as discussed in Fig.10. technique for the solar concentration ratio of 10. On the other hand, at CR=40 as shown in Fig 11 -B, the WMCHS is suitable to achieve highest gained net electric power for a cooling fluid mass flow rate ranged from 8.3 to71.1 g/s. Fig.9. Variation of the solar cell electric power versus the cooling fluid mass flow rate for MMCHS and WMCHS at CR=10 Fig. 10. Variation of the friction power, net power and the solar cell power with the cooling fluid mass flowrate at CR=10 In Fig. 11-A, when the cooling fluid mass flow rate is greater than 27.7g/s, MMCHS is preferable to achieve a higher net gained power than WMCHS. If the pumping system is not capable of delivering a cooling mass flow rate greater than 27.7g/s per meter width of channel, hence the WMCHS is the suitable cooling 109 Fig.11 variation of the net gained electric power with the cooling fluid mass flow rate at (A) CR=10 and (B) CR=40 IV. Conclusions A comprehensive 2D- mathematical model has been developed to predict the performance of the CPV systems using two different designs of manifold microchannel heat sinks (MMCHS) and conventional wide microchannel heat sink (WMCHS). For the MMCHS, the manifold pitch effect is numerically investigated. It is found that the manifold pitch of µm is most appropriate for the cooling of CPV system. Such proposed manifold pitch achieves the lowest solar cell temperature, best temperature uniformity of the solar cell and the minimum friction power. Furthermore, using MMCHS attains better

124 temperature uniformity than the WMCHS. The difference between the maximum local solar cell temperature and the lower local solar cell temperature is about 15 o C and 65 o C at CR=10 and 40 respectively for the WMCHS. However, when using MMCHS, the maximum local solar cell temperature difference is about 0.3 o C and 0.8 o C at CR=10 and 40 respectively. Furthermore, at the cooling fluid mass flow rate greater than 27.7g/s, MMCHS is achieving a higher net gained power than WMCHS. If the cooling mass flow rate is greater than 27.7g/s per meter, the WMCHS is providing a hiher net gain power at solar concentration ratio of 10. Acknowledgement The author would like to thank the Egyptian government especially Ministry of Higher Education (MoHE). We also would like to express our gratitude to Egypt-Japan University of Science and Technology (E-JUST) for offering the facilities and tools. Nomenclature Cp : Specific heat of cooling fluid (J.kg -1.K) G(t) : Incident solar radiation (W.m -2 ) h : Heat transfer coefficient (W.m -2.K) H : microchannel height (m) k : Thermal conductivity (W.m -1.K) L : Microchannel length, solar cell length (m) m : Unit cooling fluid mass flow rate (Kg.s -1 ) N : Number of inlet or outlet ports Nu : Nusselt number Nu=h.Dh /Kf P : Pressure (Pa), electrical power (W) T : Temperature ( o C) T : Temperature ( o C). Ut : Overall heat transfer coefficient from the top surface of solar cell to ambient (W.m -2.K) V : Velocity vector (m.s -1 ). Vw : Wind velocity (m.s -1 ). Greek symbols α : Absorptivity β : Backing factor and solar cell temperature coefficient (K -1 ) ε : Emissivity τ : Transmittivity µ : Fluid viscosity (Pa.s) σ : Stephan-Boltzmann constant 5.67*10-8 (W. m - 2.K 4 ) ρ : Fluid density (kg.m -3 ) δ : Thickness (m) : Solar cell and thermal efficiency Subscripts a : Ambient conv. : convection eff : Effective el : Electrical f : fluid g : Glass in : Inlet out : Outlet ref : Reference condition, G=10 3 W.m -2,T=25 o C 110 Sc Sc, x T th w w References : Solar cell : Local solar cell : Tedlar : Thermal : Wall : water Agrawal, S., Tiwari, A., Experimental validation of glazed hybrid micro-channel solar cell thermal tile. Sol. Energy 85, (2011). Ahmed, M., Eslamian, M. Laminar forced convection of a nanofluid in a microchannel: Effect of flow inertia and external forces on heat transfer and fluid flow characteristics. Appl. Therm. Eng. 78, (2015). ANSYS FLUENT 14.5 Theory Guide [WWW Document], n.d. URL project/neptunius/docs/fluent/html/th/node322.htm (accessed ). Baloch, A. a. B., Bahaidarah, H.M.S., Gandhidasan, P., Al-Sulaiman, F. a.. Experimental and numerical performance analysis of a converging channel heat exchanger for PV cooling. Energy Convers. Manag. 103, 14 27(2015). Dehghan, M., Daneshipour, M., Valipour, M.S., Rafee, R., Saedodin, S.. Enhancing heat transfer in microchannel heat sinks using converging flow passages. Energy Convers. Manag. 92, (2015). Du, B., Hu, E., Kolhe, M.. Performance analysis of water cooled concentrated photovoltaic (CPV) system. Renew. Sustain. Energy Rev. 16, (2012). G. N. Tiwari and Swapnil Dubey, Fundamentals of Photovoltaic Modules and Their Applications. The Royal Society of Chemistry, (2010). Harpole, G.M., Eninger, J.E.. Micro-channel heat exchanger optimization, in: 1991 Proceedings, Seventh IEEE Semiconductor Thermal Measurement and Management Symposium. IEEE, pp (1991). Hedayatizadeh, M., Ajabshirchi, Y., Sarhaddi, F., Safavinejad, A., Farahat, S., Chaji, H., Thermal and Electrical Assessment of an Integrated Solar Photovoltaic Thermal (PV/T) Water Collector Equipped with a Compound Parabolic Concentrator (CPC). Int. J. Green Energy 10, (2013). Jayakumar, J.S., Mahajani, S.M., Mandal, J.C., Vijayan, P.K., Bhoi, R., Experimental and CFD estimation of heat transfer in helically coiled heat exchangers. Chem. Eng. Res. Des. 86, (2008).

125 Kalteh, M., Abbassi, A., Saffar-avval, M., Frijns, A., Darhuber, A., Experimental and numerical investigation of nano fl uid forced convection inside a wide microchannel heat sink. Appl. Therm. Eng. 36, (2012). Kalteh, M., Abbassi, A., Saffar-Avval, M., Harting, J., Eulerian Eulerian two-phase numerical simulation of nanofluid laminar forced convection in a microchannel. Int. J. Heat Fluid Flow 32, (2011). Kandlikar, S.G., Heat Transfer and Fluid Flow in Minichannels and Microchannels,(2014). Lelea, D., Laza, I., The water based Al2O3 nanofluid flow and heat transfer in tangential microtube heat sink with multiple inlets. Int. J. Heat Mass Transf. 69, ,( 2014). Rahimi, M., Karimi, E., Asadi, M., Valeh-e-Sheyda, P. Heat transfer augmentation in a hybrid microchannel solar cell. Int. Commun. Heat Mass Transf. 43, ,(2013). Xu, Z., Kleinstreuer, C., Concentration photovoltaicthermal energy co-generation system using nanofluids for cooling and heating. Energy Convers. Manag. 87, (2014). Xu, Z., Kleinstreuer, C., Computational Analysis of Nanofluid Cooling of High Concentration Photovoltaic Cells. J. Therm. Sci. Eng. Appl. 6, (2014). Yang, K., Zuo, C., A novel multi-layer manifold microchannel cooling system for concentrating photovoltaic cells. Energy Convers. Manag. 89, (2014). Zhao, J., Song, Y., Lam, W.-H., Liu, W., Liu, Y., Zhang, Y., Wang, D. Solar radiation transfer and performance analysis of an optimum photovoltaic/thermal system. Energy Convers. Manag. 52, (2011). Zhou, J., Yi, Q., Wang, Y., Ye, Z., Temperature distribution of photovoltaic module based on finite element simulation. Sol. Energy 111, (2015). Ramos-Alvarado, B., Li, P., Liu, H., Hernandez- Guerrero, A. CFD study of liquid-cooled heat sinks with microchannel flow field configurations for electronics, fuel cells, and concentrated solar cells. Appl. Therm. Eng. 31, (2011). Reddy, K.S., Lokeswaran, S., Agarwal, P., Mallick, T.K., Numerical Investigation of Micro-channel based Active Module Cooling for Solar CPV System. Energy Procedia 54, ,(2014). Rejeb, O., Dhaou, H., Jemni, A., A numerical investigation of a photovoltaic thermal (PV/T) collector. Renew. Energy 77, (2015). Rohsenow, W.M., Hartnett, J.R., Forced convection internal flow in ducts. McGraw-Hill, New York (1998). Rosell, J., Chemisana, D., Tadrist, L., Iban, M., Effect of a hybrid jet impingement / micro-channel cooling device on the performance of densely packed PV cells under high concentration (2011). Royne, A., Dey, C.J., Mills, D.R., Cooling of photovoltaic cells under concentrated illumination: A critical review. Sol. Energy Mater. Sol. Cells 86, , (2005). Sarangi, S., Bodla, K.K., Garimella, S. V., Murthy, J.Y., Manifold microchannel heat sink design using optimization under uncertainty. Int. J. Heat Mass Transf. 69, (2014). Whitfield, G., Bentley, R., Weatherby, C., Hunt, A., Mohring, H.-D., Klotz, F., Keuber, P., Miñano, J., Alarte-Garvi, E., The development and testing of small concentrating PV systems. Sol. Energy 67, 23 34(1999). 111

126 Thermodynamic Analysis of Parabolic Solar Collector Driven Double-Effect Absorption Cooling System Fatih Yigit 1*, Ahmet Kabul 2, Onder Kizilkan 3 1,2,3 Suleyman Demirel University, Faculty of Technology, Energy System Engineering Department, Isparta, 32000, Turkey * Abstract This paper presents an analysis of thermodynamic performance of a double effect absorption cooling system, LiBr- H2O is used as fluid couple, driven by solar energy is supplied trough parabolic collector. The analysis is performed by using a software program Engineering Equation Solve (EES). First, cooling demand of an isolated supermarket with 1000 m 2 area is assumed 100 kw based on the location of the supermarket. To meet the cooling need, a double effect absorption cooling is considered and the double effect absorption cooling system is analyzed thermodynamically. Also some parametric studies are carried out by the variation of required thermal energy for generator, evoparator temperature, solar radiation intensity and parabolic solar collector s technical properties. Keywords: thermodynamic analysis, parabolic collector, solar energy, absorption cooling I. Introduction The rapid population growth and enhancement on the technological field for the last two decades, and people s demands for higher life standards and comfort level give rise to increasingly excessive energy consumption. One of the necessaries to comfort level and health is air-conditioning due to all people need fresh air (Evangelos et al., 2016). According to the International Institute of Refrigeration in Paris, the amount of electricity consumption of airconditioning process is approximately 15% of all the electricity produced in the world. On the other hand, electricity consumption for air conditioning systems has been estimated around 45% of the whole residential and commercial buildings (Kalkan et al., 2012). There are a few system using for air-conditioning however among these systems, especially vapor compression cooling systems consume too much electricity (Yilmazoglu, 2010). So these systems support to climatic change and global warning by using fluid (CFC) and excessive electricity consumption. CFC (Chloro Fluoro Carbon) and HCFC (hydrochlorofluorocarbon) gasses which are used in conventional cooling application utilization was band due to increase global warning by weakening ozone layer (Bejan, 1997). Additionally, it is uncertain that what kind of damages will emerge because of utilization of inorganic gasses which are used in conventional cooling systems. When this situation is considered, using double effect absorption cooling system which is an environmentally friendly system compared with conventional refrigeration systems is very beneficial to our world and country s feature. Double effect absorption cooling systems have relatively very low electricity consumption thanks to using pump rather than compressor for compression process (Onan et al., 2010). Thus, double effect absorption cooling systems with low electrical energy consumption and using renewable energy such as geothermal, solar, and waste heat is very important for sustainable energy and environment. Advantage of Absorption Cooling System (Sencan, 1999); Quiet operation because of pump quite quiet when compare compressor. It requires little maintenance. They can provide a complete productivity for variable cooling loads. Reduces energy costs thanks to use renewable energy sources such as sun, geothermal and waste heat. The absorption cooling system with H2O-LiBr might not work under 75 C because of crystallization of the LiBr. Therefore, generator temperature must be on the 75 C. To reach high generator temperature must be use like evacuated or parabolic trough solar collectors. Assumed that heat loss of the parabolic trough solar collectors only via radiation heat transfer so this type solar collectors have great advantages for solar cooling applications (Ozturk, 2008; Sencan, 1999). Turkey, which has got an advantageous location has a great potential of solar energy that can be used for numerous applications such as electricity generation, heating, cooling, etc. Solar energy is generally being used for water heating especially in southern of part of Turkey. The utilization of the solar energy can lead reduction of fossil fuel consumption and thus a reduction of carbon dioxide emissions. On the other hand, the high potential of solar energy can be utilized for cooling applications in summer times, by using double effect absorption cooling systems driven by 112

127 solar energy instead of electricity (Ozturk, 2008). There are a number of studies made by different researchers in the literature about absorption refrigeration systems. Onan et al. (2010) designed a solar assisted absorption refrigeration system (SAARS) for acclimatizing of villas in Mardin and they also analyzed the performance of the system under different temperatures by using MATLAB. F. Assilzadeh et al. (2005) studied a solar cooling system that has been designed for Malaysia and similar tropical regions using evacuated tube solar collectors and they carried out LiBr absorption unit by using TRNSYS program. Aman et al. (2014) developed a thermodynamic model which is based on a 10 kw air cooled ammonia water absorption chiller driven by solar thermal energy. They conducted both energy and exergy analyses to evaluate the performance of this residential scale cooling system. Praene et al. (2011) designed a solar-driven 30 kw LiBr/H2O single-effect absorption cooling system and installed at Institut Universitaire Technologique of Saint Pierre. Ratlamwala et al. (2012) investigated a parametric study which are undertaken and the effects of some operating conditions such as geothermal temperature, geothermal mass flow rate, concentration of ammonia water vapor, temperature of inlet stream to the very high temperature generator (VHTG), and pressure of the first turbine on the outputs of the system. Solum et al. (2011) examined effect of thermodynamic quantities of any doubleeffect absorption system operating by means of double fluid, LiBr-water on system performance by using an engineering program which name is EES. This paper presents thermodynamic performance analysis of a solar assisted double effect absorption cooling system using LiBr-H2O fluid couple. In the analysis, parabolic trough solar collectors are used for utilizing of solar energy. The analysis is performed by using the Engineering Equation Solve (EES) (F-Chart, 2016) software program. Parametric studies also are carried out for some variables such as thermal energy demand for generator, solar radiation intensity, evaporator temperature and PTSC technical properties. II. System Description The investigated supermarket which is to be cooled has got the dimensions of 1000 m 2 in area, 3.7 m in height and 3700 m 3 in volume, respectively. It is assumed that the supermarket is isolated with 0.04 m insolation material which has a got a heat conduction coefficient of W/mK. The solar assisted proposed double effect absorption cooling system consists of a solar parabolic trough collector, two generators, two heat-exchangers, a condenser, an evaporator, an absorber and four expansion valves. The proposed system is given in Figure 1. Technical properties of the parabolic trough solar collector used 113 the solar assisted double effect absorption cooling system given table 1. Tab. 1: Properties of the parabolic trough solar collector Single collector width 3.5 m Receiver inner diameter 0.04 m Receiver outer diameter 0.05 m Cover diameter 0.09 m Receiver emissivity 0.92 Glass cover emissivity 0.87 Temperature of the sun 5739 K Solar radiation intensity 500 W/m 2 System design parameters and assumptions are described below. The fluid couple used in the system is LiBr- H2O Evaporator temperature is considered to be 12 C Condenser and absorber temperature is considered to be 42 C. The effectiveness of heat exchanger for a counter flow heat exchanger is 0.6 on average Pressure losses in the system was neglected. Absorber, generator, condenser and evaporator are isolated for heat loss and gain III. Thermodynamic Analysis The performance of the parabolic solar collector driven absorption cooling system is mathematically modeled using mass, energy and exergy balance equations. In order to carry out the thermodynamic analysis for the system, some assumptions are made: The system processes are steady state. Kinetic and potential energies of the changes are ignored. The fluid couple at pump inlet is saturated liquid. The pumps are adiabatic. Potential and kinetic energies are neglected. The dead state pressure and temperature are taken P0= kpa and T0= 5 C General mass balance equation can be written as (Cengel and Boles, 2006): m in = m out (1) The general energy balance can also be written as: E in = E out (2) where E in is the ratio of net energy transfer to the system, E out is the ratio of net energy transfer from the system. For steady-flow processes the general exergy balance is defined as: Q + m inh in = W + m outh out (3)

128 Fig 1. The double effect absorption cooling system Where Q is the ratio of net heat, W is the ratio of net work, and h is the specific enthalpy. The rate of useful energy delivered by solar collector is defined as (Tiwari, 2003; Kalogirou, 2009) Q u = F R [SA a A r U L (T i T a )] (4) Q u = m C p (T o T i ) (5) where F R is the heat removal factor, S is the heat absorbed by the receiver, A a is the aperture area, A r is the receiver area, and U L is the solar collector overall heat loss coefficient. The general exergy balance equation can be defined as (Dincer and Rosen, 2007): Eẋ in = Eẋ out + Eẋ dest (6) The exergy balance equation can be expressed more explicitly as: Eẋ Q Eẋ W = m ine in m oute out + T 0 S gen (7) 114 where, Eẋ Q and Eẋ W terms are the exergies of heat and work, e is specific exergy, T0 is the state temperature and S gen is the rate of entropy generation. The terms in Equation 7 are described below (Dincer and Rosen, 2007): Eẋ dest = T 0 S gen (8) Eẋ Q = Q ( T T 0 T ) (9) Eẋ W = W (10) The specific exergy can be expressed as (Cengel and Boles, 2006; Bejan, 1997): e = (h h 0 ) T 0 (s s 0 ) (11) where, h is enthalpy, s is entropy, T is the temperature and subscript 0 stands for reference state properties. The exergy efficiency can be expressed as (Dincer and Rosen, 2007): η ex = Eẋ out Eẋ in = 1 Eẋ dest Eẋ in (12)

129 where, Eẋ out is the rate of total energy output, Eẋ in is the rate of the total energy input. The solar exergy is defined as (Petela, 2005): Eẋ solar = S A a ( ( T 0 T sun ) ( T 0 T sun )) (13) where Eẋ solar is the function of the outer surface temperature of the sun. IV. Result and Discussion In this paragraph, all the simulation results are presented. Firstly, coefficient of performance curve of system are given as a function of solar radiation. The next results are related to the exergy efficiency for each component of the double effect absorption cooling system. In this section some parametric studies are given for understanding performance change of the system. Table 2 shows the calculated thermodynamic properties of the solar assisted double effect absorption cooling system. Tab. 2: Thermodynamic properties of each point of the solar assisted double effect absorption cooling system Points Fluid Phase T P Cp h m s x Ex [ᵒC] [kpa] [kj/kg.k] [kj/kg] [kg/s] [kj/kg] [%] [kw] 0 H2O H2O-LiBr Strong mix H2O-LiBr Strong mix H2O-LiBr Strong mix H2O-LiBr Strong mix H2O-LiBr Mean mix H2O-LiBr Mean mix H2O-LiBr Mean mix H2O-LiBr Weak mix H2O-LiBr Weak mix H2O-LiBr Weak mix H2O Vapor H2O Liquid &Vapor H2O Liquid &Vapor H2O Vapor H2O Liquid H2O Liquid &Vapor H2O Vapor H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid H2O Comp. Liquid Figure 2 shows how the Coefficient of Performance (COP) and Exergy destruction of the system change with generator I temperature. COP COP Ex DestTotal T GI ( C) Fig. 2: COP according to the change in the Generator I temperature Ex DestTotal (kw) Figure 3 presents the COP of the absorption cooling system with increase absorber temperature. It shows clearly that COP increase with absorber temperature but exergy destruction of the system goes down. COP COP Ex DestTotal T E ( C) Fig. 3: COP according to the change in the Evaporator temperature Ex DestTotal (kw) 115

130 In Figure 4 seen that with the increase of the condenser temperature exergy destruction of the system decrease and COP goes up. According to Figure 5 as increase the absorber temperature COP is decrease on the other hand exergy destruction of the system first goes up than sharply goes down. The cause of the sharply decrease of exergy destruction might be crystallization of the LiBr. COP COP Ex DestTotal Ex DestTotal (kw) Fig. 7: Exergy Efficiency for each component according to the change in the Absorber temperature T C ( C) Fig. 4: COP according to the change in the Condenser temperature COP COP Ex DestTotal Ex DestTotal (kw) Fig. 8: Exergy Efficiency for each component according to the change in the Condenser temperature T A ( C) Fig. 5: COP according to the change in the Absorber temperature In Figure 6 to 9 the variation of exergy efficiency with different system parameters are given. Each figure show the effects of evaporator temperature, absorber temperature, condenser temperature and generator temperature on the exergy efficiency of each system components. Fig. 9: Exergy Efficiency for each component according to the change in the Generator I temperature Fig. 10: Exergy Efficiency for each component according to the change in the Evaporator temperature 116 V. Conclusions In this study, solar assisted double effect absorption refrigeration system analyzed using first and second laws of thermodynamics. According to the calculations, absorption refrigeration system s COP value was found to be 0.87 when the generator I temperature was 125 C. Additionally, some parametric studies have been performed to see the variation of exergy destruction rates and exergy efficiencies of the system components. It was found that with the increase of the first generator temperature, system performance increased. Consequently, parabolic trough collector integrated systems are becoming attractive for mid-temperature

131 thermal applications such as absorption refrigeration and also power generation systems. Additionally, it can be expected that these systems can redeem itself in a short period of time related to the technological improvements in PTSC systems. Nomenclature F R A a A r U L E x h m P PTSC Q s T W x References : heat removal factor, : aperture area, : receiver area, : solar collector overall heat loss coefficient. : Exergy (kw) : Specific enthalpy (kj/kg) : Mass flow rate (kg/s) : Pressure (kpa) : Parabolic Trough Solar Collector : Heat load (kw) : Specific entropy (kj/kg.k) : Temperature (C) : Work (kw) : Concentration Aman J., Ting D.S.K., Henshaw P., Residential Solar Air Conditioning: Energy and Exergy Analyses of an Ammonia-water Absorption Cooling System, Applied Thermal Engineering, 62, , Assilzadeh F., Kalogirou S.A., Ali Y., Sopian K., Simulation and Optimization of a Libr Solar Absorption Cooling System with Evacuated Tube Collectors, Renewable Energy, 30, , Bejan, A., Advanced Engineering Thermodynamics, John Wiley and Sons, New York, USA, Cengel Y.A., Boles M.A., Thermodynamics an Engineering Approach Eighth Edition, Published by McGraw-Hill Education, New York, USA, Dincer I., Rosen M.A., Exergy: Energy, Environment and Sustainable Development, Elsevier Science; 1st ed., Oxford, UK, Onan C., Ozkan D.B., Erdem S., Exergy Analysis of a Solar Assisted Absorption Cooling System on an Hourly Basis in Villa Applications, Energy, 35, , Ozturk, H.H., Gunes Enerjisi Ve Uygulamalari, Birsen Publishing, 2008, (In Turkish). Petela R., Exergy analysis of the solar cylindricalparabolic cooker, Solar Energy, 79, , Praene J.P., Marc O., Lucas F., Miranville F., Simulation and experimental investigation of solar absorption cooling system in Reunion Island, Applied Energy, 88, , Ratlamwala T.A.H., Dincer I., Gadalla M.A., Thermodynamic analysis of a novel integrated geothermal based power generation-quadruple effect absorption cooling-hydrogen liquefaction system, International Journal of Hydrogen Energy, 37, , Sencan A., Absorpsiyonlu Sogutma Sisteminin Tasarimi ve S.D.U Oditoryumunda Uygulanabilirliginin Arastirilmasi, MSc Thesis, Süleyman Demirel University, Isparta, Turkey, 1999 (In Turkish). Solum C., Koc I., Altuntas Y., Cift Etkili Libr-H2O Akiskanli Absorpsiyonlu Sogutma Sisteminde Termodinamiksel Buyukluklerin Sistem Performansina Etkileri, Journal of Aeronautics and Space Technologies, 1, 19-26, 2011 (In Turkish). Tiwari G.N., Solar Energy: Fundamentals, Design, Modelling and Applications, Alpha Science International Ltd., Pangbourne UK, Yilmazoglu M.Z., Tek Etkili Bir Absorpsiyonlu Sogutma Sisteminin Termodinamik Analizi, Journal of Faculty of Engineering and Architecture Gazi University, 25, , 2010 (In Turkis). Evangelos B., Christos T., Kimon A. A., Exergetic, energetic and financial evaluation of a solar driven absorption cooling system with various collector types, Applied Thermal Engineering, 102, , F-Chart Software. Engineering Equation Solver (EES). accessed on Kalkan N., Young E.A., Celiktas A., Solar Thermal Air Conditioning Technology Reducing the Footprint of Solar Thermal Air Conditioning, Renewable & Sustainable Energy Reviews, 16, , Kalogirou S.A., Solar energy engineering: processes and systems, 1st ed., Academic Press, Oxford, UK,

132 Energy and Exergy Analyses of a Biomass Fired Regenerative ORC System Ozum Calli 1*, Can Ozgur Colpan 2, Huseyin Gunerhan 3 1 Izmir University of Economics, Vocational School, Ventilation-Air Conditioning Technology Department, Balcova, Izmir, 35330, Turkey 2 Dokuz Eylul University, Faculty of Engineering, Mechanical Engineering Department, Tinaztepe, Buca, Izmir, 35397, Turkey 3 Ege University, Faculty of Engineering, Mechanical Engineering Department, Bornova, Izmir, 35040, Turkey * Abstract Heat generated from the combustion of biomass can be used as an energy source in an Organic Rankine Cycles (ORC). In this paper, an integrated biomass fired regenerative ORC system is examined using energy and exergy analyses. For this purpose, several control volumes enclosing the components of the system are formed. Applying exergy balances, exergy destruction in each control volumes are calculated. Various parameters including turbine inlet temperature, mass flow rate of dry biomass, fuel-air ratio, and type of biomass are investigated and to what extent which parameter affects the electrical efficiency and exergetic efficiency is determined. Some suggestions are given for increasing the electrical and exergetic efficiencies. Keywords: Organic Rankine Cycle, ORC, energy, exergy, biomass, regenerator I. Introduction As the human population grows and technological devices advance, dependence on energy increases significantly. Organic Rankine Cycles (ORCs) represent an attractive solution for the energy requirement, where traditional applications are technologically and economically unfeasible. The principle of electricity generation by means of an ORC is similar to the conventional Rankine cycle. The difference between these two cycles is that an organic working fluid (e.g. R134a, R245fa, R123, and hydrocarbons such as iso-pentane and iso-octane) with favourable thermodynamic properties at lower temperatures and pressures is used instead of water in the ORC (Qiu et al., 2011). For instance in geothermal ORC applications boiling point of ORC fluids chosen as working fluids are less than water (e.g. Quoilin et al., 2013). This fluid is selected such that it can utilize the heat from lower temperature sources. The fluid that takes the heat from this low temperature source is used to drive a turbine to generate electricity. There are various types of energy sources that can be used in an ORC system such as geothermal, solar, waste heat from industry, and biomass. This fluid is selected such that it can utilize the heat from lower temperature sources. The fluid that takes the heat from this low temperature source is usedtodrive a turbine to generate electricity. There are various types of energy sources that can be used in an ORC system such as geothermal, solar, waste heat from industry, and biomass. Organic Rankine cycle was investigated in terms of many parameters in the studies found in the literature. The most investigated parameteristhe working fluid and the most preferred fluid is R134a (Guo et al., 2010; Chen et al., 2010; Maizza and Maizza, 1996; Marion et al., 2012; Saleh et al., 2007). On the other hand, the selection of the most suitable working fluid depends on the operating conditions of the cycle. Lakew et al. (2010) investigated different working fluids for power production at different operating conditions and heat sources with different temperatures for a subcritical Rankine cycle. Their studies showed that R227ea gives the highest power for a heat source temperature range of C and R245fa produces the highest power in the range of C. Assessment of different configurations of ORC is another topic that has been widely investigated in the literature. Forinstance, Al-Sulaiman (2014) conducted exergy analysis of a combined steam Rankine cycle and organic Rankine cycle, which are both integrated with parabolic trough solar collectors. As a result of this study, it is found that the main source of exergy destruction is the solar collector. In addition, it is shown that as the solar irradiation increases, the exergetic efficiency increases. The highest and lowest exergetic efficiencies are obtained when R134a (26%) and R600 (20%) are used as the working fluid in the combined cycle, respectively. Biomass fired ORC systems can efficiently convert the chemical energy of biomass into electricity. Biomass is an important renewable energy source, available nearly everywhere and has the advantage of continuity by contrast solar and wind is intermittent. As the sunlight is presenced, biomass stored carbon continiously. In addition, biomassisoften economically favorable. There are a few studies on the biomass fired ORC in the literature. For example, Liu et al. (2011) investigated 2 kwe biomass-fired combined heat and 118

133 power (CHP) system based on anorganic Rankine cycle (ORC) by using three different environmentally friendly refrigerants, namely HFE7000, HFE7100 and n-pentane, as the ORC working fluids. They found that the electrical efficiency of the CHP system mainly depends on the temperature of the hot water entering the biomass boiler and the temperature of the cooling water entering and exiting the ORC condenseras well as the type of the ORC fluid. In another study by the same research group (Liu et al., 2012), a 0.8 kwe biomass-fired CHP system was investigated experimentally and it was found that the electricity generation efficiency is 1.41%, which islower than that predicted by the thermodynamic modelling. With an evaporator temperature of 120 C, the thermodynamic modelling study gives the electrical efficiency in the range of 8 9%. There are two main factors responsible for the apparent difference in the electrical generation efficiency between the experiments and thethermodynamic modelling: The first one is the expander efficiency. In the model, the value of this efficiency is assumed as 85%; whereas the experimental results show that this efficiency is only 53.92%. The second one is the alternator efficiency. In themodel, it is assumed that this efficiency is 90%; but the experimental results show that it is only 50.94%. Huang et al. (2013) investigated regenerativeandnon-regenerativebiomass-orc with dry and wet working fluids and found that the highest electric power was obtained for there generative system with methylcyclohexane applied as the dry working fluid. The aim of this paper is to analyze the energetic and exergetic performances of a regenerative biomass fired ORC by investigating to what extent which parameters affect. II. System description Fig. 1 shows the schematic diagram of the regenerative biomass-fired ORC system studied in this study. The system can be divided into two sections as the biomass side and the ORC cycle. In the biomass side, biomass fuel and air enter the burner and as a result of the combustion process, heat is generated. This heat is transferred to the biomass side working fluid. The working fluid in the biomass side then enters the heat exchanger that connects the biomass side and the ORC. In the ORC, the organic fluid first gets the heat from the biomass side through the heat exchanger. This fluid then expands in theturbine, producing mechanical energy, further transformed into electric energy through a generator. The fluid expanded in the turbine enters the regenerator, which is used to increase the electrical efficiency of the ORC. The exit of the regenerator enters the condenser and pump consecutively before entering the regenerator again. The regenerator increases the heat exchanger inlet temperature and decreases the heat gained from the burner. The fluid leaving the regenerator enters the heat exchanger completing the cycle. 119 Fig. 1: Schematic diagram of a regenerative biomass-fired ORC system. III. Mathematical Model The mathematical model of the integrated biomass fired ORC system is developed using energy and exergy analyses, which are discussed in thefollowingsubsections. A commercially available software, Engineering Equation Solver (EES), is used for the solution of the equations. The main assumptions made to carry out the energy and exergy analyses of the ORC system are listed below. The system is assumedtowork at steady state. The pressure drops along the components and the lines connecting the components are neglected except in the pump and ORC turbine. The heat loss from the components to the surroundings is neglected. The kinetic and potential energy and exergy changes are neglected. Air and the combustion gases are assumed to be ideal gases; and the biomass side fluid (Thermal oil) is assumed to be incompressible fluid. III.1. Energy analysis In this subsection, the energy analysis of the system considered is presented. Energy balance for steady state control volumes can be shown as follows: Q cv W cv = n o (h + v2 g. z) i 2 + g. z) o n i (h + v2 2 + (1) where Q cv and W cv are heat transfer rate and power rate of the control volume, respectively, and n i and n o are the molar flow rate of the working fluid at the inlet and outlet of the control volume, respectively. In this study molar unit system is selected for convenience as some of the modeling equations are written as a function of the molar compositions of the

134 chemical species. Hence, most of the equations given in this study are indicated according to molar unit system. v, g and z denote the velocity, gravitational acceleration and elevation according to a reference point, respectively. In the burner, the combustion process occurs in which air and biomass fuel react. The combustion gases are emitted to the atmosphere. The chemical reaction for the combustion of biomass, which mainly consists of C, H, and O atoms, can be shown as follows: C x H y O z + (λ γ )(O N 2 ) x CO 2 + (y/2) H 2 O + (α γ )O 2 + (3.76 λ γ)n 2 (2) In this equation, λ and α denote the theoretical air and excess air coefficient, respectively. γ is the stoichiometric air coefficient for the complete burning reaction of C x H y O z when there is no excess air. The relation between the excess air coefficient and the theoretical air can be shown as: α = λ 1 = λ + 2z 4x y 4γ Energy balance for the burner can be shown as: (3) n air h air + n biomass h biomass = n excessgases h excessgases + n biomassfluid (h outlet h inlet ) (4) Energy balance around the control volume enclosing the heat exchanger that connects the biomass side with the ORC is: n biomassfluid (h 1 h 2) = n ORCfluid (h 4 h 3) (5) Energy balance for the regenerator can be written in a similar way using Eq. (1). Using the energy balance for the condenser, heat transfer rate from the condenser to the cooling water can be shown as follows. Q condenser = n ORCfluid (h 6 h 7) = n cw (h 10 h 9) (6) Applying the energy balance for the control volumes enclosing the turbine and the pump, the power output of the turbine (Eq. (7)) and the power input to the pump (Eq.(8)) can be found, respectively. In these equations, η s,t denote the isentropic efficiency of the turbine, which can be defined as the actual work output of the turbine to the work output of the turbine if the turbine undergoes an isentropic process. η s,p is the isentropic efficiency of the pump, which is the ratio of the work input for an isentropic process, to the work input for the actual process. W turbine = η s,t (h 4 h 5,s ) (7) W pump = (h 8,s h 7) η s,p = υ 7 (P 8 P 7 ) η s,p (8) 120 III.2. Exergy analysis Exergy analysis is generally used to quantify the magnitudes of the irreversibilities in the thermal energy systems. Exergy balance, which is derived combining the energy and entropy balances, is applied to the components of the system to find the exergy flow rates at each state and the exergy destruction of each component. Exergy destruction can also be regarded the potential work lost due to irreversibilities. At the steady state conditions, the exergy destruction rate of a control volume can be found applying the exergy balance for a control volume, as follows: Ex d = (1 T 0 ) Q T j W cv + (n ex ) i (n ex ) o (9) j where n, T j, T 0, Q, Ex and Ex d are the molar flow rate, temperature of the boundary where heat transfer occurs, temperature of the environment, heat transfer rate between control volume and the environment, exergy flow rate, and rate of exergy destruction. The summation of the exergy destruction rate of the each component is called the total exergy destruction of the system. The contribution of the exergy destruction of each component in the total exergy destruction rate can be found by calculating the exergy destruction ratio as follows. y 1,i = Ex d,i Ex d,total (10) where Ex d,i and Ex d,total denote the exergy destruction rate of the component i and the total exergy destruction of the system. Alternatively, the exergy destruction rate of a component can be compared to the chemical exergy rate of the fuel (Colpan, 2005), which is taken as biomass for this study: y 2,i = Ex d,i Ex biomass (11) where Ex biomass and Ex d,i denote the exergy flow rate of the fuel and exergy destruction rate of the component i, respectively. The exergy includes the physical and chemical exergy, if the kinetic and potential exergies are neglected. ex = ch ex + ph ex (12) where ph ex is the specific physical exergy and ch ex is the specific chemical exergy at a given state. In general, if the chemical composition of a substance does not change at the inlets or exits of a control volume, the chemical exergy is not needed to be calculated to find the exergy destruction in that control volume. For the system studied, chemical exergy is included in the calculations only in the burner because

135 chemical reaction takes place in the combustion process. Physical flow exergy rate at a given state defined as: ph ex = (h h 0) T 0 (s s 0) (13) where h and s are specific enthalpy in molar basis and specific entropy in molar basis. s 0 and h 0 denotes specific enthalpy and specific entropy in molar basis for the dead state which defines conditions of the reference environment. In this study, The specific chemical exergy of an ideal gas mixture is defined as: ch ex = ex och + R T o ln x i (14) Here, x i is the mole fraction of species i and och ex is the standard specific chemical exergy (in molar basis) at the reference temperature and pressure. The chemical exergy of biomass can be defined as (Szargut, 2005): Ex biomass = β (n biomass (LHV biomass + w h fg ) (15) LHV biomass = biomass HHV h fg (21) According to Dulong s formula (Cho et al., 1995), the higher heating value of the biomass is a function of the dry-biomass weight percentages of carbon, oxygen, hydrogen, and sulphur, as shown in Eq. (22). HHV biomass = C (H O/8) (22) The exergy efficiency is defined as the ratio of the ratio of desired exergy outputs to exergy inputs expended to generate these outputs. As exergy efficiency is defined different each component due to having different working principles. For pump: η ex,pump = 8 ex 7 ex (23) w in For turbine: η ex,tur = w out 4 ex 5 ex For heat exchanger: (24) β = [ (H C ) (O C ) ( (H C ))] ( ( O C )) (16) η ex,hx = n ORC,fluid (ex 4 3) ex n bio,fluid (ex 1 2) ex (25) where n biomass is the molar flow rate of the biomass, w is the percentage of the moisture in the biomass, h fg is the molar specific enthalpy of vaporization of water and biomass LHV is the molar lower heating value of the biomass. C, H, O, and S denote the drybiomass weight percentages of carbon, oxygen, hydrogen, and sulphur. For regenerator: η ex,reg = (ex 3 8) ex (ex 5 6) ex (26) For the ORC and entire system exergy efficiencies can be found using Eqs. (27) and (28), respectively. III.3. Performance Assessment Parameters As a result of the energy analysis of the integrated system, the heat input to the burner and ORC, the net power output of the system, the heat transferred to ORC and the electrical efficiency of the ORC and the entire system can be found using Eqs. (17), (18), (19), (20) respectively. η ex,system = η ex,orc = W net Ex biomass W net n ORC,fluid (ex 1 2) ex IV. Results and discussion IV.1. Validation (27) (28) Q burner = n biomass LHV biomass (17) W net = W turbine W pump (18) Q ORC,in = n ORC,fluid (h 4 h 3) = n bio,fluid (h 1 h 2)(19) η el,orc = W net Q ORC,in (20) where n ORC,fluid and n bio,fluid denote working fluid circulating throughout the ORC and the fluid providing the heat transfer from biomass side to ORC. A computer code for the modeling equations presented in Section 3 is developed using the Engineering Equation Solver (EES) software. As a case study, the code developed is run considering the experimental data of a lab-scale ORC unit given in a study found in the literature (Gusev et al., 2014) as shown in Table 1. The results of the code are compared with those of the experiment for validating the model. The results of this comparison are given in Table 2. As can be seen from this table, the deviation between the numerical and experimental studies is less than 5%. The lower heating value of the biomass can be calculated knowing the higher heating value of the biomass, as shown in Eq. (21). 121

