Advances in Energy Research, Vol. 2: Selected Papers from ICAER 2017 (Springer Proceedings in Energy) [1st ed. 2020] 9811526613, 9789811526619

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Springer Proceedings in Energy

Suneet Singh Venkatasailanathan Ramadesigan   Editors

Advances in Energy Research, Vol. 2 Selected Papers from ICAER 2017

Springer Proceedings in Energy

The series Springer Proceedings in Energy covers a broad range of multidisciplinary subjects in those research fields closely related to present and future forms of energy as a resource for human societies. Typically based on material presented at conferences, workshops and similar scientific meetings, volumes published in this series will constitute comprehensive state-of-the-art references on energy-related science and technology studies. The subjects of these conferences will fall typically within these broad categories: – – – – – – –

Energy Efficiency Fossil Fuels Nuclear Energy Policy, Economics, Management & Transport Renewable and Green Energy Systems, Storage and Harvesting Materials for Energy

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More information about this series at http://www.springer.com/series/13370

Suneet Singh Venkatasailanathan Ramadesigan •

Editors

Advances in Energy Research, Vol. 2 Selected Papers from ICAER 2017

123

Editors Suneet Singh Department of Energy Science and Engineering Indian Institute of Technology Bombay Mumbai, Maharashtra, India

Venkatasailanathan Ramadesigan Department of Energy Science and Engineering Indian Institute of Technology Bombay Mumbai, Maharashtra, India

ISSN 2352-2534 ISSN 2352-2542 (electronic) Springer Proceedings in Energy ISBN 978-981-15-2661-9 ISBN 978-981-15-2662-6 (eBook) https://doi.org/10.1007/978-981-15-2662-6 © Springer Nature Singapore Pte Ltd. 2020 This work is subject to copyright. All rights are reserved by the Publisher, whether the whole or part of the material is concerned, specifically the rights of translation, reprinting, reuse of illustrations, recitation, broadcasting, reproduction on microfilms or in any other physical way, and transmission or information storage and retrieval, electronic adaptation, computer software, or by similar or dissimilar methodology now known or hereafter developed. The use of general descriptive names, registered names, trademarks, service marks, etc. in this publication does not imply, even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and therefore free for general use. The publisher, the authors and the editors are safe to assume that the advice and information in this book are believed to be true and accurate at the date of publication. Neither the publisher nor the authors or the editors give a warranty, expressed or implied, with respect to the material contained herein or for any errors or omissions that may have been made. The publisher remains neutral with regard to jurisdictional claims in published maps and institutional affiliations. This Springer imprint is published by the registered company Springer Nature Singapore Pte Ltd. The registered company address is: 152 Beach Road, #21-01/04 Gateway East, Singapore 189721, Singapore

Preface

This proceedings contains selected papers presented in the 6th International Conference on Advances in Energy Research (ICAER 2017), which was held at IIT Bombay, Mumbai, India, from 12 to 14 December 2017. The biennial conference has been organised since 2007, for providing a common platform for the researchers in the field of energy and allied domains. The conference was inaugurated by the honourable Union Minister of Petroleum and Natural Gas and Skill Development and Entrepreneurship, Shri. Dharmendra Pradhan, and was presided over by Prof. Devang Khakkar, Director, IIT Bombay. The Department of Energy Science and Engineering (DESE) has been organising the biennial conference, which serves as an excellent forum to present new findings, exchange novel ideas, discuss new developments and finally reflect on the challenges that lie ahead in line with the vision of the department “To develop sustainable energy systems, solutions and workforce for the future”. DESE has developed several novel education programmes focussing on the application of science and engineering to problems in energy. Various aspects of energy research, including but not limited to renewable energy, energy storage, energy efficiency and modelling, energy policy and conventional energy, are covered in this conference. This conference throws light on various recent accomplishments by researchers worldwide in the areas of solar thermal, thermal storage, solar PV with new materials, novel batteries, biofuelbased transportation and rural energy needs, to name a few. More than 420 submissions were received, and a rigorous peer review process was followed for acceptance of the papers. About 150 papers were accepted for oral presentation, and around 110 papers were accepted in the poster category in the conference based on the reviews received. This proceedings is divided into two volumes. Volume 1 contains papers from topics related to solar photovoltaics, energy storage and conversion and energy efficiency and management. Volume 2 contains papers from topics related to renewable energy other than solar photovoltaics; IC engines, biofuels and other conventional energy; and power electronics and microgrids.

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Preface

We would like to take this opportunity to thank all the invited speakers, delegates, sponsors, the members of the organising and advisory committee and most importantly the students and the staff of DESE for their dedicated efforts in organising this conference. These papers represent the most recent research on the subject. The editors would like to thank all the authors and the anonymous referees for paying attention to the quality of the publications. We express our gratitude for the financial support and sponsorship from government agencies and industries— ONGC, SERB-DST, EESL, NCPRE-IITB, IMASE-IITB, Pine Instruments, BioLogic, Cummins, HHV and NPCIL. The awards were sponsored by the Royal Society of Chemistry (RSC) and Springer. The publication of the issue will surely amplify the conference outcome and generate a much larger discussion and scientific progress. The contents of this proceedings reveal the breadth of current activities in different themes related to energy. We hope they form a useful starting point for beginners as well as practitioners in this discipline. Mumbai, India December 2017

Suneet Singh Venkatasailanathan Ramadesigan (Organising Secretaries, ICAER 2017)

Contents

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Mathematical Modeling of Heat Losses from Cylindrical Cavity Receiver in Solar Parabolic Dish . . . . . . . . . . . . . . . . . . . . . . . . . . . R. Sinha and N. P. Gulhane

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Performance Evaluation of Latent Heat Storage Filled with Paraffin Wax for Solar Thermal Applications . . . . . . . . . . . . . D. Gudeta, S. R. Jena, P. Mahanta and P. S. Robi

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Performance Analysis of Spiral and Conical Receivers for the Paraboloidal Dish Collector Using CFD . . . . . . . . . . . . . . . Rashmi R. Joshi, Sandeep S. Joshi, Nilesh S. Wakchaure and Akshay C. Suryawanshi

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Performance Analysis of Phase Change Material Storage System for Solar Thermal Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sneha Murali, R. P. Saini and Ambuj Punia

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Exergy and Energy Analysis of a Packed Bed Thermal Energy Storage System with Different Heat Transfer Fluids . . . . . . . . . . . . Ambuj Punia, R. P. Saini and Sneha Murali

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Performance Analysis of Parabolic Trough Solar Collector with ‘U’-Tube and Helical Coil Receivers . . . . . . . . . . . . . . . . . . . . Mohd. Mubashshir Naved, Sandeep S. Joshi and Nikhil A. Bhave

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Performance Evaluation of an Improved Dual Purpose Solar Collector . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . P. P. Krishnaraj and P. Arun

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Experimental Investigation on Farmer-Friendly Hybrid Dryer for Indoor Drying of Mushroom . . . . . . . . . . . . . . . . . . . . . . . . . . . K. Sharma, S. Kothari, N. L. Panwar, N. Rathore and K. Samar

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Wind Speed Forecasting Using New Adaptive Regressive Smoothing Models . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Parikshit G. Jamdade, Prasad A. Godse, Prathamesh P. Kulkarni, Sujay R. Deole, Sudesh S. Kolekar and Shrinivas G. Jamdade

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10 Thermal Performance Analysis of a Heat Pump-Based Photovoltaic/Thermal System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103 S. Vaishak and Purnanand V. Bhale 11 Overall Performance of N Partially Covered Photovoltaic Thermal-Compound Parabolic Concentrator (PVT-CPC) Collector with Different Concentration Ratio . . . . . . . . . . . . . . . . . 113 Rohit Tripathi, Abhishek Tiwari and G. N. Tiwari 12 Thermo-Hydraulic Performance of Solar Air Heater Roughened with V-Shaped Ribs Combined with V-Shaped Perforated Baffles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 123 Vijay Singh Bisht, Anil Kumar Patil and Anirudh Gupta 13 Highly Efficient Solar Steam Generation Using Carbon Cloth System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 133 M. W. Higgins, A. R. Shakeelur Rahman and Neetu Jha 14 Floating Absorber Integrated with Compound Parabolic Concentrator for Effective Solar Water Desalination . . . . . . . . . . . 141 Chandan and Bala Pesala 15 Study of Performance of Solar Flat Plate Collector Using Al2O3/Water Nanofluids . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149 Pankaj Raj, Geleta Fekadu and Sudhakar Subudhi 16 Thermo-Hydraulic Performance of Solar Air Heater Duct Provided with Conical Protrusion Rib Roughnesses . . . . . . . . . . . . 159 Tabish Alam, Ashok Kumar and Nagesh B. Balam 17 Flocculation–Solar Distillation—an Integrated Energy-Efficient Technology for Desalination of Seawater . . . . . . . . . . . . . . . . . . . . 169 Devlina Das and Nilanjana Mitra 18 Macro-Encapsulation of PCM Integrated with Double-Pass Solar Air Heater System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 181 Arun K. Raj, M. Srinivas and S. Jayaraj 19 Studies on Biomass Torrefaction for Energy Densification of the Fuel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 195 Pradeep Kumar Budde and Jay Pandey 20 Experimental and Theoretical Investigation of Different Coating on the Performance of the Parabolic Trough Collector . . . . . . . . . . 205 K. H. Motwani and J. R. Patel

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21 Optically Enhanced Solar Selective and Thermally Stable Absorber Coating for Concentrated Solar Thermal Application . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 217 S. R. Atchuta, B. Mallikarjun and S. Sakthivel 22 Liquid Desiccant Dehumidification Using Solar Regenerated System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 229 Geleta Fekadu and Sudhakar Subudhi 23 Wind Flow Simulation Over a Hilly Terrain for Wind Energy Harvesting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 239 Ganesh Kumar, Ajay Gairola and Raghvendra Pratap Singh 24 Theoretical Modeling of Phase Change Material-Based Space Heating Using Solar Energy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 251 Ashwath Vaidhyanathan and N. D. Banker 25 Investigations on Improving the Efficiency of Solar Air Heater Using Extended Surfaces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 261 S. Babu, S. Senthilvel, F. Paul Gregory and T. Gopi 26 Productivity Enhancement of Passive Type Solar Still Using Copper and Aluminum Based Absorber Plate with Al2O3 NanoFluid in Water Basin . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 273 Amrit Kumar Thakur and V. P. Chandramohan 27 Design and Performance Investigation of Wind Turbine Blade for Solar Updraft Tower Under Low Wind Speeds . . . . . . . . . . . . 283 Ramakrishna Balijepalli, V. P. Chandramohan and K. Kirankumar 28 Numerical Analysis of Heat Transfer Enhancement in Artificially Roughened Solar Air Heater . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 297 Vishakha Chopade and Sharad D. Patil 29 Design and Weight Minimization of Small Wind Turbine Blade for Operation in Low-Wind Areas . . . . . . . . . . . . . . . . . . . . 311 Aarti More and Anindita Roy 30 Thermohydraulic Performance of Packed Bed Solar Air Heater . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 323 Parag Jyoti Bezbaruah, Doljit Borah, Rupam Patowary and Debendra Chandra Baruah 31 Numerical Investigation on Triangular Fin-Based Solar Air Heater . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 341 Parag Jyoti Bezbaruah, Aabir Das, Rajat Subhra Das and Bikash Kumar Sarkar

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32 Comparative Performance Assessment of a Solar Hybrid Dryer with Traditional Drying Techniques . . . . . . . . . . . . . . . . . . . . . . . . 351 Bhaskar Ranjan Tamuli and Pradyumna Kumar Choudhury 33 Numerical Study of Blade Profiles of Vertical Axis Wind Turbine (VAWT) with Bidirectional Wind Flow in Highway Roads . . . . . . 361 C. ArunPrakash, P. PonsuganthIlangovan, Nitin Joy and R. Subramanian 34 CCS Combined with Geothermal Energy Generation—Hybrid Geothermal Energy Concept . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 369 Nandlal Gupta and Manvendra Vashistha 35 Effect of Preheating and Fuel Injection Pressure on Performance Parameters of Diesel Engine Running with Biodiesel . . . . . . . . . . . 379 Menelik Walle Mekonen, Niranjan Sahoo and Santosh Kumar Hotta 36 Experimental Investigation on Range of Fuel Premixing Ratio for Stable Engine Operation of Dual Fuel Engine Using Port Injection of Gasoline/Methanol and Direct Injection of Diesel . . . . 393 Mohit Raj Saxena and Rakesh Kumar Maurya 37 Used Temple Oil, a Source for Biodiesel Production . . . . . . . . . . . . 405 Sharanabasappa Saddu, Sangshetty B. Kivade and P. Ramana 38 An Experimental Study on Late PCCI Technique for Reducing NOx and Smoke Under Optimum Operating Conditions on DI Diesel Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 413 S. Parodwad Onkar and M. Sutaria Bharatkumar 39 An Assessment of Properties of Briquettes Produced from Blends of Cascabela Thevetia Seed Shell, Maize Corn Cob and Black Liquor . . . . . . . . . . . . . . . . . . . . . . . . 425 Santhosh Ujjinappa and L. K. Sreepathi 40 Experimental Investigation of In-situ Biodiesel Production from Castor Seeds (Ricinus communis) Using Combination of Microwave and Ultrasound Intensification . . . . . . . . . . . . . . . . . 435 Kartikkumar Thakkar, Keyur Shah, Pravin Kodgire and Surendra Singh Kachhwaha 41 Investigations on the Effects of Diethyl Ether as Fuel Additive in Diesel Engine Fueled with Tamarind Seed Methyl Ester . . . . . . 447 V. Dhana Raju, P. S. Kishore and R. Subbarao 42 Effect of Nitromethane–n-Butanol–Diesel Blends on Diesel Engine Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 457 Naveen Kumar Sain, Ashish Nayyar, Chandan Kumar, K. B. Rana and B. Tripathi

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43 Experimental Study on CI Engine Performance for Optimum Blending Ratio of Blended Kusum Biodiesel . . . . . . . . . . . . . . . . . . 467 A. G. Poshetti and M. S. Tandale 44 Preparation and Characterization of Biodiesel Extracted from Acidic Oil: A by-Product of Soybean Oil Refining Process . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 479 Abhijeet P. Shah, Pankaj S. Ghatage and Rahul S. Khanase 45 Thermodynamic Analysis of Diesel Engine Primed Trigeneration Configurations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 489 S. D. Bagade, M. N. Shelar and S. R. Mahajan 46 Investigations on Performance of CI Engine with Waste Palm Oil Biodiesel-Diesel Blends Using Response Surface Methodology . . . . 505 Jagannath Hirkude and Vivek Belokar 47 Design and Optimization of Air–Biogas Mixing Device for Dual Fuel Diesel Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 515 Akash Chandrabhan Chandekar and Biplab Kumar Debnath 48 Energy Response Function of Stilbene and BC501 Neutron Detection System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 529 Annesha Karmakar, S. Prasad and A. Kelkar 49 Surface Remodelling of Zeolite 4A Bodies for CO2 Capture: A Case Study . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 541 Debashis Panda, Sanjay Kumar Singh and E. Anil Kumar 50 Experimental Investigation on the Feasibility of Sugarcane Bagasse for Gasification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 551 Joel George, P. Arun and C. Muraleedharan 51 Quasi-Dimensional Thermodynamic Simulation Study of Downsizing on a Four-Cylinder Turbocharged Engine . . . . . . . . 563 Prajit Ravi, V. Devanandh, Sunil Kumar Pandey, K. Senthilnathan, Krishnan Sadagopan and Brijesh P. Patel 52 Numerical Simulation of Coal Char Gasification with CO2 in a Drop Tube Furnace . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 577 Hrusikesh Barik, Manaswita Bose, Tao Xu, Mahmud Kibria and Sankar Bhattacharya 53 Experimental Analysis of Performance and Emissions of Nanofluid Dosed Pure Neem Biodiesel (PNB)—Eucalyptus Oil (EO)-Water (W)-Surfactant (S) Emulsion Fuel on Diesel Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 587 V. W. Khond, V. M. Kriplani, S. D. Butaley, Amol Pitale and Pramod Walke

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54 A Study on Conversion of Glycerol into Solketal Using Rice Husk-Derived Catalyst . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 599 Jaspreet Kaur, Poonam Gera, M. K. Jha and Anil Kumar Sarma 55 Mathematical Model of Design and Performance Evaluation of a 210 MW CFB Boiler for Indian Lignite . . . . . . . . . . . . . . . . . . 607 S. Naga Kishore, T. Venkateswara Rao and M. L. S. Deva Kumar 56 Experience of Self-powered Neutron Detectors at TAPS-3&4 . . . . . 623 Manish Raj, Rajarshi Das, A. S. Pradhan, P. N. Prasad and A. K. Balasubramanian 57 Study of Kinetics and Reactivity Parameters of Indian Coal and Biomass Blends . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 633 Ankit Kumar, Manjula Das Ghatak, Sujan Saha and Prakash D. Chavan 58 Impact of Coal Quality on Post-combustion, Amine-Based CO2 Capture in Indian Coal Power Plants . . . . . . . . . . . . . . . . . . . . . . . 643 Pranav C. Phadke, Anand B. Rao and Munish K. Chandel 59 3D Kinetic Model for Simulation in Real Time for Full-Scope Simulator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 655 Suresh Kandpal, M. P. S. Fernando, A. S. Pradhan, P. N. Prasad and A. K. Balasubrahmanian 60 Flux Mapping System for Large PHWRs with Boiling at the Coolant Exit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 669 Abhishek Chakraborty, M. P. S. Fernando, A. S. Pradhan, P. N. Prasad and A. K. Balasubrahmaniam 61 CFD Simulation on the Effect of Hydrogen Mass Fraction and Initial Temperature in a CI Engine Under Hydrogen-Diesel Dual Fuel Mode . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 679 S. Sirajuddin and R. Manimaran 62 Multi-objective Optimization of Performance and Emissions Characteristics of CI Engine Using Cottonseed Oil as an Alternative Fuel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 689 Milind A. Pelagade, Madhavi S. Harne and Ramakant Shrivastava 63 Effect of Compression Ratio on the Performance and Emission Characteristics of a Raw Biogas Fueled Spark Ignition Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 701 Santosh Kumar Hotta, Niranjan Sahoo, K. Mohanty, P. Mahanta and A. J. Chaudhari

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64 Evaluating the Sensitivity of Biomass Feedstocks to Producer Gas Composition Using Stoichiometric Equilibrium Model . . . . . . . . . . 715 P. Pradhan, S. M. Mahajani and A. Arora 65 Estimation and Characterization of Tar from an Open-Top Downdraft Gasifier . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 725 Priyanka Tripathi, Sadhan Mahapatra and S. Dasappa 66 CO2 Capture Using Crude Glycerol-Derived Deep Eutectic Solvents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 735 R. Alok, S. S. Dawn, N. Priscilla, R. Priyanka and A. Joshua 67 NOx Reduction with Coherence of Particulate Matter for Single-Cylinder Diesel Engine Using Proportional EGR Technique . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 745 Chetan V. Bhusare and Kiran V. Chandan 68 Two-Step Modeling for Growth of Microorganisms in Stirred Tank Photobioreactor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 753 Raj Kumar Saini, Pramod P. Wangikar and Manaswita Bose 69 RBFN-Based MPPT Technique for PV System with High Voltage Gain Four-Phase Interleaved Boost Converter . . . . . . . . . . . . . . . . 763 K. Jyotheeswara Reddy and N. Sudhakar 70 Analysis and Comparative Study of Various Charging Methods Implemented for Solid-State Marx Generator . . . . . . . . . . . . . . . . . 773 Neelam S. Pinjari, S. Bindu and Ruchi D. Singh 71 Reliability Modeling of Multiphase DC–DC Boost Converter . . . . . 787 D. Umarani and R. Seyezhai 72 A Novel ANN-SMC-Based Maximum Power Point Tracking for Efficient DC Stage Conversion of a Solar PV Power Plant . . . . 803 Bijit Kumar Dey, Nirabhra Mandal and Ankur Bhattacharjee 73 Coordinated Control of DC Electric Springs for Reduction of Main Grid Dependability . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 815 S. Hari Charan Cherukuri, B. Saravanan and K. S. Swarup 74 A PI with Fuzzy-Based Multifunctional DSTATCOM Operating Under Stiff Source . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 825 Sampath Kumar Pappula and Sushama Malaji 75 A Novel Three-Phase Five-Level Inverter Control and Its Performance Analysis for a Grid-Connected Solar PV Power System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 839 Nirabhra Mandal, Bijit Kumar Dey, Abhishek Paul and Ankur Bhattacharjee

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76 Enhancement of Machine Performance by Deploying Superconductors with Numerical Analysis and Updated Characteristics—A Novel Approach . . . . . . . . . . . . . . . . . . . . . . . . 851 Sasidharan Srinivasan, Sethuraman Sivakumar and Krishna Kumar Rathinam 77 Analysis of Three-Phase Quasi-Switched Boost Inverter Topology for Renewable Energy Applications . . . . . . . . . . . . . . . . . . . . . . . . 863 P. Sriramalakshmi, A. Arvindh, S. R. Sanjay Kumar, M. Prasanth and V. T. Sreedevi 78 Unbalanced Voltage Mitigation with Reactive Power Control of Grid-Tied Solar PV System . . . . . . . . . . . . . . . . . . . . . . . . . . . . 877 Swathy Pillai, Sushil Thale and Akshay Purohit Author Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 893

About the Editors

Suneet Singh is a faculty at the Department of Energy Science and Engineering, Indian Institute of Technology Bombay (IITB), India. He received his MTech and PhD in Nuclear Engineering from the IIT Kanpur and the University of Illinois at Urbana-Champaign, USA respectively. He completed his postdoctoral research at Idaho National Lab, USA. He has received the Bhaskara Advanced Solar Energy (BASE) Fellowship 2014 from the Indo-US Science & Technology Forum (IUSSTF). His research interests include stability analysis of nuclear reactors, advanced numerical methods for fluid flows and neutron diffusion, analytical solution of multilayer heat conduction problems, and solar thermal heat transfer. Venkatasailanathan Ramadesigan is a faculty at the Department of Energy Science and Engineering, Indian Institute of Technology Bombay (IITB), India. He received his MS in Chemical Engineering from the University of South Carolina, USA, and PhD in Energy, Environmental, and Chemical Engineering from Washington University in Saint Louis, USA. His research interests include modelling and simulation of chemical and electrochemical processes, electrochemical large/grid-scale energy storage systems, system integration, nonlinear parameter estimation, and system-level optimization and control, as well as numerical and applied mathematics.

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Mathematical Modeling of Heat Losses from Cylindrical Cavity Receiver in Solar Parabolic Dish R. Sinha and N. P. Gulhane

Abstract The recent experimental investigations on model receiver with constant heat flux boundary condition have shown that the temperature profile along the cavity walls is non-uniform and seen to vary with cavity inclination. This paper presents a mathematical analysis of heat losses from cylindrical cavity receiver applied to constant heat flux boundary conditions. In addition, the empirical correlations for the radiation Nusselt number and total heat loss Nusselt numbers, with its influencing parameters like Grashof number (Gr), cavity inclination angle (θ ), temperature ratio (Ta /Tw ), and conductance parameter (γ ), are proposed. The mathematical analysis and empirical correlations are based on experimental results in previous published data. In mathematical analysis, the heat loss by natural convection is observed to be more sensitive to the cavity inclination angle in comparison with heat loss by radiation and conduction. The heat loss by radiation and conduction are not constant as initially estimated; they increase with increase in cavity inclination. It led us to conclude that it may not be accurate to predict convection heat loss using previously developed correlations based on the isothermal wall condition. Secondly, even though the variation in heat loss by radiation and conduction with cavity inclination is small, it needs to be considered for accurate design of solar parabolic dish receiver system. Keywords Cavity receiver · Receiver inclination · Boundary condition · Nusselt number correlation

R. Sinha (B) · N. P. Gulhane Department of Mechanical Engineering, VJTI, Mumbai, India e-mail: [email protected] R. Sinha K. J. Somaiya College of Engineering, Mumbai, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_1

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R. Sinha and N. P. Gulhane

Units and Symbol Nomenclature A d D g h K L P q T θ β γ ε θ υ σ

Area (m2 ) Cavity diameter (m) Aperture diameter (m) Gravitational acceleration (m/s2 ) Heat transfer coefficient (W/m2 -K) Conductivity (W/m °C) Length of cavity (m) Power (W) Heat flux (W/m2 ) Temperature (K) Angle of inclination of the receiver (°) Thermal expansion coefficient (1/°C) Conductance parameters Cavity cover emissivity Angle of inclination of the receiver (º) Kinematic viscosity (m/s2 ) Stefan–Boltzmann constant (W/m2 K4 )

Subscripts a ap cond conv f In ins rad t w

Ambient Aperture Conduction Convection Film Input Insulation Radiation Thickness Wall

Abbreviations AR OR Gr Nu

Aspect ratio Opening ratio Grashof number Nusselt number

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3

1.1 Introduction The concentrating solar technology is most important as it can be used for process industry as well as for generating power. The parabolic dish receiver assembly is one such useful system. It usually consists of a reflector in the form of a dish with downward-facing receiver at the focus of the dish. Generally, a cavity receiver is used since it can maximize the solar radiation absorption of the concentrated solar flux and minimize heat losses [1]. The heat losses in solar parabolic dish system include convective and radiative losses through the cavity opening and conduction through the solid structure and through the insulation used behind the cavity surfaces to reduce conduction. The heat loss due to conduction is smaller and can be calculated by analytical method. The heat loss by radiation is dependent on the cavity wall temperature, the shape factors and emissivity/absorptivity of the receiver walls, while conduction is dependent on the receiver temperature and the material used for insulation. The heat loss by radiation and conduction are observed to be independent of the cavity inclination [2, 3]. The heat loss by convection depends on the air temperature within the cavity, the cavity inclination and the external wind velocity, and complex to estimate [4, 5]. A detailed review on the convection heat loss for different shapes of cavity, normally used in different types of engineering systems, is also available in the literature [6]. A Nusselt number correlation was also proposed, relating cavity inclination and aperture size for a cylindrical cavity receiver [7]. On the basis of experimental result, another correlation of Nusselt number variation with Grashof number and the surface temperature of cylindrical cavity were also developed [8]. The Australian National University (ANU) conducted number of experiments to study the convection heat loss of a cylindrical cavity with isothermal surface boundary [9– 11]. Subsequent experiments had confirmed the same conclusion [12]. As a result of extensive investigation, a series of Nusselt number correlations were obtained from experimental results. A brief review on the natural convection in cavity receiver is available in the literature [6, 12–16]. The heat transfer coefficient for a rectangular cavity was also investigated by many researchers experimentally [17, 18]. Recent investigation result on electrically heated cylindrical cavity models tested the effect of constant heat flux and cavity inclination on the convective losses and concluded that the heat losses are also dependent on the surface boundary conditions [19]. In order to better understand the mechanism of heat loss under constant heat flux, many researchers have studied both experimentally and numerically. Moreover, the literature survey shows that most of those investigations on the heat loss of cavity were limited to the isothermal and/or adiabatic surfaces’ boundary conditions. AbbasiShavazi conducted the experimental investigations on model receiver to study the convection heat loss from a cylindrical cavity receiver, applied to a range of constant heat flux boundary condition [20]. However in the constant heat flux boundary condition, the temperature distribution along the cavity walls was non-uniform and was a function of cavity inclination. The heat loss by natural convection was observed to be more sensitive to the cavity inclination as compared to radiation and conduction heat losses.

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The cavity inclination has a less effect on the radiation and conduction heat losses. Also, these losses were not constant as initially estimated. It led us to conclude that it may not be sufficiently accurate to predict convection heat loss using previously developed correlations based on isothermal boundary wall condition. Secondly, even though the variation in radiation and conduction heat losses with cavity inclination is small, it needs to be considered for accurate design of cavity receiver. However, to the author’s best knowledge, very few empirical correlations are available in the literature to estimate heat loss from cavity receiver subjected to constant heat flux boundary condition. This paper presents a mathematical analysis of heat loss through solar cavity receiver, subjected to a range of constant heat flux boundary conditions. The empirical correlations of the total heat loss Nusselt numbers and radiation heat loss Nusselt are proposed, incorporating the effecting parameters like Grashof number (Gr), inclination angle of the receiver (θ ), cavity emissivity (ε), area ratio (Ta /Tw ), temperature ratio, conductance parameters (γ ), and ambient temperature. The model is developed from experimental results in previous published data [20]. The empirical correlation can be used to estimate total heat loss and radiation heat loss from a cavity receiver, subjected to constant heat flux boundary condition.

1.2 Analysis of Heat Loss Based on Constant Input Power In recent experimental investigation, the heat loss from model cavity receiver purpose, a laboratory-scale cylindrical cavity models, was investigated. The model receiver geometric aspect ratios (cavity length to diameter) of 1 and 2 and aperture opening ratios (aperture diameter to cavity diameter) of 0.5 and 1 were considered. A constant heat flux was applied to model receiver and non-uniform temperature was observed inside cavity receiver [20]. During the experiment, the total power input was maintained constant (equal to the power loss from cavity receiver). The system was operated at steady state; under steady state, the power delivered to the heating cable is lost by conduction, convection, and radiation to the surroundings, as represented in Eq. 1.1, Pin = Pcond + Pconv + Prad

(1.1)

The input power (Pin ) set at the value by using power controller, and the convective (Pconv ) loss must be indirectly estimated from the value of total input power (Pin ), conduction power loss (Pcond ), and radiation power loss (Prad ). Power input Pin can be expressed as where V is the voltage in Volt and I is current in Ampere. Pin = V I

(1.2)

To analyze the nature of heat loss from cavity receiver, heat loss by convection and radiation at different inclination was taken from experimental result. Figure 1.1 shows

1 Mathematical Modeling of Heat Losses from Cylindrical …

5

100 90

y = -0.0032x2 + 0.7719x + 32.464 R² = 0.997

Heat Loss (W)

80 70

y = 92.385e-0.014x R² = 0.9923

60

Prad

50

Pconv

40

Pcond

30

Poly. (Prad)

y = 0.2271x + 27.997 R² = 0.9775

20

Expon. (Pconv)

10 Linear (Pcond)

0 0

20

40

60

80

100

Inclina on angle (θ) Fig. 1.1 Variations of heat losses with cavity inclination (θ) at constant heat flux q = 3195 w/m2 (P = 150 W) [20]

the variations of conduction, radiation, and convection power losses with inclination under constant input power (150 W) and aspect ratio 2 (AR = L/d) as measured in the experiment [20]. Convection and radiation power loss is taken from experimental data for constant heat input 150 W, and conduction heat loss has been estimated by subtraction heat loss by convection and radiation from heat input as in Eq. (1.1). The convection and radiation data is best fitted with second degree of polynomial and regression factor more than 0.98. However, the heat loss by conduction varies linearly with the cavity inclination. As shown in Fig. 1.1, the heat losses by radiation and conduction increases slightly with the increase of cavity inclination, while the heat loss by natural convection decreases greatly with increase in cavity inclination. This observation reveals that the natural heat loss by natural convection is more sensitive to the cavity inclination angle in comparison with the heat loss by radiation and conduction. This can also be observed from mathematical equation that for convection at zero inclination (θ = 0) heat loss is maximum, and it decreases greatly with inclination with the order of second degree of polynomial. The rate of radiation heat loss is higher as compared to conduction heat loss, but with increase in inclination (θ ), rate of radiation heat loss decreases. This variation in heat transfer is similar to the previous experimental results [19].

1.3 The Development of Nusselt Number Correlation The main influencing parameters to estimate heat losses through cavity receiver such as Grashof number (Gr) opening ratio Aap /Aw , angle of inclination (θ ), cavity wall temperature (T w ), material thickness (t), conductivity (k), and emissivity (ε) plays major role.

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From the power balance inside a cavity receiver (Eq. 1.1), we have Pin = Pcond + Pconv + Prad where (Pconv ) is the power loss by natural convection from cavity, (Pcond ) is the power loss by conduction which is loss from the heated cavity surface to the surrounding through insulation and (Prad ) is the power loss by radiation from the heated cavity surface to the ambient air, and (Pin ) is the total input power and can be expressed as Eq. 1.1. A model receiver was examined experimentally to estimate the heat losses from a model solar cavity receiver. For this purpose, laboratory-scale cylindrical cavity models with geometric aspect ratios (cavity length to diameter) of 1 and 2 and aperture opening ratios (aperture diameter to cavity diameter) of 0.5 and 1 were considered. Receiver was subjected to constant heat flux boundary condition, and non-uniform temperature was observed inside cavity receiver [20]. The radiative and total Nusselt number empirical correlations have been developed, relating the influencing parameters like Grashof number (Gr), cavity inclination angle (θ ), cavity emissivity (ε), temperature ratio (Ta /Tw ), area opening ratio (Aw /Aap ), and conductance parameter (γ ). To develop the empirical correlations of Nusselt number versus the inclination, the total heat flux q is maintained constant and is taken at the power input (Pin ) during the experiment. The Grashof number for this calculation is defined as [20]. Gr = gβqd 4 /ka ϑa2

(1.3)

The parameters like g is the gravitational acceleration, β is the coefficient of thermal expansion of air, q is the total heat flux at power input, and ϑa is the kinematic viscosity of air. Fluid properties are taken corresponding to film temperature. Where film temperature, Tf = (Tw + Ta )/2 [20]. As opening ratio Aap /Aw is same in experiment and emissivity (ε) is 0.86–0.88, average value 0.87 was taken for all the data, so the variable parameter is Grashof number (Gr), angle of tilt (θ ), temperature ratio (Ta /Tw ), and conduction resistance parameters (γ ). The ambient temperature ((Ta ) is taken as 298 K. The ratio of effective thermal resistance of the solid to the thermal resistance of the fluid is taken as conduction resistance parameters (γ ) and is given by (γ ) = [(tw /kw ) + (tins /kins )]/(d/ka )

(1.4)

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1.3.1 Development of Total Heat Loss Nusselt Number Correlation from Experimental Result The empirical correlations for the total heat loss have been developed, relating the influencing parameters like Grashof number (Gr), inclination angle of the receiver (θ ), cavity emissivity (ε), temperature ratio (Ta /Tw ), area opening ratio (Aap /Aw ), and conductance (γ ). For the development of the total Nusselt number correlation, the total Nusselt number is given by Nutotal = Pin d/Aw ka (Tw − Ta )

(1.5)

The total Nusselt number (Nutotal ) is function of Grashof number (Gr), inclination of the receiver (θ ), cavity emissivity (ε), temperature ratio (Ta /Tw ), area opening ratio (Aw /Asp ), and conductance parameter (γ ). The empirical correlations for total Nusselt number relating the four influencing parameters, i.e., Grashof number (Gr), angle of tilt (θ ), temperature ratio (Ta /Tw ), and conductance resistance parameters (γ ) are expressed as:  ft  gt Nutotal = atGrbt (1 + cos θ )ct (Aap /Aw )dt (1 + ε)et 1 − (Ta /Tw )4 1/(1 + γ ) (1.6)    2.157 −43.43247 1/(1 + γ ) Nutotal = 108.185184 Gr−0.8764 (1 + cos θ )0.1202 1 − (Ta /Tw )4 (1.7) The total Nusselt number (Nutotal ) is calculated at different inclination and the corresponding wall temperature from Eq. 1.7 and cavity inclination (θ ). The actual total heat loss (Ploss ) is calculated from Eq. 1.8 at calculated total Nusselt number (Nutotal ), where Tw is the mean cavity surface temperature, and Aw is the total heat transfer area of the cavity wall. The variation of total heat transfer coefficient (h total ) at different inclination is estimated from Eq. (1.9) at estimated power loss (Ploss ) from Eq. 1.8. The ambient temperature is taken as 298 K. Nutotal = Ploss d/Aw ka (Tw − Ta )

(1.8)

Ploss = h total Aw (TW − Ta )

(1.9)

The change in total Nusselt number with cavity inclination is given in Fig. 1.2.

