Development of Solar Thermal Harvesting Technology Which Meets the Needs of Industry

Qiyuan Li

A thesis in fulfilment of requirements for the degree of Doctor of Philosophy

School of Mechanical and Manufacturing Engineering Faculty of Engineering The University of New South Wales

December 2016 PLEASE TYPE 1. THE UNIVERSITY OF NEW SOUTH WALES

2. Thesis/Dissertation Sheet

Surname or Family name: Li

First name: Qiyuan Other name/s:

Abbreviation for degree as given in the University calendar: PhD of Engineering (Research)

School: Mechanical and Manufacturing Engineering Faculty: Engineering

Title: Development of solar thermal harvesting technology which meets the needs of industry

Abstract 350 words maximum: (PLEASE TYPE) Rooftop integrated solar thermal systems can potentially supply medium to high temperature (100-400 °C) heat to industrial and commercial applications. A key barrier for conventional concentrated solar systems is that integration with rooftops is relatively complex and cumbersome. To avoid wind loading issues and for ease of installation, low-profile collectors with form factors similar to photovoltaic panels were examined in the present study. In order to advance the field of solar thermal technology for commercial and industrial applications, this study investigated the potential for utilizing compact optical and thermal concentrators. As a major aspect of this study, an innovative low-profile optical concentrator (<10 cm in height) was designed, developed and systematically analysed. As another potential path for innovation, thermal concentrators (e.g. passive solar receivers consisting of heat pipes and highly conductive fins or solar receivers integrated with active thermoelectric heat pump elements) were explored in this study. Also within this research, the effect of modifying the absorber from a standard selective surface (black chrome-coated copper tube) to a selective volumetric absorber (nanofluid contained within ITO-coated glass tube) was investigated. Several avenues for improvement for both receivers and the entire design were identified from these analyses. Due to the fact industrial demand may not match the transient operation time of concentrating thermal collectors, the integration of latent thermal storage in its limited free space was also investigated. To minimize the cost, shell-and-tube storage units were analysed for their technical feasibility in a characteristic industrial application. The optimum configuration was found to provide a solar fraction of ~60% and an annual charging efficiency of ~100%, indicating technical feasibility for this integrated collector/storage (ICS) system. Lastly, the techno-economic feasibility of these innovations for air-conditioning application was investigated. Overall, this study reveals that the proposed low-profile collector with/without an integrated storage could meet the needs of industrial and commercial heating applications. Due to the fact that solar-derived industrial heat production is an emerging market, it was ultimately found that a nominal amount of further research and manufacturing cost reductions could potentially help these types of systems to realize unsubsidized economic viability going forward.

Declaration relating to disposition of project thesis/dissertation I hereby grant to the University of New South Wales or its agents the right to archive and to make available my thesis or dissertation in whole or in part in the University libraries in all forms of media, now or here after known, subject to the provisions of the Copyright Act 1968. I retain all property rights, such as patent rights. I also retain the right to use in future works (such as articles or books) all or part of this thesis or dissertation. I also authorise University Microfilms to use the 350 word abstract of my thesis in Dissertation Abstracts International (this is applicable to doctoral theses only).

…………………………………………………………… ……………………………………..……………… ……….…………………...…….…19/12/2016 Signature Witness Date

The University recognises that there may be exceptional circumstances requiring restrictions on copying or conditions on use. Requests for restriction for a period of up to 2 years must be made in writing. Requests for a longer period of restriction may be considered in exceptional circumstances and require the approval of the Dean of Graduate Research.

FOR OFFICE USE ONLY Date of completion of requirements for Award:

3. THIS SHEET IS TO BE GLUED TO THE INSIDE FRONT COVER OF THE THESIS

ORIGINALITY STATEMENT

‘I hereby declare that this submission is my own work and to the best of my knowledge it contains no materials previously published or written by another person, or substantial proportions of material which have been accepted for the award of any other degree or diploma at UNSW or any other educational institution, except where due acknowledgement is made in the thesis. Any contribution made to the research by others, with whom I have worked at UNSW or elsewhere, is explicitly acknowledged in the thesis. I also declare that the intellectual content of this thesis is the product of my own work, except to the extent that assistance from others in the project's design and conception or in style, presentation and linguistic expression is acknowledged.’

Signed ……………………………………………......

Date ……………………………………………...... 19/12/2016

‘I hereby grant the University of New South Wales or its agents the right to archive and to make available my thesis or dissertation in whole or part in the University libraries in all forms of media, now or here after known, subject to the provisions of the Copyright Act 1968. I retain all proprietary rights, such as patent rights. I also retain the right to use in future works (such as articles or books) all or part of this thesis or dissertation. I also authorise University Microfilms to use the 350 word abstract of my thesis in Dissertation Abstract International (this is applicable to doctoral theses only). I have either used no substantial portions of copyright material in my thesis or I have obtained permission to use copyright material; where permission has not been granted I have applied/will apply for a partial restriction of the digital copy of my thesis or dissertation.’

Signed ……………………………………………......

Date ……………………………………………...... 19/12/2016

AUTHENTICITY STATEMENT

‘I certify that the Library deposit digital copy is a direct equivalent of the final officially approved version of my thesis. No emendation of content has occurred and if there are any minor variations in formatting, they are the result of the conversion to digital format.’

Signed ……………………………………………......

Date ……………………………………………...... 19/12/2016

iii

Acknowledgements

I would like to express my deepest gratitude to my supervisor, Dr. Robert Taylor, for providing me the opportunity and scholarships to pursue the degree and expand my knowledge. His extraordinary guidance and continuous support ensured that I could achieve the objectives throughout this project. I have been extremely lucky to have a supervisor like him. I am also greatly grateful to my co-supervisor, Dr. Jason Scott and project leader Prof. Gary Rosengarten for their consistent guidance and constructive advice.

I wish to thank my parents and all of my family for their unconditional love and consistent support. Special mention to my wife Lu Ying Jin and our first baby, they energized me to confront challenges with passion and responsibility.

I gratefully acknowledge receiving assistance and advice from our solar thermal research group, including Cheng Zheng, Sara Mesgari, Natasha Hjerrild, Felipe Crisostomo, Yasitha L. Hewakuruppu during my study. My sincere appreciation is extended to the solar lab manager Karl Morrison and workshop officers Martyn Sherriff, Ian Cassapi and Terry Flynn for their assistance in prototyping. My special thanks also go to Ali Shirazi, S. Saeed Mostafavi Tehrani, Albert Woffenden, and Yinfeng Wang for their helpful cooperation and assistance on experiments and publications.

Last but not least, I would like to take this opportunity to thank the School managers/officers Mary Stringer, Loyce Davis, Prof. Zhongxiao Peng and all the MME School staff for the great and friendly workplace created by their warmhearted and caring management. Additionally, I also would like to gratefully acknowledge the Australian Renewable Energy Agency (ARENA) for funding Micro Urban Solar Integrated Concentrator (MUSIC) project.

Abstract

Global demand for heating accounts for more than 50% of primary energy consumption. Rooftop integrated solar thermal systems do not have to be relegated to only supplying hot water, they can also supply medium to high temperature (100-400 °C) heat to industrial and commercial applications. At present the suite of technologies which can achieve high temperatures is limited, but there is a huge potential market for solutions which can provide high quality heat (to offset volatile gas prices behind the meter), while also minimizing our ecological footprint (e.g. land-use). A key barrier for most concentrated solar systems is that integration with rooftops is relatively complex and cumbersome in comparison with photovoltaic (PV) panels. To avoid wind loading issues and for ease of installation, low-profile collectors with form factors similar to PV panels are desirable.

In order to advance the field of solar thermal technology for commercial and industrial applications, this study investigated the potential for utilizing compact optical and thermal concentrators. As a major aspect of this study an innovative low-profile optical concentrator was designed, developed (<10 cm in height with a novel internal tracking mechanism) and systematically investigated to demonstrate its potential to deliver heat energy in the 100-250 °C range.

Thermal concentrators (<10 cm in height), were explored as another key aspect of this study to achieve high energy concentration for solar thermal applications. While heat pipes with highly conductive fins were found to be the most promising ‘passive’ concentrator, thermoelectric heat pumps integrated into the receiver of solar thermal collectors showed the highest development potential as an ‘active’ energy concentrator.

Within this system, experimental and numerical methods were used to investigate the effect of modifying the absorber from a standard selective surface to a selective volumetric absorber (consisting of a multi-walled carbon nanotube nanofluid contained within an ITO-coated glass tube). A comparative analysis was used to gauge its feasibility against a conventional surface absorber (consisting of a black chrome-coated copper tube). The analysis revealed that the vacuum-packaged volumetric receiver has a 14-20% lower thermal efficiency than a conventional selective surface over the same

operating range (from 80-200 °C). Several avenues for improvement for both receivers and the entire design were identified from these detailed analyses. The analysis revealed that the instantaneous thermal efficiency could be enhanced from ~45% to ~55% by using lower emissivity coating (e.g. TiNOX) when working at 200 °C, while the annual thermal efficiency could be increased from 15% to 25% by integrating prism arrays to improve the daily optical efficiency.

As the concentrator has a limited operation time, the integration of thermal storage was also studied in detail. Due to limited ‘free’ space in the design, it was found that the best storage option would be a latent heat thermal energy storage (LHTES) unit. To minimize the cost, relatively simple shell-and-tube storage units were analyzed for their techno-economic feasibility in a characteristic industrial application – a dairy production that requires constant thermal power at 120-150 ˚C. Solar fractions of up to 65% were achievable with some design alternatives, but the optimum techno-economic design had a solar fraction of ~35% where an annual charging efficiency of almost 100% was achievable (using a non-evacuated receiver and 7 storage pipes in each module). It was also found that if the capital cost of the integrated collector and LHTES tank can be reduced by 50% from ~700 $/m2 to ~350 $/m2 through mass production and further design optimization, industrial process heat with a levelized cost of energy (LCOE) price of ~0.065 $/kWh can be derived from this ISC system.