136 Table 1: Data taken from the experimental lab-scale ORC unit Parameter Value Type of working fluid in the ORC R245fa Type of heat transfer fluid Therminol 66 Mass flow rate of the working fluid in the ORC 0.3 kg/s (m ORC) Regenerator inlet pressure (P 5 ) Turbine inlet temperature (T 4 ) Heat exchanger pressure at the ORC side (P 3 ) Outlet temperature of the heat transfer fluid (T 2 ) Inlet temperature of the cooling water (T 9 ) Temperature difference between the inlet and outlet of cooling water stream (T 10 T 9 ) Heat transferred to the ORC side 171 kpa K 931 kpa 383 K K 3.7 K kw Table 2: Comparison between the experimental and numerical results Experimental Value Numerical Value Deviation Rate Turbine output 4.7 kw 4.9 kw %4 power (W T) Heat exchanger inlet K K %1 temperature at the ORC side (T 3 ) Regenerator outlet K K %0.4 temperature (T 6 ) IV.2. Parametric Studies After validation of the model, effects of some of the important input parameters on the performance of the system are investigated. These parameters include the turbine inlet temperature, excess air ratio, mass flow rate of the dry biomass, and the biomass types. In these studies, the baseline conditions used are given in Table 3. Using the data given in Table 3, the results that give all the thermodynamic properties of each state within the system are given in Table 4. Table 3: Baseline conditions used in the parametric studies Parameter Value Burner Type of fuel Wood Chemical composition of fuel CH 1.44 O 0.66 Mass flow rate of dry biomass kg/s Temperature of exhaust gases 400 K Excess air coefficient 0.2 Heat Transfer Fluid Type of heat transfer fluid Therminol VP-1 Mass flow rate of the heat transfer fluid 0.5 kg/s Pressure of the heat transfer fluid 780 kpa Temperature of the heat transfer fluid 673 K entering the heat exchanger ORC Type of working fluid in the ORC R134a Mass flow rate of the working fluid in the 0.3 kg/s ORC (m ORC) Pressure of working fluid entering the heat 3300 kpa exchanger (P 3 ) Condenser pressure (P 5 ) 700 kpa Turbine inlet temperature (T 4 ) 473 K Temperature difference between the inlet 10 K and outlet of cooling water stream (T 10 T 9 ) Isentropic efficiency of the turbine 0.68 Isentropic efficiency of the pump 0.8 Ambient temperature 298 K Ambient pressure 100 kpa 122 Table 4: Thermodynamic properties of each state of the integrated ORC system State Substance T (K) P h s n ex (kpa) (kj/kmol) (kj/ kmolk) (kmol/s) (kj/ kmol) 1 Therminol VP Therminol VP R 134a R 134a R 134a R 134a R 134a R 134a Water Water IV.2.1. Variation of Turbine Inlet Temperature Turbine inlet temperature (TIT) is one of the key operating parameters that affects the performance of the integrated system. The effect of the TIT on the electrical and exergy efficiencies are investigated and the results are shown in Fig. 2a. As shown in this figure, as the TIT increases, the net power output increases. As a result of this increase, the electrical and exergy efficiencies increase. Please note that as the heat transferred to the ORC is not a function of the TIT for this study as shown in Figure 2b, the trend of the electrical and exergetic efficiencies mainly depends on that of the net power output. To understand the reason of the trend of the change of exergy efficiency of the overall system shown in Fig. 2a, the exergy destrucion rates and effiencies of each component of the system are calculated. In this way, the components that have the highest exergy destruction rate (i.e. irreversibility rate) and lowest exergetic efficiency are found. Hence, the components that have more potential for improvement of the performance of the overall system are identified. Figures 3 and 4 show that the main reason of the increase in the exergy efficiency with the increase in TIT is the comperatively higher increase in the exergy efficiency of the heat exchanger or decrease in the exergy destruction rate. It can be seen from these figures that for this component, as the temperature increases from 377 to 470 K, exergy destruction rate decreases by 25%. This decrease can be atrributed to the increase in the change of the molar specific exergy change of the working fluid between the inlet and outlet the heat exchanger. Exergy destruction rate decrease in the condenser is the second significant reason for the increase in the exergetic performance of the overall system. The exergy destruction rate in this component mainly decreases because of the decrease in the heat transfer rate from the working fluid to the cooling water (i.e. enthalpy of the state 6 decreases while the enthalpy of state 7 does not change with turbine inlet temperature). On the other hand, these figures show that the exergy destruction rate of the regenerator and the turbine increases from kw to kw

137 and 2.97 kw to 3.5 kw, respectively, in this temperature range. Although there is an increase in the exergy destruction rate in these components, the exergy efficiency of the overall system increases as the total exergy destruction rate decrease is more than total exergy destruction rate increase. Exergy destruction rate and exergy efficiency of the pump and burner do not change because of the fact that inlet and outlet conditions of their control volumes do not change with respect to the turbine inlet temperature in this study; hence these components do not have an effect on the exergetic performance of the integrated system. Fig. 3: Change of exergy efficiency with the turbine inlet temperature Fig. 2: Change of (a) efficiency and (b) energy transfer with the turbine inlet temperature 123 Fig. 4: Change of exergy destruction rate of (a) heat exchanger and burner, and (b) pump, turbine, condenser, and regenerator with the turbine inlet temperature

138 IV.2.2. Variation of Excess Air Ratio A combustion process is complete if all the carbon, hydrogen, and sulfur (if any) in the fuel burns to CO 2, H 2 O, and SO 2, respectively. The minimum amount of air which allows the complete combustion of the fuel is called stoichiometric air. In this case, the products do not contain any oxygen. In practice, additional amount of air, which is called excess air, is fed to the burner. This excess air results in oxygen appearing in the products. Excess air also increases turbulence, which increases mixing in the combustion chamber. As there is more mixing of the air and fuel, these components have more chance to react. Hence, excess air helps to prevent the fuel from remaining unburned. On the other hand, supplying more than theoretical air provides safety. For ensuring complete safety, it is essential to control the levels of CO and check the amount of unburned hydrocarbon fuel. CO is a toxic gas that can be lethal in higher concentrations. Hydrocarbons which contains unburned fuel can cause explosions. The addition of excess air greatly lowers the formation of CO and unburned hydrocarbons by allowing them to react with O 2. High excess air ratio also reduces air pollution. As toxic compounds such as sulfur dioxide, carbon monoxide, nitrogen oxides can occur in high concentrations, smog, acid rain, and respiratory problems can occur (Basu et al., 2000). Excess air ratio is one of the important parameters affecting the performance of the system studied. In this study, the effects of this ratio on the energy and exergy efficiencies of the integrated system, and exergy efficiency and exergy destruction of each components are examined. Fig. 5a shows that the electrical and exergy efficiencies of the ORC increase with increase of the excess air ratio. This increase can be explained as follows. Heat transferred to ORC decreases in higher excess air ratios. As the excess air ratio increases, enthalpy of the state 2 increases whereas enthalpy of the state 1 does not change. When energy balance is applied to the control volume enclosing the heat exchanger, it is clearly seen that enthalpy of the state 3 (heat exchanger inlet of the ORC side) increases whereas enthalpy of the state 4 (turbine inlet) does not change. As the net power output does not change with the excess air ratio, both the electrical and exergy efficiencies of the ORC increase. The decreases in the exergy destructions in the heat exchanger, the regenerator and the condenser are responsible for the increase in the exergy efficiency of the ORC. The electrical and exergy efficiencies of the overall system also do not change as the heat occurred in burner does not change with excess air ratio. In addition, the increase in the exergy destruction rate of the burner is equal to the total decrease in the exergy destruction rates of other components; thus the exergy efficiency of the overall system does not change. On the other hand, excess air ratio has no effects on the exergy efficiency and exergy destruction rate of the turbine and pump. When the exergetic efficiencies of the components are observed, it can be seen that exergy efficiency of 124 the heat exchanger slightly increases as the decrease in the specific molar exergy difference between state 1 and 2 is greater than the decrease in the specific molar exergy difference between state 4 and 3. On the other hand, exergy efficiency of the regenerator increase significantly because the increase in the molar exergy difference between state 5 and 6 is less than the increase in the difference between state 3 and 8. Fig. 5: Change of efficiency (a) and (b) energy transfer with the turbine excess air ratio

139 Fig. 7: Change of exergy destruction rate of (a) burner, heat exchanger, and (b) pump, turbine, regenerator, condenser with the excess ratio IV.2.3. Variation of mass flow rate of dry biomass Fig. 6: Change of exergy efficiency of (a) burner, heat exchanger, and (b) pump, turbine, regenerator with the excess air ratio In this section, the effect of the mass flow rate of the dry biomass on the performance of the system is examined. An increase in the mass flow rate of dry biomass means increases in the heat gain from the burner and the heat transferred to ORC, which can also be interpreted from Eqs.17 and 19. These increases cause decreases of the electrical and exergy efficiencies because of the fact that the net output power of the cycle does not change with the change of mass flow rate of dry biomass. The trends of the changes of these efficiencies are shown in Fig. 8a. From the energy balance around a control volume enclosing the regenerator, it can be shown that an increase in the enthalpy of state 9 causes a decrease in the enthalpy of state 6. Enthalpy of the state 10 does not change with mass flow rate of dry biomass. Hence, heat transfer by the condenser increases with an increase in the mass flow rate of dry biomass. Exergy efficiency of the regenerator and heat exchanger denotes the ratio of exergy lost in the hot stream to the exergy gained in the cold stream as shown in Eq.26. The most considerable decrease of the exergy efficiency with an increase in the mass flow rate of biomass is in the regenerator. As the mass flow rate increases, exergy recovered from cold stream of the regenerator (stream that comes from pump) decreases more than its hot stream (stream that comes from turbine). Hence, the ratio becomes lower. Exergy efficiency of the heat exchanger slightly decreases with the increase of the mass flow rate because of the increase of the hot stream (stream that comes from burner) is more than the increase of the cold stream (stream that circulating in ORC). Exergy destruction rate of the regenerator increases with the increase of mass flow rate of dry biomass because of the fact that total molar specific exergy rates of the state 6 and 3 (outlet of the regenerator) increase with the increase of the mass flow rate of dry biomass while state 5 and 8 does not change. Exergy 125

140 destruction of heat exchanger also increases since the decrease of the specific molar exergy of the outlet conditions of the heat exchanger is more than decrease of the specific molar exergy of the intlet conditions of the heat exchanger. Hence exergy destruction rate of the heat exchanger decreases as shown in Eq. 9. Fig. 9: Change of exergy efficiency with the mass flow rate of dry biomass As shown in Fig. 10, decrease or increase of exergy destruction rate of the regenerator depends on mass flow rate. The increase in the specific flow exergy of turbine side outlet of the regenerator is less than the decrease in that of the pump side outlet of the regenerator until the mass flow rate value becomes 5 g/s. Hence exergy destruction increases. When the mass flow rate is higher than 5 g/s, the increase in the specific flow exergy of the turbine side outlet of the regenerator is more than the decrease in that of the pump side outlet of the regenerator. Thus exergy destruction increases. Fig. 8: Change of (a) efficiency and (b) energy transfer with the mass flow rate of dry biomass 126

141 burner. The more heat gained is, the more exergy destruction occured is (Fig. 14). Fig. 10: Change of exergy destruction rate of (a) heat exchanger, burner, and (b) pump, turbine, condenser, regenerator with the mass flow rate of dry biomass Fig. 12: Change of efficiency by the biomass type Variation of Biomass Fuels The lowest efficiency is gained when wheat straw is used as biomass fuel because of the fact that chemical structure of the wheat straw consists of the highest mass ratio of oxygen while paper s lowest. High oxygen mass ratio provides high higher heating value (HHV) and the high enthalpy of biomass. Hence the output heat increases as shown in Fig. 11. Fig. 13: Change of exergy efficiency by the biomass type Fig. 11: Change of energy transfer by the biomass type Power consumed and produced by the pump and the turbine does not change with the type of the biomass. As the net output power does not change, electrical efficiency of the ORC and system decrease when the biomass with high oxygen mass ratio is used as a fuel (Fig. 12). Accordingly, type of the biomass has no effects on the turbine and pump exergy efficiencies as shown in Fig. 13. Biomass with high oxygen mass ratio provides high amount of heat gained from Fig. 14: Change of exergy destruction rate by the biomass type V. Conclusions In this study effects of the change of various input parameters including turbine inlet temperature, mass flow rate of dry biomass, excess air ratio, and type of 127

142 biomass are examined to find out in what way and to what extent which parameter causes to change of the energy and exergy efficiency. Increase and decrease of the efficiencies are important as well as how this changes are occured. Because there are some important points that are related to the electrical and exergy efficiencies significantly. For instance, heat transfer to the ORC system, power output of the system, exergy destruction of the each components, etc. Hence this output parameters trends are examined with respect to change of input parameters. There are some significant conclusion of the study: High turbine inlet temperature provides high power output. Turbine inlet temperature has no effect on heat transferred to ORC or heat occurred in the burner. Hence electrical and exergy efficiencies of the ORC and entire system increase with the increasing turbine inlet inlet temperature. Variation of the excess air ratio does not affect the electrical and exergy efficiencies of the system because of the fact that change of the excess air ratio does not change the net output work and the lower heating value of the biomass. Higher mass flow rate of the dry biomass causes lower electrical and exergy efficiencies. The more biomass burns, the more heat occured. Increase of the heat generation causes the increase of the exergy destruction. Hence, both energy and exergy efficiency decreases in higher mass flow rate of the dry biomass. Type of the biomass has no effects to the power output of the turbine and the the power input of the pump. Because of the fact that turbine inlet temperature is taken input parameter. Hence exergy efficiencies of these components don t change with respect to the type of the biomass. Type of the biomass affects the amount of the heat generation and amount of the heat transfer to ORC system. Wheat straw provides high amount of heat gained from burner. In future studies, system can be modelled with more detailed conditions and thermal optimization can be applied. Furthermore, exergoeconomic analysis can be done. References Qiu G., Liu H., Riffat S., Expanders for micro-chp systems with organic Rankine cycle, Applied Thermal Engineering, 31, (2011). Quoilin S., Den Broek MV., Declaye S., Dewallef P., Lemort V., Techno-economic survey of Organic Rankine Cycle (ORC) systems, Renewable and Sustainable Energy Reviews, 22, (2013). Guo T., Wang HX., Zhang SJ., Selection of working fluids for a novel low-temperature geothermallypowered ORC based cogeneration system, Energy Conversation Management, 52, (2011). thermodynamic cycles and working fluids for the conversion of low-grade heat, Renewable and Sustainable Energy Reviews, 14, (2011). Maizza V., Maizza A., Working fluids in non-steady flows for waste energy recovery systems, Applied Thermal Engineering, 16, (1996). Marion M., Voicu I., Tiffonnet AL., Study and optimization of a solar subcritical organic Rankine cycle, Renewable Energy, 48, (2012). Saleh B., Koglbauer G., Wendland M., Fischer J., Working fluids for low-temperature organic Rankine cycles., Energy, 32, (2007). Lakew AA., Bolland O., Working fluids for lowtemperature heat source, Applied Thermal Engineering, 30, (2010). Al-Sulaiman FA., Exergy analysis of parabolic trough solar collectors integrated with combined steam and organic Rankine cycles, Energy Conversion Management, 77, (2014). Liu H., Shao Y., Li J., A biomass-fired micro-scale CHP system with organic Rankine cycle (ORC) - Thermodynamic modelling studies, Biomass and Bioenergy, 35, (2011). Qiu G., Shao Y., Li J., Liu H., Riffat SB., Experimental investigation of a biomass-fired ORC-based micro- CHP for domestic applications, Fuel, 96, (2012). Huang Y., Wang YD., Rezvani S., McIlveen-Wright DR., Anderson M, Mondol J., Zacharopolousa A., Hewitta NJ., A techno-economic assessment of biomass fuelled trigeneration system integrated with organic Rankine cycle, Applied Thermal Engineering, 53, (2013). Colpan CO., Exergy analysis of combined cycle cogeneration systems, Ms.C. Thesis, Middle East Technical University (2005). Szargut J., Exergy Method: Technical and Ecological Applications, WIT Press (2005). Cho KW., Park HS., Kim KH., Lee YK., Lee KH., Estimation of the heating value of oily mill sludges from steel plant, Fuel, 74, (1995). Gusev S., Ziviani D., Bell I., De Paepe M., Den Broek MV., Experimental comparison of working fluids for organic Rankine cycle with single-screw expander, 15 th International Refrigeration and Air Conditioning Conference, Purdue (2014). Basu, P., Cen, K.F., Jestin L., Boilers and burners, Springer, New York (2000) Chen H., Goswami DY., Stefanakos EK., A review of 128

143 Transient Analysis of an Absorption Solar Refrigerator with External and Internal Irreversibilities Yasmina Boukhchana 1*, Ali Fellah 2, Ammar Ben Brahim 3 1 Research Unit of Applied Thermodynamics, Department of Chemical and Processes Engineering, National School of Engineers of Gabes, University of Gabes, St Omar Ibn El-Khattab, 6029 Gabes, Tunisia Affiliation 2 Research Unit of Applied Thermodynamics, Technology Department, High Institute of Applied Sciences and Technology University of Gabes, 6029 Gabes, Tunisia 3 Research Unit of Applied Thermodynamics, Department of Chemical and Processes Engineering, National School of Engineers of Gabes, University of Gabes, St Omar Ibn El-Khattab, 6029 Gabes, Tunisia * Abstract The transient analysis of a solar absorption refrigeration cycle with external and internal irreversibilities is presented in this paper. The model consists of a flat plate solar collector, a refrigerator with three finite-size heat exchangers, namely, the evaporator between the refrigeration load and refrigerant, the condenser between the refrigerant and the ambient, and the generator between the solar collector and the refrigerant, and finally the refrigerated space. The total thermal conductance of the three heat exchangers is fixed.an empirical function is used to model the internal entropy generation of the cycle. The parameters of this function are estimated by fitting data obtained by simulation to the predictions of the THR model. The model is based on the first and second laws of thermodynamics, heat transfer equations at finite thermal source and sink capacities and entropy generation terms in order to consider the internal and external irreversibilities of the cycle. A thermodynamic analysing and optimization of the absorption cycle is then performed, reporting the operating conditions for minimum time to reach a prescribed cold-space temperature, thus maximum refrigeration rate, specifically, the optimal temperature of hot space and the optimal way of allocating the thermal conductance inventory. The results are presented in normalized charts for general applications. The collector temperature presents major influence on the conceptual and functional characteristics compared to the stagnation temperature influence. On the other hand the thermal load in the refrigerated space and the thermal conductance of the walls has analogous effects, therefore important to be considered in actual design. As a result, the model is expected to be a useful tool for simulation, design, and optimization of solar collector based energy systems. Keywords: Solar energy, Refrigeration, Absorption, irreversibilities, Optimization, Transient regime. I. Introduction Absorption refrigeration systems that could be used with solar energy or other sources of thermal energy such as waste heat are being developed for application in air-conditioning systems. The performance of absorption systems were studied expensively by detailed computer simulation (M. O. MC Linden and S. A. Klein, 1985; G. Grossman and al. 1987; K. Gommed and G. Grossman, 1990). The development of such computer codes require considerable effort and they also need as input the thermophysical properties of the working fluids. For preliminary design studies and for performance data representation it is useful to develop simplified models for absorption cooling systems. Such models can be used to represent performance characteristics of absorption machines when they form sub-components of a larger thermal system simulation programme. Several idealized models were developed recently using the three-heat-reservoir (THR) configuration of the absorption cycle (J. Chen and Z. Yan, 1989; 1989 bis; N. E. Wijeysundera, 1996). These models which take into account the external heat transfer irreversibilities of the cycle are able to provide realistic performance limits for the coefficient of performance (COP) and the cooling capacity of absorption refrigeration systems (N. E. Wijeysundera, 1996). However, if the THR models are to predict the performance of real absorption machines closely, the internal irrevesibilities of the cycle in addition to the external irrevesibilities have to be included in the analysis. Such models were used to obtain the optimum performance of commercial absorption chillers (J. M. Gordon and K. C. Ng, 1995; H. Tong Chua and al. 1996). Also by using a few fitting parameters, these models were able to reproduce performance data for absorption chillers (J. M. Gordon and K. C. Ng, 1995). Nevertheless, all those studies focus on the systems steadystate properties and ignore completely their dynamic behavior. Steady-state models are useful under many conditions although under strongly dynamic conditions that are often seen in real-life operation, these models can become unacceptably inaccurate 129

144 (Browne MW, Bansal PK. 2002). However, steady state models do not provide time dependent information on the thermal behavior of absorption refrigerators and are therefore not suitable for transient system simulations. In contrast, the model presented in this work allows the simulation of the dynamic absorption refrigerator behavior. It extends the range of applicable models for transient system simulations, where the time constants of the refrigerator significantly influence the system performance. The dynamic model of an irreversible absorption refrigerator allows the simulations of its transient behavior for changing input conditions or design parameters. This is important because absorption refrigerators usually have a high thermal mass, consisting of their internal heat exchangers, the absorbing solution and the externally supplied heat transfer media. The contribution of this work is the analysis of the transient irreversible three heat reservoir absorption heat transformers with Newton s heat transfer law. Thus, a transient mathematical model for a solar collector driven refrigeration plant is introduced. Finding an optimum heat transfer rate received from the solar collector to the generator and investigating the effect of time in solar collector stagnation temperature and collector temperature and heat rate are derived by minimizing the time required to reach a certain operation temperature in the refrigerated space. This issue becomes more important in large scale cooling applications in which the thermal inertia of the refrigerated space becomes very large. II. The transient model The main features of the absorption refrigerator-refrigerated space model are shown in Figure 1. The cycle has negligible work input. The cycle is driven by the heat transfer rate QH received from the source temperature TH, which is determined by the operation temperature of the generator. The refrigeration load QL is removed from the refrigerated space, at TL, and the heat transfer rate Q0 is rejected to the ambient, T0. The refrigerator shown in Figure 1 operates irreversibly due to the entropy-generation mechanisms that are present (for example, heat transfer, mixing, and throttling). The irreversible model takes into account the internal and external irreversibilities, which are fundamental features that will be present in the design of real absorption refrigerators. The instantaneous heat transfer interactions are given by Q UA T T (1) H H H HC Q UA T T (2) L L L LC Q UA T T (3) 0 0 0C 0 Q H Additionally, is proportional to the collector efficiency, where, without loss of generality, and negligible heat loss between the solar collector and the generator, as follows: Fig. 1: Problem sketch QH SCASCG (4) where AS.C represents the collector area, G is the irradiance at the collector surface. The efficiency of a flat plate collector can be calculated as: (Bejan and all., 1995; Sokolov and Hershgal, 1993) SC ab T H T 0 (5) where a and b are two constants that can be calculated, as discussed by Sokolov and Hershgal (Bejan and all., 1995; Sokolov and Hershgal, 1993bis). Eq. (5) can be rewritten by introducing the collector stagnation temperature Tst as follows: b T T (6) SC St H where Tst (for which SC 0 ) is given by: T T a b (7) St 0 The equation for heat input QH can be rewritten by combining Eqs. (4) and (6) as follows: Q A Gb T T (8) H SC st H The first and second law read: Q Q Q (9) H L Generator Condenser + Absorber Evaporator 0 0 ds Q0 Q Q dt T T T in H L 0C HC LC T HC T 0C T LC Q 0 Q L T 0 T L Q H T H Irreversible refrigerator (10) 130

145 We account for the transient cooling of the refrigerated space by writing the first law, dtl M C. UA T0 T Q Q W dt air v air L load L where W 0 L (11) UA T T accounts for the rate of heat gain from the ambient through the walls of the refrigerated space and Qload is the thermal load or rate of heat generated inside the refrigerated space. By writing the set of Eqs. (1) (10) for the absorption refrigerator and (11) for the refrigerated room, we take into account the fact that the thermal inertia of the refrigerated space is large enough such that the transient operation of the refrigerator can be neglected when compared to the time evolution of the temperature inside the refrigerated space. Generally, it is difficult to model all internal entropy generation sources in order to get an analytical variation law. We have chosen to consider the following approaches (Wijeysundera, 1997; Gordon and Ng, 2000). The entropy of the working fluid is represented by using linear variation law with temperature: ds dt in T T T T 1 HC 0C 2 0C LC, (12) where the parameters are to be estimated by fitting detailed simulation data to predictions. 1 2 To obtain the best estimates of the parameters 2 and from simulated performance data (Boukhchana and all., 2014, 2015) the following least-square procedure is used. According to the cycle model mentioned above, the rate of entropy generated by the cycle is described quantitatively by the second law as: ds Q0 Q Q dt T T T Tot H L 0 H L (13) The factors (UA)H, (UA)L, and (UA)0 represent the overall thermal conductances of the heat exchangers. The overall thermal conductance of the walls of the refrigerated room is given by (UA)W. The proposition here is to use the model to optimize the distribution of finite resources and the generator heat input QH, aiming to achieve maximum refrigeration rate, QL, in the transient regime. For that, since (UA)H, (UA)L, and (UA)0 are commodities in short supply, it makes sense to recognize the total external conductance inventory, UA (hardware), as a constraint: UA UA UA UA (14) H L In addition, we define the ratio w, which accounts for the size of the heat transfer area of the refrigerated room, as follows: UA w UA w The nondimensional version is (15) Q y (16) H H HC Q z (17) L L LC 1 1 Q y z (18) 0 0C Q B (19) H st H Q Q Q (20) H L 0 0 ds Q0 Q Q d in H L 0C HC LC dsin d ds Q Q Q0 d tot H L HC C C LC d L w 1 Q Q d H L L load L (21) (22) (23) (24) where we have appropriately defined the following nondimensional groups: TH T H, L T T L 0 0 T T T, 0 HC LC 0C T T T HC LC C QH QL Q0 Qload QH, QL, Q0, Qload UAT UAT UAT UAT tua A Gb S, B SC, S M C UA M C air v, air air v, air The conductance allocation ratios are y UA UA UA H, z UA L (25) (26) (27) (28) (29) We are interested in how the imperfect features (finite temperature differences) identified in the model influence the overall performance of the refrigeration plant.

146 III. Numerical method and results set point temperature (, L set = 0.97). The problem consists of integrating Eqs. (23) and (24) in time and solving the non-linear system (16) (22) at each step time. The objective is to minimize the time θset to reach a specified refrigerated space temperature, L, set, in transient operation. To generate the results shown in Fig. 2 8 some selected parameters were held constant and others were varied. The numerical method calculates the transient behavior of the system, starting from a set of initial conditions, then the solution is marched in time and checked for accuracy until a desired condition is achieved (temperature set points or steady state). The equations are integrated in time explicitly using an adaptive time step, 4th 5th order Runge Kutta method (Yang and all., 2005). Newton Raphson s method with appropriate initial guesses was employed for solving the above set of non-linear equations. During the integration of the ordinary differential equations, one time the set of fixed parameters H, st, B, y, z, w HC and Q load is defined Eqs (16) and (19) give. The system of Eqs (16) (22), at each time step of integration of Eqs (23) and (24), deliver and. Q, Q L, 0 0C LC To test the model and for conducing the analysis presented in this section, we assuming a small absorption refrigeration unit with a low total thermal conductance (UA = 400 W/K), we considered a total heat exchanger area A = 4m 2 and an average global heat transfer coefficient U = 0.1 kw/m 2 K in the heat exchangers and Uw = kw/m 2 K across the walls, which have a total surface area Aw = 54m 2, T0 = 25 C and Qload=0.8 kw. Considering a typical air conditioning application, the refrigerated space temperature to be achieved was established at TL,set = 16 C. Thus, the resulting dimensionless parameters that were kept fixed initially were: L, set =0.97. Q load =0.007, w=0.2, Fig. 2 shows that during the heat up period, the temperature of the evaporator starts to decrease linearly then it decreases very slowly. Here, the reaction of the evaporator is seen strongly affected by the generator behavior. His temperature starts rising linearly, then it becomes stable. As the temperature of the generator is higher causing more heat is absorbed in the evaporator. While, the temperature of the evaporator is decreasing very slowly the temperature of the generator still maintained quit constantly, indicating that the equilibrium state has reached (Abdullah and Hien, 2010). Also, there is an intermediate value of the collector size parameter B, between and 0.175, such that the temporal temperature gradient is maximum, minimizing the time to achieve prescribed L Fig. 2: The behavior refrigeration space temperature, in time ( H = 1.3, L st = 1.6) B=0.07 B=0.04 B=0.1 Fig. 3 and 4 show the behavior of θset versus B, while varying y and ΓH. The results stress the importance of identifying Bopt, mainly for lower values of ΓH. For ΓH = 1.3, there is a narrower range of values for B where the system operates in optimal conditions, outside of which the performance deteriorates dramatically. This effect is reduced as ΓH increases, as is demonstrated with the results for ΓH = 1.4. The existence of an optimum with respect to the thermal energy input is not due to the irreversible equations that model the system alone. However, an optimal thermal energy input results when the irreversible equations are constrained by the recognized total external conductance inventory, UA (hardware), Eq. (14), which is finite, and the operating temperature of the generator, ΓH. These constraints are the physical reasons for the existence of the optimum point. During the transient operation, to reach the desired ΓL,set = 0.97, there is an internal and a total entropy generated by the cycle, which is obtained by integrating Eqs. (21-23) in time. set y=0.2, z=0.3 y=0.3, z=0.2 y=z= B Fig. 3: Time to reach a refrigerated-space temperature setpoint for different thermal conductance allocations for H = 1.3 and st =

147 10 9 Bopt is simply the optimal collector size for which in the presence of a finite UA θset is minimum, which represents neither maximum efficiency nor minimum total entropy generated by the cycle. 8 set x y=0.3, z= 0.2 y=z=0.25 y=0.2, z= B Fig. 4: Time to reach a refrigerated-space temperature setpoint for different thermal conductance allocations for H = 1.4 and st = 1.6. Fig. 5 and 6 show the internal and total entropy generated by the cycle up to θset, versus B, while varying y and ΓH. We see that there are a minimum for internal and total entropy generated by the cycle for a certain dimensionless collector size parameter B. Note that Bopt, identified for minimum time to reach ΓL,set, does not coincide with the Bopt where minimum internal and total entropy occurs, although the values are close. (a ) S in S tot y=0.2, z=0.3 y=0.3, z=0.2 y=z= B (a) S in x 10-4 y=0.3, z= y=0.2, z=0.3 y=z= B B (b ) Fig. 6: Internal and total entropy generated during the time to reach a refrigerated-space temperature setpoint for different thermal conductance allocations and for H = 1.4, st = 1.6. According to our initial proposition, we seek the set of optimal values (Bopt, yopt) that minimize θ to reach ΓL,set, thus maximizing in the transient regimes. Figures 7 and 8 illustrate the behavior of θset,min and Bopt(θset,min) versus y, while varying ΓH, therefore identifying the set ( Bopt, yopt) which corresponds to our original set of fixed parameters, w, and L, set 0.02 y=0.2, z=0.3 y=0.3, z=0.2 y=z= Q L Q load. The results show that the thermal conductance should be divided equally between the generator and evaporator for maximum Q (y = 0.25). L S tot y=0.3, z=0.2 y=0.2, z=0.3 y=z= B (b) Fig. 5: Internal and total entropy generated during the time to reach a refrigerated-space temperature setpoint for different thermal conductance allocations and for H = 1.3, st =

148 Fig. 7: Minimum time to reach a refrigerated-space temperature setpoint for different coupling temperatures, with respect to the variation of the thermal conductance allocation. B opt set-min y H =1.3 H =1.4 H =1.3 H = y Fig. 8: Optimal collector size to reach a refrigerated-space temperature setpoint for different coupling temperatures, with respect to the variation of the thermal conductance allocation. V. Conclusions In this article, a transient irreversible model to study the absorption refrigeration cycle was presented and used to demonstrate the existence of an optimal way of allocating the thermal conductance inventory and an optimal collector size for maximum refrigeration rate. The model accounts for the internal and external irreversibilities. This means that these optima are fundamental features that will be present (and deserve to be identified and exploited) in the design of actual absorption refrigerators, no matter how complicated these designs may be. Appropriate dimensionless groups were identified and the generalized results reported in charts using dimensionless variables. The importance of the analysis of the absorption refrigeration system in the transient regime is this stressed. The most important conclusion is that 1. The maximum refrigeration rate, for minimum time to achieve a specified refrigeration load temperature, requires a narrow range of collector size parameter, mainly for low coupling temperatures. 2. The Optimal collector size and minimum time to reach a specified refrigerated-space temperature are influenced analogously by the thermal conductance of the walls. 3. In general, half of the total supply of thermal conductance has to be divided equally between the generator and evaporator, for maximum refrigeration rate. 4. Optimal size collector identified for minimum time to reach set point temperature in the refrigerated space does not coincide with Bopt where minimum total entropy occurs. Nomenclature A : Area, (m 2 ) a, b : Constant in Eq.(5) B : Dimensionless collector size parameter C : Specific heat, (kj/kg K) G : Irradiance on collector surface, (W/m 2 ) M : Mass of air in the refrigerated space, (kg) Q : Heat transfer rate, (W) S : Entropy generation rate, (kj/k) t : Time, (s) T : Temperature, (K) U : Global heat transfer coefficient, (W/m 2 K) W, y, z : Conductance fraction Greek letters Γ : Dimensionless temperature θ : Dimensionless time : Efficiency of a flat plate collector Superscripts 0 : Ambient air : Air H : Heat source L : Refrigerated space load : Cold space thermal load opt : Optimum SC : Solar collector Set : Setpoint St : Collector stagnation temperature References Abdullah M.O., Hien T.C., Comparative analysis of performance and technoeconomics for a H2O NH3 H2 absorption refrigerator driven by different energy sources. Applied Energy, 87, , (2010). Bejan A., Vargas J.V.C. and Sokolov M., Optimal Allocation of a Heat Exchanger Inventory in Heat Driven Refrigerators, International Journal of Heat and Mass Transfer, 38, , (1995). Boukhchana Y., Fellah A. and Ben Brahim A., Numerical Study of Entropy Generation in an Irreversible Solar-Powered Absorption Cooling Systems, 9 ème Congrès Francophone de Génie des Procédés, Agadir, Maroc, April 28-30, (2014). 134

149 Boukhchana Y., Fellah A. and Ben Brahim A., Transient modeling and simulation of an ammonia-water absorption solar refrigerator, International Journal of Mechanics and Energy, 3(1), 33-43, (2015). Browne MW, Bansal PK. Transient simulation of vapour-compression packaged liquid chillers. Int J Refrige, 25, , (2002). Engineering, 17, 12, , (1997). Yang W.Y., Cao W., Chung T.S., Morris J., Applied numerical methods using MATLAB, Wiley-Interscience, A John Wiley & Sons, Inc.; (2005). Chen J. and Yan Z., Equivalent combined systems for three-heat-source heat pumps. J. Chem. Phys. 90(9), (1989). Chen J. and Yan Z., An optimal endoreversible three-heat-source refrigerator. J. Appl. Phys. 65(l), l-4 (1989 bis). Gommed K. and Grossman G., Performance analysis of staged absorption heat pumps: water-lithium bromide systems. ASHRAR Trans., (1990). Gordon J. M. and Ng K. C., A general thermodynamic model for absorption chillers: theory and experiment. Heat Recovery CHP 15(l), (1995). Gordon J.M., Ng K.C., Cool Thermodynamics, Cambridge Int. Science Publishers, Cambridge, (2000). Grossman G., Gommed K.and Gadoth D., A computer model for simulation of absorption systems in flexible and modular form. ASHRAE Trans. 93(2), (1987). Linden M.O. MC and Klein S. A., Steady state modeling of absorption heat pumps with a comparison to experiments. ASHRAE Trans. 2(B), (1985). Sokolov M., Hershgal D., Optimal coupling and feasibility of a solar powered year-round ejector air conditioner, Solar Energy, 50, 6, , (1993). Sokolov M., Hershgal D., Solar-powered compression-enhanced ejector air conditioner, Solar Energy, 51, , (1993bis). Tong Chua H., Han Q., Choon Ng K., and Gordon J. M., Thermodynamic modeling and experimental evidence for the optimization and maximum-efficiency operation of absorption chillers. ECOS 96, Efkiency, Cost, Optimizarion,Simulation and Environmental Aspects of Energy Systems, June 25-27, Stockholm, , (1996). Wijeysundera N. E., Performance limits of absorption cycles with external heat-transfer irreversibilities. Appl. Thermal Engng 16(2) (1996). Wijeysundera NE., Performance of three-heat-reservoir absorption cycles with external and internal irreversibilities, Applied Thermal 135

150 A Study on Adsorption Characteristics of Activated Carbon-R134a and Activated Carbon-R404a Pairs Muhsin Kilic*, Ersan Gonul Uludağ University, Engineering Faculty, Department of Mechanical Engineering, Bursa, TR16059, Turkey * Abstract As one of environmentally friendly refrigeration methods, solid adsorption refrigeration, which can be powered by low-grade renewable and waste heat resources, has tracked much interest over the world. The physical adsorption process occurs mainly within the pores and surface of the adsorbent. It required the knowledge of adsorption characteristics when the temperatures and pressures are varying. The objective of this study is to evaluate adsorption characteristics of R134a and R404a on activated carbon experimentally by a constant volume variable pressure method at different adsorption temperatures ranging from 293 to 333 K and for pressures up to about 5 MPa. These data are useful for the design of adsorption cooling and refrigeration systems and are unavailable in the literature. Two samples of commercially available activated carbon with widely varying surface areas were chosen. The shapes of the isotherms obtained from the experimental data were similar in all cases and comparable to those reported in the literature. Adsorption parameters were evaluated from the isotherms using the Dubinin-Astakhov (DA) equation. The concentration dependence of the isosteric enthalpies of adsorption is extracted from the data. Further, the enthalpy of adsorption data were extracted, and correlations are provided for the two specimens investigated. Keywords: Adsorbent, activated carbon, adsorption system, refrigerant, R134a, R404a. I. Introduction Optimizing energy, protecting environment and sustainable development are all the main themes of the contemporary world in the 21 st century. As one kind of environmentally friendly refrigeration method, the research and developments on the adsorption refrigeration systems have attracted more attention in recent decades (Wang et al., 2009, Wang et al., 2010). The main heat sources for adsorption machines are waste heat and solar energy. Physical adsorption working pairs are usually preferred when solar energy is the heat source (Anyanwu and Ezekwe, 2003, Solmuş et al., 2010). It provides an alternative to conventional vapor compressor refrigeration, because the former can be driven by low grade heat sources such as solar energy and industrial waste heat. In addition, they have minimal moving parts. In contrast to vapor absorption cycles, adsorption cycles dispense with the heat exchangers (Wang et al., 2010). The properties of adsorbent/adsorbate pairs as well as the operating conditions have significant effects on the system performance (Solmuş et al., 2014, Wang et al., 2009). The isosteric heat of adsorption is a specific combined property of an adsorbent/adsorbate combination. The equilibrium adsorption properties at several adsorbent temperatures and adsorption chamber adsorbate pressures were studied for a wide range of pairs (Chan et al., 1984, Wang et al., 2009, Solmuş et al., 2011, Solmuş et al., 2011, Saham et al., 2008, Saha et al., 2007, Saha et al., 2009, Wang and Wang, 1999). Meanwhile, on the refrigerant field considerable impetus already exists to use natural and/or ozone friendly refrigerants. If the need is to use refrigerants that result in system pressures above atmospheric pressures, that are also non-toxic and ozone friendly, the choice narrows down to partly halogenated hydro fluorocarbon refrigerants such as R-134a (tetrafluoroetan CF3CH2F) and R-404a (CHF2CF3 / CH3CF3 / CF3CH2F) which is near a zeotropic blend of HFC-125/HFC-143a/HFC-134a. Thus, R-134a and R-404a based adsorption refrigeration cycles provide a perfect match for the current aspirations and expectations from adsorption cooling systems. The design of these refrigeration systems requires data on isotherms and the heats of adsorption for indenting heating inventories (Riffat et al., 1997, Solmuş et al., 2014, Anyanwu and Ezekwe, 2003, Banker et al., 2004). For example, in an adsorbtion cooling system, when the adsorbate gas is adsorbed by a solid adsorbent in a thermal compressor, the heat of adsorption has to be removed using a heat sink. Similarly, when the adsorbate gas is desorbed at a higher temperature and pressure there is a need to add the heat of adsorption. Therefore, the variation of the heat of adsorption as a function of loading, which in turn depends on the pressure and temperature at which adsorption/desorption occurs, has to be considered. It is known that the dependence of isosteric heat of adsorption on the loading is a measure of energetically homogeneous nature of the adsorbent surface. Detailed literature review on adsorption working pairs for refrigeration is given by Wang et al.(2009). Isosteric heat of adsorption is traditionally expressed 136