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20 15

Poly. (Series1)

10

y = 0.0017x2 - 0.2757x + 26.814 R² = 0.9864

5 0 0

20

40

60

80

100

Fig. 1.2 Variation of total Nusselt number (Nutotal ) with cavity inclination (θ)

1.3.2 Development of Radiation Heat Loss Nusselt Number Correlation from Experimental Result The empirical correlations for the radiation have been developed, with the relating parameters like Grashof number (Gr), inclination angle of the receiver (θ ), emissivity of cavity (ε), ratio of temperature (Ta /Tw ), area opening ratio (Aap /Aw ), and conductance parameters (γ ). The ambient temperature is taken as 298 K. The radiation heat loss is small as compared to convection heat loss as given in Eq. 1.10.   Prad = εa σ Aap Tw4 − Ta4

(1.10)

  εa = 1/ 1 − (1 − εw )(1 − Aap /Aw )

(1.11)

The heat loss by radiation is also an important factor that affects the performance of a cavity receiver. The heat loss by radiation Nusselt number is given by Eq. 1.12 (Wu et al. 2013). Where εa is the effective emissivity of cavity, σ is the Stefan– Boltzmann constant, and Aap is the area of opening. Effective emissivity is calculated using Eq. 1.11. Nur = Prad d/[ Aw ka (Tw − Ta )]

(1.12)

By substituting Prad from Eq. (1.10) into Eq. (1.12), we have        Nurad = εw σ Aap Tw4 − Ta4 d / 1 − (1 − εw )(1 − Aap /Aw ) Aw ka (Tw − Ta ) (1.13)

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160 140 120 Series1

100 y = -0.0044x2 + 0.9312x + 90.917 R² = 0.9838

80 60

Poly. (Series1)

40 20 0 0

20

40

60

80

100

Cavity Incilna on (θ)

Fig. 1.3 Variation of radiation Nusselt number (Nurad ) with cavity inclination (θ)

Nusselt number can be expressed with the relation given in Eq. 1.14.  fr  gr Nurad = aGrbr (1 + cos θ)cr (Aap /Aw )dr (1 + ε)er 1 − (Ta /Tw )4 1/(1 + γ ) (1.14) With Grashof number from Eq. 1.3, emissivity 0.87, conductance ration (γ ) from Eq. 1.4 and Nusselt number from Eq. 1.13, Nusselt number radiation correlation is found to be  81.49707  −0.28799 1/(1 + γ ) Nurad = 10−11 aGr1.912495 (1 + cos θ )−0.20818 1 − (Ta /Tw )4

(1.15) The change in radiation Nusselt number with cavity inclination is given in Fig. 1.3.

1.4 Results and Conclusions This paper presents a mathematical analysis of cylindrical cavity receiver heat loss under constant heat flux cavity wall boundary conditions. • To analyze the nature of heat loss from cavity receiver, convection and radiation heat loss at different cavity orientations has been taken from experimental result [20]. Figure 1.1 shows the variations of conduction, radiation, and convection power losses with inclination under constant input power (150 W) and aspect ratio 2 (AR = L/d) as measured in the experiment. Convection and radiation heat loss is taken from experimental data for constant power input 150 W, and conduction heat loss has been estimated by subtraction heat loss by convection and radiation from power input.

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• This can be observed from mathematical analysis result that the heat loss by convection is maximum at zero inclination and decreases greatly with cavity inclination in the order of second degree of polynomial (Fig. 1.1). The conduction heat loss increases linearly with the inclination. The radiation heat losses increases in the order of second degree of polynomial, and the heat loss by radiation decreases with increase in cavity inclination. The rate of heat loss by natural convection is more sensitive to the cavity tilt angle as compared to heat loss by radiation and conduction. The heat losses by radiation and conduction are not constant as initially estimated; they increase with increase in cavity inclination. This variation is very similar to the previous experimental results [19]. • The empirical correlations for radiation and total heat loss Nusselt numbers relating influencing parameters, the Grashof number (Gr), cavity inclination angle (θ ), temperature ratio (Ta /Tw ), and conductance parameter (γ ) are proposed on the basis of experimental results of cavity temperature profile in model receiver obtained at 200 W power input [20]. The total Nusselt number (Nutotal ) is highest at the 0◦ and slightly decreases with increase in inclination and minimum at 90° as shown in Fig. 1.2. This result is similar to heat loss from actual helical coil type cavity receiver, where maximum heat loss was observed at 0◦ and minimum at 90° [12]. The radiation Nusselt number (Nurad ) is minimum at 0◦ and increases with inclination as shown in Fig. 1.3. • Experimental investigation is necessary to model heat losses from cavity receivers. Initially, solar cavity receivers used for high-temperature applications were generally modeled as plain walls with uniform temperature boundary conditions, where radiation and conduction were independent of inclination and observed to be constant. • It led us to conclude that it may not be sufficiently accurate to predict convection heat loss using previously developed correlations based on isothermal boundary wall condition. Secondly, even though the variation in heat loss by radiation and conduction with cavity inclination is small, it needs to be considered for accurate design of solar parabolic dish receiver system. • A very few experiments are carried out on actual helical coil type cavity receiver, where temperature is observed to be non-uniform. More experimental investigation is required for accurate heat loss prediction from cavity receiver as it is most important parameter to decide efficiency and cost-effectiveness of concentrated solar power parabolic dish receiver system. Acknowledgements A mathematical modeling and Nusselt number of heat loss correlation form cavity receiver are proposed on the basis of literature survey and the research work done in the past as well as in recent years. The authors gratefully acknowledge the contribution of all the researchers and professors working in this area for giving valuable insight into the topic and guided me in right direction. The author acknowledges the contribution of K.J. Somaiya management for sponsoring Ph. D research. The authors gratefully acknowledge the contribution of the VJTI, Mumbai Research Fund and for giving the opportunity to explore research scope in this area.

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References 1. J.A. Harris, T.G. Lenz, Thermal performance of concentrator/cavity receiver systems. Sol. Energy 34(2), 135–142 (1985) 2. W.B. Stine, C.G. McDonald, Cavity receiver heat loss measurements, in Proceedings of ASME Solar Energy Division Conference. Denver, Colorado (1988) 3. U. Leibfried, J. Ortjohann, Convective heat loss from upward and downward-facing cavity solar receivers: measurements and calculations. ASME J. Sol. Energy Eng. 117(2), 75–84 (1995) 4. A.M. Clausing, Convective losses from cavity solar-receivers—comparisons between analytical predictions and experimental results. ASME J. Sol. Energy Eng. 105(1), 29–33 (1983) 5. W.B. Stine, C.G. McDonald, Cavity receiver convective heat loss, in Proceedings of the international solar energy society (ISES) solar world conference (Kobe, Japan, 1989) 6. S.Y. Wu, L. Xiao, Y. Cao, Y.R. Li, Convection heat loss from cavity receiver in parabolic dish solar thermal power system: a review. Sol. Energy 84(8), 1342–1355 (2010) 7. A.A. Koenig, M. Marvin, Convection heat loss sensitivity in open cavity solar receivers. Final Report, DOE Contract No: EG77-C-04-3985, Department of Energy, USA (1981) 8. D.L. Siebers, J.S. Kraabel, Estimating convective energy losses from solar central receivers. Sandia National Laboratories Report, SAND 84-8717 (1984) 9. T. Taumoefolau, K. Lovegrove, An experimental study of natural convection heat loss from a solar concentrator cavity receiver at varying orientation, in Proceedings of 40th Conference of the Australia and New Zealand Solar Energy Society (ANZSES) (Newcastle, Australia, 2002) 10. T. Taumoefolau, S. Paitoonsurikarn, G. Hughes, K. Lovegrove, Experimental investigation of natural convection heat loss from a model solar concentrator cavity receiver. ASME J. Sol. Energy Eng. 126(2), 801–807 (2004) 11. S. Paitoonsurikarn, K. Lovegrove, A new correlation for predicting the free convection loss from solar dish concentrating receivers, in Proceedings of 44th ANZSES Conference (Australia, 2006), pp. 1–9 12. M. Prakash, S.B. Kedare, J.K. Nayak, Investigations on heat losses from a solar cavity receiver. Sol. Energy 83, 157–170 (2009) 13. M. Prakash, S.B. Kedare, J.K. Nayak, Numerical study of natural convection loss from open cavities. Int. J. Therm. Sci. 51(1), 23–30 (2012) 14. J.O. Juárez, J.F. Hinojosa, J.P. Xamán, M.P. Tello, Numerical study of natural convection in an open cavity considering temperature-dependent fluid properties. Int. J. Therm. Sci. 50, 2184–2197 (2011) 15. M. Prakash, Numerical study of natural convection heat loss from cylindrical solar cavity receivers, Hindawi Publishing Corporation, ISRN Renewable Energy Volume 2014, Article ID 104686, 7 pages (2014) 16. J. Samanes, J. García-Barberena, F. Zaversky, Modeling solar cavity receivers: a review and comparison of natural convection heat loss correlations. Energy Procardia. 69, 543–552 (2015) 17. W. Chakroun, M.M. Elsayed, S.F. Al-Fahed, Experimental measurements of heat transfer coefficient in a partially/fully open tilted cavity. ASME J. Sol. Energy Eng. 119(4), 298–303 (1997) 18. W. Chakroun, Effect of boundary wall conditions on heat transfer for fully opened tilted cavity. ASME J. Heat Transf. 126(6), 915–923 (2004) 19. S.-Y. Wu, J.-Y. Guan, L. Xiao, Z.-G. Shen, L.-H. Xu, Experimental investigation on heat loss of a fully open cylindrical cavity with different boundary conditions. Exp. Thermal Fluid Sci. 45, 92–101 (2013) 20. E. Abbasi-Shavazi, G.O. Hughes, J.D. Pye, Investigation of heat loss from solar cavity receiver. Energy Procedia. 69, 269–278 (2015)

Chapter 2

Performance Evaluation of Latent Heat Storage Filled with Paraffin Wax for Solar Thermal Applications D. Gudeta, S. R. Jena, P. Mahanta and P. S. Robi

Abstract Owing to the non-uniform availability of solar radiation, designing of a latent heat storage found necessary so as to bridge the supply and demand gap. In the current investigation, the charging and discharging characteristics of a 10 MJ capacity, paraffin wax-based latent heat storage are analyzed numerically. Validations with experimental results showed reasonably good agreement following which parametric studies are conducted and detailed discussion on results are presented. Keywords Charging and discharging · Latent heat storage · Paraffin wax · Melt fraction

2.1 Introduction Due to fluctuation in solar heat flux owing to adverse weather conditions, storage of solar thermal energy is necessary and it provides a means to utilize the thermal energy for nocturnal use. Directions of research involving latent heat thermal energy storage (LHS) have become manifold ascribed to its many practically relevant applications which include building heating, providing hot water for domestic needs, refrigeration applications, drying equipment and waste heat recovery. It has innumerous advantages comparable to sensible energy storage (SES) device in terms of effectiveness and storage capacity. Study by Sharma et al. [1] revealed that energy storing capacity of LHS is 14 times higher than SES. There are various types of storage media for LHS out of which, paraffin wax is cheap and easily available. Agarwal and Sarviya [2] investigated a shell and tube type latent heat storage (LHS) for solar drying application using paraffin wax as the storage medium. In D. Gudeta · S. R. Jena · P. Mahanta · P. S. Robi (B) Department of Mechanical Engineering, Indian Institute of Technology, Guwahati 781039, India e-mail: [email protected] D. Gudeta e-mail: [email protected] P. Mahanta e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_2

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their study, the effect of flow rate and temperature of heat transfer fluid (HTF) during charging and discharging process were also analyzed. Kabeel et al. [3] conducted an experimental study and evaluated the energy yield and performance of a solar desalination system by comparing a conventional system with an improved LHSbased system. The result indicated that the LHS-based system showed an improved performance in terms of daily freshwater yield. Melting and solidification in a double pipe heat exchanger were investigated by Jesumathy et al. [4] using paraffin wax as the means of storage. They reported that charging is dominated by natural convection and discharging is dominated by conduction. Performance of a finned solar air heater is experimentally studied by Kabeel et al. [5]. The effects of variation in mass flow rate on daily and instantaneous efficiency were measured. The result revealed that the efficiency increased by 10.8–13.6%. Hu et al. [6] performed a wide investigation on PCMs for investigating the thermal management of electronic devices and emphasized on using the modularized thermal storage unit as an improvement over current PCM-based heat sinks for cooling applications in high duty electronic equipment. Korti and Tlemsani [7] researched a latent heat-based energy storage system with various types of paraffin and the effect of inlet temperature and flow rate of HTF were studied. It was also noted that addition of engine oil to paraffin improved the charging and discharging process by 42.4 and 66%, respectively. Experimental investigations have been performed by Sobolˇciak et al. [8] for various compositions of linear low-density polyethylene, paraffin wax and expanded graphite using both conventional and non-conventional methods. From the testing, it is found that thermal conductivity was improved by adding extended graphite. Melting process of industrial grade paraffin wax-based energy storage was studied by Saraswat et al. [9] both experimentally and numerically (using OpenFOAM). They highlighted about using the copper pipes along with the PCM to enhance heat dissipation rate. Salunkhe and Krishna [10] reviewed the recent works on latent heat storage materials (LHSMs) and their thermophysical properties. Further, they discussed in detail the various factors affecting the life of a LHSM. Experimental studies involving paraffin wax-based latent heat storage with an application in forced convection solar dryer was reported by Rabha and Muthukumar [11] and the exergy and energy efficiencies were reported as 18.3–20.5% and 43.6–49.8%, respectively. Wahid et al. [12] provided a comprehensive review of literatures based on the various features of the PCMs, their latest developments and future directions with an application towards the building architecture. Naghavi et al. [13] analyzed a solar water heating system by implementing an evacuated tube heat pipe solar collector along with latent heat storage (LHS). Results indicated that the system efficiency in the summer was found to be 38–42% and it showed a fluctuation of about 8% in the rainy season. Németh et al. [14] prepared microcapsules containing paraffin wax and studied the various process parameters. Khan et al. [15] performed parametric investigation in a shell and tube-based thermal energy storage to study the performance of LHS. It was observed that increase in inlet temperature increases the efficiency of the storage. Comparison study between a naturally cooled and a storage-based latent heat cooled PV solar panels conducted by Tana et al. [16] was found that the panel temperature of the latent heat cooled shell reduced by 15 °C in comparison with a naturally cooled PV panel. It is observed from

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the above literatures that most of the studies were either numerical or experimental. However, studies including both methods were a few. The objective of the present study is to perform the numerical analysis of a latent heat storage system (LHS) and, subsequently, to compare the results with the experimental model of 10 MJ capacities using paraffin wax as PCM. Various performance parameters have been analyzed during the charging/discharging process and are reported in terms of energy stored/released, variation of melt fraction, etc.

2.2 Numerical Details 2.2.1 Modelling and Assumptions The three-dimensional sectional view of a shell and tube heat exchanger is presented with paraffin wax as the phase change material (PCM) and water as the heat transfer fluid (HTF). The tube radial-thickness and tube internal diameter are maintained as 4 mm and 14 mm, respectively, whereas the shell length and diameter are maintained as 1000 mm and 300 mm, respectively. The data related to the thermophysical properties of paraffin wax are taken from Niyas et al. [15] and are mentioned as follows. The properties of PCM, namely, thermal conductivity (k), specific heat capacity (C p ), density (ρ), dynamic viscosity (μ) and latent heat of fusion (L) are maintained as 0.25 W m−1 K−1 , 2000 J kg−1 K−1 , 780 kg m−3 , 0.0041 kg m−1 s−1 and 168 kJ kg−1 , respectively. The melting temperature and the melting range are maintained at 315.15 K and 3 K, respectively. While simulating the model, the following assumptions were considered (Fig. 2.1).

Fig. 2.1 Flow domain

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1. HTF is an incompressible and Newtonian fluid. 2. PCM is homogeneous and isotropic, its initial temperature is assumed to be uniform throughout the domain. 3. Viscous dissipation is neglected in the flow and phase change is assumed to occur in a temperature range.

2.2.2 Governing Equations The model considers conjugate heat transfer across the thickness of the tube which involves the heat transfer in heat transfer fluid (HTF) as well as in the phase change material (PCM). For incorporating the effect of latent heat, an effective heat capacity (EHC) method is used in which C P, EFF is calculated keeping into consideration the latent heat of fusion. The governing equation for mass conservation, momentum conservation and energy conservation are written in Eqs. (2.1)–(2.3) as follows. The quantities are in their dimensional form. ∂V ∂U + =0 ∂x ∂y   2  ∂p  2 ∂U ∂U ∂U 1 − ∂ x + μ ∂∂ xU2 + ∂∂ yU2 +U +V = 2 ∂t ∂x ∂y ρ +ρgβ(T − T∞ ) + (1−θ) A V (θ 3 +ε) MUSH     2 ∂T ∂ T ∂2T ∂T ∂T =k + +U +V ρC p ∂t ∂x ∂y ∂x2 ∂ y2

(2.1)

(2.2)

(2.3)

In the above equations, U, V, T, β, θ , AMUSH and p, etc., represent stream wise velocity component (m/s), span wise velocity component (m/s), temperature (K), volume expansion coefficient, melt fraction, mushy zone constant and pressure (Pa), respectively.

2.2.3 Initial and Boundary Conditions The initial temperature for HTF and PCM are set as 298 K, whereas no flow boundary condition is set for velocity. At the inlet, boundary condition for HTF is set to be U = 0.05 m/s and T = 333 K. In order to avoid heat losses to the atmosphere, adiabatic boundary condition is set at the shell outer surface and the heat losses to the ambient through this surface are assumed to be negligible. While solving, only one-fourth of the flow domain is solved and symmetry boundary condition is used to reduce the number of mesh elements.

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Fig. 2.2 Experimental setup depicting various components and their arrangement

2.2.4 Meshing and Numerical Treatment Computational domain for the current investigation consists of 406,380 volume elements. Finite elements method based COMSOL Multiphysics 4.3a is used to solve continuity, momentum and energy equation. BDF time discretization scheme is used in conjunction with a nonlinear time-dependent PARADISO solver. Convergence criteria for velocity and temperature are set to a prescribed limit of 10−3 .

2.3 Experimental Setup The LHS system comprises of 17 copper tubes with a shell that is made up of stainless steel. Thermo-foam insulation is provided at the shell outer surface to reduce heat loss to the ambient. T-type thermocouples were used to measure the temperature at the different location of the LHS (Fig. 2.2).

2.4 Result and Discussions 2.4.1 Validation Study Validation study is performed for the present investigation in terms of temporal variation of volumetric-averaged temperature, with the experimental results for charging/discharging process and the results are revealed in Figs. 2.3. Numerical results

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Fig. 2.3 Variation of volumetric-averaged temperature with time a charging and b discharging

provided a reasonably good comparison with the experimental data and moreover, the profile showed a similar trend with the reported results of Niyas et al. [17]. After performing this validation, results are reported in terms of variation of volume-averaged temperature, melt fraction, sensible, latent and total energy stored/released with time as obtained in the simulation.

2.4.2 Contours of Average Melt Fraction Instantaneous melt fraction contours of PCM are presented at different time instances as depicted in Fig. 2.4. As charging is a convection dominated process, faster melting rate is observed compared to a relatively slower rate of solidification during discharging. Simulation data substantiates that time taken for complete melting is t = 110 min (6600 s), whereas complete solidification of paraffin wax is obtained at t = 150 min (9000 s).

2.4.3 Variation of Average Temperature and Melt Fraction During Charging and Discharging The HTF is circulated through the pipe at 333 K/298 K during charging/discharging process and the variation of average temperature is recorded with time as shown in Fig. 2.5. It can be comprehended from the figures that the average temperature varies smoothly over time for charging process, whereas during release of energy, the average temperature profile followed a steep decrease in the beginning which is followed by a flat slope. Further, volume-averaged melt fraction is also plotted, as seen in Fig. 2.6, to understand the melting and solidification process comprehensively.

2 Performance Evaluation of Latent Heat …

19

(a)

(f)

(b)

(g)

(c)

(h)

(d)

(i)

(e)

(j)

Fig. 2.4 Instantaneous contour of melt fraction during charge (as shown in a, b, c, d and e) and discharge (as shown in f, g, h, i and j)

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Fig. 2.5 Variation of average temperature with time during a charging and b discharging

Fig. 2.6 Variation of average melt fraction with time during a charging and b discharging

When melt fraction becomes zero, it suggests pure solid paraffin wax and when its value becomes one, it represents pure liquid paraffin suggesting complete melt.

2.4.4 Energy Stored/Released The change in sensible, latent and total energy in both charging and discharging process are shown in Fig. 2.7. During charging process, when HTF passes through the tubes, PCM starts melting. In the initial phase, sensible heat transfer occurs which follows a latent transfer of heat during phase change and, subsequently, sensible transfer of heat from liquid paraffin. When the PCM reached 333 K, the sensible, latent and total energy stored are found to be 3.152 MJ, 7.5 MJ and 10.71 MJ, respectively. Similarly, while discharging, when averaged temperature of PCM reached 302 K

2 Performance Evaluation of Latent Heat …

21

Fig. 2.7 Variation of energy stored/released with time during charging/discharging

the sensible, latent and total energy released are found to be 2.751 MJ, 7.5 MJ and 10.2 MJ, respectively.

2.5 Conclusions Numerical analysis of a 3D shell and tube type LHS was studied and the storage characteristics are analyzed for both the charging and discharging process. After performing validation study with the experimental results, reasonably good agreement is found. Further, performance parameters of the LHS system are evaluated in terms of volume-averaged temperature, melt fraction, energy stored/released and a few key findings are summarized below. • Time taken for complete charging and discharging are found to be 110 min and 150 min, respectively. • Charging, being convection dominated process, is faster than discharging. • The total energy stored and released during the process are 10.71 MJ and 10.2 MJ, respectively.

References 1. A. Sharma, V.V. Tyagi, C.R. Chen, D. Buddhi, Review on thermal energy storage with phase change materials and applications. Renew. Sustain. Energy Rev. 13, 318–345 (2009) 2. A. Agarwal, R.M. Sarviya, An experimental investigation of shell and tube latent heat storage for solar dryer using paraffin wax as heat storage material. Eng. Sci. Technol., Int. J. 19, 619–631 (2016)

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3. A.E. Kabeel, M. Elkelawy, H.A.E. Din, A. Alghrubah, Investigation of exergy and yield of a passive solar water desalination system with a parabolic concentrator incorporated with latent heat storage medium. Energy Convers. Manag. 145, 10–19 (2017) 4. S.P. Jesumathy, M. Udayakumar, S. Suresh, S. Jegadheeswaran, An experimental study on heat transfer characteristics of paraffin wax in horizontal double pipe latent heat storage unit. J. Taiwan Inst. Chem. Eng. 45, 1298–1306 (2014) 5. A.E. Kabeel, A. Khalil, S.M. Shalaby, M.E. Zayed, Improvement of thermal performance of the finned plate solar air heater by using latent heat thermal storage. Appl. Therm. Eng. 123, 546–553 (2017) 6. J. Hu, R. Hu, C. Yuan, B. Duan, M. Huang, X. Luo, Fabrication and thermal characterization of the modularized thermal storage unit. IEEE Trans. Compon., Packag. Manuf. Technol. 6(8), 1198–1207 (2016) 7. A.I.N. Korti, F.Z. Tlemsani, Experimental investigation of latent heat storage in a coil in PCM storage unit. J. Energy Storage 5, 177–186 (2016) 8. P. Sobolˇciak, H. Haneen Abdelrazeq, N.G. Özerkan, M. Ouederni, Z. Nógellová, M.A. AlMaadeed, M. Karkri, I. Krupa, Heat transfer performance of paraffin wax based phase change materials applicable in building industry. Appl. Therm. Eng. (2016) 9. A. Saraswat, R. Bhattacharjee, A. Verma, M.K. Das, S. Khandekar, Investigation of diffusional transport of heat and its enhancement in phase-change thermal energy storage systems. Appl. Therm. Eng. 111, 1611–1621 (2017) 10. P.B. Salunkhe, D.J. Krishna, Investigations on latent heat storage materials for solar water and space heating applications. Journal of Energy Storage 12, 243–260 (2017) 11. D.K. Rabha, P. Muthukumar, Performance studies on a forced convection solar dryer integrated with a paraffin wax–based latent heat storage system. Sol. Energy 149, 214–226 (2017) 12. M.A. Wahid, S.E. Hosseini, H.M. Hussen, H.J. Akeiber, S.N. Saud, A.T. Mohammad, An overview of phase change materials for construction architecture thermal management in hot and dry climate region. Appl. Therm. Eng. 112, 1240–1259 (2017) 13. M.S. Naghavi, K.S. Ong, I.A. Badruddin, M. Mehrali, H.S.C. Metselaar, Thermal performance of a compact design heat pipe solar collector with latent heat storage in charging/discharging modes. Energy 127, 101–115 (2017) 14. B. Németh, A.S. Németh, J. Tóth, A. Fodor-Kardos, J. Gyenis, T. Feczkó, Consolidated microcapsules with double alginate shell containing paraffin for latent heat storage. Sol. Energy Mater. Sol. Cells 143, 397–495 (2015) 15. Z. Khan, Z. Khan, K. Tabeshf, Parametric investigations to enhance thermal performance of paraffin through a novel geometrical configuration of shell and tube latent thermal storage system. Energy Convers. Manag. 127, 355–365 (2016) 16. L. Tana, A. Datea, G. Fernandesa, B. Singh, S. Ganguly, Efficiency gains of photovoltaic system using latent heat thermal energy storage. Energy Procedia 110, 83–88 (2017) 17. H. Niyas, S. Prasad, P. Muthukumar, Performance investigation of a lab–scale latent heat storage prototype–Numerical results. Energy Convers. Manag. 135, 188–199 (2017)

Chapter 3

Performance Analysis of Spiral and Conical Receivers for the Paraboloidal Dish Collector Using CFD Rashmi R. Joshi, Sandeep S. Joshi, Nilesh S. Wakchaure and Akshay C. Suryawanshi Abstract Paraboloidal dish collectors are widely used two-axis tracking collectors for cooking and allied applications. These collectors operate with 40–70% thermal efficiency for medium temperature applications. Conventionally, the cooking vessel is mounted at the focal point of the paraboloidal dish collector and the cooking vessel itself acts as a receiver. In order to improve the performance of these collectors, various receiver designs have been investigated by many researchers in the past. In the current study, the conical and spiral receivers have been suggested for a paraboloidal dish collector of aperture area 1.5625 m2 , concentration ratio of 10. The CFD simulation of the receivers with water as a working fluid is carried out using semi-implicit pressure linked model. The receivers’ geometry is created using SolidWorks software. The system is installed at Nagpur [21° N, 79° E]; the experiments have been performed in the month of March and April. The study shows that, for the paraboloidal dish collector, the spiral receiver performs better as compared to the conical receiver; with different heat transfer liquids and appropriate thermal storage facilities, the system can be used for indoor cooking applications. Keywords Parabolic dish collector · Spiral receiver · Conical receiver

R. R. Joshi (B) · S. S. Joshi Shri Ramdeobaba College of Engineering and Management, Nagpur, Maharashtra, India e-mail: [email protected] S. S. Joshi e-mail: [email protected] N. S. Wakchaure Department of Production Engineering, KKWIEER, Nashik, India A. C. Suryawanshi Knest Manufacturers LLP, Pune, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_3

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R. R. Joshi et al.

Abbreviations AP As ε σ β ʋ Cp L D mw Ta T in T out Ib

Aperture area (m2 ) Absorber/receiver area (m2 ) Emissivity Stefan–Boltzmann constant (W/m2 K4 ) Coefficient of thermal expansion (1/K) Kinematic viscosity (m2 /s) Specific heat of water (KJ/kg-K) Length of coil (m) Span diameter (m) Mass flow rate of water (kg/s) Ambient temperature (°C) Inlet temperature of HTF to coil (°C) Outlet temperature of HTF from coil (°C) Beam solar radiation falling on the collector (W/m2 )

3.1 Introduction Currently, several countries are focusing on the utilization of renewable energy sources to provide a solution to resolve a global warming problem and alleviate the potential of energy crisis. Solar energy is abundantly available source. To utilize the maximum amount of solar energy reaching to the earth surface different solar devices and solar concentrator arrangements are investigated by many researchers in past decades. The solar collectors collect the solar thermal energy at the receivers. The collected energy is extracted by circulating fluid over the receiver and used for various applications. These applications mainly include cooking, industrial and domestic process heating, VAR systems and space heating [1–4]. For the effective use of concentrated solar power and to reduce the time for indoor cooking and other applications, different receiver arrangements are carried out. Recently, many studies have been conducted on various receiver designs, concentrators and working fluids. Daabo et al. [5] analyzed the optical efficiency as well as the flux distribution of the three different geometries: cylindrical, conical and spherical, cavity receivers. The study objective was to analyze their behavior using an advanced ray-tracing method. The optical efficiency reached by the conical receiver is highest. Prakash et al. [6] carried out experimental and numerical studies to analyze convective losses occurring from a downward facing cylindrical cavity receiver of length 0.5 m, internal diameter of 0.3 m and a wind skirt diameter of 0.5 m. The numerical study is performed on fluid inlet temperatures between 50 and 300 °C and receiver inclinations of 0°, 45° and 90° using the Fluent CFD software. Sagade et al. [7] Designed a truncated cone-shaped helical coiled receiver for solar parabolic dish collector made up of copper and their study reports the average rise

3 Performance Analysis of Spiral and Conical …

25

of 67 °C in the receiver temperature. Suman et al. [3] worked on the performance enhancement techniques such as geometrical modifications on the absorber plate, use of solar selective coatings and nanofluids. The modification in surface geometry of the absorber plate by the use of extended surfaces/ribs/corrugation results in enhanced performance. The use of conical cavity receivers in CSP is comparatively new technique. For the solar cookers with paraboloidal collectors, very less literature is available on spiral receiver. In the present study, the performance of spiral and conical receiver is analyzed for a paraboloidal dish collector.

3.2 Numerical Study Geometry of both the receivers are created using SolidWorks [8] software with utilization of sweep command and the dimensions of the both receivers are as follows, length of copper tube is 15 m, diameter of the copper tube is 6 mm and span diameter is 35 cm, which are shown in Figs. 3.1 and 3.2. Meshing is carried out in ICEM CFD [9] and the details of it are as follows, the elemental type of mesh is Quadrilateral at the surfaces and hexahedral faces at the inlet and outlet.

3.2.1 Solution Setup Procedure In solver settings, pressure-based approach is selected for the given problem as pressure-based approach was developed for low-speed incompressible flow. Model selection: The model selected as a laminar model because the Reynolds number is less than 2300 and the laminar numerical model is assumed for Rayleigh number lower than 109 [5].

L= 15m outlet

inlet D = 35cm

Fig. 3.1 Geometry of spiral receiver and meshing

26

R. R. Joshi et al. L= 15m

outlet

H=30cm

inlet D=35cm

Fig. 3.2 Geometry of conical receiver and meshing

Material selection: Material used for fluid flow is considered as liquid (water), and for the receivers’ solid part, copper is used. The material properties used for the simulations are taken from Kothandaraman and Subramanyam [10]. Boundary conditions: (a) Mass flow and temperature given as input boundary condition to water. (b) Pressure output as an outlet boundary condition. (c) Heat flux (W/m2 ) with direct solar radiation and hconv as top wall boundary condition. (d) Heat flux (W/m2 ) with concentration ration and hconv as bottom wall boundary condition. The solutions are obtained by solving the continuity equation, the momentum equation and the energy equation simultaneously. The semi-implicit pressure linked equation (SIMPLE) scheme of the fluent software is used [9]; The SIMPLE scheme is suited for the complicated 3D problem. SIMPLE algorithm has been used, with a first-order upwinding scheme for the discretization of equations. A convergence criterion of 10−3 was imposed on the residuals of the continuity equation, momentum equation. A convergence criterion of 10−6 is considered for energy equation [11–13].

3.3 Experimental Study The schematic diagram of the experimental setup is shown in Fig. 3.3, and specifications are mentioned in Table 3.1. It consists of a paraboloidal dish collector with four segments, each segment of 625 × 625 mm2 , water supply tank, non-return valves fitted in the pipeline to define the flow direction, and control valve used to regulate the flow rate through the circuit. The paraboloidal dish collector is mounted on a stand with the use of nut and bolts as shown in Fig. 3.4. Stand having four wheels at

3 Performance Analysis of Spiral and Conical …

27

Fig. 3.3 Schematic of experimental setup Table 3.1 Specifications of paraboloidal dish collector

Parameter’s

Numerical value

Aperture area

1.5625 m2

Focal length

0.460 m

Reflecting material

Anodized aluminum sheet

Reflectivity of material

80%

Total dish size

1250 × 1250 mm2 2.Copper coil spiral receiver

1. Parabolic dish collector 1 2 Inlet

Outlet to storage tank

3. Copper coil Conical receiver outlet

inlet

Fig. 3.4 Actual photographs of experimental setup

28

R. R. Joshi et al.

the bottom for the movement of collectors. There is provision of shadow sensor nut for the optimum tilt. Water as the working fluid passed through the receivers shown in Fig. 3.4 and get heated due to the concentration of solar radiation and that heat is stored in storage unit and utilized for allied applications.

3.3.1 Analytical Procedure The energy balance equation used for the calculation of convective and radiative losses and the outlet temperature of water. Q in (with concentration) + Q in (without concentration) =Q convection + Q radiation + Q liquid (3.1) Q convection = h × As × (Ts − Ta )

(3.2)

  4 Q radiation = ε × σ × As × Ts4 − Tsurr

(3.3)

Q liquid = m × C p × (Tout − Tin )

(3.4)

The convective heat transfer coefficient k × Nu D

(3.5)

g × β × (Ts − Ta ) × Lc3 × Pr v2

(3.6)

h = Rayleigh number Ra =

The efficiency of the collector is calculated as ηc =

Q liquid Ib×Ap

(3.7)

3.4 Results and Discussion The results are based on the simulations conducted using the model described above. Figure 3.5 represents the CFD simulation analysis for the spiral and conical receivers.

3 Performance Analysis of Spiral and Conical …

29

108 0 C

95 0 C

30 0 C

29 0 C

Fig. 3.5 Counter plots of outlet water temperature of spiral and conical receivers

Maximum obtained temperature values are shown in Table 3.2 for maximum solar intensity (W/m2 ). From the above study, it is proposed that the spiral receiver perform effectively for segmental parabolic dish collector than that of conical receiver. Figures 3.6 and 3.7 represent the results of experimental, analytical and CFD simulation for spiral receiver and conical receiver. The maximum outlet temperature achieved by CFD simulation results and the maximum temperature is achieved at 1.30 p.m. Figure 3.8 shows the comparison of outlet water temperature for spiral and conical receivers over the time of day with CFD simulation results. Table 3.2 Maximum values of various parameters of spiral and conical receivers Parameters Solar intensity

Spiral receiver (W/m2 )

Conical receiver

1257

1211

Ambient temperature (T a °C)

43.4

41.2

Inlet temperature of water (T in °C)

33.4

31.2

Outlet water temperature by experimental results (T out °C)

90.3

81.9

Outlet water temperature by analytical results (T out °C)

110.03

88.07

Outlet water temperature by CFD simulation results (T out °C)

122.6

104.67

The experimental results was validated against analytical results with a maximum % deviation of water outlet temperature

15

7

The CFD simulated work validated against the analytical results with a maximum % deviation of water outlet temperature

11

12

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Fig. 3.6 Outlet water temperature of spiral receiver

Fig. 3.7 Outlet water temperature of conical receiver

As shown in Fig. 3.9, convective heat loss from receiver surface to surrounding increases with temperature difference of surface temperature and ambient temperature. Figure 3.10 shows the Nusselt number relationship with (T s − T a ) and Fig. 3.11 represents the residual plot where solution converges with increase in number of iterations.