To investigate the real-world viability of these designs, a system-level techno-economic performance analysis of the collector (with/without possible improvements) was undertaken using a validated, hybrid CFD /TRNSYS (Transient System Simulation Tool, version 17) model. For the solar air-conditioning case study (located in Sydney, Australia), a double-effect chiller was coupled with the proposed collector array to supply both the cooling and heating demand of a large hotel building. The maximum cooling and heating demands of the building were 965kW and 520 kW, respectively. Accordingly, an absorption chiller with a nominal cooling capacity of 1,163 kW was selected. The solar field was sized to achieve a total solar fraction of 50%. The simulation results showed that a specific collector area of 3 m2 per kW cooling and an optimal storage tank specific volume of 40 L/m2 were sufficient to cover 50% of the load requirement of the building. An economic analysis of the system indicated that a LCOE price of ~0.6 $/kWh can be derived from this solar cooling system.

Overall, this study reveals that the proposed low-profile, concentrated solar thermal collector with/without integrated storage can meet the technical needs of industrial and commercial heating applications. This study also identified key avenues of research which could yield improvements for these types of systems. Due to the fact that solar-derived industrial heat production is an emerging market, it was ultimately found that with a few improvements and cost reductions (through mass production) these types of systems can potentially realize unsubsidized economic viability going forward.

vii

List of publications

Journal Papers:

[1] Li, Q., Shirazi, A., Zheng, C., Rosengarten, G., Scott, J., & Taylor, R. A. (2016). Energy concentration limits in solar thermal heating applications. Energy, 96, 253-267. [2] Li Q., C. Zheng, S. Mesgari, Y.L. Hewkuruppu, N. Hjerrild, F. Crisostomo, G. Rosengarten, J.A. Scott, R.A. Taylor, (2016). Experimental and numerical investigation of volumetric versus surface solar absorbers for a concentrated solar thermal collector, Solar Energy, 136 349-364. [3] Gu, X., Taylor, R. A., Li, Q., Scott, J. A., &Rosengarten, G. (2015). Thermal analysis of a micro solar thermal collector designed for methanol reforming. Solar Energy, 113, 189-198. [4] S. Mesgari, R.A. Taylor, N.E. Hjerrild, F. Crisostomo, Q. Li, J. Scott, (2016). An investigation of thermal stability of carbon nanofluids for solar thermal applications, Solar Energy Materials and Solar Cells, 157, 652-659. [5] LI, Q., Tehrani S.S.M., Taylor, R. A., Techno-Economic Analysis of a Concentrating Solar Collector with a Built in Shell and Tube Latent Heat Thermal Energy Storage (2016) (Revision has been submitted on Energy on 19th Dec, 2016, ) [6] Li, Q., Zheng, C., Shirazi, A., Bany Mousa O., Moscia F., Scott, J., Taylor, R. (2016). Design and Analysis of a Medium-Temperature, Concentrated Solar Thermal Collector for Air-conditioning Applications (Revision has been submitted on ‘Applied Energy’ by 19th Dec 2016) Patents:

[1] Zheng, C., Taylor, R. A., LI, Q.; Rosengarten, G., A radiative energy concentrator (Australian Standard Patent Application No. 2015210329, Date of filing: 3th August 2015) pericles.ipaustralia.gov.au/ols/auspat/applicationDetails.do?applicationNo=2015210329 [2] LI, Q., Taylor, R. A., Zheng, C., Rosengarten, G., A radiative energy concentrator (Chinese innovative Patent Application No. 201510479100.1, Date of filing: 6th August 2015) http://cpquery.sipo.gov.cn/ [3] LI, Q., Tehrani S.S.M., Taylor, R. A., A radiation energy collection system (Australian Patent Application No. 2016903413, Date of filing: 26th August 2016) http://pericles.ipaustralia.gov.au/ols/auspat/applicationDetails.do?applicationNo=2016903413 Conference papers:

[1] R. Taylor, Q. Li, G. Rosengarten, Optical Design & Engineering Bright ideas for the future of industrial heating. SPIE. Doi: 10.1117/2.1201602.006362 [2] Li, Q., Zheng, C., Mesgari, S., Hewakuruppu, Y., Hjerrild, N., Crisostomo, F., . . . Taylor, R. (2015). Experimental investigation of a nanofluid absorber employed in a low-profile, concentrated solar thermal collector. In SPIE Micro+Nano Materials, Devices, and Systems. Sydney, New South Wales, Australia. doi:10.1117/12.2202513 [3] Li, Q., Zheng, C., Gu, X., Rosengarten, G., Hawkes, E., Yang, M., . . . Taylor, R. (2014). Design and analysis of a low-profile, concentrating solar thermal collector. In N. Kasagi (Ed.), International Heat Transfer Conference 15. Kyoto, Japan: Bengell House. doi:10.1615/IHTC15.sol.008611 [4] Zheng, C., Li, Q., Rosengarten, G., Hawkes, E., & Taylor, R. (2015). A new optical concentrator design and analysis for rooftop solar applications. In R. Winston, & J. Gordon (Eds.), Proceedings of SPIE Vol 9572, Non-imaging Optics: Efficient Design for Illumination and Solar Concentration XII Vol. 9572. San Diego, California, United States: SPIE. doi:10.1117/12.2186477

viii

Contents

ACKNOWLEDGEMENTS ...... IV ABSTRACT ...... V LIST OF PUBLICATIONS ...... VIII CONTENTS ...... IX LIST OF FIGURES ...... XIII LIST OF TABLES ...... XX NOMENCLATURE ...... XXII CHAPTER 1 ...... 1 1. INTRODUCTION ...... 1 1.1 Research Background ...... 1 1.2 Objectives ...... 3 1.3 Thesis structure ...... 4 CHAPTER 2 ...... 5 2. LITERATURE REVIEW ...... 5 2.1 Medium temperature solar thermal applications ...... 5 2.1.1 Solar derived industrial/commercial heating applications ...... 6 2.1.2 Solar-derived thermal-driven cooling applications ...... 8 2.2 Solar thermal collectors: The state-of-the-art ...... 10 2.2.1 Solar thermal collector overview...... 10 2.2.2 Performance and cost ...... 12 2.3 Optical concentration ...... 13 2.3.1 Refractive concentrators ...... 14 2.3.2 Reflective concentrators ...... 17 2.3.3 Hybrid concentrators: utilizing both refractive and reflective elements 23 2.4 Thermal concentration ...... 26 2.5 Mitigating radiative loss in medium temperature collectors ...... 27 2.5.1 Selective surface absorbers...... 27 2.5.2 Volumetric absorbers ...... 29 2.6 Integrated storage ...... 30 2.6.1 Thermal Storage Overview...... 31 ix

2.6.2 Solar collector/storage integration ...... 31 2.7 Chapter Summary ...... 33 CHAPTER 3 ...... 35 3. INNOVATIVE OPTICAL CONCENTRATION ...... 35 3.1 Optical design and analysis...... 36 3.1.1 Design of the Double Lens ...... 37 3.1.2 Design of the Cylindrical Lens-CPC system ...... 39 3.1.3 Design of the Fresnel Lens-CPC system ...... 41 3.2 Optical Performance ...... 42 3.3 Chapter Summary ...... 45 CHAPTER 4 ...... 46 4. COLLECTOR/PACKAGE DESIGN AND ANALYSIS ...... 46 4.1 Collector/Package design ...... 47 4.1.1 Design 1: Planar vacuum chamber ...... 47 4.1.2 Design 2: Evacuated glass tubes ...... 49 4.2 Performance Analysis ...... 53 4.2.1 Theoretical thermal performance analysis...... 53 4.2.2 Collector thermal performance simulation ...... 58 4.2.3 Thermal analysis summary ...... 61 4.3 Chapter Summary ...... 64 CHAPTER 5 ...... 65 5. INNOVATIVE THERMAL CONCENTRATION ...... 65 5.1 Passive concentrators ...... 67 5.1.1 Thermal conductive concentrator ...... 67 5.1.2 Heat pipe thermal concentrator ...... 72 5.1.3 Radiative concentrators ...... 77 5.2 Active thermal concentrators ...... 80 5.2.1 Forced convection thermal concentrators ...... 80 5.2.2 Heat pump thermal concentrators ...... 83 5.2.2.1 Vapor compression thermal concentrator ...... 84

5.2.2.2 Thermoelectric heat pump thermal concentrator ...... 89

5.3 Summary ...... 94 CHAPTER 6 ...... 96 6. INNOVATIVE SOLAR ABSORBERS ...... 96 6.1 Absorber design and analysis ...... 97 6.1.1 Optical design ...... 97 6.1.2 Thermal design ...... 101 6.1.3 Optical efficiency ...... 103 6.2 CFD analysis ...... 104 6.2.1 2-D Simulation ...... 105 6.2.2 3-D Simulation ...... 108 6.3 Absorber prototyping ...... 113 6.3.1 Nanofluid preparation ...... 113 6.3.2 Absorber assembly ...... 115 6.4 Summary ...... 117 CHAPTER 7 ...... 119 7. COLLECTOR PERFORMANCE TESTING ...... 119 7.1 Experimental Setup ...... 119 7.1.1 Non-vacuum prototype and loop ...... 119 7.1.2 Vacuum glass insulated prototype and loop ...... 121 7.2 Experimental Results and Discussion ...... 124 7.2.1 Thermal Efficiency Results ...... 124 7.2.2 Incident Angle Modifier Test ...... 131 7.3 Summary ...... 133 CHAPTER 8 ...... 134 8. INNOVATIVE ICS SOLAR SYSTEM ...... 134 8.1 Integrated collector/storage design and analysis ...... 134 8.1.1 Design overview ...... 134 8.1.2 Optical and thermal performance ...... 136 8.1.3 Shell and Tube LHTES ...... 138 8.2 Integrated collector and LHTES system ...... 142 8.2.1 Description of case study ...... 142

8.2.2 Design alternatives ...... 144 8.2.3 Annual performance analysis and system size ...... 145 8.2.3.1 System Sizing via technical metrics ...... 145