151 as a function of concentration due to its dependence on temperature is relatively weaker (Chakraborty et al., 2006, Chan et al., 1984, Parakash et al., 2000, Saha et al., 2007). For adsorption of fluids below their thermodynamic critical point, its magnitude is larger than the heat of vaporization of the adsorbate, which has a strong temperature dependence (Saha et al., 2009, Chakraborty et al., 2006). As a result, the difference between the two is a property of relevance in the design of adsorption refrigeration systems. It is a matter of regret that adsorption data are unavailable from the manufacturers of adsorbents. The characteristics of a new adsorbent like a kind of activated carbon may show differently than the known ones. In order to design adsorption based cooling cycle it is inevitable to evaluate adsorption isotherms of the assorted adsorbent/adsorbate pair as well as the isosteric heat of adsorption. In the view of the above mentioned perspectives, the present study reports an experimental study to obtain isotherm data for the adsorption of R134a and R404a refrigerants on the two different type of commercially available activated carbon (AC) specimens. Adsorption isotherms of R134a and R404a on the activated carbon specimens were measured over a temperature range of C and pressures up to about 5 bar using constant volume variable pressure (CVVP) method. Moreover, the isosteric heats of adsorption are evaulated from the present experimental data. II. Experimental Facility II.1. Setup tanks were heated by using hot water circulation at 60 C during 6 h, while the vacuum process is still running. At the end of regeneration process, the test system is purged with helium gas and evacuated further to achieve low vacuum conditions. The evacuation and helium purging are continued several times to ensure that there is no residual gas left in the system. Based on the measurements, there is no measurable interaction between the inert gas and the adsorbent. After evacuation, the charging cell is pressurized with the assorted refrigerant and left until it reaches an equilibrium state. During charging, it is necessary to keep the charging pressure lower than the saturation pressure of the refrigerant to ensure no condensation is occurred. At this state the initial pressure and temperature in the charging cell are measured before adsorption. Once equilibrium is achieved, the needle valve between the charging and adsorption tank is opened. The pressure and temperature in the adsorption tank are recorded to calculate the uptake of the assorted refrigerant by ensuring thermal equilibrium present. This process was repeated for the each charging step until the high pressure reached. By the use of a specimen, each isotherm was measured at a constant temperature over a range of pressure from 0 to about 5 bar. For each specimen with the known initial dry mass, experiments were performed at constant temperatures chosen at the range of 20 to 60 ºC for pressures up to about 5 bar. Experimental study was performed by the use of commercially available two different type of activated carbon (AC) specimens. Physical characteristics of the adsorbents used in the tests are presented in Table 1. The constant volume variable pressure (CVVP) experimental setup comprises (i) a charging tank with a volume of 3000 cc, (ii) an adsorption tank with a volume of 3000 cc, (iii) temperatures of the both the charging and adsorption tanks are controlled independently by the separate circulating water systems, (iv) a pressure transducer with an uncertainty of 0.15% of full scale and a pressure ranging from 0 to 1.6 MPa, (v) Pt 100 type thermometers with an uncertainty of 0.2% for temperature measurement, (vi) separate sensors with an uncertainty of 0.2% used with the adsorbent species for direct temperature measurement, (vii) a vacuum pump that achieves vacuum level of 0.5 mbar, and (viii) a computer used to control the test system and record the data. The volume of both charging and adsorption tanks are inclusive of the volumes of related piping and valves. II.2. Measurements Prior to adsorption process, the specimen of the adsorbent is placed in an oven for 24 h to desorb any residual gas. The oven temperature is kept constant at 120 C. Before starting adsorption test the system was evacuated to take out any gases and moisture from the bed using a vacuum pump to 5 mbar. The 137 Tab. 1: Physical characteristics of the adsorbents used in the tests. Activated Carbon Type I: ACG Type II: ACP Size (mm) D=4 Density (kg/m 3 ) Micro Pore Volume(cm 3 /g) Specific Surface Area (m 2 /g) Pore Diameter (Å) Shape Granulated Cylindrical Pellets II.3. Assessment of overall uncertainty There are some uncertainties associated with instrumentation, average adsorption cell temperature during adsorption and the void correction. Moreover, certain errors introduced due to the mathematical calculations. It is expected that the overall uncertainty will be within 3%. III. Mathematical approach The starting point for this analysis is the use of Dubinin Astakhov (D A) model of adsorption isotherm in the following form (Saha et al., 2009, El-Sharkawy et al., 2006, Akkimaradi et al., 2001):

152 W = W 0 exp { [ RT E (p s p )]n } (1) with W = Cυ a and W 0 = C 0 υ 0 (2) Here E is the characteristic energy of the assorted adsorbent/adsorbate pair which can be evaluated experimentally. The parameter n is an exponential constant which gives the best fitting of the experimental isotherms. The quantity C denotes the specific mass of adsorption (kg of adsorbate per unit mass of adsorbent), and v a is the specific volume of the adsorbed phase, which is given by υ a = υ b exp ( Ω(T T b )) (3) where Ω = ln(b/υ b )/ (T c T b ) (4) The quantity b denotes the van der Waals volume, vb is the saturated liquid specific volume at the normal boiling point (Saha et al., 2009, El-Sharkawy et al., 2006, Akkimaradi et al., 2001). T is the temperature with suffixes c and b referring to critical and normal boiling points, respectively. The parameter v0 can be obtained by using Eq. (3) at T = 0. Table 2 shows the properties and parameters of the adsorbates used in the present experimental study. Tab. 2: Properties and parameters of the adsorbates. R-134a R-404a Molecular Weight (MW) Boiling Point at 1 atm (Tb) C C Critical Temperature (Tc) C C Critical Pressure (Pc) 4059 kpa 3729 kpa Critical Density kg/m kg/m 3 b vb (m 3 /kg) v0 (m 3 /kg) Ω Eq. (1) can be rewritten as follows: ln p = ln p s E/(RT)[ln(c 0 ν 0 cν a )] 1/n (5) Differentiating Eq. (5) with respect to 1/T for the isosteric conditions (i.e. C is constant). Noting that va is also a function of temperature, one can get the following equation: ln p (1/T) = ln p s (1/T) (E R ) [ln(c 0ν 0 Cν a )] 1 n (ETΩ (nr))[ln(c 0 ν 0 Cν a )] ((1 n))/n (6) Isosteric heat of adsorption is defined by the Clausius Clapeyron relation at constant concentration as: Q ads CC = R( ln p)/ (1/T) (7) and for the heat of vaporization defined as h fg = R( ln p s )/ (1/T) (8) Substituting Eq. 7 and 8 into Eq.6, the following equation for the heat of adsorption can be derived (El-Sharkawy et al., 2007). Q ads = h fg + (E)[ln(C 0 ν 0 Cν a )] 1/n +(ETΩ n)[ln(c 0 ν 0 Cν a )] ((1 n))/n (9) The standard procedure for evaluation of isosteric heat of adsorption as described by Eq.7, is to plot the isosters on ln p versus 1/T plane. Normally, a constancy of slope is observed at temperatures well over the critical point of the adsorbate. As a result the classical treatment of isosteric heat of adsorption being shown as a function of relative uptake is a good approximation for adsorbent adsorbate combinations which broadly follow the Dubinin s isotherms (Saha et al., 2009, El-Sharkawy et al., 2006, Akkimaradi et al., 2001). Due to non-ideality of the gas phase, during an adsorbate molecule uptake to the assorted adsorbent is affected by the pressure and temperature changes (Saha et al., 2009, El-Sharkawy et al., 2006, Akkimaradi et al., 2001, Lin et al., 1999, Saham et al., 2008). In order to consider the effect of pressure and temperature changes, heat of adsorption can be calculated by using the Eq.9. IV. Results and discussions The experimental data was used to evaluate the adsorption parameters for the granulated activated carbon (ACG)-R134a, pellet activated carbon (ACP)-R134a, granulated activated carbon (ACG)-R404a and pellet activated carbon (ACP)-R404a pairs. Derived objection function is optimized by the use of a homemade code based on a genetic algorithm. Table 3 shows that computed values of the adsorption parameters (W 0, C 0, E and n) for the ACG-134a, ACP-R134a, ACG-R404a and ACP-R404a pairs. Tab. 3: Computed adsorption parameters of the different adsorbent-adsorbate pairs. Pairs W0 C0 E n (m 3 /kg) (kg/kg) (kj/kmol) ACG-R134a 0.380x ACP-R134a 0.211x ACG-R404a 0.338x ACP-R404a 0.192x Comparison of the experimental and the computed isotherms of the adsorbent-adsorbate pairs at 30 C are given in Fig.1. It can be seen from the Fig.1. that the measured results and the computed data obtained from D-A equation (Eq. 1) with the parameters given in Table 2 are in a very good agreement. The shapes of the isotherms obtained from the experimental data were similar in all cases and comparable to those reported in the literature for commercially available different adsorbents (Saha et al., 2009, El-Sharkawy et al., 2006, Akkimaradi et al., 2001, Lin et al., 1999, Saham et al., 2008). Then, Eq.5 and Eq.9 with the parameters provided in Table 3 were used to evaluate isosters and isosteric heat of adsorption of the adsorbent-adsorbate pairs. Fig.2 138

153 shows the isosters of adsorption of R134a and R404a on activated carbon specimens for C/C0 = 0.7. It can be seen from the Fig.2 that the variation of ln(p) with 1/T presents lineer variation for each adsorbent-adsorbate pairs. It is interesting to see that the isosters of the adsorbates of R134a and R404a with activated carbon is almost overlap each other, while the slopes of the lines are almost same for the all adsorbent-adsorbate pairs. molecules first penetrate into narrow pores of adsorbent, resulting in a stronger interaction between adsorbate and adsorbent. This implies a higher value of isosteric heat of adsorption at lower loading. After completely filling the smaller pores, adsorbate molecules are gradually accommodated in larger pores, in which the adsorption affinity becomes weaker. Therefore a monotonic decrease in isosteric heat of adsorption as a function of adsorbate uptake. Fig. 1: Comparison of the experimental and the computed isotherms of R134a and R404A on adsorbents at 30 C. Fig. 3: Variation of isosteric heat of adsorption of R134a on granulated activated carbon with temperature. Fig. 2: Obtained isosters of adsorption of R134a and R404a on granulated and pellet activated carbon for C/C0 = 0.7. Fig. 4: Variation of isosteric heat of adsorption of R404a on granulated activated carbon with temperature.. The maximum value of adsorption capacity decreases with the increase of adsorbent temperature. Comparing relative uptakes of the two refrigerants of R134a and R404a, the uptake magnitude of R134a on the same adsorbent is greater than the one of R404a. Meanwhile, the granulated activated carbon (ACG) have higher uptake comparing to the pellet activated carbon (ACP) at the same conditions considered in this study. A careful inspection on the Figs. 3 to 6, it can be seen that the isosteric heat of adsorption decreases with increasing adsorbate uptake for all the cases. The isosteric heat of adsorption varied with the temperature and the maximum value of isosteric heat obtained at the lowest temperature. In addition, the temperature have more effect on isosteric heat, where the maximum value of isosteric heat obtained with low temperature at 0 C. The adsorbate 139 Fig. 5: Variation of isosteric heat of adsorption of R134a on pellet activated carbon with temperature. Comparing the magnitude of the isosteric adsorption heat, ACP-R134a pair has the greatest one among the pairs tested in this study. It is also observed that replacing R134a with R404a do not significant effect on the magnitude of the isosteric adsorption heat on

154 the same adsorbent as seen Figs. 3 and 4 for ACG, and also Figs. 5 and 6 for ACP. Acknowledgements The authors gratefully acknowledge that this study is supported by Scientific and Technology Research Council of Turkey (TÜBİTAK), under project number: 112M163, and this study is also supported by Scientific Research Projects Funds of Uludağ University, under project number: KUAP(MH)-2015/60. Fig. 6: Variation of isosteric heat of adsorption of R404a on pellet activated carbon with temperature. In addition to all, the magnitude of the heat of adsorption is greater than that of the enthalpy of vaporization of R-134a and R-404a in the all range of the experimental tests performed as seen from the Figs. 3 to 6. V. Conclusions In this study, the adsorption properties of R-134a and R-404a on commercially available two different type specimens of activated carbon for adsorption process has been experimentally studied. The experiments were conducted over a temperature range from 20 C to 60 C and pressure up to about 5 bars. The measured experimental data were used to obtain the parameters in the Dubinin-Astakhov (D-A) equation for corresponding adsorption process. Adsorption characteristics such as isotherms were evaluated from the D-A equation with the obtained specific parameters for the each adsorbate-adsorbent pairs. Further, the isosteric heat of adsorption were obtained, and correlation parameters were provided for the each adsorbate-adsorbent pairs investigated. Comparison between the correlated results and the experimental data shows very good agreement. It is observed that the adsorption capacity per kg of adsorbent increases rapidly with increasing relative pressure at the beginning of the adsorption process. The maximum value of adsorption capacity decreases with the increase of adsorbent temperature. The isosteric heat of adsorption varied with the temperature and the maximum value of isosteric heat obtained at the lowest temperature. Comparing two refrigerants of R134a and R404a, the uptake magnitudes of R134a on the same adsorbent is greater than the one of R404a. Meanwhile, the granulated activated carbon (ACG) have higher uptake comparing to the pellet activated carbon (ACP) at the same conditions. In addition, it is also observed that the magnitude of the heat of adsorption is greater than that of the enthalpy of vaporization of R134a and R404a in the range of experimental conditions studied. This aspect is important in the design of thermal compressors in which the coolant requirements for removing the enthalpy of adsorption have to be assessed. 140 Nomenclature ACG : Activated carbon granulated ACP : Activated carbon pellet b : The van der Waals volume C : Adsorption uptake per kg of adsorbent (kg adsorbate/kg adsorbent) C0 : Maximum adsorption uptake per kg of adsorbent (kg adsorbate/kg adsorbent) E : Characteristic energy of adsorption pair (kj/kmole) MW : Moleculer weight (kg/kmole) p : Pressure (kpa or bar) R : Universal gas constant (8.314 kj/kgk) Qads : Isosteric adsorption heat (kj/kg) T : Temperature (K or ºC) va : Adsorbed phase specific volume (kg/m 3 ) vb : Saturated liquid specific volume at the normal boiling temperature (kg/m 3 ) n : Exponential constant W : Adsorbed volume per unit mass of adsorbent (m 3 adsorbate/kg adsorbent) W0 : Maximum adsorbed volume per unit mass of adsorbent (m 3 adsorbate/kg adsorbent) Subscripts a : adsorbed phase b : boiling point c : critical point CC : constant concentration ads : adsorption fg : vaporization enthalpy 0 : maximum s : saturation References Akkimaradi, B.S., Prasad, M., Dutta, P., Srinivasan, K., Adsorption of 1,1,1,2-tetrafluoroethane on activated charcoal, J. Chem. Eng. Data, 46, (2001). Banker, N.D, Srinivasan, K., Prasad, M. Performance analysis of activated carbon + HFC 134a adsorption coolers, Carbon, 42, (2004). Chakraborty, A., Saha, B.B., Koyama, S., Ng, K.C., On the thermodynamic modeling of the isosteric heat of adsorption and comparison with experiments, Appl. Phys. Lett., 89, , (2006). Chan, C.K., Tward, E., Boudale, K.I., Adsorption isotherms and heat of adsorption of hydrogen, helium, neon and nitrogen on activated charcoal, Cryogenics,

155 24, (1984). El-Sharkawy, I.I., Kuwahara, K., Saha, B.B., Koyama, S., Ng, K.C., Experimental investigation of activated carbon fibers/ethanol pairs for adsorption cooling system application, Appl. Therm. Eng., 26, (2006). Reviews, 14, (2010). Wang R.Z., Wang B.Q., Adsorption mechanism and improvements of the adsorption equation for adsorption refrigeration pairs, International Journal of Energy Research, 23, (1999). Lin, S.H., Lin, R.C., Prediction and experimental verification of HFC-134a adsorption by activated carbons. J. Environ. Sci. Health, 34 (1), (1999). Prakash, M., Mattern, A., Prasad, M., Sant, R., Subramanya, P., Srinivasan, K., Adsorption parameters of activated charcoal from desorption studies, Carbon, 38 (8), (2000). Riffat, S.B., Williams, M.D., Corr, S., Adsorption heat pump using HFC refrigerants, International Journal of Energy Research, 21, (1997). Saha B.B., Habib K., El-Sharkawy I.I., Koyama S., Adsorption characteristics and heat of adsorption measurements of R-134a on activated carbon, International Journal of Refrigeration, 32, (2009). Saham B,B,, Chakraborty A,, Koyama S,, Yoon S,H, Mochida I,, Kumja M, Yap C., Ng K.C., Isotherms and thermodynamics for the adsorption of nbutane on pitch based activated carbon, International Journal Heat and Mass Transfer, 51, (2008). Saha, B.B., Koyama, S., El-Sharkawy, I.I., Habib, K., Srinivasan, K., Dutta, P., Evaluation of adsorption parameters and heats of adsorption through desorption measurements. J. Chem. Eng. Data 52 (6), (2007). Solmuş I., Kaftanoğlu B., Yamalı C., Baker D., Experimental investigation of a natural zeolite water adsorption cooling unit, Applied Energy, 88, (2011). Solmuş I., Yamalı C., Kaftanoğlu B., Baker D., Çağlar A., Adsorption properties of a natural zeolite water pair for use in adsorption cooling cycles, Applied Energy, 87, (2010). Solmuş, I., Yıldırım, C., Theoretical analysis of the performance of an adsorption cooling system for various working pairs, J. of Thermal Science and Technology, 34(2), (2014). Wang, L.W., Wang, R.Z., Oliveira, R.G., A review on adsorption working pairs for refrigeration, Renewable and Sustainable Energy Reviews, 13, (2009). Wang, D.C., Li, Y.H., Li, D., Xia, Y.Z., Zhang, J.P., A review on adsorption refrigeration technology and adsorption deterioration in physical adsorption systems, Renewable and Sustainable Energy 141

156 Performance Investigation of a Geothermal Powered Organic Rankine Cycle for Natural Working Fluids Mustafa Alptekin 1,2*, Onder Kizilkan 1, Ahmet Kabul 1, Resat Selbas 1 1 Suleyman Demirel University, Faculty of Technology, Department of Energy Systems Engineering, Isparta, 32260, Turkey 2 Hakkari University, Faculty of Engineering, Department of Mechanical Engineering, Hakkari, 30000, Turkey. * Abstract Importance of geothermal energy, which is one of renewable energy resources, has been rapidly increasing in our country as well as all over the world. As power generation can be done using different resources, electric generation can be performed from a power plant with organic Rankine cycle, which operates low temperature, from inactive geothermal resources in our country. In this study, energy and exergy analyses of a geothermal powered Organic Rankine Cycle (ORC) were conducted, and net work output, total exergy destruction rate, thermal and exergy efficiencies of overall system were calculated using original data of a geothermal power plant in Denizli province. There are two organic Rankine cycles in the system, and the analyses were performed for organic working fluids n-pentane, R245fa and R600a. it was investigated effect of condenser pressure and turbine inlet temperature on system performance. As the condenser pressure increases, the thermal and exergy efficiencies decreases, and the turbine inlet temperature is directly proportional with system performance. Among the working fluids examined, R600a and R245fa demonstrate the best exergetic performance. For R245fa, the total exergy destruction rate and exergy efficiency are calculated to be 3675 kw and 76%, respectively. Keywords: Geothermal energy, energy and exergy analysis, Organic Rankine Cycle (ORC), natural fluids. I. Introduction In the recent years, there has been a significant increase in the usage of renewable and low-grade waste heat. The geothermal energy is considered as an alternative source instead of fossil fuels since it is reliable and one of least-expensive renewable energy source. The utilization of this energy may be by using organic Rankine cycle (ORC), which converts this energy to useful power. ORC has some advantages such as environment-friendly, safety, system components are available and high flexibility (El-Emam and Dincer, 2013; Long et al., 2014). Solar radiation, biomass combustion, geothermal energy and industrial waste heat Kaska, 2014] can be used as required heat source of an ORC (Kaska, 2014). There are many studies about ORC in the literature. Heberle and Brüggemann (2010) performed the exergy analysis of a combined heat and power generation system for low grade geothermal resources (below about 180 C). They investigated the first and second law efficiencies of the system for different operating conditions, and found that isopentane and R227ea can be preferred in series and parallel circuits, respectively. Kanoglu, (2002) carried out exergy analysis of a dual-level binary geothermal power plant, which had a power output of 12.4 MW, using actual plant data of the power plant. As a result of, he determined that the condenser had the highest exergy destruction rate, and the thermal and exergy efficiencies were found to be 5.8% and 29.1%, respectively. Unverdi and Cerci (2013) investigated the performance of Germencik geothermal power plant. The working fluid was water and geothermal water had a source temperature of 205 C. They compared this system with the other geothermal power plants in the world, and the exergy efficiency of the overall system was calculated as 35.4%. El-Emam and Dincer (2013) performed exergy and exergoeconomic analyses of geothermal regenerative organic Rankine cycle with optimization. They conducted the effect of operating parameters on the system energetic and exergetic efficiencies and economic parameters, and found the energy and exergy efficiency values to be 16.37% and 48.8%, respectively, for a net out power of 5 MW. Thermodynamic and economic analysis for the pre-feasibility of a binary geothermal power plant were performed by Budisulistyo and Krumdieck (2015). They conducted thermodynamic and economic analyses for key cycle design options and component selection parameters, and used n-pentane, R134a and R245fa as working fluid. The profitability analysis was done for the top three options by them. They stated that a standard Rankine cycle with a 2-stage turbine using n-pentane is the most thermo-economical design for the particular brine resource and re-injection conditions. Coskun et al. (2011) proposed a modified exergoeconomic model for geothermal power plants using exergy and cost accounting analyses in a case study for Tuzla geothermal power plant system (Tuzla GPPS), which has a total installed capacity of 7.5 MW. They conducted the analysis using actual system data to assess energy and exergy efficiencies, exergy losses 142

157 and loss cost rates. Besides, they determined that exergy efficiency values vary between 35% and 49%, and studied to provide a more comprehensive evaluation of the system six new exergetic cost parameters. Dagdas et al. (2005) performed thermodynamic optimization of a power plant using actual data, and obtained some important results. They found the optimum flashing pressure of 200 kpa, and determined isobutene as a working fluid. In addition, they calculated to be 8.80% and 38.58% the first and second law efficiencies of the power plant, respectively. Akpinar and Hepbasli (2007) evaluated, constructed and tested with a comparison of exergetic (second law analysis) analysis of two Ground Source Heat Pumps (GSHPs) in Turkey based on the actual operational data by using entropy and exergy balance equations. They determined exergy (second law) efficiency values for both systems and exergy destructions in each of the system components to assess the improvement potential, and indicated that this method presented here can be applied to other GSHP systems worldwide as a useful tool. Kecebas and Gokgedik (2015) carried out both conventional and advanced exergy analyses of an existing geothermal binary power system. Therefore, in-depth information was collected about the irreversibilities in the system and its parts. They used to simulate the Bereket Geothermal Power Plant, in Denizli, the Engineering Equation Solver (EES) and GateCycle software packages. They found that condensers have the highest improvement potential for both conventional and advanced exergy analysis, and the modified exergy efficiency and the total system efficiency are to be 18.26% and 9.60%, respectively, in the real conditions. Yamankaradeniz (2016) performed thermodynamic performance assessment of a geothermal district heating system (GDHS) by using advanced exergy analysis to identify the interactions among system components and the potential for improvement. He applied new exergetic parameters to the Bursa GDHS in Turkey. He concluded that the advanced exergetic analysis is a more meaningful and effective tool than the conventional one. In addition, he found the exergy efficiencies are 25.24% and 26.34% for the conventional and advanced ones, respectively. Tan and Kecebas (2014) assested thermodynamic and economic evaluation of a geothermal district heating system (GDHS) using advanced exergy-based methods. They splited into endogenous/exogenous and unavoidable/avoidable parts the exergy destruction and the total operating cost within each component of the system by the help of the advanced exergetic and exergoeconomic analyses. They found that the exergetic efficiency and the exergoeconomic factor of the overall system for the Sarayköy GDHS is 43.72% and 5.25%, respectively, according to the conventional tools while these values are 45.06% and 12.98%, respectively, according to the advanced tools. Yildirim and Ozgener (2012) performed a review study about thermodynamics and exergoeconomic analysis of two geothermal power plants in Turkey. They investigated the effects of thermal fluids used in 143 power plants on energy and exergy efficiencies. They presented improvement suggestions, and conducted exergoeconomic analyses while power plants investment costs and equipment maintenance costs are taken into consideration. Liu et al. (2015) carried out parametric optimization and performance analyses of geothermal organic Rankine cycles using R600a/R601a mixtures as working fluids. They optimized evaporator and condenser pressures and cooling water temperature rise, and analyzed ORC power output, parasitic power consumption, heat exchanger areas and turbine sizes using R600a/R601a. They found that a geothermal ORC using R600a/R601a generates 4 11% more power than with pure R600a, and determined that the evaporator and condenser area per unit power output using R600a/R601a are higher than that using pure R600a or R601a. In this study, thermodynamic analysis of a geothermal powered ORC system is investigated using actual data for different natural working fluids. The working fluids are selected to be R245fa, and R600a while the system uses n-pentane as working fluid in actual case. Fot these there fluids, the thermal efficiency, exergy efficiency and exergy destructions of the ORC based power plant are found. Besides, the effects of turbine inlet temperature and condenser pressure on system performance are also parametrically analyzed. II. System Description In Fig. 1, schematic view of the geothermal powered ORC power generation system is presented. There are two different cycles in the system, and condenser is a water-cooled condenser. The same fluid operates in both power cycles and as mentioned earlier, analyses are performed for different working fluids which are R245fa, R600a and n-pentane. Fig.1: Schematic view of ORC system Firstly, geothermal water which has have a temperature of 145 C transfers to some of amount of its heat to the organic working fluid in evaporator at first and second cycles, respectively. Secondly, it splits in half, and transfers to rest amount of its heat to the organic working fluid in preheaters at the first and second cycles. Finally, it is pumped back to underground as waste water after it exits from preheaters. The organic working fluid enters to turbine as saturated vapor in both two cycles after it

158 gets heat of the geothermal water in the preheaters and evaporators. After the working fluid leaves the turbine as superheated vapor at condenser pressure and leaves the condenser as saturated liquid. The working fluid is circulated by pump, and enters the preheater as compressed liquid. For the thermodynamic assestment of the system, the some assumptions are made as follows: 1) All the processes are assumed as steady state. 2) The pumps and turbines are adiabatic. 3) The heat transfer to/from ambient and pressure losses in the piping system and in the preheaters, evaporators, condensers of ORC system are neglected 4) The working fluid at the inlet of ORC pump is assumed as saturated liquid. 5) Potential and kinetic energy variations are neglected. 6) The dead state pressure P0 and temperature T0 are considered to be kpa and 27 C, respectively. III. Thermodynamic Evaluation Energy and exergy analysis generally involves applying the first and second laws of thermodynamics and the principles of conservation of mass, while energy analysis usually excludes considerations of the second law of thermodynamics. Neglecting kinetic and potential energies, the conservation of mass for steady-state processes can be expressed as follows (Dincer and Rosen, 2007) m in = m out (1) The first law of thermodynamics is an express of energy principle. It is expressed for steady state processes as follows (Cengel and Boles, 2007; Bejan, 1997); Q + (m h) in = W + (m h) out (2) Neglecting chemical, kinetic and potential exergies, exergy balance in a control volume in which a steady state process occurs can be written as (Dincer and Rosen, 2007; Akpinar and Hepbasli (2007)); E x Q E x W = (m ε) out (m ε) in + T 0 S gen (3) where; E x dest = T 0 S gen (4) where E x Q, E x W and ε represent exergy of heat, exergy of work and thermomechanical exergy (flow exergy), respectively. These expressions are shown as follows Kaska (2014); E x Q = Q ( T T 0 T ) (5) E x W = W (6) 144 ε = (h h 0 ) T 0 (s s 0 ) (7) S in + S gen = S out (8) where 0 subscript expressed reference conditions. The first and second law efficiencies of the all system are calculated as follows (Dincer and Rosen, 2007); η I = W Q (9) η II = E x W E x Q (10) The above equations were applied to the geothermal powered ORC system, and were obtained mass, energy and exergy balance equations for each component. Tab. 1: Input values to the system Parameters Values Pump isentropic efficiency 85 % Turbine isentropic efficiency 85 % Condenser temperature 30 C Turbine inlet temperature 100 C Geothermal water inlet temperature 145 C Inlet temperature to condensers of cooling water 20 C Outlet temperature from condensers of cooling water 27 C Pressure of geothermal water 600 kpa Mass flow rate of working fluid in the first ORC 58.6 kg/s Mass flow rate of working fluid in the second ORC 58.4 kg/s Inlet mass flow rate of geothermal water kg/s Preheater I Capacity 8770 kw Preheater II Capacity 8555 kw Ambient pressure kpa Ambient temperature 25 C IV. Results and discussions Analyses were made using Engineering Equation Solver (EES) program. The calculated thermodynamic data of the ORC were determined for n-pentane, R245fa and R600a working fluids. Table 2 shows at each location of the ORC system thermodynamic properties such as temperature, enthalpy and entropy for n-pentane. The analyses were performed using actual data of geothermal power plant in Denizli. IV.1. Effect of condenser pressure (Pcon) The geothermal powered ORC system was thermodynamically modelled and conducted the first and second law of thermodynamics analyses. The Figures 2 to 5 illustrate the effect of condenser pressure variation on net work output, total exergy destruction rate, thermal and exergy efficiencies. The turbine inlet temperature and isentropic efficiencies of pumps and turbines were kept constant as 100 C and 85 %, respectively. Figures 2 and 3 show the net work output and thermal efficiency of overall system as function of condenser pressure for different organic working fluids. When the condenser pressure increases from 150 kpa to 450 kpa, both the net work output and thermal efficiency decrease for all working fluids. While R600a has the highest thermal efficiency,

159 n-pentane has the lowest efficiency, respectively. The maximum net work output and thermal efficiency was obtained to be about 10 MW and 21%, respectively. Net work output - kw Tab. 2: Data of the ORC system for n-pentane n-pentane R245fa R600a P con, kpa Fig. 2: Net power output as a function of condenser pressure for different organic working fluids Thermal efficiency (h) m P Ref. Substance T ( C) (kg/s) (kpa) h (kj/kg) s ε (kj/kg.k) (kj/kg) 1 n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane n-pentane Water Water Water Water Water Water Water Water Water Water Water Water Water Water Water Water ,24 0,2 0,16 0,12 0,08 0,04 n-pentane R245fa R600a P con, kpa Fig. 3: Thermal efficiency as a function of condenser pressure for different organic working fluids The effect of condenser pressure on the exergy destruction rate and exergy efficiency of ORC system. The effect of condenser pressure on the exergy destruction rate and exergy efficiency of ORC system is illustrated in Figs. 4 and 5. While the total exergy 145 destruction rate increases with increasing of condenser pressure, the exergy efficiency decreases as the condenser pressure increases for all working fluids. It is observed that the exergy destruction rate decreases from 7629 kw to kw when the condenser pressure increases for n-pentane. Moreover, this fluid have the highest destruction rate. R600a have the largest exergy efficiency and the lowest total exergy destruction rate. Although increment in destruction rate for R600a is larger than that of R245fa, R600a have larger than R245fa exergy efficiency at high pressures. Total exergy destruction rate - kw Exergy efficiency n-pentane R245fa R600a P con, kpa Fig. 4: Exergy destruction rate as a function of condenser pressure for different organic working fluids 1 0,9 0,8 0,7 0,6 0,5 0,4 0,3 0,2 0,1 n-pentane R245fa R600a P con, kpa Fig. 5: Exergy efficiency as a function of condenser pressure for different organic working fluids IV.2. Effect of turbine inlet temperature (Tturb,in) The effect of turbine inlet temperature variation on net work output, total exergy destruction rate, thermal and exergy efficiencies are presented in Figs The Figures 6 and 7 illustrate variation of the net work output and the thermal efficiency as the turbine inlet temperature increases 80 C to 110 C. Both the net work output and efficiency increase with increasing turbine inlet temperature. The highest thermal efficiency is obtained for n-pentane while the lowest one is for R245fa. Furthermore, the thermal efficiency of n-pentane and R245fa increase from about 10% to 14% and from about 10% to 14%, respectively, while the turbine inlet temperature increases.

160 The effect of turbine inlet temperature on the total exergy destruction rate and exergy efficiency of ORC system is presented for different working fluids in Figs. 8 and 9. While the total exergy destruction rate decreases for all working fluid with increasing turbine inlet temperature, the exergy efficiency of overall system increases with increasing turbine inlet temperature. It is observed that the lowest exergy destruction rate occurs for the case of R245fa. Therefore, this fluid have the highest exergy efficiency. n-pentane and R600a are compared to each other that for n-pentane destruction rate is higher than that of R600a at low turbine inlet temperatures. However, it is noticed that exergy destruction in n-pentane is lower than that of R600a at high temperatures. Net work output - kw n-pentane R245fa R600a T turb in, C Fig. 6: Net power output as a function of turbine inlet temperature for different organic working fluids Thermal efficiency (h) 0,16 0,14 0,12 0,1 n-pentane R245fa R600a 0, T turb in, C Fig. 7: Thermal efficiency as a function of turbine inlet temperature for different organic working fluids Total exergy destruction rate - kw n-pentane R245fa R600a T turb in, C Fig. 8: Total exergy destruction rate as a function of turbine inlet temperature for different organic working fluids 146 Exergy efficiency Fig. 9: Exergy efficiency as a function of turbine inlet temperature for different organic working fluids V. Conclusions Geothermal energy powered an Organic Rankine Cycle is evaluated in terms of first and second laws of thermodynamics. Energy and exergy analysis of ORC system are carried out using actual power plant data. The main results of the study can be summarized as follows: The thermal and exergy efficiencies are directly proportional with turbine inlet temperature whereas they decrease with increasing condenser pressure. R600a and R245fa have the best performance for effect of condenser pressure and turbine inlet temperature, respectively. For condensation pressure of 150 kpa and turbine inlet temperature of 100 C, the thermal and exergy efficiencies of R600a are 20.7% and 91% while those of n-pentane are 9% and 52.7%, respectively. It is found that turbine inlet temperature has significant effect on both system performance. Higher thermal and exergy efficiencies are obtained by increasing turbine inlet temperature. Used working fluid n-pentane in actual plant, has the lower performance than R600a and R245fa for two parameters. Therefore, it is proposed that R600a or R245fa should be used. Nomenclature E x h m P ORC Q s S T W : Exergy (kw) : Specific enthalpy (kj/kg) : Mass flow rate (kg/s) : Pressure (kpa) : Organic Rankine Cycle : Heat load (kw) : Specific entropy (kj/kg.k) : Work (kw/k) : Temperature (C) : Work (kw) Greek letters ε : Specific exergy (kj/kg) η I : Thermal efficiency : Exergy efficiency η II 0,78 0,77 0,76 0,75 0,74 0,73 n-pentane R245fa R600a 0, T turb in, C

161 Subscripts con : Condenser dest : Destruction gen : Generation in : Inlet out : Outlet 0 : Referance state References Akpinar E.K. and Hepbasli A., A comparative study on exergetic assessment of two ground source (geothermal) heat pump systems for residential applications, Building and Environment, 42, (2007). Bejan A., Advanced Engineering Thermodynamics, Wiley, New York, (1997). Budisulistyo D. and Krumdieck S., Thermodynamic and economic analysis for the pre-feasibility study of a binary geothermal power plant, Energy Conversion and Management, 103, (2015). Coskun C., Oktay Z. and Dincer İ., Modified exergoeconomic modeling of geothermal power plants, Energy, 36, (2011). Long R., Bao Y.J., Huang X.M. and Liu W., Exergy analysis and working fluid selection of organic Rankine cycle for low grade waste heat recovery, Energy, 73, (2014). Tan M. and Kecebas A., Thermodynamic and economic evaluations of a geothermal district heating system using advanced exergy-based methods, Energy Conversion and Management, 77, , (2014). Unverdi M. and Cerci Y., Performance analysis of germencik geothermal power plant, Energy, 52, (2013). Yamankaradeniz N., Thermodynamic performance assessments of a district heating system with geothermal by using advanced exergy analysis, Renewable Energy, 85, (2016). Yildirim D. and Ozgener L., Thermodynamics and exergoeconomic analysis of geothermal power plants, Renewable and Sustainable Energy Rewiews, 16, (2012). Cengel Y.A. and Boles M.A., An Engineering Approach Thermodynamics, Fifth Edition, 946, (2007). Dagdas A., Ozturk R. and Bekdemir S., Thermodynamic evaluation of denizli kizildere geothermal power plant and its performance improvement, Energy Conversion and Management, 46, (2004). Dincer I. and Rosen M.A., Environment and sustainable development, Elsevier Science, 472, (2007). El-Emam R.S. and Dincer I., Exergy and exergoeconomic analyses and optimization of geothermal organic rankine cycle, Applied Thermal Engineering 59, (2013). Kanoglu M., Exergy analysis of a dual-level binary geothermal power plant, Geothermics, 31, (2002). Kaska O., Energy and exergy analysis of an organic Rankine for power generation from waste heat recovery in steel industry, Energy Conversion and Management, 77, (2014). Kecebas A. and Gokgedik H., Thermodynamic evaluation of a geothermal power plant for advanced exergy analysis, Energy, 1-10 (2015). Liu Q., Shen A. and Duan Y., Parametric optimization and performance analyses of geothermal organic rankine cycles using r600a/r601a mixtures as working fluid, Applied Energy, 148, (2015). 147

162 An Experimental Investigation on Exergy Analysis of an Ejector Expansion Refrigeration System Nagihan Bilir Sag 1*, Halil Kursad Ersoy 2, Arif Hepbasli 3 1,2 Department of Mechanical Engineering, Faculty of Engineering, Selcuk University, Alaeddin Campus, Konya, Turkey 3 Department of Energy Systems Engineering, Faculty of Engineering, Yasar University, Bornova, Izmir, Turkey * Abstract Usage of ejector as an expander for expansion work recovery in the conventional refrigeration cycle was experimentally investigated. In order to identify the magnitudes and locations of irreversibilities within the components of the ejector expansion refrigeration cycle, an exergy analysis was employed. The exergetic performance of the ejector refrigeration system was compared with that of a conventional vapor compression refrigeration system of the same cooling capacity under the same external conditions. It was found that the ejector expander system exhibited a lower total irreversibility and a higher exergy efficiency in comparison with the conventional system for all operating temperatures. When the ejector was used as the expander in the refrigeration system, the total irreversibility was lower than in the conventional system by %, while the exergy efficiency values were % higher than in the conventional system. Keywords: Refrigeration, ejector, throttling loss, exergy efficiency, irreversibility I. Introduction Energy consumption is increasing day by day in the world. So, the efficient use of energy is very important for the future of humanity. There are a lot of researches on efficiency improvement of systems that consume energy. One of the goals of these researches is to increase the energy efficiency of conventional vapor compression refrigeration systems. Recently conducted studies to improve the performance of vapor compression refrigeration systems have concentrated on reducing the throttling losses that increase irreversibility in the expansion valve. For this purpose, using a simple and low-cost ejector with no moving parts instead of expansion valve is a topic recently investigated in order to find out whether the performance of the system would increase. The idea of using an ejector instead of expansion valve was, for the first time, introduced by Gay in 1931 (Pottker, 2012). Kornhauser (1990) determined that using ejector in R12 conventional refrigeration system improves the cooling coefficient of performance (COP) by 21%, theoretically. However, an improvement of only % could be achieved experimentally (Menegay and Kornhauser, 1996). In a study by Bilir and Ersoy (2009), the performance of R134a ejector refrigeration system according to the conventional system could be theoretically improved up to of 22.3%. Additionally, in this study, it was calculated that even under off-design conditions, when an ejector was used in a refrigeration system, the COP of the system exhibited a higher value than a conventional cycle. In another study done by Ersoy and Bilir (2010), the exergy efficiency of the ejector system, irreversibility analysis, and the effects of ejector components on the system performance were theoretically investigated. In this study, it was found that while the efficiency of the ejector components increased, the system performance increased and the area ratio of the ejector decreased. It was determined that when ejector was used as an expander in the transcritic CO2 refrigeration cycle, the total irreversibility of the system reduced by 39.1% (Ersoy and Bilir, 2012). Tas et al. (2015) made a detailed review study on the ejector refrigeration cycle. They stated that the important factors affecting the system performance were operating conditions and refrigerant type. In a theoretical study, where R134a was used as a refrigerant in a bus air condition system that incorporated an ejector and double evaporator, Unal and Yilmaz (2015) reported that the system performance had a 15% higher COP than the conventional double-evaporator refrigeration system. There are very few experimental studies on the use of ejector instead of expansion valve in the conventional vapor refrigeration system with R134a as a refrigerant. A study by Harrell and Kornhauser published in 1995 stated that the COP improvement of an experimentally tested R134a refrigeration system varied between 3.9% and 7.6% (Pottker, 2012). The reason why this improvement remained low was associated with the ejector design based on singlephase flow knowledge and lack of two-phase ejector flow data. Experimental studies on the ejector refrigeration system were conducted under the leadership of Wongwises (Disawas ve Wongwises, 2004; Wongwises ve Disawas, 2005; Chaiwongsa ve Wongwises, 2007; Chaiwongsa ve Wongwises, 2008). 148