3 Performance Analysis of Spiral and Conical …

31

Fig. 3.8 CFD simulation results of outlet water temperature

Fig. 3.9 External convective heat loss (W)

(Degree Celsius)

Fig. 3.10 Nusselt number

(Degree Celsius)

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Fig. 3.11 Residual plot represents the solution convergence

3.5 Conclusions Performance of a paraboloidal dish collector with ‘spiral and conical’ receivers is analyzed using CFD. The experiments are carried out at Nagpur [21° N, 79° E] in the month of March and April. The CFD results show good agreement with the experimental outcomes. • The highest outlet temperature of water is achieved in spiral receiver. • The highest surface temperature at the focal point of 232 °C is achieved in spiral receiver (T a = 42 °C, I g = 1220 W/m2 ). • For the segmental paraboloidal dish collector, the ‘spiral receiver’ is performing better as compared with the ‘conical receiver.’ • The collector efficiency of 59.23% and 45% is recorded in spiral and conical receivers, respectively. • The heat losses from the ‘conical receiver’ are more as compared to ‘spiral receiver.’ • The ‘spiral receiver’ of CFD model is suitable for the analysis of segmental paraboloidal dish collector. The study shows that, for the paraboloidal dish collector, the spiral receiver performs better; with different heat transfer liquids, the system can be used for indoor cooking applications.

References 1. S. Skouri et al., Design and construction of sun tracking systems for solar parabolic concentrator displacement. J. Renew. Sustain. Energy, Rev. 60, 1419–1429 (2016) 2. F. Kreith, J.F. Kreider, Principles of Solar Engineering, 2nd edn. (McGraw-Hill, New York, 1978)

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3. S. Suman et al. Performance enhancement of solar collectors-A review. J. Renew. Sustain. Energy Rev. 49, 192–210 (2015) 4. S.P. Sukhatme, Solar Energy, Principle of Thermal collection & Storage, 3rd edn. (McGraw Hill Education, New Delhi, India, 2014) 5. A.M. Daabo et al., The effect of receiver geometry on the optical performance of a small scale solar cavity receiver for parabolic dish applications. Energy, Elsevier 114, 513–525 (2016) 6. M. Prakash, S.B. Kedare, J.K. Nayak, Investigations on heat losses from a solar cavity receiver. Solar Energy, 83, 157–170 (2008) 7. A.A. Sagade, N.N. Shinde, Experimental analysis of effect of variation of convection heat losses on performance of parabolic dish collector with nickel chrome coated receiver. Sustainable technologies world congress IEEE, 107–111 (2011) 8. SOLIDWORK software 2013 copyright 9. Fluent–Inc., 2015, FLUENT 15.1 User’s Guide, FLUENT. Inc, FLUENT 15.1 10. C.P. Kothandaraman, S. Subramanyam, Heat and Mass Transfer Data Book, 4th edn. (New Age International Publishers, New Delhi, 1998) 11. S. Paitoonsurikarn et al., Numerical investigation of natural convection loss from cavity receivers in solar dish applications. J. Sol. Energy Eng., ASME, 133 (2011) 12. M. Wang et al., The impact of geometrical parameters on the thermal performance of a solar receiver of dish-type concentrated solar energy system. J. Renew. Energy 35, 2501–2513 (2010) 13. K.S. Reddy et al., Effect of wind speed and direction on convective heat losses from solar parabolic dish modified cavity receiver. J. Sol. Energy 131, 183–198 (2016)

Chapter 4

Performance Analysis of Phase Change Material Storage System for Solar Thermal Applications Sneha Murali , R. P. Saini

and Ambuj Punia

Abstract This paper focuses on the storage of solar thermal energy in a shell and tube type latent heat storage system. Paraffin wax was used as phase change material to store heat and water as the heat transfer fluid in the experimental study. Firstly, the performance of the simple shell and tube system has been analyzed based on parameters, such as the charging time, cumulative energy stored and charging and system efficiency of the heat storage unit. Influence of factors such as the inlet temperature and mass flow rate of the heat transfer fluid on the performance of the system has been studied. A better performance was observed at higher inlet temperature and mass flow rate. In the second part of the experimentation, longitudinal fins were used to enhance the heat transfer rate and a comparative study on the performance of the system was carried out. Results show a drastic reduction in the charging time and an increase in charging efficiency by about five times, by the use of fins. Total energy stored is higher with the use of fins, especially at higher inlet temperature of the HTF and higher system efficiency was obtained with the use of fins. Keywords Solar thermal · Phase change material · Latent heat · Thermal energy storage · Renewable energy

Abbreviations PCM Phase change material HTF Heat transfer fluid LHTESS Latent heat thermal energy storage system

S. Murali (B) · R. P. Saini · A. Punia Alternate Hydro Energy Centre, Indian Institute of Technology, Roorkee, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_4

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4.1 Introduction Energy usage has increased tremendously over the years, which has increased the demand for energy. Excessive usage of fossil fuels has resulted in damage to the ecosystem in addition to the depletion of the resources. It is high time to move toward clean and renewable energy sources such as solar energy. Due to the intermittent nature of solar energy, there arises a need to store the available energy, so that it can be used to meet the demand during times when it is not available. In a latent heat thermal energy storage system (LHTESS), energy is stored and released at an almost constant temperature and the density of energy storage is much higher than that for sensible heat storage [1]. One of the most important aspects in the design of a LHTESS is the selection of the phase change material (PCM). It depends on the application where it is to be used, which is used to decide the melting temperature range of the PCM to be selected. Numerous PCM are available and they are broadly categorized as organics, inorganics and eutectics. The performance of the PCM storage system is influenced by various geometrical as well as thermal parameters of the heat storage unit and the PCM that stores the thermal energy. The direction of flow of the HTF, the shell configuration and dimensions, number of HTF tubes, its geometry and dimension [2–6] affect the performance. Studies have also been carried out on the influence of thermophysical properties of the PCM [7, 8]. Further, owing to the lower thermal conductivity of the PCM, there arises a need to enhance the heat transfer rate from the HTF to the PCM. Finned HTF pipes and different fin configuration were studied [9, 10]. Other enhancement techniques studied are the use of foam, expanded graphite, encapsulation, nano-enhanced PCMs and multiple PCM’s and direct contact method [11–17]. Based on the literature review, it is found that in most of the studies, the basis for measuring performance has not been defined clearly. The experimental studies carried out were not focused on the charging efficiency which is also one of the important parameters to measure performance. The present experimental study is carried out to investigate the performance of a PCM storage on the basis of the charging time, the charging efficiency, the energy stored and released. Further, the effect of parameters like the inlet temperature and mass flow rate of the HTF on performance is also carried out. Investigations have been carried out to find the effect of longitudinal fins on the performance of the heat storage system.

4.2 Methodology Experimental studies were carried out with water as the HTF and commercial grade paraffin wax as the PCM.

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37

4.2.1 Heat Storage Unit Considered for the Experiment The heat storage unit consisted of a cylindrical-shaped outer shell of mild steel, 590 mm in length and 110 mm inner diameter, open at both ends to accommodate the PCM. Two windows were provided on the shell to visualize the melting of the PCM. Provision was also made for pouring the molten PCM into the shell. The inner tube assembly had an upper plenum chamber 100 mm in diameter onto which the copper HTF tube of 29 mm diameter, threaded at the ends was welded. This assembly could be inserted into the shell from the top. The lower plenum chamber was a flanged cylindrical chamber with a hole at the center, through which the HTF tube goes in and was tightened with a nut from the bottom. The upper and the lower cover plates cover the upper and the lower plenum chamber, respectively, with a rubber gasket in between to prevent leakage. These were tightened to the shell. Later fins, 330 mm long, 30 mm wide and 1 mm thick were welded onto the HTF tube.

4.2.2 PCM Used for the Study Commercial grade paraffin wax was chosen as the PCM for the study. Its properties were studied by carrying out DSC testing, while some of its properties and that of the HTF were obtained from the literature available [18–20]. The DSC curve obtained is shown in Fig. 4.1 and the properties of paraffin wax and water are shown in Tables 4.1 and 4.2, respectively.

Fig. 4.1 DSC curve of paraffin wax obtained when heating at 5 °C/min

38 Table 4.1 Properties of paraffin wax

Table 4.2 Properties of water [18–20]

S. Murali et al. Property

Value obtained

Melting temperature

54.9 °C

Latent heat

176 kJ/kg

Density

845.4 kg/m3

Specific heat

1.92 kJ/kg K

Thermal conductivity

0.18 W/m-K

Property

Values

Density

997 kg/s

Specific heat capacity at 90 °C

4.208 kJ/kg K

Specific heat capacity at 80 °C

4.198 kJ/kg K

Specific heat capacity at 70 °C

4.191 kJ/kg K

4.2.3 Experimental Setup and Procedure The schematic diagram and the actual photograph of the experimental setup are as shown in Figs. 4.2 and 4.3, respectively. The setup consisted of the heat storage system, T-type thermocouples, data logger connected to a computer, water heater, storage tanks for water, PID controller, rotameter, control valves and pump. The heat storage system was a heat exchanger used for storing and retrieving heat. The PCM was filled inside the shell of the storage unit through the opening provided. The HTF was allowed to flow through the central tube from the upper plenum chamber. The

Fig. 4.2 Schematic diagram of the experimental setup

4 Performance Analysis of Phase Change …

39

Fig. 4.3 Photograph of the experimental setup

outer surface of the heat storage unit was insulated with the help of two layers of foam which is 5 mm thick (one layer). The temperature of the PCM and the inlet and outlet temperature of the HTF were measured by using T-type thermocouples, which calibrated using Fluke 9142 Field Meteorological Well. An error of about 2.23% was observed in the working temperature range. A total of 12 thermocouples were used inside the heat storage unit, while 2 of them were used to measure the inlet and outlet temperature of the HTF. The axial and radial locations of these thermocouples are depicted in Fig. 4.4. The thermocouples were connected to the data logger, which in turn was connected to a computer. Computer software was used to record and display the data. Before the start of the experiment, leakage testing of the heat storage unit was carried out by pouring water into the shell and circulating water through the HTF tube. Trial runs were carried out. The charging cycle involves the transfer of heat from the HTF to the PCM. It began when the entire PCM was in the solid state and approximately at a constant temperature. The HTF was heated in the storage tank and the circulation of the HTF was started when the required inlet temperature was reached. During the charging cycle valves, V3 and V4 were kept closed, while the others were kept open. At a constant inlet temperature mass flow rate was varied by adjusting the bypass valve V5 and at a constant mass flow rate, inlet temperature was varied. The data logger recorded temperature at an interval of 1 min. The charging cycle was completed after the entire PCM melted. The discharge cycle involves the transfer of heat from the PCM to the HTF and was started immediately after the charging cycle was completed.

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S. Murali et al.

(a)

(b) A B C

3

1

D 2

Fig. 4.4 a Axial positions of thermocouples and b radial positions of thermocouples

4.3 Results and Discussions 4.3.1 Charging Cycle The charging cycle was carried out at inlet temperatures of 90, 80 and 70 °C and mass flow rate of 0.085 and 0.068 kg/s. Cumulative Energy Stored Influence of inlet temperature of the HTF. The cumulative energy stored in the PCM remained almost the same with slightly higher energy stored for a higher inlet temperature but as time progressed, higher amount of energy was stored when the inlet temperature was higher. Cumulative energy decreased in the order of HTF inlet temperature, considered as 90 °C, 80 °C and 70 °C. Without fins, a total of heat of 702.85 kJ, 130 kJ and 104.5 kJ were stored at 90 °C, 80 °C and 70 °C with a charging efficiency of 2.69%, 0.639% and 0.52%, respectively, whereas with fins a total of 723.01 kJ, 708.59 kJ and 551.34 kJ were stored at 90 °C, 80 °C and 70 °C with a charging efficiency of 16.34%, 16.6% and 20.89%, respectively. The influence of inlet temperature of the HTF on the cumulative energy stored is shown in Fig. 4.5. Influence of mass flow rate of the HTF. The influence of mass flow rate of the PCM was more prominent when fins were used than for a simple shell and tube type. For both the cases, the cumulative energy stored was a little higher at a lower mass initially, as the HTF had more time to transfer its energy to the PCM, but toward the end of the charging cycle cumulative energy stored was higher at a higher mass flow rate. For the shell and tube type, the cumulative energy stored was 727.23 kJ and 702.85 kJ and with fins it was 749.6 kJ and 723.01 kJ, when mass flow rates were

4 Performance Analysis of Phase Change …

(b) Cumulative energy stored in the PCM in kJ

Cummulative energy stored in the PCM in kJ

(a)

41

250 90

70

80

200 150 100 50 0

0

50

100

150

250 90

70

150 100 50 0

200

80

200

0

20

Time in minute

40

60

80

100

Time in minute

Fig. 4.5 Influence of inlet temperature of HTF on the cumulative energy stored for a shell and tube and b with fins

(b) 250

0.085 kg/s

Cummulative energy stored in the PCM in kJ

Cummulative energy stored in the PCM in kJ

(a) 0.068 kg/s

200 150 100 50 0

0

50

100

Time in minute

150

200

250

0.085 kg/s

0.068 kg/s

200 150 100 50 0

0

50

100

Time in minute

Fig. 4.6 Influence of mass flow rate of HTF on the cumulative energy stored for a shell and tube and b with fins

0.085 kg/s and 0.068 kg/s, respectively. The influence of mass flow rate of the HTF on the cumulative energy stored is shown in Fig. 4.6. Influence of fins. A higher amount of energy was stored in the PCM when fins were used in the system as fins enhanced the rate of heat transfer from the HTF. Around 749.6 kJ of heat was stored with fins and 727.23 kJ was stored without the use of fins. The influence of fins on the cumulative energy stored is shown in Fig. 4.7. Charging Time Influence of inlet temperature of the HTF. The influence of inlet temperature of the HTF on charging time is shown in Fig. 4.8. For both the cases, charging time was found to be lesser when the inlet temperature of the HTF was higher. For the simple shell and tube type without fins, melting temperature was not reached when the inlet temperature was 70 °C and 80 °C even after 164 min and 199 min, respectively, while the entire PCM melted when the HTF inlet temperature was 90 °C after a 246 min. The PCM reached the melting temperature range when fins were used at all three inlet temperature. It took around 198 min, 165 min and about 96 min for charging to complete with an inlet temperature of 70 °C, 80 °C and 90 °C, respectively.

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Cummulative energy stored in kJ

250

with fin

200 150 100 50 0

0

20

40

60

80

100

Time in minute Fig. 4.7 Influence of fins on cumulative energy stored

(a)

(b) 90

80

70

Charging time in minute

Charging time in minute

300 250 200 150 100 50 0 0.00

20.00

40.00

60.00

PCM temperature in ◦C

80.00

250 90

80

70

200 150 100 50 0 0.00

20.00

40.00

60.00

80.00

100.00

PCM Temperature in ◦C

Fig. 4.8 Influence of inlet temperature of HTF on charging time for a shell and tube and b with fins

Influence of mass flow rate of the HTF. The influence of the mass flow rate of the HTF on the charging time for the shell and tube and with fins is shown in Fig. 4.9. The influence of mass flow rate of the HTF was more profound for the shell and tube type as it took 195 min to finish charging at 0.085 kg/s, while it took 246 min when it was 0.068 kg/s. For the case with fins, charging finished at almost the same time for both the mass flow rates, at 94 min and 96 min for a mass flow rate of 0.085 kg/s and 0.068 kg/s, respectively. Influence of fins. The influence of fins on the charging time is shown in Fig. 4.10. The charging time reduced by more than half when fins were used. It took 94 min for charging cycle with fins compared to 195 min without the use of fins.

4 Performance Analysis of Phase Change …

(b)

250 0.085 kg/s

0.068 kg/s

Charging time in minute

Charging time in minute

(a)

43

200 150 100 50 0 0.00

20.00

40.00

60.00

80.00

100 90 80 70 60 50 40 30 20 10 0 0.00

PCM temperature in ◦C

0.085 kg/s

20.00

0.068 kg/s

40.00

60.00

80.00

100.00

PCM Temperature in ◦C

Fig. 4.9 Influence of mass flow rate of the HTF on charging time for a shell and tube and b with fins

Charging time in minute

250

without fin

with fins

200 150 100 50 0 0.00

20.00

40.00

60.00

80.00

100.00

PCM temperature in ◦C

Fig. 4.10 Influence of fins on charging time

4.3.2 Discharging Cycle It was started immediately after the charging cycle was completed. HTF at ambient temperature was circulated at mass flow rate of 0.085 kg/s and 0.068 kg/s. The discharge cycle was completed after a steady state was reached. Cumulative Energy Released Influence of mass flow rate of the HTF. The influence of the mass flow rate of the HTF on the cumulative energy released by the PCM for shell and tube and with fins is shown in Fig. 4.11. Mass flow rate affected the cumulative energy released by the PCM in both the cases, but it was more for the shell and tube case than with fins. The cumulative heat energy released by the PCM was 643.62 and 671.7 kJ for shell and tube and 699.86 and 701.15 kJ with fins at mass flow rate of 0.068 and 0.085 kg/s. Influence of fins. The influence of fins on the cumulative energy released by the PCM is shown in Fig. 4.12. The cumulative energy discharged was higher with the

S. Murali et al. 200 180 160 140 120 100 80 60 40 20 0

0.085 kg/s 0.068 kg/s

0

50

100

150

Cummulative energy released by the PCM in kJ

Cummulative Energy released by the PCM in kJ

44 250 200 150 100

0.085 kg/s

50 0

0.068 kg/s 0

50

100

150

200

Time in minute

Time in minute

Fig. 4.11 Influence of mass flow rate of HTF on cumulative energy released for a shell and tube, b with fins 250

Cumulative energy discharged by the PCM in kJ

Fig. 4.12 Influence of fins on cumulative energy released

200 150 100

without fins 50 0

with fins 0

50

100

150

Time in minute

use of fins, but the influence of fins was not much prominent during discharging cycle compared to the charging cycle.

4.3.3 Charging and System Efficiency The charging and system efficiency was influenced by parameters like the inlet temperature and mass flow rate of the HTF as well with the use of fins. The charging and system efficiency for the different cases are summarized in Tables 4.3 and 4.4, respectively.

4 Performance Analysis of Phase Change … Table 4.3 Charging efficiency for the cases investigated

Table 4.4 System efficiency for the cases investigated

45

Inlet temperature and mass flow rate

Charging efficiency without fins

Charging efficiency with fins

90 °C −0.085 kg/s

3.17

20.43

90 °C –0.068 kg/s

2.69

16.34

80 °C –0.068 kg/s

0.639

16.6

70 °C –0.068 kg/s

0.52

20.89

Inlet temperature and mass flow rate

System efficiency without fins

System efficiency with fins

90 °C –0.085 kg/s

48.67

69.16

90 °C –0.068 kg/s

27.39

83.7

80 °C –0.068 kg/s

24.64

54.93

70 °C –0.068 kg/s

18.89

82.04

4.4 Conclusions 1. In all the cases under the present investigation, during the charging cycle, temperature of the topmost layer increased and reached a maximum value close to 75 °C after the melting temperature range due to effect of natural convection, followed by the bottom layer because of the heat transfer from the hot HTF being stored in the lower plenum chamber. 2. The PCM layer close to the HTF tube melted first, followed by the adjacent layers radially. At the end of the charging cycle, all the radial positions reached the same temperature in one axial position. 3. The inlet temperature of the HTF is found to be an important parameter which influenced the performance of the heat storage system. A higher inlet temperature enhanced the performance in terms of higher energy storage (723 kJ at 90 °C and 708.59 at 80 °C), less charging time (195 min at 90 °C and 246 min at 80 °C) and higher system efficiency (27.39% at 90 °C and 24.64% at 80 °C). 4. A higher mass flow rate also improved the performance but its influence was not as prominent as the inlet temperature of the HTF (charging time: 94 min at 0.085 kg/s and 96 min at 0.068 kg/s). 5. Use of fins significantly improved the performance of the heat storage system. The charging time and system efficiency is increased drastically when fins were incorporated. Its effect was more pronounced during the charging cycle than in the discharging cycle. From the experimental results obtained, the charging time without fins is found to be 195 min, whereas it is 94 min with fins. Charging efficiency increased to 20.43% with fins, while it was only 3.17% without fins. System efficiency also is found to be higher with fins as 69.16% than without fins as 48.67%.

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6. The results obtained from the study may be helpful to investigate further, the PCM-based solar thermal storage system for higher temperature requirement.

References 1. A. Sharma, V.V. Tyagi, C.R. Chen, D. Buddhi, Review on thermal energy storage with phase change materials and applications. Renew. Sustain. Energy Rev. 13, 318–345 (2009). https:// doi.org/10.1016/j.rser.2007.10.005 2. Z. Gong, A.S. Mujumdar, Finite-element analysis of cyclic heat transfer in a shell and tube latent heat energy storage exchanger. Appl. Therm. Eng. 17(4), 583–591 (1997). https://doi. org/10.1016/S1359-4311(96)00054-3 3. B. Zivkovic, I. Fujii, An analysis of isothermal phase change of phase change material within rectangular and cylindrical containers. Sol. Energy 70(1), 51–61 (2001). https://doi.org/10. 1016/S0038-092X(00)00112-2 4. W.W. Wang, L.B. Wang, Y.L. He, The energy efficiency ratio of heat storage in one shell-andone tube phase change thermal energy storage unit. Appl. Energy 138, 169–182 (2015). https:// doi.org/10.1016/j.apenergy.2014.10.064 5. M. Esapour, M.J. Hosseini, A.A. Ranjbar, Y. Pahamli, R. Bahrampoury, Phase change in multitube heat exchangers. Renew. Energy 85, 1017–1025 (2016). https://doi.org/10.1016/j.renene. 2015.07.063 6. Y. Pahamli, M.J. Hosseini, A.A. Ranjbar, R. Bahrampoury, Analysis of the effect of eccentricity and operational parameters in PCM-filled single-pass shell and tube heat exchanger. Renew. Energy 97, 344–357 (2016). https://doi.org/10.1016/j.renene.2016.05.090 7. Y.B. Tao, V.P. Carey, Effects of PCM thermophysical properties on thermal storage performance of a shell-and-tube latent heat storage unit. Appl. Energy 179, 203–210 (2016). https://doi.org/ 10.1016/j.apenergy.2016.06.140 8. D. Cano, C. Funéz, L. Rodriguez, J.L. Valverde, L.S. Silva, Experimental investigation of a thermal storage system using phase change materials. Appl. Therm. Eng. 107, 264–270 (2016). https://doi.org/10.1016/j.applthermaleng.2016.06.169 9. A.A.A. Abduljalil, S. Mat, K. Sopian, M.Y. Sulaiman, A.T. Mohammad, Internal and external fin heat transfer enhancement technique for latent heat thermal energy storage in triplex tube heat exchangers. Appl. Therm. Eng. 53, 147–156 (2013). https://doi.org/10.1016/j. applthermaleng.2013.01.011 10. S. Tiari, S. Qiu, M. Mahdavi, Discharging process of a finned heat pipe–assisted thermal energy storage system with high temperature phase change material. Energy Convers. Manag. 118, 426–437 (2016). https://doi.org/10.1016/j.enconman.2016.04.025 11. M. Martinelli, F. Bentivoglio, A.C. Soupart, R. Couturier, J.F. Fourmigue, P. Marty, Experimental study of a phase change thermal energy storage with copper foam. Appl. Therm. Eng. 101, 247–261 (2016). https://doi.org/10.1016/j.applthermaleng.2016.02.095 12. S. Pincemin, R. Olives, X. Py, M. Christ, Highly conductive composites made of phase change materials and graphite for thermal storage. Sol. Energy Mater. Sol. Cells 92, 603–613 (2008). https://doi.org/10.1016/j.solmat.2007.11.010 13. N. Soares, A.R. Gaspar, P. Santos, J.J. Costa, Experimental study of the heat transfer through a vertical stack of rectangular cavities filled with phase change materials. Appl. Energy 142, 192–205 (2015). https://doi.org/10.1016/j.apenergy.2014.12.034 14. T.T. Ping, Y.C. Chieh, Characteristics of phase-change materials containing oxide nanoadditives for thermal storage. Nanoscale Res. Lett. 7(1), 611 (2012). https://doi.org/10.1186/ 1556-276X-7-611 15. K.A. Taha, M.M. Rahman, Comparison between the single-PCM and multi-PCM thermal energy storage design. Energy Convers. Manag. 83, 79–87 (2014). https://doi.org/10.1016/j. enconman.2014.03.047

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16. M. Fang, G. Chen, Effects of different multiple PCMs on the performance of a latent thermal energy storage system. Appl. Therm. Eng. 27(994), 1000 (2007). https://doi.org/10.1016/j. applthermaleng.2006.08.001 17. W. Wang, S. He, S. Guo, J. Yan, J. Ding, A combined experimental and simulation study on charging process of Erythritol–HTO direct-blending based energy storage system. Energy Conserv. Manag. 83, 306–313 (2014). https://doi.org/10.1016/j.enconman.2014.03.054 18. A. Mete, M.Y. Yazici, Experimental study of thermal energy storage characteristics of a paraffin in a horizontal tube-in-shell storage unit. Energy Conserv. Manag. 73, 271–277 (2013). https:// doi.org/10.1016/j.enconman.2013.04.030 19. K.D. Reddy, P. Venkataramaiah, T.R. Lokesh, Parametric study on phase change material based thermal energy storage system. Energy Power Eng. 6, 537–549 (2014). https://doi.org/10.4236/ epe.2014.614047 20. M.K. Rathod, J. Banerjee, Thermal performance enhancement of shell and tube latent heat storage unit using longitudinal fins. Appl. Therm. Eng. 75, 1084–1092 (2015). https://doi.org/ 10.1016/j.applthermaleng.2014.10.074

Chapter 5

Exergy and Energy Analysis of a Packed Bed Thermal Energy Storage System with Different Heat Transfer Fluids Ambuj Punia , R. P. Saini

and Sneha Murali

Abstract The requirement for energy is increasing at a very high rate, the fulfilment of which is nearly impossible with the help of fossil fuels or conventional sources of energy. Also, the heavy exploitation of fossil fuels has become a major contributor to global warming. Renewable energy sources can help to mitigate this situation. But being intermittent in nature, there arise a need to store them. This paper focuses on the storage of solar thermal energy. In the present study, exergy and energy evaluation of a packed bed solar thermal energy storage using different heat transfer fluids, namely air, water and oil has been carried out. From the experimental investigation, it has been observed that the average exergy and energy efficiency when air is used as the heat transfer fluid are better than when other two fluids are used. While for an individual case of charging and discharging, water shows a sharp temperature profile followed by oil and air. The overall exergy or second law efficiency and total energy efficiency for the system have been tabulated. Further, the future scope of study has also been given which can be done with the help of different heat storage materials and innovative heat transfer fluids. Keywords Solar thermal · Packed bed · Sensible heat storage · Exergy and energy analysis · Renewable energy

5.1 Introduction Packed bed energy storage system is an efficient way to store energy from the sun in the form of heat. The thermal energy stored can be utilized for various applications where heat energy is required as well as for electricity generation. Packed bed thermal energy storage system generally consists of a storage tank filled with storage material. The hot fluid after taking the heat energy from the sun goes inside the packed bed, thereby delivering energy in the form of heat to the storage material. This process is the charging cycle. In the discharging cycle, cold fluid is passed through the packed A. Punia (B) · R. P. Saini · S. Murali Alternate Hydro Energy Centre, Indian Institute of Technology Roorkee, Roorkee, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_5

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Fig. 5.1 Schematic of a packed bed energy storage system [1]

bed, and heat retrieval from the storage material takes place. The heat extracted is then delivered to the required space for further use. A generally used storage material is concrete, because of its cheapness and easy availability. The fluid which gives and takes away the heat to and from the storage system is called the heat transfer fluid (HTF). Schematic of the packed bed thermal storage system is shown in Fig. 5.1. Various studies have been conducted on packed bed thermal energy storage system taking into account various parameters. Zanganeh et al. [2] designed a 100 MWhth thermal energy storage in which they used rocks as the storage material and air as the heat transfer fluid. Initially, they built a pilot-scale model of 6.5 MWhth and tested it experimentally. The experimental results were then used to design the prototype. They stated that the overall thermal efficiency was 89% after steady state conditions were reached. Allen et al. [3] fixed the Biot number and established a cost optimum method to establish the particle size and bed length of a rock bed. Al-Nimr et al. [4] developed a mathematical model in which dynamic thermal behaviour of packed bed energy storage system was investigated, and the analytical solutions were validated experimentally by investigating different parameters axially along the bed such as the inlet fluid temperature, mass flow rate and void fraction. Hughes et al. [5] using gravel as the storage material and air as the heat transfer fluid developed correlation to measure the performance of the system. It was found that the system performance was found to be insensitive for NTU > 10. Apart from the literature given above, various investigations were also done to evaluate the performance of the packed bed thermal energy storage system. Choudhury and Garg [6] evaluated system performance for different volume flow rate of air and water, and they found that the inclusion of rock bed resulted in 11 °C higher temperature of the storage tank. Chaiet et al. [7] used exergy and energy analysis as the performance measure to evaluate the efficiency of the storage system. They found that the flow direction plays an important role in determining the exergy and energy efficiencies of the storage system. Sagara and Nakahara [8] with their theoretical model analysed the large-sized storage elements with small-sized ones. Lesser frictional loss in the case of large-sized

5 Exergy and Energy Analysis of a Packed …

51

elements leads to the performance measure of the storage system as less pumping power was required to force air through the storage tank. Mawireet et al. [9] developed a single-phase model to evaluate the thermal performance of various storage materials, in which, alumina in terms of energy and exergy stored and fused silica glass in terms of stratification was found to be the best material. From the literature study, it was found that although various studies have been conducted on packed bed thermal energy storage system, the exergy and energy analysis for medium-sized storage material has not been done comprehensively. The next sections discuss the exergy and energy analysis carried out on the storage system with the three different heat transfer fluids, namely air, water and oil.

5.2 Experimental Set-up The experimental set-up used for the investigation is shown in Fig. 5.2. The storage tank consists of a steel cylindrical tank of dimensions 300 mm diameter and 600 mm length inside which the crushed stone with a maximum size of 3 cm diameter is filled randomly. Two storage tanks were used to store hot and cold fluids during charging and discharging, respectively. Apart from this, a valve arrangement was made so that cold and hot fluids do not mix during different cycles. A pump in case of water and oil and a blower in case of air were used to pass the fluid through the storage tank. Eighteen T-type thermocouples were used which were placed at different axial and

Fig. 5.2 Final assembly of the experimental set-up

Reference and Thermocouple temperature(°C)

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120

T1

100

T2 T3

80

T4 T5

60

T6 T7

40

T8

20

T9 T10

0 0

20

40 60 80 Set Temperature(°C)

100

120

Fig. 5.3 Calibration curve showing the thermocouples

radial positions. The calibration of the thermocouples was done up to 120 °C with an accuracy of 0.1% deviation. The calibration curve for the thermocouples is shown in Fig. 5.3. A 5 mm thickness of insulation was used to reduce the heat losses to the surrounding to a minimum. Heat transfer fluid plays an important role in delivering the energy it gets from the sun. Under the present work, three heat transfer fluids have been identified, namely air, water and oil. Selection of air has been done on the basis of its easy availability and stability at high temperatures, while water and oil have been selected on the basis of high thermal capacity. Table 5.1 gives various properties of the heat transfer fluids which were used for the study. A heat coil (adjustable to 1.5–3 KW) was used to heat water and oil, while for the case of air, a separate air heater of 2 KW was fabricated to maintain the desired flow rate and temperature. Two different mass flow rates (0.033 kg/s and 0.083 kg/s) at two different inlet temperatures were used for all the three heat transfer fluids. Table 5.1 Properties of various heat transfer fluids used for the study [10] Property

Air

Water

Oil

Density (kg/m3 )

1.226

1000

958

Specific heat capacity (kJ/kg K)

1.005

4.186

1.52

Thermal conductivity (W/mK)

0.0314

679.1

0.128

Kinematic viscosity (×10−6 m2 /s)

23.06

0.294

0.986

5 Exergy and Energy Analysis of a Packed …

53

5.3 Procedure Under the present study, the exergy and energy analysis of a packed bed energy storage system is evaluated and compared with three different heat transfer fluids. For the present study, it is assumed that there is only one-dimensional heat transfer in the axial direction. The equations used to calculate the exergy and energy during charging and discharging are discussed below.

5.3.1 Energy Analysis During Charging In the current section, the energy given by the hot incoming fluid to the storage material and the energy stored in the storage material is discussed. Energy given by the hot incoming fluid can be evaluated by the following equation: ⎡ Q in = m˙ f Cpf ⎣

tin

tin dtin −

tout

⎤ dtout ⎦

(5.1)

tout

where m ˙ f is the mass flow rate of incoming fluid. Cpf is the specific heat of incoming fluid. Energy stored in the storage material after charging was completed using the following equation: Q st = m st Cpst [t2 − t1 ]

(5.2)

where mst C pst t2 t1

is the mass of the storage material. is the specific heat of the storage material. is the final temperature of the storage material. is the initial temperature of the storage material (generally taken as 25 °C).

Energy efficiency of the storage system during charging is defined as follows: ηst,c =

Energy stored in the storage system Energy given to the storage system

(5.3)

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5.3.2 Exergy Analysis During Charging Exergy is the available or useful energy which can be used to measure the actual performance of the system. Exergy given by the hot incoming fluid to the storage system during charging is expressed as follows: ψf,c

  tin,m = m˙ f Cpf tin,m − tout,m − 298 ln tout,m

(5.4)

where m ˙f C pf t in,m t out,m

is the mass flow rate of incoming fluid. is the specific heat of incoming fluid. is the mean inlet temperature of the incoming fluid. is the mean outlet temperature of the outgoing fluid.

Exergy stored in the storage material after charging is expressed as follows:   t1 ψs,c = m˙ s Cps t1 − t2 − 298 ln t2

(5.5)

where mst C pst t2 t1

is the mass of the storage material. is the specific heat of the storage material. is the final temperature of the storage material. is the initial temperature of the storage material (taken as 25 °C).

Exergy efficiency during charging is represented by ηex,c =

Energy stored in the storage material after charging Energy given by the hot fluid during charging

(5.6)

In a similar way, the equations for the discharge cycle can also be evaluated and the corresponding discharge efficiencies can be calculated.

5.4 Results The results obtained from the calculations carried out from the experimental data are discussed and tabulated. For one cycle of charging and discharging, exergy efficiency or second law efficiency has been calculated based on the calculation of exergic efficiencies for the whole cycle. Furthermore, the same is repeated for the energy efficiency of the system for all the three heat transfer fluids used. Table 5.2 gives the overall energy analysis of the system, whereas Table 5.2 gives the overall exergy analysis data for the whole system taking into account all the three heat transfer fluids.

5 Exergy and Energy Analysis of a Packed …

55

Table 5.2 Overall energy and exergy analysis for the packed bed storage system S. no.