8.2.3.2 System sizing via economic analysis ...... 151

8.3 Conclusion ...... 156 CHAPTER 9 ...... 158 9. SYSTEM TECHNO-ECONOMIC ANALYSIS ...... 158 9.1 Collector future improvements ...... 158 9.2 Solar Absorption Cooling ...... 166 9.3 SHC System Analysis ...... 171 9.4 Economic Analysis ...... 175 9.5 Conclusion ...... 178 CHAPTER 10 ...... 180 10. CONCLUSION AND RECOMMENDATIONS FOR FUTURE WORK ...... 180 Appendix A ---Tracking code ...... 185 Appendix B--- A MATLAB code for analysis of heat pump ...... 192 Appendix C--- Prototype assembly drawings ...... 194 Appendix D--- Prototype parts and price list ...... 195 Appendix E--- Collector mass production cost estimation ...... 199 REFERENCES ...... 201

xii

List of Figures

Figure 1.1 Annual industrial heating energy consumption (in Europe 2005) [4] ...... 3

Figure 2.1 Categorization of tracking approaches for concentrating solar collectors .... 10

Figure 2.2 Categorization of energy concentration techniques ...... 14

Figure 3.1 Ray tracing results of a single lens ...... 38

Figure 3.2 Optical design (a) radii of curvature (unit: mm) and refractive indices chosen for double lens (b) optical simulation of double lens ...... 38

Figure 3.3 Geometry of optical system (unit: mm) ...... 39

Figure 3.4 Simulation the performance of CPC with different incident angle ...... 40

Figure 3.5 Simulation of Lens with CPC with different incident angle ...... 40

Figure 3.6 Dimensional drawing of the optimised double linear Fresnel lens design (units: mm) ...... 41

Figure 3.7 Simulation of Lens with CPC with different incident angle ...... 42

Figure 3.8 Optical efficiency comparison as a function of incidence angle ...... 43

Figure 3.9 Transverse IAM comparison as a function of incidence angle ...... 44

Figure 4.1 Drawings of the proposed design (a) 3D view of collector; (b) exploded view of the collector; (c) sketch of the collector in the cross section ...... 48

Figure 4.2 Ray tracing diagram for package with (a) 0o incident light angles; (b) 45o incident light angles ...... 49

Figure 4.3 Collector design: (a) 3D end view of the collector; (b) transverse cross-sectional view ...... 50

Figure 4.4 Ray tracing results for (a) 0o incident light angles; (b) 45o incident light angles...... 51

Figure 4.5 Open Loop Control System: PC, Arduino microprocessor (μP) board, Motor Board and stepper motor ...... 51

Figure 4.6 Closed Loop Control System with position measurement feedback...... 53

xiii

Figure 4.7 Heat loss schematic of the collector (a) modes of internal heat transfer (b) equivalent thermal resistance network ...... 54

Figure 4.8 Schematic of the three dimensional model built in ANSYS Icepak ...... 58

Figure 4.9 Temperature distribution of fluid Therminol VP-1: Tin = 200 ℃, Tout= 255 ℃,

푇 = 228 ℃ (v_wind = 10 m/s, Ta = 20 ℃) ...... 60

Figure 4.10 The temperature distribution on the collector cross section (a) V_wind = 0 m/s; (b) V_wind = 10 m/s ...... 61

Figure 4.11 Efficiency of the collector with different ambient wind speed (퐼= 1000 W/m2, 퐼푏푒푎푚= 900 W/m2) by theoretical and simulation analysis and performance comparison with Chromasun MCT ...... 62

Figure 4.12 Collector daily efficiency with variation of sun incident angle (퐼= 1000 W/m2, 퐼푏푒푎푚= 900 W/m2, outlet temp.: 220 ℃) ...... 63

Figure 4.13 Stagnation temperature as a function of effective optical efficiency and concentration ratio for 퐼= 1,000 W/m2, 퐼푏푒푎푚= 900 W/m2 – dashed arrows indicate directions of possible improvement...... 63

Figure 5.1 General schematic of thermal concentration ...... 66

Figure 5.2 Categorization of thermal concentration techniques ...... 67

Figure 5.3 Thermal analysis of a conductive thermal concentrator in a vacuum: (a) schematic of conductive thermal concentration; (b) equivalent thermal resistance network ...... 68

Figure 5.4 The effective thermal concentration and working efficiency of conductive concentrators: (a) copper concentrator, kcu=380 W/m-K; (b) CNT concentrator, kCNT =3,000 W/m-K ...... 71

Figure 5.5 Thermal analysis of a heat pipe with a flat plate absorber concentrator in a vacuum: (a) schematic of heat pipe thermal concentration; (b) equivalent thermal resistance network ...... 75

Figure 5.6 The effective thermal concentration ratio and working efficiency of heat pipe with absorber concentrator: (a) heat pipe with Cu absorber in a vacuum; (b) heat pipe with CNT absorber in a vacuum ...... 76

xiv

Figure 5.7 Thermal analysis of a radiative concentrator in a vacuum: (a) schematic of radiative thermal concentration; (b) equivalent thermal resistance network ...... 78

Figure 5.8 The effective thermal concentration of radiative thermal concentrators ...... 79

Figure 5.9 Thermal analysis of a forced convection thermal concentrator: (a) schematic of forced convection thermal concentration; (b) equivalent thermal resistance network81

Figure 5.10 The effective thermal concentration of forced convection thermal 2 concentrators, the output heat exchanger, Uex = 7.5 kW/m K; ...... 83

Figure 5.11 Schematic of solar assisted heat pump system ...... 84

Figure 5.12 The effective thermal concentration, system efficiency and COP of a vapor compression refrigeration cycle: (a) Cr|eff and  variation with evaporating temperature and compressor pressure ratio; (b) COP comparison between Rankine and Carnot refrigeration cycle systems with varied evaporating temperature and compressor pressure ratio ...... 88

Figure 5.13 Schematic of thermoelectric heat pump thermal concentrator ...... 89

Figure 5.14 The effective thermal concentration, system efficiency and COP of an available thermoelectric heat pump thermal concentrator: (a) ZTm (figure of merit) = 1, TH-TC= 10 °C, kCu absorber=380 W/m-K, output heat exchanger Uex = 7.5 kW/m2K; (b) ZT = 1, TH-TC= 70 °C, kCu absorber=380 W/m-K, output heat exchanger Uex = 7.5 kW/m2K ...... 91

Figure 5.15 The effective thermal concentration, system efficiency and COP of advanced thermoelectric heat pump thermal concentrator: (a) ZT = 2, TH-TC= 10 °C, kCNT Absorber =3,000 W/m-K, output heat exchanger Uex = 7.5 kW/m2K; (b) ZT = 2, TH-TC= 70 °C, kCNT Absorber =3,000 W/m-K, Uex = 7.5 kW/m2K ...... 92

Figure 6.1 Cross section of concentrator and receivers (glass tube or coated copper tube) ...... 97

Figure 6.2 Ray tracing results for the collector: (a) details of ray paths through the copper tube, (b) details of ray paths to the glass tube ...... 98

Figure 6.3 Solar weighted absorbance of water-based MWCNT nanofluids at different concentrations and receiver depths ...... 100

Figure 6.4 Design proposal of (a) a vacuum glass tube insulated surface absorber (b) a vacuum glass tube insulated volumetric absorber ...... 101

Figure 6.5 Meshing of the 2D transverse collector cross–section...... 106

Figure 6.6 Temperature distribution for the collector cross section with coated copper tube receivers (without vacuum insulation) ...... 106

Figure 6.7 Absorber heat loss at different absorber temperatures: (a) black chrome-coated copper tube receiver (b) nanofluid volumetric absorber (contained within a glass tube) ...... 107

Figure 6.8 Overall heat loss and overall heat loss coefficient for different absorber conditions (a) overall heat loss for absorber in a vacuum; (b) overall heat loss coefficient for absorbers with/without vacuum insulation ...... 107

Figure 6.9 Meshing of the 3D receiver tube interior ...... 109

Figure 6.10 Temperature distribution of the working fluid with an inlet temperature of 100 °C: (a) Therminol 55 and MWCNTs-Therminol 55 nanofluid on the axial cross sections of the absorber without vacuum; (b) Therminol 55 on the radial cross sections at middle of the BCCCT absorber without vacuum (at 0.75 meters); (c) MWCNTs-Therminol 55 nanofluid on the radial cross sections of nanofluid absorber without vacuum (at 0.75 meters) ...... 110

Figure 6.11 Temperature distribution of the receivers with an inlet temperature of 200 °C (a) Therminol 55 and MWCNTs-Therminol 55 nanofluid on the axial cross-sections of the vacuum insulated BCCCT and nanofluid absorber; (b) Therminol 55 on the radial cross-sections at middle of the vacuum insulated BCCCT absorber (at 0.75 meters); (c) MWCNTs-Therminol 55 nanofluid on the radial cross-sections at middle of the vacuum insulated nanofluid absorber (at 0.75 meters) ...... 111

Figure 6.12 CFD simulation efficiency results ...... 112

Figure 6.13 TEM image of potassium persulfate functionalized MWCNTs ...... 114

Figure 6.14 Absorbance of the prepared nanofluid (UV-vis-NIR spectra of KPS treated CNTs in DI) ...... 114

xvi

Figure 6.15 Pictures of the solar absorbers: (a) copper tube receiver mounted in the CPC; (b) full optical assembly of the surface receiver; (c) MWCNT nanofluid receiver mounted in the CPC; (d) full optical assembly of the volumetric receiver; ...... 115

Figure 6.16 Vacuum glass chamber prototype design: (a) 3D view of absorber inside the vacuum glass tube; (b) cross-sectional view of vacuum flange with O-rings ...... 116

Figure 6.17 Pictures of the solar absorbers in vacuum chamber: (a) surface absorber inside the vacuum glass tube; (b) both vacuum packaged receivers, side-by-side ...... 117

Figure 7.1 View of the non-vacuum collector prototype undergoing testing ...... 121

Figure 7.2 Prototype with vacuum insulation ...... 121

Figure 7.3 Testing platform for the collector prototype: (a) Photo of the testing platform and (b) a flow diagram of the high temperature loop...... 123

Figure 7.4 Comparison of the experimental performance by the black chrome-coated copper tube and nanofluid receivers (a) tested without vacuum insulation (b) tested with vacuum insulation...... 126

Figure 7.5 Comparison between the experimental and CFD collector efficiency values for the black chrome-coated copper tube and MWCNT nanofluid receivers ...... 127

Figure 7.6 Instantaneous solar collector efficiency comparison ...... 130

Figure 7.7 Incidence angle modifier (IAM) outdoor test (a) prototype tested with longitudinal incidence angle of 40°; b) prototype tested with transverse incidence angle of 45° ...... 132