163 In these studies, the evaporator was wet-type and ejector partially recirculated the refrigerant on the low pressure side. However, when using the ejector, the amount of the improvement of the cooling performance was not clearly stated in these studies. An ejector-expander refrigeration cycle with noseparator and dual evaporator was experimentally compared with a conventional dual-evaporator refrigeration cycle by Lawrence and Elbel (2013). It was also found that the ejector refrigeration system with dual evaporator exhibited an exergy efficiency that was 8.5% higher than that of the conventional dual evaporator refrigeration system. Yılmaz (2015) experimentally found that the coefficient of performance of the dual-evaporator refrigeration cycle with the ejector increased by 8% compared to that of the conventional dual-evaporator refrigeration cycle. Bilir Sag et al. (2015) experimentally investigated the coefficient of performance of the ejector and the conventional refrigeration systems for fresh food refrigerator (at a condenser temperature of 40 ºC and evaporator temperature of 5 ºC) under the same external conditions. They found that the coefficients of performance were % higher than that of the conventional system while the exergy efficiency values were % higher than in the conventional system. In the literature, there have been a limited number of experimental investigations of the exergy efficiency of the R134a ejector-expander refrigeration systems. The main objective of this contribution is to experimentally carry out an exergetic comparison between two refrigeration cycles, a conventional vapor-compression refrigeration cycle (VCRC) and an ejector-expander refrigeration cycle (EERC), using R134a refrigerant and that the system should be suitable for an automobile air conditioner (condenser temperature at 60 ºC and evaporator temperature at 10 ºC). II. Experimental Setup A schematic view of an ejector expander refrigeration cycle, along with its P-h diagram, is given in Fig. 1. The way, in which this system operates, was presented in the previous studies of two of the authors of this article (Bilir and Ersoy, 2009). The experimental refrigeration setup shown in Fig. 2 was used for measuring the COP of both the VCRC and EERC under the same cooling capacity and external operating conditions. The same compressor, evaporator and condenser were used in both refrigeration systems. The differences of the EERC in comparison with the VCRC include the use of an ejector instead of an expansion valve as the main throttling element, the use of liquid-vapor separator at the diffuser exit of the ejector, and the use of a small expansion valve to ensure the pressure drop ( P 70 kpa) for a small throttling between the separator and the evaporator. Explanation and operation of the experimental set together with the details of the elements and properties of the measuring tools used in the experimental set are described in the study by Ersoy and Bilir Sag (2014). In investigating operating conditions, the software was written using the Engineering Equation Solver (EES) software package (Klein, 2011) in accordance with the mathematical model (Bilir and Ersoy, 2009; Ersoy and Bilir, 2013) of the ejector expander system. The important dimensions of ejector geometry (diameters of the motive nozzle throat and exit, the constant area mixing chamber diameter) were determined. Dimensions of the constant area ejector can be seen in Fig 3. 8 Qc 9 Condenser 5 P P cp Primary flow 2 Ejector 2 b 1 b 3 m Secondary flow Compressor Expansion Evaporator valve m b 2 b h 10 Q e 11 Fig. 1: Ejector expander refrigeration cycle with its P-h diagram. 149

164 P T T Inverter Power meter 4 P T 5 T P Compressor Oil separator Sight glass CO/EC P T Filter Dryer Condenser T Water heater 9 8 FM EO/CC CO/EC Sight glass City water T P T Accumulator Ejector P T FM T 6 Separator EO/CC 3 P 1 EO/CC 2 EO/CC P T T 10 FM EO/CC Expansion valve 7 CO/EC P T Evaporator T T 11 FM T P FM TCP Temperature sensor Pressure sensor Flow meter Temperature control panel Expansion valve TCP Electrical heater Brine tank Pump CO/EC CO/EC EO/CC On in conventional mode and Off in ejector mode On in ejector mode and Off in conventional mode Fig. 2: Schematic view of experimental setup for the VCRC and EERC Suction nozzle Constant-area mixing chamber Diffuser Flow from condenser (Primary flow) o 2 1.7º o Stream to separator (mixed flow) Primary nozzle Flow from evaporatory (Secondary flow) Fig. 3: Dimensions of the ejector and its schematic diagram (in mm) 150

165 III. Exergy Analysis The main purpose of using the ejector as an expander is to reduce the irreversibility of the conventional refrigeration system. Exergy analysis is important to determine the reduction of irreversibility for every element and the whole system of refrigeration cycle that uses an ejector instead of an expansion valve. Exergy analysis is also crucial in terms of determining whether there is any improvement in the exergy efficiency in comparison with the conventional system while it predicts the distribution, source and magnitude of irreversible losses in energy systems and hence, provides guidelines for efficient energy usage (Kotas, 1985). In the expansion valve for EERC: m m m 6 7 e (21) E x ( Ex E 7 ) (22) exp, EERC 6 x exp, Ex / E x (23) EERC 7 6 In the separator: m m cp m 3 e (24) m m 4 6 m cp m e (26) x Ex Ex E ) E sep ( 3 4 x6 (27) sep ( Ex 4 Ex 6 ) / E x3 (28) (25) The physical exergy of every point in a cycle is expressed as: E x m ( h h ) T ( s ) (1). 0 0 s0 The reference state values were taken to be ambient pressure of 100 kpa and the temperature of 27 ºC (Ersoy ve Bilir, 2010). The exergy destructions and efficiencies of each element of the cycle can be determined from the following equations (Kotas, 1985; Dincer and Rosen, 2007; Alsuhaibani et al.,2012). In the compressor: m m m 4 5 cp (2) E xcp ( Ex Ex ) W 4 5 cp (3) ( E x Ex 4 ) / (4) cp 5 W cp In the condenser: m 5 (5) m 1 m cp m 9 m w m 8 (6) E x c ( Ex 5 Ex 1) ( Ex 8 E x9) (7) Ex Ex ) /( Ex E ) (8) c ( x5 In the evaporator: m 2 m 7 m e (9) m 10 m 11 m br (10) E x e ( Ex 7 Ex 2) ( Ex 10 E x11) (11) Ex Ex ) /( Ex E ) (12) e ( x2 In the ejector: m m 1 cp (13) m m 2 e (14) m m cp m 3 e (15) x ( Ex Ex E 3 ) (16) E ej 1 2 x Ex /( Ex E 2 ) (17) ej 3 1 x In the expansion valve for VCRC: m m m 1 7 e (18) E x ( Ex E 7 ) (29) exp, VCRC 1 x (20) exp, VCRC Ex7 / Ex1 151 The total exergy destruction in the ejector expansion refrigeration system is thus calculated: E x Ex Ex Ex Ex Ex E, x (29) tot des cp c e When the total exergy destruction is calculated for a VCRC, and E are taken to be zero. E x ej xsep The exergy efficiency is defined as the ratio of total exergy output rate to the total exergy input rate (Aroraa and Kaushikb, 2008; Matawala, 2012). E x Ex 1 ( Ex Ex ) W, (30) sys o i tot des ej loss exp Here, the total exergy destruction rate cp sep E x tot, des is the amount lost because of irreversibility and the unused exergy. The exergy loss rate E is the exergy xloss rate dissipated from the system to the environment, which could be used by other systems (Matawala, 2012). The exergy loss from the system to the environment due to various causes (discharging of the condenser cooling water to the ambient and heat losses due to piping) is computed: E x 9 (31) loss ( Ex Ex 8) Ex heatloss IV. Results and Discussion In this study, the exergetic evaluations of the experimental results, that were obtained from the use of the ejector designed and are shown in Fig. 3, were made for a condenser temperature of 60 C and an evaporator temperature of 10 C (design condition). In order for the ejector system to operate at the design condition, the temperatures and volumetric flow rates of the condenser inlet water and brine fluid were kept constant. Thus, the system was set to the necessary external conditions. In order to carry out the experiments, the external operating conditions were set as follows: the condenser inlet water temperatures were 40.3, 42.9 and 46.2 ºC, the condenser water volumetric flow rate was 0.44 m 3 /h, the temperature and volumetric flow rate of brine fluid were 20 ºC and 0.58 m 3 /h,

166 respectively. First, the cooling capacity of the ejector expander refrigeration system was determined under the external conditions. Then, the experiments of the vapor compression refrigeration cycle were performed under the same external conditions and cooling capacity as the EERC. The same cooling capacity in the conventional system was obtained by the inverter of the compressor. Following this method, the experiments of the ejector expander and conventional system were performed under the same cooling capacity and the same external conditions. temperature, which leads to the increase in the condenser saturation temperature and saturation pressure. For the given operating conditions, the COP values for the ejector expander cycle are % higher than in the conventional cycle. The uncertainties associated with the COP due to the accuracy of the instrument measurements for both the conventional and ejector systems were calculated to range between ±2.73% and ±2.8%. The input and output properties for each component of the EERC and VCRC are given on Table 2 for the design conditions (condenser temperature ~ 60 ºC, evaporator temperature ~ 10 ºC). From Table 1, it is seen that the condenser and evaporator exit pressures are close to each other when the two cycles are compared. Fig. 5 shows the variations of the cooling capacity of the EERC and VCRC with the condenser water inlet temperature. While the water inlet temperature increases, it is observed that the cooling capacity of both the systems also increase. As the condenser water inlet temperature increases, the pressure of the condenser exit and the motive nozzle exit increases. This causes the flow rate of the refrigerant sucked from the evaporator to increase. Thus, it is also predicted to increase the cooling capacity of the system. The cooling capacity of the conventional system was adjusted according to the ejector system. The inverter of the compressor in the VCRC was used to provide it. Fig. 5 also shows the variations of the coefficients of cooling performance (COP) of the EERC and VCRC with the condenser water inlet temperature. It was found that for the condenser water inlet temperature of 46.2 C, the COP for the conventional system was 1.858, whereas it was for the ejector system. Under these conditions, the use of an ejector as an expander provided an improvement of 14.2% in the COP. The reasons of the obtained higher coefficient of performance from the ejector system are the work recovery in this system and the fact that the refrigerant entering the evaporator is almost as saturated liquid. According to Fig. 5, the decrease in the COP for both systems as the condenser inlet water temperature increases is an expected result because of the increased water Fig. 5: Variation of COP values with the condenser inlet water temperature Fig. 6 shows the irreversibility of every component and total irreversibility for the ejector and conventional systems under the design condition for the same external operating conditions and the cooling capacity (4.47 kw). It can be seen that amount of the irreversibility of every component of the ejector system is lower than that of the conventional system. According to Fig. 6, the use of an ejector as an expander in the system reduces irreversibility by 12.66% compared with the total irreversibility of the VCRC. The use of the ejector instead of the expansion valve reduces the total amount of the irreversibility of the refrigeration cycle, as expressed theoretically in a study by Ersoy and Bilir (2010). Fig. 6: Comparing the irreversibility of every component and total irreversibility of the ejector and conventional 152

167 refrigeration cycles Tab. 1: Inlet and outlet variables of every component in both the ejector-expander and conventional refrigeration systems for the same cooling capacities under the same external conditions and (design conditions). EERC VCRC m (kg.s-1) P (kpa) T ( C) h (kj.kg-1) s (kj.kg-1.k-1) m (kg.s-1) P (kpa) T ( C) h (kj.kg-1) s (kj.kg-1.k- 1) Condenser inlet Condenser outlet Condenser cooling water inlet Condenser cooling water outlet Compressor inlet/separator vapor outlet Compressor outlet Ejector inlet (primary flow) Ejector inlet (secondary flow) Ejector outlet (exit of diffuser)/separator inlet Evaporator inlet Evaporator outlet Brine inlet Brine outlet Expansion valve inlet Expansion valve outlet/separator liquid outlet Fig. 7 shows the variation of the total destruction for the ejector and conventional systems with the condenser water inlet temperature. As the condenser water inlet temperature increases, the amount of the total destruction of both systems also increases. This can be explained as follows: as the condenser water inlet temperature increases, the saturated temperature of the condenser increases. Accordingly, the compressor pressure ratio also increases. Hence, the total irreversibility in the systems components and the total exergy destruction increase. In Fig. 7, it is seen that the total exergy destruction amount decreases by % when ejector, instead of the expansion valve, is used in the system. This result shows us, the purpose of using an ejector has been reached experimentally. It is shown in Fig. 7 that as the condenser inlet water temperature increases from 40.3 ºC to 46.2 ºC, the total exergy destruction of the ejector system increases by 15.6%, while that of the conventional system increases by 22.7%. Fig. 8 shows a comparison of the exergy efficiencies for both the conventional and ejector cycles. According to Fig. 8, as the temperature of cooling water at the entrance to the condenser increases, the exergy efficiency for both refrigeration cycles decreases. On the other hand, it is determined that as the condenser cooling water inlet temperature increases the total exergy destruction for both cycles increases (Fig. 7). For this reason, it is an expected result that the exergy efficiency decreases as the condenser water inlet temperature increases. The consumed power of the compressor increases as the condenser cooling water inlet temperature increases for the constant brine temperature and flow rate. With regard to the exergy efficiency equation, it is clear that the exergy efficiency will decrease with the increase in the power consumption of the compressor. Fig. 7: Variation of total exergy destructions of the ejector and conventional cycles with the condenser water inlet temperature. 153 Fig. 8: Variation of exergy efficiency with condenser water inlet temperature for the EERC and VCRC From Fig. 8, it is observed that the exergy efficiency of the EERC is higher than the conventional system.

168 This is because, according to the ejector efficiency equation, when the exergy output of both systems is almost equal, the exergy input (the compressor consumed power) of the ejector system is less than that of the conventional system. It is clear from Fig. 8 that the exergy efficiency increases by % when ejector instead of the expansion valve is used in the system under the same external conditions. A similar result was reported by Lawrence and Elbel, (2013) for the ejector cycle with dual evaporators without separator. The uncertainty associated with the exergy efficiency due to the accuracy of the instrument measurements for both the conventional and ejector systems was found ±2.1%. Fig. 9 shows a Grassmann (exergy flow and loss) diagram of the exergy balance of the ejector cycle operating under the design conditions. It can be seen that the compressor has the highest irreversibility, and the expansion valve operating in small pressure range has the lowest irreversibility. According to the Grassmann diagram, the total destruction of the system is 64.49%. The exergy loss that spreads to the environment from the system due to various reasons (discharging of the condenser cooling water to the environment and heat loss to the environment from the pipeline) is 21.26%. Consequently, the exergy efficiency of the ejector system is 14.25%. In addition, it is seen in Fig. 9 that the irreversibility rate in the evaporator is W, while that in the evaporator of the conventional system is W under the same operating conditions for both systems (Fig. 6). Accordingly, the use of the ejector in the system decreases the irreversibility in the evaporator by 59.6% and increases the exergy efficiency of the evaporator by 51.53%. This is because, as mentioned previously, the refrigerant enters as almost saturated fluid to the evaporator of the ejector system. This improvement in the evaporator and expansion work recovery provided by the ejector creates a combined effect on the system performance. Fig. 9: Grassmann diagram for the EERC under the design conditions. V. Conclusions In this study, we have experimentally investigated usability of the ejector as an expander to reduce the irreversibility of the R134a conventional refrigeration system. We have performed some experiments on an experimental setup that can operate as a conventional or an ejector refrigeration system of the same cooling capacity under the same external operating conditions. b) The total irreversibility was lower than in the conventional system by % while the exergy efficiencies were % higher than in the conventional system. c) The coefficients of performance were determined to be % higher than those of the conventional system. d) Fort future works, exergoeconomics and exergoenvironmental analyses and assessments are recommended. We may summarize some concluding remarks obtained from the results of the present study as follows: a) As the condenser water inlet temperature increased, the amount of the total destruction of both systems also increased. 154 Acknowledgements The authors would like to thank the Scientific and Technical Research Council of Turkey (TUBITAK) for their financial support of the project with the Grant number 110M044. In addition, the present paper

169 constitutes part of the PhD thesis of Nagihan Bilir Sag. Nomenclature E x : Exergy destruction rate (kw) h : Specific enthalpy (kj.kg-1 ) : Mass flow rate (kg.s-1 ) P : Pressure (kpa) Q : Cooling capacity rate (kw) s : Specific entropy (kj.kg-1.k-1 ) T : Temperature (ºC or K) Greek letters : Exergy efficiency ( ) Subscripts b : Primary and secondary flow entrance state to the suction chamber br : Brine c : Condenser ci : Condenser inlet cp : Compressor dest : Destruction e : Evaporator EERC : Ejextor Expander Refrigeration Cycle ei : Eveporator inlet ej : Ejector exp : Expansion valve i : Input m : Constant area mixing chamber o : Output sep : Separator sys : System tot : Total w : Water VCRC : Vapour Compression Refrigeration Cycle 0 : Reference environment m References Alsuhaibani Z., Ersoy H.K., Hepbaşlı A., Exergetic and sustainability performance assessment of geothermal (ground source) ejector heat pumps, International Journal of Exergy, 11 (3), (2012). Aroraa A., Kaushikb S.C., Theoretical analysis of a vapour compression refrigeration system with R502, R404A and R507A, International Journal of Refrigeration, 31 (6), (2008). Bilir N., Ersoy H.K., Performance improvement of the vapour compression refrigeration cycle by a two phase constant area ejector. Int. J. Energy Res. 33 (5), (2009). Bilir Sag N., Ersoy H.K., Hepbasli A., Halkaci H.S., Energetic and exergetic comparison of basic and ejector expander refrigeration systems operating under the same external conditions and cooling capacities, Energy Conversion and Management, 90, (2015). (2007). Chaiwongsa A P., Wongwises S., Experimental study on R-134a refrigeration system using a two-phase ejector as an expansion device. Appl. Therm. Eng. 28, (2008). Dincer I., and M.A., Rosen., Exergy: Energy, Environment and Sustainable Development. Oxford, UK: Elsevier (2007). Disawas S., Wongwises S., Experimental Investigation on The Performance of The Refrigeration Cycle Using a Two-Phase Ejector as an Expansion Device, International Journal of Refrigeration, 27 (6), (2004). Ersoy H.K., Bilir N., The influence of ejector component efficiencies on performance of ejector expander refrigeration cycle and exergy analysis. Int. J. Exergy 30, (2010). Ersoy H.K., Bilir N., Performance characteristics of ejector expander transcritical CO2 refrigeration cycle. Proc. IMche Part A J. Power Energ. 226, (2012). Ersoy H.K., Bilir N., Ejector Design and Experimental Investigation of its Effects on Performance of a Compressorbased Refrigerator in Which the Ejector is Used as an Expander. The Scientific and Technological Research Council of Turkey (TUBITAK), MAG Project 110M044 (2013). Ersoy H.K., Bilir Sag N., Preliminary experimental results on the R134a refrigeration system using a twophase ejector as an expander. Int. J. Refrigeration, 43, (2014). Kornhauser A.A., The use of an ejector as a refrigerant expander. Proceedings of the 1990 USNC/IIR-Purdue refrigeration conference, Purdue University (1990). Klein, S.A EES (Engineering Equation Solver), Academic Professional Version D, F-Chart Software Madison, WI, USA. Kotas T.J., The exergy method of thermal plant analysis. London: Butterworths (1985). Lawrence N., Elbel S., Theoretical and practical comparison of two-phase ejector refrigeration cycles including First and Second Law analysis, International Journal of Refrigeration, 36, (2013). Matawala V.K., Exergoeconomic optimization of an industrial aqua ammonia vapour absorption refrigeration unit, Ph.D. thesis, The Maharaja Sayajirao University of Baroda, Vadodara (2012). Chaiwongsa P., Wongwises S., Effect of throat diameters of the ejector on the performance of the refrigeration cycle using a two-phase ejector as an expansion device. Int. J. Refrigeration, 30, Menegay P., Kornhauser A.A., Improvements to The Ejector Expansion Refrigeration Cycle, Proceedings of the 31th Intersociety Energy Conversion Engineering Conference, Washington DC,

170 (1996). Pottker G. Potentials for COP increase in vapour compression systems. PhD thesis, University of Illinois at Urbana-Champaign (2012). Tas H., Bilgin N., Senturk B., Gungor A., Soğutma Çevrimlerinde Ejektör Kullanımının Araştırılması, Tesisat Mühendisliği Dergisi, 149, (2015). Wongwises S., Disawas S., Performance of the twophase ejector expansion refrigeration cycle, International Journal of Heat and Mass Transfer, 48, (2005). Yilmaz T., Unal S., Thermodynamic analysis of the two-phase ejector air-conditioning system for buses, Applied Thermal Engineering, 79, (2015). 156

171 Thermodynamic Assessment of Ozone Friendly Cascade Refrigeration System Using Natural Refrigerants H. Cenk Bayrakci 1, Onder Kizilkan 2 *, Ahmet Kabul 3, Selin Cekin 4 1,2,3,4 Süleyman Demirel University, Faculty of Technology, Department of Energy Systems Engineering, 32260, Isparta, Turkey * Abstract The purpose of this study is to make energy and exergy analysis of two-stage (cascade) refrigeration system which could be reached by spending less energy at lowers temperatures without causing extreme global warming and environment pollution by using natural refrigerants. In cascade system, R-744 and R-600a refrigerants were selected for low temperature circuit and for high temperature circuit respectively. Cooling load was determined as 10 kw at the system. The temperatures were taken 40 at condenser side and -30 at evaporator side. The effect of sub-cooling and superheating were investigated on system performance. Energy and exergy analysis were completed by using EES software. The results were presented by tables and graphically. Keywords: Natural refrigerant, Cascade system, Energy, Exergy I. Introduction Because of the global warming, usage of natural refrigerants is being common in large scale. CO2 refrigeration systems especially preferred at vehicle air conditioners and cascade refrigeration applications. Cascade systems are combined refrigerating systems used in industrial refrigerating plants at very lower temperature applications or at super refrigerating. If a system is to be operated at very lower temperatures, it means lower evaporation and lower condenser temperature, accordingly condensing temperature. In a system, if the heat is aimed quickly removed by condenser and if the refrigerant is aimed to condense totally at low temperatures, this could only be performed by the refrigerating of this system s condenser by other system. Cascade refrigerating systems are gained with the combined operation of two systems that is, with a system s condenser refrigerating other system s evaporator. In cascade refrigeration systems there are two different refrigeration systems as mentioned above. The biggest advantage of cascade system is to be able to provide different properties with different refrigerants. A different refrigerant is used at the system when obtaining low temperatures and as to the system s condenser refrigerating other system s evaporator, it is used a different refrigerant. The first system is called low temperature system and the other one is called high temperature system. In these systems refrigerating applications could be carried out between 70 C and 100 C (Dincer, 2003). In cascade systems there are many studies about the usage of different refrigerants in technical literature. 157 Kim and Kim (2002) have carried out an experimental study for cascade systems in which CO2 (R744), R744/134a and R744/290 have been used. Furthermore, they have compared these experimental data with a simulation suitable for system conditions. They have explained refrigerant flow rate, compressor power, cooling capacity; performance coefficient (COP) values changes with graphics (Kim and Kim, 2002). There are a number of studies about cascade refrigeration systems in the literature. Lee et al. (2006) were made thermodynamic analysis for the most suitable condensing temperature in the cascade condenser of a cascade cooling system which is operated with CO2/NH3 refrigerant couple in their study. In the study, they have taken the temperature difference in the system of evaporation temperature, condensing temperature and cascade condenser as design parameters. In the designed system they have written the exergy balances of each component separately and they have explained the temperature difference in evaporation temperature, condensing temperature and cascade condenser with graphics. With the results, they have presented equations concerning with the system s maximum performance coefficient depending on the temperature difference in evaporation temperature, condensing temperature and cascade condenser, by optimum condensing temperature of cascade condenser and again with the parameters. Gong et al. (2009) were measured cooling performance parameters by using dual (R170 + R23 and R170+R116) and triple (R170 + R23 + R116) azeotropic mixtures. Besides, R508B (R23+R116) refrigerant has also been used for comparison in similar conditions. R404A has been used in system s high temperature circuit. These four refrigerants COP, cooling capacity and condenser

172 temperature values have been determined in various condensing and evaporation temperatures. As a consequence of the indications, they have emphasized that R170+R116 dual mixture has % 10 higher performance coefficient value compared to R508B and mixtures have better potentials at low temperatures to -80 C. Bhattacharyya et al. (2005) were analyzed optimum cases of performance parameters (compressor outlet pressure, COPheating, COPcooling, COPsystem and second law efficiency) in the course of simultaneous heating and cooling. They were explained the results with graphics; they have emphasized that COP value has increased and also underlined that propane and carbon dioxide couple have been ideal in terms of thermal efficiency in these kind of systems. By using finite times method, Agnew and Ameli (2004) have showed the usage of this method at thermodynamics in a cascade circuit where R717 (ammoniac, known as eco-friendly) and R508b alternative refrigerant couple have been used. With this method, they have carried out optimization of a cascade circuit whose high temperature circuit is multi-stage and whose low temperature is singlestage. They have suggested several data to designers for these kind of systems operating with alternative refrigerants. Bhattacharyya et al. (2009) have analyzed simultaneous heating and cooling applications in a heating / cooling purpose cascade circuit where N2O and CO2 refrigerant couple have been used. In their study, they have presented COP (low temperature, high temperature and in general separately for the system) value and the change of Second Law efficiency according to the efficacy of the heat exchanger and inter stage temperature; and also they have given COP value and the change of Second Law efficiency as graphics according to the gas cooler outlet temperature. Therefore, they have introduced that the system s all performance is distinct from heat exchanger and they have stated that the gas cooler, evaporator and internal heat exchanger s design have affected all system s performance equally. Getu and Bansal (2008), were made thermodynamic analysis of an R744 R717 cascade refrigeration system. Their systems working range were between -50 and 40 C. They were showed COP variations with changing R- 717 evaporating temperature and condensing, evaporating and differential temperatures by graphics. They developed a mathematical model (multi linear regression analysis) for a guide of setting optimum thermodynamic design parameters. Duney et al. (2014), were used natural refrigerant propylene (R1270), which was been proposed for transcritical cascade refrigeration system and analyzed. In their study, Propylene was used in the low temperature (LT) cycle and carbon dioxide was used in the high temperature (HT) cycle of the cascade transcritical refrigeration system. They made also a thermodynamic analysis for a transcritical CO2/propylene (R744 R1270) cascade system for cooling and heating applications. They used EES software for making analysis. In the study variation of three important design parameters i.e. gas cooler outlet temperature TC, evaporating temperature TE 158 and overlap temperature in cascade heat exchanger is considered in order to determine system COP, optimum temperature in cascade heat exchanger and optimum mass flow ratio of LT and HT cycles. They developed regression equations for Topt, COPmax and optimum mass flow ratio to help thermal engineers to design an optimized transcritical cascade system. Yan et al (2015), were made thermodynamic analysis of an internal auto-cascade refrigeration cycle (IARC) with mixture refrigerant R290/R600a. R290/R600a mixture was a zeotropic mixture and used in domestic refrigerator-freezers. According to their study, performances of the IARC are evaluated by using a developed mathematical model, and then compared with that of the conventional refrigeration cycle (CRC). According to the simulation results, the IARC with R290/R600a has % improvement in coefficient of performance (COP), % improvement in volumetric refrigeration capacity and % reduction in pressure ratio of compressor compared with those of the CRC under the same given operating conditions. Therefore, many investigations of the cascade refrigeration system are attracting attention The purpose of this study is to make energy and exergy analysis of two-stage (cascade) refrigeration system which could be reached by spending less energy at lowers temperatures without causing extreme global warming and environment pollution by using natural refrigerants. II. System Description Natural refrigerants are naturally occurring substances, such as hydrocarbons (propane, isobutane), CO2, ammonia, water and air. These substances can be used as cooling agents (heat transfer medium) in refrigerators and air conditioners, don't harm the ozone layer and have no or negligible climate impact (Refrigerants Naturally, 2015). This kinds of fluids have been used as refrigerants for many years, however, they are now finding their way into applications where previously fluorocarbons were the preferred option (AGDE, 2015).They are now being used more extensively due to their low impact on the environment (Linde, 2015). The advantages of natural refrigerants are they do not damage the ozone layer and have a negligible impact on the greenhouse effect. From an economic perspective, these refrigerants are inexpensive, in some cases even cheaper than HFCs. Also, natural refrigerants are extremely energy-efficient, sometimes up to 40% more than HFCs (Shecco, 2015; Kizilkan, 2015). In this study, some different natural refrigerants and some chlorine based refrigerants were used in theoretical analysis. Analyses were made by using EES software (Klein, 2015). Schematic diagram of the system was shown in Figure 1. In Table 1, it could be seen that the Physical, safety and environmental properties of investigated the refrigerants.

173 III. Thermodynamic Modelling The performance characteristics of the vapor compression refrigeration cycle for the cold storage facility are assessed by applying first and second law analysis of thermodynamics. The balance equations are used to determine the work and heat interactions, energy and exergy efficiencies and exergy destruction rates for each system component. The general mass balance equation for a steady-state and steady-flow processes can be written as (Cengel and Boles, 2006) m in = m out (1) The energy balance equation is given below: E in = E out (2) Equation (2) can be written as: Q + m inh in = W + m outh out (3) Fig. 1: Schematic diagram of the cascade system where, m is the mass flow rate, E is the rate of net energy, Q is the rate of net heat, W is the rate of net work, and h is the specific the subscripts in and out stand for inlet and outlet respectively. ASHRAE number Tab. 1: Physical, safety and environmental properties of investigated refrigerants Critical Critical ODP GWP Molecular Safety Pressure Temperature (relative (relative formula group (kpa) ( C) to R11) to CO2) 159 Atmospheric life time (year) R1270 CH3CH=CH A R290 CH3CH2CH A R600 CH3CH2CH2CH A R600a CH(CH3) A R717 NH B R12 CCl2F A R22 CHClF A R134a CH2FCF A The second law of thermodynamics overcomes with concepts of entropy and exergy. Exergy analysis of systems allows determining irreversibility and available energy (exergy) in the system. These analyses reveal the efficiency of the systems in terms of first and second law of thermodynamics for a steady state operation, the general exergy balance equation can be defined as (Dincer and Rosen, 2007). Eẋ in = Eẋ out + Eẋ dest (4) In equation 4, the exergy balance equation can also be written as: Eẋ Q Eẋ W = m ine in m oute out + T 0 S gen (5) where, Eẋ Q and Eẋ W are the exergies of heat and work, respectively, e is the specific exergy, T0 is the dead state temperature and S gen is the entropy generation rate. In equation 5, the exergy of heat, the exergy work and entropy generation are given below (Kotas, 1985) Eẋ Q = Q ( T T 0 T ) (7) Eẋ W = W (8) Eẋ dest = T 0 S gen (6) The specific exergy is given below with relative to the environment conditions: e = (h h 0 ) T 0 (s s 0 ) (9) where s is entropy, P is the pressure and the subscript 0 indicates properties at the reference state. The performance of the cascade refrigeration system can be calculated using energy and exergy efficiency definitions: COP = Q E W C,low+W C,high (10)

174 η ex = Eẋ Q E Eẋ W C,low +Eẋ W C,high (11) where Q E represents evaporator refrigeration capacity and W C,low and W C,high represents compressor capacity of lower and higher cycles, respectively. The governing balance equations are given for all system components in Table 2 according to the reference points shown in Figure 1. Tab. 2: Energy and exergy balance equations for system components. Component Mass balance Energy balance Exergy balance Higher cycle compressor m 5 = m 6 = m high W C,high = m high (h 6 h 5 ) Ex 5 + W C,high = Ex 6 + Ex dest,wc,high Condenser m 6 = m 7 = m high Q con = m high (h 6 h 7 ) Higher cycle throttling valve Cascade heat exchanger Lower cycle compressor Lower cycle throttling valve m 7 = m 8 = m high h 7 = h 8 m 8 = m 5 = m high m 2 = m 3 = m low Q HEX = m low (h 2 h 3 ) Q HEX = m high (h 5 h 8 ) m 1 = m 2 = m low W C,low = m low (h 2 h 1 ) Ex Qcon = Q con [1 T 0 ] T con Ex 6 = Ex 7 + Ex Qcon + Ex dest,con Ex 7 = Ex 8 = Ex dest,hctv Ex 2 + Ex 8 = Ex 5 + Ex 3 + Ex dest,hex Ex 4 +Ex Qevap = Ex 1 + Ex dest,wc,low m 3 = m 4 = m low h 3 = h 4 Ex 2 + Ex 8 = Ex 3 + Ex 6 + Ex dest,lctw Evaporator m 4 = m 1 = m low Q evap = m low (h 1 h 4 ) Ex Qevap = [( T 0 T evap ) 1] Ex 4 + Ex Qevap = Ex 1 + +Ex dest,evap IV. Results and Discussion In order to simulate the cascade refrigeration cycle for natural refrigerants, the following assumptions were made: All operations are steady state and steady flow. Pressure losses through pipelines are neglected. Heat losses and heat gains from or to the system are neglected. The changes in potential and kinetic energies are neglected. The pump operations are adiabatic and isentropic. The directions of heat transfer to the system and work transfer from the system are taken positive. Using the balance equations, and under the assumptions given above, the analyses are performed for different natural refrigerants using EES software (Klein, 2015). The results of thermodynamic analyses of the cascade refrigeration system for the given cooling load are given in Table 3 for all refrigerants. The table is divided into two parts, natural refrigerants and chlorine based refrigerants. It can be seen from the table that the best COP value is obtained using R717 followed by, R600, R600a and R290 in natural refrigerants. The performances of these refrigerants are very similar to that of R12, R22 and R134a. Also the trend of exergy efficiency is the same as COP. For the exergy destruction rates, the highest destruction is occurred using R717 and R600 for the given refrigeration duty. Furthermore, the electricity consumption and pressure ratio of the compressor for all refrigerants are given in Table 3. Tab. 3: The results of thermodynamic analyses of the refrigeration system Eẋ Refrigerant COP η Dest, ex kw R R Natural R600a R R Chlorine R based R134a Also, some parametric studies were carried out to see the variation of performance indicators with different parameters. In Fig. 2, variation of condenser temperature with COP can be seen. If the COP values examined, the best COP value was obtained for R717 refrigerant. With the increasing condenser temperatures, COP values decreasing. R290 has the worst value for COP. In Fig. 3, variation of evaporator temperature with COP in can be seen. If the COP values examined, the best COP value was obtained for R717 refrigerant again like before. With the increasing evaporator temperatures, COP values also increasing. R290 has the worst value for COP again. 160

175 Fig. 2: Variation of condenser temperature with COP Fig. 5: Variation of evaporator temperature with exergy efficiency In Figure 6 and Figure 7, variation of condenser and evaporator temperatures with exergy destruction in the high pressure side and low pressure side could be seen respectively. With the increasing condenser temperature and evaporator temperature total exergy destruction values are decreasing. Fig. 3: Variation of evaporator temperature with COP In Fig. 4, variation of exergy efficiency with condenser temperature in the upper system could be seen. If the exergy efficiency values examined, the best value was obtained for R717 refrigerant again like before. With the increasing condenser temperature, exergy efficiency values are decreasing. R290 has the worst value for exergy efficiencies. Fig. 6: Variation of condenser temperature with exergy destruction Fig. 4: Variation of condenser temperature with exergy efficiency In Fig. 5, variation of evaporator temperature with exergy efficiency in the subcooling system could be seen. Same situation occurs here like Figure 3. R717 has best values for exergy efficiency. R290 has the worst values. 161 Fig. 7: Variation of evaporator temperature with exergy destruction V. Conclusion Thermodynamic assessment of cascade refrigaretion cycle was carried out for natural refrigerants while CO2 refrigerant was used for the lower cycle.