Inlet temperature (°C)

Mass flow rate (kg/s)

Energy efficiency during charging (%)

Energy efficiency during discharging (%)

Energy efficiency of system (%)

Exergy efficiency during charging (%)

Exergy efficiency during discharging (%)

Exergy efficiency of system (%)

1

90

0.0105

54.51

53.15

28.97

40.20

26.88

10.8

2

75

0.0105

71.59

61.08

43.72

55.36

36.19

20.03

3

90

0.0136

59.65

57.11

34.06

44.97

27.52

12.26

4

75

0.0136

60

59.94

35.96

56.41

36.01

20.31

1

90

0.083

51.78

64.4

33.34

28.65

91.48

26.2

2

75

0.083

74.86

96.1

71.94

19.89

37.58

7.47

3

90

0.033

60.19

93.98

56.56

36.89

64.15

23.66

4

75

0.033

49.12

91.03

44.71

27.97

56.31

15.74

1

90

0.083

47.92

66.16

31.7

29.3

57.14

16.74

2

75

0.083

47.88

86.97

41.64

31.67

74.15

23.48

3

90

0.033

43.64

87.5

38.18

31.54

66.53

20.98

4

75

0.033

50.33

69.63

35.04

31.22

74.28

23.19

Air

Water

Oil

Figures 5.4, 5.5, 5.6 and 5.7 show the effect of inlet temperature on exergy and energy efficiency during charging and discharging. Air and water again were found to be the better HTF during charging and discharging exergy and energy analysis. air

60

water

oil

Exergy Efficiency(%)

50

40

30

20

10

0

0

20

40

60

80

100

Inlet temperature(°C)

Fig. 5.4 Exergy efficiency versus inlet temperature plot during charging

120

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80

water

Oil

Energy Efficiency(%)

70 60 50 40 30 20 10 0

0

20

40

60

80

100

120

Inlet Temperature(°C)

Fig. 5.5 Energy efficiency versus inlet temperature plot during charging air

100

water

oil

Exergy Efficiency(%)

90 80 70 60 50 40 30 20 10 0

0

5

10

15

20

25

30

35

40

Inlet temperature(°C) Fig. 5.6 Exergy efficiency versus inlet temperature plot during discharging

From second law efficiency data, it can be concluded that although water has the highest exergic performance for the packed bed thermal energy storage system, average exergy efficiency of the system during discharging is higher in case of oil which lies in the range of 65–70%, while for water it is reduced to 60–65% and for air it is further reduced to 30–35%. Thus, from the average exergic performance analysis, oil proves to be the best among the three heat transfer fluids. From Table 5.2, it is found that average energy efficiency is better in case of air (60–65%), which can be the case of varying loads, followed by water (55–60%) and

5 Exergy and Energy Analysis of a Packed … air

120

57 water

oil

Energy Efficiency(%)

100 80 60 40 20 0

0

5

10

15

20

25

30

35

40

Inlet Temperature(°C)

Fig. 5.7 Energy efficiency versus inlet temperature plot during discharging

oil (45–50%), respectively. From average energy efficiency point of view, thus air is the better heat transfer fluid.

5.5 Conclusions From the experimental study carried out on the packed bed thermal energy storage system, it can be concluded that for the temperature limit up to 100 °C, water is found to be the better heat transfer fluid due to its highest specific heat and higher density among the heat transfer fluids considered, i.e. air and oil. The following are the specific findings of the experimental study conducted. • The average energy efficiency of the system during charging is found to be better in case of air which ranges from 60 to 65%, while in case of water is found to be from 55 to 60% and for oil it is from 45 to 50%. • The average value for energy efficiency of the system during discharging is found to be better in case of water ranging from 85 to 90%, while for oil, it is 75–80% and for air, it is 35–40%. • The average exergy efficiency of the system during charging is found to be the highest in case of air 45–50%, while in case of water it is observed to be in the range of 25–30%, and for oil it is nearly 31%. • The average exergy efficiency of the system during discharging is higher in case of oil which lies in the range of 65–70%, while for water it is reduced to 60–65% and for air it is further reduced to 30–35%.

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• The overall energy efficiency of the storage system taking air as the heat transfer fluid lies in the range of 28–50%, which in case of oil is found to be in the range of 30–40%, and the efficiency in case of water lies in the range of 35–70%. • The overall exergy efficiency of the storage system for air lies in the range of 10–20%, while for water it lies in the range 7–26%, and for oil it is approximately 20%. Average exergy efficiency obtained is better in case of oil as compared to the other two heat transfer fluids considered.

References 1. H. Singh, R.P. Saini, J.S. Saini, A review on packed bed solar energy storage systems. Renew. Sustain. Energy Rev. 14(3), 1059–1069 (2010). https://doi.org/10.1016/j.rser.2009.10.022 2. G. Zanganeh, A. Pedretti, S.A. Zavattoni, M.C. Barbato, A. Haselbacher, A. Steinfeld, Design of a 100 MWhth packed-bed thermal energy storage. Energy Procedia 49, 1071–1077 (2014). https://doi.org/10.1016/j.egypro.2014.03.116 3. K.G. Allen, T.W.V. Backström, D.G. Kröger, Packed rock bed thermal storage in power plants: design considerations. Energy Procedia 49, 666–675 (2014). https://doi.org/10.1016/j.egypro. 2014.03.072 4. M.A. Al-Nimr, M.K. Abu-Qudais, M.D. Mashaqi, Dynamic behaviour of a packed bed energy storage system. Energy Convers. Manag. 37(1), 23–30 (1996). https://doi.org/10.1016/01968904(95)00024-8 5. P.J. Hughes, S.A. Klein, D.J. Close, Packed bed thermal storage models for solar air heating and cooling systems. J. Heat Transf. 98(2) (1976). https://doi.org/10.1115/1.3450552 6. C. Choudhury, H.P. Garg, Performance calculations for closed loop air to water solar hybrid heating systems with and without a rock bed in the solar air heater. Renew. Energy 3(8), 897–905 (1993) 7. L. Chai, L. Wang, J. Liu, L. Yang, H. Chen, C. Tan, Performance study of a packed bed in a closed loop thermal energy storage system. Energy 77, 871–879 (2014). https://doi.org/10. 1016/j.energy.2014.09.073 8. K. Sagara, N. Nakahara, Thermal performance and pressure drop of rock beds with large storage materials. Sol. Energy 47(3), 157–163 (1991) 9. A. Mawire, M. McPherson, R.R.J. Van Den Heetkamp, S.J.P. Mlatho, Simulated performance of storage materials for pebble bed thermal energy storage (TES) systems. Appl. Energy 86(7–8), 1246–1252 (2009). https://doi.org/10.1016/j.apenergy.2008.09.009 10. L. Moens, D.M. Blanke, D.L. Rudnicki, M.J. Hale, Advanced thermal storage fluids for solar parabolic trough systems. J. Sol. Energy Eng. 125(1), 112–116 (2003)

Chapter 6

Performance Analysis of Parabolic Trough Solar Collector with ‘U’-Tube and Helical Coil Receivers Mohd. Mubashshir Naved, Sandeep S. Joshi and Nikhil A. Bhave

Abstract The parabolic troughs are the most popular line focusing collectors. They are being used for process heating and medium temperature applications since long. In these collectors, the sunlight is concentrated at the receiver fixed at the focal line of the parabola. In general, the tubular types of receivers are used in the line focusing collectors. The working liquid is allowed to flow through the receiver tube. In the current study, the performance of a parabolic trough collector with the modified receivers is analyzed by experiments. Two types of receivers, viz. ‘U’-tube receiver and ‘helical coil receiver,’ enclosed in an evacuated glass tube are used. The system is installed at Nagpur [21.14° N, 79.08° E]; experiments have been performed in the month of February. The highest output temperature of 90 °C and the corresponding efficiency of 90.92% is obtained with helical coil receiver geometry. The details of receiver geometries and the experiments are discussed in this article. Keywords Parabolic trough collector · Helical coil receiver · U-tube receiver

Nomenclature ω β δ Φ cw mw T out T in d l

Hour angle Tilt angle Declination angle Latitude of the location Specific heat of water at constant pressure(J/kg K) Mass flow rate of water (kg/s) Outlet water temperature (°C) Inlet water temperature (°C) Outer diameter of receiver tube (m) Length of receiver tube (m)

Mohd. M. Naved (B) · S. S. Joshi · N. A. Bhave Shri Ramdeobaba College of Engineering and Management, Nagpur, Maharashtra, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_6

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6.1 Introduction Looking at the current energy scenario, globalization, and depleting conventional energy sources, there is an urgent need to utilize the renewable energy sources. Solar energy is the most popular and abundant renewable energy source. Several techniques have been developed to use the solar energy for various applications. In the medium temperature-concentrated solar applications, the parabolic trough collectors are most popular due to its economical manufacturing. Since inception, to improve the performance of the trough collectors, several modifications have been suggested so far. The major components in the trough collector system are the receivers, reflectors, and the working fluid. Most of the past work includes the use of different working fluids and the receiver geometry. Many researchers focused on the use of different efficiency-enhancing techniques with the use of different reflectors, receivers, and heat transfer fluids. Bader et al. [1] proposed a novel solar trough concentrator design, a cylindrical cavity-receiver containing a tubular absorber that uses air as the heat transfer fluid. A simple Monte Carlo ray-tracing technique has been established to study the behavior of PTC. To increase the overall heat transfer performance and reliability of tube receiver for parabolic trough solar collector system, Wang et al. [2] introduced asymmetric outward convex corrugated tube of parabolic trough receiver. The research indicated that the usage of such receivers can enhance the heat transfer performance and reduce the thermal strain effectively. Gong et al. [3] inserted pin fin arrays into the absorber tube of parabolic trough receiver to increase the overall heat transfer performance. A thermal and optical analysis of U-type evacuated tube and a cylindrical absorber was carried out by Korres et al. [4]. The thermal and optical efficiency and the overall heat loss coefficient of the collector were calculated for several incident angles of the sun rays. Daabo et al. [5] studied optical efficiency as well as the flux distribution of cylindrical, conical, and spherical receivers. The study concluded with having maximum optical efficiency and uniform flux distribution with conical receiver. Evaluation was carried out by Kearney et al. [6] to investigate the feasibility of utilizing a molten salt as the heat transfer fluid and for thermal storage in a parabolic trough. Evaluation showed that higher temperatures can be attained using molten salt heat transfer fluid. Jaramillo et al. [7] tested a low twist ratio twisted tape inserts and found the boost in performance of a parabolic trough collector. In order to improve the heat transfer and increase the efficiency of parabolic trough collector, Abad et al. [8] used absorber filled with metal (copper) foam and changing volume flow rate. Study resulted in a decrease in overall loss coefficient.

6.2 Experimental Setup The schematic diagram of the experimental setup is shown in Fig. 6.1. It consists of a parabolic trough, a direct absorption collector, storage tank, non-return valves fitted in the pipeline to define the flow direction, and control valve used to regulate

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Fig. 6.1 Schematic of experimental setup

the flow rate through the circuit. Trough is mounted on the mild steel stand with the use of nut and bolt for fixing. Water is supplied using a water tank placed on the stand above the height of the supply end of receiver tube through the supply pipe. The receivers consist of helical copper tube and a U-tube inserted in an evacuated tube which is mounted centrally on the parabolic collector. Due to helical structure, the surface area of helical coil receiver is more than U-tube receiver (Figs. 6.2 and 6.3).

Glass Tube

Absorber

U tube receiver Reflector

Fig. 6.2 U-tube copper receiver

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Evacuated tube

Helical Copper coil

Fig. 6.3 Helical tube copper receiver

6.3 Experimental Method The experimentation is started with cleaning the system, i.e. wiping out the dust present on parabolic trough which can hamper the performance of the system followed by flushing of absorber collector. The evacuated tube is made of an outer glass tube of diameter 58 mm and an inner glass tube of diameter 45 mm with selective coating; it is having a length of 1650 mm. Passage between outer and inner tubes was evacuated; due to this the heat transfer losses are reduced in large amount and helps in increasing the inner temperature of the core passage area. The setup is made at North–South as horizontal axis parallel to the focal axis of trough because more amount of solar radiation is reflected by reflector in this direction. Manual tracking was carried out to track the sun for getting the maximum solar irradiations. Tracking is done manually by adjusting the trough with 15° every one hour. The experiment has been performed for 7 h over the day from 1000 to 1600 h. Pyranometer is used in the experimentation to measure the global radiation. To reduce convective losses, glazing is used over an entire collector area. The experiments have been performed for the entire month. During experiments, all the temperatures, mass flow rate, and the radiation intensities are measured using well-calibrated instruments. The radiation intensities are recorded using a Class-I pyranometer (Model—Kipp Zonen/PM-10). The average values of the entire month have been considered for the subsequent calculations so that the effect of instrumental error if any can be neglected (Fig. 6.4).

6.4 Result and Discussion While carrying the experiments, the optimum mass flow rate of water is kept constant. A thermal analysis is done to calculate the behavior of the collector for different receiver geometry. Equations 6.1–6.3 have been used to estimate the thermal performance of the system (Figs. 6.5, 6.6 and 6.7).

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Reflector

Insulated storage tank

Fig. 6.4 Experimental setup

Fig. 6.5 Measured irradiation and ambient temperature Fig. 6.6 Water outlet temperature

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Fig. 6.7 Instantaneous collector efficiency

Heat gained by the water qu = m w cw (Tout − Tin )

(6.1)

Tilt factor which is the ratio of beam radiation flux falling on the tilted surface to that falling on the horizontal surface can be calculated as rb =

sin δ sin(Φ − δ) + cos δ cos ω cos(Φ − β) sin Φ sin δ + cos Φ cos δ cos ω

(6.2)

Instantaneous collector efficiency is the ratio of useful heat gain by water to the solar energy concentrated by parabolic trough and given by η=

qu Ib × rb × π × d × l

(6.3)

6.5 Conclusion The parabolic trough collectors are able to produce heat at high temperatures suitable for various thermal applications. The performance of parabolic trough collector with U-tube and helical coil receiver is investigated in this study. Water is used as working medium. The experiments were carried out at Nagpur [21.14 °N, 79.08 °E] in the month of February. The maximum outlet temperature of 90 °C and 87 °C was observed for the helical coil and U-tube receiver, respectively. The maximum instantaneous collector efficiency of 90.92 and 88.33% was recorded for the helical coil and U-tube receiver, respectively. The study shows that the helical coil receiver gives the best performance with parabolic trough collector as compared to the conventional tubular receiver systems. Some of the major shortcomings of this system are manufacturing difficulties and the initial investment of the helical tube receiver for

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large-scale installations, difficulties in the cleaning of the receiver tube. In addition, the copper tube receiver adds additional weight to the tracking system which may increase the auxiliary power consumption. Further research work is recommended to overcome the mentioned shortcomings. It is further suggested to make use of different selective coating materials for the receivers, different working fluids, and alternative low-cost material for the collector and receiver systems.

References 1. R. Bader, M. Barbato, A. Pedretti, A. Steinfeld, An air-based cavity-receiver for solar trough concentrators. J. Sol. Energy Eng. 132(3), 7–14 (2010) 2. F. Wang, Z. Tang, X. Gong, J. Tan, H. Han, B. Li, Heat transfer performance enhancement and thermal strain restrain of tube receiver for parabolic trough solar collector by using asymmetric outward convex corrugated tube. J. Energy 144(6), 275–292 (2016) 3. X. Gong, F. Wang, H. Wang, J. Tan, Q. Lai, H. Han, Heat transfer enhancement analysis of tube receiver for parabolic trough solar collector with pin fin arrays inserting. J. Sol. Energy 144(3), 185–202 (2017) 4. D. Korres, C. Tzivanidis, A new mini-CPC with a U-type evacuated tube under thermal and optical investigation. J. Renew. Energy 148(17), 560–568 (2017) 5. A. Daabo, S. Mahmoud, R.K. Al-Dadah, The effect of receiver geometry on the optical performance of a small scale solar cavity receiver for parabolic dish applications. J. Energy 114(3), 513–525 (2016) 6. D. Kearney, U. Herrmann, P. Nava, B. Kelly, R. Mahoney, J. Pacheco, R. Cable, N. Potrovitza, D. Blake, H. Price, Assessment of a molten salt heat transfer fluid in a parabolic trough solar field. J. Sol. Energy Eng. 125(9), 225–232 (2013) 7. O.A. Jaramillo, Mónica Borunda, K.M. Velazquez-Lucho, M. Robles, Parabolic trough solar collector for low enthalpy processes: an analysis of the efficiency enhancement by using twisted tape inserts. J. Renew. Energy 93(9), 125–141 (2016) 8. M.T. Jamal-Abad, S. Saedodin, M. Aminy, Experimental investigation on a solar parabolic trough collector for absorber tube filled with porous media. J. Renew. Energy 17(3), 502–526 (2017) 9. J. Abad, Experimental investigation on a solar parabolic trough collector for absorber tube filled with porous media. J. Renew. Energy 17(3), 502–526 (2017) 10. S. Sukhatme, J. Nayak, Solar Energy Principles of Thermal Collection and Storage, 3rd edn. (Tata McGraw-Hill Education, New Delhi, India, 1996) 11. A. Mwesigye, P. Meyer, Heat transfer and entropy generation in a parabolic trough receiver with wall-detached twisted tape inserts. Int. J. Therm. Sci. 99(8), 238–257 (2017) 12. R. Farjallah, Thermal performance of the U-tube solar collector using computational fluid dynamics simulation. J. Sol. Energy Eng. 138, 111–132 (2016) 13. A. Jaramillo, Parabolic trough solar collector for low enthalpy processes: an analysis of the efficiency enhancement by using twisted tape inserts. J. Renew. Energy 93(9), 125–141 (2016) 14. P. Sasa, Experimental investigation and parametric analysis of a solar thermal dish collector with spiral absorber. J. Appl. Therm. Energy 11(4), 129–141 (2016)

Chapter 7

Performance Evaluation of an Improved Dual Purpose Solar Collector P. P. Krishnaraj and P. Arun

Abstract A dual purpose solar collector permits the simultaneous heating of two fluids utilizing the incident solar energy on the collector. An experimental study on an improved dual purpose solar collector integrated with fins and wire mesh in absorber plate has been performed. Simultaneous heating of both air and water is obtained using such a system. The hot air can be used for drying of crops, fruit, timber, etc., while the water can be used for domestic purposes. The performance study of the system has been reported for the typical summer day conditions of Kozhikode (11.25 °N, 75.77 °E). The variations of solar intensity, air temperature rise, water temperature rise, absorber plate temperature, etc., have been illustrated for three modes of system operation. For the system working as a single mode solar air heater, maximum temperature difference between inlet and outlet air was obtained as 33.5 °C for a plate temperature of 104 °C. For the system working as a single mode solar water heater, maximum temperature difference between inlet and outlet water was obtained as 43 °C for a plate temperature of 90 °C. For the system operating in the dual purpose mode, maximum temperature rise of 20.8 and 40.5 °C was observed for the air and water streams, respectively. Keywords Dual purpose solar collector · Thermal performance · Solar air heater · Solar water heater

7.1 Introduction Development of a nation’s power sector has a direct impact on its economy. However, considering the rapidly depleting conventional energy resources, there is an increased need for utilization of renewable energy in a nation’s energy mix. Solar P. P. Krishnaraj (B) Rajagiri School of Engineering and Technology, Kochi 682039, India e-mail: [email protected] P. Arun National Institute of Technology, Calicut 673601, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_7

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energy is major renewable energy resource particularly suited for India. Solar thermal systems and solar photovoltaic systems are major classifications of solar energy systems. Solar thermal systems develop heat (low-grade energy), whereas solar photovoltaic system develops electrical energy (high-grade energy) from solar insolation. Flat plate collectors are widely used for low and medium temperature solar thermal applications. They are typically used for heating water in solar water heating systems or air as in the case of a solar air heater. Several methods have been reported regarding the thermal performance improvement of flat plate collectors. For example, a numerical study on a double pass flat plate solar air heater considering the performance and entropy generation with the incorporation of longitudinal fins was reported by Naphon [1]. The numerical results were obtained at mass flow rates ranging from 0.02 to 0.1 kg/s. It was understood that thermal efficiency increased with increase in height and number of fins. Improvement of thermal efficiency of a flat plate collector with a pin fin integrated absorber plate was given by Peng et al. [2]. An experimental study on solar collector with staggered absorber sheets and fins attached on absorber plate was done by Ucar and Inalli [3]. Irreversibility was higher in case of conventional solar collector in which collector efficiency was the lowest. The experiments also revealed that using staggered sheets and fins increases efficiency of the solar collector. An experimental analysis on single and double pass solar air heaters was done by Mahmood et al. [4]. Necessary modification to absorber plate was done by incorporating longitudinal fins and porous matrix in the lower channel. Thermal efficiency of double pass solar air heater was obtained as 63% and single pass solar air heater was obtained as 59%. A mathematical study was conducted by Ammari [5], for a flat plate solar air heater that has slats provided between the absorber plate and bottom plates. Solar air heater with slat showed greater improvement in performance over that of common type heater. The efficiency of a solar collector could be enhanced by using metal slats between absorber plate and bottom plate of solar air heater as reported by Matrawy [6]. The collector showed better performance than that of finned type of collector. An experimental study on thermosyphon type solar water heater was performed by Sae-Jung et al. [7]. The maximum temperature of storage tank was obtained as 55 °C, and the maximum collector outlet temperature was obtained as 68 °C. The average thermal efficiency of system, for which the experiment was performed, was obtained as 56%. An experimental study was conducted by Karaghouli and Alnaser [8] on thermosyphon solar water heater. At a peak intensity of 695 W/m2 , average storage tank temperature was obtained as 50 °C. The average efficiency of the system was obtained as 38%. All the above-mentioned studies are related to solar collector working with single working fluid, either air or water. A flat plate solar collector for simultaneous heating of multiple fluids like air and water is termed as a dual purpose solar collector. An experimental and theoretical investigation on dual purpose solar collector was conducted by Assari et al. [9, 10]. Air enters through the upper channel and leaves through the bottom channel of the dual purpose collector as studied by them. A thermosyphon type solar water heating system was used. Efficiency of dual purpose system was reported to be 3–5% higher than single purpose system. An experimental study of dual purpose solar heating

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system was also done by Omid et al. [11]. The system was used to heat the air simultaneously or separately. The efficiency of dual purpose system was observed to be higher than the single purpose system. In this study, a new and improved dual purpose solar collector has been designed with integrated fins and wire mesh arrangement in the absorber plate so as to increase its thermal performance. An improved thermal performance and higher temperature of heat transfer fluids were obtained using such a system over the existing designs. Further, the system is compact consuming relatively lesser space. The system generates hot air which can be used for crop drying, timber drying, fruit drying applications and hot water can be used for domestic purposes. In the present study, illustration of the system performance in the tropical climate of Kozhikode (11.25 °N, 75.77 °E) operating at three different modes has been presented.

7.2 Experimental Setup and Procedure The performance of an improved dual purpose solar collector, with integrated fins and wire mesh arrangement in the absorber plate has been investigated in this work.

7.2.1 Description of the System The cross-sectional view of the dual purpose collector is represented in Fig. 7.1. Dual purpose solar collector consists of a copper absorber plate of dimensions 2 m × 1 m and 1 mm thickness. The height of upper channel is 7 cm and that of lower channel is 10 cm. The bottom and sides of the collector are insulated with a 2 cm thick layer of glass wool. Five copper fins of length 150 mm, thickness 2 mm and height 30 mm are fixed longitudinally at equal spacing at the bottom of copper absorber plate. In order to increase the absorptivity, top surface of the absorber plate and lower channels are coated with black paint. A top cover made of toughened glass of thickness 5 mm

Fig. 7.1 Cross-sectional view of dual purpose solar collector

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Fig. 7.2 Dual purpose solar collector test setup

is provided for reducing convective heat loss from the collector. Steel wire mesh layers are fixed on the lower channel so as to increase the heat transfer coefficient between the plate and air. The air is passed through the inlet pipe by means of a blower and flow rate is adjusted by means of a valve. Header and riser configuration, commonly used in solar water heater assembly is used for water circulation in the collector. Six copper tubes of diameter 12.5 mm, thickness 0.5 mm having center to center distance of 120 mm are used as riser. Copper tubes of diameter 25.4 mm and thickness 0.71 mm are used as header. Water inlet to the collector assembly is taken from bottom left side and outlet from the collector is taken from the top right side. Inlet and outlet are connected to an insulated storage tank of 100 L capacity. The photograph of the test setup installed in the Solar Energy Centre of National Institute of Technology Calicut located at Kozhikode is shown in Fig. 7.2.

7.2.2 Test Procedure In this study, the performance of the dual purpose collector was experimentally investigated for the summer months of Kozhikode. The results of the dual purpose operation have been compared with the collector operations as single mode air heater

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and single mode water heater. The test was conducted at a constant air mass flow rate of 0.009 kg/s. A venturimeter with U-tube differential manometer is used for the airflow rate measurements. A 600 W electric blower was used to pump the air through the inlet pipe and was controlled manually. Measurement of air and water temperature was done with J-type thermocouples with an accuracy of ±0.1 °C. The airflow rate was maintained at 0.009 kg/s, air was passed for enough time, and steady state readings were recorded. The experiments on dual purpose solar collector were carried out for several days. The representative observations for four successive days (28th, 29th, 30th, and 31st of March) have been illustrated. These days had almost comparable climatic conditions such as wind speed, ambient temperature, cloudiness, and humidity.

7.3 Results and Discussion Experimental studies were conducted on the modified dual purpose solar collector. The following modes of operation were considered: (1) single mode solar air heater, (2) single mode solar water heater, and (3) dual mode air and water heater. The variation of solar intensity, absorber plate temperature, and temperature rise of air and water versus time of the day and mass flow rates were obtained and plotted. Comparisons of dual purpose solar collector working in three different modes were also carried out. The efficiency of solar collector was calculated with the following formula: η=

mC ˙ p T I Ac

T = Tout − Tin where m˙ Cp I av

is the mass flow rate of the fluid (air or water) in (kg/s). is the specific heat at constant pressure (kJ/kgK). is the average of intensity of solar radiation, measured with pyranometer (W/m2 ). T is the difference of temperature between the collector outlet and inlet (°C). Ac is the area of flat plate solar collector (m2 ).

7.3.1 Dual Purpose Collector Operation Variation of solar insolation on four days, i.e., on 28th, 29th, 30th, 31st of March 2016, is represented in Fig. 7.3. It is observed that there is a gradual increase in

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Fig. 7.3 Variation of solar intensity with time of day

solar insolation from morning 7.30 a.m. to noon 1 p.m., and then, it decreases from noon till evening 6 pm. The highest daily solar radiation obtained was 1180 W/m2 at 13.00 h. Variation of absorber plate temperature versus time of days for which the experiment was carried out is represented in Fig. 7.4. It can be seen that the rise in absorber plate temperature was in accordance with that of the solar intensity. The maximum value of temperature was obtained as 85.6 °C for the peak solar intensity of 1180 W/m2 . Temperature difference of inlet and outlet air versus time of days for a fixed mass flow rate of 0.009 kg/s is represented in Fig. 7.5. The difference in temperature Fig. 7.4 Variation of absorber plate temperature with time of day

Fig. 7.5 Variation of temperature rise of air with time of day

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Fig. 7.6 Variation of temperature rise of water with time of day

between inlet and outlet air increases to a peak value of 20.8 °C with increasing solar radiation. The maximum value of temperature difference was occurred between 12 h and 13.30 h for a minimum mass flow rate of 0.009 kg/s. Variation of temperature rise of water versus time of days is shown in Fig. 7.6. Temperature rise of water is greater than that of air due to higher value of heat capacity of water compared to air. Since the radiation and received energy enhances in the mid hours of day, rise in temperature of water is maximum during these hours. The maximum value of T was obtained to be 40.5 °C at 13.00 h at a peak value of solar intensity, i.e., 1180 W/m2 .

7.3.2 Single Mode Water Heater Operation Variation of absorber plate temperature versus time of days for which the experiment was carried out is represented in Fig. 7.7. The maximum value of peak temperature was obtained as 90 °C for the peak solar intensity of 1100 W/m2 . Variation of temperature rise of water versus time of days for which experiment is carried out as shown in Fig. 7.8. Temperature rise of water in single mode solar collector is greater than that of water in dual mode, since whole energy absorbed Fig. 7.7 Variation of absorber plate temperature rise of air with time of day

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Fig. 7.8 Variation of temperature rise of air with time of day

Fig. 7.9 Variation of energy efficiency with time of day

by the plate is taken by a single fluid, i.e., water. The maximum value for T was obtained to be 43 °C at 14.00 h. Variation of energy efficiency versus time of days is represented in Fig. 7.9. Solar radiation intensity and absorber plate temperature increase from morning to midday hours, and therefore, the heat absorbed by water increases during these hours. Later in the day, heat gained by water does not increase compared to that of heat absorbed by the plate, and also due to higher heat losses at higher temperatures, system efficiency decreases. The maximum value of energy efficiency is obtained as 51%.

7.3.3 Single Mode Air Heater Operation Variation of solar insolation on four days, i.e., on 25th, 26th, 27th and 28th of April 2016, is represented in Fig. 7.10. It is observed that there is a gradual increase in solar insolation from morning (7.30 a.m.) to noon (1 p.m.), and thereafter it decreases from 1 p.m. to 6 p.m. The highest daily solar radiation obtained was 1120 W/m2 at 13.00 h. Variation of absorber plate temperature versus standard local time of the days for which the experiment was carried out as represented in Fig. 7.11. It is observed that there is a gradual rise in absorber plate temperature which varied in accordance with that of solar intensity. The maximum value of plate temperature was obtained as 104 °C for a peak solar intensity of 1120 W/m2 .

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Fig. 7.10 Variation of solar intensity with time of day

Fig. 7.11 Variation of absorber plate temperature rise of air with time of day

Temperature difference of inlet and outlet air versus time of day for a fixed mass flow rate of m˙ = 0.009 kg/s is represented in Fig. 7.12. It is observed that the temperature difference between inlet and outlet air increases to a peak value of 34 °C with increasing solar radiation. The maximum value of temperature difference occurred between 12 h and 13.30 h for a minimum mass flow rate of 0.009 kg/s. Variation of thermal efficiency of solar air heater versus standard local time of the days for which the experiment was carried out was represented in Fig. 7.13. It is observed that there is a gradual rise in thermal efficiency from morning 7.30 h to noon since the heat gained by the air from the absorber plate increases, later as the absorber plate temperature decreases the heat gain by air decreases and thermal efficiency decreases. The maximum value of thermal efficiency obtained was 48% Fig. 7.12 Variation of temperature rise of air with time of day

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Fig. 7.13 Variation of thermal efficiency with time of day

Fig. 7.14 Variation of temperature rise of air with mass flow rate for I = 600 W/m2

at 12.30 h at a peak value of absorber plate temperature of 104 °C and solar intensity of 1120 W/m2 . Temperature difference of inlet and outlet air versus mass flow rate obtained from experimental analysis at three different mass flow rates for I = 600 W/m2 is represented in Fig. 7.14. It is observed that the temperature difference between inlet and outlet air has a peak value of 11.5 °C with mass flow rate of 0.009 kg/s. Temperature difference between inlet and outlet air decreases with increasing mass flow rate as the resident time of air inside the collector is reducing. Temperature difference of inlet and outlet air versus mass flow rate obtained from experimental analysis at different mass flow rates for I = 1100 W/m2 is represented in Fig. 7.15. It is observed that the temperature difference between inlet and outlet air has a peak value of 21.5 °C with mass flow rate of 0.009 kg/s. Temperature difference between inlet and outlet air decreases with increasing mass flow rate as the resident time of air inside the collector is reducing.

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Fig. 7.15 Variation of temperature rise of air with mass flow rate for I = 800 W/m2

7.3.4 Comparison of Different Modes of Operation The maximum rise in temperature of air and water in single mode solar collector is compared with dual purpose solar collector for three representative days of almost similar solar intensity as represented in Figs. 7.16 and 7.17. Rise in temperature of air for the system working as dual purpose solar collector is compared with that of system working as single mode solar collector. The maximum rise in temperature of air for the system working as dual mode solar collector was obtained as 21 °C in Fig. 7.16. The maximum rise in temperature of air for system Fig. 7.16 Comparison of rise in temperature of air in single and dual mode with time of day

Fig. 7.17 Comparison of rise in temperature of water in single and dual mode with time of day

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working as single mode solar air heater is obtained as 33.5 °C. The rise in temperature of air in single mode solar collector is higher as compared to that of dual mode solar collector since total amount of heat from the plate was absorbed by single fluid (air). Rise in temperature of water for the system working as dual purpose solar collector is compared to that of system working as single mode solar collector. The maximum rise in temperature of water for the system working as dual purpose solar collector was obtained as 41 °C in Fig. 7.17. The maximum rise in temperature of water in single mode solar collector is higher as compared to that of dual mode solar collector since the total amount of heat from the plate was absorbed by single fluid (water).

7.4 Conclusion A new and improved design of a dual purpose solar collector (DPSC) setup was fabricated. For this, an existing solar air heating system was appropriately modified. Calculations were performed to obtain the design details of the additional components needed for the dual purpose collector. The system was incorporated with fins in the lower duct, and wire mesh packing was provided on the absorber plate. Based on the finalized design, a solar collector utilizing air and water as fluids, integrated with fins and wire mesh, was fabricated. An extensive experimental study was conducted on the fabricated setup of DPSC. In case of DPSC, for a fixed mass flow rate of 0.009 kg/s, the absorber plate temperature varied in similar manner as that of solar intensity and the maximum value of plate temperature is obtained as 104 °C at a peak solar intensity of 1120 W/m2 . For the system working as a single mode solar air heater, maximum temperature difference between inlet and outlet air was obtained as 33.5 °C for a plate temperature of 104 °C. The intermediate drop and rise of air temperature difference can be attributed to rise and drop in the value of solar intensity. The maximum value of efficiency is obtained as 48%. The experimental results suggest that with increase of mass flow rate for fixed solar intensity (800 W/m2 ) and ambient temperature of 35 °C, temperature rise shows a decreasing trend and thermal efficiencies show an increasing trend with respect to mass flow rates. For the system working as a single mode solar water heater, maximum temperature difference between inlet and outlet air was obtained as 43 °C for a plate temperature of 90 °C. The maximum value of efficiency is obtained as 51%. A lot of potential exists for improving the thermal performance of the system. The optimization of the system considering various parameters such as height of channels, height of storage tank, number of fins, and porous medium used can be done.