Figure 7.8 Transverse and longitudinal incident angle modifier ...... 132

Figure 8.1 ICS module design: (a) End view of the internal components (b) Cross-sectional view ...... 135

Figure 8.2 Ray tracing results for (a) 0˚ incident light angles and (b)45˚ incident light angles...... 137

Figure 8.3 Collector thermal efficiency with and without vacuum insulation ...... 138

xvii

Figure 8.4 Generalized cylindrical latent heat thermal storage system (a) 3D view of cylindrical tank with 5 tubes in shell; (b) cross section view of cylindrical tank with 1 tube (D0) and 7 tubes (D2) in shell ...... 140

Figure 8.5 Comparison of vacuum and non-vacuum collector performance for a typical day in January with D0: (a) collector thermal output, (b) available energy for storage 146

Figure 8.6 Transient performance of the system with 100 vacuum collectors over a typical day in January (LHTES with D0) ...... 147

Figure 8.7 Total amount of annual stored energy in various design alternatives for different number of collectors ...... 148

Figure 8.8 Solar fraction versus the number of collectors for different design alternatives ...... 149

Figure 8.9 LHTES fraction versus total number of collectors for various design alternatives ...... 150

Figure 8.10 Annual LHTES charging efficiency versus total number of collectors for various design alternatives ...... 151

Figure 8.11 Levelized cost of heat energy for various design configurations ...... 155

Figure 8.12 Levelized cost of heat energy and payback time for various design configurations and specific cost of ICS module ...... 156

Figure 9.1 Prism designed to redirect beam from right side of normal of prism normal by (a) 15 degrees and (b) 45 degrees ...... 159

Figure 9.2 Ray tracing results for (a) 0o incident light angle; (b-c) 45o and 52.5o incident light angle with the same prism array ...... 160

Figure 9.3 The transverse IAM of the proposed designs ...... 162

Figure 9.4 Collector design: (a) 3D end view; (b) transverse cross-sectional view..... 163

Figure 9.5 Collector efficiency over the year for different prism configurations as a function of operating temperature ...... 165

Figure 9.6 General layout of a solar-assisted absorption chiller system for air-conditioning applications [36] ...... 170

xviii

Figure 9.7 The load profile of the hotel building throughout a year in Sydney ...... 172

Figure 9.8 Solar fraction variation with specific storage tank volume ratio for different collector specific areas for: (a) SHC1 and (b) SHC2 layouts ...... 173

Figure 9.9 Comparison of the capital cost of the proposed SHC systems (employing both the current and improved collector designs) ...... 177

Figure 9.10 SHC system LCOC and LCOE for the current and improved collector designs using various specific capital costs ...... 178

xix

List of Tables

Table 1.1 Typical characteristics and cost comparison of solar thermal technologies and competitors [2-4, 10, 11] ...... 2

Table 2.1 Industrial sectors and processes with the greatest potential for solar thermal utilization [14, 19] ...... 7

Table 2.2 Summary of commercialized or practical demonstrated state-of-the-art solar collectors (100-400 °C) ...... 12

Table 2.3 Overview of the costs of typical solar thermal collectors [48] (~1,000/850 W/m2 of GHI/DNI were assumed in the calculation) ...... 12

Table 2.4 Summary the studies of representative refractive solar concentrators ...... 16

Table 2.5 Summary of studies on reflective solar concentrators...... 18

Table 2.6 Summary the studies of representative hybrid solar concentrators ...... 24

Table 2.7 Properties of mid-temperature selective surfaces [112, 113] ...... 28

Table 2.8 Summary of DASC studies ...... 30

Table 2.9 Summary of ICS with built in LHTES system studies ...... 32

Table 3.1 Comparison of two concentrator designs ...... 45

Table 4.1 Control software initialization parameters ...... 52

Table 4.2 Dynamic energy balance formulas of all the components [161] ...... 55

Table 4.3 Parameters of this collector ...... 56

Table 4.4 Heat transfer coefficient correlations [163]...... 57

Table 4.5 Materials and characteristics of the model ...... 59

Table 5.1 Thermal concentration parameters ...... 73

Table 5.2 Analysis results summary for passive thermal concentrators ...... 79

Table 5.3 Analysis results summary for active thermal concentrators ...... 93

Table 6.1 Black body-weighted emittance of materials as a function of surface temperature ...... 102 xx

Table 6.2 Details of the material and optical properties along with the geometric dimensions of the components used in this collector [81, 196, 207] ...... 104

Table 7.1 The coefficients of tested collector (with vacuum insulation) and other compared collectors ...... 129

Table 8.1 Physical parameters of the HTF and PCMs ...... 143

Table 8.2 Various design alternatives for LHTES unit in one solar collector ...... 144

Table 8.3 Economic assumptions ...... 152

Table 8.4 Prediction of capital cost (CC), annual energy produced for various configurations of proposed ICS ...... 153

Table 9.1 Details of the material and optical properties along with the geometric dimensions of the components used in this collector ...... 161

Table 9.2 Summary of comparison to our previous experimentally-validated collector design ...... 166

Table 9.4 Input parameters used for simulation of the solar -assisted heating and cooling system (SHC) ...... 171

Table 9.5 System sizing results ...... 174

Table 9.6 Economic analysis - assumptions and justifications ...... 176

xxi

Nomenclature

A Area (m2), Ambient

A Absorbance

Cp Specific heat capacity (J/kg K)

Cr Geometrical concentration ratio d Inner diameter (m)

D Outer diameter (m)

E Amount of stored/discharged energy [J]

F Fin efficiency, View factor

F′ Collector efficiency factor

G Global solar irradiation (W/m2) h Enthalpy (J/kg), Heat transfer coefficient (W/m2K)

K Incidence angle modifier k Thermal conductivity (W/m-K)

L Litre, Length (m)

m Mass flow rate (kg/s) n System economic lifetime (year)

N Number

NIR Near Infrared (0.7-2.5 um)

'' 2 q Flux density (W/m )

Q Heat transfer rate (W) R Reflectance, Thermal resistance (m2K/W) r Radius (m)

Rp Compressor pressure ratio

S Specific entropy (J/Kg K)

xxii

T Temperature (K), Transmission, t Time (s)

2 UL Collector heat loss coefficient (W/ m K)

UV Ultraviolet (0.01-0.4 um)

V Wind speed

Vis Visible (0.4-0.7 um)

W Width

W Supplied electric power (W)

ZT Figure of merit

Greek symbols

σ Stefan-Boltzman constant (W/m2K4)

α Absorptance, Seebeck coefficient of thermoelectric module (V/K)

τ Transmittance

ρ Reflectance, Density (kg/m3)

θ Incident angle (degree)

 Efficiency

λ Wavelength

 Dynamic viscosity (kg/m.s)

 Diameter (m)

Subscripts

a Ambient

ac Acceptance

aper Aperture

air Air

xxiii

abs Absorber b Beam radiation, Base

B Bond b_m Mean value of beam radiation bf Base fluid

Col Solar collector

Conv Convection comp compressor cond Condenser

Cu Copper d Diffuse d_m Mean value of diffuse radiation dh Directional-hemispherical e Electricity eff Effective ev Evaporation ex Exterior f Fluid g Glass tube, glass cover gl Global

H Hot side of thermoelectric module hp Heat pipe

HS Heat sink in Inlet loss Heat loss m Mean n Project life time (years)

xxiv

o Outlet

Opt Optical

p Electricity power (W)

P Pressure, Phase change material

r Receiver, Refrigerator

S Solid

Top Top of tank (Z/L=1)

Rad Radiation

sp Spreading

t Thickness

TIM Thermal interface material

th Thermal

u Useful

w Water

Z Axial direction

Abbreviations

AR Annual Revenue ($)

AC Annual Cost ($)

BCCCT Black chrome-coated copper tube

CFD Computational fluid dynamics

CC Capital cost ($)

CDA Carbon dioxide abatement

CPC Compound parabolic concentrator

CST Concentrated solar thermal

DNI Direct normal irradiance

xxv

EFC Evacuated flat plate collector

ETC Evacuated tube collector

FCS Fuel consumption savings

HTF Heat transfer fluid

ICS Integrated collector/storage

ICSSWH Integrated collector/storage solar water heater

LCOC Levelized cost of cooling energy

LCOE Levelised cost of energy

LCOH Levelized cost of heating energy

LHTES Latent heat thermal energy storage

MWCNT Multi-walled carbon nanotube

PCM Phase change material

PV Photovoltaics

PTC Parabolic trough collector

SAHP Solar-assisted heat pump

SHC Solar heating and cooling

STEG Solar thermoelectric generator

SWA Solar weighted absorbance

TEC Thermoelectric cooler

TES Thermal energy storage

xxvi

Chapter 1

1. Introduction

1.1 Research Background

Although the sun continuously creates – and provides free delivery on – an enormous amount of clean, renewable energy, less than 0.0005% of this resource is currently harvested for human use on Earth [1]. Due to increasing energy demand and environmental pressures along with technological improvements, solar energy utilization is rapidly expanding [1]. In this field, Photovoltaics (PV) may garner more recognition, but solar thermal technology still represents the majority (~70%) of the global installed capacity (410 GWth of 590 GW in total) [2]. Of the solar thermal market, domestic hot water collectors cover the lion’s share – with more than 289.5

GWth installed in China alone [2, 3]. For reference, the global installed capacity of PV was ~178 GWe at the end of 2014 [2, 3]. It is clear that domestic hot water heating (requiring temperature lower than 100 °C) currently dominates the solar industry – a fact which can be explained by their low levelised energy cost (~0.06 $/kWth-h) [2, 3], as given in Table 1.1. In many locations this is competitive (before subsidies) with conventional hot water systems (i.e. gas and electricity-based hot water ranges from

0.04-0.15 $/kWth-h) [2, 3].

However, in recent years, the average annual growth rate of solar water heating been slowing from ~14% to ~7% in 2014 [2, 3]. The meteoric growth of PV (~50% per year over past 10 years) may soon result in a change of solar leadership, particularly with declining solar water heating markets in Europe and China. A major factor in the recent success of solar PV has been its continued march down the manufacturing learning curve and the associated $/We (installed) price reductions [2, 3].