176 According to the graphics and results it could be seen clearly that the best values given by natural refrigerants (especially R717 and R600) in the cascade refrigeration systems. Because of low GWP, low exergy destruction, good COP and exergy efficiency values and ozone friendly properties, natural refrigerants could prefer by the refrigeration systems manufacturers. Because of the other refrigerants have higher GWP values and lower COP and exergy efficiency values, they are non-preferable for these systems. However, in the point of view of energy and exergy efficiency, the best alternative refrigerants for higher cycle are found to be R717 and R600. Such a comparison of energetic and exergetic performance of these refrigerants gives valuable and practical knowledge for refrigeration sector designers. References Agnew B., Ameli S.M., A finite time analysis of a cascade refrigeration system using alternative refrigerants. Applied Thermal Engineering, 24, , AGDE, (2015). Australian Government Department of the Environment. lications/natural-refrigerants-case-studies, accessed Bhattacharyya S., Mukhopadhyaya S., Kumar A., Khurana R.K., Sarkar J., Optimization of a CO2 C3H8 cascade system for refrigeration and heating, International Journal of Refrigeration, 28, , Bhattacharyya S., Garaia A., Sarkarb J., Thermodynamic analysis and optimization of a novel N2O CO2 cascade system for refrigeration and heating, International Journal of Refrigeration, 32, , Cengel Y.A., Boles M.A., Thermodynamics: an engineering approach, 5 th ed., McGraw-Hill, New York, USA, Dincer I., Refrigeration Systems and Applications, Wiley, England, Gong M., Sun Z., Wu J., Zhang Y., Meng C., Zhou Y., Performance of R170 mixtures as refrigerants for refrigeration at -80 0C temperature range, International Journal of Refrigeration, 32, , Kim S.G., Kim M.S., Experiment and simulation on the performance of an auto cascade refrigeration system using carbon dioxide as a refrigerant, International Journal of Refrigeration, 25, , Kizilkan O., A Comparative Investigation of Natural Refrigerants: A Case Study for Cold Storage Application, SDU International Journal of Technological Sciences, 7(3), 1-15, Klein, S.A., Engineering Equation Solver (EES), Version D, F-Chart Software, Kotas T.J., The exergy method of thermal plant analysis, Butter-Worths, London, UK, Lee, T.S., Liu, C., Chen, T., Thermodynamic analysis of optimal condensing temperature of cascadecondenser in CO2/NH3 cascade refrigeration systems, International Journal of Refrigeration, 29, , Linde, (2015). The Linde Group, Industrial gases. l_refrigerants/index.html, accessed Refrigerants Naturally, (2015). c/o HEAT International, accessed Shecco, (2015). Beyond HFCs accessed Yan G., Hu H., Yu J., Performance evaluation on an internal auto-cascade refrigeration cycle with mixture refrigerant R290/R600a, Applied Thermal Engineering, 75, , Dincer I., Rosen, M.A., Exergy: Energy, Environment and Sustainable Development, 1 st ed., Elsevier Science, Oxford, UK, Duney, A.M., Kumar, S., Agrawal, G.D., Thermodynamic analysis of a transcritical CO2/propylene (R744 R1270) cascade system for cooling and heating applications, Energy Conversion and Management 86, , Getu, H.M., Bansal, P.K., Thermodynamic analysis of an R744 R717 cascade refrigeration system, International Journal of Refrigeration, 31,45 54,

177 Thermodynamic Analysis of an Integrated System with A Concentrating Collector for Multi-Generation Purposes Yunus Emre Yuksel 1*, Murat Ozturk 2 1 Afyon Kocatepe University, Education Faculty, Department of Elementary Science Education, ANS Campus, Afyon, 03200, Turkey 2 Suleyman Demirel University, Faculty of Technology, Department of Mechatronics Engineering, Cunur, West Campus, Isparta, Turkey * Abstract In this paper, thermodynamic analysis of an integrated system with concentrating collector is investigated for power, heating, cooling and domestic hot water production. The renewable energy based integrated system consists of four sub-systems; i-) a concentrating collector cycle, ii-) an energy storage process, iii-) a Rankine cycle, and iv-) a double effect absorption cooling system. The integrated system for multi-generation purposes is examined in two operating modes, i-) solar mode and ii-) storage system mode. Thermodynamic analysis based on the energy and exergy efficiency, and also exergy destruction rate for whole system and its components are presented for two operating modes. The overall energy and exergy efficiencies of the integrated system are calculated as 51.32% and 46.75%, respectively for the solar mode, whereas these efficiencies are found to be 47.44% and 45.43%, respectively for the storage system mode. In addition, the parametric studies including the variation of ambient temperature from 0 0 C to 30 0 C, and solar radiation flux from 500 W/m 2 to 1000 W/m 2 are presented for the integrated system and its components to investigate and compare the system efficiency. Keywords: Solar energy, concentrating collector, energy, exergy, multi-generation, efficiency. I. Introduction Nowadays, world energy production is based approximately 80% on fossil fuels, such as coal, oil and natural gas (Carvalho, et al., 2011). The problem about fossil fuels is not only shortening of them but damaging the environment as well. Gases such as CO2, SOx, and NOx emitting from burning of fossil fuels constitute greenhouse gases. Therefore, the effects of global warming and climate changes increase day by day, and also these effects can be easily seen by human being. It is time to change this current energy infrastructure with alternative energy sources which are abundant on earth and harmless to the environment. Solar energy is a reliable energy resource. Also, this renewable energy source is a well-known proven renewable energy system, because of its availability. According to the environmental view-point, solar energy systems do not have negative effects and harmful emission to the environment compared to the fossil energy sources, which continuously increase the earth s average ambient temperature and pollution (Al-Sulaiman, et al., 2011). There are a few solar thermal systems that can be used to produce electricity via thermal power plants, such as a solar tower system (STS), a parabolic dish collector (PDC) and a parabolic trough solar collector (PTSC). The PDC system is used to focus the concentrated solar energy on a working fluid to generate heat energy, and then change it to electricity in a conventional generator. The system uses one or more parabolic dishes called reflector to reflect the direct solar energy onto the receiver sub-system located the focus point of the concentrating collector. The thermal energy is collected in a heated working fluid. The high temperature working fluid is transferred to the steam generator through pipes to produce high pressure and super-heated steam. This super-heated steam is sent to a conventional high-efficiency steam turbine in order to generate electricity in this case. Also, the PDC system is the most advanced solar energy technology and has been utilized in large solar power plants for three decades. As a result, they are rather efficient at thermal energy absorption and power conversion processes (Al-Sulaiman, et al., 2011). Thus, in this paper, the PDC system is considered for increasing the temperature of the working fluid. In order to produce cooling, an absorption chiller system should be used integrated with the renewable energy resources, such as solar, geothermal and biomass. The other important advantages of the absorption cooling system should be given as i-) this system does not cause ozone layer depletion, ii-) this system use natural refrigerants possibly having less CO2 emissions, and iii-) this system is independent of the electric grid. The most common commercially suitable absorption refrigeration systems are single and double effect systems. In double effect absorption refrigeration systems, a secondary fluid (absorbent) is used to circulate and to absorb the primary fluid (refrigerant). The success of the absorption relies on the selection of an appropriate combination of absorbent and 163

178 refrigerant (Minciuc, et al., 2003). The most widespread absorbent and refrigerant combinations in absorption refrigeration systems are LiBr-H2O and ammonia-water. The LiBr-H2O pair is the most suitable one for air-conditioning and chilling applications. Gomri (2010) have also performed second law comparison of single effect and double effect vapor absorption refrigeration systems, and concluded that the exergy efficiency of double effect absorption system is higher than the single effect system. Also, in this paper, lithium bromide and water are chosen as working fluid mixture. Zhao et al. (2003) have investigated a new type double effect absorption cooling system based on the energy and exergy analysis. Also, balance equations of the mass, energy and exergy are given for the whole system and its components. Abo Elazm et al. (2011) have presented a study about comparison of single and double effect absorption coolers showing that coefficient of the performance of double effect absorption cooler is higher than single effect absorption system. In this paper, double effect absorption system is considered for cooling application. Increasing population and high energy demand need new alternatives for current energy infrastructure. In order to meet rising energy demand, energy sources should be used more efficiently. Integrated systems offer higher advantages than single output systems in terms of efficiency (Ahmadi, et al., 2013). Several studies have been conducted on multi-generation energy production systems. Buck and Fredmann (2007) have analyzed the efficiency of a tri-generation system based on a micro turbine assisted by a solar power tower. They have conducted an economic analysis on the use of the single and double effect absorption process. The authors have recommended that using the double effect process because it has showed better thermal performance and lower operating cost compared to the single effect absorption process. Khaliq et al. (2009) have presented a trigeneration system using waste heat. They have studied the energy, exergy efficiencies and electrical to thermal energy ratio with respect to both waste heat temperature and pressure of process heat. Kavvadias and Maroulis (2010) have analyzed the multi-objective optimization of a new tri-generation system for power, heating and cooling production. This optimization study has been carried out on technic, economic, energetic and environmental performance indicators in a multi-objective optimization framework. The results have indicated that tri-generation system should be more economically attractive, energy and exergy efficiently and environmental friendly than conventional system. Dincer and Zamfirescu (2012) have carried out energy and exergy based analyses for renewable energy based multigeneration, considering different options for generating such outputs as power, heat, hot water, heating and 164 cooling, hydrogen and fresh water. They have compared single and cogeneration systems in terms of payback time, it is found that cogeneration systems have 2.8 less payback time than single generation system. Ozturk and Dincer (2013) have presented a solar based multi-generation system with hydrogen, electricity, cooling and heating outputs. They have presented exergy destruction rate, energy and exergy efficiencies of each component and the whole system. Also, overall system performance depending on reference temperature has been analyzed. Several researchers have studied the utilizing of integrated systems in energy generation to improve the thermodynamic and environmental performance. Ozturk and Dincer (2013) have researched the integrated systems having rising interest in the last few decades so as to reduce energy consumption and accomplish more sustainable energy production. Al-Sulaiman (2013) has carried out an energy-based analysis of a concentrating solar collector integrated with steam and binary vapor cycles as a prime power for electricity generation. The author has applied an energy efficiency and power production analysis to find the best design parameters of the integrated system. Caliskan et al. (2013) have conceptually modelled hybrid renewable energy based hydrogen and electricity production and storage systems and analyzed them in detail with energy, exergy and sustainable approaches. As a case study they have designed a hybrid wind-solar renewable system. Results show that maximum energy efficiencies of wind turbine, solar PV panel, electrolyzer, and PEMFC are 26.15%, 9.06%, 53.55%, and 33.06% respectively. Maximum exergy efficiencies of the same subsystems are 71.70%, 9.74%, 53.60%, and 33.02% respectively. Ghosh and Dincer (2014) have proposed a novel multi-generation system, which combines three different renewable energy resources, such as solar, wind and geothermal energy, for production multi-outputs as power, heating and cooling, drying and fresh water. The mathematical expressions of the mass, energy, entropy and exergy balance are also given. Also, the meteorological parameters that affect the renewable energy systems for multi-generation are considered. Padilla et al. (2014) have conducted an exergy analysis to parabolic trough collectors (PTC) in order to investigate the effects of operational and environmental parameters on performance of PTC. The main parameters considered for the analysis are: inlet temperature and mass flow rate of heat transfer fluid, wind speed, pressure or vacuum in annulus and solar radiance. According to results, inlet temperature of heat transfer fluid, solar irradiance and vacuum in annulus have significant effect on the thermal and exergetic performance, however the effect of wind speed and mass flow rate of heat transfer fluid is negligible. Al-Ali and Dincer (2014) have presented a new multigenerational integrated geothermal-solar system to produce electrical power, cooling, space

179 heating, hot water and heat for industrial use. They have applied energy and exergy analyses and compared the results of single generation, cogeneration, trigeneration and multigeneration systems. To investigate the effects of operating conditions and environmental parameters, a parametric study is exercised. As a comparison, energy efficiencies of single generation and multigeneration are found as 16.4% and 78% respectively, 26.2% and 36.6% in exergy efficiency respectively. Mamaghani et al. (2015) have modelled a molten carbonate fuel cell-gas turbine (MCFC-GT) hybrid plant in view of energetic, exergetic, economic and environmental analyses. They have optimized the system by using a multi-objective optimization. Target of multi-objective optimization has been 51.7% exergy efficiency and the total cost of million USD per year. They have concluded that operating pressure has the most significant effect on the exergetic efficiency of the plant according to sensivity analysis on variations of system parameters. Chitsaz et al. (2015) have analysed a novel trigeneration system driven by a solid oxide fuel cell (SOFC) in terms of exergy efficiency, exergy destruction rate and greenhouse gas emissions. In the study, four operation cases have been investigated: electrical power generation, electrical power and cooling cogeneration, electrical power and heating cogeneration, and trigeneration. A maximum improvement in the exergy efficiency has been found as 46% in trigeneration system equipped with solid oxide fuel cell as a prime mover compared to the case when the SOFC is used as a standalone unit. Khalid et al. (2015) have analyzed three new developed HVAC systems for heating and cooling applications. To compare these three systems, energy and exergy analyses have been applied for each case and the effects of parameters on energy and exergy efficiencies have been evaluated. The maximum overall efficiency has been found in natural gas operated system with vapour absorption system chiller at 27.5% while minimum energy efficiency has occurred in photovoltaic and solar thermal operated system with vapour compression chiller at 19.9%. A multigeneration energy system based on sawdust biomass fuel with five useful outputs has been analysed in terms of energy and exergy analyses by Soltani et al. (2015). Energy and exergy efficiencies of the multigeneration system are found as 60% and 25%, respectively, while corresponding energy and exergy efficiencies of a biomass system with only electricity generation are 11% and 13%, respectively. Of the several heat recovery options from exhaust gases, electricity generation and wood drying result in the highest exergy efficiency while district heating and drying lead to the highest energy efficiency. Hassoun and Dincer (2015) have developed a new organic Rankine cycle based multigenerational system to meet the demands of a net zero energy building and assessed such a system for an application to a net zero energy house in Lebanon. Energy and exergy analyses have been applied and a parametric study 165 has been conducted. In addition, exergoeconomic analysis and an optimization study for optimizing the total system cost to the overall system efficiency have been carried out. Bade and Bandyopadhyay (2015) have used a pinch analysis based methodology to integrate gas turbine and regenerator with a process plant to minimize fuel consumption. Thermodynamic analysis of this combined heat and power plant has been applied on gas turbine pressure ratio versus power to heat ratio diagram. By using this novel diagram, it is expected to optimize the integration of gas turbine with a process plant. The specific objectives of this study are to investigate a thermodynamic analysis of the multi-generation system supported by the PDC system with an energy storage sub-system, a Rankine cycle and a double effect absorption cooling sub-system, and to reduce the environmental impacts and system cost. The other main sub-objectives of this paper should be detailed as listed below; To develop an advanced Engineering Equation Solver (EES) software code for analyzing a novel integrated system using the PDC as a prime mover. To determine the exergy contents for each stream of the integrated system. To calculate the exergy efficiencies and destructions of the system components and whole system for two operating modes. To conduct a complete parametric study to analyze the impacts of the varying some significant parameters on the integrated system performance. II. System design Fig. 1 shows the schematic representation of the integrated system with the PDC sub-system, the Rankine cycle, the energy storage sub-system and the double effect absorption cooling system. The PDC system collects the solar radiation and then concentrates in order to boil the working fluid for obtaining thermal energy. The working fluid leaving a collector pump at point 6 enters the PDC system to be heated up to 600 C. This outlet temperature is assumed the maximum suitable temperature for the selected working fluid in the PDC system. The mass flow rate of the PDC system without looping is kg/s. At point 1, heated working fluid leaves the concentrating collector and goes through the heat exchanger-i (HEX-I) and HEX-V for solar energy storage and electricity generation, respectively. At point 15, water enters a hot storage tank and then water is pumped into HEX-III at night time. Heat transfer coefficient of the heat water storage tank integrated with multi-generation system is taken as W/m 2 K. The working fluid heated by the PDC

180 system is used to heat the working fluid in the Rankine cycle sub-system for power generation. The Rankine cycle considered in this system has one turbine and feed water unit with a condenser, a pump and a heat exchanger. At point 16, water vapor leaving heat exchanger with approximately 420 C goes through turbine in order to be expanded and produce electricity. After this process, temperature and pressure of water vapor decrease at point 17 where steam enters the condenser-i. Steam at this point is generally high quality saturated liquid-vapor mixture. Steam is condensed at constant pressure in the condenser-i, and leaves the condenser-i as saturated liquid. Water entering pump-iii at point 18 is compressed isentropically to the operating pressure of the HEX-III and V for the night time and solar time, respectively. Figure 1. Schematic diagram of the integrated system with concentrating collector The Rankine cycle of the integrated system produces electricity, and a part of this producing electricity is used for operations of the system devices. The waste heat from the HEX-II is used to produce cooling using by the double effect absorption cooling sub-system. In order to use waste heat of system, the double effect lithium-bromide-water absorption system is chosen instead of a conventional refrigeration system. In this paper, required energy for the double effect absorption system is supplied from the Rankine cycle and the energy storage sub-system. As it can be seen from Fig. 1, the most important components of the double effect absorption system are the high and low pressure generator, the high and low temperature heat exchanger, the solution and refrigerant pump, an absorber, a condenser and an evaporator. It should be noted that, integrated system is modeled according to the optimum operating parameters for the double effect 166 absorption sub-system. Since the solar energy inputs to the system changes with time, the solar-based integrated system has a dynamic process characteristic. Solar radiation increases from zero at the sunrise to its maximum at solar noon time, after that decreases until it equals to zero at the sunset. To reach a continuously processing solar based system, an additional prime mover or an energy storage sub-system should be integrated with the solar system. In this paper, a thermal energy storage system is combined with the solar-based integrated system. This energy storage sub-system stores the excess parts of the solar thermal energy during the solar time, and provides operating the integrated system at night time. Hence, design parameters of the solar thermal energy storage sub-system are very important for a continuously running solar thermal system. In this paper, it is assumed that, 65% of the solar thermal energy during the solar-time is stored in the energy

181 storage sub-system with the purpose of meeting the heat loss from the heat storage tank and to provide heat in the heat exchanger for continuously energy production. Also, a thermocline or single tank system is selected as the hot-storage tank. III. Assumptions In order to make thermodynamic modeling of the efficiency and structure of the integrated system, some assumptions should be accepted, and numerical analysis also should be made for the enthalpy, entropy, temperature, pressure, mass flow rate and exergy of the inlet and outlet flows. The following assumptions are used for this paper; All the system components operate in the steady state conditions. References temperature and pressure are assumed as 25 C and 1 bar, respectively. All flow steams are ideal gases. The changes in the kinetic and potential terms in the energy and exergy balance equations are negligible. No pressure and heat losses are considered in the flow channels. The turbine and pumps in the integrated system are adiabatic. There is no chemical reaction in the system components. IV. Thermodynamic Analysis Thermodynamic analysis is used to evaluate the performance of the system and its components in terms of the first and second laws of thermodynamic. In this section, thermodynamic analysis consisting of the mass, energy, and exergy analysis is employed in order to evaluate performance and improvement potential of the integrated system with concentrating collector system. In the most general principle, a balance equation for a quantity in the given process should be written as follows; Input + Generation Output Consumption = Accumulation (1) This equation gives the quantity balance for a process. The difference between input to the system with generated quantity in the system boundary and output quantity from the system with consumed quantity in the system is equal to the accumulated quantity. In the steady state condition, the accumulation terms in the Eq. (1) are equal to zero, because all properties in the process are unchanging with time (Dincer & Rosen, 2013). IV.1. Mass balance analysis Mass balance is the first analysis in evaluating any system thermodynamically. In the steady state condition, mass balance equation for any system 167 can be written as follows; m i = m e (2) where m is the mass flow rate and subscripts i and e indicates the inlet and exit flow of the matter, respectively. IV.2. Energy balance analysis Energy balance is a key analysis for any system that is investigated. According to the first law of thermodynamic, energy is neither created nor destroyed in a system. So the sum of energy content in a process is always constant, therefore energy is conserved. By neglecting the kinetic and potential energy changes, energy balance analysis should be written as follows showing that the sum of input energy is equal to the sum of output energy; i E i + Q i = e E e + W (3) where Q and W are the heat and work transfer rate, respectively. Neglecting potential and kinetic energy, the above equation should be written as follows; Q + m i h i = W + m eh e (4) where h is the specific enthalpy. The energy analysis and balance equation is performed for each main components of the integrated system as given below. IV.2.1. Concentrating collector The PDC system has a parabolic shape with covered reflecting materials, and a receiver is placed to the focal point of the concentrating collector. Reflected solar energy from the reflecting surface to the collector receiver should be given as follows (Kalogirou, 2009) (Duffie & Beckman, 2006); Q R = ρ r,c α r,c I ds A C (5) where ρ r,c and α r,c are the reflectivity and absorptivity of the concentrating collector reflecting surface material, I ds is the direct solar radiation and A C is the area of the reflector. The produced useful power from the PDC system should be calculated as follows; Q u = F R [C o (ρ r,r α r,r )Q R U L (T c T o )A R εσ(t R 4 T o 4 )A R ] (6) where F R is the heat removal factor, C o is the collector concentrating ratio, ρ r,r and α r,r are the reflectivity and absorptivity of the receiver, U L is the overall heat loss coefficient of the concentrating collector, T c and T R are the collector and receiver temperature, respectively, ε is the receiver emissivity and σ is the Stefan-Boltzmann constant.

182 To analyze the concentrating collector performance on the basis of thermodynamic assessment, design parameters of the collector are given in Table 1. Tab. 1: Concentrating collector parameters Parameter Values ρ r,c 0.9 ρ r,r 0.85 α r,c 0.9 α r,r 0.85 A C 100 m 2 A R 0.4 m 2 F R 0.9 C o 250 U L 25 W/m 2 K ε 0.2 IV.2.2 Rankine cycle As seen from Fig. 1, the high temperature working fluid goes to the Rankine cycle turbine at point 16 and after expansion leaves from here at point 17. To analyze the inlet and outlet enthalpies and turbine power output, the energy balance equation of the turbine should be written as follows; m 16h 16 = m 17h 17 + W T (7) Energy balance equation for the condenser-i in the Rankine cycle is given by; m 17h 17 = m 18h 18 + Q cond I (8) The Rankine cycle pump work should be written using by the energy balance equation; W pump IIII = m 16(h 18 h 19 ) (9) In order to exchange heat energy more from the PDC to the Rankine cycle, HEX-V is used. Inlet and outlet enthalpies of the HEX-V should be calculated by simply applied to the energy balance equation; m 7(h 7 h 8 ) = m 16(h 16 h 20 ) + Q loss,hex V (10) IV.2.3 Double-effect absorption cooling sub-system To analyze the inlet and outlet conditions of the generator-i (or high temperature generator), energy balance equation should be written as follows; m 22h 22 + m 30h 30 + Q gen I = m 23h 23 + m 31h 31 + m 34h 34 (11) An energy balance equation of the generator-ii (or low temperature generator) is given as follows; m 33h 33 + m 34h 34 + Q gen II = m 35h 35 + m 37h 37 + m 38h 38 (12) An energy balance equation of the condenser-ii in the cooling sub-system should be expressed as follows; m 36h 36 + m 37h 37 = m 24h 24 + Q con II (13) The following energy balance equation should be used to calculate the heat absorbed from the evaporator; m 25h 25 + Q eva = m 26h 26 (14) To obtain the heat rejected from the absorber, the following energy balance equation should be used; m 26h 26 + m 40h 40 = m 27h 27 + Q abs (15) IV.3. Exergy balance analysis Exergy is defined as the maximum useful work that could be obtained from the system at a given state in a specified environment. Exergy analysis provides more meaningful information about a system or a process than energy analysis. Because energy analysis deals with conserved quantities, it provides one sided view of a process or a system. Unlike energy analysis, exergy analysis shows inefficiencies and wastes occurring in the process. According to the second law of thermodynamic, different kinds of energies exist, all energies are not equal, and quality decreases in any system. Exergy analysis based on the second law of thermodynamic can be used to investigate a system for more economical and effective use of energy sources (Yuksel & Ozturk, 2014). Thus, integrated systems offering more effective use of energy sources should be evaluated with both energy and exergy analyses. By applying exergy analysis, where and how inefficiencies occur can be found, and so efficiency of the system can be increased. Based on the first and second law of thermodynamics, an exergy balance equation should be written as follows (Dincer & Rosen, 2013); i m inex in + E x Q = e m outex out + E x W + E x D (16) where E x Q and E x W are the heat and work exergy flow rates through the boundary at temperature Tj at location j, respectively. E x D is the exergy destruction rate. Heat exergy flow rate is given as; E x Q = (1 T o T i ) Q i (17) where T i is the temperature in the i th given state. E x W = W (18) Exergy is generally comprised of four parts which are physical exergy (exph), chemical exergy (exch), kinetic exergy (exke), and potential exergy (expe). The 168

183 specific exergy should be given as follows; ex = ex ke + ex pe + ex ph + ex ch (19) In this study, we have neglected the kinetic, potential and chemical exergy, as elevation difference is low and speeds in the process are negligible, and there is no chemical reaction. Physical exergy can be defined as maximum effective work as a process interacts with the environment (Bejan, et al., 1996). The physical exergy rate of the i th flow is written as follows; ex ph = (h i h o ) T o (s i s o ) (20) The exergy rate of the material flow should be calculated as follows; E x i = m ex i (21) The exergy destruction rate for each components of the solar-based integrated system are written based on the given above procedure, and shown in Table 2. Tab. 2: Exergy destruction rate equations System Exergy destruction rate equations components Parabolic Q E x D,PDC = E x 6 E x 1 + E x Solar collector HEX- I E x D,HEX I = E x 2 + E x 14 E x 3 Ex 15 Collector E x D,CP = E x 5 E x 6 + W CP pump Pump-I E x D,P I = E x 13 E x 14 + W P I Hot storage Ex D,HST = E x 15 E x 9 tank Cold storage E x D,CST = E x 12 E x 13 tank Turbine E x D,tur = E x 16 E x 17 W T Condenser-I Q E x D,Con I = E x 17 E x 18 E x Con I Generator-I E x D,Gen I = E x 22 + E x 30 E x 23 E x 31 E x 34 Expansion E x D,EV I = E x 32 E x 33 Valve-I Generator-II E x D,Gen II = E x 33 E x 34 + E x 35 + E x 37 + E x 38 Condenser-II Q E x D,Con II = E x 36 E x 37 E x 24 E x Con II Evaporator Q E x D,Eva = E x 25 E x 26 + E x Eva Absorber Q E x D,Ab = E x 26 + E x 40 E x 27 E x Ab IV.4. Energy and exergy efficiencies of the integrated system The definition of the energy efficiency of the system is the ratio of useful energy outputs by the system to the total energy inputs. In this paper, solar based integrated system is considered in two modes. Energy efficiency equations of the main sub-system should be written as follows for the solar mode (SM); Ƞ Rankine SM = W Rankine SM W p III Q HEX V+W p III Ƞ storage SM = Q storage tank Q HEX I+W p I Ƞ absorption SM = Q cooling SM+Q heating SM Q HEX II+W p IV (22) (23) (24) 169 Ƞ system SM = (W Rankine SM W p I W p III W p IV W CF) + Q PDC Q cooling SM+Q heating SM+Q storage SM Q PDC (25) The energy efficiency equations for the storage system mode (SSM) are given as follows; Ƞ Rankine SSM = Ƞ storage SSM = W Rankine SSM Q HEX III+W p III Q HEX III+Q HEX IV Q storage tank+w p II Ƞ absorption SSM = Q cooling SSM+Q heating SSM Q HEX IV+W p IV (26) (27) (28) Ƞ system SSM= (W Rankine SSM W p II W p III W p IV) Q storage + Q cooling SSM+Q heating SSM Q storage (29) The evaporator provides the cooling applications and the energetic coefficient of performance (COP en ) of the double effect absorption system should be written as follows; COP en = Q cooling Q gen I+W p IV (30) Exergy efficiency equations of the integrated system for the solar system mode should be written as follows; Ψ collector SM = E Q x PDC Q E x solar +W CF Ψ Rankine SM = W Rankine SM W p III Q E x HEX V +W p III Q Ψ storage SM = E x storage tank Q E x HEX I +W p I Q Ψ absorption SM = E x cooling SM Q +E x heating SM Q E x HEX II +W p IV Ψ system SM = (W Rankine SM W p I W p III W p IV W CF) Q E x cooling SM (31) (32) (33) (34) Q + E x PDC Q Q +E x heating SM +E x storage SM Q (35) E x PDC Exergy efficiency equations of the integrated system for the storage system mode should be given as; Ψ Rankine SSM = Ψ storage SSM = E x HEX III W Rankine SSM Q E x HEX III +W p III Q Q +E x HEX IV Q E x storage tank +W p II Ψ absorption SSM = E Q Q x cooling SSM +E x heating SSM Q E x HEX IV +W p IV Ψ system SSM = (W Rankine SSM W p II W p III W p IV) Q Q E x cooling SSM +E x heating SSM E x storage Q + E x storage (36) (37) (38) Q (39) Exergetic coefficient of performance (COP ex ) of the double effect absorption cooling system should be written as follows;

184 COP ex = E Q x cooling Q E x gen I +W p III (40) The exergy efficiency equations for each component of the solar based integrated system are given in Table 3. IV. Results and discussion In this chapter, the outcomes of thermodynamic assessment of the integrated system using renewable energy are exhibited and discussed. In this analysis, the mass flow rates of the working fluids are kept constant. The thermodynamic data at each state number, such as mass flow rate (kg/s), pressure (kpa), temperature ( C), enthalpy (kj/kg), entropy (kj/kgk), energy rate (kw) specific exergy (kj/kg) and exergy rate (kw) for both solar and storage sub-system modes of the multi-generation system are calculated by using the EES software program (Klein, 2007). The performance of each component of the integrated system is analyzed through different thermodynamic parameters, i.e., exergy destruction rate (kw), exergy destruction ratio (%), exergy efficiency (%) and power or heat transfer rate (kw). These thermodynamic parameters are calculated by the equations, and analysis results are presented in Table 4. Tab. 3: Exergy efficiency equations for the solar based integrated system System components Exergy efficiency equations Parabolic dish collector Ψ PDC = E x 1 E x 6 Q E x Solar HEX- I Ψ HEX I = E x 15 E x 14 E x 2 E x 3 Collector pump Ψ CP = E x 6 E x 5 W CP Pump-I Ψ P I = E x 14 E x 13 W P I Hot storage tank Ψ HST = E x 9 E x 15 Cold storage tank Ψ CST = E x 13 Turbine Ψ Tur = E x 12 W T E x 16 E x 17 Q Condenser-I Ψ Con I = E x Con I E x 17 E x 18 Generator-I Ψ Gen I = E x 31+E x 34 E x 30 E x 22 E x 23 Expansion valve-i Ψ EV I = E x 33 E x 32 Generator-II Ψ Gen II = E x 34 E x 35 E x 37 +E x 38 E x 33 Expansion valve-ii Ψ EV II = E x 36 E x 35 Condenser-II Ψ Con II = Evaporator Ψ Eva = E x Col Absorber Ψ Ab = Q E x Con II E x 36 +E x 37 E x 24 Q E x 26 E x 25 Q E x Ab E x 26 +E x 40 E x 27 Moreover, the overall exergy destruction rate is analyzed to ensure better understanding of the magnitudes of the useful work losses in the all sub-systems. In order to investigate the system performance, the exergy efficiency analysis can assist to identify the ineffectiveness within the integrated system. The system performance should 170 also be improved by reducing the heat losses in the integrated system by the re-design or modification studies. Table 6 shows that the highest exergy destruction rate occurs in the PDC system with 2042 kw, and its exergy efficiency is 80.73% for the solar mode. The component having the highest exergy efficiency is HEX-I with 98.69% for the solar mode. According to the thermodynamic assessment results, it is necessary to improve the development aims on this concentrating collector model for the more efficiency solar based integrated system design. As seen in Table 7, for storage sub-system mode, the highest exergy destruction rate belongs to the heat storage tank with kw, and also exergy efficiency of this component is 12.69%. Two components have very high exergy efficiencies; these are expansion valve-iv and expansion valve-i with 99.67% and 99.27%, respectively. In addition, exergy destruction rate of these components for the storage system mode are 0.31 kw and 0.65 kw, respectively. Tab. 4: Thermodynamic analysis results for solar mode of the renewable energy based integrated system devices Devices Exergy destruction rate (kw) Exergy destruction ratio (%) Exergy efficiency (%) Power or heat transfer rate (kw) Parabolic dish collector HEX-I Hot storage tank HEX-V HEX-II Collector pump Pump-I Pump-III Condenser-I Turbine Generator-I Generator-II HEX-VI HEX-VII Pump-IV Condenser-II Expansion valve-i Expansion valve-ii Expansion valve-iii Expansion valve-iv Evaporator Absorber Energy and exergy efficiencies of each sub-system of the renewable energy based integrated system for the solar mode and storage system mode are illustrated in Fig. 2 and 3, respectively. As it is observed, energy and exergy efficiency of the whole system for both modes are higher than all other sub-systems, because integrated systems have higher efficiency values than single output systems. According to the results shown in Fig. 2 and 3, the

185 Rankine sub-system has the highest energy and exergy efficiency among all sub-systems for each mode, mainly due to the work producing by working fluid passes successively through the Rankine turbine, and the concentrating collector has the lowest energy and exergy efficiency for the solar mode principally for irreversibility associated with the high temperature difference between the collector and ambient air. The absorption cooling sub-system has the lowest energy and exergy efficiency for the storage system mode, mainly due to the temperature differences of working fluid streams, and also due to the pressure drops in the cooling system components. It is recommended that, it should likely be important to focus development studies on the concentrating collector and absorption cooling sub-system. Otherwise, the thermodynamic analysis results illustrated that the cooling sub-system does not show important exergy destruction rate, principally this sub-system does not directly use the fuel energy but instead utilizes waste heat produced by the concentrating collector sub-system. The energetic COP of the cooling sub-system is calculated as 0.98 for two modes. This is much greater than the exergetic COP which are for solar mode and for storage system mode, respectively. parameters have important roles in the integrated system outputs. The effects of the varying ambient temperature from 0 C to 30 C on the exergy destruction rate and exergy efficiency of the integrated system for the solar mode are presented in Fig. 4 through the Fig. 8. Exergy destruction rate and exergy efficiency of the PDC system for the solar mode with respect to reference temperature is shown in Fig. 4. As seen from this figure, when the ambient temperature increases, the exergy destruction rate of the PDC system decreases from 1190 kw to 1135 kw, and exergy efficiency of the collector increases from 14.42% to 15.36%, respectively. This originates from the decrease in the temperature difference between ambient and concentrating collector. Ex D,collector (kw) Ex D,collector ycollector 0,154 0,152 0,15 0,148 0,146 ycollector , T 0 ( o C) Fig. 4: Exergy destruction rate and exergy efficiency of the parabolic dish collector system for solar mode depending on the reference temperature changes Fig. 2: Energy and exergy efficiencies for sub-systems of the integrated system for solar mode Fig. 5 illustrates the variation in the ambient temperature and their corresponding impacts on the exergy destruction rate and exergy efficiency of the Rankine sub-system for the solar mode. While the ambient temperature increases from 0 C to 30 C, exergy destruction rate of the Rankine sub-system decreases from 1700 kw to 850 kw, and also the exergetic efficiency increases from 37.21% to 41.05%. The main reason for this outcome, output exergy rate from the Rankine sub-system increases by the increasing ambient temperature, and consequently exergy efficiency of the system increases simultaneously Fig. 3: Energy and exergy efficiencies for sub-systems of the integrated system for storage system mode To analyze the efficiency variation of the renewable energy based integrated system in terms of exergy destruction rate and exergy efficiency, main design parameters such as ambient temperature and solar radiation intensity are investigated. These design 171 Ex D,Rankine-SM (kw) Ex D,Rankine-SM y Rankine-SM T 0 ( 0 C) Fig. 5: Exergy destruction rate and exergy efficiency of the Rankine system for solar mode depending on the reference temperature changes yrankine-sm

186 Similarly, in Fig. 6, by increasing the ambient temperature of the storage tank sub-system for continuously system operation, the exergy destruction rate decreases from 952 kw to 412 kw, and the exergy efficiency increases exponentially from 9.1% and 21.84%, respectively. This is due to fact that the increase in the ambient temperature for the heat storage tank sub-system requires less heat energy. In contrast, as seen in Fig. 7, exergy destruction rate of the absorption cooling system for solar mode increases by the increasing ambient temperature and exergetic efficiency decreases. This is due to increasing the temperature difference between ambient air and cooling sub-system requires more cooling load. Ex D,HST-SM (kw) T 0 ( o C) Fig. 6: Exergy destruction rate and exergy efficiency of the hot storage tank for solar mode depending on the reference temperature changes Ex D,absorption-SM (kw) Ex D,HST-SM y HST-SM Ex D,absorption-SM yabsorption-sm Fig. 7: Exergy destruction rate and exergy efficiency of the absorption cooling system for solar mode The impacts of varying ambient temperature on the whole system exergy destruction rate and exergy efficiency is shown in Fig. 8. It can be observed that, increasing ambient temperature decreases the exergy destruction rate, and increases the exergy efficiency for the solar mode. Because the solar radiation intensity varies during the solar daylight, the variations of the integrated system efficiency are analyzed. Fig. 9 demonstrates the variations of exergy destruction rate and exergy efficiency of the concentrating collector sub-system for different values of solar radiation intensity. As seen in this figure, exergy destruction rate and exergy efficiency of the PDC system for the solar , T 0 ( o C) 0.1 0,184 0,18 0,176 0,172 0,168 yhst-sm yabsorption-sm 172 mode increase with increasing solar radiation intensity. This is because increasing the solar radiation intensity increases the outlet temperature of the working fluid for the PDC sub-system. Ex D,system-SM (kw) , T 0 ( o C) Fig. 8: Exergy destruction rate and exergy efficiency of the whole system for solar mode depending on the reference temperature changes Ex D,collector-SM (kw) Fig. 9: Exergy destruction rate and exergy efficiency of the parabolic dish collector system for solar mode depending on the solar radiation intensity changes The effects of the increasing solar radiation intensity on the exergy destruction rate and exergy efficiency of the whole system for the solar mode are given in Fig. 10. Similarly, these two thermodynamic properties increase with increasing solar radiation intensity. Ex D,system-SM (kw) Ex D,system-SM ysystem-sm 0,47 0,46 0,45 0,44 0, I b (W/m 2 ) Ex D,collector-SM y collector-sm Ex D,system-SM y system-sm I b (W/m 2 ) Fig. 10: Exergy destruction rate and exergy efficiency of the whole system for solar mode depending on the solar radiation intensity changes ysystem-sm ycollector-sm ysystem-sm

187 Similar results between the exergy destruction rate and exergy efficiency based on the variable ambient temperature for the integrated-system components for the storage system mode are obtained, and the parametric studies results are shown in Figs Ex D,Rankine-SSM (kw) T 0 ( o C) Fig. 11: Exergy destruction rate and exergy efficiency of the Rankine system for storage system mode depending on the reference temperature changes Ex D,HST-SSM (kw) Ex D,Rankine-SSM y Rankine-SSM Ex D,HST-SSM y HST-SSM T 0 ( o C) Fig. 12: Exergy destruction rate and exergy efficiency of the hot storage tank for storage system mode depending on the reference temperature changes For storage mode, as seen in Figs. 11, 12 and 14, increase of the ambient temperature decreases the exergy destruction rate of the Rankine system, storage tank system and whole system, respectively, and increases the exergy efficiency of these systems. In contrast, as seen in Fig. 13, the exergy destruction rate of the absorption cooling sub-system for the storage system mode increases, and exergy efficiency decreases with increasing ambient temperature. Therefore, this situation improves the exergetic COPs of the integrated systems. Maximum exergy efficiency of whole system for the solar mode and storage system mode which are 46.8% and 46.4%, respectively, occurs at 30 C ambient temperature. At this temperature condition, exergy destruction rates of whole system are about 3500 kw and 2000 kw, respectively, for the both solar and storage system mode yrankine-ssm yhst-ssm Ex D,absorption-SSM (kw) , T 0 ( o C) Fig. 13: Exergy destruction rate and exergy efficiency of the absorption system for storage system mode depending on the reference temperature changes Ex D,system-SSM (kw) Fig. 14: Exergy destruction rate and exergy efficiency of the whole system for storage system mode depending on the reference temperature changes V. Conclusions Ex D,absorption-SSM yabsorption-ssm Ex D,system-SSM ysystem-ssm 0,158 0,156 0,154 0,152 0,148 In this paper, thermodynamic analysis results of an integrated system powered by the PDC system are presented. Because energy analysis cannot provide adequate information about the energy losses, exergy analysis is performed in order to see real efficiencies and destructions of whole system and its components. In addition, parametric studies are conducted in order to understand how environment temperature affects the exergy efficiencies of each systems and whole system. Also, solar radiation intensity impacts on the exergy destruction rate and exergy efficiency of the integrated system considered for the solar mode are investigated for a more efficient system design. According to the thermodynamic analysis results based on the first and second laws, integrated systems offer higher efficiency than single output systems. Finally, these concluding points can be drawn from the analyses: Exergy efficiencies of the Rankine cycle, PDC system, hot storage tank, absorption cooling sub-systems and whole system for the solar mode are found as 43.05%, 15.36%, 17.43%, 16.46% and 46.75%, respectively. 0,16 0,15 0,47 0,46 0,45 0,44 0, , T 0 ( o C) yabsorption-ssm ysystem-ssm 173

188 Exergy efficiencies of the Rankine cycle, storage tank, absorption cooling sub-systems and whole system for the storage system mode are calculated as 38.48%, 18.27%, 14.73% and 45.62%, respectively. Increasing ambient temperature causes an increase in the exergy efficiency of the all sub-systems except for the absorption cooling sub-system and whole system for both solar and storage system modes. Increasing solar radiation intensity causes an increase in the exergy efficiency of the PDC sub-system and whole system for the solar mode. With regard to exergy analysis results, the highest exergy destruction ratio occurs in the PDC system and hot storage tank for the solar and storage system mode with 51.30% and 36.61%, respectively. The expansion valve III in the absorption cooling sub-system has the lowest exergy destruction rates in the multi-generation system for both solar and storage system modes, because the absorption cooling components create relatively low exergy destruction in the proposed system. Nomenclature A C : Collector area (m 2 ) C : Concentrating ratio I ds : Direct solar radiation (W/m 2 ) E : Energy rate (kw) COP en :Energetic coefficient of performance COP ex :Exergetic coefficient of performance ex : Specific exergy (kj/kg) ex ch : Chemical exergy (kj/kg) ex ph : Physical exergy (kj/kg) E x : Exergy rate (kw) E x D : Exergy destruction rate (kw) E x Q : Exergy transfer associated with heat transfer (kw) E x W : Exergy transfer associated with work (kw) F R : Heat removal factor h : Specific enthalpy (kj/kg) HEX : Heat exchanger U L : Heat loss coefficient (W/m 2 K) m : Mass flow rate (kg/s) T : Temperature (K) PDC : Parabolic dish collector SM : Solar mode SSM : Storage system mode Q : Heat rate (kw) W : Work rate (kw) Greek Letters : Energy efficiency : Exergy efficiency ρ r,c : Collector reflectivity ρ r,r : Receiver reflectivity : Collector absorptivity α r,c 174 α r,r ε σ Subscript ch e i ke pe ph R u o References : Receiver absorptivity : Receiver emissivity : Stefan-Boltzmann constant : Chemical exergy : Exit : Inlet : Kinetic exergy : Potential exergy : Physical exergy : Receiver : Useful : References state Ahmadi, P., Dincer, I. & Rosen, M., Development and assessment of an integrated biomass-based multi generation energy system. Energy, Issue 56, pp Al-Ali, M. & Dincer, I., Energetic and exergetic studies of a multigenerational solar-geothermal system. Applied Thermal Engineering, Issue 71, pp Al-Sulaiman, F., Energy and sizing analyses of parabolic trough solar collector integrated with steam and binary vapor cycles. Energy, Issue 58, pp Al-Sulaiman, F., Dincer, I. & Hamdullahpur, F., Exergy modeling of a new solar driven trigeneration system. Solar Energy, 85(9), pp Al-Sulaiman, F., Hamdullahpur, F. & Dincer, I., Trigeneration: A comprehensive review based on prime movers. International Journal of Energy Research, 35(3), pp Bade, M. & Bandyopadhyay, S., Analysis of gas turbine integrated cogeneration plant: Process integration approach. Applied Thermal Engineering, Issue 78, pp Bejan, A., Tsatsaronis, G. & Moran, M., Thermal Design and Optimization. s.l.:john Wiley & Sons Inc. Buck, R. & Friedmann, S., Solar-assisted small solar tower trigeneration systems. Journal of Solar Energy Engineering, 129(4), pp Caliskan, H., Dincer, I. & Hepbasli, A., Energy, exergy and sustainability analyses of hybrid renewable energy based hydrogen and electricity production and storage systems: Modeling and case study. Applied Thermal Engineering, Issue 61, pp Carvalho, M., Serra, L. & Lozano, M., Optimal synthesis of trigeneration systems subject to environmental constraints. Energy, 36(6), pp