References 1. P. Naphon, A numerical study on the performance and entropy generation of the double pass solar air heater with longitudinal fins. Renew. Energy 30, 1345–1357 (2005)

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2. D. Peng, X. Zhang, H. Dong, K. Lv, Performance study of a novel solar collector. Appl. Therm. Eng. 20, 2594–2601 (2010) 3. A. Ucar, M. Inalli, Thermal and exergy analysis of solar air collectors with passive augmentation techniques. Int. Commun. Heat Mass Transf. 33, 281–1290 (2006) 4. A.J. Mahmood, L.B.Y. Aldabbagh, F. Egelioglu, Investigation of single and double pass solar air heater with transverse fins and packed wire mesh layers. Appl. Energy 87, 3759–3765 (2010) 5. H.D. Ammari, A mathematical model of thermal performance of a solar air heater with slats. Renew. Energy 28(10), 1597–1615 (2003) 6. K.K. Matrawy, Theoretical analysis of an air heater with box type absorber. Sol. Energy 63(3), 191–198 (1998) 7. P. Sae-Jung, T. Krittayanawach, P. Deedom, B. Limmeechokchai, An experimental study of thermosyphon solar water heater in Thailand. Energy Procedia 442–447 (2015) 8. A.A. Karaghouli, W.E. Alnaser, Experimental study on thermosyphon solar water heater in Bahrain. Renew. Energy 24, 389–396 (2001) 9. M.R. Assari, H. Basirat Tabrizi, I. Jafari, Experimental and theoretical investigation of dual purpose solar collector. Sol. Energy 85, 601–608 (2011) 10. M.R. Assari, H. Basirat Tabrizi, H. Kavoosi, M. Moravej, Design and performance of dual purpose solar collector, in Proceedings of 3rd International Energy, Exergy and Environment Symposium, IEEE-3, University of Evora, Portugal (2006) 11. O. Nematollahi, P. Alamdari, M. R. Assari, Experimental investigation of dual purpose solar heating system. Energy Convers. Manag. 78, 359–366 (2014)

Chapter 8

Experimental Investigation on Farmer-Friendly Hybrid Dryer for Indoor Drying of Mushroom K. Sharma, S. Kothari, N. L. Panwar, N. Rathore and K. Samar

Abstract This investigation deals with thermal performance evaluation, technoeconomic analysis and quality analysis of dried mushroom flakes in a developed hybrid system using both solar and electrical energy. Hybrid dryers are efficient and economic means of continuous drying of agricultural products which require drying at temperature range of 45–60 °C without compromising in the quality of the final product. Heat storage capacity of water is higher than air, so hot water is directly circulated in the drying chamber through a heat exchanger to generate heat inside the drying chamber resulting in substantial saving of electricity. Mushrooms are considered for the investigation as they are highly perishable agricultural products having good off seasonal utilization. They are a source of powerful nutrients and a leading source of essential antioxidant which are relished for their characteristic aroma and flavour. Hence, drying is must for proper storage of mushrooms to increase its shelf life. Mushroom flakes were dried within 8 h from the moisture content of 82.0% (w.b.) to 10.67% (w.b.). For mushrooms, besides the investigation of the drying times and rates, the product rehydration capacity and quality are also evaluated. Keywords Hybrid drying · Mushroom · Performance evaluation · Techno-economic analysis · Quality analysis

8.1 Introduction Today, mushroom cultivation is one of the biggest markets in the world with mushroom being an important horticulture cash crop. The importance of mushroom lies in the fact that its production does not require any arable land, only some nonagricultural land to build infrastructure for its preparation of substrate [1]. But this indoor crop is highly perishable, so its producers are obliged to sell it at a lower price during their maximum yielding time which calls for its preservation. Thus, K. Sharma (B) · S. Kothari · N. L. Panwar · N. Rathore · K. Samar Department of Renewable Energy Engineering, College of Technology and Engineering, Maharana Pratap University of Agriculture and Technology, Udaipur, Rajasthan, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_8

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by its preservation in peak growing season the landless and marginal farmers can gain profit in the lean season [2]. Drying proves to be a cheap and better method for increasing the shelf life which is a prerequisite process for proper storage of mushrooms. It provides the advantages of substantial reduction in weight, volume and also minimizes packaging, storage and transportation costs [3, 4]. Mushroom drying is a very beneficial process as the dried product is of good market value as it is easy to reconstitute and contains considerable amount of nutrients and after drying, it can be sold in off-season in reasonable price by converting it into value added products like soup base, mushroom paste, etc. The mushroom of the Pleurotus genus is more delicate and sensitive than the Agaricus genus and they start deteriorating immediately within one day after the harvest [2]. Hence, viewing at the production, requirement of processing, the potential of market for export and also the seasonal availability of oyster mushroom, it is selected for the present research work. To overcoming spoiling problems of vegetables, food grains and fruit, various preserving methods are used and renewable sources are best for this purpose by which we can save energy for preservation and keeping the product in their natural flavour [5]. Most of the agricultural products are dried at temperature range of 45–65 °C. Solar energy can be used to heat air up to this range of temperature needed for drying of most of the agricultural products, efficiently and economically without compromising in the quality of final product. In India, the average solar radiation available is 5 kW/m2 /day for 250–300 days in a year with approximately 8–10 h of full sunshine hours [6]. The intermittent nature of solar energy led to non-continuous drying of products causing absorption of moisture by the product during night time leading to increase in the total fungal count and total microbes count. To overcome this problem, drying can be made continuous by using electric heater during night times or off sunshine hours for the drying purpose. Thus, keeping in view of the off seasonal utilization of mushroom, investigation was carried out for indoor drying of mushrooms through hybrid drying system as it is well recognized that testing of solar equipment under indoor controlled simulation is preferred to outdoor variable conditions for obtaining more reliable steady state results [7].

8.2 Materials and Methods 8.2.1 System Description The hybrid drying system consisted of a drying chamber, solar collector (ETC type), water storage tank, a 0.5 hp motor pump, a heat exchanger, an exhaust fan with volumetric air flow rate of 4.5 m3 /h, a 2 kW fin type electric heater, and insulated pipes. The design considerations of hybrid dryer are provided in Table 8.1. Drying chamber is made from galvanized iron sheet. To prevent the heat loss from drying chamber, it was made insulated. Drying chamber comprises of three trays above the electric heater on which product, which is to be dried, is placed.

8 Experimental Investigation on Farmer-Friendly Hybrid Dryer … Table 8.1 Design considerations of hybrid dryer

83

S. No.

Parameters

Specifications

1

Capacity (kg)

3 kg fresh mushroom slices

2

Initial moisture content (mi )

87% (w.b.)

3

Final moisture content (mf )

7% (w.b.)

4

Loading rate (LR)

1 kg/m2

5

Solar insolation (I t )

540 W/m2 h (March month)

6

Ambient temperature (T a )

30 °C

7

Drying temperature (T d )

60 °C

8

Ambient relative humidity (RH)

50%

9

Drying time (t d )

8h

Temperature sensor is provided above the first tray to measure its temperature. The velocity of the drying air was kept constant as 2.8 (+0.1) m/s. An evacuated tube collector (ETC) is used for water heating. Hot water is stored in storage tank and then supplied to drying chamber through insulated pipes. That hot water is utilized to heat air at bottom of the drying chamber. A plate and fin type heat exchanger uses heat transfer mechanism for heating air. The temperature at inlet of drying chamber is adjustable so that it can be maintained at required amount of temperature. At the bottom of drying chamber, electric heater is provided to heat the incoming air inside the drying chamber as per the required amount of temperature. Temperature controller is provided with the heater to adjust the heat input. If temperature of incoming hot air is 50 °C and the required temperature is 60 °C, then it will add up and drying chamber will get 60 °C temperature, so that constant drying can take place. Exhaust fan is provided to remove the humidity inside the drying chamber. The technical specifications of drying system are presented in Table 8.2. Table 8.2 Dimensions of hybrid dryer S. No.

Particulars

Length (m)

Width (m)

Height (m)

1

Drying chamber (Insulation 0.04 m thick)

0.70

0.50

1.00

2

Plate and fin type heat exchanger

0.14

0.65

0.33

3

Temperature controller unit

0.20

0.38

0.34

4

Tray

0.60

0.45

0.05

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8.2.2 Experimental Setup and Measurement In the present study, oyster mushrooms which were procured from the Department of Plant Pathology, Rajasthan College of Agriculture, Udaipur, were used. Three kilogram cleaned mushrooms of approx. uniform size were separated from bundles of mushrooms and then were placed in the trays of the dryer without overlapping as the dryer which was developed in the Department of Renewable Energy Engineering. The schematic of the developed drying system is shown in Fig. 8.1. The initial moisture content of mushroom slices was determined by using the oven method at 110 °C for 24 h [8]. It ranged between 82 and 87% (w.b.). An average moisture content of 82% was considered for the research calculations. This drying experiment was conducted in the month of March 2015 with collector facing south to absorb maximum solar radiation incident on it. The global solar radiation incident on a horizontal surface was measured using a Solarimeter (Make: M/s. Surya Solar Systems, Ahmedabad). The dataTaker DT82E (manufactured by Thermo Fisher Scientific Australia Pty Ltd., Australia) having temperature range of −45 to 70 °C was used to measure temperature of all trays, ambient temperature, dryer inlet and exit air. The digital thermometer (manufactured by M/s. Mextech Instruments) was used to measure the temperature of hot water from the collector. Weight measurement was done by using electronic balance (Make: Adair Dutt and Co. Pvt. Ltd.). Total experimental uncertainties in measurement from instruments for different parameters during the experimental study are presented in Table 8.3.

Fig. 8.1 Schematic of hybrid dryer

8 Experimental Investigation on Farmer-Friendly Hybrid Dryer … Table 8.3 Uncertainties of the experimental instruments

Parameters

85 Uncertainty

Solarimeter

±0.17

Digital thermometer

±0.17

DataTakerDT82E

±0.15

8.2.3 Quality Analysis of Hybrid Dried Mushroom Rehydration Capacity Test The quality of rehydrated mushrooms in terms of colour, flavour and texture is highly important in a dehydrated product [9]. This characteristic was expressed in terms of a rehydration ratio, calculated as the ratio of the rehydrated mass to the dehydrated mass, and a rehydration fraction as shown in Eqs. (8.1) and (8.2) [10]. Rehydration ratio RR = C/D

(8.1)

Coefficient of rehydration COR =

C × (100 − A) , D × (100 − B)

(8.2)

where A B C D

Moisture content of samples before dehydration, IMC (% wb) Moisture content of dehydrated samples (% wb) Weight of rehydrated sample (g) Test weight of dehydrated sample (g).

Organoleptic Evaluation In our study, the quality of the dehydrated mushroom samples was evaluated on the basis of organoleptic evaluation for colour, taste, flavour and rehydration studies on which mainly their commercial acceptability depends. All the samples were placed before consumer panel of 10 judges and colour, taste and flavour were assessed using 9 point hedonic scale [11].

8.2.4 Assumptions for Techno-Economic Evaluation 1. 2. 3. 4.

The life of hybrid dryer is assumed as 10 years (n). The discount rate is assumed to be 8%, (i). The Annual repair and maintenance cost is 3% of cost. The cost of fresh oyster mushroom is Rs. 70/kg during season.

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5. 6. 7. 8. 9.

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The cost of dried oyster mushroom is Rs. 800/kg. Labour charge is assumed as Rs. 100 per day. The dryer can be operated 300 days in a year. The cost of system is Rs. 70,000. The total material cost for drying of mushroom flakes is Rs. 63,000 per year.

8.3 Results and Discussion 8.3.1 Thermal Performance The hybrid drying system was developed with the aim of supplying constant hot air for drying of agricultural products. The experiment was successful. The thermal performance of the system is evaluated on the basis of no load and full-load conditions of the dryer [12]. No Load Performance with Electric Heater The no load test with no product and a preset temperature of 60 °C in drying unit was carried out to know the trends of various operating parameters with respect to time [13]. The temperature profile at various positions in dryer is presented in Fig. 8.2. It was observed that the insolation was increased up to 13.00 h and then decreases rapidly. Similarly the ambient temperature increased with the day time and decreases in late hours. The temperature of water along the collector length was also increased gradually. The temperature at collector outlet and dryer outlet was also increased up to 13:00 h and then decreased. The dryer temperature was stable and maintained at 60 °C but with the dryer outlet temperature higher than the drying cabinet temperature during the hours of 12:00 h 1600 h. This increase

Fig. 8.2 Performance curve for no load testing with electric heater

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Fig. 8.3 Performance curve for full-load testing

occurred as a result of no work done in the dryer and also because of the heat which gets accumulated during the period when the ambient temperature was at its peak. During the last hours of drying, the dryer outlet temperature comes close to the dryer cabinet temperature and attains equilibrium at the end. The electricity consumption by using motor pump was approximately 2 units in 8 h. Full-Load Performance A full-load measurement highlights latent problems in the system [14]. The drying temperatures in the drying cabinets were shown in Fig. 8.3 and similar trend was also reported by [15]. The temperature of the drying chamber was set at 60 °C as required for drying of food products. Electric heater was provided to supply heat when temperature was below 60 °C that is in the initial hours of operation and also during the late hours of operation if required. It was observed that during the full load the maximum temperature at ambient was 37 °C. The maximum solar radiation was 1057 W/m2 at 13.00 h. During the initial hours of drying, much of the product moisture gets evaporated with moving hot air; hence, much work is done during this period. Meanwhile, when the dryer temperature becomes higher than the preset temperature thermostat breaks the power supply to the heating element making only the fan and motor to be at work, blowing the excess heat out of the drying chamber. This makes the outlet temperature higher than the drying cabinet temperature as experienced throughout the experimental period. During the last hours of drying, the exit temperature comes close to the drying cabinet temperature, and finally when all the product moisture gets evaporated an equilibrium is attained inside the dryer with a safe storage moisture level of product.

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8.3.2 Variation in Moisture Content The drying product took 8 h to get the desired level of moisture content for safe storage, i.e. 10.67% (w.b). In the first 4 h of drying, product moisture went from 82 to 53% (w.b.) and in the last 4 h of drying, moisture ranged from 52 to 10.63% (w.b.). Three kilogram of fresh mushroom flakes were reduced to 0.54 kg after dehydration. At initially, moisture evaporates rapidly and then it is removed slowly.

8.3.3 Variation in Drying Rate The drying rate for the mushroom flakes was estimated from the difference in its moisture content in a given time interval and expressed as gram water evaporated per hour [16]. The average drying rate during the drying varies over the range of 0.022–1.555 g water evaporated/g of dry matter/hr.

8.3.4 Variation in Electricity Consumption The total electricity consumption using electric heater, motor, exhaust fan, temperature controller and indicator during full-load testing of hybrid dyer was 4.1 kWh as compared to 8 kWh for electricity-based dryers. Electricity consumption is higher in starting as compared to during 13:00 h. as in initial hours water temperature in collector tank unit is low, due to this heat requirements are met through electric heater. Up to 11 a.m., electricity is consumed by the electric heater as heat is provided through heater as temperature inside dryer is less than 60 °C, after that electricity is only consumed by motor so electricity consumption decreases slightly after that. Energy consumed in first two hours of operation by electric heater, motor, exhaust fan, temperature controller and indicator is 1.5 kW. Hence, electricity consumed per hour will be 0.75 kW. If the system is fully operated on electric heater then it will consume about 2.0 kW/day. In the present study, the heater is operated only for two hours and it consumes about 0.5 kW energy revealing that electric heater consume 0.25 kW/h. Hence, about 50% of electricity can be saved using hybrid drying system and this is clear from Fig. 8.4.

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Fig. 8.4 Pattern of electricity consumption

Table 8.4 Economic indicator for hybrid dryer

S. No.

Economic indicator

Value

1

Net present worth

Rs. 1,55,456

2

B/C ratio

2.22

3

Payback period

2.08 yr

8.3.5 Techno Economics of Hybrid Dryer The techno economics were carried out on the basis of assumptions made in Sect. 8.2.4 and as described by Seveda et al. [17]. The various economic indicators are presented in Table 8.4. The benefit cost ratio was found to be 2.22 with a payback period of 2.08. It can be inferred that the developed dryer is technically as well as economically feasible. The higher percentage of the internal rate of return, i.e. 46.98% indicated the good economical return of the investment.

8.3.6 Quality Analysis of Dehydrated Oyster Mushroom Rehydration Studies The five grams sample of selected dehydrated oyster mushroom flakes was rehydrated by steeping them in cold water for 10 min. The weight of sample before and after rehydration was measured. It may be noted that higher rehydration ratio indicates better product [10]. The rehydration ratio and coefficient of rehydration (COR) were determined as explained in Sect. 8.3.6. The minimum rehydration ratio is obtained for sun dried sample,

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Table 8.5 Reconstitution capacity test S. No.

Type of mushroom sample

Dehydrated weight (g)

Rehydrated weight (g)

Rehydration ratio

Coefficient of rehydration (COR)

1

Hybrid dried

5

9.5

1.9

0.38

2

Sun dried

5

8.0

1.6

0.31

3

Commercially dried

5

8.3

1.7

0.29

Table 8.6 Organoleptic evaluation of dehydrated mushroom flakes S. No.

Type of mushroom samples

Average score of colour

Average score of taste

Average score of flavour

Overall acceptability

1

Hybrid dried

8.2

8.5

8.7

8.4

2

Sun dried

7.1

7.0

6.9

7.0

3

Commercially dried

6.9

7.2

7.4

7.2

and minimum coefficient of rehydration is obtained for commercially dried sample. Table 8.5 shows the results of reconstitution capacity test. It was observed that the rehydrated mushrooms had good appearance; even so they did not recover the same appearance of fresh mushrooms, because they were not completely rehydrated and suffered some changes under the drying conditions. Organoleptic Evaluation of Dehydrated Mushroom Flakes The colour, taste and flavour of dehydrated oyster mushroom flakes were tested with the help of consumer panel of 10 judges for all the samples. The 9 point hedonic scale was used which ranges from like extremely (9) to unlike extremely (1). All the samples were presented before the consumer panel along with commercially available dehydrated mushroom samples. The samples scored more than six values were adjudged good [16]. Table 8.6 shows the result of the consumer panel. It was found that commercially available dehydrated samples and sun dried samples scored less than the samples obtained in the case study.

8.4 Conclusions With the help of hybrid drying technology, the available solar energy can be used to get constant drying air temperature preventing any food spoilage and quality deterioration. This novel concept helps in maintaining a working temperature of drying air of about 60 °C inside the drying chamber as required for drying of food products. The total saving of 50% of electricity by using hybrid drying technology

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proved it a beneficial system over other drying technologies. Three kilogram of fresh mushroom flakes were reduced to 0.54 kg after dehydration. The techno-economic indicators of the developed dryer showed that it is technically as well as economically feasible with a high percentage of the internal rate of return, i.e. 46.98% indicating good economical return of investment. The rehydrated mushrooms had good appearance; even so they did not recover the same appearance of fresh mushrooms. The higher rehydration ratio of the hybrid dried mushroom flakes indicates it a better dried product than the commercially available and the sun dried products. Also the commercially available dehydrated samples and sun dried samples were found to score less than the samples obtained in the custody in case of organoleptic evaluation. Acknowledgements The authors are grateful to the Dean, College of Technology and Engineering, Udaipur, for providing all sorts of facilities required for the study.

References 1. B.K. Mehta, S.K. Jain, G.P. Sharma, A. Doshi, H.K. Jain, Cultivation of button mushroom and its processing: a techno-economic feasibility. Int. J. Adv. Biotechnol. Res. 2(1), 201–207 (2011) 2. https://www.slideshare.net/ANILRAUT1/anil-thesis-on-mushroom. Assessed on 3 June 2017 3. F. Fumagalli, A.M. Silveira, Quality evaluation of microwave-dried Packham’s triumph pear. Drying Technol. 23, 2215–2226 (2005) 4. K.O. Falade, O.J. Solademi, Modeling of air drying of fresh and blanched sweet potato slices. Int. J. Food Sci. Technol. 45, 278–288 (2010) 5. A.H. Patel, S.A. Shah, H. Bhargav, Review on solar dryer for grains, vegetables and fruits. Int. J. Eng. Res. Technol. 2(1) (2013) 6. A. Nagori, Design and development of hybrid dryer based on solar and electrical energy. https:// www.scribd.com/document/258679212/5-Introduction.2013. Assessed on 2 June 2017 7. S. Singh, S. Kumar, Testing method for thermal performance based rating of various solar dryer designs. Sol. Energy 86, 87–98 (2012) 8. I. Doymaz, Drying kinetics and rehydration characteristics of convective hot-air dried white button mushroom slices. J. Chem. 8 (2014) (Article ID 453175) 9. D. Kumar, Lalmani, R. Singh, Performance study regarding dehydration and quality characteristics of mushroom (Pleurotus ostreatus). Agric. Sci. Res. J. 4(2), 39–43 (2014) 10. M. Kulshreshtha, A. Singh, D. Vipul, Effect of drying conditions on mushroom quality. J. Eng. Sci. Technol. 4(1), 90–98 (2009) 11. P.J. Kanu, E.H. Sandy, B.A. Kandeh, J.Z. Bahsoon, Z. Huiming, Production and evaluation of breakfast cereal-based porridge mixed with sesame and pigeon peas for adults. Pak. J. Nutr. 8, 1335–1343 (2009) 12. J. Burguillos, J.C. Elauria, I. DeVera, Design, development and performance testing of a novel indirect solar dryer. J. Nat. Stud. 15(1), 1–18 (2016) 13. M.A. Leon, S. Kumar, S.C. Bhattacharya, A comprehensive procedure for performance evaluation of solar food dryers. Renew. Sustain. Energy 6, 367–393 (2002) 14. https://www.ibm.com/support/knowledgecenter/SSGMGV_3.1.0/com.ibm.cics.ts31.doc/ dfht3/dfht32w.htm. Assessed on 6 June 2017 15. T.O. Aduewa, A.S. Ogunlowo, S.T. Ojo, Development of hot-air supplemented solar dryer for white Yam (Dioscorea rotundata) slices. J. Agric. Vet. Sci. 7(12), 114–123 (2014)

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16. S. Upendra, S.K. Jain, R.C. Verma, A. Doshi, M.K. Jaipal, Dehydration characteristics and quality analysis of button mushroom slices (Agaricus bisporus). Agric. Eng. Today 31(3, 4), 43–46 (2007) 17. M.S. Sevada, N.S. Rathore, Techno-economics of solar tunnel dryer—a case study. J. Agric. Eng. 41, 12–17 (2004)

Chapter 9

Wind Speed Forecasting Using New Adaptive Regressive Smoothing Models Parikshit G. Jamdade, Prasad A. Godse, Prathamesh P. Kulkarni, Sujay R. Deole, Sudesh S. Kolekar and Shrinivas G. Jamdade

Abstract This paper proposes two new adaptive regressive smoothing models for wind speed forecasting. Generally, wind speed forecast (WSF) largely depends on the numerical models used which contribute the most error in WSF. This paper first uses most commonly used moving average and regressive smoothing models for WSF. Second, two adaptive models are proposed for WSF. In this paper, model’s quality measures are calculated by using mean error (ME), mean absolute error (MAE), mean square error (MSE), mean absolute percentage error (MAPE) and root means square error (RMSE) statistics. Simulations are performed to compare different models. It proves that the proposed models can improve the wind speed forecasting by effectively identifying and adjusting the errors from wind speed. Keywords Wind speed forecast · Regressive smoothing models · Adaptive regressive smoothing models · Quality measures

9.1 Introduction The energy need of countries is increasing day by day. Majority of energy is generated by fossil fuels. But fossil fuels are causing environmental pollution and this leads to use of renewable energy sources more and more. Wind energy is cheapest and commercially viable option for energy generation form renewable energy sources. But wind power is highly random in nature, so when it penetrates, power of wind speed exceeds a certain value and it will affect the quality of electric power as well as power system operation. If we are able to forecast the wind speed and the wind generated output power more accurately, the power system scheduling can be done to adjust the scheduling plans promptly, which reduces the impact of wind power on P. G. Jamdade (B) · P. A. Godse · P. P. Kulkarni · S. R. Deole · S. S. Kolekar Department of Electrical Engineering, PVG’s College of Engineering and Technology, Pune, Maharashtra, India e-mail: [email protected] S. G. Jamdade Department of Physics, Nowrosjee Wadia College, Pune, Maharashtra, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_9

93

94 Table 9.1 Geographical position of selected locations

P. G. Jamdade et al. Location

Latitude

Longitude

Pune

73.8569

18.5204

Nashik

73.7898

19.9475

Ahmednagar

74.7496

19.0952

the electric power grid [1, 2]. It also lower downs the power system operating costs and increase wind power penetration power limit. To install wind power station, it is important to investigate and analyze the geographical positions of wind speed in the locations [3–6]. In this paper, monthly average wind speed data between the years 2012 and 2016 are used for three cities Pune, Ahmednagar and Nashik (Maharashtra, India) . Geographical positions of three locations are given in Table 9.1. The wind speed forecasting can effectively reduce the adverse effect of wind farm on power grid, and it will also ensure the safe and economic operations of the power system [5–8]. In this study, we have determined the wind speed forecasting of three geographical locations in India using adaptive regressive models. Moving average model, weighted moving average model [9], regressive smoothing model [10, 11], regressive smoothing with trend model [12, 13] and regressive adaptive smoothing models [14–16] were the methods used for forecasting. To evaluate model’s performance, mean error (ME), mean absolute error (MAE), mean square error (MSE), mean absolute percentage error (MAPE) and root means square error (RMSE) statistics were calculated [2, 17]. The paper unfolds as follows: Section 9.2 describes the brief information about the methodology used in this paper. Section 9.3 includes information about the quality measures used in this study. In Sect. 9.4, results and discussions are presented. Finally, Sect. 9.5 concludes the paper.

9.2 Methodology In this paper, we are using six different models for wind speed forecasting. These models are based on the time series of wind speed, and when they are extrapolated over a time period it can more precisely predict the future value. They naturally do not consider small factors affecting wind speed and hence can be accurate for some conditions, while for others these models can be adjusted to agree with them. Regression forecasting models are designed to predict the theoretical relationship between a dependent variable and a number of independent variables that are known or can be computed. They are statistical models. They can either be linear or multiple regression models.

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Model 1—Moving Average (MA) Model MA is one of the basic moving averages used for forecasting data series. A moving average (MA) is found by finding the current average wind speed over a set number of periods. Most often, the recent wind speed is used to compute the moving average in the time series instead of the entire historical values as it provides significant advantage while deducing the stochastic process. In this paper, three-month moving average is used to forecast the future value. The moving average model (MA) is calculated as Fn =

n 1 At n t=1

(9.1)

where n number of observation, At actual value at observation t. Model 2—Weighted Moving Average (WMA) Model This average is estimated by multiplying each of the previous month’s data by a certain weight. Therefore, it puts more weight on recent data and less weight on past data. The use of this type of moving average gives better volatility estimates than the simple moving average. The weighted moving average model (MA) is calculated as Ft+1 =

n i=1 wi dt−i+1 n i=1 wi

(9.2)

Model 3—Regressive Smoothing (RS) Model Regressive moving averages decrease the lag by applying more weight to recent wind speed values. The weighting applied to the most recent wind speeds depends on the number of periods in the moving average. This model is similar to moving average, but the moving average removes the older wind speeds as new values become available. This regression model calculates the average of all historical values, starting at the point you specify. In this paper, we are using regressive smoothing model of order one (model 3a), regressive smoothing model of order one with linear trend (model 3b). Ft+1 = α

t−1  n=0

where α

smoothing constant,

(1 − α)n yt−n + (1 − α)t y1

(9.3)

96

Ft

P. G. Jamdade et al.

forecast value for observation t. For calculating linear trend, following expression is used. Tt = β(Ft − Ft−1 ) + (1 − β)Tt−1

(9.4)

where β smoothing constant for trend, T t smoothed trend for observation t. Model 4—Adaptive 1-Regressive Smoothing (ARS) Model In this paper, adaptive regressive smoothing model is defined as per Eq. 9.5 of order one. Ft+1 = α

t−1 

(1 − α)n yt−1 + (1 − α)t (yt−1 + (yt−1 − yt )

(9.5)

n=0

Model 5—Adaptive 2-Regressive Smoothing (ARS) Model On similar terms with respect to the above model, an alternate adaptive regressive smoothing model is defined as per Eq. 9.6 which is also of order one. Ft = α

t 

(1 − α)n yt + (1 − α)t (yt + (yt − yt−1 )

(9.6)

n=0

9.3 Quality Measures In order to compare quality of the above-proposed models against the old classical models, several quality measures are used. To measure the accuracy of the above models, the quality measures used are the mean error (ME) (Eq. 9.7), the mean absolute error (MAE) (Eq. 9.8), mean square error (MSE) (Eq. 9.9), moving average percentage error (MAPE) (Eq. 9.10), root mean square error (RMSE) (Eq. 9.11). Mean error near-zero shows there is not a bias in forecast. MAE shows the proximity between the forecast and actual value at observation. MAPE shows the percentage error in forecast. RMSE says there is not a bias in errors and fits a normal distribution. These measures are denoted by the following equations. ME =

n 1 et n t=1

(9.7)

9 Wind Speed Forecasting Using New Adaptive Regressive Smoothing …

MAE =

97

n 1 |et | n t=1

(9.8)

n 1 2 e n t=1 t  n  100   et  ( ) MAPE = n t=1  At    n 1  RMSE =  et2 n t−1

MSE =

(9.9)

(9.10)

(9.11)

where n At Ft et

number of observation, actual value at observation t, forecast value for observation t and error value at observation t.

9.4 Results and Discussions The descriptive statistics for three locations are shown in Table 9.3. The forecast models are given in Figs. 9.1, 9.2 and 9.3 for Pune, Nashik and Ahmednagar locations, respectively. 16

AV Method 3a Method 5

14

Method 2 Method 3b

Method 1 Method 4

Wind Speed in m/s

12 10 8 6 4 2 0

0

10

20

30

Month Fig. 9.1 Wind speed forecast for Pune location

40

50

60

98

P. G. Jamdade et al. 14

AV Method 3a Method 5

12

Method 1 Method 4

Method 2 Method 3b

Wind Speed in m/s

10 8 6 4 2 0 0

10

20

30

40

50

60

Month Fig. 9.2 Wind speed forecast for Nashik location 10

AV Method 3a Method 5

9

Wind Speed in m/s

8

Method 1 Method 4

Method 2 Method 3b

7 6 5 4 3 2 1 0

0

5

10

15

20

25

Month Fig. 9.3 Wind speed forecast for Ahmednagar location

Figures 9.1, 9.2 and 9.3 present the variational values in comparison with the actual data values to indicate the trend in variation of wind speeds. These figures show the cognizance of the proposed adaptive models in tracking the time series of the actual data. In this paper, monthly averaged wind speed data are used in developing the prediction models for the next time interval. For a good forecasting model, quality measure values should be low. In Table 9.2, the mean error (ME), the mean absolute error (MAE), mean square error (MSE),

9 Wind Speed Forecasting Using New Adaptive Regressive Smoothing …

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Table 9.2 Quality measures Location—Pune Quality measure

Method 1

Method 2

Method 3a

Method 3b

Method 4

Method 5 −0.006

ME

2.658

2.656

−0.058

−0.001

−0.001

MAE

3.032

2.967

0.237

1.969

0.195

0.173

MSE

14.216

13.455

0.299

6.301

0.059

0.050

MAPE

26.757

30.442

−5.018

−17.305

−0.579

1.229

RMSE

3.771

3.668

0.547

2.510

0.244

0.224

Method 4

Method 5

Location—Ahmednagar Quality measure

Method 3a

Method 3b

ME

Method 1 1.882

Method 2 1.885

−0.038

−0.006

0.001

−0.003

MAE

2.032

2.000

0.150

1.239

0.121

0.109

MSE

7.515

7.234

0.078

1.239

0.121

0.109

MAPE

32.422

33.729

−1.868

−8.777

−0.563

0.691

RMSE

2.741

2.689

0.280

1.113

0.348

0.330

Method 3a

Method 3b

Method 4

Method 5

−0.026

−0.003

−0.001

3.059

0.279

0.274

Location—Nashik Quality measure

Method 1

Method 2

ME

2.592

2.593

−0.0396

MAE

2.905

2.862

0.315

MSE

13.557

12.946

0.213

13.639

0.119

0.112

MAPE

24.524

24.524

−5.618

−48.138

−4.185

4.252

RMSE

3.682

3.598

0.461

3.693

0.345

0.334

moving average percentage error (MAPE), root mean square error (RMSE) are used for comparative analysis. Based upon the data in Table 9.3, it can be said that adaptive 1 and adaptive 2 models are the best fit for all locations as they shown lesser values of errors and quality measures (ME, MAE, MSE, MAPE and RMSE). For location Pune, reductions in ME, MAE, MSE, MAPE and RMSE values are 0.75%, 5.17%, 0.351%, 1.905% and 5.93%, respectively, for adaptive models. For location Nashik, reductions in ME, MAE, MSE, MAPE and RMSE values are 0.05, 9.44, 0.82, 17.06 and 9.08%, respectively, for adaptive models. For location Table 9.3 Statistics for three locations

Location Mean

Pune 5.565

Nashik 5.4583

Ahmednagar 4.6542

Maximum

13.8

12.8

9

Minimum

0.7

0.5

1.4

Std. Dev.

2.9009

2.5825

2.0099

Kurtosis

0.3570

0.2648

0.6539

Median

5.7

5.4

4.5

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Ahmednagar, reductions in ME, MAE, MSE, MAPE and RMSE values are 0.015%, 5.36%, 1.45%, 1.668% and 12.05%, respectively, for adaptive models.

9.5 Conclusions Wind speed forecasting is a key element for wind energy production. Wind speed forecasting is a difficult task due to its intermittence, high variability and nonlinearity. This paper proposes two new adaptive regressive smoothing models for WSF. As compared to old classical models like three-month moving average, three-month weighted moving average and regression smoothing models, the newer adaptive 1 and 2 models provide better accuracy. The proposed models can improve the wind speed forecasting by effectively identifying and adjusting the errors from wind speed.