To compete with solar PV, it is clear that solar thermal technologies must follow a similar trajectory of technical and economic improvement. Additionally, solar thermal technology can also benefit from moving into new applications – e.g. moving beyond hot water [4]. At present, there is a growing interest to push towards higher

temperatures (100-400 °C) for industrial process heating applications and solar thermal-driven air-conditioning (double and triple effect absorption chillers) [5-9]. Medium temperature (100-300 °C) collectors can be used for industrial process heating, solar cooling, seawater desalination and other commercial applications. Solar thermal collectors which match the needs of these applications are under-developed, but there is a clear and present need to investigate and develop this technology and to make solar energy a viable option for these applications.

Table 1.1 Typical characteristics and cost comparison of solar thermal technologies and competitors [2-4, 10, 11] Technologies Typical Capital cost Typical energy cost characteristics (USD/W) (LCOE-US $/kWh) Conversion ~0.1(Ground-mounted Solar PV efficiency: 10–46% 1.3-2 utility-scale) (high end is CPV) ~0.25(Residential rooftop) Nature gas and oil ~100% / 0.04 (Gas)-0.09 (Oil) (heater) Electricity (heater) ~100% / 0.15 0.150–0.635 Solar thermal: (China) 0.02–0.1 (China) (Domestic hot water Efficiency: ~50-70% 1.1-2.14 0.05-0.5 Systems)

Temperature range: Solar thermal: 50–400 °C; 0.04-0.16 (low temperature (Industrial process 0.47-1 Efficiency: ~40-60% applications) heat) at 200 °C 0.17-0.37 (Trough and 7.1-9.8 (Trough, Fresnel, 6 hours storage); Concentrating solar 10-16% (Trough) 6 hours storage) 0.20–0.29 (Tower, 6–7 thermal power 7-20% (Tower) 6.3-10.5 (6–15 hours storage); (CSP) hours storage) 0.12–0.15 (Tower, 12–15 hours storage) Recent studies [1-4] have reported that about one half of energy demand is associated with fulfilling thermal requirements for buildings and industrial processes. As shown in Figure 1.1, an estimated 27% of total industrial heat demand is for the supply of 100-400 °C [4, 8, 12, 13]. This represents a vast, untapped potential market for high quality solar heat which is currently met by gas and electricity [12]. Chemical processing, kilning, drying, curing, sterilization, and distillation activities require higher temperatures. To achieve these temperatures, sunlight must either be absorbed inside a very well designed thermal package, or it must be concentrated to reduce the losses from the absorber.

Since heat – particularly high temperature heat – is hard to transport over large distances, rooftop heat production represents an inimitable opportunity for industries to exploit their unused real estate (large, flat rooftops) and convert volatile energy expenses into stable, internal capital investments [7]. Unfortunately, to date only a few collectors have been commercially developed which can fit the needs of the industrial process heat market.

Figure 1.1 Annual industrial heating energy consumption (in Europe 2005) [4] One barrier for concentrated solar systems is that they require tracking systems, which are relatively heavy, complex, and cumbersome to integrate with rooftops in comparison with PV racks [9]. Another challenge is that they need to compete with the relatively low cost of conventional systems (e.g. heat generated from gas or electricity). If these challenges can be overcome, solar thermal technology could tap into this huge commercial and industrial (behind the meter) market for high quality heat [7, 8].

1.2 Objectives

This aim of this research is to investigate and develop new and cost-effective ways to obtain medium-temperature solar-derived heat to meet the needs of industrial processes

and commercial/residential buildings. As noted above, this aim necessitates working within several constraints, including being: thin, reliable, efficient, low-cost, lightweight, aesthetic, and easily integrated with rooftops. To be competitive in this market, the following performance characteristics are required:

 A collector which can supply 100-250 °C thermal energy with an efficiency above 45%;

 A stationary collector platform which does not require external moving parts which can mount on standard PV racks;

 A system that can providing heat energy with relatively low levelised cost of

energy (e.g. < 0.25 $/kWthh );

To achieve these objectives, both innovative optical and thermal concentration mechanisms were explored as innovative means to convert the relatively low density incoming solar radiation into heat in a low-profile package. A mix of theory, modelling and experiments was used to systematically assess and optimize for this aim among a wide range of potential materials and configurations.

1.3 Thesis structure

This thesis is organized into ten chapters. Following this introductory chapter, chapter 2 provides a review of the recent development and innovative technologies available in the literature for low profile solar thermal collectors, storage integrated systems, and work on incorporating these into industrial applications. Chapter 3 presents the design and analysis of an innovative optical concentration system. Chapter 4 gives a detailed design of the solar thermal collector developed in this work. A theoretical and numerical analysis of how thermal concentration could be beneficially employed in this type of technology is reported in chapter 5. In chapter 6, a new volumetric solar receiver is design and analyzed. Chapter 7 presents an experimental analysis of a full-scale prototype. In chapter 8, an innovative integrated latent heat storage collector is presented and investigated. The techno-economic analysis of this collector for air-conditioning and process heating applications is also reported in chapter 9. Finally, the main research findings and contributions are concluded in chapter 10.

Chapter 2

2. Literature Review

As stated in chapter 1, solar thermal systems can fulfill a substantial amount of heat demand (100-400 °C) in industrial and commercial applications [4-6]. If well-designed, new technologies are developed for this purpose it could open a huge commercial and industrial market for high quality heat (to offset volatile gas prices), while also minimizing our ecological footprint (e.g. land-use) [14, 15]. In this chapter, a comprehensive literature review on medium temperature solar thermal collectors and systems is presented. The review particularly focuses on the solar thermal technologies published in academic literature and on what is available on company websites.

As the motivation for developing these technologies, the review begins with medium temperature solar thermal applications, followed by the improvements and developments in the field of medium temperature solar thermal collectors over the last ~10 years. Different types of solar concentrators and absorbers are examined, and a critical comparison and assessment is carried out among different technologies/materials based on the performance and feasibility of what is currently available. In addition, to ensure these technologies can manage the transient solar resource for these applications, a brief overview of thermal storage technologies and integrated collector/storage technologies is also presented.

By reviewing currently available technologies/ideas and their technical advantages/ challenges, the research gaps for developing medium temperatures solar collector and system are identified.

2.1 Medium temperature solar thermal applications

Beyond the low-temperature applications, medium temperature level (100-250 °C) solar thermal applications have received interest in the recent years in both residential/commercial and industrial sectors [14, 16]. The review in this section will be concentrated on the most recent studies on the largest potential applications where solar

thermal energy could have an impact – industrial heating and thermal-driven air-conditioning.

2.1.1 Solar derived industrial/commercial heating applications

Industrial process heat represents a large piece of the global energy pie. About 13% of industrial heating applications require low temperatures (below 100 °C), 27% require 100-200 °C, and the remaining applications require high temperature (e.g. steel, glass, and ceramic industries) [17]. Most industrial processes require both heating of a fluid stream (e.g. hot air streams, hot water) and heating of stationary reservoirs (e.g. ovens, liquid baths). Existing heating systems for industrial processes are based on steam or hot water from a gas, coal, and other non-renewable fuels –fired [7, 9, 18]. Integration of solar thermal systems into conventional industrial processes is a feasible way of off-setting some of this fuel consumption to create more sustainable, cleaner heat for industry.

Several studies on the use of solar heat for industrial processes (SHIP) have been carried for different countries or regions [14, 19-21]. From a number of studies on industrial heat-demand, several industrial sectors (e.g. food and beverages, chemicals, textile, plastic and paper) have been identified to have favorable conditions for the implementation of solar energy [19]. The most important industrial processes using heat at a medium temperature level are: sterilizing, pasteurizing, drying, hydrolyzing, distillation and evaporation, washing and cleaning, and polymerization [19, 22]. Some of the most important processes and the related studies for each are outlined in Table 2.1. Solar thermal collectors can supply up to 100% of heating demand (called the solar fraction) for many of these applications, according to practical and theoretical operational studies [16, 21].

Despite the large potential for solar energy to meet industrial thermal demand, there are several barriers to large-scale deployment. The most noteworthy barriers are cost, variability of output, energy storage, and process integration. Integration of solar heat systems into industrial applications requires storage and control strategies to handle the non-continuous supply of solar energy. The accurate design and sizing of a solar thermal system (e.g. collector storage size) should take into account the local weather

conditions and the specific demand profile. This can be carried out by dynamic system simulation software packages, such as TRNSYS.

Table 2.1 Industrial sectors and processes with the greatest potential for solar thermal utilization [14, 19]

Main applications Author or Characteristics (e.g. Performance Industrial (Temperature company Demand, installed (Technical and sector level °C) (year) capacity) economic)

Daily demand: 36 MWh Reduces 120,000 Lacteas per day for milk drying m3 gas Cobreros (180 °C) and other consumption and S.A. processes; Installed 290 tons CO Boiler feed water (2012) 2 capacity:1 MWth (2,040 emission per Food, Dairy (60-90); Drying [20] m2 PTC collector area); year and (60-180), Beverage Sterilisation Daily demand: 8000 L Industries (140-150) Moguntia hot water (~60 °C); Meat Installed capacity:150 Solar fraction: Spices kWth (215 m2 FPC 45% (2007) collector area); Storage [21] volume: 10 m3

Annual system Installed capacity: Washing (60-90), efficiency: nominal collector power Bleaching Frey et al. 27.8%; Textile of 70 kWth (60-100); Drying (2015)[23] (PTC);Storage volume: Primary energy (100-160) 10-70 m3 savings :22.8 MWh per year

Bleaching Numerical study; Total (120-150) demand in India: 25 Sharma et PJ/year (50-250 °C); Solar fraction: Paper drying (90-200); al. (2015) Boiler feed water Assuming Installing 25-30% [18] 2 (60–90); Bleaching capacity: 1.11 million m (130–150) of FPC collector area

Electroplating Installed capacity (in Metal and (50-60); Chile): 26 MWth Solar fraction: mining Electrolyte (50); 39,300 m² Flat plate 85-100% industry Cleaning/drying collector; 4,300 m³ processes (50-160) Storage

However, based on the available literature, few studies investigate anything other than low-temperature solar heat integration (<100 °C). Additionally, it was found that to date no techno-economic investigation has been carried out for medium temperature solar-assisted process heating systems. Thus, there is a need to assess the system

techno-economic feasibility to promote medium temperature solar systems in industrial process heating applications.