189 Chitsaz, A., Mahmoudi, S. & Rosen, M., Greenhouse gas emission and exergy analyses of an integrated trigeneration system driven by a solid oxide fuel cell. Applied Thermal Engineering, Issue 86, pp Dincer, I. & Rosen, M., Exergy: energy, environment and sustainable development. 2 dü. s.l.:elsevier, Oxford, UK. Dincer, I. & Zamfirescu, C., Renewable-energy-based multigeneration systems. International Journal of Energy Research, 46(15), pp Duffie, J. & Beckman, W., Solar engineering of thermal processes. s.l.:john Wiley & Sons, Inc.. Elazm, M. A., Shahata, A., Elsafty, A. & Aboelnasr, M., Parametric thermodynamic analysis for single and double effect absorption systems. Istanbul, Turkey, 10th International Conference on Sustainable Energy Technologies. Ghosh, S. & Dincer, I., Development and analysis of a new integrated solar-wind-geothermal energy system. Solar Energy, Issue 107, pp Gomri, R., Thermal seawater desalination: possibilities of using single effect and double effect absorption heat transformer systems. Desalination, 253(1-3), pp Hassoun, A. & Dincer, I., Analysis and performance assessment of a multigenerational system powered by Organic Rankine Cycle for a net zero energy house. Applied Thermal Engineering, Issue 76, pp Kalogirou, S., Solar energy engineering: processes and systems. s.l.:elsevier. Kavvadias, K. & Maroulis, Z., Multi-objective optimization of a trigeneration plant. Energy Policy, 38(2), pp system. Applied Thermal Engineering, Issue 77, pp Minciuc, E. et al., Thermodynamic analysis of trigeneration with absorption chilling machine. Applied Thermal Engineering, 23(5), pp Ozturk, M. & Dincer, I., Thermodynamic analysis of a solar-based multi-generation system with hydrogen production. Applied Thermal Engineering, Issue 51, pp Ozturk, M. & Dincer, I., Thermodynamic Assessment of an Integrated Solar Power Tower and Coal Gasification System for Multi-generation Purposes. Energy Conversion and Management, Issue 76, pp Padilla, R. et al., Exergy analysis of parabolic trough solar receiver. Applied Thermal Engineering, Issue 67, pp Soltani, R., Dincer, I. & Rosen, M., Thermodynamic analysis of a novel multigeneration energy system based on heat recovery from a biomass CHP cycle. Applied Thermal Engineering, Issue 89, pp Yuksel, Y. & Ozturk, M., Thermodynamic modelling of an integrated energy system for polygeneration design. Istanbul, Turkey, Clean Energy Conference. Zhao, Z., Zhou, F., Zhang, X. & Li, S., The thermodynamic performance of a new solution cycle in double absorption heat transformer using water/lithium bromide as the working fluids. International Journal of Refrigeration, Issue 26, pp Khalid, F., Dincer, I. & Rosen, M., Development and analysis of sustainable energy systems for building HVAC applications. Applied Thermal Engineering, Issue 87, pp Khaliq, A., Kumar, R. & Dincer, I., Performance analysis of an industrial waste heat-based trigeneration system. International Journal of Energy Research, 33(8), pp Klein, S., Engineering Equation Solver (EES), Academic Commercial, F-Chart Software. Mamaghani, A., Najafi, B., Shirazi, A. & Rinaldi, F., Exergetic, economic, and environmental evaluations and multi-objective optimization of a combined molten carbonate fuel cell-gas turbine 175

190 Heat Recovery Analysis of a Rotary Kiln in Cement Industry Ahmet Yakup Cumbul 1, Mehmet Akif Ezan 1*, Ismail Hakki Tavman 1, Arif Hepbasli 2 and M. Ziya Sogut 3 1 Dokuz Eylul University, Department of Mechanical Engineering, Izmir, Turkey 2 Yasar University, Department of Energy Systems Engineering, Izmir, Turkey 3 Bursa Orhangazi University, Department of Mechanical Engineering, Bursa, Turkey * Abstract Energy efficiency of a cement production process is quite small due to large amounts of heat loss from the systems. Rotary kilns have been widely used in the cement industry to produce clinker. The surface temperature of the rotary kiln reaches up to 300 C. Considering the higher temperature difference and higher surface area of the kiln (Diameter: 4.8 m, Length: 70 m), heat losses to the environment become significant while those from the furnace determine the overall efficiency of the cement production. In the current work, a heat recovery unit was proposed to recovery the heat losses from the rotary kiln first. A mathematical model was then developed using the Engineering Equation Solver (EES) software package to resolve the coupled balance equations. Finally, parametric results were obtained from varying the ambient conditions and working parameters of the system. It was determined that the useful heat rate recovered from the rotary kiln could reach up to 350 kw. Keywords: Rotary Kiln, heat recovery, mathematical model I. Introduction Portland cement is one of the most widely used construction material for buildings. Cement is a fine powder material and it is the core ingredient of concrete. In an industrial cement production process, the raw materials are heated inside large rotating furnaces, which is also known as rotary kiln, to produce cement clinker. Four essential elements, silicon, aluminum, iron, and calcium are used in the production of cement. Most cement production plants are settled nearby the mineral deposits of limestone. Other ingredients that are necessary for cement production are smaller amounts of clay, sand, mill scale, shale, bauxite and fly ash. The raw materials undergo several processes before entering the rotary kilns. Materials are broken into small pieces inside a crusher and then blended to prepare a mixture in a proportion of ingredient. Materials are then ground into powder and sent through the preheater and increase the temperature of the mixture before entering the furnace. Rotary kilns are slightly inclined cylindrical vessels in which the counter-current hot gases are passed through the materials and increase the temperature of the materials. The final product, which is known as cement clinker, gradually move through the rotary kiln to the bottom output and reaches to the cooling unit. The cement production process is illustrated in Figure 1. Inside the rotary kiln, the highest temperature of the raw materials reaches nearly 1500 C while the supplied flame temperature is up to 1800 C, which is roughly one-third of the surface temperature of the sun. The rotary kilns are commonly made of mild steel, which becomes weaken above 480 C. In cement plants, rotary kilns are monitored by remote temperature sensors to avoid such overheating problems and thermal deteriorations. The outer surface of the rotary kilns is exposed to the ambient to provide a natural heat rejection and provide a continuous production process. When the surface temperature of rotary kiln tends to increase up to a pre-defined upper limit, an additional cooling unit, generally an air blower unit, becomes activate to reject excessive amount of heat. Figure 1. Schematic representation of production process (Liu et al. 2015) In a recent report of The European Cement Association (Cembureau), it is revealed that the cement production sharply increases since 2001 (CEMBUREAU, 2015). Figure 2 shows the evolution of World cement production by regions. It is clear that the growth in production maintains in South America, Africa, and Asia in These areas produced the 4.4%, 4.8% and 80.4% of world cement production, respectively. According to the same report. China holds the first place in cement production with 2,438.0 million tons per year, which corresponds the 56.5% of the world cement production. European 176

191 Union countries and Turkey, on the other hand, hold positions at 3 rd and 6 th places, respectively with the production capacities of and 71.2 million tons per year, respectively. Figure 2. World cement production by regions (CEMBUREAU, 2015) Cement production is one of the most energy-consuming industries in the world. To produce one kilogram of cement clinker approximately kj of energy is required without considering the heat losses (Luo, 2015).In a real process, on the other hand, due to the significant amount of heat losses, the energy requirement increases up to kj/kg (Luo, 2015). Waste heat recovery units can be utilized to reduce the energy consumptions and improve the system efficiency. Engin and Ari (2005) assessed the possible heat recovery approaches for a dry type cement rotary kiln system. It is stated that the total energy loss of a cement plant is approximately 40% of the total input. Detailed thermodynamic analysis revealed that 19.15% is through hot flue gas, 5.61% is from cooler stack and 15.11% is due to the heat loss (convective and radiative losses) from the kiln surface. It is claimed that with a proper waste heat recovery design, 15.6% of energy could be recovered. Söğüt et al. (2010) developed a mathematical model to examine the performance of a heat recovery unit for the rotary kiln. It is found that 73% of waste heat could be recovered with the proposed heat exchanger geometry and transferred to the working fluid. Karamarković et al. (2013) stated that in a magnesium production company, the heat losses from the rotary kiln and exhaust gases are 26.35% and 18.95% of the input energy, respectively. To reduce the heat loss, they have proposed an annular duct heat exchanger. The annular heat exchanger can reduce the fuel consumption of the kiln by 12% and increases the energy and exergy efficiency of the system by 7.53% and 3.81%, respectively. Liu et al. (2014) established a combined analytical model that consist of raw material preheating & decomposition (I), clinker calcination (II) and clinker cooling processes (III). The impact of each process on the overall energy efficiency of the system is evaluated. It is found that process (I) has the highest impact on the 177 plant efficiency. It is followed by the process (III) and process (II). Caputo et al. (2011), on the other hand, analyzed a heat recovery unit to capture waste heat from the external surface of the rotary kiln. Parametric heat transfer and economic analyses have been conducted to find the optimum design and working conditions. Results confirm that the proposed heat exchanger model is appropriate both technically (efficient) and feasible (low-cost).wang et al. (2013) stated that in a cement production plant, 85% of total energy is consumed in a rotary kiln to produce the clinker. Due to convection and radiation effects on the external surface of the rotary kiln, the heat loss can reach up to 15%. They have conducted an experimental study for a prototype rotary kiln and examined the influence of working conditions on the performance of the heat recovery unit. It is found that increasing the temperature of the working fluid tends to reduce heat losses through the ambient. Recently, Luo et al. (2014) proposed a thermoelectric waste heat recovery unit to produce electricity directly from the temperature difference between the surface of the rotary kiln and the ambient. They have developed a mathematical model to predict the possible power generation and the savings by utilizing a waste heat recovery unit on a rotary kiln with dimensions of 4.8 m in diameter and 72 m long. The proposed unit produces approximately 221 kw electricity which corresponds nearly 32% of the heat loss through the kiln surface. In this study, a mathematical model is developed to predict the heat recovery from a rotary kiln under steady-state conditions. Comparative results are obtained by varying the surface temperature of the rotary kiln, mass flow rate of the working fluid, and also ambient conditions. II. Material and method III.1. Definition of the problem In the current work, the annual surface temperature variation of a rotary kiln is obtained from a cement production plant in Turkey. The outer diameter and the total length of the furnace are 4.8 m and 70 m, respectively. Considering the construction limitations on the plant site the length of the heat exchanger is decided to be 6 m. It is also known that the surface temperature of the kiln varies along the furnace. That is, parametric analyses have been conducted by varying the thermal boundary condition on the kiln so that the influence of the surface temperature on the useful (recovered) heat is obtained. The geometry of the proposed heat exchanger is given in Figure 3. The heat exchanger is designed as two half cylinders that are positioned around the kiln. The openings on the each side of the heat exchanger are intentionally designed to prevent overheating of the furnace. The surface temperature of the surface can be monitored remotely from the side opening and if the temperature reaches the upper limit, the blowers nearby the furnace are turn on to reduce the temperature. There is a total of 158 pipes

192 (carbon-steel) with an inner diameter of D in = 55.4 mm. It is proposed that there are two collectors to supply and collect the heat transfer fluid (HTF) to the pipes. The total mass flow rate of the HTF is varied in the range of 1 kg/s to 4 kg/s depending on the surface temperature of the kiln. The air gap between the furnace wall and the pipes is assumed to be Δr gap =15 cm. The outer face of the tubes is insulated (Δr ins =15 cm) to avoid heat loss through the ambient. calculated regarding the mass flow rate and the enthalpy variation of the HTF, q m h h (3) useful HTF out in HTF The heat lost from the outer surface of the insulation is composed of radiation and convection, q, A, T, T 4 4 lost sur ins ins outer ins outer sky h, total Ains, outer T ins, outer T (4) where sky temperature is defined according to the modified Swinbank equation (Hendricks & Sark, 2011) T T 0.32T (5) sky 1.5 (a) General view The total heat transfer coefficient (h,total) covers both the forced (by the wind) and natural (by buoyancy) components. The heat lost inside the air gap is expressed as q m c T T (6) lost, gap air air out in air To solve the Eqs. (1) to (6) useful heat rate and heat lost inside the air gap are defined regarding overall heat transfer coefficient. Useful heat rate is given as follows, (b) Close view Figure 3. Geometry of the proposed heat recovery unit III.2. Mathematical model & solution method Under steady-state working conditions, heat balance for the proposed system can be written as follows, q q q q (1) rad, kiln useful lost, gap lost, sur where q rad,kiln represents the net radiative heat transfer rate between the outer surfaces of furnace and pipes. q useful is the total rate of useful heat transferred by HTF. The last the terms on the right-hand side, on the other hand, represent the heat losses from the system. q lost, the gap is heat lost from kiln to the ambient air inside the gap and q lost,sur is the lost from the outer surface of the insulation. The rate of radiative heat exchange between two concentric cylinders is defined as (Cengel & Ghajar, 2011), q rad, kiln 4 4 Tkiln Tpipe, out Akiln 1 1 pipe Dkiln kiln pipe Dhx (2) Useful heat rate, on the other hand, can be 178 q useful UA T, (7) lm HTF where UA is calculated from the overall thermal resistance which is defined between the working fluid and the outer surface of the pipe. ΔT lm,htf is a log-mean temperature difference. Similarly, heat loss inside the air gap is defined in terms of the overall heat transfer coefficient as below, q UAT (8) lost, gap lm, air The numerical code is developed in Engineering Equation Solver (EES) software to solve the heat balance equations that are given in Eqs. (1) to (8). The efficiency of the heat recovery unit is defined regarding to the rate of useful heat to the rate of radiative heat transfer from the kiln surface as q useful (9) q rad, ki ln III. Results and discussions The chemical reactions inside the rotary kiln significantly affect the surface temperature of the furnace. Experimental measurements indicate that there are five different temperature zones along the length of the furnace: T kiln = 473 K, 523 K, 573 K, 623 K and 653 K. The ambient (T ) and sky (T sky)

193 temperatures vary throughout the year. That is, according to the monthly average temperature data of Izmir, Turkey (MGM, 2015), three different ambient temperatures are considered (T = 273 K, 283 K, 293 K and 303 K). Figure 4 represents the influence of ambient temperature on the performance of the heat recovery unit for T kiln = 573 K. In Figure 4(a), the variations of the outlet temperature of the HTF are given regarding the mass flow rate and also ambient temperature. It is clear that the outlet temperature decreases as the flow rate and outdoor temperature increase. Lowering the ambient from 303 K to 273 K, cause nearly 2 K reduction in outlet temperature of the HTF. On the other hand, the mass flow rate has a significant effect on the outlet temperature of the HTF. Increasing the flow rate from 2.7 to 3.5 kg/h the outlet temperature reduces nearly by 3 K. From Figure 4(b), one can see that the efficiency decreases from 61% to 54% by decreasing ambient temperature from 303 K to 273 K. On the other efficiency does not significantly influenced by varying mass flow rate for the selected parameters. Figure 4(c), indicates that increasing the flow rate from 2.7 to 3.5 kg/h the useful heat rate increases approximately by 3%. Besides, decreasing the ambient from 303 K to 273 K reduces the useful heat rate by 12%. (c) Useful heat rate Figure 4. Effect of ambient conditions on the performance of heat recovery unit Figure 5 shows the effect of surface temperature of the kiln on the performance of the heat recovery unit for T = 293 K. In Figure 5(a), the variations of the outlet temperature of the HTF are given as a function of mass flow rate and also furnace temperature. The outlet temperature increases as the flow rate decreases or the furnace temperature increases. The influence of kiln surface temperature becomes significant for lower values. At a flow rate of 1.5 kg/s, when the kiln surface rises from 473 K to 523 K the outlet temperature increases by 9 K. The outlet temperature reaches up to 400 K for T kiln = 653 K at the lowest flow rates. (a) HTF outlet temperature (a) HTF outlet temperature (b) Efficiency of the heat exchanger (b) Efficiency of the heat exchanger 179

194 (c) Useful heat Figure 5. Effect of kiln surface temperature on the performance of heat recovery unit Regarding the outlet temperature of the HTF, the difference between the highest and lowest furnace temperature cases is nearly 75 K at 1.5 kg/s. Increasing the flow rate lessen the temperature gap. At a flow rate of 3 kg/s the difference between the highest and lowest furnace temperature cases is less than 15 K. Furthermore, it is also evident that the effect of mass flow rate becomes significant at higher furnace temperatures. Increasing the flow rate from 1.5 kg/s to 5 kg/s reduces the outlet temperature by 35 K. Figure 5(b), shows that the efficiency slightly changes by varying mass flow rate but strongly influenced by the kiln temperature. At a flow rate of 5 kg/s, the efficiency of the heat recovery unit is obtained as 29%, 48%, 60%, 68% and 72% at T kiln = 473 K, 523 K, 573 K, 623 K and 653 K, respectively. Figure 5(c), on the other hand, depicts that the useful heat rate hardly depends on the furnace wall temperature. Increasing the furnace temperature from 473 K to 653 K, the useful heat rate improves more than 10 times. IV. Conclusions We have proposed a numerical model to predict the steady-state heat transfer of a heat recovery unit for a rotary kiln in a cement production plant in this study. We have considered an insulation thickness of 15 cm around the tubes while we have not included the heat bridges in the model. We have drawn the following concluding remarks from the results of the present study: a) The ambient temperature slightly affects the performance of the recovery unit. At lower ambient temperature heat lost through the environment increases. b) The useful heat rate recovered from the rotary kiln can reach up to 350 kw. c) The useful heat is recovered around the kiln could be directly used in the preheater of the furnace or could be integrated with secondary systems for providing heating or hot water to the cement plant. 180 d) In a cement production facility, the thermal management of the control system is crucial to ensure the continuity of the production line. e) The proposed recovery unit may also be integrated with an absorption cooling system to reject excessive heat from the control cabinet. f) The accuracy of the current model may be increased using geometry-specific Nusselt correlations. As a further study, the authors suggest performing an in-depth CFD (computational fluid dynamics) analysis to predict the effects of convection inside the air gap. Moreover, the heat transfer inside the tube can be calculated by dividing the tube into small segments, so that the variation of tube wall temperature along with the flow direction could be taken into account. References Caputo, A. C., Pelagagge, P. M., & Salini, P. (2011). Performance modeling of radiant heat recovery exchangers for rotary kilns. Applied Thermal Engineering, 31(14), CEMBUREAU, Activity Report 2014, The European Cement Association, May 2015 <retrieved from: Cengel, Y. A., & Ghajar, A. J. (2011). Heat and Mass Transfer: A Practical Approach, McGraw-Hill Education. Engin, T., & Ari, V. (2005). Energy auditing and recovery for dry type cement rotary kiln systems A case study. Energy conversion and management, 46(4), Hendricks, J. H. C., & Sark, W. G. J. H. M. (2013). Annual performance enhancement of building integrated photovoltaic modules by applying phase change materials. Progress in Photovoltaics: Research and Applications, 21(4), Karamarković, V., Marašević, M., Karamarković, R., & Karamarković, M. (2013). Recuperator for waste heat recovery from rotary kilns. Applied Thermal Engineering, 54(2), Liu, Z., Wang, Z., Yuan, M. Z., & Yu, H. B. (2015). Thermal efficiency modelling of the cement clinker manufacturing process. Journal of the Energy Institute, 88(1), Luo, Q., Li, P., Cai, L., Zhou, P., Tang, D., Zhai, P., & Zhang, Q. (2015). A Thermoelectric Waste-Heat-Recovery System for Portland Cement Rotary Kilns. Journal of Electronic Materials, 44(6), MGM, 2015, <retrieved from: Söğüt, Z., Oktay, Z., & Karakoç, H. (2010).

195 Mathematical modeling of heat recovery from a rotary kiln. Applied Thermal Engineering, 30(8), Wang, K., Du, W. J., & Cheng, L. (2013). Experimental Investigation on a Heat Recovery Device Installed on Cement Rotary Kiln. Applied Mechanics and Materials, 345,

196 SOLAR ENERGY 182

197 Experimental Analysis of Latent Thermal Energy Storage for Solar Heating Applications: Preliminary Results Onder Kizilkan 1*, Ahmet Kabul 2, Sefika Yildirim 3, Gamze Yildirim 4 1,2,3,4 Süleyman Demirel University, Faculty of Technology, Department of Energy Systems Engineering, 32260, Isparta, Turkey * Abstract In this study, the preliminary results of an experimental study is given for latent thermal energy storage. For this aim, an experimental setup is built for solar assisted thermal energy storage using phase change materials. The system consists of two solar air collectors and a heating coil. The phase change material is selected to be Sodium Acetate Trihydrate and it is located in a small tank inside of the heating coil. During the experiments, the air is heated up by air collectors by the help of solar energy. Then it enters to the heating coil where it gives some amount of its thermal energy to the phase change material and the rest of the air is blew to the room. The results are showed that after sunset time, the room is heated for two hours using the latent heat of PCM. Keywords: Thermal energy storage, solar energy, phase change material, latent heat I. Introduction Technological developments and increase in world population rising energy consumption rapidly. Nowadays, energy production and sustainability are important issues for humanity. In the world, energy is mostly provided from fossil fuels. However the burning of fossil fuels brought the largest environmental issue ever, which is climate change caused by CO2 emission. On this occasion, scientists had begun to research in renewable energy technologies in order to turn the tide of climate change and achieve a sustainable development for human beings (Gok, 2010). Renewable energy sources are inexhaustible resource has the potential to be different and to renewable and conventional energy sources. Production of energy from this source due to the clean and environmentally friendly sources of renewable energy sources in our country and around the world is rapidly evolving. The most important energy source in renewable forms of energy is the sun. Solar energy is abundant, renewable and free energy source. But constantly to have wavy and intermittent power characteristics of solar energy in terms of electricity supply constitute some problems. Overcoming these problems and energy storage applications in order to increase the usage rate of renewable energy is on the agenda. Especially wind and solar origin of different energy storage methods are used for the purpose of performance enhancement in renewable energy sources. Solar energy potential of renewable energy sources, taking into consideration the heat without harming the environment with solar energy produced is used for heating and storing this energy phasechange material (PCM) in the night environment. 183 Energy storage plays important roles in conserving available energy and improving its utilization, since many energy sources are discontinuous in nature. Short term storage of only a few hours is essential in most applications, however, long term storage of a few months may be required in some applications. Solar energy applications require an efficient thermal storage. Thus, the successful application of solar energy depends, to a large extent, on the method of energy storage used (Khudhair et al, 2004). Thermal energy storage (TES) is one of the key technologies for energy conservation, and therefore, it is of great practical importance. One of its main advantages is that it is best suited for heating and cooling thermal applications. TES can contribute significantly to meeting society s needs for more efficient, environmentally benign energy use. TES is a key component of many successful thermal systems, and a good TES should allow little thermal losses, leading to energy savings, while permitting the highest reasonable extraction efficiency of the stored thermal energy (Dincer et al, 2002). Amongst above thermal heat storage techniques, latent heat thermal energy storage is particularly attractive due to its ability to provide high energy storage density and its characteristics to store heat at constant temperature corresponding to the phase transition temperature of phase change material (PCM). Phase change may be in the following form: solid solid, solid liquid, solid gas, Liquid gas. Phase Change Materials (PCM) is latent heat storage material. As the source temperature rises, the chemical bonds within the PCM break up as the material changes phase from solid to liquid (as is the case for solid-liquid PCMs which are of particular interest here). The phase change is a heat-seeking

198 (endothermic) process and therefore, the PCM absorbs heat. Upon storing heat in the storage material, the material begins to melt when the phase change temperature is reached. The temperature then stays constant until the melting process is finished (Sharma et al., 2004) A literature survey about energy storage using PCM show that there is an increasing interest on TES applications with PCMs. Bhargava (1983) examined a water heater utilizing a material which changes phase for storage of solar energy. Serale et al (2014) analyzed some of the thermo-physical and rheological properties and material behavior that interest flatplate solar thermal collectors with slurry PCM as the heat carrier fluid. Concepts of solar thermal systems filled with a slurry phase change material were proposed and a prototypal system as presented. Possible advantages and drawbacks of this technology was also discussed. Arjun and Hayavadana (2014) worked on thermal energy storage materials, development of PCMs, classification, working principle of PCMs and working of PCMs in clothing. The study also summarizes the evaluation of textiles containing PCMs and different applications. Guichard et al. (2015) studied a new configuration of a complex roof using PCM installed on a dedicated test cell. For the first time and in field conditions, an experimental device using phase change material was set up at Reunion Island, location having a tropical and humid climate. They concluded that in the tested configurations and for a non-ventilated air layer, the measured temperatures were on either side of the PCM's melting point. D'Avignon and Kummert (2016) conducted an experimental study carried out to assess the performance of PCM storage tank in various operating conditions in a dynamic test bench. Soares et al. (2016) evaluated the heat transfer through small thermal energy storage (TES) units filled with different phase change materials (PCMs): free-form and microencapsulated PCMs. They reported that the experimental results were very useful for benchmarking and validation of numerical models to be used in the design and optimization of new TES systems for buildings. Li et al. (2016) studied on the influences of thermal conductivity enhancers on heat transfer performance inside the PCM during the melting and solidification processes for solar chimney application. Heinz and Streicher (2006) investigated different ways of the integration of PCMs into a thermal energy storage. Different PCM materials, with and without enhancement of the thermal conductivity, were used, and their performance concerning the resulting charge/discharge power of a storage tank was tested experimentally. In this study, it is aimed to utilize solar based latent thermal energy storage technique for the heating of a laboratory which is located in Suleyman Demirel University, Isparta. For this goal, two air solar collectors are used for heating up the air and then it is used to transfer its heat energy to PCM which is 184 placed inside of a fan coil unit, inside the laboratory. Sodium Acetate Trihydrate is selected to be the phase change material for its good properties. The results given here are the preliminary outcomes of the experimental measurements. II. Phase Change Materials The aim to use PCMs for the purpose of storing thermal energy is to make use of the latent heat of a phase change, usually between the solid and the liquid state. Since a phase change involves a large amount of latent energy at small temperature changes, PCMs are used for temperature stabilization and for storing heat with large energy densities in combination with rather small temperature changes. The successful usage of PCMs is on one hand a question of a high energy storage density, but on the other hand it is very important to be able to charge and discharge the energy storage with a thermal power, that is suitable for the desired application. One major disadvantage of latent thermal energy storage is the low thermal conductivity of the materials used as PCMs, which limits the power that can be extracted from the thermal energy storage (Heinz and Streicher, 2006) Some of the important properties required for PCMs are; High latent heat of fusion per unit mass, so that a lesser amount of material stores a given amount of energy. High specific heat that provides additional sensible heat storage effect and also avoid sub cooling. High thermal conductivity so that the temperature gradient required for charging the storage material is small. High density, so that a smaller container volume holds the material. A melting point in the desired operating temperature range. The phase change material should be nonpoisonous, non-flammable and nonexplosive. No corrosiveness to construction material. PCM should exhibit little or no super cooling during freezing (Ravikumar and Srinivasan, 2008) III. System Description The solar assisted thermal energy storage system is designed for heating of a laboratory located at Suleyman Demirel University, in Isparta. The laboratory is 46.9 m 2 in area (width: 7 m, length: 6.7 m, height: 3 m) and it is aimed to be heated by the heat energy stored in the phase change material by solar energy. The schematic drawing of the test facility is given in Fig. 1.

199 For the dimensioning of the experimental system, the calculations are made first which are given in section IV. According to the calculations, the air speed of fan is found to be 7 m/s in the solar collector system so that it is kept constant during the experiments. Each channel diameter 100 mm, and the length of the fancoil unit is 1300 mm. In order to ensure sufficient heat, two solar air collectors are mounted (Fig. 2). Fig. 1: The schematic drawing of the test facility Fig. 2: The installation of experimental setup For the phase change material, Sodium Acetate Trihydrate (CH3COONa.3H2O) is selected for latent energy storage because of its good heat transfer properties. The PCM is located in a separated cylinder inside the fan-coil unit so that the heated air flows over the PCM and transfers its heat energy to it (Fig. 3). The properties of the Sodium Acetate Trihydrate is given in Tab. 1. Tab. 1: Properties of sodium acetate trihydrate Phase Solid Color colorless Density ~ 1.42 gr/cm 3 (20 C) Solubility ~ 613 g/lt (20 C,H 2O) Flashpoint temperature > 250 C (anhydrous) Melting point ~ 58 C ph ~ (50 g/lt,h 2O,25 C) 185

200 Re = m D A μ (9) where μ is the dynamic viscosity. In equation 9, A and D can be found from: A = s W (10) D = 2 s (11) Fig. 3: Sodium acetate trihydrate located inside the fan-coil IV. Thermodynamic Calculations For the thermodynamic calculations of the solar air collector, the mathematical formulation given in reference Kalogirou (2009) is used. Also, similar design calculations can be found in some other literature (Duffie and Beckman 2013; Tiwari 2003). The useful energy absorbed by the solar collector is defined as (Kalogirou 2009): Q u = A c F R [S U L (T i T a )] (2) where FR is the heat removal factor, S is the absorbed solar radiation, UL is the heat loss coefficient, Ti is the inlet air temperature, Ta is the ambient temperature. F R = m c p A c U L {1 exp [ U L F A c m c p ]} (3) where m is the mass flow rate of air, cp is the specific heat of air, F is the collector efficiency factor. where s is the depth of air channel, W is the collector width. h[1] = σ (T p+t b ) (T p 2 +T b 2 ) ( 1 ɛp )+( 1 ɛ b ) 1 (12) where, ɛ p is the emissivity of absorber plate, ɛ b is the emissivity of back plate, Tp is the temperature of absorber plate, Tb is the temperature of back plate. Fan selection is made and the air flow should be done in the laboratory for the design of the device must be known. Therefore intended to be heated the quantity of air required for laboratory; Q L = V m H d (13) where, Q L is the air flow required for laboratory, Vm is the volume of laboratory, Hd is the number of hour weather cycle For the pressure drop of the whole system, the formula given below is used: ΔP t = ΔP s + ΔP d = ( l R + Z) + P E (14) T 0 = T i + ( 1 U L ) [S U L (T i T a )][1 exp ( A c U L F m c p )] (4) where T0 is the exit air temperature, Ti is the inlet air temperature and Ta is the ambient temperature. S = G t (τ α) (5) Here Gt is the total insolation, (τ α) is the effective coefficient. In equation 4, F is given below: h F = (6) h+u L where h is the convection heat transfer coefficient. h = h[2] + ( 1 ( 1 h[1] )+( 1 h[2] )) (7) where h[1] is the radiation heat transfer coefficient from the absorber to the back plate, h[2] is the convection heat transfer coefficient. h[2] = ( k D ) (Re)0.8 (8) Here k is the heat transfer coefficient, D is the hydraulic diameter, Re is the Reynolds number. 186 Where, l is channel length (m), R is unit pressure drop (Pa/m), Z is pressure drop of fittings elements (Pa), PE is the total pressure drops (Pa) due to devices such as filters, heaters, measurement devices, etc. Additionally, ΔP s and ΔP d are the static and the dynamic pressure drops, respectively. V. Result and Discussion The experiment were made in June during a sunny day. The measurements are made for solar radiation intensity, outer temperature, indoor blowing air temperature at the exit of fan-coil and indoor temperature. In Fig. 1, the variation of solar radiation intensity is given with variation of time for Isparta region. As can be seen from the figure, the radiation intensity is about 650 W/m 2 at 11 am while it reaches to a maximum value of about 1100 W/m 2 between 1 pm to 3 pm. These values show that Isparta city has a great potential of solar energy. In Fig. 2, the variation of blowing air temperature at the exit of solar collectors is given with the variation of time. The first measurements have shown that the air temperature after the collectors was lower than the melting temperature of PCM. As seen from Fig. 2, maximum air temperature after the collectors was

201 measured about 44 C. This is obviously lower than the melting point of sodium acetate trihydrate (58 C). It was expected according to the calculations that, the air temperature at the exit of the collectors to be about 60 C. This unexpected result is due to improperly designed channels inside the collectors and also the number of collectors can be increased to 3. Another option to solve this problem in order to store the solar energy in PCM is to utilize a different PCM which has got a melting temperature around C. Fig. 3: Variation of outdoor temperature with time Fig. 1: Variation of solar radiation intensity with time Fig. 2: Variation of blowing air temperature with time In Fig. 3 and 4, the variation of outdoor and indoor temperatures are given with the variation of time, respectively. It is obvious from the figures that, the temperature difference after the sun rise is getting bigger when compared to the day time. From this situation, it can be said that in spite of not reaching to the melting temperature of PCM at the exit of air collectors, a small amount of sensible heat has been transferred to the ambient. Fig. 4: Variation of indoor temperature with time To ensure the effect of solar energy storage system after the sunrise, the measurements were continued till 9 pm in the evening. Figs. 5-7 show the variation of indoor, outdoor and blowing air temperatures between 7 pm 9 pm. As can be seen from the figures that after sun rise, the indoor temperature decreases from 28 C to 26 C while the outdoor temperature decreases from 23 C to 21 C. As mentioned earlier, this temperature difference is due to sensible heat storage. Fig. 5: Variation of outdoor temperature with time after sunset 187

202 References Arjun D., Hayavadana J., Thermal Energy Storage Materials (PCMs) for Textile Applications, Journal of Textile and Apparel Technology and Management, 8(4), 1-11, Bhargava A.K., A solar water heater based on phasechanging material, Applied Energy, 14(3), , Fig. 6: Variation of indoor temperature with time after sunset Duffie J.A., Beckman W.A., Solar Engineering of Thermal Processes, John Wiley & Sons Inc., New Jersey, USA, Gok O., Increasing Energy Efficiency In Dishwashers By Using Thermal Energy Storage In Phase Change Materials, PhD Thesis, Cukurova University, Turkey, D'Avignon K., Kummert M., Experimental assessment of a phase change material storage tank, Applied Thermal Engineering, 99, , Dincer I., Rosen M.A., Thermal energy storage, systems and applications. John Wiley and Sons, England, Fig. 7: Variation of blowing air temperature with time after sunset VI. Conclusions Utilization of thermal energy storage applications are being increased for sustainable use of energy. Latent heat thermal energy storage with PCMs is particularly attractive due to its high energy storage density amongst TES techniques. In this study, a solar assisted latent thermal energy storage application was established experimentally for heating of a laboratory at Suleyman Demirel University, Isparta. Sodium acetate trihydrate was used as PCM. The preliminary results of the experimental measurements showed that, the blowing air has been heated up to a maximum 44 C which is not adequate for melting of the PCM which has got a melting temperature of 58 C. There were two options discussed to solve the problem. The first one was to use 3 solar air collectors instead of 2 and the other one is to use a different PCM which has got a melting point range between C. The next step of this study will be mounting automatic control system to TES application and also a different PCM will be used. Guichard S., Miranville F., Bigot D., Damour B.M., Boyer H., Experimental investigation on a complex roof incorporating phase-change material, Energy and Buildings, 108(1), 36-43, Heinz A., Streicher W., Application of Phase Change Materials and PCM-Slurries For Thermal Energy Storage, Proceedings of the Ecostock Conference, USA, Kalogirou S.A., Solar Energy Engineering: Processes and Systems, Academic Press, Oxford, UK, Khudhair A.M., Farid M.M., A review on energy conservation in building applications with thermal storage by latent heat using phase change materials, Energy Conversion and Management, 45, , Li Y., Liu S., Shukla A., Experimental analysis on use of thermal conductivity enhancers (TCEs) for solar chimney applications with energy storage layer, Energy and Buildings, 116, 35-44, Ravikumar M., Dr. Pss. Srinivasan, Phase change material as a thermal energy storage material for cooling of building, Journal of Theoretical And Applied Information Technology, , Seralea G., Casconea Y., Capozzolia A., Fabriziob E., Perinoa M., Potentialities of a Low Temperature Solar 188

203 Heating System Based on Slurry Phase Change Materials (PCS), Energy Procedia, 62, , Sharma S.D., Kitano H., Sagara K., Phase Change Materials for Low Temperature Solar Thermal Applications, Res. Rep. Fac. Eng. Mie Univ., 29, 31-64, Soares N., Gaspar A.R., Santos P., Costa J.J., Experimental evaluation of the heat transfer through small PCM-based thermal energy storage units for building applications, Energy and Buildings, 116, 18-34, Tiwari G.N., Solar Energy: Fundamentals, Design, Modelling and Applications, Alpha Science International Ltd., Pangbourne, UK,

204 A review of Solar Energy Status in Iraq and Current Status Ahmed Emad*, Ahmet Kabul, Onder Kizilkan Süleyman Demirel University, Faculty of Tchnology, Department of Energy Systems Engineering, 32260, Isparta, Turkey * Abstract There is no secret how renewable energy is important and useful by increasing concern into improve its efficiency and utilize it as an alternative energy clean and sustain comparing with fossil energy and other energy recourses. the paper release how to utilize solar energy in Iraq depending on studies of sun radiation and lights incidence area, and annual hot weather temperature per a year, in addition study of changing climate,so we could use these previous studies to analyze and compare of using solar energy types, depends of results duration life time usage sustain and costs, this paper show different ways to use renewable solar energy in Iraq,and looking for finding a way to use a better kind of renewable energy with high efficiency and long life duration usage of power, whereas Iraq suffers from short age power usage and the effects of war led to collect all efforts to find a way to improve the power energy production. Keywords: Renewable energy, solar thermal energy, Concentrated solar power I. Introduction Renewable energy sources: are energy resources that are inexhaustible within the time horizon of humanity. Renewable types of energy can be subdivided into three areas: solar energy, planetary energy and geothermal energy. Even if the use of fossil fuels can be reduced significantly, and accepting that nuclear power is no long-term alternative, the question remains as to how the future supply of energy can be secured. The first step is to significantly increase the efficiency of energy usage, i.e. useful energy must be produced from a much smaller amount of primary energy, thus reducing carbon dioxide emissions. Renewable energies will be the key to this development, because they are the only option that can cover the energy demand of Earth in a climatically sustainable way (Quaschning, 2005). Now, Renewable energy become an abundant, wellestablished technology and the main ingredient is free. It is a well-known fact that eight countries have 81% of all word crude oil reserves, six countries have 70% of all natural gas reserves, and eight countries have 89% of all coal reserves. More than half of Asia, Africa and Latin America import over half of all their commercial energy. Most of these countries export crops that fetch low prices, but import energy at high prices, which leads to a drain on foreign exchange earnings. This problem is worsened by the fact that power generation is continuously increasing in these countries (Table 1) additionally, the world population keeps increasing at 1.3 2% per year, so that we are doubling our population every 60 years. Therefore in the year 2060, we expect our population will be in excess of 12 billion (Sari, 2004). Tab. 1: World total final consumption (Mtoe) (Source: (IEA, 2016) a Coal Oil Gas Electricity Heat Renewables Total final consumption a Average annual growth rate, in percent. The sun is an excellent source of radiant energy, and is the world s most abundant source of energy. It emits electromagnetic radiation with an average irradiance of 1353 W/m2 on the earth s surface. To put this into perspective, if the energy produced by 25 acres of the surface of the sun were harvested, there would be enough energy to supply the current energy demand of the world (Chaichan and Abaas, 2012). Recently, there is an increase needed for energy, especially electrical energy. Not only are oil prices increasing but pollution continues to rise due to the burning of fossil fuels, and the probability of oil supply depletion remains. All of these Issues encourage the investigation of using solar, wind and other Renewable energies for the generation of electrical power (Kazem and Chaichan, 2012), Renewable resources gained more attention in the last two decades due to persisting energy demand coupled with decrease in fossil fuel resources and its environmental effect to the earth. This paper reviewed and discussed the status of renewable energy used and developments in Arab countries especially in Iraq, how to apply and improve these systems as an essential element for the 190