References 1. D. Ambach, C. Croonenbroeck, Space time short to medium term wind speed forecasting. Stat. Methods Appl. 25(1), 5–20 (2016) 2. B. Doucoure, K. Agbossou, A. Cardenas, Time series prediction using artificial wavelet neural network and multi-resolution analysis: application to wind speed data. Renew. Energy 92, 202–211 (2016) 3. J. Wu, J. Wang, D. Chi, Wind energy potential assessment for the site of Inner Mongolia in China. Renew. Sustain. Energy Rev. 21, 215–228 (2013) 4. C. Wan, Z. Xu, P. Pinson, Z.Y. Dong, K.P. Wong, Probabilistic forecasting of wind power generation using extreme learning machine. IEEE Trans. Power Syst. 29(3), 1033–1044 (2014) 5. N. Chen, Z. Qian, I. Nabney, X. Meng, Wind power forecasts using gaussian processes and numerical weather prediction. IEEE Trans. Power Syst. 29(2), 656–665 (2014) 6. S.G. Jamdade, P.G. Jamdade, Evaluation of wind energy potential for four sites in Ireland using Weibull distribution model. J. Power Technol. 95(1), 48–53 (2015) 7. H. Quan, D. Srinivasan, A. Khosravi, short-term load and wind power forecasting using neural network based prediction intervals. IEEE Trans. Neural Netw. Learn. Syst. 25(2), 303–315 (2014) 8. T. Hong, P. Wang, L. White, Weather station selection for electric load forecasting. Int. J. Forecast. 31(2), 286–295 (2015) 9. S.D. Kwon, S.D. Kwon, Uncertainty analysis of wind energy potential assessment. Appl. Energy 87, 856–865 (2010) 10. A.D. Papalexopoulos, T.C. Hesterberg, A Regression based approach to short-term system load forecasting. IEEE Trans. Power Syst. 5(4), 1535–1550 (1990) 11. J. Wu, J. Wang, H. Lu, Y. Dong, X. Lu, Short term load forecasting technique based on the seasonal exponential adjustment method and the regression model. Energy Convers. Manag. 70, 1–9 (2013) 12. S.G. Jamdade, P.G. Jamdade, Extreme value distribution model for analysis of wind speed data for four locations in Ireland. Int. J. Adv. Renew. Energy Res. Ger. 1(5), 254–259 (2012) 13. S.G. Jamdade, P.G. Jamdade, Analysis of wind speed data for four locations in Ireland based on Weibull distribution’s linear regression model. Int. J. Renew. Energy Res. Iran 2(3), 451–455 (2012) 14. M. Dong, C. Lou, adaptive electric load forecaster. Tsinghua Sci. Technol. 20(2), 164–174 (2015)

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15. C.D. Zuluaga, M.A. Alvarez, E. Giraldo, Short-term wind speed prediction based on Robust Kalman filtering: an experimental comparison. Appl. Energy 156, 321–330 (2015) 16. G. Santamaría-Bonfil, A. Reyes-Ballesteros, C. Gershenson, Wind speed forecasting for wind farms: a method based on support vector regression. Renew. Energy 85, 790–809 (2016) 17. S. Makridakis, S.C. Wheelwright, R.J. Hyndman, Forecasting: Methods and Applications, 3rd edn. (Wiley Publication, New York, 1998)

Chapter 10

Thermal Performance Analysis of a Heat Pump-Based Photovoltaic/Thermal System S. Vaishak and Purnanand V. Bhale

Abstract Increased energy consumption and environmental pollution have necessitated the use of renewable energy sources. Among the various renewable energy sources, solar energy seems to be the most promising one that can meet the energy demand in the nearby future. Photovoltaic (PV) is the best-known method for generating the electricity from solar, and these modules have an efficiency in the range of 6–18%. The efficiency of the PV module mainly depends on the cells semiconductor material, irradiation, and temperature. For effectively utilizing the available solar radiation and to improve the performance of the module, it is required to extract the heat accumulated on the PV module and use it for other applications. It was identified that a photovoltaic/thermal (PV/T) system which simultaneously produced heat and electricity have greater potential to be used as a hybrid system for domestic and commercial applications. Based on this, here the thermal performance of a heat pump-based PV/T was evaluated and analyzed via a steady-state numerical model. Keywords Photovoltaic/thermal system · Heat pump · Solar assisted heat pump

10.1 Introduction Among the various diversifying technology to utilize solar energy, photovoltaics and solar assisted heat pump are identified as two promising technologies. For a PV technology, conversion efficiency decreases with increase in temperature, whereas heat pump operates more efficiently at higher evaporator temperature. Hence, these two technologies can be favorably integrated to increase the performance of both systems and to improve the utilization of available solar irradiance and space. During operation, a portion of the solar radiation is converted to electricity and the major part goes to thermal energy. The latter serves as a heat source for refrigerant evaporation a process that lowers more considerably the operating temperature of the PV. S. Vaishak · P. V. Bhale (B) Sardar Vallabhbhai National Institute of Technology, 395007 Surat, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_10

103

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For analyzing the system performance, a number of numerical simulation and experimental works have been conducted and reported by various researchers. For simulating the performance of a dry expansion evaporator, a distributed modeling has been widely used [1–3]. Later a generalized distributed parameter model was developed by Tso et al. [4] to analyze both the steady and the transient behavior of evaporators. Ji et al. [5] also developed a dynamic distributed parameter model of their system. Chen and Wei [6] also done a numerical analysis of their novel system by incorporating a PV/T solar collector acting as the evaporator. Furthermore, Tsai [7] proposed a mathematical model in MATLAB/Simulink environment by considering the reciprocal energy exchange between the heat pump and PV evaporator. More recently, Gunasekar et al. [8] developed an artificial neural network to predict the energy performance of the similar systems. The study carried out by Zondag et al. [9] showed that a one-dimensional model is reasonably accurate when compared to other models considered for the study. Here, the thermal performance of a heat pump-based PV/T was evaluated and analyzed via a steady-state numerical model.

10.2 System Description and Working Principle The proposed system is a heat pump-based PV/T system based on the fact that a heat pump operates more efficiently at higher temperature whereas PV module operates more efficiently at lower temperature. The novelty of the proposed system lies in the design of the PV/T evaporator where a glass to glass PV structure is used instead of a glass to tedlar PV structure which is commonly used in various literatures for the design of the PV evaporator. In some studies, the tedlar is directly replaced by thermal absorber in order to improve the thermal efficiency of the system. But in such cases, the problem of delamination (because of uneven thermal expansion coefficient of various layers of PV evaporator) and hence the degradation of PV module will be severe which reduces the reliability of the system. This issue is not addressed in most of the literature. In this context, a glass to glass PV module design provides better heat management and longer panel life due to less delamination risk. Also, the tensile strength of this module is superior in relation to typical modules with back sheet. In addition to this, with the glass on both sides, the glass to glass modules have much better moisture barrier properties as compared to glass-back sheet modules. It was reported that for an air-based PV/T glass to glass arrangement gives better thermal and electrical performance than a glass to tedlar arrangement. Glass to glass PV modules can be integrated with a heat pump system, by laying the evaporator coils beneath the PV panels. By doing so, a higher evaporator temperature is obtained because the radiation falling on non-packing area of glass to glass module is transmitted through the glass cover and falls directly on the thermal absorber. However, in case of glass to tedlar all the radiation is absorbed by the tedlar and then heat is carried away by the conduction to the thermal absorber which acts as the heat source for heat pump. During operation, the refrigerant from the condenser passes through an expansion valve directly into the PV evaporator where it gets

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105

Table 10.1 Physical properties of different layers of PV evaporator Material Front glass EVA Cell Back glass

τ

A

δ (m)

0.7

0.91

0.05

0.005

0.35

0.9

0.08

0.0005

0.09

0.8

0.0003

0.7

0.05

0.005

k(W/m°C)

148 0.7

Table 10.2 Specifications of the system

Area of collector, Ac

1.3 m2

Inside diameter of refrigerant tube, Di

8 mm

Length of evaporator tube, L

52 m

Packing factor, b

0.9

Cooling capacity of compressor

944 W

Displacement of compressor

9.42 CC/rev

Condenser capacity

150 L

Length of condenser coil

15 m

evaporated by incident solar energy. A portion of the solar radiation is converted into electricity by the PV cells and the major part goes to the refrigerant as thermal energy. This process lowers the operating temperature of the PV cells considerably when compared to conventional PV modules. On the other hand, because of direct insolation, the evaporator temperature will be higher than the conventional air source heat pump. Sometimes the ambient air also acts as an additional heat source. The lowpressure vapor refrigerant is then compressed in the compressor. Following which this high-pressure vapor refrigerant is finally condensed in the immersed coil watercooled condenser. Detailed specifications of the system are given in Tables 10.1 and 10.2.

10.3 Mathematical Model The following assumptions are made for evaluating the thermal performance of the system: • For the simulation period, the system is considered to be in quasi steady state. • Pressure drop in various pipelines of the system is neglected. • Condition of the refrigerant is considered to be saturated at the exits of evaporator and condenser. • Refrigerant compression is assumed to be polytropic. • Refrigerant expansion in the capillary tube is considered to be isenthalpic. • Water tank is assumed to be non-stratified.

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Following the above assumptions, the governing equations for each component of the proposed system are as below.

10.3.1 PV Evaporator Based on principle of conservation of energy, the energy balance for each component is as follows:    k f g × T f g − T f EVA  (10.1) α f g × I = h f ga × T f g − Ta + δfg     k f g × T f g − T f EVA b × kEVA × T f EVA − Tcell = (10.2) αEVA × τ f g × I + δfg δEVA   b × kEVA × T f EVA − Tcell = ηcell × αcell αcell × τ f g × τEVA × I + δEVA b × kcell × (Tcell − TbEVA ) × τEVA × τ f g × I + (10.3) δcell b × αEVA × τ f g × τcell × τEVA × I + (1 − b) × αEVA × τ f g × τEVA × I   kEVA × TbEVA − Tbg b × kcell × (Tcell − TbEVA ) = (10.4) + δcell δEVA b × αbg × τ f g × τEVA × τcell × τEVA × I + (1 − b) × αbg × τ f g × τEVA × τEVA × I     kbg × Tbg − Tab kEVA × TbEVA − Tbg = (10.5) + δEVA δbg b × αab × τ f g × τEVA × τcell × τ E V A × τbg I + (1 − b) × αab × τ f g   kbg × Tbg − Tab × τEVA × τEVA × τbg × I + = h r × (Tab − Tr ) δbg (10.6) where α is absorptivity, τ is transmissivity, h is convective heat transfer coefficient (W/m2 K), k is thermal conductivity of material (W/mK), δ is the thickness of material layer (m), I is the solar radiation, b is the packing factor of PV cell, and η is the efficiency of cell. The subscript a, ab, bEVA,bg, cell, f EVA, f g, r corresponds to ambient, absorber, back EVA, back glass, cell, front EVA, front glass, and refrigerant, respectively.

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For two phase flows in horizontal pipes, the heat transfer coefficient (h r ) is given by: hr =

0 × 0082 × kl × (Re2 × J × x × h f g /L e )0×40 Di

(10.7)

where J is 778 (dimensional constant), x is the quality change of the refrigerant in the evaporator, and h f g is the latent heat. The evaporator heat gain (Q u ): Q u = h r × At × (Tab − Tr )

(10.8)

where At (m2 ) is the evaporator surface area. The same can be expressed as the change in enthalpy of refrigerant in the evaporator: Q e = m˙ r × (h 1 − h 4 )

(10.9)

where m r (kg/s) is the mass flow rate of the refrigerant.

10.3.2 Compressor For a constant speed compressor, the mass of refrigerant circulated is given as m˙ r =

Vd × ηv V1

(10.10)

where the displacement volume (Vd ) can be expressed as: Vd =

i × π × Db2 × S × N 4 × 60

(10.11)

where i is the number of compressor bores, Db is the bore diameter, S is the stroke length, and N is the speed of the compressor. The volumetric efficiency (ηv ) of the compressor for this simulation is assumed to be 70%. The compression work can be calculated as given below: P1 n Wc = m˙ r ηv n − 1



P2 P1

(n−1)/n

 −1

(10.12)

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10.3.3 Condenser The total heat rejection in the condenser is given by: Q C = ACo × UCo × (TC − TW )

(10.13)

where ACo is the heat transfer area of the condenser coil and UCo denotes the overall heat transfer coefficient, which is given by: 1 U co =  

1 B hW + h Co

(10.14)

where B = ACo /ACi . The heat transfer coefficient (h Co ) can be estimated using the following equation: h Co = 0.0265

kl di



G r di μl

0.8 

C pr μl kl

0.3 (10.15)

The heat transfer coefficient on water side (h W ) is given by: h W = 0.5

kw dO



gβw td O3 ρW C pw μw kw

 (10.16)

The condenser heat gain can also be expressed as Q c = Q e + Wc

(10.17)

From the energy balance of a non-stratified tank, Mw C pw

dTw = Qc dt

(10.18)

Finally, the COP of the system is given as COP =

Qc Wc

(10.19)

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109

10.4 System Simulation and Result Analysis 10.4.1 Computation Procedure A Commercially available software was used to determine the performance of the system based on the above detailed analysis of the each component. Initial temperature of the water in the storage tank and the meteorological conditions were the datas inputted for the simulation. The properties of the chosen working fluids were directly available with the software. In order to solve the numerical model, the initial temperature of the refrigerant in the condenser and evaporator was assumed. Then the steady-state condition for the evaporator and condenser was checked. If this condition was satisfied, then the program evaluates the performance of the system, otherwise, the program was repeated with updated values of above temperature until the steady-state conditions were achieved. The weather dataset of Ahmedabad (Gujarat) for November 2, 2016, was used for the simulation study. Figure 10.1 shows the variation of ambient temperature and total amount of solar irradiance on the tilted surface.

Fig. 10.1 Variation of solar irradiance and ambient temperature

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10.4.2 Results and Discussion Variation of Evaporator and Condenser Heat Gain Figure 10.2a shows the variation of evaporator and condenser heat gain with time. It can be observed that evaporator heat gain has a direct relationship with solar irradiance and ambient temperature. For the simulated period, it increases from 772.3 W at 09.00 to a maximum of 937.1 W at 11.00. Even though at 12.00 noon the solar irradiance and ambient temperature was higher than that at 11.00, the evaporator heat gain was found to be slightly less at 12.00 noon when compared to that at 11.00. This is because of the reduced heat gain from the ambient due to reduced wind velocity at 12.00 noon. Increased evaporator heat gain increases the COP system for a constant condenser operating condition. It was also identified that a 10 °C evaporator temperature makes the PV to operate at their ideal operating temperature of 25 °C. It can be seen that the condenser capacity increases with increase in solar radiation and ambient temperature. Condenser capacity was 946.3 W in the morning when the solar radiation and ambient temperature were 561.4 W/m2 and 29.2 °C, respectively. At 12.00 noon, the condenser capacity increased up to a maximum value of 1257 W. Even though the solar radiation was relatively same at 9.00 and 14.00, high ambient temperature results in higher condenser capacity in the afternoon periods. Variation of Compressor Power Figure 10.2b shows the variation of compressor power during the simulated period. During the simulated period, the compressor power increased from 221.6 to 342.4 W. The increase in compressor power is mainly due to the increase in condenser water temperature with time. The increase in condenser temperature leads to a higher compressor power consumption because of increased compression ratio. Ji et al. [4] reported that the power consumption of compressor reduces mildly with the rise in solar radiation, but the influence of this parameter is less when compared to condenser water temperature.

Fig. 10.2 a Variation of evaporator and condenser heat gain with time. b Variation of compressor power consumption with time

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111

Fig. 10.3 a Variation of COP with time. b Variation of condenser water temperature with time

Variation of Heat Pump COP Figure 10.3a shows the variation of COP of the system with time. The average value of COP was found to be 4.25 with a maximum value of 5.35 at 09.00 and a minimum value of 2.96 at 15.00, respectively. It can be observed that as time progresses, COP decreases. This is because of increase in condenser temperature which results in a rise of compressor power. With constant condenser supply water temperature, it is expected that the COP of the system increases with increase in solar irradiance and ambient temperature. Variation of Condenser Water Temperature Figure 10.3b shows the variation of condenser water temperature with time. During the simulated period, the condenser water temperature was found to be increased from 25 to 68 °C. At least 40 °C water is required to support practical heating applications. Rise in water temperature will result in increase in compressor power consumption which will in-turn reduces the COP of the system.

10.5 Conclusion It was identified that a heat pump-based PV/T can significantly increase utilization of solar heat energy and improve the electrical efficiency of the PVs. This work proposed an experimental setup for the same for analyzing the actual performance of the system under the climatic conditions of India. Via numerical simulation of the system, the performance of the proposed system was evaluated and analyzed. The following conclusions can be deducted: • The evaporator heat gain and condenser capacity increase with the increase in solar irradiance, ambient temperature, and wind velocity. • A mild decrease in compressor power was observed with increase in solar radiation increases, but the same was greatly affected by condenser water temperature.

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• COP decreases gradually with the rise in condenser water temperature. • During operation, the cell temperature can go below the ambient temperature when the heat load on the evaporator is low. This may result in the formation of condensate on the panel surface.

References 1. J. MacArthur, E. Grald, Unsteady compressible two-phase flow model for predicting cyclic heat pump performance and a comparison with experimental data. Int. J. Refrig 12(1), 29–41 (1989) 2. H. Wang, S. Touber, Distributed and non-steady-state modelling of an air cooler. Int. J. Refrig. 14(2), 98–111 (1991) 3. T.T. Chow, K.F. Fong, G. Pei, J. Ji, M. He, Potential use of photovoltaic-integrated solar heat pump system in Hong Kong. Appl. Therm. Eng. 30(8–9), 1066–1072 (2010) 4. X. Jia, C.P. Tso, P. Jolly, Y.W. Wong, Distributed steady and dynamic modelling of dry-expansion evaporators. Int. J. Refrig 22, 126–136 (1999) 5. J. Ji, H. He, T. Chow, G. Pei, W. He, K. Liu, Distributed dynamic modeling and experimental study of PV evaporator in a PV/T solar-assisted heat pump. Int. J. Heat Mass Transf. 52(5–6), 1365–1373 (2009) 6. H. Chen, P. Wei, Numerical study on a novel photovoltaic/thermal heat pump system. Energy Procedia 547–553 (2011) 7. H.-L. Tsai, Design and evaluation of a photovoltaic/thermal-assisted heat pump water heating system. Energies 7(5), 3319–3338 (2014) 8. N. Gunasekar, M. Mohanraj, V. Velmurugan, Artificial neural network modeling of a photovoltaic-thermal evaporator of solar assisted heat pumps. Energy 93, 908–922 (2015) 9. H.A. Zondag, D.W. de Vries, W.G.J. van Helden, R.J.C. van Zolingen, A.A. van Steenhoven, The yield of different combined PV-thermal collector designs. Sol. Energy 74(3), 253–269 (2003)

Chapter 11

Overall Performance of N Partially Covered Photovoltaic Thermal-Compound Parabolic Concentrator (PVT-CPC) Collector with Different Concentration Ratio Rohit Tripathi, Abhishek Tiwari and G. N. Tiwari Abstract In the present paper, overall gain performance of 1/4 covered seriesconnected N identical photovoltaic thermal-compound parabolic concentrators (PVT-CPC) collectors have been carried out at constant flow rate of water. Further, a comparison has been made on the basis of four different concentration ratio or cases (i–iv), respectively. The ratio of aperture and receiver, Aa : Ar has been considered as 1:1 [case (i)], 2:1 [case (ii)], 3:1 [case (iii)], and 4:1 [case (iv)]. It has been observed that case (iv) have most preferable to obtain maximum overall thermal energy and exergy, whereas case (i) is most suited for obtaining electrical gain and chosen for delivering minimum thermal gain. Case (i) is acting like a conventional N-PVT collector due to concentration property which is not working in collector system because aperture area is equal to receiver area. Case (ii) has been found to be best for overall exergy point of view due to lower input energy from case (iii–iv). Keywords PVT · Compound parabolic concentrator (CPC) · Concentration ratio (C)

R. Tripathi (B) · G. N. Tiwari Centre for Energy Studies, Indian Institute of Technology Delhi, Hauz Khas, New Delhi 110016, India e-mail: [email protected] R. Tripathi School of Electrical Electronics & Communication Engineering, Galgotias University, Greater Noida, U.P, India A. Tiwari University School of Information and Communication Technology, Guru Gobind Singh Indraprastha University, Sector 16C, Dwarka, New Delhi, Delhi 110078, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_11

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11.1 Introduction Due to limited availability of fossil fuels, present life is suffering the energy demand. To fulfil the energy demand, another way is renewable energy which is the only hope in the globe now. Renewable energy is mainly classified into solar, wind and biogas energy where prime choice and availability is solar energy only. It is environment friendly also which does not produce CO2 emissions during production. For producing the electrical energy, photovoltaic (PV) is the prime choice for users which produces electrical energy only. But, for producing electrical as well as thermal energy from solar, photovoltaic thermal (PVT) is the best choice [1–3]. The earliest design of PVT has been developed in 1978 [4]. In hybrid energy-based system, the PVT collector is designed in two parts: one is PV module and other is tube-in-plate configuration-based channel box which is directly exposed to the sun. Top surface of second part is covered by PV module. In this study, glass to glass or semitransparent PV module has been chosen for generating more overall energy. The received solar heat is directly transmitted to the black copper plate through semitransparent PV module and covered glass. This direct heat from non-packing of glass and indirect heat from back of solar cell increase the temperature of flowing fluid via absorber copper plate. A detailed study of PVT using techniques of convectional thermal plane collector has been presented [5]. The detailed study of this hybrid collector has been carried out on the effect of airflow and duct design [2, 6], respectively. Series-connected PV-based flat plate collector has designed concluded the results [7]. Two types of hybrid flat plate collector have been discussed, namely tube-in-plate configuration and parallel plate configuration. Tube-in-plate configuration and parallel plate configuration have been compared and concluded that tube-in-plate configuration has been chosen best for thermal gain [3, 8]. Experimental study of PVT system has been presented and concluded that it is best configuration for producing both energies simultaneously. The PVTs have been designed with glass and without glass covered for performance analysis [9]. Concentrator technique deals a best possibility of enhancing thermal gain. Less than four concentration ratio is known as low concentrator and CPC is best for low concentrator. It has essentially two half parabolas face-to-face configuration based. Basic modelling for hybrid flat plate collector and electrical efficiency have been developed and obtained a conclusion that glass to glass PV module gives higher efficiency [10]. It was also suggested that if the thermal performance was not totally important then the collector should not be covered by glass. If CPC is integrated over PVT collector then it is expected that the thermal gain can be enhanced, due to this, the overall energy from the system as PVT-CPC can also be enhanced. To mind this concept, the PVT-CPC has been designed for generating electrical and thermal energies. The analytical modelling of series-connected N identical hybrid collectors has been presented and discussed four different cases based on PV coverage area on collector [8, 10–13]. Further, validation has been performed for experimental and theoretical results for a fully covered PVT-CPC collector at New Delhi, India [14]. They have found fair agreement between the results. The life cycle cost analysis

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has been studied for N-PVT-CPC collector connected in series by using different fluids other than water [15–18]. In this study, energy matrices and exergoeconomic and carbon credits have been estimated for the proposed cases. In present paper, an attempt has been performed for obtaining higher thermal energy by considering different concentration ratio on N-PVT-CPC collector connected in series.

11.2 System Description The series connection of partially covered N-PVT-CPC collector has been presented in present study. The collector area is 1 m2 where PV module area is 0.25 m2 and glass area is 0.75 m2 . The side cut view section of first collector of partially covered NPVT-CPC collector has been shown in Fig. 11.1 [19]. All collectors are connected in cascaded pattern. The first outlet temperature of PV module in first collector (T fom1 ) handles into the inlet of the glass upper portion of first collector and the outlet water temperature of the first concentrated hybrid water collector is (T fo1 ). T fi is as inlet at first collector and T foN is handled as outlet at Nth collector. The aperture and receiver area of collector has been considered as Aa and Ar, respectively. Four cases that have been considered on the basis of concentration ratio of collector are as follows: case (i): concentration ratio (C) = 1, case (ii): C = 2, case (iii): C = 3, case (iv): C = 4. The solar irradiance falls on aperture area of collector. It reflects from both reflector, falls on receiver area. The direct gain of solar is accepted through the non-packing area of PV module and glass portion and comes to the black absorber copper plate. The indirect gain of cells is transferred to the absorber plate. So, temperature of absorber plate increases. It increases the temperature of the flowing water through tubes below the absorber plate. The design parameters for proposed system have been given in Table 11.1. X

X Absorption plate Glazed surface

Solar cell

Tfom1 Air gap

Inlet,

1

Outlet

X′

L

Xo′

Insulation Cut section of metallic tubes

Fig. 11.1 Cross section side view of 1/4 covered PVT-CPC first water collector

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Table 11.1 Details of basic specification of parameters for partially covered PVT-CPC collector Ar = 1 m2 Arm = 0.25

m2

Utc,a = 9.17 W/m2 °C

Utc,p = 5.58 W/m2 °C

UL,m = 7.87

Utp,a = 4.8 W/m2 °C

W/m2

Arc = 0.75 m2

PF1 = 0.3782

Aa = 2

°C

cf = 4179 J/kg K

PF2 = 0.9512

m˙ f = 0.018 kg/s

Aam = 0.5 m2

PFc = 0.9842

UL1 = 3.47 W/m2 °C

Aac = 1.5 m2

h pf = 100 W/m2

ULc = 4.7 W/m2 °C

F

UL,m = 7.87 W/m2 °C K i = 0.166 W/m °C

m2

h i = 5.7

W/m2

= 0.9680

Frc = 0.8693

m2

h i

= 5.8

W/m2

Frm = 0.8110

m2

h o = 9.5

W/m2

K g = 0.816 W/m °C

ρ = 0.84

K p = 6 W/m °C τg = 0.95

11.3 Modelling Certain assumptions have been taken to write the basic energy balance equations for proposed number of hybrid collectors connected in series as per [8]. 11.3.1 Energy balance equation for solar cell of semitransparent PV module (Fig. 11.1a, b)    ραc τg βc Ib Aam = Utc,a (Tc − Ta ) + Utc,p Tc − Tp Arm + ρηm Ib Aam

(11.1)

11.3.2 Energy balance for absorber plate below the photovoltaic module       ραc τg2 (1 − βc )Ib Aam + Utc,p Tc − Tp Arm = F  h pf Tp − Tf Arm + Utp,a Tp − Tf (11.2) 11.3.3 Energy balance for flowing water as fluid below the absorber plate

m˙ f cf

  dTf dx = F  h pf Tp − Tf Arm bdx dx

(11.3)

F  is the collector efficiency factor which is calculated by [1] and [4]. Now, the outlet water temperature at the end of very first collector of N-PVT-CPC collector has been following as Tfo1

      Ta (AFR UL )1 1 − K kN Ib (AFR (ατ ))1 1 − K kN (AFR UL )1 + + Tfi 1 − = m˙ f cf m˙ f cf m˙ f cf (1 − K k ) (1 − K k ) (11.4)

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By the method, the outlet water temperature at the end of Nth collector of PVTCPC is followed [10], from Eqs. 11.1–11.4. One can get outlet fluid temperature like (T foN ) as

TfoN

  N N Ib (AFR (ατ ))1 1 − K p Ta (AFR UL )1 1 − K p   +   + K pN Tfi = m˙ f cf m˙ f cf 1 − Kp 1 − Kp

(11.5)

With help of Eq. (11.5), the rate of thermal energy gain from partially covered N-PVT-CPC water e has been evaluated by following equation Q˙ uthe,N = m˙ f cf (TfoN − Tfi )

(11.6)

The rate of useful thermal exergy from the partially covered N-PVT-CPC water collector has been evaluated by following equation (TfoN + 273) Q˙ uthx,N = m˙ f cf (TfoN − Tfi ) − m˙ f cf (Ta + 273) ln (Tfi + 273)

(11.7)

The temperature-dependent electrical efficiency of solar cells in PV modules of partially covered N-PVT-CPC collectors has been evaluated by following equation    ηcN = η0 1 − β0 T¯c − T0

(11.8)

Equation (11.8) can be further rewritten as following

Now, ηcN =

η0 1 −

β0 {(XIb (Utc,p +Utc,a )

1−

 + YTa + ZTfi ) − T0 }  + βγ ) (α )

η0 β0 Ib (Utc,p +Utc,a

(11.9)

where X, Y, Z, α, β and γ are defined. The temperature-dependent electrical efficiency of PV modules of N-PVT-CPC water collector has been evaluated by following equation ηmN = τg βc ηcN

(11.10)

The rate of useful electrical gain from proposed system has been evaluated by following equation Q˙ xel,N = Arm Ib

N 1

ηmN

(11.11)

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The rate of an overall thermal energy gain is obtained from proposed system and can be calculated by following expression Q˙ xel,N Q˙ u,overall = Q˙ uthe,N + 0.38

(11.12)

The rate of an overall exergy is obtained from proposed system and can be evaluated by following expression Q˙ ux,overall = Q˙ uthx,N + Q˙ xel,N

(11.13)

11.4 Results and Discussions The solar irradiance (beam) and an ambient air temperature on horizontal surface for a clear day (type ‘a’) in January month for New Delhi, India have been obtained from IMD, Pune, India. The solar irradiance has been calculated at 30° on south facing for New Delhi location. The variation of solar irradiance and ambient air temperature with the time of the day has been shown in Fig. 11.2. The hourly variation of average solar cell temperature and electrical efficiency of PV module of all cases for proposed system have been shown in Fig. 11.3. Here, it is seen that the electrical efficiency of PV module is totally inversely proportional to the average solar cell temperature, as expected. It is observed that case (iv) has maximum solar cell temperature due to high concentration ratio or high input energy and minimum electrical efficiency, whereas minimum solar cell temperature and 1000

16

2

Beam radiation, I b (W/m )

Ib

15

Ta

14

800

13 700

12

600

11

500

10

400

9 8

300

6

100

o

7

200

Ambient air temperature, Ta ( C)

900

5 7

8

9

10

11

12

13

14

15

16

17

Time of the day (hr)

Fig. 11.2 Hourly variation of beam radiation and ambient air temperature of a typical day

o

200

η mN for C=3

η mN for C=4

average Tc for C=3

average Tc for C=4

119 0.14

180

0.12 for C=2 η

mN

120 100

0.08

60 40 20 7

0.10

Average Tc for C=1

80

average Tc for C=2

140

η mNfor C=1

160

8

9

10

11

12

13

14

15

16

0.06

Electrical efficiency of PV module (in fraction)

Avergae solar cell temperature, Tc ( C)

11 Overall Performance of N Partially Covered Photovoltaic Thermal …

0.04

17

Time of the day (hr)

Fig. 11.3 Hourly variation of average solar cell to electrical efficiency of PV module of N-PVT-CPC collector for four different cases

o

Outlet fluid temperature, TfoN ( C)

maximum electrical efficiency have been obtained for case (i) due to less input energy to other cases. Hourly variation of outlet fluid temperature at Nth collector of PVT-CPC has been shown in Fig. 11.4. Case (iv) has been found to achieve highest temperature for water due to maximum input energy or high concentration ratio. The maximum outlet water temperature has been obtained around 95 °C at 0.018 kg/s mass flow rate and four number of collectors (N = 4). Hourly variation of electrical gain, overall thermal energy and exergy have been shown in Figs. 11.5 and 11.6. Here, the maximum electrical gain has been found for case (i) due to lowest temperature of solar cell. The maximum overall thermal c=1 c=2 c=3 c=4

January, mf=0.018 kg/s, N=4

105 100 95 90 85 80 75 70 65 60 55 50 45 40 35 30 25 20 15 10 7

8

9

10

11

12

13

14

15

16

17

Time of the day (hr)

Fig. 11.4 Hourly variation of outlet fluid (water) temperature at Nth collector of N-PVT-CPC for all cases

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Electrical gain (kWh)

c=1 c=2 c=3 c=4

January, mf=0.018 kg/s, N=4

0.12 0.10 0.08 0.06 0.04 0.02 7

8

9

10

11

12

13

14

15

16

17

Time of the day (hr)

Overall thermal energy/exergy (kWh)

Fig. 11.5 Hourly variation of electrical gain at Nth collector of PVT-CPC for all cases overall thermal energy: C=1 C=2 C=3 C=4 Overall exergy: C=1 C=2 C=3 C=4

January, mf=0.018 kg/s, N=4

7 6 5 4 3 2 1 0 7

8

9

10

11

12

13

14

15

16

17

Time of the day (hr)

Fig. 11.6 Hourly variation of overall thermal energy and exergy for all cases of N-PVT-CPC collector

energy and exergy have been achieved for case (iv) only, due to maximum input solar energy on collector area.

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11.5 Conclusions Under study, significant outcomes of the work are concluded as following • The maximum outlet water temperature has been obtained around 95 °C at 0.018 kg/s mass flow rate and N = 6 in a clear day, month of January, New Delhi, India [14]. • Maximum electrical efficiency and electrical gain have been obtained for case (i) [concentration ratio (C) = 1:1] of N-PVT-CPC collector connected in series, due to lowest solar cell temperature [14]. • The overall thermal energy and exergy has been noted maximum for case (iv) [concentration ratio (C) = 4:1] of partially covered N-PVT-CPC collector, which 3.6, 1.8 and 2.9 times higher than for case (i), case (ii) and case (iii), respectively [8].