2.1.2 Solar-derived thermal-driven cooling applications

Besides industrial process heating applications, solar thermal systems can also drive air-conditioning applications. This is a very logical choice since the cooling demand of buildings is highly temporally correlated to the availability of solar flux. The east coast of Australia represents a particularly attractive location for solar air-conditioning systems since it receives between six to nine hours of sunshine a day, with an annual solar exposure of 1,200-2,400 kWh/m2, which is more than sufficient for solar air-conditioning applications [24].

The available technologies on the market for thermally driven cooling systems are absorption and adsorption chillers, solid and liquid desiccant cooling systems, and ejector refrigeration cycles [25]. Of these, absorption chillers are considered the most desirable method for harnessing solar thermal energy due to their commercial readiness, reliability, and relatively high coefficient of performance. In addition, absorption chillers are already commercially available for large-scale applications (traditionally gas-fired), so their mass-produced cost is lower than other emerging thermally-driven air-conditioning systems [26, 27]. Absorption chiller systems can be retrofitted to utilize solar thermal energy as the driving heat source. Since much of the technology has already been proven at commercial scales, solar cooling via absorption chillers represents a promising (and potentially sustainable) near-term alternative to conventional air-conditioning systems [28, 29].

There are three types of absorption chillers commercially available on the market − single-, double-, and triple-effect chillers. The advantage of moving towards a higher effect cycles is to enhance the COP of the chiller, assuming a high temperature heat source is available. The most common working fluid pair used in absorption chillers for air-conditioning applications is lithium bromide-water (LiBr-H2O), where LiBr is the absorbent and water is the refrigerant [26]. The driving heat source temperature for single-effect chillers is about 80-100 °C, while their COP is limited to around 0.7 [30]. Double- and triple-effect chillers, on the other hand, require driving temperatures of

around 180 and 240 °C, respectively, and can reach COPs of up to 1.2 and 1.8, respectively [30].

The majority of solar absorption chillers installed around the world are based on single-effect chillers and low-temperature solar thermal flat plate or evacuated tube collectors (FPCs and ETCs) [31, 32]. This configuration is usually considered as the most promising design in European climates [33]. The main drawback of solar single-effect chillers is the low COP of the chiller, requiring a large collector area to provide the thermal energy demand. In addition to the cost of large collector areas, this may be a significant limiting factor for the use of such systems in buildings with limited available rooftop area.

The combination of high-temperature solar thermal collectors and multi-effect absorption chillers is becoming more attractive due to their higher COP compared to single-effect chillers [34, 35]. This means that multi-effect chillers require less solar thermal energy (and potentially less collector area) to supply a given amount of cooling. However, they also require high driving temperatures which can only be achieved by concentrating solar thermal collectors. Numerous studies have been conducted in the field of solar assisted absorption chillers [36]. Al-lili et al. [37] conducted a thermal and economic analysis of solar driven absorption cycles in Abu Dhabi. The results showed that ~50% of solar fraction was achieved and the collector area was found to be the key parameter influencing payback period of the initial investment. Shirazi et al. [5, 6, 36] investigated the feasibility of solar heating and cooling (SHC) absorption systems based on combining three types of LiBr–H2O absorption chillers (single-, double-, and triple-effect) with commercially available solar thermal collectors (including a linear rooftop concentrating solar thermal collector, manufactured by Chromasun Inc. and a typical PTC). The studies revealed that solar driven absorption systems are not economically profitable under current market prices due to high capital cost of solar collectors and chillers. To achieve a satisfactory payback period (<5 years) and solar fraction (>45%), around 50% of the capital cost reduction (or financed by government subsidies) of solar collectors and chillers should be required [6]. The results also indicated that current commercially available concentrating solar thermal collectors are less feasible compared with evacuated flat plate collector (EFC) due to their higher capital cost and lower annual efficiency. This could mainly be explained by the fact that

concentrating systems are more complex and only beam radiation can be harvested by concentrating collectors [36].

Moreover, the building integration of concentration technologies is difficult for rooftop applications because they are usually quite large, have high wind loading, and do not aesthetically fit into the building’s architecture [33, 38]. Therefore, research on new solar collectors for building integration that produce higher temperatures and exhibit high solar collection efficiency is an important research topic.

2.2 Solar thermal collectors: The state-of-the-art

2.2.1 Solar thermal collector overview

Since the sun moves in an arc throughout the day, concentration and tracking plays a critical role in medium temperature solar thermal collectors. Most of the recent advances in medium temperature solar collectors involve system simplification and performance enhancement of the concentration and/or tracking system. A review of approaches to track and concentrate sunlight and a survey of the existing designs was conducted for this work. In fact, concentrating solar collectors can be classified by their tracking technologies, as is shown in Figure 2.1.

Sun Tracking

Active Semi-active Non- (full (tracking tracking motion) integrated) or passive

Refractive Liquid Internal Beam or Catadioptic EFP/ETC CPC prism motion steering Reflective array

Concentrator Prism+ Reflector or receiver +lens motion Lens

Figure 2.1 Categorization of tracking approaches for concentrating solar collectors There are a few collectors which have been commercially developed that can fit the needs of the industrial process heat market. To attain higher temperatures (>100 °C) for commercial and industrial applications, optical concentrators and/or vacuum insulation

are usually required. Due to their relatively high concentration ratio, parabolic troughs, including the PolyTrough 1800 solar collector [39], the SOLITEM PTC1800 [40], and the SopoNovaTM MicroCSP [41] can deliver heat at a temperature of 200 °C with between 53-57% efficiency with a DNI of 800-850 W/m2 [39, 42]. However, the solar tracking mechanism in these collectors is achieved by rotating the whole bulky, heavy module to maintain normal light incidence (classified as active tracking). As a result, trackers significantly add capital cost (a premium cost of ~$0.36 W-1) and annual maintenance costs (~2% of tracker cost) [43]. Moreover, they are also very hard to integrate on industrial buildings’ rooftops. These barriers limit the commercial uptake of solar thermal energy in the industrial heating market.

To reduce the complexity and cost of collectors, some attention has recently been devoted to the development of ‘stationary’ rooftop solar thermal systems, which could be categorized as: i) non-tracking collector, such as evacuated flat plate (EFP) and evacuated tube collector) (ETC); ii) passive tracking collectors (tracking can be functionalized by optics components, e.g. CPC); or iii) semi-active tracking collectors (tracking integrated/performed inside the module, e.g. rotating internal reflectors or moving the receivers). As summarized in

Table 2.2, TVP Solar (a Swiss company who developed a high vacuum, non-tracking flat plate solar thermal panel) has recently released results from a solar air cooling pilot installation in Abu Dhabi. The results demonstrated that their flat panels successfully reach the 180 °C with average daily system efficiency (net of piping losses, panels were stationary) of around 45% required for the double effect absorption chiller [36]. A rooftop solar, semi-active tracking concentrating collector, developed by Chromasun, can deliver heat at temperatures of up to 200 °C at 44% efficiency (when the DNI is 850 W/m2) [44]. A passive tracking External Compound Parabolic Concentrator (XCPC) collector developed at the University of California, Merced uses passive tracking to achieve a 200 °C fluid temperature with an efficiency of 36–40%, at DNI of 800 W/m2 [45].

Thus, low profile rooftop ‘stationary’ collectors (with/without internal tracking where the module itself remains fixed) seem to represent a promising and cost-effective approach to meet the needs of commercial and industrial applications.

Table 2.2 Summary of commercialized or practical demonstrated state-of-the-art solar collectors (100-400 °C) Collector Author/company Collector type Performance categories (year) ~50% at 200 °C Evacuated Flat Stationary TVP Solar (Global irradiance of Plate(EFPC) 1,000 W/m2) [46] External Compound Winston et al.(2014) 40% 200 °C (DNI of Passive-Tracking Parabolic Concentrator [45] 800W/m2) (XCPC) Linear Fresnel 44% at 200 °C (DNI of Semi-active Chromasun [44] reflector 850 W/m2) tracking Stationary spherical (internal Cohen and Grossman ~50% at 200 °C (DNI reflector + tracking tracking) [47] of 950 W/m2) absorber Small size PTC NEP Solar Inc. (PolyTrough 1800) Small size PTC 53-57% at 200 °C Full-active Solitem Inc. (SOLITEM PTC1800) (DNI of 800-850 tracking Small size PTC W/m2) [39, 42] Sopogy Inc. (SopoNovaTM MicroCSP) 2.2.2 Performance and cost

Table 2.3 provides an overview of the costs of selected solar collectors from those mentioned above. Since evacuated tube collectors have been mass produced in China, they currently have a relatively low cost, around 130 $/m2. However, this type of collector was designed for domestic water heating and low temperature process heating applications. The main technologies used for medium temperature heat, concentrating solar collectors (400-650 US $/ m2), are more expensive than conventional ETC, but permit higher temperature ranges.

Table 2.3 Overview of the costs of typical solar thermal collectors [48] (~1,000/850 W/m2 of GHI/DNI were assumed in the calculation) Specific Collector Cost Cost thermal power Location categories (US $/ m2) (US $/ KW ) (KW/m2) th China 130 200-220 Evacuated tube 0.6-0.65 Europe 450-900 690-1,500 collector (ETC) 0.3 India 333 1,133 Parabolic 0.5-0.56 Europe 650 1,160-1,300 trough collector 0.22-0.28 India 445 1,580-2,040 (PTC) 0.55-0.7 Mexico 400-629 570-1,100 Linear Fresnel 0.5-0.56 Europe 650-900 1,160-1,800

In the future, solar thermal collectors can be made even more cost-effective if tailored to the specific process heating needs. The IEA (2009) suggests that costs can be reduced by as much as 20% when a country’s total installed capacity doubles [49]. Furthermore, large-scale applications can benefit from the economies of scale and lowered investment costs, thus increasing the project’s economic viability [22].