205 sustainable economic development of these countries, despite their wealth in oil and gas. The paper presents a review of the renewable energy resources in Arab countries and sheds light on some achieved and/or ongoing renewable energy projects in the (table.2). Also show how Iraq suffers from electricity reduction, and many challenges will have to face the future increases in electrical prodution.as well as the impact of war, Years of underinvestment in the power sector. In Iraq, the electric power generated is not enough to meet the power demand of domestic and industrial sectors. Tab. 2: Solar-energy resources (kwh/m 2 /day) Country Solar Country Solar energy energy Algeria 5 7 Oman 5 6 Bahrain 5 8 Palestine 4 6 Egypt 5 9 Qatar 5 6 Iraq 5 6 Saudi Arabia 6 8 Djibouti 4 6 Sudan 5 8 Jordan 5 7 Somalia 6 9 Kuwait 5 8 Syria 5 6 Lebanon 4 6 Tunisia 5 7 Libya 5 7 UAE 5 6 Mauritania 6 Yemen 4 6 Morocco 5 7 II. Literatures overview By concerning of sustainable renewable energy in Arab countries especially in Iraq this paper refer to how this region could utilize available solar energy in real, associating with other reviews and studies of other researches. Quaschning (2005), described the most important technical systems for using renewable energy sources, and introduces important calculation and simulation methods for these. The main focus was on technologies with high development potentials such as solar thermal systems, photovoltaics and wind power. Sari, (2004), discussed the growing need of energy in both developed and developing countries, and the acute population growth. In the results, it was observed that renewable energy penetration into the energy market was much faster than was expected in recent years and by Chaichan and Abaas (2012), focused on the feasibility of improving concentrating solar power (CSP) planet efficiency, by manufacturing a diminished prototype. They investigated three states, coloring the central target with selective black color, enclose the target by a glass box and coloring the glass encloser by selective black color. The tests were conducted in Baghdad- Iraq summertimes weathers The results showed improvement in thermal storage when using the glass encloser. This improvement varied form month to month. Kazem and Chaichan, (2012), reviewed and discussed the status and future of renewable energy in Iraq. The uses of renewable energy sources, such as solar, wind and biomass have been reviewed. Concluding with recommendations for the utilization of these energy resources, investigated found that solar, wind and biomass energy were not being utilized sufficiently at 191 present, but according to them, these energies could play an important role in the future of Iraq s renewable energy. The authers showed aims to review and discuss the status and future of renewable energy in Iraq. The uses of renewable energy sources, such as solar, wind and biomass, have been reviewed. Chedid and Chaaban (2003), reviewed and discussed status of renewable-energy (RE) developments in Arab countries (AC) as an essential element for the sustainable economic development of these countries, despite their wealth in oil and gas. In conclusion, they have given a detailed status of RE developments in AC. Also RE resources were introduced, and success stories in selected countries were extracted and discussed. Valenzuela (2010), studied and simulated the viability of introducing a 50 MW solar power plant in the locality of Barletta, Italy. The study was divided in two main parts: The first theoretical one which was about solar energy and explained the two main processes with which profit could be taken from the sun: photovoltaic energy and thermal energy. In the second part, the 50 MW plant was studied and simulated to arrive to the final design. Dickes et al., (2014) have desigened a lab-s solar power plant and for experimental purposes and dynamic analysis. The test rig included an Organic Rankine Cycle (ORC) unit, a field of parabolic trough collectors and a thermal energy storage system presents the results of an experimental campaign conducted on the ORC module alone. This power unit, designed for a 2.8 kw net electrical output. Al-Karaghouli and Kazmerski (2010) addresses the need for electricity of rural areas in southern Iraq and proposes a photovoltaic (PV) solar system to power a health clinic in that region, analysis shows that the optimal system s initial cost, net present cost, and electricity cost, respectively. These values for the PV system are compared with those of a generator alone used to supply the load. Using the HOMER software computer model, determined that the most economic system for a remote health clinic in southern Iraq having a daily load. Ismael Mohammed Saeed et al (2016) worked at study the current and future energy issue such as the energy policy revolution, the power sector expansion strategy and the corresponding environmental impact in Iraq. It aims to introduce the capabilities of renewable and nuclear energy and deals with the environmental impact of renewable and nuclear energy, especially in greenhouse-gas mitigation. Also review the capabilities of renewable and nuclear energy in Iraq. Long-range Energy Alternative and Planning (LEAP) System was used to analyse the electricity generation including the future expansion plan and simulates the environmental impact for every technology used in electricity generation. III. Solar Energy Potential of Iraq Solar energy is a highly renowned alternative energy type. The intensity of the Sun s irradiation that reaches 92 billion the tons of globe petroleum. A calculation from 2002 states that the energy received from the sun in one hour was greater than the world

206 used in one year. Due to the latter, the investigation of this source has lasted for years and continues to be a matter of great importance and relevance today There are two main ways of taking profit of the energy, sun s which are photovoltaic plants and solar collector plants (Valenzuela, 2010). Iraq is well-known for long hours of sunshine. Studies have shown that Iraq receives more than 3000 hours of solar radiance per year in Baghdad alone. The hourly solar intensity varied between 416 W/m 2 in January to 833 W/m 2 in June. Even the hours of sunshine in Spain cannot compete with the levels observed in Iraq (Kazem and Chaichan, 2012). Iraq has excellent solar energy potential, ranging from 1800 to 2390 kwh/m2/yr of direct normal irradiation, and much of the flat Iraqi landscape is appropriate as shown in (Figure 1.) Also shown are the locations where the Ministry of Electricity plans to issue concessions for the 3,500 MW in total capacity via fossil-fuel power plants. Fig. 1: Iraq s solar irradiation (Source: The German Aerospace Center (DLR), Iraq Ministry of Electricity) There are two basic principles for converting solar energy. Photovoltaic systems convert solar energy directly into electrical energy. Tracking systems are used in concentrated photovoltaic (CPV) in particular. These use photovoltaic in power plants for the central energy supply in order to generate higher efficiency. Solar thermal systems have reflectors that concentrate the sun s rays in an absorber to heat a fluid that generates water vapor using a heat exchanger. The water vapor is used to drive turbines and generators as in conventional thermal power plants. Solar thermal Power plants include parabolic trough, Fresnel and solar tower power plants. Dish Stirling power plants operate differently here concentrated sunlight directly drives a Stirling engine that in turn drives a generator. One thing all these solar applications have in common is a single-axis or Double-axis tracking system that continuously aligns the reflectors with the course of the sun (Schaeffler, 2013). Iraq is one of the hottest countries in the world (with summer temperatures up to C), and summer temperatures are steadily increasing. About 50% of overall electricity demand is due to air conditioning, People in Baghdad, especially, are desperate to buy, and hopefully have enough electricity to use, air conditioners, as noted frequently in media. Lack of electricity during the critical summer months affects national productivity and makes it difficult to work in the stifling heat. As a result of the electricity shortages and demand for air conditioning, 90% of Iraqi households rely on some sort of diesel power generation operated by private independent operators (Figure2). 192

207 Fig. 2: Ad-hoc power distribution grids established by entrepreneurs who run diesel generators to feed demand when the grid is down (Undp, 2015) III.1. Photovoltaic (PV) systems Due to uniform distribution of solar radiation throughout Iraq, solar PV technology is suitable for producing electricity through-out Iraq. Solar PV technology is also suitable for off-grid electricity generation in power plants in rural desert areas. The efficiency of PV cells is influenced by high air temperature and dust contamination. Due to the dusty weather in Iraq, it is important to investigate the type of dust, density of dust, rate of accumulation of dust, and the effect of dust on the PV performance. Iraqi experiments using photovoltaic (PV) cells were unsuccessful. In Iraq, photovoltaic cells were used in community street lights but were unsuccessful because the cells had a low efficiency factor and Iraqi weather is characterized by dusty days (Figure 3 and 4). In addition, power from PV systems is currently more expensive comparing with large-scale concentrating solar power These factors acted and reduced the range of use of PV cells, though the cells did find limited application in individual home rooftop systems, community water pumping stations, and areas where the terrain makes it difficult to access the power grid (Kazem and Chaichan, 2012). Fig. 3: Solar street lights installed in Iraq Also Iraqi government started building dams in early 1950, to reduce the flood impact and developing the agricultural sector in the country. The company that constructed the first dam proposed to the Iraqi government to exploit the dam also for producing electricity, and the government accepted. This was a good step in electricity generation in Iraq. The first hydroelectric power station to supply the national grid network was the Dokan station in Initially, it applied 84 MW installed power and later increased it to 400 MW. Iraq has five main hydroelectric plants, with about total initial capacity of 2500 MW. The electricity produced by hydroelectric plants is affected by many factors, thus a continuous steady production is impossible. Among the factors is the height of stored water in the dam, which is decided by the rate of rainfall. Dams in Iraq are also used for irrigation; this means faster depletion of the stored water in drought years, which happened in Iraq frequently. Tables 3 and 4 show the currently operational dams and those under con-struction. Iraq aims to increase her installed capacity from hydroelectricity to 5500 MW after completion of these projects(saeed.et al,2016) Tab. 3: Dams with hydroelectric plants being constructed in Iraq and their maximum capacities (Movr, 2016) Name Location Capacity/MW Status Mosul Ninawa 1010 Active Haditha Anbar 660 Active Dokan Sulaimaniyah 400 Active Derbandikhan Hexagonal 240 Active Samara Salahuddin 75 Active Hamrin Dyala 50 Active Hindiya Karbala 15 Active Kufa Karbala 6 Active Total 2456 Fig. 3: Electrcity generation from PV sysmtems in Iraq 193

208 Tab. 4: Under-construction dams with hydroelectric plants in Iraq with expected capacity (Movr, 2016) Name Location Capacity/ MW Status Bekhme Erbil 1500 Under Construction Mandawa Erbil 620 Under Construction Taq Taq Erbil 300 Under Construction Al-Baghdadi Anbaar 300 Under Construction Badush Ninawa 171 Under Construction Bakerman Ninawa 24 Under Construction Total 2615 III.2. Solar thermal systems Among the different technologies being developed to this end, Concentrated Solar Power systems is a promising renewable technology. A standard CSP technology uses solar collectors and a tracking system in order to concentrate solar rays during sunshine hours. This concentrated beam is then absorbed and used as the heat source for a thermodynamic cycle. (Figure 4) illustrates the working principle of such system using parabolic trough collectors as sol receivers (Dickes et al., 2014). Fig. 5: Concentrating solar power distribution in the world (Trec, 2016) Fig. 6: Global hours of bright sunlight (Wikipedia, 2016) Fig. 4: Parabolic trough collectors as sol receivers Concentrated solar power (CSP) is expected to be very well suited to the long days of sunshine and the high temperatures found in Iraq (Figure 5-6). Investigations are underway in Iraq to improve the use of CSP during high temperature weather conditions. The use of solar water heater systems by domestic loads has increased. PV and/or CSP system implementations have shown that their efficiency and reliability depend on many factors, including orientation (longitude and latitude), environment (solar intensity, temperature, humidity, wind, dust, rain, pollution, etc.) And the PV technology used. Thus, before committing to a large-scale (in megawatts) PV or CSP project, a thorough investigation of the above factors is essential (Kazem and Chaichan, 2012). III.3. Hybrid Power Systems (HPS) Solar energy is copious, but capturing it is not cheap, which is the primary reason that solar power contributes only a tiny fraction of global energy production. Solar remains many times more expensive than power derived from fossil fuels, even as oil and natural gas prices rise. The intermittency issue could be solved by combining a solar thermal plant with a natural gas plant, enabling the solar power shortage during Iraq s winter months to be covered by gas-generated power. Maximum solar output during Iraq s intensely sunny summers, meanwhile, would coincide with peak demand and lessen the need for additional peaking capacity from gas. Colocating with a gas plant would also answer the solar plant s need for accessible transmission lines, a critical siting criterion for many renewable projects. Most importantly, combining a solar parabolic trough plant with a gas-powered plant would reduce costs because both can utilize the same steam turbine, generator, and associated equipment (Doyle and Jaafar, 2010). Hybrid power systems (HPS) are any autonomous electricity generating systems combining renewable energy sources and conventional generators. Winddiesel systems, which combine wind turbines and diesel generators are a subclass of HPS. The purpose of these systems is to produce as much energy as 194

209 possible from the renewable sources while maintaining an acceptable power quality and reliable supply. Furthermore, the fuel savings and lower generation costs obtained with the HPS should at least balance the high investment costs (Figure7) for renewable energy generators, controllers, dump loads, storage units, converters, etc. (Pereira, 2000). Financial support for studies that investigate renewable energy in Iraq and its applications is required. Introducing solar thermal collectors in public buildings to produce hot tap water can be considered a first step towards reducing dependence on fossil fuel resources. Focusing and working on using hybrid systems by combining renewable energy source with conventional generator its suitable efficiency comparing with climate change in this area and standalone renewable energy systems. Referemces Al-Karaghouli A., Kazmerski L., Optimization and lifecycle cost of health clinic PV system for a rural area in southern Iraq using HOMER software, National Renewable Energy Laboratory, 1617 Cole Blvd., Golden, CO 80401, USA, Fig. 7: current and potential future costs for largescale solar power systems in high-resource locations vs. fossil power Most of the rural areas in southern Iraq are still undeveloped and in a chaotic state after the invasion, and there is a need to provide these areas with electricity. Small standalone photovoltaic (PV) electrification systems can play a strategic role in the region s development. The region enjoys a huge amount of solar radiation during the entire year. Although capable of providing plentiful and reliable electricity, HOMER software developed by the National Renewable Energy Laboratory (NREL) to assist the design of micropower systems. HOMER is a computer model that simplifies the task of evaluating design options for both off-grid and grid-connected power systems for remote, stand-alone, and distributed-generation (DG) applications. HOMER s optimization and sensitivity analysis algorithms allowone to evaluate the economic and technical feasibility of a large number of technology options and to account for variation in technology costs and energy resource availability. HOMER models both conventional and renewable-energy technologies (Al- Karaghouli and Kazmerski, 2010). IV. Conclusions Chaichan M.T., Abaas K.I., Practical investigation for measurement of concentrating solar power prototype for several target cases at Iraqi summertime weathers, Anbar Journal for Engineering Sciences, 5(1), 76-87, 2012 Chedid R., Chaaban F., Renewable-energy developments in Arab countries: a regional perspective, Applied Energy 74, , Dickes R., Dumont O., Declaye S., Quoilin S., Bell I., Lemort V., Experimental Investigation of an ORC System for a Micro-Solar Power Plant, Proceedings of the 2014 Purdue Conferences, Purdue University, USA, Doyle P., Jaafar K., Iraq Has an Opportunity to Become a Solar Leader, Iraqi Solar, Developmentsarticle.pdf, IEA, International Energy Agency, Kazem H.A.,Chaichan M.T., Status and future prospects of renewable energy in Iraq, Renewable and Sustainable Energy Reviews, 16(8), , By all reviews and studies above it is concluded that that Iraq is suitable area for applying sustainable renewable energy so that, by study these reviews we can illustrate the factors that could be play an important rules to improve sustainable renewable energy in Iraq: The solar energy density in Iraq is among the highest in the world. Additionally, there is significant wind energy potential in several areas in Iraq. Government support is required for implementing small, renewable energy pilot projects, especially those that serve people in rural areas. 195 Movr, Ministry of Water Resources Iraq, Pereira A.L., Modular supervisory controller for hybrid power systems, Risø National Laboratory, Roskilde, https://www.researchgate.net/publication/ _Modular_supervisory_controller_for_hybrid_power_ systems, Quaschning V., Understanding Renewable Energy Systems, Earthscan, United Kingdom, Saeed I.M., Ramli A.T., Saleh M.A., Assessment of sustainability in energy of Iraq, and achievable

210 opportunities in the long run, Renewable and Sustainable Energy Reviews, 58, , Sari A, Renewable energy for a clean and sustainable future, Energy Sources, 26, , Schaeffler Technologies AG & Co. KG Issued: 2013, April. Trec, Undp, United Nations Development Programme, NDPIQ_catalysing_solarenergy_ProDoc.docx/_jcr_c ontent/renditions/page.html, Valenzuela J, Performance of a 50 MW concentrating Solar Power Plant, Politecnico Di Bari University, Wikipedia, https://en.wikipedia.org/wiki/solar_power_in_africa#/ media/file:solargis-solar-map-world-map-en.png,

211 Effect of Solar - Geothermal Heat Exchanger Design and Fluid Type on the Thermodynamic Performance of a Power Plant Anil Erdogan 1*, Duygu Melek Cakici 1, Can Ozgur Colpan 2 1 Dokuz Eylul University, Graduate School of Natural and Applied Sciences, Mechanical Engineering Department, Izmir, Turkey 2 Dokuz Eylul University, Faculty of Engineering, Mechanical Engineering Department, Izmir, Turkey * Abstract A design problem for shell and tube heat exchanger that combines a Parabolic Trough Solar Collector (PTSC) and Organic Rankine Cycle (ORC) was formed for the given fluids using their temperatures, pressures and mass flow rates. The design problem was modeled and solved using Engineering Equation Solver (EES). Using a thermal model for the PTSC, first, glasscover temperature of the PTSC was calculated, and then useful energy gain and the temperature of the fluid leaving the solar collector were found. The design problem of heat exchanger formed was solved using Logarithmic Mean Temperature Difference (LMTD) method; and heat transfer surface area, overall heat transfer coefficient, pipe side and shell side heat transfer coefficients, and the pressure drop along the heat exchanger were found. A parametric study was performed to study the effect of some of the important parameters (e.g. pipe diameter, pipe length, and baffle spacing) on the output parameters (e.g. overall heat transfer coefficient and pressure drop). The results show that R600 as the tube side fluid and Therminol VP1 as the shell side fluid give better performance. When the solar irradiation intensity changes between 800 and 2700 kwh/m 2, overall heat transfer coefficient decreases from 1579 W/m 2 -K to 1491 W/m 2 -K, heat transfer surface area increases from 7 m 2 to m 2, and pumping power increases from kw to kw. Keywords: Parabolic trough solar collector (PTSC), shell and tube heat exchanger, logarithmic mean temperature difference (LMTD), engineering equation solver (EES) I. Introduction Solar collectors absorb the incoming solar radiation and transfer the heat absorbed into a fluid (e.g. air, water, or thermal oil) circulating through the tubes of the collector (Mills, 2004). This fluid is then directly used or transfers its heat to another fluid for the desired purpose (e.g. production of hot water or steam). In order to reach high temperatures with high efficiency, Parabolic Trough Solar Collectors (PTSC) are generally preferred. These collectors are light structures and have low cost and used for process heat applications between 50 C and 400 C (Duffie and Beckman, 2013; Fernández-García et al., 2010). PTSCs are made of a reflective material sheet which has a form of parabolic shape. The schematic of a PTSC is given in Figure 1. As can be seen in this figure, the receiver consists of a metal tube (colored black) surrounded by a glass cover which is used to reduce the undesirable heat losses. PTSCs provide higher solar concentration levels according to flat plate collectors. Collector performance, which depends on design and material, is significantly affected by factors such as reflectivity, receiver, absorptivity, heat transfer fluid, and its flow rate tracking mechanism, among others (Kalogirou, 2004; Kalogirou, 2009). Tracking mechanism must be safe and able to follow the sun with certain degree of accuracy. It returns the collector to its original position at the end of the day. Also it should work during periods of sparse cloud cover. In addition, tracking mechanisms are used for protection of collectors. When the collectors turn, mechanisms protect from hazardous environmental and working conditions (e.g. storm, overheating, and the failure of working fluid) (Nuwayhid et al., 2001). Tracking mechanisms are divided into two categories, namely mechanical and electronic controlled systems. Electronic controlled tracking mechanisms are generally used for PTSCs. These system improved accuracy and reliability. Additionally, electronic controlled tracking mechanisms together with the sensor and computer controlled motors measure the solar radiance on the receiver (Islam et al., 2015). Fig. 1: Schematic of a PTSC (Modified from Islam et al., 2015) There are some studies on the PTSC design and modeling in the literature. Kalogirou et al. (1996) conducted a performance test of a PTSC according to 197

212 ASHRAE standards. The collector efficiency and incidence angle were measured and as a result of these measurements the collector s acceptance angle was obtained in the range of ±0.5. In addition, when the maximum error of the tracking mechanisms was ±0.2, the system worked continuously at the maximum possible efficiency. In the study by Odeh et al. (1998), efficiency of PTSCs was determined for its operation with Syltherm 800 oil and water as the working fluids. Absorber emissivity effects and internal working fluid convection effects were evaluated. An efficiency equation was developed and used in a simulation model to evaluate the performance of the system for different radiation conditions and different absorber tube sizes. Herrmann et al. (2004) proposed a concept, where a less expensive liquid medium such as molten salt is utilized as storage medium rather than the heat transfer fluid itself. Detailed performance and cost analyses were performed. The study concluded that specific cost for two tank molten salt storage is in the range of US $30-40/ kwh depending on storage size. Kalogirou et al. (2004) discuss the different types of solar collectors (flat plate, compound parabolic, evacuated, parabolic trough, Fresnel lens, heliostat field collectors) and their applications. Optical, thermal and thermodynamic analyses of the collectors, and a description of the methods used to evaluate their performance were given. Brooks et al. (2005) characterized the performance of PTSC. Low temperature testing was performed using water as the working fluid. Evacuated glass-shield and unshielded receiver were tested and it was found that the thermal efficiencies are 53.8% and 55.2%, respectively for these receivers. Brooks et al. (2005) developed a PTSC in a similar size to the small-scale commercial modules. In this study, the working collector length is 5 m, aperture width is 1.5 m, and rim angle is In addition, optical error analysis was conducted to estimate the intercept factor. In the study by Qu et al. (2006), a performance model of a solar collector based on a linear and tracking parabolic trough reflector was programmed using the software EES. The model included fundamental radiative and convective heat transfer, and mass and energy balance relations. Typical performance showed that when the hot water at 165 o C flows through a 6 m by 2.3 m PTSC with 900 W/m 2 solar insulation and 0 incident angle, the collector efficiency is estimated as 35%. Patnode et al. (2006) developed a model for a solar collector driven Rankine cycle using TRNSYS simulation program. Both the PTSC and Rankine cycle models were validated with the measured temperature and flow rate data of the SEGS VI plant for the years between 1998 and These models were used to evaluate the effects of solar field collector degradation, flow rate control strategies, and alternative condenser design on the plant performance. A heat exchanger is a device that is used to transfer heat between two or more fluids. Liquids flow in separate plates or tube surfaces and they do not leak. Heat exchangers are used in many different 198 engineering applications such as air conditioning, space heating, waste heat recovery, and geothermal power plant systems. Heat exchangers are typically classified according to the flow arrangement (e.g. parallel flow, counter-flow and cross-flow), number of fluids (one fluid or two fluids), and construction type (e.g. shell and tube, plate, and fin) (Kakaç et al., 2002; Shah and Sekulic, 2002). Shell and tube heat exchangers are preferred for space heating, power production, and chemical processing applications. The main advantages of this heat exchanger type over other types can be listed as follows: There is substantial flexibility regarding their materials to accommodate corrosion and other concerns. There is substantial flexibility regarding their materials to accommodate corrosion and other concerns; they can be used in systems with higher operating temperatures and pressures; and tube leaks are easily located and plugged since pressure test is comparatively easy (Perry et al., 1997; Subramanian, 2010a; Subramanian, 2010b). However, this heat exchanger requires more space and cleaning and maintenance is difficult since a tube requires enough clearance at one end to remove the tube nest. Shell and tube heat exchangers are classified and built according to the widely used Tubular Exchange Manufactured Association (TEMA) standards (Harrison, 2007). This type of heat exchangers differ according to the number of tube passes. The simplest form is one tube pass as shown in Figure 2a. Two tube passes and three tube passes configurations are shown in Figure 2b and Figure 2c, respectively. Baffles are often installed to increase heat transfer coefficient of the shell side fluid. Additionally, baffles are fixed to the tubes to reduce tube vibration. In TEMA standards, these heat exchangers are classified according to front end, shell types, and rear end head types. Some common types of front ends are Type A (head channel and removable cover) and Type B (integral cover), shell types are Type E (one shell pass) and Type F (two shell passes), rear end head types are Type U (U tube bundle) and Type M fixed tubesheet like B stationary head (Harrison, 2007; Perry et al., 1997). Fig. 2: Types of shell and tube heat exchangers (a) one tube pass, (b) two tube passes, and (c) three tube passes The thermal design of shell and tube heat exchangers is done according to the principles of thermodynamics, heat transfer, and fluid dynamics. As a result of this

213 design, shell types, flow arrangement, geometry of the heat exchanger, and tube and shell materials are determined for the specified heat transfer. The mass flow rate of shell and tube side fluids, inlet temperatures, and outlet temperature of one of the fluids are generally used as input parameters in this type of design problem. There are some studies on the design, modeling, and optimization of the shell and tube heat exchangers in the literature. For example, Kara et. al. (2004) created 240 alternative exchanger configurations and the computer program that they developed selects the optimum configuration among the all possible exchanger configurations. The shell diameter, baffle spacing, number of passes are the parameters that can be changed in this program. The program then determines the overall dimensions of the shell and the optimum heat transfer surface area required to meet the specified heat transfer duty by calculating minimum or allowable shell-side pressure drop. The results of their study showed that triangular tube pitch layout with one or two tube passes yields the best performance. Reppich et al. (1995) developed a computer based design model to determine the optimum dimensions of segmentally baffled shell and tube heat exchangers by calculating optimum shell side and tube side pressure drops from the equations provided in his work. The six optimized dimensional parameters are number of tubes, tube length, shell diameter, number of baffles, baffle spacing, and baffle cut. The proposed model also includes a cost analysis. Selbas et al. (2006) created a mathematical model of the heat exchanger, which was coded and solved in MATLAB environment. Dittus-Boetner equation and Bell and Delaware method were used to calculate the tube side heat transfer coefficient and shell side heat transfer coefficient, respectively. LMTD method was used to analyze the heat exchanger. They also performed a cost analysis to study the effect of six different variables (outer tube diameter, tube layout, number of tube passes, shell diameter, baffle spacing, and baffle cut) on the heat transfer surface area and shell side fluid velocity. In the paper by Walraven et al. (2014), the system optimization of different configurations of ORCs with both plate heat exchangers and shell-andtube heat exchangers were compared. They also created a mathematical model for integrated ORC systems, and compared the performance of the systems when different fluids are used. Their study showed that ORCs with all plate heat exchangers perform mostly better than ORCs with all shell-andtube heat exchangers. Different tube configurations were investigated in the paper by Walraven et al. (2014). Their study concluded that the 30 tube configurations should be used for the single phase heat exchangers and the 60 tube configurations for the two phase heat exchangers. Fettaka et al. (2013) developed a mathematical model of the shell and tube heat exchangers, and nine different optimization works were carried out by using different geometric parameters (tube layout, number of the tube passes, baffle spacing, baffle cut, tube to baffle diametrical clearance, tube length, tube outside diameter, tube wall thickness, and shell to baffle diameter clearance). 199 Taguchi method is a commonly used experimental design method based on statistical approaches to determine the optimum conditions for a given problem. In recent years, the use of Taguchi method in the world has increased rapidly, especially in the industry. This method was developed and pioneered by the Japanese engineer, Dr. Genichi Taguchi (Ghani et al., 2004). This technique suggests a simple and systematic way to optimize design for performance quality and cost. Two important criteria used in Taguchi design are signal to noise (S/N), which measures quality; and orthogonal arrays, which establish many design factors simultaneously. Although this method has been generally used for experimental studies, it has been shown that it could be effectively applied to mathematical modeling studies (e.g. Ghani and Choudhury, 2004; Sasmito et al., 2015; Yang and Tarng, 1998). The literature survey discussed above shows that studies that include the shell and tube heat exchanger design for PTSC systems are limited; and there is no performance or optimization study on shell and tube heat exchangers that combine a geothermal brine fueled ORC and a PTSC. In this study, a shell and tube heat exchanger that combines a PTSC and ORC systems was designed. For this purpose thermal models for the PSTC and heat exchanger were first developed. Then, a parametric study was conducted to study the effect of important design parameters (e.g. outer diameter, tube length, baffle spacing, number of passes, and tube and shell side fluids) on the output parameters (e.g. the overall heat transfer coefficient and the total pressure drop). In addition, an optimization study was carried out using Taguchi method to find the optimum design parameters that maximizes the performance and minimizes the cost. In this method, six key parameters were selected such as tube diameter, tube length, baffle spacing, number of passes, shell side fluid, and tube side fluid. Three values for each parameters were taken for the first stage of Taguchi method. Then according to the results found, a further refinement of the optimum parameters was done using second stage Taguchi analysis. In addition, the effect of solar irradiation intensity on the optimum parameters was assessed. II. Mathematical Modeling As the objective of this study is to design and optimize a shell and tube heat exchanger that combines a PTSC and an ORC system under different solar irradiation intensity, mathematical models for both PTSC and heat exchanger are developed. The schematic of the integrated system is shown in Figure 3. The modeling approach and equations for both of these components are given in detail in the following subsections.

214 Fig. 4: Geometric properties of the PTSC Fig. 3: The schematic of the integrated system consisting of a PTSC, an ORC, and a shell and tube heat exchanger. II.1. Mathematical Modelling of PTSC In this section, the modeling approach and equations for the PTSC are given. The aim of the PTSC model developed is to find the exit temperature of the PTSC system for a given set of input parameters consisting of the design and operating parameters of PTSC as well as the meteorological data. Modeling of PTSC discussed in this section is based on the approach and equations given in the Refs. (Duffie and Beckman, 2013; Kalogirou, 2009). The rate of useful energy delivered by a single collector can be found using the Eq. 1. Q u = F R (SA a A r U L (T i T a )) (1) Where F R is the heat removal factor, S is the heat absorbed by the receiver, Aa is the aperture area, Ar is the receiver area, UL is the solar collector overall heat loss coefficient, Ti is the entering fluid temperature, and Ta is the ambient temperature. Then, the heat absorbed by the receiver is defined as: S = G b η r (2) Where G b is the direct irradiation intensity and η r is the receiver efficiency, which can be found using Eq. (3). η r = ργταk (3) Where ρ is the reflectance of the mirror, γ is the intercept factor, τ is the transmittance of the glass cover, α is the absorptance of the receiver, and K is the incidence angle modifier. Geometric properties of the PTSC are shown in Fig. 4. Where m is the mass flow rate of the collector fluid and c p is the heat capacity of the collector fluid. F is the collector efficiency factor defined as: F = 1 U L 1 + D o +( D o U L h fi D i 2k +lnd o ) D i (5) Where D i and D o are the inner and outer diameters of the receiver, respectively. These diameters and the diameter of the glass cover (Dg) are shown in Figure 4. In this equation, k is thermal conductivity of the receiver tube. h fi is the heat transfer coefficient inside the receiver tube, which can be calculated using Eq. (6). h fi = Nu k fi D i (6) Where Nu is the Nusselt number of the fluid flowing through the receiver. If the flow inside the receiver tube is turbulent (Re>2300), Nusselt number can be evaluated using the following correlation (Duffie and Beckman, 2013): Nu = 0.023(Re 0.8 )(Pr 0.4 ) (7) Where Re is Reynolds number of the flow inside the receiver, Pr is the Prandlt number of the collector fluid. If the flow is laminar, (Re<2300) Nusselt number is taken as constant as The solar collector heat loss coefficient is defined as: U L = ( 1 A r ) (h c,g a +h r,g a )A g (8) Where A r is receiver area, A g is glass cover area. A r and A g can be found as: A r = πd o L PTSC (9) A g = πd g L PTSC (10) The heat removal factor is given by: F R = m h c p (1 exp ( U L F A r )) (4) A r U L c p m 200 Fig. 5: Schematics of receiver tube and its diameters.

215 Aperture area, length of solar collector and width of the PTSC are shown in Fig. 6. E b = εσt s 4 (17) Where ε is a radiative property of the surface termed the emissivity. The net rate of radiation heat transfer from the surface is: " q rad = εσ(t 4 s T 4 sur ) (18) Where T sur (T sur = T a ) is surrounding temperature. Then, the net radiation heat transfer is in following the form: " q rad = h r (T s T sur ) (19) Then, from Eqs. (20) and (21), the radiation heat transfer coefficient, h r,g a, for the glass cover to the ambient is calculated as follows. Fig. 6: Design parameters of the PTSC The convection heat transfer coefficient, h c,g a, for the glass cover to the ambient can be calculated as: h c,g a = Nu k D g (11) Where Nu is the Nusselt number of air and k is the thermal conductivity of air. Before calculating the Nusselt number, temperature of glass cover, Tg, is assumed to be close to the ambient temperature. Then, the average temperature ( T ave ) is found between the ambient temperature and glass cover temperature. T ave = T a+t g 2 (12) Then, Reynold number (Re) is calculated according to the average temperature. εσ(t g 4 T a 4 ) = h r (T g T a ) (20) h r,g a = ε g σ(t g + T a )(T g 2 + T a 2 ) (21) Where ε g is the glass cover emittance and σ is the Stephan-Boltzman constant. T g is the glass cover temperature and T a is the ambient temperature. These temperatures are shown in Fig. 7. The net radiation exchange between two black surfaces is given by Q 12 = A 1 F 12 σ(t 1 4 T 2 4 ) A 2 F 21 (T 1 4 T 2 4 ) (22) Where T 1 and T 2 are the temperatures of surfaces associated with the surfaces A 1 and A 2. The term 1 = 1 A 1 F 12 A 2 F 21 represents the resistance due to the geometric configuration of the two surfaces. If surface is a real surface, ε = α, and ρ = 1 α = 1 ε. Re = ρ V D g μ (13) Where ρ is the density of air, V is the wind velocity, D g is glass cover outer diameter, and μ is the kinematic viscosity of the air. Then, Nusselt number is calculated as follows. Nu = (Re) 0.52 (0.1 < Re < 1000 (14) Nu = 0.3(Re) 0.6 (1000 < Re < 50000) (15) The maximum emissive power at a given temperature is the blackbody emissive power (E b ). Emissive power is prescribed by the Stefan-Boltzmann law. E b = σt s 4 (16) Where T s (T s = T g ) is the surface temperature and σ is the Stefan-Boltzmann constant. A black body is an ideal emitter. But the energy emitted by a real surface is less than a black body at the same temperature and is defined as follows. 201 Fig. 7: Different temperature types in the receiver tube The general form of radiation heat transfer coefficient, h r, between the receiver tube and the glass cover is defined as: h r = σ(t 1+T 2 )(T 2 1 +T2 2 ) 1 ε1 +A 1 A2 ( 1 (23) ε2 1) Where T 1 is the receiver temperature, T 2 is the glass

216 cover temperature, ε 1 is the receiver emittance, ε 2 is the glass cover emittance, A 1 is the aperture area, and A 2 is the glass cover area. For our case, the subscript 1 and 2 are replaced by g, and r. h r,r g can be calculated as follows. h r,r g = σ(t g+t r )(T g 2 +Tr 2 ) 1 εr +A r Ag ( 1 εg 1) (24) Finally, since U L is based on the assumed T g value, we need to check if the assumption made was correct. Therefore, T g can be obtained from energy balance: T g = A rh r,r g T r + A g (h r,g a +h c,g a )T a A r h r,r g +A g (h r,g a +h c,g a ) (25) To find the exit temperature of the PTSC (or the inlet temperature of the heat exchanger for the shell side), an energy balance around the PTSC should be applied, as shown in Eq. (26). T h,i = T i,oil + Q u (26) m PTSC c p,oil II.2. Mathematical Modeling of a Shell and Tube Heat Exchanger The modeling of the heat exchanger is done using the LMTD method. Using this model, a design problem is formed to calculate the overall heat transfer coefficient, the heat transfer surface, the pressure drop across the heat exchanger and the pumping power. The following assumptions are made in this model: The heat exchanger runs under steady-state conditions. Heat losses to the surroundings are negligible. The temperature of each fluid at the inlet and exit is uniform across the cross-sectional area of the shell and tube. Changes in the kinetic and potential energies of the flowing streams from inlet to exit can be neglected. The temperature change of the fluid between the inlet to exit can be considered negligible. The temperature of the ORC side fluid can be found applying an energy balance around a control volume enclosing the heat exchanger as shown in Eq. (27). material conductivity, and R" fi and R" fo are the tube side and shell side fouling factors, respectively. Fouling factor is found according to the fluid types shown in Table 1: Table 1. Fouling Factors for different fluid types (Harrison, 2007) Fluid R" f (m 2 K/W) Sea water and treated boiler feedwater (below 50 C) Sea water and treated boiler feedwater (above 50 C) River water (below 50 C) Fuel oil Refrigerant liquids Steam (non-oil bearing) Firstly, velocity of the fluid in the inner tube is found for calculating h i (heat transfer coefficient of the tube side). V i = 4 m N π d i 2 ρ T npass (29) m is the mass flow rate of tube side fluid, N T is the number of tubes and n pass is the number of passes. Then the Reynolds number is calculated using the following equation. Re = ρ V i D i μ (30) According to Eq. (30), flow regime is identified as either laminar or turbulent. According to the identified regime, Nusselt number is calculated using the equations given in Table 2. Table 2: Nusselt number for laminar and turbulent flow in a circular tube (Kakaç, et al., 2002) Regime Equation Condition Laminar Nu = 4.36 Re 2300 f (Re 1000)Pr Nu = 8 Turbulent ( f 1 8 ) < Re (Pr 3 1) < 5x10 4 Where f = (0.79ln (Re) 1.64) 2 After finding the velocity and the Nusselt number for the inner tube as discussed above, heat transfer coefficient h i can be calculated using Eq. (31). h c,o = h c,i + Q m c (27) h i = Nu k d i (31) To perform the heat transfer analysis of an heat exchanger, the major parameters are the heat transfer rate q, heat transfer surface area A, overall heat transfer coefficient U, and cold and hot fluid inlet and outlet temperatures. Overall heat transfer coefficient is calculated as follows: 1 = D o 1 + ln(d o D i ) U o D i h i 2πkL + 1 h o + D o D i R" fi + R" fo (28) Where D i and D o are the inner and outer diameters of the tube, respectively; h i and h o are the tube side and shell side heat transfer coefficients, k is the tube 202 For calculating the shell side heat transfer coefficient (h o ), equivalent diameter should be identified firstly. If the pipe alignment is triangular (Kakaç et al., 2002): D e = 1.27 d o (P t d o 2 ) (32) If pipe alignment is square (Kakaç et al., 2002): D e = 1.10 d o (P t d o 2 ) (33)