11.6 Units and Symbols

αc

Absorptivity of the solar cell

Tc

Solar cell temperature (°C)

m˙ f

Mass flow rate of water (kg/s)

Tp

Absorption plate temperature (°C)

τg

Transmissivity of the glass

Lp

Thickness of absorption plate (m)

β0

Temperature coefficient of efficiency (K−1 )

Kp

Thermal conductivity of plate (W/m K)

Cf

Specific heat of water (J/kg K)

Tfi

Inlet fluid (water) temperature (°C)

Utc,a

Overall heat transfer coefficient from cell to ambient (W/m2 K)

Tf

Flowing fluid (water) temperature (°C)

Utc,p

Overall heat transfer coefficient from cell to plate(W/m2 K)

Tfom

Outlet water temperature at the end of PV module (°C)

Arm

Area of receiver covered by PV module (m2 )

ηo

Efficiency at standard test condition

hi

Heat transfer coefficient for space between the glazing and absorption plate (W/m2 K)

Tfoc

Outlet water temperature at the end of portion covered by glass (°C)

ηc

Solar cell efficiency

TfomN

Outlet water temperature at the end of Nth PV module (°C)

ηm

PV module efficiency

TfoN

Outlet water temperature at the end of Nth PVT-CPC water collector (°C)

b

Breath of receiver (m)

h i

Heat transfer coefficient from bottom of PVT to ambient (W/m2 K)

Arc

Area of receiver covered by glass (m2 )

ho

Heat transfer coefficient from top of PVT to ambient (W/m2 K)

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References 1. A.K. Bhargava, H.P. Garg, R.K. Agarwal, Study of a hybrid solar system-solar air heater combined with solar cells. Energy Convers. Manag. 31(5), 471–479 (1991) 2. T.T. Chow, Performance analysis of photovoltaic–thermal collector by explicit dynamic model. Sol. Energy 75, 143–152 (2003) 3. S. Dubey, G.N. Tiwari, Analysis of PV/T flat plate water collectors connected in series. Sol. Energy 83, 1485–1498 (2009) 4. J. Prakash, Transient analysis of a photovoltaic-thermal solar collector for co-generation of electricity and hot air/water. Energy Convers. Manag. 35, 967–972 (1994) 5. E.C. Kern, M.C. Russell, Combined photovoltaic and thermal hybrid collector systems, in Proceedings of the 13th IEEE Photovoltaic Specialists (1978), pp. 1153–1157 6. G.N. Tiwari, Solar Energy: Fundamentals, Design, Modeling and Applications (Narosa Publishing House, New Delhi, India, 2016) 7. S.D. Hendrie, Evaluation of combined photovoltaic/thermal collectors, in International Conference ISES, vol. 3 (Atlanta, Georgia, USA, 2009), pp. 1865–1869 8. R. Tripathi, G.N. Tiwari, I.M. Al-Helal, Thermal modelling of N partially covered photovoltaic thermal (PVT)–Compound parabolic concentrator (CPC) collectors connected in series. Sol. Energy 123, 174–184 (2016) 9. G.N. Tiwari, R.K. Mishra, S.C. Solanki, Photovoltaic modules and their applications: a review on thermal modeling. Appl. Energy 88, 2287–2304 (2011) 10. R. Tripathi, G.N. Tiwari, Energetic and exergetic analysis of N partially covered photovoltaic thermal- compound parabolic concentrator (PVT-CPC) collectors connected in series. Sol. Energy 137, 441–451 (2016) 11. R. Tripathi, G.N. Tiwari, V.K. Dwivedi, Overall energy, exergy and carbon credit analysis of N partially covered photovoltaic thermal (PVT) concentrating collector connected in series. Sol. Energy 136, 260–267 (2016) 12. R. Tripathi, G.N. Tiwari, V.K. Dwivedi, Overall energy and exergy performance of partially covered N-photovoltaic thermal (PVT)-compound parabolic concentrator (CPC) collectors connected in series, in IEEE International Conference on Power Electronics, Intelligent Control and Energy Systems (2016), pp. 12–17 13. V. Saini, R. Tripathi, G.N. Tiwari, I.M. Al-Helal, Electrical and thermal assessment of N partially covered PVT-compound parabolic concentrator collector connected in series, for different solar cell materials. Appl. Therm. Eng. 128, 1611–1623 (2018) 14. R. Tripathi, G.N. Tiwari, Annual performance evaluation (energy and exergy) of fully covered concentrated photovoltaic thermal (PVT) water collector: An experimental validation. Sol. Energy 146, 180–190 (2017) 15. R. Tripathi, G.N. Tiwari, V.K. Dwivedi, Energy matrices evaluation and exergoeconomic analysis of series connected N partially covered (glass to glass PV module) concentrated photovoltaic thermal collector: At constant flow rate mode. Energy Convers. Manag. 145, 357–370 (2017) 16. R. Tripathi, G.N. Tiwari, Annual energy, exergy, and environmental benefits of n half covered concentrated photovoltaic thermal (CPVT) air collectors, in Advances in Smart Grid and Renewable Energy. Lecture Notes in Electrical Engineering, vol. 435 (2017), pp. 113–124 17. R. Tripathi, G.N. Tiwari, Exergy and carbon credits for series connected N photovoltaic thermal—compound parabolic concentrator (PVT-CPC) collector: at constant outlet temperature. Int. J. Eng. Sci. Res. Technol. 6(9), 678–696 (2017) 18. R. Tripathi, G.N. Tiwari, T.S. Bhatti, V.K. Dwivedi, 2-E (Energy-Exergy) for partially covered concentrated photovoltaic thermal (PVT) collector. Energy Procedia 142, 616–623 (2017) 19. R. Tripathi, G. N. Tiwari, Energy matrices, life cycle cost, carbon mitigation and credits of open-loop N concentrated photovoltaic thermal (CPVT) collector at cold climate in India: A comparative study. Solar Energy 186, 347–359 (2019)

Chapter 12

Thermo-Hydraulic Performance of Solar Air Heater Roughened with V-Shaped Ribs Combined with V-Shaped Perforated Baffles Vijay Singh Bisht, Anil Kumar Patil and Anirudh Gupta Abstract A 3-D CFD numerical study of solar air heater roughened with V-shaped ribs and V-shaped perforated baffles has been presented in this work. A numerical study has been conducted in ANSYS FLUENT by employing RNG k-epsilon turbulence model with enhanced wall treatment condition. CFD results are validated with the previous experimental work. The Reynolds number varied from 4000 to 18,000, and open area ratio (β) is varied from 12 to 36%. In this study, energy balance has been used to obtain the value of the heat transfer coefficient and thermal efficiency which results that the thermo-hydraulic performance of roughened solar air heater is much superior than the smoother one. The highest value of the Nusselt number and thermal efficiency corresponds to open area ratio (β) of 12%. The highest value of thermo-hydraulic performance parameter is obtained for open area ratio (β) of 24% at all Reynolds number. Airflow behavior has been analyzed by velocity contour and turbulence kinetic energy contour. Keywords Solar energy · CFD · Solar air heater · Energy balance · Thermal efficiency · Thermo-hydraulic performance

12.1 Introduction Solar energy is an eco-friendly, permanent, and freely available form of energy. The solar air heater is a simple device that directly utilizes solar energy for different purposes like space heating, agricultural, and industrial applications. The thermohydraulic efficiency of the solar air heater is low, and it can be raised by enhancing by V. S. Bisht (B) Department of Thermal Engineering, Faculty of Technology, UTU, Dehradun, India e-mail: [email protected] A. K. Patil Department of Mechanical Engineering, DIT University, Dehradun, India A. Gupta Department of Mechanical Engineering, BTKIT, Dwarahat, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_12

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the application of turbulence promoters on the undersurface of the absorber plate of the solar air heater duct. Researchers carried out both experimental and computational investigations on different turbulent promoters (artificial roughness) to achieve higher thermal efficiency. CFD analysis is particularly beneficial where safety is considered and the domain of investigation is large in size [1]. The main merits of CFD over the experimental approach are cost reduction and capacity to yield large data in a smaller time period [2]. Moreover, the CFD captures the fluid flow visualization which is a tedious job in the experimental work. In this part, a brief review of the past work is discussed. Chaube et al. [3] performed a computational study of solar air heater roughened with nine different turbulent promoters in ANSYS FLUENT. They have reported that heat transfer with chamfered ribs is more than the other ribs. Kumar and Saini [4] used ANSYS FLUENT for the 3-D computational study of solar air heater roughened with arc-shaped ribs. They concluded that at relative arc angle 30° and e/D (relative roughness height) value of 0.0426, the highest value of the overall enhancement ratio was achieved. Yadav and Bhagoria [5] used CFD FLUENT for 2-D computational investigation of solar air heater roughened with triangular ribs. They selected the RNG k-epsilon model in this study. They reported that the highest value of Nusselt number was found at the pitch ratio of 10. Singh et al. [6] have examined a CFD performance of solar air heater roughened with periodic transverse ribs. They found maximum enhancement in heat transfer with saw tooth rib roughness and trapezoidal rib roughness. Kumar and Kim [7] performed the computational study of solar air heater roughened with discrete multi-V-pattern rib geometries. They reported that RNG k-epsilon turbulence model results were in good acceptance with the previous empirical relations. Kumar and Kumar [8] used elliptical ribs turbulent promoters for the computational study of solar air heater. In their study, RNG k-epsilon model of FLUENT is used to explore the effect of various system parameters on thermo-hydraulic efficiency. They reported that the maximum enhancement in heat transfer was found at the pitch ratio of 6 and e/D value of 0.045. V shaping of ribs and obstacle was found to be effective turbulence promoters [9, 10]. The main goals of the present study are to explore the thermal behavior of air flowing through solar air heater duct roughened by the combination of V-shaped ribs and V-shaped perforated baffles.

12.2 CFD Study In this study, a solar air heater duct is modeled according to ASHRAE standards 93-77 [11]. The duct of dimensions 2400 mm × 300 mm × 25 mm in size is selected for the numerical study. The entrance section, test section, and exit section of duct have the length of 525 mm, 1000 mm, and 875 mm, respectively. The absorber plate exposed to a uniform flux of 1000 W/m2 , the V-shaped ribs and V-shaped perforated baffles are attached to the absorber plate; these roughness elements are assumed to be adiabatic. The duct wall face opposite to absorber plate and other two wall

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Table 12.1 Geometrical and operating parameters Geometrical and operating parameters

Symbolic representations

Value/range

Hydraulic diameter of the duct

D

46.15 mm

Relative roughness height (for ribs)

e/D

0.043 mm

Height of duct

h

25 mm

Relative roughness height (for baffles)

h/D

0.285 mm

Angle of attack (for both ribs and baffles)

α

60°

Open area ratio

β

12–36% in three steps

Number of holes in a V-shaped baffle

n

15

Reynolds number

Re

4000–18,000

Prandtl number

Pr

0.71

Insolation

I

1000 W/m2

faces are assumed to be insulated and smooth. At the wall-fluid attachment zone, no-slip conditions are established. In this investigation, velocity inlet and pressure outlet boundary conditions at duct inlet and exit are considered, respectively. The geometrical and the operating parameters considered in this investigation are shown in Table 12.1.

12.2.1 CFD Methodology CFD methodology comprises three steps, namely the preprocessor, the solver, and the post-processor. The turbulent promoters considered in the study are a combination of V-shaped ribs and V-down perforated baffles as shown in Fig. 12.1.

Fig. 12.1 V-shaped ribs and V-down perforated baffles on absorber plate

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Fig. 12.2 Mesh of the computational domain of absorber plate with V-shaped ribs and V-shaped perforated baffles

After nomenclature of different parts of ducts in design modeler, the geometry has been imported in ANSYS ICEM, where meshing of geometry was completed. The meshed computational domain is shown in Fig. 12.2.

12.2.2 Grid Independence Test and the Validation of Model It is clear from the introduction section that for the computational study of solar air heater in ANSYS FLUENT solver, RNG k-epsilon is the most suitable turbulence model. In this study, for all geometries, the RNG k-epsilon turbulence model with enhanced wall treatment condition is used. For verification of computational domain, a grid independence test is carried out with the open area ratio (β) of 12% and Reynolds number (Re) of 12,000. It is found that for the increment of elements from 9.62 million to 13.12 million, the percentage change in the Nusselt number is only 0.36%, and therefore, 13.121 million grid cells have been selected for further investigation. The results of the Nusselt number of the smooth duct from RNG kepsilon are compared with the Dittus–Boelter correlation as given in Eq. (12.1) and as shown in Fig. 12.3. Nus = 0.023 Re0.8 Pr0.4

(12.1)

The average absolute deviation of Nusselt number obtained using RNG k-epsilon model is 7.7% from Dittus–Boelter correlation. The computational model is also validated by comparing the Nusselt number results from the RNG k-epsilon model with the Nusselt number results obtained in the experimental work by Chamoli and Thakur [12] as shown in Fig. 12.3. The average absolute deviation between the computational and the experimental data for the Nusselt number is found to be 7.86%.

12 Thermo-Hydraulic Performance of Solar Air Heater Roughened … 120

RNG k-epsilon turbulence model

100

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40 20 0 4000

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Present Study 8000

12000

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Re Fig. 12.3 Validation of numerical results with Dittus–Boelter equation and the previous experimental work

12.3 Data Processing In this investigation, the thermo-hydraulic performance of solar air heater roughened with V-shaped ribs combined with V-shaped perforated baffles has been studied. The open area ratio [12] for baffles is given as: β=

n(π/4)d 2 be B

(12.2)

where n is a number of drilled holes in baffles, d is diameter holes in mm, b is the length of the baffle in mm and e B height of baffle in mm. The Nusselt number given as: Nu =

hD k

(12.3)

From the energy balance of steady flow through solar air heater duct, the energy carried away by fluid (Qc ) is expressed as: Q c = hAp (Ts − Tb )

(12.4)

where Ap is the surface area of the absorber plate, T s is the average surface temperature of the absorber plate and T b is the bulk mean temperature of air is defined as:  Tb =

Ti + To 2

 (12.5)

where T i and T o are the inlet temperature and average outlet temperature of air flowing through the solar air heater duct, respectively. Qu is the useful heat gain and is given by: ˙ p (TO − Ti ) Q c = Q u = mC

(12.6)

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Thermal efficiency [13] is defined as: 

Qu I Ap

ηt =

 (12.7)

Thermo-hydraulic performance parameter [14] expressed in terms of Nusselt number and friction factor of the rough and smooth duct as:  η=

Nur Nus

 (12.8)

 1/ 3 fr fs

12.4 Results and Discussions The thermo-hydraulic performance of the solar air heater duct with V-shaped ribs and V-shaped perforated baffles is discussed in this section. The velocity contours and the turbulence kinetic energy contours are also exhibited to discuss the fluid flow behavior of air flowing through the solar air heater duct.

12.4.1 Heat Transfer The influence of β on Nusselt number for different values of Reynolds number is shown in Fig. 12.4. For all roughness geometries including smooth duct, the Nusselt number increases with the increase in Reynolds number. With the introduction of V-shaped ribs combined with V-shaped perforated baffles, the higher heat transfer rates are achieved in comparison with the smooth duct. At the open area ratio (β) equals to 12%, the 95

β = 12%

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β = 24%

55

β = 36%

35 15 4000

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Re Fig. 12.4 Effect of β on Nusselt number for different values of Reynolds number

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highest value of the Nusselt number is recorded at all the Reynolds number. It can be seen from Fig. 12.4 that an increase in the value of the open area ratio (β) reduces the heat transfer rates. It may be due to the reduction in the air velocity with the increase in the size of perforations that may result in the poor mixing of fluid.

12.4.2 Thermal Efficiency and Thermo-Hydraulic Performance Parameter Figure 12.5 shows the effect of open area ratio (β) on thermal efficiency. Figure 12.5 exhibits an increase in thermal efficiency of all roughness geometries with the increase in the Reynolds number due to an increase in turbulence intensity. It can be observed that the increase in the value of open area ratio (β) leads to the decrement in the thermal efficiency; however, there is a substantial gain in the efficiency by the application of the combination of these V-shaped turbulent promoters. The synergy of the secondary flow formation due to V shaping of rib roughness and the jet released from the perforations in V-shaped baffles produces higher turbulence near the wall and thereby enhances the heat transfer rates so the thermal efficiency. Figure 12.6 shows the effect of open area ratio (β) on thermohydraulic performance parameter for different values of Reynolds number. It can be seen from Fig. 12.6 that the open area ratio (β) of 24% produces the maximum value of thermo-hydraulic performance parameter for all values of Reynolds number; this is because as there is an increase in the value of open area ratio (β), less pumping power is required and so frictional losses decreases and thermo-hydraulic performance relatively increases.

0.85 β = 24%

ηth

0.75

Smooth Duct

0.65

β = 12%

0.55 0.45 4000

β = 36% 8000

12000

16000

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Fig. 12.5 Effect of β on thermal efficiency for different values of the Reynolds number

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η

β = 24%

0.9

β = 12%

0.7

β = 36%

0.5 4000

8000

12000

16000

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Re Fig. 12.6 Effect of β on thermo-hydraulic performance for different values of Reynolds number

12.4.3 Fluid Flow Behavior The fluid flow behavior can be understood by analyzing the velocity and turbulent kinetic energy contours. Figure 12.7 shows the velocity contour where the accelerated flow can be noticed in the zone between the V-shaped perforated baffles and V-shaped rib. The intense fluid mixing in the said region may be attributed to higher heat transfer rates. The turbulence kinetic energy contour shown in Fig. 12.8 reflects that the strength of turbulence plays an important role in heat transfer near the surface. A higher turbulence kinetic energy can be seen in the regions between the V-shaped perforated baffles and V-shaped ribs due to brisk mixing of jet flow, secondary flow, and the main flow; hence, the greater localized heat transfer rates are expected in these regions.

Fig. 12.7 Velocity contour of airflow through the solar air heater duct

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Fig. 12.8 Turbulence kinetic energy contour of airflow through solar air heater duct

12.5 Conclusions A 3-D CFD analysis is performed to investigate thermo-hydraulic characteristics of solar air heater roughened with V-shaped rib and V-shaped perforated baffles. RNG k-epsilon turbulence model with enhanced wall treatment condition is used in this study. The following conclusions can be drawn from this study: • CFD numerical study is an effective way to examine the complete fluid flow behavior of the computational domain under consideration. Complex geometries which are difficult to design and manufacture can be easily investigated by using the CFD approach. • The Nusselt number increases with the increase in Reynolds number for all geometries. Solar air heater duct roughened with V-shaped rib combined with V-shaped perforated baffles has attained significantly higher values of Nusselt number than that of a smooth duct. Roughness geometry having β equal to 12% yields highest value of Nusselt number irrespective of the change in Reynolds number. The Nusselt number decreases with an increase in the value of β. • Thermal efficiency is increased with an increase in Reynolds number in all cases. The value of thermal efficiency decreases with the increase in the value of the open area ratio (β). The highest and lowest value of thermal efficiency corresponds to open area ratio (β) equal to 12%, and 36%, respectively. Thermo-hydraulic performance parameter achieved its maximum value for the open area ratio (β) of 24%. • The combination of V-shaped ribs and perforated V-shaped baffles can be used as roughness geometries in a solar air heater to attain a substantial gain in its performance.

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References 1. H.K. Versteeg, W. Malalasekera, An Introduction to Computational Fluid Dynamics: The Finite Volume Method, 2nd edn (Pearson Education, 2007) 2. D.A. John, Computational Fluid Dynamics (McGraw Hill International Editions 2005) 3. A. Chaube, P.K. Sahoo, S.C. Solanki, Analysis of heat transfer augmentation and flow characteristics due to rib roughness over absorber plate of a solar air heater. Renew. Energy 31(3), 317–331 (2006) 4. S. Kumar, R.P. Saini, CFD based performance analysis of a solar air heater duct provided with artificial roughness. Renew. Energy 34(5), 1285–1291 (2009) 5. A.S. Yadav, J.L. Bhagoria, A CFD analysis of a solar air heater having triangular rib roughness on the absorber plate. Int. J. ChemTech Res. 5(2), 964–971 (2013) 6. S. Singh, B. Singh, V.S. Hans, R.S. Gill, CFD (computational fluid dynamics) investigation on Nusselt number and friction factor of solar air heater duct roughened with non-uniform cross-section transverse rib. Energy 84, 509–517 (2015) 7. A. Kumar, M.H. Kim, CFD analysis on the thermal hydraulic performance of an SAH duct with multi V-shape roughened ribs. Energies 9(6), 415 (2016) 8. R. Kumar, A. Kumar, A parametric study of the 2D model of solar air heater with elliptical rib roughness using CFD. J. Mech. Sci. Technol. 31(2), 959–964 (2017) 9. A.K. Patil, J.S. Saini, K. Kumar, Heat transfer and friction characteristics of solar air heater duct roughened by broken V-shape ribs combined with staggered rib piece. J. Renew. Sustain. Energy 4(1), 013115 (2012) 10. V.S. Bisht, A.K. Patil, A. Gupta, Review and performance evaluation of roughened solar air heaters. Renew. Sustain. Energy Rev. 81(1), 954–977 (2018) 11. ASHRAE Standard, Method of testing of determines thermal performance of solar collector (1977), pp. 93–77 12. S. Chamoli, N.S. Thakur, Correlations for solar air heater duct with V-shaped perforated baffles as roughness elements on absorber plate. Int. J. Sustain. Energy 35(1), 1–20 (2016) 13. J.A. Duffie, W.A. Beckman, Solar Egineering of Thermal Processes (Wiley, 2013) 14. M.J. Lewis, Optimizing the thermohydraulic performance of rough surfaces. Int J Heat Mass Trans 18, 1243–1248 (1975)

Chapter 13

Highly Efficient Solar Steam Generation Using Carbon Cloth System M. W. Higgins, A. R. Shakeelur Rahman and Neetu Jha

Abstract Production of steam using renewable resources has drawn in a great deal of interest in recent times. Photothermal systems of various designs and applications have thus been developed with considerable success. Nevertheless, it is a necessity to develop a system which is low cost and scalable to fit real-world applications. We have developed a system of carbon cloth and wood which provides an excellent solar steam generation with thermal efficiencies up to 60% under 1 sun illumination. The carbon cloth system improves the evaporation rate of water by 180%. Solar steam can be got almost instantly upon illumination. The high thermal stability of carbon cloth makes it beneficial in large-scale applications with very high incident intensities. It is envisioned that this system can be readily scaled in practical solar steam generation applications due to its simplicity, low cost, and high thermal stability. Keywords Solar steam · Carbon cloth

13.1 Introduction In the past decade, scientists around the globe are attempting to produce steam using renewable energy resources. One such method employed is the use of solar energy for steam generation. In this technique, incident solar energy is applied to evaporate water molecules and create steam. This is a very inefficient process. Zhang et al. [1] Water is a poor absorber of light and allows the incident solar irradiation to pass through the water surface. The transmitted light gets scattered by water molecules and results in a substantial wastage of incident energy. In this method, the water vaporization is a result of an increase in bulk temperature of water or commonly defined as bulk evaporation. To tackle this issue of poor solar thermal efficiency of water, researchers have employed techniques like the addition of light-absorbing nanoparticles in water, which help in maximizing the absorption of incoming solar M. W. Higgins · A. R. Shakeelur Rahman · N. Jha (B) Department of Physics, Institute of Chemical Technology Mumbai, N.P. Marg, Mumbai 400019, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_13

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energy. Neumann et al. [2] The light-absorbing nanoparticles help in converting the absorbed solar irradiation into heat and increase the temperature of surrounding water molecules to produce steam. Another significant approach used to increase the solar thermal efficiency is the use of floating substrate containing broadband lightabsorbing nanoparticles. Wang et al. [3], Liu et al. [4], Zhou et al. [5], Li et al. [6] In this technique, the substrate floats above water. The heat produced by the radiationless transition is localized at air–water interface. The bulk water below the substrate that is in direct contact with the hot zone gets evaporated instantly without much bulk heating of water. The substrate technique thus promotes surface evaporation of water and significantly improves the evaporation rates of water. In this study, we present carbon cloth (CC) as a substrate for a highly efficient solar-driven steam generation. To obtain high efficiency in solar steam generation, the substrate should possess broadband solar absorption, low thermal conductivity, and porosity. The CC offers excellent broadband absorption in the 400–2000 nm wavelength range. Secondly, the low thermal conductivity of CC helps in the creation of hot zone and low dissipation of heat away from it. Finally, the porosity of CC helps in the capillary action of underlying bulk water towards the hot zone. CC showcased a thermal efficiency of 60.2% at 1 kW/m2 and 93% at 18 kW/m2 . These results showcase CC as a promising material for a high-performance solar steam generation.

13.2 Materials Carbon cloth was procured from AvCarb, USA. (Price $154/m2 ). A wooden ring of thickness 1 cm was made from wooden blocks of balsa wood.

13.3 Characterization X-ray diffraction pattern of CC was measured using XRD Bruker D8 Advance with 2θ range of 5°–80°. Field emission scanning electron microscopy (FESEM) was used to examine the surface morphology of CC. The absorbance and reflectance spectra were measured using agilent spectrophotometer.

13.4 Evaporation Performance Evaluation The evaporation performance was calculated by recording the mass loss of water against time. Water was filled in vessel of 4 cm in diameter and illuminated using a solar simulator (PET 50AAA). The evaporation set up was kept on weighing balance and weight loss of water was studied after steady state. A thermal camera was used

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to record the IR images of the system in time. The temperature of bulk water was observed using a standard laboratory thermometer

13.5 Results and Discussion The density of carbon cloth was measured to be 1.5 g cm−3 and thus a wooden ring was used to support CC and keep it afloat in water. SEM images were used to visualize the structure of carbon cloth. SEM images showcase a porous woven fabric structure with a pore size of 250 μm (Fig. 13.1a). The thickness of each fibre was approximately ~7 μm (Fig. 13.1b). The XPS survey scan of CC display carbon and oxygen peaks. (Fig. 13.1c). The atomic percentages of carbon and oxygen were measured to be 57.8% and 34.2%, respectively. XRD pattern of CC display peaks at 25.4° and 44.5° correspond to (002) and (101) planes of carbon with turbostratic structure, respectively (Fig. 13.1d). Figure 13.2a shows the reflectivity of carbon cloth. Carbon cloth has a very low reflectivity of less than 5% in the wavelength range 200–2000 nm.

Fig. 13.1 a, b SEM images showcases the porosity and structure of CC. c survey scan of CC. d X-ray diffraction pattern of CC

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Fig. 13.2 a Reflectivity of carbon cloth, b absorption spectra of carbon cloth, c experimental setup for solar steam generation evaluation, and d optical photograph of steam generation

The high absorbance and very low reflectance of CC as shown in Fig. 13.2b make CC an excellent material to absorb solar radiations. Meanwhile, the thermal conductivity was also measured to be 0.21 Wm−1 K−1 . This helps in isolating the heat at the air– water interface. In conclusion, the porosity, low thermal conductivity, and excellent absorptance make CC idyllic for solar-driven steam generation. The steam generation was estimated using the setup as shown in Fig. 13.2c. Upon solar illumination, a hot zone is produced as shown in Fig. 13.2d. The water molecules surrounding the hot zone escape instantly to produce steam. The solar-driven steam generation was examined by means of a solar simulator. The power density of solar illumination was kept constant at 1 sun throughout the study. In 1 sun intensity, the water evaporation rate was 0.49 kg m−2 h−1 . (As shown in Fig. 13.3) The evaporation rate improved to 0.89 kg m−2 h−1 in identical conditions when CC was added. Thus, the evaporation rate of water with CC is almost ~2 times higher without it. The tests were repeated 15 times and the results showed meagre variations. The photothermal efficiency of CC can be defined as: ´ = m˙ ∗ h LV /I

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Fig. 13.3 Mass change versus time graph of carbon cloth under 1 sun illumination

where m, ˙ hLV , and I are rate of water evaporation, total heat enthalpy and light intensity. The photothermal efficiency of CC is 60.2% at 1 sun. In higher concentrations such as 18 suns, the evaporation rate of water with CC was 23 kg m−2 h−1 . The evaporation rate of water with CC increased ten-folds in 18 suns which accounted to a photothermal efficiency of 93%. To investigate further the excellent performance of CC, we measured its temperatures using IR photography. Figure 13.4 depicts the surface temperatures during the test. Initially, the sample temperatures were ~26 °C for bare and ~27.7 °C with CC. After 1 sun illumination for 30 min, the temperature distributions of the samples displayed distinct variations. The surface temperature of CC surface increased to ~33 °C during the test compared to ~27 °C of pure water test. It can be noted from the diagram that the heat gets localized (highest temperature) at the top of the CC surface and the temperature decreases as we move away from the top surface. This was confirmed by measuring the water temperatures of both samples at the top and bottom of the vessel during the experiment. The water temperatures at the top and bottom were similar without CC under 1 sun illumination. With the presence of CC on the top, the temperature of water at the top and bottom varied with distance. The temperature at the top was 33 °C, whereas the temperature of water at the bottom remained at 27 °C. This confirms that CC helps in minimizing the bulk heating of water and promotes surface heating of water. Hence, with IR photography and water temperatures at different distances, it can be confirmed that CC helps in localizing heat at the top of the water surface which contributes to high thermal efficiency. In summary, we have demonstrated the application of carbon cloth as an excellent material for solar-driven steam generation processes. Carbon cloth with its broadband absorption of incident light helps in absorbing the majority of incoming solar radiation. The low thermal conductivity of carbon cloth helps in localizing the heat at

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Fig. 13.4 a, b Thermal images at 0 and 30 min, respectively, in the absence of carbon cloth under 1 sun illumination. c, d Thermal images at 0 and 30 min, respectively, in the presence of carbon cloth under 1 sun illumination

the surface and prevents loss of heat by dissipation. The porosity of the carbon cloth helps in continuous supply of water molecules to the hot zone and continuous steam generation at very high thermal efficiency. The carbon cloth also remained thermal stable throughout the experiment even at high intensities which make it a candidate for industrial applications. Thus, carbon cloth inspires further development of highly thermally stable photothermal materials for highly efficient solar-driven water evaporators with applications in seawater desalination, sterilization among others.

References 1. L. Zhang, B. Tang, J. Wu, R. Li, P. Wang, Hydrophobic light-to-heat conversion membranes with self-healing ability for interfacial solar heating. Adv. Mater. 27(33), 4889–4894 (2015) 2. O. Neumann, A.S. Urban, J. Day, S. Lal, P. Nordlander, N.J. Halas, Solar vapor generation enabled by nanoparticles. ACS Nano 7(1), 42–49 (2013) 3. Y. Wang, L. Zhang, P. Wang, Self-floating carbon nanotube membrane on macroporous silica substrate for highly efficient solar-driven interfacial water evaporation. ACS Sustain. Chem. Eng. 4(3), 1223–1230 (2016)

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4. Y. Liu, S. Yu, R. Feng, A. Bernard, Y. Liu, Y. Zhang, H. Duan, W. Shang, P. Tao, C. Song, T. Deng, A bioinspired, reusable, paper-based system for high-performance large-scale evaporation. Adv. Mater. 27(17), 2768–2774 (2015) 5. L. Zhou, Y. Tan, J. Wang, W. Xu, Y. Yuan, W. Cai, S. Zhu, J. Zhu, 3D self-assembly of aluminium nanoparticles for plasmon-enhanced solar desalination. Nat. Photonics 10(6), 393–398 (2016) 6. R. Li, L. Zhang, L. Shi, P. Wang, MXene Ti3 C2 : An effective 2D light-to-heat conversion material. ACS Nano 11(4), 3752–3759 (2017)

Chapter 14

Floating Absorber Integrated with Compound Parabolic Concentrator for Effective Solar Water Desalination Chandan and Bala Pesala

Abstract Accessibility to clean water is one of the most important challenges currently being faced by humanity. More than 780 million people in the world do not have access to clean water. Though earth has plenty of water, 97% of it is saltwater in oceans which needs appropriate treatment technologies to convert this to potable water. The current technologies of water desalination such as reverse osmosis consume a significant amount of energy, leading to the water-energy conundrum. To overcome this limitation, recently, several technologies based on nanoparticleenhanced steam generation have been explored demonstrating extremely high conversion efficiencies (>40%) in the laboratory scale. These methods effectively utilize the plasmonic resonances of nanoparticles to increase the absorption cross section for the sunlight. However, they typically need to be operated under high concentration (>10X) requiring continuous tracking which results in increased system cost and complexity. Here, a novel solar-powered desalination system is proposed using CPCbased concentrator (with concentration ratio 2) combined with a low-cost absorber. The system demonstrates huge potential with an efficiency of 39% achieved at 2X concentration. Keywords Compound parabolic concentrator · Desalination system · Photothermal efficiency · Solar absorber

Chandan · B. Pesala (B) Academy of Scientific and Innovative Research, Chennai 600113, India e-mail: [email protected] Chandan CSIR—Structural Engineering Research Center, Chennai 600113, India B. Pesala CSIR—Central Electronics Engineering Research Institute, Chennai 600113, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_14

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14.1 Introduction Solar irradiation is a sustainable power source falling on the surface of the earth. Because of its sustainable nature, effective harnessing of sunlight has gained momentum. Recently, steam generation using solar power has attracted much attention because of its vast applications in the field of large-scale power generation [1], vapor absorption chillers [2] and desalination systems [3]. Conventional solar-powered desalination methods (single still, double still, cone type still, inclined still with multiple wick system, multi-effect solar still, etc.) rely on bulk heating of water by suitable absorber surface, leading to water evaporation and collection of the condensate. The major drawback of these systems is the relatively low thermal efficiencies (4X) suitable for steam generation

(a)

(b) 6 Local Concentration Ratio Average Concentration Ratio

Concentration Ratio

Concentration Ratio

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Local Concentration Ratio Average Concentration Ratio

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Fig. 14.3 Variation of local concentration on the surface of the absorber at an angle of incidence of a 0° b 10° c 20° d 30°

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Fig. 14.4 Absorption spectrum of carbon paper in UV–visible and NIR spectrum

14.3 Experimental Results 14.3.1 Characterization The absorber material was characterized using UV–visible-NIR spectrophotometer with integrating sphere attachment. For measuring UV–visible reflection and transmission spectrum, Ocean Optics 2000 spectrometer was used while NIR spectrum was measured by Flame NIR spectrometer. To measure the absorptivity, the values of transmission and reflection were subtracted from unity. The equation is given as A = 1 − T − R where A is absorptivity, T is transmissivity and R is reflectivity of the sample (Fig. 14.4).

14.3.2 Evaporation Experiments of CPC Under Solar Simulator Figure 14.1 shows the setup for the steam generation rate using CPC. The change in mass of water from the tank was measured using a precision balance which can sustain maximum load of 220 g with least count of 0.001 g. The carbon paper-based absorber having dimension 2.5 cm × 10 cm was placed at the base of the CPC, and water was pulled from the tank to the carbon paper using cotton wick placed just below the carbon paper. The average power available under the solar simulator was 0.94 kW/m2 . To perform experiments on variation of the efficiency at different angles of incidence, the CPC was tilted by the help of mountings placed at the longer edge. The height of the mountings is varied from 0 to 2.5 cm. Temperatures were measured using thermocouple at different locations of the absorber. To calculate the thermal efficiency, the equation proposed by Ghasemi et al. [12] was used as shown in Eq. (14.4)

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(b)

0

50 0° 11° 30°

Efficiency (%)

Mass Evaporated (gm)

(a)

-2

-4

40 30 20 10 0

-6 0

10

20

30

40

50

60

0

5

10

15

20

25

30

Angle of Incidence (Degree)

Time (min)

Fig. 14.5 a Steam generation rate curve for CPC at different angles of incidence b variation of thermal efficiency with different angles of incidence

η=

mh ˙ lv C I Aaperture

(14.4)

where m˙ is the mass evaporated from the CPC—tank combined, h lv is the latent heat of vaporization of water from the surface of the absorber, C is the geometric concentration ratio, I is the irradiance falling on the CPC and Aaperture is the aperture area of the CPC. All the steam generation rate experiments were performed at 940 W/m2 . To quantify the performance of the CPC, 50 g of water was taken in the tank on which the CPC was mounted. Then, CPC was kept under the solar simulator to measure the amount of vapor evaporated which was measured by precision balance. The results are shown in Fig. 14.5a. Since the water is evaporated from the surface of the absorber to the ambient, it has been plotted on the negative y-axis representing the mass of water gone out of the system. Once the readings were taken at 0° angle of incidence, the CPC was tilted at different angles to evaluate the performance (Fig. 14.5b). The maximum vapor generation rate of 1.01 kg/m2 h was achieved at 0° and 11° and angle of incidence at an efficiency of 39% while minimum vapor generation rate of 0.68 kg/m2 h was achieved at 30° angle of incidence. This dip in efficiency can be because of nonuniform spread of radiation over the entire surface of the absorber. Further, the low efficiency of the system could be because of the parasitic losses associated with the CPC-based system. The advantage of CPC-based system can be seen in terms of vapor generation rate where it is almost constant at different angles of incidence, thus could be a potential replacement of the tracking-based floating absorber desalination systems.