Recent assessments in Europe suggest cost reductions of ~40% by 2020, and an industry roadmap for Europe targets a system price (inclusive of storage) of ~400-500 $/KWth (200-250 US $/ m2) for low-medium temperature collectors (<250 °C) [22, 48]. Key factors for cost reductions are automation of production processes, modular designs for easier installation and integration onto industrial roofs, optimized tracking systems, standardization and certification, and material replacement of copper and steel with aluminum and polymers [21].

In order to develop a cost-effective solar thermal collector, the review particularly focuses on the solar thermal technologies which are actively investigated by researchers, including optical/thermal concentration, solar absorbers, storage, and integrated collector/storage systems.

2.3 Optical concentration

A solar collector which uses reflectors, lenses or other optical elements to concentrate the radiant energy passing through the aperture onto an absorber which has a surface area smaller than the aperture. Heat production and electricity generation comprise the main current applications for optical solar concentration. Generally, solar concentration seeks to achieve cost reduction and/or efficiency enhancement. Cost reductions can be achieved because expensive receivers are replaced by inexpensive optics. Efficiency benefits result (in solar thermal systems) from significantly reducing heat-loss area. PV-conversion efficiency can increase linearly with the logarithm of the concentration at low series resistance [50].

High flux concentration ratios, high optical efficiency and high manufacture tolerance are the key features required for a successful solar concentrator design [51]. The concentration ratio (Cr=1/sin 휃, where 휃 is the half acceptance angle) of concentrators can be classified as: low for Cr≤10, medium for 10100.

Kalogirou (2012) [52] revealed that the value of maximum optical concentration ratio for a single-axis tracking and full tracking are 216 and 46,747 [50]. Although it is never needed or even approached, the temperature limit for a solar thermal system is the effective surface temperature of the sun.

Figure 2.2 summarizes different optical concentration technologies based on the literature review. The higher concentration domain is still dominated by parabolic and Fresnel mirrors and Fresnel lenses, although a variety of advanced non-imaging concentrators have been developed, such as hybrid (combining refractive and reflective components [53]) concentrators and/or multi-stage concentrators.

Optical Concentrators

Refractive Reflective Hybrid

Fresnel Composite Multi- lens Parabolic reflector Optics stage

Gradient Parabolic Light Fibre Fresnel Liquid CPC index trough guide optic

Figure 2.2 Categorization of energy concentration techniques

2.3.1 Refractive concentrators

As categorized in Table 2.4, studies on Fresnel lens (linear/point focus), liquid lens/prism, and gradient-index lenses are reviewed in this section.

Fresnel lenses are typically produced from plastics, e.g. Polymethylmethacrylate (PMMA), and first became available in the 1950s [54]. Based on the relatively low cost of Fresnel lenses, they have huge potential to reduce solar concentrator costs for concentrated solar thermal (CST) and photovoltaic (CPV) systems [54]. Fresnel collectors can be sub-classified into: a) line-focus systems and b) point-focus systems. In order to get high concentration ratios and decrease their focal length, convex shapes

can be employed to effectively increase their concentration ratio [55]. Khalil and Munadhila [56] studied a solar collector which consisted of two flat linear Fresnel lens and absorbers. Each Fresnel lens was 0.37 meters in width and tracked the sun in two directions. The maximum daily thermal efficiency reached 0.58. According to an analytical and experimental study by Franc et al. [57], efficiencies of between 37% - 51% were achievable with line-focusing acrylic Fresnel lenses when the outlet temperature was 100 °C. Their study also examined collector performance versus different types of insulation technologies. A Fresnel lens collector with an evacuated absorber tube was developed and tested by Zhai et al. [58]. This system has an efficiency of ~50% when the beam radiation was 800 W/m2 and the working temperature approached 200 °C. The thermal performance of a line-focus Fresnel lens solar collector using different cavity receivers has been investigated by Xie et al. (2012) [59]. They found that a triangular cavity receiver had the best thermal performance. The highest experimental heat removal factor is about 0.80 when the operation temperature was 180 °C [59]. A point-focus solar collector using a high concentration imaging PMMA Fresnel lens prototype was proposed and tested by Xie et al. (2011) [60]. Different cavity receivers with synthetic heat transfer oil flow were simulated and analyzed. The results indicated that the optimum receiver shape (conical cavity receiver) processed higher thermal efficiency and lower heat loss than others. The thermal efficiency could reach 50% when inlet fluid temperature was 150 °C. A photovoltaic-thermal module for a Fresnel linear concentrator [61] was developed by Chemisana et al. (2011). This optical concentrator had a domed linear Fresnel lens as the primary concentrator and a compound parabolic reflector as the secondary concentrator. Overall, the Fresnel lens collector with good insulation represented a solution for supplying a high temperature (100-370 °C) application, and a high efficiency (50%) was achievable when the collector output was around 200 °C. However, the focal lengths of these tested systems were about 50 cm to 80 cm, which means they are (arguably) too large and unwieldy for rooftops. Therefore, the major design aspect for Fresnel lens collectors is the maximization of concentration ratio while minimizing the size and shape.

Optofluidic techniques can be used for guiding and collecting light in solar energy applications. For instance, liquid lenses provide not only a much larger focal range and

shorter focal depth than traditional lenses, but also can provide dual-axis tracking (daily and seasonal changes of the Sun’s orbit) without inefficient and costly mechanical moving parts [62]. Liu (2012) [63] developed a liquid lens (prism). By manipulating the lens’ shape creatively, the lens could adaptively track the seasonal and daily changes of the incident radiation. Two transparent cubic containers were filled with two immiscible conducting liquids, the liquids are formed as sloped to a desired shape when different voltages were applied. For the design by Liu, the sunlight could be refracted 4 times when it passed through the prototype. In theory, their device could achieve beam tracking and steering within an angle of -29 deg to +29 deg [63]. However, their experiment showed that the device can track and steer the light only up to 19.9 deg. Some challenges were identified in their research. One problem was that the light beam was blocked by the edge of the sidewall. Another problem was that the liquid-liquid interface was not flat when voltage was applied to the device, which may have caused an optical loss.

Table 2.4 Summary the studies of representative refractive solar concentrators Feature Performance Author Categories W or D f θ η and other remarks (year) Cr η ac th (m) (m) opt (o) Zhai et 50% (I =800 W/m2, 0.32 ~0.4 20 / / b Fresnel lens al.(2010) [58] To=200 °C) (line- focus) Franc, et 37-51% (I =950 W/m2, 0.37 0.5 9.6 / / g al.(1985) [57] To=100 °C) 0.85- Xie et al. 0.86*0.8 ~28 48% (I =850W/m2 , ~1.0 0.9 / b Fresnel lens (2011) [60] 6 0 T =150 ℃) (lens) o (point- Yeh N. and focus) ~30 Curve-based, Yeh P.(2016) 0.23 0.28 / / 0 point-focused Fresnel lens [64] Electrowetting for wide Liu, et al. Liquid / / / / +/-29 angle beam tracking and (2012) [63] lens/prism steering (Opto- Droplet manipulation to Tsou et al. fluidic) / / 9 / +/-25 concentrate and track the (2013) [65] sun Spherical gradient-index Gradient-in Kotsidas et al. ~30 0.99 1.09 0.97 / lenses was analysed dex lenses (2011) [66] ,000 theoretically

Note that: W, D, f, Cr, , θac and represent width, diameter, focus length, concentration ratio, optical efficiency, acceptance angle and thermal efficiency of optical concentrator, respectively. Droplet manipulation on a liquid crystal and polymer composite film (LCPCF) as a concentrator and tracker was proposed by Tsou et al. (2013) [65] for a concentrating PV system. The change of curvature of the liquid lens resulted in a tunable concentration

ratio (e.g. lower concentration ratio under strong illumination and higher concentration ratio under weak illumination), whilst the change of position of the liquid lens helped to track the sun up to 25 deg.

Conventional approaches to lens design have been based on the geometric optics of how light refracts at the interface between two materials. However, this solar concentration may be improved with the aid of metamaterials – custom designed materials which can force light to bend in almost any manner by spatially changing the structure and materials of the optical medium [67, 68].

For instance, gradient-index lenses (extremely high concentration ratio) would markedly simplify and downsize solar trackers. Spherical gradient-index lenses was analysed theoretically by Kotsidas et al. (2011) [66], who found that the concentration properties (~30,000 concentration ratio is achievable) could approach fundamental limits. However, the development of accurate and affordable fabrication procedures remains very challenging. Kundtz (2010) [69] proposed an extreme-angle broadband metamaterial lens. It was demonstrated that this Gradient Index (GRIN) lens had a field-of-view approaching 180° as the index of refraction was varied throughout the body of the lens. The design and analysis process of GRIN lenses were carried out and the ray tracing results showed that rays at extreme angles could be focused onto the image plane. Although GRIN lenses have significant potential advantages over conventional lenses, they are far less prevalent in practical applications because of difficult fabrication issues [68].

2.3.2 Reflective concentrators

Reflective solar concentrators use smooth surfaces with high specular reflectance across the solar spectrum. The reflecting surfaces are usually highly polished metals or metal coatings. Reflective concentrating technologies exist in four common forms, namely parabolic troughs, parabolic dish concentrators, concentrating linear Fresnel reflectors, and heliostat mirrors. Linear Fresnel Reflector (LFR), parabolic trough concentrating (PTC) and compound parabolic concentrating (CPC) technologies are based on linear reflective solar concentration. Compared with heliostats and parabolic dishes, these type of concentrators are easily integrated into rooftop applications [70]. Studies on representative reflective solar concentrators are summarized in Table 2.5.