217 Where P t is the tube pitch and this parameter changes according to the tube diameter (Harrison, 2007). Cross sectional area of the shell perpendicular to the flow direction A s, can be calculated as follows. A s = (P t d o ) e D s P t (34) Where D s is the shell diameter, e is the baffle spacing. Shell diameter is calculated using Eq. (35) D s = CL 1 d o 2 CTP [A(PR)2 ] L (35) CL and CTP depend on the geometry of heat exchanger as shown in Table 3. Table 3: CL and CTP Values (Kakaç et al., 2002) Tube Layout Angle CL For 45 o and 90 o 1 For 30 o and 60 o 0.85 Number of Passes CTP One Pass 0.93 Two Passes 0.9 Three Passes 0.85 Velocity of the shell side fluid is can be found as follows m V o = (36) ρ A s Reynolds number of the shell side is: Re = ρ V o D e μ (37) To calculate the heat transfer coefficient of the shell side, ho, Nusselt number should be first calculated using the equation developed by McAdams, which is shown in Eq. (38) (Kakaç et al., 2002) Nu = h od e k = 0.36 Re 0.55 Pr 1 3 ( μ b μ w ) 0,14 (38) Where μ b and μ w are the kinematic viscosities of shell fluid at bulk temperature and wall temperature, respectively. Wall temperature is defined as: T wall = T b,h+t b,c 2 (39) Where T b,h, and T b,c are the average temperatures of hot and cold fluids between the inlet and exit. In this study, for modeling and analysis of shell and tube heat exchanger, Logarithmic Mean Temperature Difference Method (LMTD) is used. For a counter flow type heat exchanger, the logarithmic mean temperature difference is calculated as follows. T lm = (T h,i T c,o ) (T h,o T c,i ) ln (T h,i T c,o) (T h,o T c,i ) (40) Where T h,i and T h,o are the inlet and outlet temperatures of hot and cold fluid; T c,i and T c,o are inlet and outlet temperatures of cold fluid respectively. The effective mean temperature difference: T m = F T lm (41) Where F is the correction factor for the heat exchanger and the figures to find this factor can be found in Ref. (Kakac et al., 2002). In this figures, F is a function of P and R, which can be calculated using Eqs. (42) and (43). P = (T c,o T c,i ) (T h,i T c,i ) R = (T h,i T h,o ) (T c,o T c,i ) (42) (43) Finally, heat transfer rate is calculated using Eq. (44). q = UA T lm (44) Where A is the heat transfer surface area. This term can be calculated as follows. A = πd o LN T (45) Where L is the tube length and N T is the number of tubes. Pressure drop across the heat exchanger and pumping power are important parameters in the design of a heat exchanger. The pressure drop in the shell side of a heat exchanger can be calculated as follows (Shah and Sekulic, 2002). P shell = f 2 fric G shell (N b +1) D shell (46) 2 ρ D e φ shell Where G shell is the mass velocity of shell side fluid, N b is the number of baffles, f fric is the friction factor and φ shell is the viscosity ratio between bulk and wall temperatures. These parameters are calculated as follows. f fric = exp ( ln(re)) (47) G shell = m A s (48) φ s = ( μ b μ wall ) 0.14 (49) N b = ( L + 1) (50) e To calculate the pressure drop across the tubes of the PTSC, head loss due to frictional losses in the pipe is first calculated using Eq. (51). h l = f L d V 2 2g (51) 203

218 f is the friction factor, which can be found from the Moody chart (White, 2009). The pressure change across the PTSC tubes for a single row can be found as follows. P PTSC = P o P i = h l ρg (52) The total pressure including both the PTSC and the shell side of the heat exchanger can be found as shown below. Finally, the pumping power can be calculated using Eq. (53). W pump = m vδp total η pump (53) Where η pump is the pump efficiency. ΔP total can be calculated using Eq. (54): ΔP total = P PTSC + P shell (54) III. Results and Discussion In this section, the results and discussion on the parametric studies are presented. The parametric studies include the effect of outer diameter of the heat exchanger tube, tube length, baffle spacing, number of passes, and tube layout angle on the performance parameters including number of tubes required, overall heat transfer coefficient, and total pressure drop. It was considered that in the solar farm, there are 5 rows, and there are 8 modules of PTSC in each row. The type of the collector selected is SkyTrough collector (SkyFuel, 2009). The geometric data of PTSC is given in Table 4. Table 4: Geometric data of SkyTrough PTSC (SkyFuel, 2009) Single collectors width 6 m Single collectors length 14 m Receiver inner diameter m Receiver outer diameter m Glass cover diameter m Transmissivity of the receiver 0.94 Absorptivity of the receiver 0.97 Reflectivity of the aperture surface 0.96 Intercept angle 1º Receiver emittance 0.92 Glass cover emittance 0.87 Intercept factor 0.93 In this study, some assumptions were made as follows: The receiver temperature is taken T r = 300 C Entering temperature of heat transfer fluid is taken T i = 290 C The dead state properties are taken as T a = 25 C and P a = kpa As the baseline condition for the simulations, one shell and one tube pass type heat exchanger is used in the system shown in Fig. 2. The input parameters of the heat exchanger model for the baseline conditions are 204 given in Table 5. The parametric studies were conducted considering the minimum and maximum values for outer diameter of the tube, tube length, and baffle spacing considering the standards given in TEMA standards (Harrison, 2007). These values are shown in Table 6. Number of passes is taken as 1, 2, or 3; whereas tube layout angle is taken as 30, 45, 60, or 90. Table 5: Input parameters of the heat exchanger model for the baseline conditions Name of the parameter Value PTSC (Shell) side fluid Therminol VP1 ORC (Tube) side fluid R134a Number of passes 1 Tube length 12 m Inner diameter of the tube m Wall thickness m Tube pitch m Tube layout angle 90 Baffle spacing 1.5 m Thermal conductivity of tube 63.9 W/m K Mass flow rate of cold fluid kg/s Mass flow rate of hot fluid 27.5 kg/s Hot fluid pressure kpa Cold fluid pressure 4370 kpa Outlet temperature of the hot fluid 563 K Inlet temperature of the cold fluid K Table 6: The values of the design parameters used in the parametric study Outer Diameter of the Tube (m) Tube Length (m) Baffle Spacing (m) Minimum Maximum III.1 The effect of outer diameter of the tube on the performance of the system The effect of outer diameter of the tube on the performance of the system was assessed using the data given in Table 5. The different values for this diameter were taken from TEMA standards (Harrison 2007). In TEMA standards, tube diameters have 10 values, which are m, m, m, m, m, m, m, m, m, and m. Using the code developed in EES, a parametric study was conducted for these values for the outer diameter of the tube. As a result of this study, the change of overall heat transfer coefficient, and pressure drop with respect to this diameter was found. The parametric study was repeated for different number of passes (1, 2 and 3) and layout angles (30, 45, 60, and 90 ), and the results for these studies are shown in Figures 7 and 8, respectively. Fig. 8a and 8b show the change of overall heat transfer coefficient, and total pressure drop for different values of outer diameter of the tube and the number of passes, respectively. The results show that increasing the outer diameter of the tube, the overall heat transfer coefficient fluctuate. It can be seen that taking this parameter as small as possible ( m), the overall heat transfer coefficient required get its maximum value. The reason of this trend can be attributed to change the flow regime. On the other

219 hand, as it can be seen from Figure 8b, the effect of this parameters on the total pressure drop is more significant when the outer diameter of the tube is less than m. When this diameter is m, the total pressure drop is at its maximum values. This finding can be explained as follows. As the outer diameter increases, the equivalent diameter (Eq. 32) increases and thus pressure drop decreases. These figures also show that the number of passes does not have a significant effect on the results but it only changes the number of tubes required. Taking the number of passes as 1 yield slightly lower overall heat transfer coefficient compared to a heat exchanger with 2 or 3 tube passes. (b) Fig. 9: The effect outer diameter of tube on the (a) overall heat transfer coefficient, (b) total pressure drop for different tube layout angles III.2 The effect of tube length on the performance of the system (a) Ten different values of tube lengths can be found in TEMA standards. These values are m, m, m, m, m, 7.32 m, 8.53 m, 9.75 m, 10.7 m, and m (Harrison, 2007). Fig. 10a and 10b show the change of the overall heat transfer coefficient and the total pressure drop for different values of tube length respectively. The results show that increasing the tube length, the overall heat transfer coefficient increases. This trend may be due to the change in the flow from laminar to turbulent when the tube length increases; and thus the tube side heat transfer coefficient increases. When the tube length is taken as high as possible (11.58 m), the overall heat transfer coefficient gets it maximum value. (b) Fig. 8: The effect outer diameter of tube on the (a) overall heat transfer coefficient, (b) total pressure drop for different number of passes. Fig. 9a and 9b show the change of overall heat transfer coefficient, total pressure drop for different values of outer diameter of the tube and the tube layout. The results show that increasing the outer diameter overall heat transfer coefficient, and total pressure drop fluctuate. If tube layout angle is selected 30 or 60 instead of 45 or 90, both overall heat transfer coefficient and total pressure drop increase. (a) (b) Fig. 10. The effect of the tube length on the (a) overall heat transfer coefficient, (b) total pressure drop for different number of passes. (a) Figs. 11a and 11b show the overall heat transfer coefficient and total pressure drop for different values 205

220 of tube length and layout angle, respectively. These figures present that when the tube layout angle is 30 or 60, the overall heat transfer coefficient, the total pressure drop are higher. The tube layout angle depends on the tube layout constant. If tube layout angle is 30 or 60, the value of CL and shell diameter are lower than the values of those when the tube layout angle is 45 or 90. Thus, shell side fluid velocity, overall heat transfer coefficient, total pressure drop slightly increase. where the baffle spacing is 0.32 m and then this parameters almost remain constant with a further increase in the baffle spacing. This result is mainly due to the change in the flow regime. When the baffle spacing changes between 0.08 m to 1.5 m, the Reynolds number decreases from to 72189; hence the flow regime gets closer to laminar flow when we increase the baffle spacing. (a) (a) (b) Fig. 11. The effect of the tube length on the (a) overall heat transfer coefficient, (b) total pressure drop for different tube layout angles. III.3 The effect of baffle spacing on the performance of the system Minimum and maximum baffle spacing values are shown in Table 6. Minimum baffle spacing is found using the TEMA standards (Harrison, 2007); whereas the maximum baffle spacing is taken as 29.5d o 0.75 where d o is in meters (Fettaka et al., 2013). (b) Fig. 12. The effect of the baffle spacing on the (a) overall heat transfer coefficient, and (b) pressure drop for different number of passes. Figs. 13a and 13b show the change of overall heat transfer coefficient, and total pressure drop for different values of baffle spacing and tube layout angle, respectively. If tube layout angle is 45 and 90, overall heat transfer coefficient decreases. It can also be seen from the Figs. 13b that the effect of tube layout angle on the pressure drop is not significant. Figs. 12a and 12b show the change of overall heat transfer coefficient, and total pressure drop for different values of baffle spacing and number of passes, respectively. The results show that increasing the baffle spacing decreases the overall heat transfer coefficient. This trend may be due to the fact that as the baffle spacing decreases, the effect of the turbulence and thus the shell side heat transfer coefficient decrease. When the baffle spacing is 0.08 m, overall heat transfer coefficient is at its maximum value (1040 W/m 2 K). On the other hand, if the baffle spacing is 1.5 m, overall heat transfer coefficient takes its minimum value (115 W/ m 2 K). These results show that the baffle spacing should be selected as low as possible (0.08 m). The effect of number of passes on the results seems to be negligible. Fig. 12a and 12b shows that increasing the baffle spacing first sharply decreases the total pressure drop up to the point (a) (b) Fig. 13. The effect of the baffle spacing on the (a) heat transfer surface area (b) pressure drop for different tube layout angles. IV. Conclusions In this study, the design of a shell and tube heat 206

221 exchanger, which combines a PTSC and an ORC, was done by applying the principles of thermal sciences. For this purpose, thermal models of the PTSC and heat exchanger were first developed and then solved using Engineering Equation Solver for a case study. Parametric studies were conducted to find the effect of some of the key design parameters on the output parameters of the model such as overall heat transfer coefficient and the total pressure drop. The main conclusions derived from the parametric studies are listed below. Outer tube diameter does not have significant effect on the pressure drop. However, increasing the diameter decreases the overall heat transfer coefficient. When the tube length increases, the overall heat transfer coefficient increases; but the pressure drop increases. If baffle spacing decreases, Reynolds number of shell side increases and the flow regime might become turbulent. Thus, the overall heat transfer coefficient increases as the spacing decreases. If baffle spacing is taken between 0.08 m and 1.5 m, the overall heat transfer coefficient decreases from 1040 W/m 2 K to 115 W/m 2 K. The parametric studies showed that if tube layout angle is chosen 30 or 60 instead of 45 or 90 both the overall heat transfer coefficient and total pressure drop increase. The number of passes does not have a significant effect on the overall heat transfer coefficient and the pressure drop. Acknowledgements The authors would like to thank Mehmet Akif Ezan and Gokhan Fidan for their guidance in some parts of the modelling of the heat exchanger. Nomenclature As cross sectional area of the shell perpendicular to the flow direction (m 2 ) A area (m 2 ), heat transfer surface area (m 2 ) b collector width (m) cp specific heat capacity (kj/kg K) CL tube layout constant CTP tube count constant D diameter (m) Ds shell diameter (m) e baffle spacing (m) F correction factor F collector efficiency factor ffric friction factor Fr heat removal factor g gravitational acceleration (m/s 2 ) Gb direct irradiation intensity (kw h/m 2 ) Gshell mass velocity of shell side fluid (kg m/s) h heat transfer coefficient (W/m 2 K), enthalpy (kj/kg) hc,g-a convective heat transfer coefficient for the glass cover to the ambient (W/m 2 K) hfi heat transfer coefficient of inside receiver tube (W/m 2 K) 207 hi tube side heat transfer coefficient (W/m 2 K) hloss head loss (m) ho shell side heat transfer coefficient (W/m 2 K) hr,g-a radiative heat transfer coefficient for the glass cover to the ambient (W/m 2 K) K incidence angle modifier k thermal conductivity (W/m K) L single collector length (m), length of a heat exchanger (m) m mass flow rate (kg/s) Nb number of baffles npass number of passes Nt Number of tubes Pr Prandtl number Pt tube pitch (m) Q heat transfer rate (W) q reflectivity Rf " fouling factor (m 2 K/W) Re Reynolds number S heat absorbed by the receiver (W/m 2 ) T temperature (K) U overall heat transfer coefficient (W/m 2 K) UL heat loss coefficient for the solar collector (W/m 2 K) Tlm logarithmic mean temperature difference (K) Greek Symbol α absorptivity γ specific weight (N/m 3 ) ε emissivity ηpump isentropic efficiency of the pump μ viscosity φ viscosity ratio between bulk and wall temperatures ρ reflectance of the mirror, density (kg/m 3 ) σ Stefan Boltzmann constant (W/m 2 K4 ) ϑ specific volume (m 3 /kg) transmittance of the glass cover τ o Subscript a ambient, aperture b bulk c cold e equivalent fric friction g glass cover h hot i inner, inlet o outer, outlet r receiver s surface sur surroundings w wall References Brooks, M., Mills, I., Harms T., Design, Construction and Testing of a Parabolic Trough Solar Collector for a Developing-Country Application, In Proceedings of the ISES Solar World Congress, Orlando, FL, (2005). Brooks, M. J., Mills, I., Harms, T.M., Performance of a Parabolic Trough Solar Collector. Journal of Energy in Southern Africa 17(3): 71 80, (2005).

222 Duffie, J., Beckman, W., Solar Engineering Of Thermal Processes, Wiley, (2013). Fernández-García, E., Valenzuela, Z.L., Pérez, M., Parabolic-Trough Solar Collectors and Their Applications. Renewable and Sustainable Energy Reviews 14(7): , (2010). Fettaka, S., Thibault, J., Gupta Y., Design of Shelland-Tube Heat Exchangers Using Multiobjective Optimization, International Journal of Heat and Mass Transfer 60: , (2013). Ghani, J.A, Choudhury, I.A., Hassan H.H., Application of Taguchi Method in the Optimization of End Milling Parameters. Journal of Materials Processing Technology 145(1): 84 92, (2004). Harrison, J., Standards of the Tubular Exchangers Manufacturers Association. Eight Ed., New York: Tubular Exchangers Manufacturers Association, (2007). Herrmann, U., Kelly, B., Price H., Two-Tank Molten Salt Storage for Parabolic Trough Solar Power Plants, Energy 29(5-6): , (2004). Islam, M.K., Hasanuzzaman, M., Rahim, N.A., Modelling and Analysis of the Effect of Different Parameters on a Parabolic-Trough Concentrating Solar System, RSC Adv. 5(46): (2015). Kakaç, Sadik, Liu, H. Pramuanjaroenkij, A. Heat Exchangers: Selection, Rating, and Thermal Design. Second Ed., CRC Press, (2002). Kalogirou, S., Parabolic Trough Collector System for Low Temperature Steam Generation: Design and Performance Characteristics. Applied Energy 55(1): 1 19, (1996). Kalogirou S., Solar Thermal Collectors and Applications, Progress in Energy and Combustion Science, 30(3), , (2004). Kalogirou S., Solar Energy Engineering Processes and Systems. Elsevier Inc., (2009). Kara Y. A., Ozbilen G., A Computer Program for Designing of Shell-and-Tube Heat Exchangers. Applied Thermal Engineering 24(13): (2004). Collectors. Solar Energy 62(6): , (1998). Patnode A., Simulation and Performance Evaluation of Parabolic Trough Solar Power Plants. Ph.D Thesis, University of Wisconsin-Madison, (2006) Perry, R.H., Green D.W., Maloney J.O., Chemical Engineers Handbook, Seventh Ed., (1997). Ming Q., Archer David H., Sophie V. M.,. A Linear Parabolic Trough Solar Collector Performance Model. Renewable Energy Resources and a Greener Future 8(3), (2006). Reppich M., Zagermann S., A New Design Method for Segmentally Baffled Heat Exchangers. Computers & Chemical Engineering 19(95): , (1995). Sasmito, A. P., Jundika C. K., Shamim T., Arun S. M., Optimization of an Open-Cathode Polymer Electrolyte Fuel Cells Stack Utilizing Taguchi Method, Applied Energy. (2015). Selbaş, R., Kizilkan O., Reppich M., A New Design Approach for Shell-and-Tube Heat Exchangers Using Genetic Algorithms from Economic Point of View. Chemical Engineering and Processing: Process Intensification 45(4): (2006). Shah, R.K., Sekulic D.P., Fundamentals of Heat Exchanger Design, Wiley (2002). SkyTrough Next-Generation Solar Parabolic Trough Technology. 4. Link: ghbrochure.pdf. (Last Access: ) Subramanian, R S., How to Design a Shell-and-Tube Heat Exchanger.(2010a) Link:http://web2.clarkson.edu/projects/subramanian/ ch302/notes/designshelltube.pdf (Last Access: ). Subramanian, R. S., Shell-and-Tube Heat Exchangers. (2010b) Link:http://web2.clarkson.edu/projects/subramanian/ ch302/notes/shelltube.pdf (Last Access: ). Walraven, D., Laenen, B., William D., Comparison of Shell-and-Tube with Plate Heat Exchangers for the Use in Low-Temperature Organic Rankine Cycles. Energy Conversion and Management 87: (2014). Mills D. Advances in Solar Thermal Electricity Technology, Solar Energy 76(1-3): (2004). Nuwayhid R. Mrad Y.F., Abu-Said R. The Realization of a Simple Solar Tracking Concentrator for University Research Applications, Renewable Energy 24: (2001) Odeh S.D., Morrison G.L., Behnia, M.. Modelling of Parabolic Trough Direct Steam Generation Solar 208 Walraven, D., Laenen, B., William, D., Optimum Configuration of Shell-and-Tube Heat Exchangers for the Use in Low-Temperature Organic Rankine Cycles. Energy Conversion and Management 83: , (2014). White, F. M., Fluid Mechanics, Seventh Ed., Mc Graw Hill, (2009). Yang, W.H., Tarng, Y.S., Design Optimization of

223 Cutting Parameters for Turning Operations Based on the Taguchi Method, Journal of Materials Processing Technology 84(1-3): (1998). 209

224 Thermal Regulation Enhancement of Concentrated Photovoltaic Systems Using Phase- Change Materials Mohamed Emam 1*, Mahmoud Ahmed 1, Shinichi Ookawara 1, 2 1 Egypt-Japan University of Science and Technology (E-JUST), Department of Energy Recourses Engineering, New Borg-El- Arab city, Alexandria, 21934, Egypt. 2 Tokyo Institute of Technology, Tokyo, Japan. * Abstract Concentrated photovoltaic system (CPV) is one of the most promising applications of solar energy. However, due to high concentration ratios (CR), a significant increase of its temperature occurs which reduces the conversion efficiency and increases the potential to damage the CPV system. To avoid such problems, thermal regulation of CPV system is of great importance. It can be achieved by integrating the phase change materials (PCMs) with phase transition temperature close to the CPV optimal operating temperature. One of the main obstructions for such application is how to enhance the low thermal conductivity of the PCMs in order to achieve a fast thermal dissipation response. In the present work, the insertion of metal fins inside the PCMs for improving heat transfer is proposed. To investigate the thermal performance of the proposed CPV-PCM system, a comprehensive 2-D model for CPV layers integrated with PCM is developed. This model couples a thermal model for CPV layers and thermo-fluid model that takes into account the phase-change phenomenon using enthalpy method. The model is numerically simulated at different internal fin arrangements at CR = 20. The numerical results are validated using the available experimental and numerical results. It is found that the use of fins increases the heat transfer inside the PCM and achieves a significant reduction of solar cell temperature compared with that of the system without using fins. Keywords: Concentrated photovoltaic, thermal regulation, PCM, solar cell efficiency I. Introduction Concentrated photovoltaic (CPV) systems are widely recognized as the most efficient form of Photovoltaic (PV) power generation due to its high solar energy gain with small capital cost, viz., for getting more PV power output by using less solar cell material than other conventional non-concentrated PV systems. In CPV systems, relatively inexpensive materials such as plastic lenses or mirrors are used to capture the incident solar radiation on a fairly large area and concentrate that energy onto small solar cell (Du et al. 2012). However, due to high concentration ratios (CR), a significant increase of solar cell temperature occurs which reduces the conversion efficiency and increases the potential to damage the CPV system (Ma et al. 2015). Therefore, thermal regulation of CPV systems is of great importance. Many researchers are seeking to develop an effective cooling system to mitigate the impact of excessive temperature rise in the CPV conversion efficiency by removing heat from CPV module surfaces to keep a good performance as much as possible. The effective cooling method would achieve high efficiency, long lifetime, and enhance the possibility of using concentrators. Numerous previous investigations had been carried out to incorporate phase change materials (PCM) within PV systems for thermal regulation. The PV- PCM system absorbs a considerable amount of energy as a latent heat during the phase transition from solid to liquid over a very narrow range of transition temperature which increases the electric conversion efficiency by preventing the overheating of the system during the day and releasing it during the night. In addition, PV-PCM system stands out from the conventional thermal control systems with its compactness, lightness, and high efficiency. However, thermal management of CPV systems using PCM is relatively rare. Therefore, the incorporation of PCMs with phase transition temperature near to the PV normal operating temperature of 25 C for thermal regulation of CPV systems under relatively high concentration ratios is a new contribution for thermal regulation of CPV systems with relatively high CRs. One of the main obstructions for such application is how to enhance the low thermal conductivity of the PCMs to achieve a quick thermal dissipation response, especially at relatively high CRs. The literature demonstrate that thermal performance improvements achieved using metal fins were significant since the temperature distribution in the PCM container became more uniform than systems without fins, and the PV module temperature rises were clearly restrained. A review of experimental and computational studies to improve the low thermal conductivity of PCMs that were conducted over many decades had been carried by (Fan & Khodadadi 2011). (Huang et al. 2004; Huang et al. 2006b; Huang et al. 2011) investigated the effect of fin spacing, width, and fin type on the PV-PCM system performance. It was 210

225 noticed that the insertion of fins improved the effective thermal conductivity of PCMs and enhanced the thermal performance of PV-PCM system. As the fin spacing reduced, the maximum temperature decreased, and the temperature uniformity of PV cells was achieved. (Malvi et al. 2011) used conductive fins, mesh or encapsulation inside the PCM layers to enhance PCM thermal conductivity. It was found that an increase of PCM conductivity by 10% can improve PV output by 3%. In the present work, the insertion of metal fins inside the PCMs for improving its thermal conductivity is proposed. To investigate the thermal performance of the proposed CPV-PCM system, a comprehensive 2- D model for CPV layers integrated with PCM is developed. This model couples a thermal model for CPV layers and thermo-fluid model that takes into account the phase-change phenomenon using enthalpy method, the conversion of solar incident radiations and real-time heat loss boundary conditions. The model is numerically simulated at different internal fin arrangements. It was reported (Huang et al. 2006a) that the 2-D model can accurately reproduce the prediction of the 3-D model for simple line-axis system with simple boundary conditions. II. Physical model The schematic diagram of the proposed 2-D CPV- PCM system with dimensions and boundary conditions is presented in Fig. 1. This system was introduced to study the effect of fins on the CPV-PCM system temperature control under high values of solar incident radiations. As shown in the figure, the PCM was placed between two aluminum flat plates and then attached to the rear side of the CPV module. The aluminum front/back walls of 3 mm thickness were included to achieve uniform temperature distribution over the front surface of the system. Moreover, they protected the PCM and provided a high rate of heat transfer so that the PCM absorbed heat easily from the CPV module. This was enhanced further by a series of aluminum fins with 3 mm thickness extended into the PCM from the front wall. The interior dimensions of the container were 125mm height by 200 mm depth. The upper and lower ends of the CPV- PCM system were assumed to be adiabatic, so the heat flows were assumed to be symmetrical and occurred through the perpendicular direction of the CPV-PCM cell only (Huang et al. 2004; Huang et al. 2006b; Park et al. 2014). As the solar radiation incident on the surface of the PV module, part of the solar incident radiation G(t) was absorbed at the front face of the CPV-PCM system and converted to electric power by the PV cells and another part was lost by convection and radiation to the surrounding. The rest of radiation energy was conducted through the PV cells and its mounting plate to PCM causing the system temperature rise while a small fraction was dissipated from the rear surface of the system. 211 Fig. 1: Schematic diagram of CPV-PCM system with internal fins. The main criteria for selection of a suitable PCM for a particular application are its phase transition temperature which should be near to the PV characterizing temperature of 25 C. In addition, other important parameters including high values of thermal conductivity and latent heat. Additionally, stability to cycling heat process must be taken into account for making an appropriate decision (Dhaidan et al. 2013). In the present work, the selected PCM is salt hydrate CaCl2.6H2O, and the thermo-physical properties of the PCM and aluminum are shown in Table 1. CPV- PCM systems without fins and with a different number of fins were investigated at CR = 20. Also, the effect of the fin length on the CPV thermal regulation was discussed. Table 1: Thermo-physical properties of selected PCM (Hasan et al. 2015) and aluminum. Thermo-physical properties CaCl 2-6H 2O (PCM) Aluminum Melting point, ( C) 29.8 N/A Heat of fusion, (kj/kg) 191 N/A Thermal conductivity Solid, (W/m C) Liquid, (W/m C) Density Solid, (kg/m 3 ) Liquid, (kg/m 3 ) Specific heat capacity Solid, (kj/ kg K) Liquid, (kj/ kg K) N/A 2675 N/A N/A Thermal expansion coefficient, (k -1 ) N/A Thermal cyclic stability Yes (Tyagi & Buddhi 2008) - Chemical classification Salt hydrate - N/A: Not Applicable. III. Mathematical model III.1. Governing equations In the current study, a comprehensive 2-D model for CPV layers integrated with PCM is developed to predict the transient temperature distribution within the CPV-PCM system with and without fins. This model comprises the energy equations for CPV layers and the thermo-fluid model for the transient analysis of PCM. The computational domain is shown in Fig. 2

226 which consists of the aluminum front plate, aluminum fins, PCM and the aluminum back plate. conversion efficiency and temperature coefficient at a reference temperature, T ref =25 C respectively. The reference solar radiation G(t) is equal to 1000W/m 2. These values are provided by the manufacturer data sheet and are available for most PV cells (Tiwari & Swapnil 2010). The front loss, E f of thermal energy by the effect of the wind speed and radiation can be determined as follow (Zelin Xu & Kleinstreuer 2014): E U f f U ( T Ta ) (7) f g K g sc 1 1 (8) h f Fig. 2: Computational domain of CPV-PCM system. The front wall heat flux q"w can be calculated from the energy balance equations of the CPV module where the total energy absorbed by the CPV cell can be written as follow: E gg (t) (1) sc sc sc The total energy absorbed by the tedlar can be written as: 1 ( t) E gg (2) T sc sc Then the total energy absorbed by both CPV cell and its tedlar can be written by: 1 G( t) G( t) E (3) sc T g sc According to the CPV cell efficiency, part of the total absorbed energy (E) at the front surface of the CPV- PCM system is converted into electricity E el, another part (E f) is lost from the front surface of CPV cell to the surrounding by means of convection and radiation. The remaining part is conducted through the CPV cells and its mounting plate to PCM causing the system temperature rise as shown in Fig. 2. The amount of thermal energy passes through the CPV cell to PCM can be estimated according to the following relation: q E E E (4) w el f The electric power produced by solar cell can be written as follows: E gg(t) (5) el sc sc sc g The front surface heat transfer coefficient for the glass surface of the CPV to the ambient can be calculated by (Agrawal & Tiwari 2011). h (9) f V w In order to deal with phase change problem, the enthalpy-porosity technique is used. In this technique, the solid-liquid interface is not tracked explicitly. The presence of the solid or liquid phase is instead monitored using a quantity known as a liquid fraction (λ). The PCM is assumed to be Newtonian, incompressible, and unsteady. Also, the PCM density variation in the buoyancy term is modeled by the Boussinesq approximation for involving thermal buoyancy. Thus, the governing equations include the continuity equation, momentum equations, and the energy equation for the 2-D transient laminar flow could be written as follows (Hosseini et al. 2012; Taylor 2007): u v 0 x y u u u P u u u v t x y x 2 2 x y v v v u v t x y S x P v v g y 2 2 x y (10) (11) T T ref Sy (12) With μ = μ s =, ρ = ρ s in solid regions of PCM, and μ = μ l, ρ = ρl in liquid region of PCM. S In Eqs 11 and 12, is the Darcy's law damping terms (as source term) that are added to the momentum equation due to phase change effect on convection. It is defined as: Where, η sc is the solar cell efficiency which is a function of solar cell temperature as reported in (Zelin Xu & Kleinstreuer 2014) and can be written as follows: 1 S A mush V (13) 1 ( T T )) (6) sc ref ( ref sc ref Where: η ref and β ref are the CPV cell electrical 212 Where ε is a small number, typically around 10-3 introduced to avoid the singularity and A mush is the mushy zone constant which describes how steeply

227 the velocity is reduced to zero when the material solidifies and its value depends on the morphology of the medium. The value of A mush used here for the computations is the standard one, i.e kg.m -3 s -1. Energy equation for melt: H H H u v t x y Energy equation for solid: H t H H l l (14) x x y y H H s s (15) x x y y The enthalpy of the material is computed as the sum of the sensible enthalpy, h, and the latent heat, ΔH: H h H Where, T (16) h href cpdt (17) Tref The latent heat content can be written in terms of the latent heat of the material, L: L (18) Where ΔH may vary from zero (solid) to L (liquid). Therefore, the liquid fraction, λ, can be defined as: H 0 T Tsolid L H T Tsolid Tsolid T Tliquid (19) L Tliquid Tsolid H 1 T Tliquid L III.2. Boundary conditions The computational domain is a 2-D rectangular cavity with dimensions (L H) in the x-y plane, and δ is aluminum plate thickness, so the no-slip boundary conditions are: u = v = 0, at x = δ, (L- δ) and y = 0, H for t 0 P = 0, at x = δ and y = 0 for t 0 The initial values of u, v and P, are set to be zero. At the exterior front boundary at x = 0, where the computational domain is exposed to heat flux, the boundary condition is: T q w k x x0 Thermally coupled boundary condition is applied at the interface between the aluminum front plate and 213 PCM at x =δ, T T k k x x Al, x PCM, x At the exterior back boundary at x = L, the boundary condition is: T k x x l h b T x l T a Where: h b is the heat transfer coefficient from back exterior wall to the surrounding; h b 2.8 3V (20) w For adiabatic boundary condition is applied on the upper and lower ends, the boundary conditions are as follow: T y T y 0, H 0 0, x, ytini The phase change occurs at a set temperature. In the case of constant specific heat capacities for each phase, the temperature field can be defined as: E c T T m T E L m cl 0 E L, E L, T T m T T m T T m ( Solid phase) s (21) ( melt zone) ( liquid phase) The latent heat value E of the PCM in the melt zone is modeled as high sensible heat value in each time step and accumulated with time. At any time, when the accumulated heat is larger than the specified latent heat L of PCM, the PCM is changed to the liquid phase. The values of CPV cell characteristics and design parameters used in the CPV-PCM are indicated in Table 2. The temperature of the aluminum front wall will be estimated from the numerical simulation. Then the temperature of CPV cell can be calculated by the following equation (Tiwari & Swapnil 2010). T sc T w qw k sc sc T k T (22) Table 1: Characteristics and design parameters used in the CPV-PCM. Parameter Value Parameter Value H sc 125 mm α sc τ g α T δ g 0.32 mm 1 G(t) 1000W/m 2 k g 1 2, 3 η ref δ sc 0.2 mm 1 k T k sc δ T 0.3 mm 1 β sc β ref , (Zhou et al. 2015); 2, (Hedayatizadeh et al. 2013); 3, (Zelin Xu & Kleinstreuer 2014); 4, (Z. Xu & Kleinstreuer 2014); 5, (Dubey & Tay 2013).

228 III.3.Computational procedures and Model validation. Firstly, the upper wall heat flux q"w can be calculated from the energy equation of the CPV module by guessing initial value for the electric conversion efficiency. The governing equations subjected to the boundary and initial conditions for the computational domain are solved by using the commercial software ANSYS 17.0 using the finite volume technique. The SIMPLE algorithm has been used to solve the pressure velocity coupling. The first-order upwind scheme was used for solving the momentum and energy equations, whereas the PRESTO (PREssure STaggering Option) scheme was adopted for the pressure correction equation. By solving the governing equations at each time step, liquid mass fraction has been updated using Eq. 19. First-order implicit time integration scheme has been employed. The time step in the calculations was selected as small as 0.5 s and the number of iterations for each time step was 400. The number of triangular 2-D elements is discretized the physical domain. This number of grids and time step are considered after careful examination of the results to achieve the grid independency and accommodate both the required solution accuracy and convergence at a relatively low run time. The convergence was checked at each time step, with the convergence criterion of 10 6 for all variables. The computational results are validated with the experimental data of Huang et al. (Huang et al. 2011) by comparing the average predicted and measured temperature on the front surface of the system with time as shown in Fig 3. The incident solar radiation and ambient temperature used were 750 W/m 2 and 19 C, respectively. The top, bottom, and back surfaces were adiabatic. Reasonable agreement is obtained between the current computational results and experiments of Huang et al. (Huang et al. 2011). Furthermore, comparisons of the predicted front surface temperature with the numerical results reported by Huang et al. (Huang et al. 2004) is conducted as shown in Fig 4 at an ambient temperature of 20 C, incident solar radiation of 1000 W/m 2 and an initial system temperature of 20 C. Good agreement is found between both results. Fig. 3: Comparison between the predicted average temperature evaluations on the front surface with the corresponding experimental results of (Huang et al. 2011) with the ambient temperature set at 19 C and incident radiations of 750 W.m -2. Fig.4: Comparison between the predicted average temperature evaluations on the front surface of the numerical results of (Huang et al. 2004) with the ambient temperature set at 20 C with incident radiations of 1000 W.m -2 and an initial system temperature of 20 C. 214 IV. Results and discussion IV.1. Solar cell temperature without PCM In order to investigate the performance of CPV-PCM system, a thermal model for the PV reference cell without PCM is developed for comparison. This model includes a complete energy balance on PV layers such as glass cover, polycrystalline silicon solar cell, and PV-tedlar as documented in Eqs 1-9. The computed temperature is compared with the nominal operating temperature of the polycrystalline silicon

229 (MSX- 60) solar cell given by the manufacturing data sheet at solar radiation of 800 W.m -2, wind speed of 1m/s, and ambient temperature of 20 o C is compared with the computed temperature at the same conditions. The comparison shows a very good agreement where the predicted temperature is about o C, and the given nominal temperature is 47 o C. The deviation from the nominal value is estimated to be about 1.2% (Hedayatizadeh et al. 2013). Figure 5 indicates the variation of steady state solar cell temperature at different values of CR ranged from 1 to 20 without using PCM as a cooling medium. It is clear that increasing the CR leads to an increase in the solar cell temperature where it increases from 50 to about 510 o C. This occurs due to the increasing of the incident solar energy as CR increases. Fig. 5: variation of solar cell temperature without PCM versus CR, at an ambient temperature of 25 C and 1 m.s -1 wind velocity. occurred. Therefore, the liquid-solid interface changes from parallel lines near to the front wall to curved one as the time elapses as shown in Fig. 6. Along the height at the center of the non-finned CPV- PCM system at points A, B, and C, heat transfer, is initially dominated by conduction with a linear temperature increase with time. After one hour, the melting interface reaches point A; then heat transfer is dominated by convection. Due to convection, the temperature increased sharply towards the solar cell temperature while the temperatures in the solid phase locations (B, C) maintained their slow conductiondominated increase. The same behavior is observed at point B and C after two hours and three hours respectively. In addition, two stages of temperature variation of the solar cell are observed during the phase change process of the cooling material.firstly, a steep increase in the temperature with time followed by a gradual increase until reaching the peak. This variation is most likely due to sensible heating of PCM by conduction heat transfer through the aluminum plate, followed by the phase transition of the PCM adjacent to the aluminum front plate which is causing a thin melt layer in the PCM, absorbing the CPV thermal energy as latent heat. During this period, the PCM acts as an insulation material for the CPV cell while the heat transfer is dominated by conduction rising the CPV front surface temperature. Lastly, further increase of time results in a decrease in cell temperature followed by an increase until reaching the complete melting point. This stage indicates the start of the heat transfer by convection which balances heat transfer by conduction in the PCM. With increasing the time, heat is removed by convection and conduction until reaching a fully liquid phase. IV.2. Thermal performance of the CPV-PCM system without fins. Thermal regulation of the CPV-PCM system depends on the thermal behavior of the PCM during melting. The transient variation of the average solar cell temperature of the non-finned CPV-PCM system with 200 mm PCM thickness and CR=20 is presented in Fig. 6. The same figure presents the temperature variation with elapsed time along the height at the center of the non-finned CPV-PCM system ( points A, B, and C locations, see Fig. 6). As indicated in Fig. 5, it was found that the temperature of the solar cell without PCM could reach 510 o C at CR=20. Using PCM salt hydrate CaCl2.6H2O with a melting point 29.8 o C with similar conditions, the CPV-PCM without fins could maintain the solar cell at an average temperature of 64 o C for 2.0 hours while the temperature at the complete melting point of PCM is around 119 o C. With continued energy input, the melting of PCM occurred from top to bottom with clear temperature stratification because of the natural convection 215 Fig. 6: Average predicted solar cell temperature and the temperature variation with time in the center vertical line of CPV-PCM system with the time evolution of the solid-liquid interface of CPV-PCM system at an ambient temperature of 25 C and 1 m.s -1 wind velocity.

230 IV.3. Thermal performance comparison for CPV- PCM systems with a different number of fins. In the present work, a detailed analysis of the effect of the insertion of a different number of aluminum fins on the thermal regulation of the CPV-PCM system is proposed. Figure 7 presents the transient variation of the average solar cell temperature of the CPV-PCM system with a different number of fins ( each is 75 mm length) and no fins. From this figure, it is noticed that the CPV-PCM system with two fins could maintain the solar cell at an average temperature of 57 o C for 2.0 hours while the temperature at the complete melting point of PCM is around 111 o C. For the CPV-PCM system with four fins, the temperature of the solar cell was maintained at 54 o C for 2.0 hours while the temperature at the complete melting point of PCM is around 109 o C. It is clear that the use of the fins increases the heat transfer inside the PCM and achieves a significant reduction of solar cell temperature compared with the system without fins. PCM temperature begins to rise quickly. This flow pattern is maintained until the PCM is fully molten. Once the PCM in the uppermost section is fully molten after two hours, the temperature of the CPV/PCM system increases rapidly. t = 0.5 hr t = 1.0 hr t = 1.75 hr t = 2.0 hrs Fig. 7: Average predicted solar cell temperature of CPV - PCM system with different number of fins and no fins at CR= 20, an ambient temperature of 25 C and 1 m.s -1 wind velocity The predicted temperature distributions during the PCM melt process within the CPV-PCM system with four fins during the PCM melt process is presented in Fig. 8. From the figure, it is noticed that when aluminum fins are added to the system, the formation of a deep cavity in the upper part of the CPV-PCM system is reduced and divided into several small shallow cavities between the fins which reduced the thermal stratification within the system. After one hour, a natural convection flow of hot molten PCM passes through the gap at the end of the fins into the upper part of the system then turns to flow downward through the gap and near to the liquid-solid interface into the lower section. After 1.75 hours, the molten PCM reaches the aluminum rear plate, and its temperature rises causing an increase in the heat transfer rate from this side to the PCM adjacent to it. Thereby, the melting velocity increases and the CPV- 216 t = 3 hrs Fig. 8: Predicted isotherms of CPV-PCM system with four fins at different times In order to investigate the effect of using fins on the temperature uniformity for the CPV-PCM system.

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