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14.4 Conclusion Here, a CPC-based novel desalination system has been demonstrated, integrated with a carbon paper-based absorber material. The absorber chosen was having very high absorptivity of 87% in the wavelength range of 400–1650 nm. When the system was tested under the solar simulator, vapor generation rate was almost constant at different angles of incidence. The main aim of integrating CPC to the carbon-based absorber material was to localize more heat to the surface of the absorber which in turn can improve the throughput of the system; at the same time, tracking can be avoided that makes the system very simple and cost effective. Future work involves scaling the system to larger dimensions and testing under real solar insolation conditions. This new CPC-based vapor generator could potentially replace the existing Fresnel concentrator-based system for desalination system. Acknowledgements The authors acknowledge the funding support of DST through sanction letter no: DST/TM/SERI/278(G)

References 1. M.J. Montes, A. Abánades, J.M. Martínez-Val, Performance of a direct steam generation solar thermal power plant for electricity production as a function of the solar multiple. Sol. Energy 83(5), 679–689 (2009) 2. P. Srikhirin, S. Aphornratana, S. Chungpaibulpatana, A review of absorption refrigeration technologies. Renew. Sustain. Energy Rev. 5(4), 343–372 (2001) 3. D.D.W. Rufuss et al., Solar stills: A comprehensive review of designs, performance and material advances. Renew. Sustain. Energy Rev. 63, 464–496 (2016) 4. O. Neumann et al., Compact solar autoclave based on steam generation using broadband lightharvesting nanoparticles. Proc. Nat. Acad. Sci. 110(29), 11677–11681 (2013) 5. L. Zhou et al., 3D self-assembly of aluminium nanoparticles for plasmon-enhanced solar desalination. Nat. Photonics 10, 393 (2016) 6. L. Zhou et al., Self-assembly of highly efficient, broadband plasmonic absorbers for solar steam generation. Sci. Adv. 2(4), e1501227 (2016) 7. Y. Liu et al., Bioinspired bifunctional membrane for efficient clean water generation. ACS Appl. Mater. Interfaces 8(1), 772–779 (2015) 8. X. Li et al., Graphene oxide-based efficient and scalable solar desalination under one sun with a confined 2D water path. Proc. Nat. Acad. Sci. 113(49), 13953–13958 (2016) 9. Q. Jiang et al., Bilayered biofoam for highly efficient solar steam generation. Adv. Mater. 28(42), 9400–9407 (2016) 10. G. Xue et al., Robust and low-cost flame-treated wood for high-performance solar steam generation. ACS Appl. Mater. Interfaces 9(17), 15052–15057 (2017) 11. K.K. Liu et al., Wood–graphene oxide composite for highly efficient solar steam generation and desalination. ACS Appl. Mater. Interface 9(8), 7675–7681 (2017) 12. G. Ni et al., Steam generation under one sun enabled by a floating structure with thermal concentration. Nat. Energy 1, 16126 (2016) 13. J. Chaves, Introduction to Non-Imaging Optics, CRC Press (2015)

Chapter 15

Study of Performance of Solar Flat Plate Collector Using Al2 O3 /Water Nanofluids Pankaj Raj, Geleta Fekadu and Sudhakar Subudhi

Abstract The present paper investigates the performance of a solar flat plate collector using Al2 O3 /water nanofluids. The enhancement in performance of collector when using Al2 O3 (20 nm) nanofluid in comparison with base fluid has been studied here. The mass flow rate is varied from 0.5 l.p.m to 2.5 l.p.m (0.5, 1.0, 1.5, 2.0 and 2.5), and volume fraction of nanofluid is varying from 0.01 to 0.15 vol.% (0.01, 0.02, 0.04, 0.075 and 0.15). Experiments are carried out with stable nanofluid. The optimum condition for maximum efficiency is found to be 2.0 l.p.m and 0.15 vol.%. The efficiency is found to increase from 46.3% for water to 73.1% for 0.15 vol.% Al2 O3 nanofluid at a 2 l.p.m flow rate. A maximum rise in temperature of 22 °C was obtained for 0.5 l.p.m and 0.15 vol.%. Keywords Solar flat plate collector (SFPC) · Nanofluids · Volume fraction · Efficiency

15.1 Introduction With increasing pollution and depleting fossil fuels, the demand for an alternate form of energy is on high. Solar technology has a promising and emerged supplement in the form of thermal and electrical energy [1]. But, lower collector’s efficiency and higher investment cost is the cause of concern [2]. Various researchers performed experiments to improve collector’s efficiency by varying collector’s design, size, orientation and absorber fluid. Nanofluid in use instead of water due to absorbing fluid showed an appreciable rise in collector’s efficiency. Meibodi et al. [3] studied the efficiency and performances of solar flat plate collector (SFPC) using SiO2 /ethylene glycol (EG)-H2 O nanofluid. The collector efficiency is increased from 4 to 8% when the nanofluid volume concentration increases from 0.5 to 1% provided that heat loss close to zero. With increasing, heat loss parameter nanofluid with lower concentration was preferred. Said et al. [4] analyzed the exergy, heat transfer, power for pumping P. Raj · G. Fekadu · S. Subudhi (B) Department of Mechanical and Industrial Engineering, IIT Roorkee, Roorkee, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_15

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and drop in pressure for a SFPC for single-wall carbon nanotubes (SWCNTs) in a nanofluid. With increasing, volume flow rate enhanced thermal conductivity, and higher value of Nusselt number was observed. While using, SWCNT reduced 4.34% of entropy generation and enhanced heat transfer coefficient to 15.33%. Faizal et al. [5] performed economic, energy and environment studies of the different oxides of metal nanofluid in SFPC. Varying volume fraction effects, the specific heat and density fluid were observed. The efficiency of solar collector where increased with the increased volume flow rate. The reduction size for nanofluid was also considered. Moghadam et al. [6] carried out the efficiency and performance analysis of SFPC using CuO-H2 O nanofluid. The efficiency increases when the mass flow rate considered is from 1 to 2 kg/min and it decreases when the mass flow is increased to 3 kg/min and the efficiency of the collector is increased by 16.7% when the CuO-Water nanofluid is used. Yousefi et al. [7] studied the efficiency outcome of MWCNT-H2 O nanofluid on SFPC. The addition of surfactant on the collector efficiency and stability was seen. The efficiency of FPC was evaluated for 0.0167, 0.033 and 0.05 kg/s mass flow rate and it increases with increase in mass flow rate and it is found highest at 57%. Nasrin and Alim [8] carried out numerical analysis for forced convection through a FPC and set a semiempirical formula. Al2 O3 –H2 O was used as nanofluid. The efficiency of collector, mid-height temperature, heat transfer rate was considered. Gan et al. [9] did experiment on the solar absorption capacity of Al2 O3 nanofluids which is less than aluminum nanofluids. In practical Al2 O3 nanofluids, it has different properties of Al2 O3 . Less solar absorption did not result in convective heat transfer to the base fluids. Alim et al. [10] analyzed pressure drop and entropy generation in SFPC using the nanofluid CuO, Al2 O3 , SiO2 , TiO2 dispersed in water. The study was conducted for a volume flow rate (1 to 4 L/s and volume fraction (1 to 4%) for the absorber area of 1.51. Entropy analysis was carried out by the second law of thermodynamics. When these nanofluids were used as absorbing medium, the properties like viscosity, thermal conductivity and density were increased but specific heat decreased. By using these nanofluids, the entropy generation was reduced, where there was enhanced heat transfer rate with the increase of concentration. The objective of this study is to evaluate the performance enhancement of solar collector using Al2 O3 /water nanofluid to that of base fluid.

15.2 Experimental Setup The actual and schematic experimental setup is indicated as in Figs. 15.1 and 15.2, respectively. A pump was being used for the forced convection system to circulation nanofluid in the entire collector system. A plate-type heat exchanger was being used to transfer the heat load from the nanofluid system to the cold water. The cold water after being heated was being used for regeneration purpose. The data were taken for various mass flow rates, and volume fraction of nanofluid and base water and later

15 Study of Performance of Solar Flat Plate Collector Using …

Fig. 15.1 Experimental setup

Fig. 15.2 Schematic line diagram of the setup

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entire data were exported into the computer through interfaces. The entire system was calibrated several times, and readings were taken. (1) Nanofluid preparation: Nanofluid was prepared using two-step methodology by mixing Al2 O3 with distilled water. The mixture was well shacked and properly sonicated to prepare a stable nanofluid. Nanofluid with five different sample concentrations was prepared (0.01, 0.02, 0.04, 0.07 and 0.15 vol.%). (2) Collector fabrication: Collector was fabricated combining various subparts in an efficient way. The system was made completely leak proof. The complete system working is described below: (1) Process 1–2: The nanofluid is pumped from the tank and made to flow into the absorber plate, i.e., inlet to the collector. The mass flow rate is controlled by the control valve and is measured using a flow meter. (2) Process 2–3: The nanofluid after passing through absorber plate gets heated and is then to the heat exchanger to provide heat to the cold water. (3) Process 3–1: The nanofluid after being cooled is sent back to the nanofluid tank. (4) Process 4–3: Cold water is made to flow to the heat exchanger to gain the heat from nanofluid. (5) Process 3–5: The hot water coming out of heat exchanger is being stored in the storage tank and drawn out as par requirement.

15.3 Results and Discussion In this chapter, the efficiency collector with time, mass flow rate of water and nanofluid, the volume fraction of nanofluid, the temperature reduction parameter were studied. The volume fraction varied from 0.01 to 0.15 vol.%, i.e., 0.01, 0.02, 0.04, 0.07 and 0.15 vol.%., and the mass flow rate was varied from 0.5 l.p.m to 2.5 l.p.m. The variation of ambient temperature, the solar intensity with respect to time has also been plotted and studied. The various parameters have been defined as follows: ηi =

mC ˙ p T I Ac

where m˙ mass flow rate of water/nanofluid Cp Specific heat capacity of water/nanofluid.

C pn f

  Tb = C pw 1.036 − 0.0298φ − 0.07261 70

(15.1)

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where Tb  T I Ac

Fluid bulk temperature 20 °C  70° Nanoparticle volume concentration Temperature rise Solar intensity Collector area.

Figure 15.3 shows the outlet temperature variation of a solar collector with the mass flow rate for both water and 0.15 vol.% Al2 O3 nanofluids. The outlet temperature was found to be increasing with decreasing the mass flow rate. The reason behind this is that with a decrease in the mass flow rate, the mean contact time between the fluid and the absorber plate increases. Thus increasing heat transfer between the absorber plate and fluid thus higher outlet temperature. Temperature rise was higher in case of nanofluid because of enhanced in property of thermal conductivity and heat transfer coefficient. Figure 15.4 depicts the variation of efficiency and solar intensity with time for water. Both the curves have a similar nature, i.e., the maximum value at the noon and minimum values in morning and evening. A maximum solar intensity of 950 W/m2 was reached, and maximum efficiency reached was 53%. Figure 15.5 shows the variation of efficiency and solar intensity with time for 0.01 vol.% Al2 O3 nanofluid at a 2 l.p.m flow rate. Both the curves have a similar nature, i.e., the maximum value at the noon and minimum values in morning and evening. A maximum solar intensity of 970.92 W/m2 was reached, and maximum efficiency reached was 60%.

75

0.15vol.% Al2O3 nanofluid Water 0.15vol.% Al2O3 Water

Maximum temperature attained

70

65

60

55

50 45 0

0.5

1

1.5

2

2.5

Mass flow rate (kg/min.) Fig. 15.3 Maximum temperature attained for a different mass flow rate of water and nanofluid

3

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1000

Efficiency(%)

800

45%

700

40%

600 500

35% 30% 25%

Water

Solar intensity

Water efficiency

Solar Intensity

20%

400 300 200

Solar intensity(W/m 2)

900

50%

100 0

Time interval (hr.) Fig. 15.4 Variation of solar intensity and collector efficiency with time for water at a 2 l.p.m flow rate 65%

1200

60%

1000

800

50% 45%

600

40%

0.01vol.% Al2O3 nanofluid 35%

Solar intensity 0.01vol.% Al2O3 nanofluid

30%

400

Solar intensity (W/m 2)

Efficiency (%)

55%

200

Solar intensity

25%

0

Time interval (hr.) Fig. 15.5 Variation of solar intensity and collector efficiency with time for 0.01 vol.% Al2 O3 nanofluid at a 2 l.p.m flow rate

Figure 15.6 shows the variation of efficiency and solar intensity with time for 0.15 vol.% Al2 O3 nanofluid at a 2 l.p.m flow rate. Both the curves have a similar nature, i.e., the maximum value at the noon and minimum values in morning and evening. A maximum solar intensity of 1091.92 W/m2 was reached, and maximum efficiency reached was 77%.

155

80.00%

1400

75.00%

1200

Efficiency(%)

70.00%

1000

65.00% 800 60.00% 55.00%

600

0.155vol.% Al2O3 nanofluid Solar Intensity

50.00% 45.00%

400

Solar insolation(W/m2)

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0.155vol.% Al2O3 nanofluid 200

Solar Intensity

40.00%

0

Time interval (hr.)

Fig. 15.6 Variation of solar intensity and collector efficiency with time for 0.15 vol.% Al2 O3 nanofluid at a 2 l.p.m flow rate

Figure 15.7 shows efficiency variations with the mass flow rate for water and different nanofluids concentrations. Efficiency was found to be increasing with the mass flow rate. The slope was less steep in case of water when compared to nanofluid. 0.02vol.% Al2O3 nanofluid

70.00%

Water 65.00%

0.04vol.% Al2O3 nanofluid 0.02vol.% Al2O3 nanofluid

Efficiency (%)

60.00%

Water 0.04vol.% Al2O3 nanofluid

55.00% 50.00% 45.00%

Al2O3 40.00% 35.00% 30.00%

Mass flow rate(lit/min.)

Fig. 15.7 Variation of collector efficiency at a different mass flow rate for water and nanofluid

Efficiency(%)

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Water 0.01 vol.% Al2O3 nanofluid 0.07 vol.% Al2O3 nanofluid 0.15 vol.% Al2O3nanofluid Al2O3 nanofluid volume frac on=0.01vol.% Al2O3 nanofluid volume frac on=0.07vol.% Al2O3 nanofluid vol. fr.=0.15vol.%

Time(hr.)

Fig. 15.8 Variation of the efficiency of nanofluid and base fluid (water) with time

With the increase in mass flow rate of nanofluid, chances of the settlement of nanofluid are decreased, thus preserving the property of nanofluid. Figure 15.8 compares the efficiency achieved by base fluid and nanofluid having various concentration (0, 0.01, 0.07 and 0. 15 vol.%). The average efficiencies found in these cases were 0.46, 0.497, 0.578 and 0.73, respectively. The trend shown by these graphs shows that volume fraction nanofluid increased the energy efficiency of the solar collector also. The reason can be attributed to the abnormal increase in thermal conductivity and heat transfer coefficient of nanofluid with increasing volume fraction and temperature.

15.4 Conclusions The important conclusions that can be drawn out from this research work are as follows: (1) Energy efficiency increases with the increase in solar intensity. (2) The efficiency of solar collector increases with an increase in volume concentration of nanofluid; maximum enhancement in efficiency was observed at 0.15 vol.% Al2 O3 nanofluid concentration. The efficiency is found to increase from 46.3% for water to 73.1% for 0.15 vol.%Al2 O3 nanofluid at 2 l.p.m flow rate. (3) At higher temperature, the higher enhancement in performance of solar collector using nanofluid with respect to water is observed. (4) The maximum temperature rise of 14°C is obtained in the case of water and of 22 °C in case of 0.15 vol.% Al2 O3 nanofluid. (5) With increasing the mass flow rate, the efficiency is found to be increasing; the rise is less in case of water in comparison to nanofluid. (6) The maximum temperature 61.3 °C is attained in case of water and 70.1 °C in case of nanofluid.

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References 1. V. Devabhaktuni, M. Alam, SSSR Depuru, RC Green II, D Nims, C Near.: Solar energy: trends and enabling technologies. Renew Sustain Energy Rev 19, 555–64 (2013) 2. S.Y. Liu, Y.H. Perng, Y.F. Ho, The effect of renewable energy application on Taiwan buildings: what are the challenges and strategies for solar energy exploitation. Renew Sustain Energy Rev 28, 92–106 (2013) 3. S.S. Meibodi, A. Kianifar, H. Niazmand, O. Mahian, S. Wongwises, Experimental investigation on the thermal efficiency and performance characteristics of a flat plate solar collector using SiO2/EG–water nanofluids. Int. Commun. Heat Mass Transfer 65, 71–75 (2015) 4. Z. Said, R. Saidur, N.A. Rahim, M.A. Alim, Analyses of exergy efficiency and pumping power for a conventional flat plate solar collector using SWCNTs based nanofluid. Energy and Buildings 78, 1–9 (2014) 5. M. Faizal, R. Saidur, S. Mekhilef, M.A. Alim, Energy, economic and environmental analysis of metal oxides nanofluid for flat-plate solar collector. Energy Conversion Manage. 76, 162–168 (2013) 6. A.J. Moghadam, M. Farzane-Gord, M. Sajadi, M. Hoseyn-Zadeh, Effects of CuO/water nanofluid on the efficiency of a flat-plate solar collector. Exp. Therm. Fluid Sci. 58, 9–14 (2014) 7. T. Yousefi, F. Veisy, E. Shojaeizadeh, S. Zinadini, An experimental investigation on the effect of MWCNT-H2 O nanofluid on the efficiency of flat-plate solar collectors. Experimental Therm. Fluid Sci. 39, 207–212 (2012) 8. R. Nasrin, M.A. Alim, Semi-empirical relation for forced convective analysis through a solar collector. Solar Energy 105, 455–467 (2014) 9. Y. Gan, L. Qiao, Radiation-enhanced evaporation of ethanol fuel containing suspended metal nanoparticles. Int J Heat Mass Transfer 55, 5777–5782 (2012) 10. M.A. Alim, Z. Abdin, R. Saidur, A. Hepbasli, M.A. Khairul, N.A. Rahim, Analyses of entropy generation and pressure drop for a conventional flat plate solar collector using different types of metal oxide nanofluids. Energy Build. 66, 289–296 (2013)

Chapter 16

Thermo-Hydraulic Performance of Solar Air Heater Duct Provided with Conical Protrusion Rib Roughnesses Tabish Alam , Ashok Kumar

and Nagesh B. Balam

Abstract Thermal performance of solar air heater (SAH) considers only useful energy gain to air, propelling through SAH duct provided with artificial roughness. Effective efficiency takes into account of energy lost due to pressure drop in the duct as well as thermal energy gain in air. In this context, an attempt has been made to study the effect of roughness parameter of conical protrusion ribs on effective efficiency. Effective efficiencies of SAH duct have been computed using developed correlations of Nusselt number and friction factor. Plots of effective efficiencies of smooth and roughened duct have been plotted as a function of temperature rise parameter and Reynolds number. Based on effective efficiency criteria, values of roughness parameters have been optimized for different range of temperature rise parameter at different insolation. It is found that there exist sets of roughness parameters which depend on the range of temperature rise parameter. Keywords Solar air heater · Conical protrusion rib · Effective efficiency

16.1 Introduction Global demand for energy is increasing rapidly due to social, cultural and industrial development and is resulting in large amounts of per capita energy consumed. Conventionally, energy is extracted from exhaustible sources of energy which include coal, crude oil and natural gas. These conventional sources of energy are limited in nature and will be eventually consumed in a little span of time. As a consequence of consumption of conventional fuel, environmental pollution has tremendously increased. In order to minimize it, conventional energy resources have to be replaced by environmentally friendly renewable energy resources. Renewable energy is usable in different forms like solar, wind, tidal, geothermal, hydro and biomass energy. T. Alam (B) · A. Kumar · N. B. Balam CSIR-Central Building Research Institute, Roorkee 247667, India e-mail: [email protected] © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_16

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Among all renewable energies, solar energy is placed in top position due to its abundant quantity, omnipresent, clean and pollution-free nature. Solar energy may be utilized indirectly or directly, depending on the tasks such as heating, cooling, electricity generation and industry. Solar energy is best suited for heating process by means of solar air heater, which is simple in design and inexpensive. The main component of SAH is a collector which converts insolation into thermal energy of air. But it is noticed that performance of smooth collector is very poor due to very low convective heat transfer coefficient in between collector and flowing air. Convective heat transfer coefficient of collector is enhanced by application of artificial roughnesses [1–3], dimple ribs [4, 5], grooves [6, 7], tabulators [8], baffles [9, 10], blocks [11, 12] and so on. Sharma and Kalamkar [13] conducted review on various artificial roughnesses and indicated that dimple or protrusion ribs offered comparatively low friction factor and leading to improvement in thermo-hydraulic performance. Furthermore, the application of protrusion roughness in SAH may be attractive because protrusion ribs do not add extra mass on collector and it can be easily fabricated by means of punching the thin collector using indentation device. Enhanced convective heat transfer coefficient is accompanied with pressure drop penalty which causes to higher pumping power requirement. Better heat transfer rate with minimum pressure drop penalty is the main requirement of efficient collectors. So, it becomes necessary to evaluate hydraulic performance. It has been shown in previous studies of the author [14] on protrusion conical rib roughnesses in SAH that considerable frictional factors are found which cannot be neglected. In this context, effective efficiency has been evaluated analytically using correlations of Nusselt number and friction factor, developed in previous published paper [14]. Based on the effective efficiency curves at various insolations, roughness parameters have been presented.

16.2 Thermo-hydraulic Performance Evaluation Criteria Cortes and Piacentini [15] hypothesized that the real output of solar air heater should be expressed in terms of a net useful gain which should account for both thermal gain as well as the pumping power consumption. Since the net useful energy output of air cannot be evaluated by subtracting pumping power directly because it involves different forms of energy. The authors defined net useful gain as the difference of thermal gain and the equivalent thermal energy that would be required to produce pumping power, given as: Net useful gain = Q u −

Pm C

(16.1)

Hence, the effective efficiency is calculated by subtracting the lost mechanical energy from the total energy gain by air. Effective efficiency (ηeff ) is given below:

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Fig. 16.1 Protrusion rib roughness in SAH



Effective efficiency, ηeff

Q u − PCm = I AP

 (16.2)

where C = η f ηm ηtr ηthp is a conversion factor for conversion of mechanical energy to thermal energy. The typical value of ‘C’ with normal values of efficiencies being, ηtp = 0.35, ηtr = 0.92, ηm = 0.88 and ηf = 0.65, works out to be typically 0.18 [15].

16.3 System and Operating Parameters The schematic diagram of protrusion ribs’ roughness has been presented in Fig. 16.1. The roughness parameters have been expressed in terms of corresponding dimensionless parameters, namely relative rib height (e/D) and relative pitch rib pitch (p/e). The values of roughness parameters have been taken based on previous studies [14]. The duct parameter, range of roughness parameters and operating parameters are selected on the basis of actual application of solar air heater duct, and values are also presented in Table 16.1.

16.4 Computation Method Thermo-hydraulic analysis of smooth SAH duct and duct provided with protrusion rib roughness is very similar. So, procedure is to determine the absorbed insolation, useful heat gain and losses to environment are adopted as in case of conventional

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Table 16.1 Values/ranges of the system and operating parameters Symbol

System parameters Parameter

Value

N

Number of glass covers

1

H

Collector height

0.025 m

W

Collector width

0.3 m

L

Collector length

1.0 m

ki

Thermal conductivity of insulation

0.037 W/m K

te

Thickness of collector edge t e

0.1 m

ti

Thickness of back insulation

0.05 m

τα

Transmittance-absorptance product

0.8

Lg

Gap between collector and glass cover

0.025 m

εg

Emissivity of transparent glass sheet

0.88

εp

Emissivity of absorber plate

0.9

tg

Thickness of glass cover

0.002 m

β ti

Tilt angle

30°

e/D

Relative rib height

0.020–0.044

p/e

Relative rib pitch

6–12

Operating parameters Vw

Wind velocity

1.0 m/s

Ta

Ambient temperature

285 K

T /I

Temperature rise parameter

0.002–0.030 K.m2 /W

I

Insolation

600–1000 W/m2

duct of SAH. However, values of Nusselt number and friction factor are calculated from the correlations of Nu number and friction factor. These correlations have been established using the CFD results, and validations of CFD results have been presented in previous published paper. Results predicted by SST k-ε turbulence model were found in good agreement with results obtained from experimental work. The details of model validation have been presented in authors’ previous paper [14]. The optimum values of roughness parameters are presented as function of insolation (I) and temperature rise parameters (T /I). The silent features of mathematic modelling are presented in flowchart (Fig. 16.2).

16.5 Results and Discussions To understand the behaviour of effective efficiency, plots of useful heat energy gain to air, pumping power and collector temperature of SAH roughened with conical protrusion rib roughness have been presented as a function of Reynolds number for

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Fig. 16.2 Flowchart

relative rib height of 0.0289 and relative rib pitch of 10 as shown in Fig. 16.3. It can be seen that useful heat gain increases with increase in Reynolds number, and useful heat energy becomes nearly constant at high Reynolds number. This is due to fact that collector temperature decreases with increase in Reynolds number. For high Reynolds number, low values of collector temperature cause low thermal loss to environment, and consequently, higher useful energy gain to air. Pumping power due to conical protrusion ribs’ roughened surface increases with increase in Reynolds number, nevertheless, significant pumping power is not observed at low Reynolds number. However, pumping power increases sharply at high Reynolds number, which is due to fact that pumping power is depended on cube of Reynolds number. Effective efficiency of the roughened surface with conical protrusion ribs have been reckoned at all potential combinations of roughness parameters at insolation (I) of 1000 W/m2 . For comparison purposes, effective efficiency (ηeff ) of smooth duct has also been computed. Figure 16.4 prepares to show the effect of relative rib height (e/D) on effective efficiency (ηeff ) as a function of Reynolds number (Re) for

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Fig. 16.3 Variation of useful heat energy gain, pumping power and collector temperature as a function of Reynolds number

Fig. 16.4 Variation of effective efficiency w.r.t. Reynolds number at different relative rib height

optimum relative rib pitch (p/e) of 10 and insolation (I) of 1000 W/m2 . It can be clearly understood that effective efficiency (ηeff ) increases with increase in Reynolds number (Re) and thereafter decreases with further increase in Reynolds number (Re). However, maximum effective efficiency has been obtained at different values of temperature rise parameter (T /I). Decrease in effective efficiency beyond certain values temperature rise parameter is due to the fact that pumping power of blower is dominant over useful heat energy gain in an air. The values of optimum relative

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Fig. 16.5 Variation of effective efficiency w.r.t. temperature rise parameter at different relative rib height

rib height depend on the range of Reynolds number which offers maximum effective efficiency. For a low Reynolds number, i.e. Re < 11175, relative rib height of 0.044 is found to be optimum, beyond this value of the Reynolds number, i.e. 11175 < Re < 13103, the optimum value of relative rib height is 0.036. Furthermore, for 13103 < Re < 16772 and 16772 < Re < 21785, relative rib height of 0.0289 and 0.020 are found optimum, respectively. Further increase in Reynolds number beyond 21785, smooth duct is found to be best which offers maximum effective efficiency. Similarly, Fig. 16.5 shows the effect of relative rib height on effective efficiency as a function of temperature rise parameter at relative rib pitch of 10 and insolation of 1000 W/m2 . The trend of plots is similar to trends of plots of effective efficiency with respect to Reynolds number, presented in Fig. 16.4. The optimum values of relative rib height depend on the temperature rise parameters which give maximum effective efficiency. Smooth SAH duct yields the highest effective efficiency at lower value of temperature rise, i.e. 0.00362 < T /I, and beyond this value of temperature rise parameter, i.e. 0.00362 < T /I < 0.00456, the optimum value of relative rib height is found as 0.020. Afterwards, optimum values of relative rib height are found as 0.0289 and 0.036 for 0.00456 < T /I 0.0596 and 0.00569 < T /I 0.0067, respectively. Beyond the temperature rise parameter of 0.0067, relative rib height of 0.044 yields maximum effective efficiency. Effect of relative rib pitch on effective efficiency has been presented as a function of Reynolds number at optimum value of relative rib height of 0.0289 and insolation of 1000 W/m2 , as shown in Fig. 16.6. Different values of optimum relative rib height have been observed for different range of Reynolds number. For low value of

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Fig. 16.6 Variation of effective efficiency w.r.t. Reynolds number at different relative rib pitch

Reynolds number, i.e. Re < 15114, optimum relative rib pitch is found as 10, and relative rib pitch of 12 is found optimum for the range of Reynolds number of 15114 < Re < 21780. For 21780 < Re, smooth duct yields the highest effective efficiency. In a similar manner, effect of relative pitch on effective efficiency has also been presented (Fig. 16.7) as a function of temperature rise parameter at optimum relative height of 0.0289 and insolation of 1000 W/m2 . The optimum values of relative rib pitch are found as 12 and 10, and these corresponding ranges of temperature rise parameter are found as 0.00355 < T /I < 0.0051 and 0.0051 < T /I. However, smooth duct is found as optimum for low value of temperature rise parameter, i.e. T /I < 0.00355. It is clear from Figs. 16.4, 16.5, 16.6 and 16.7 that conventional smooth duct yields comparatively higher effective efficiency over roughened duct with conical protrusion rib for high temperature rise parameter and low Reynolds number (up to 21785). Due to increasing value of Reynolds number, pumping power requirement in roughened duct is more than of smooth duct, and therefore, smooth duct is more effective than roughened duct for high Reynolds number.

16.6 Conclusions On the basis of effective efficiency, thermo-hydraulic performance of solar air heater duct roughened with conical protrusions ribs has been determined. Results of convectional smooth duct have also been calculated and compared roughened duct. The following conclusions are drawn:

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Fig. 16.7 Variation of effective efficiency w.r.t. temperature rise parameter at different relative rib pitch

i. ii.

iii. iv.

v.

Due to conical shaped protrusion ribs, considerable increment in effective efficiency has been found in comparison to conventional smooth duct. Effective efficiency is increased with increase in Reynolds number, obtain maximum values and then start decreasing with further increase in Reynolds number. However, maximum values of effective efficiency are obtained at different Reynolds number at particular roughness parameters. Similar trend has been found when effective efficiency varies as a function of temperature rise parameter. Smooth SAH duct yields the highest effective efficiency when 0.00362 < T /I, and however, optimum values of relative rib height are found as 0.020, 0.0289 and 0.036 corresponding to the following ranges of temperature rise parameters 0.00362 < T /I < 0.00456, 0.00456 < T /I 0.0596 and 0.00569 < T /I 0.0067, respectively. The optimum values of relative rib pitch are found as 12 and 10, and these corresponding ranges of temperature rise parameter are found as 0.00355 < T /I < 0.0051 and 0.0051 < T /I. However, smooth duct is found as optimum for low value of temperature rise parameter, i.e. T /I < 0.00355.

References 1. E.A.M. Momin, J.S. Saini, S.C. Solanki, Heat transfer and friction in solar air heater duct with V-shaped rib roughness on absorber plate. Int. J. Heat Mass Transf. 45, 3383–3396 (2002)

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2. B.N. Prasad, J.S. Saini, Optimal thermohydraulic performance of artificially roughened solar air heaters. Sol. Energy 47, 91–96 (1991) 3. S. Singh, S. Chander, J.S. Saini, Investigations on thermo-hydraulic performance due to flowattack-angle in V-down rib with gap in a rectangular duct of solar air heater. Appl. Energy 97, 907–912 (2012) 4. R.P. Saini, J. Verma, Heat transfer and friction factor correlations for a duct having dimple-shape artificial roughness for solar air heaters. Energy 33, 1277–1287 (2008) 5. M.V. Sethi, N.S. Thakur, Correlations for solar air heater duct with dimpled shape roughness elements on absorber plate. Solar Energy 86, 2852–2861 (2008) 6. A.R. Juarker, Heat and Fluid Flow Characteristics of Rib-Groove Artificially Roughened Solar Air Heater (Indian Institute of Technology, Roorkee, 2005) 7. A.R. Juarker, J.S. Saini, B.K. Ghandi, Heat transfer and friction characteristics of rectangular solar air heater duct using rib-grooved artificial roughness. Solar Energy 80, 895–907 (2006) 8. S. Abraham, R.P. Vedula, Heat transfer and pressure drop measurements in a square crosssection converging channel with V and W rib turbulators experimental. Therm. Fluid Sci. 70, 208–219 (2016) 9. R. Karwa, B.K. Maheshwari, Heat transfer and friction in an asymmetrically heated rectangular duct with half and fully perforated baffles at different pitches. Int. Commun. Heat Mass Transf. 36, 264–268 (2009) 10. R. Karwa, B.K. Maheshwari, N. Karwa, Experimental study of heat transfer enhancement in an asymmetrically heated rectangular duct with perforated baffles. Int. Commun. Heat Mass Transf. 32, 275–284 (2005) 11. T. Alam, R.P. Saini, J.S. Saini, Experimental investigation of thermohydraulic performance of a rectangular solar air heater duct equipped with V-Shaped perforated blocks. Adv. Mech. Eng., pp. 1–11 (2005) 12. T. Alam, R.P. Saini, J.S. Saini, Heat transfer enhancement due to V-shaped perforated blocks in a solar air heater duct. Appl. Mech. Mater. 619, 125–129 (2005) 13. S.K. Sharma, V.R. Kalamkar, Thermo-hydraulic performance analysis of solar air heaters having artificial roughness–a review. Renew. Sustain. Energy Rev. 41, 413–435 (2015) 14. T. Alam, M.H. Kim, Heat transfer enhancement in solar air heater duct with conical protrusion roughness ribs. Appl. Therm. Eng. (Article in Press) (2017) 15. A. Cortes, R. Piacentini, Improvement of the efficiency of a bare solar collector by means of turbulence promoters. Appl. Energy 36, 253–261 (1990)

Chapter 17

Flocculation–Solar Distillation—an Integrated Energy-Efficient Technology for Desalination of Seawater Devlina Das and Nilanjana Mitra Abstract Although single-basin solar distillation (SD) has been extensively studied for desalination, and it has been recently explored for enhancing the rate of desalination by integrating it with other technologies. The current work focuses on developing an integrated technology consisting of flocculation and solar distillation as a cost and energy-effective alternative to existing technologies for seawater desalination. Experimental studies were carried out using a solar distillation set-up (SD), Floc-SD set-up and membrane distillation-SD (MD-SD) set-up at varying parameters, namely radiation, distillate volume, temperature variation, total dissolved solids removal (%), variation in pH and global efficiency (%) for a month duration. Results suggested a higher change in latent heat of vapourization in case of Floc-SD and MD-SD which enhanced the process efficiency. This was also in accordance with the temperature values obtained at the water/vapour interphase. A higher variation in pH was noted in case SD set-up, whereas the variation was less in case of Floc-SD set-up which was attributed to the maintenance of dissolved oxygen in the latter case. Based on the results, it could be inferred as that Floc-SD was as effective as MD-SD in terms of desalination and allied pollutant remediation. Keywords Solar distillation · Flocculation · Desalination · Membrane distillation

17.1 Introduction Water scarcity is one of the main challenges of the century that many societies around the world are already facing. Throughout the last century, use and consumption of water grew at twice the rate of population growth. With an increase in regional freshwater scarcity, the practice of seawater desalination is rapidly increasing. Many developing countries show an increasing interest to the seawater and brackish water D. Das (B) Department of Biotechnology, PSG College of Technology, Coimbatore, India e-mail: [email protected]; [email protected] N. Mitra School of Bio Sciences and Technology, Vellore Institute of Technology, Vellore, India © Springer Nature Singapore Pte Ltd. 2020 S. Singh and V. Ramadesigan (eds.), Advances in Energy Research, Vol. 2, Springer Proceedings in Energy, https://doi.org/10.1007/978-981-15-2662-6_17

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desalination by solar distillation owing to its ample availability and non-extinguishing nature [1, 2]. Different types of solar stills have been developed and commercialized. Experimental and theoretical studies have been undertaken in order to modify the solar still to improve its yield [3]. The production of distillate was estimated at varying inclinations [4], climatic parameters [5], external/internal reflector enabled [6] and material modifications [7]. However, the major challenge faced in this regard is time and distillate output [8]. In this context, mature thermal desalination technologies, namely multi-stage flash (MSF), multi-effect distillation (MED) and membrane distillation have been used as an integrated technology with solar desalination [9]. Membrane distillation consists in a thermal process in which only vapour molecules pass through a porous hydrophobic membrane. The feed is in direct contact with the surface of the membrane but it does not penetrate thanks to the hydrophobic nature of the membrane. The driving force of the process is the vapour pressure difference through the membrane [10–12]. Although an efficient technology with reference to brine and dissolved solids removal, yet the major challenge related to this technology is cost [13]. Organic flocculation agents are being used for waste water treatment due to their economic and simple manageable properties. During the last years, chemically modified polysaccharides like cellulose, chitosan, and starch have attracted enormous attention due to their numerous advantages, for example, biocompatibility, diversity in molar mass or charge density, low energy and zero electricity requirements [14]. Chitosan has advantages like nontoxic, ecologically friendly, biodegradable, not caustic and, therefore, easy and safe to handle [15]. The binding and flocculating capacity of chitosan could be attributed to the hydroxyl and aliphatic amino moieties in the structure. The binding of ions is aided by the quaternary amino groups [16]. This article reports the use of a composite made of chitosan, clay and polyaniline nanoparticles for a prior pre-treatment of seawater before being desalinated using a solar distillation still. This integrated approach has been designed in order to come up with a cost and energy-efficient alternative of conventionally used solar stills and energy intensive techniques.

17.2 Experimentation 17.2.1 Instrument Details The experiments were performed using a single-basin solar still. The system consisted of a sloping pane of glass is maintained by apt frame. The basin was covered by the frame and was preserved tight for diminishing the vapour leakage. The basin of the still was made out of glass fibres and a resin layer as insulator. The glass had an area of 6.45 m2 . The absorber is covered with seawater of 5 cm in depth, characterized by a salinity of 40,000 mg/L. The side and the lower walls are insulated by a glass woollen layer of 46 mm of thickness and a thermal performance to minimize the heat

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loss into the outside. The whole of the device is fixed on a metal support. A benchscale permeate gap membrane distillation (PGMD) module was fabricated following standard protocol [17]. The half-shells were made of polycarbonate material with a thickness of 10 mm and dimensions of 250 × 160 mm. The effective area was 125 cm2 . The active layer of the membrane was made of PTFE, while the backing material was of PP. The nominal pore size and the porosity were 0.2 μm. The test module was integrated with two hydraulic loops, one for the hot feed and one for the cooling.

17.2.2 Chemicals and Flocculant Preparation All chemicals including chitosan, polyaniline nanoparticles (