Table 2.5 Summary of studies on reflective solar concentrators Feature Performance Authors η and other Categories W th (year) f or H (m) Cr η θ (o) (m) opt ac remarks Chromasun [44, 71]; Vivar Linear 0.5 0.248 20 68% ~70 48% (T =200 ℃) at el. (2012) o Fresnel [72] Reflector Lin et al. (2013) / 1.5 / 0.76 / 37% (T =150 ℃) [73] o NEP Solar PolyTrough 1.2 0.65 15 / / 58% (To=160℃) 1200 [74] [75] Parabolic IST Corp. 70% (combined 1.15 / 14.4 0.76 / Trough RMT (roof) [76] efficiency) Solitem PTC 1800 [75, 1.8 0.78 40 / 50% (To=180℃) 77] AoSol CPC [78] 0.07 and 35%-40% and Solarfocus / 0.10 1.7 / (T =100°C) CPC [44] (Height) o ESE 0.137 T = 80 °C to o VACOSOL [75] (Height) 170°C 50% (I = 1000 Buttinger et al. 0.07 0.12 1.8 0.82 +/-35 W/m2, T =150 CPC (2011) [79] (Height) o ℃) 0.35 0.32 and +/-20 Li et al. (2013) 3 and 40-46% and 0.89 and o [80] 6 (T =200 C ℃) 0.7 (Height) +/-10 o Gu et al. (2013) 0.06 30% (T =250 0.06 1.75 0.78 +/-27.5 o [81] (Height) ℃) Gudekar et al. 0.24 25% (T =120 0.96 6.3 0.63 +/-6 o (2013) [82] (Height) ℃)

Note that: W, f, Cr, ηopt , θac and represent width, focus length, concentration ratio, optical efficiency, acceptance angle and thermal efficiency of optical concentrators, respectively. LFR collectors rely on an array of linear mirrors which concentrate light on a receiver [80, 83]. Flat mirrors allow a more reflective surface in the same amount of space as a parabolic reflector, thus capturing more of the available sunlight. In addition, they are potentially much cheaper than parabolic reflectors. Zhu at el.(2013) [80] reviewed the development history and outlook of different types of linear Fresnel collectors. They pointed out that linear Fresnel collectors can be easily tailored for different target temperatures to meet varying application needs due to their flexibility in the selection of concentration ratio.

A rooftop solar concentrating collector, developed by Chromasun, is suitable for roof-top applications as it has a relatively thin profile (about 30 cm), and all moving parts are contained within a glass envelope. This semi-active tracking collector can deliver heat at temperatures of up to 200 °C with 48% efficiency under solar radiation of 1,000 W/m2 [44, 71]. This collector is able to supply high temperature for industrial and commercial use due to the relatively high efficiency. The hybrid CPV-T ANU-Chromasun micro-concentrator has been developed and tested by Vivar at el. (2012) [72]. The results showed that the combined efficiency if the system could exceed 70%. Over the span of a day, the average electrical efficiency was 8% and the average thermal efficiency was 60%. A LFR solar collector was investigated both experimentally and theoretically by Lin et al. (2013) [73] which consisted of 6 pieces of rectangle mirror elements of dimension 1m× 0.3m× 1.7mm, and a cavity receiver mounted 1.5 m above the mirror surface. The test result showed that the thermal efficiency decreased from 45% to 37% as the average surface temperature increased from 90-150 °C. There are some typical medium -scale linear Fresnel collectors, examples including the prototypes developed by Morin et al. (2006) [84], Bernhard et al. (2008) [85], and Singh et al. (1999) [86]. As was concluded by N. Said et al.(2011) [70], concentration ratios of 25-100 and temperature outputs of 250-500 °C can be achieved by LFR solar collectors.

There are many factors that affect the optical efficiency of the LFR collector, including incidence angle of light, shading and blocking between adjacent mirrors, the offset of the facula, the optical properties of different surfaces, etc. Among them, the influence of geometry factors is most evident – e.g. the mirror’s width and the interval between two mirrors next to each other should be determined carefully. Analytic optical design of linear Fresnel collectors with variable widths and shifts of mirrors have conducted by Abbas and Martínez-Val (2014) [73, 87].

As another approach, a Non-Imaging Reflective Lens (NIRL) concentrator is a device that concentrates radiation towards a static receiver by means of an array of reflectors which rotate collectively [88]. Chemisana et al. (2013) designed a transmissive Fresnel reflector solar concentrating system, integrated on a building facade to drive a double-effect absorption chiller. The results showed that the collector could deliver 150

℃ to drive double-effect absorption chillers. However, the solar concentrating system required a large aperture area [33].

PTC (parabolic trough solar collector) focuses direct solar radiation reflected by the parabolic reflector onto a receiver located on its focal line. As a class of concentrators, these are far and away the most well developed and have been used in commercial Concentrated Solar Power (CSP) plants. PTCs typically achieve geometrical concentrating ratios between 20 and 30 and temperatures between 300 to 400 °C. PTCs have also been applied to industrial process heating and heat-driven refrigeration and cooling applications which require geometrical concentration ratios between 15 and 20, and temperatures between 100 and 250 °C [76]. The PolyTrough 1200 solar collector, developed by NEP Solar, is a small roof and ground mountable parabolic trough collector (aperture of trough: 1.20 m, Focal length: 0.65 m, Geometric Concentration Ratio: 45) which can deliver heat at a temperature of 160 °C with 58% efficiency under a direct solar radiation of 800 W/m2 [74] [75]. The Roof Mounted Parabolic Trough (RMT), developed by Abengoa Solar has an aperture width of 1.148 and a geometric concentration ratio of 14.4. The peak optical efficiency of the RMT can reach 76%, and the maximum operating temperature is 205 °C [76]. SOLITEM GmbH has produced its PTC-1800 for solar process steam generation and high efficiency solar chilling. This collector module (Length: 5,090 mm, Width: 1,800 mm, Height: 260 mm, Focal length: 780 mm) was designed to operate up to 200 °C and a 50% thermal efficiency is achievable with an outlet temperature up to 180 °C [75, 77]. The PTC 1000 modular parabolic trough collector was developed by Solar-Institut Jülich (SIJ) and Solitem GmbH [75]. An anti-reflective solar-glass was employed in this small single axis-tracking parabolic trough collector. It can operate in the range of 120°C to 200°C with high efficiency. The collector has an efficiency of around 60% at a DNI of 800 W/m² and a temperature of 160 °C. Stagnation experiments have shown temperatures around 590 °C for this collector [75, 76].

A combined heat and power solar (CHAPS) collector was developed by The Australian National University with a width of 1.55 m and a focal length of 0.85 m. The testing results demonstrated that a thermal efficiency of around 58% and an electrical efficiency of around 11%, and therefore a combined efficiency of 69% under typical operating conditions (fluid inlet temperature of 65 °C, ambient temperature of 25 °C

and direct radiation of 1,000 W/m²) [89]. Therefore the system is ideally suited to lower temperature applications (< 80 °C) and the electrical system efficiency can be maintained above 10%. The ANU team also showed that temperatures up to around 150 °C were feasible, with an electrical efficiency still above 8% [75]. Smeltink, J and Blakers, A (2007) [73] developed a micro-concentrator collector which had dimensions of 1.7 m x 1.5 m x 0.2 m. It consisted of seven parabolic polished aluminum mirrors (0.15 m wide by 1.25 m long) that focus light onto the receiver tubes. A detailed theoretical and testing analysis results showed that it could produce both solar hot water and solar electricity with a combined efficiency of at least 60%. However, the parabolic trough technology still has some room for improvement in its design, manufacture and operation. Firstly, the solar tracking systems in these collectors are operated by a mechanical tracker that rotates the whole module to maintain normal light incidence. Secondly, the support frame and post of a parabolic trough should be strong enough to support both the weight of mirror reflectors and the wind force. Lastly, to remove dust or snow, the trough must be periodically turned downwards, consuming power.

In order to overcome the general drawbacks of a parabolic trough, many innovative concentrator designs have been proposed which use either Fresnel lenses or reflectors. Another alternative is non-imaging concentrators, such as the Compound Parabolic Concentrator (CPC), which was originally developed by Winston and Hinterberger (1975) [45, 90]. While standard parabolic reflectors have a very small acceptance angle, CPCs have comparatively a large acceptance angle. By using multiple internal reflections, any radiation that is entering the aperture within this wide acceptance angle finds its way to the absorber surface located at the bottom of the collector. Having a large acceptance angle represents a partial solution for sunlight tracking, since a rooftop CPC collector can be used as a static solar concentrator. Of course, the optical efficiency drops as the sun moves away from the normal position, but a typical CPC can maintain 80% to 95% optical efficiency for angles within 27 degrees (or almost 4 hours per day) [91].

Due to geometric restrictions, rooftop CPC collectors typically only achieve a very low concentration ratio (1

processes heat if they were cost effective compared to parabolic trough or other medium temperature collectors. In order to increase the performance and the working temperature of the collector, several companies and researchers have been seeking improvements. For example, both the AoSol standard CPC collector [78] and the Solarfocus CPC collector [44] were designed to operate in the range of 80-120 °C. These can supply heat with a thermal efficiency of 35%-40% at ~100 °C. The relatively low efficiency is due to the fact that these designs eschew vacuum insulation in order to keep costs down. Another innovation with huge potential to increase the performance is the flat evacuated CPC-collector. This class of technology takes advantage of the fact that when the air pressure is maintained below 1.0 × 10-2 N/m, both conduction and convection heat losses are negligible, even for medium temperature operation [79]. As an example of this technology, the evacuated tubular collector, ESE VACOSOL [75], consists of 7 twin-glass tubes designed to operate in the range of 80°C to 170°C, and has a stagnation temperature is 244°C. Buttinger et al. (2011) [79] has reported a flat evacuated CPC-collector with an aperture area of 2.0 m2 and an efficiency of about 50% at a temperature of 150 °C for radiation of 1,000 W/m2. By combining vacuum insulation, spectrally selective coatings, and non-imaging concentration, Snail et al. (1984) [92] developed the Integrated Stationary Evacuated Concentrator, or ISEC, collector. According to their experimental testing, it had an optical efficiency of 65% and achieved higher than 50% thermal efficiencies at fluid temperatures of 200 °C, without using a tracking system [92].

By combing the external CPC and a U-shaped evacuated tube, two truncated CPC solar collectors were developed by Li et al. (2013) [80]. Two CPC with 3x and 6x concentration ratios were evaluated to deliver daily thermal efficiencies of about 40% and 46%, respectively, at outlet temperatures of 200 °C. Gu et al. (2013) [81] developed a novel, portable, CPC-based solar thermal collector for methanol reforming. The experimental results showed that the collector stagnated at temperatures of 320 °C , and it could provide solar heat with an efficiency of ~17-42% for methanol reforming that requires temperatures of around 220-300 °C. Lu et al. (2013) developed a medium temperature evacuated-tube CPC for an absorption cooling system. These solar collectors were found to have higher efficiency compared to that of traditional solar collectors [93]. The collector efficiency of the newly developed collector could reach

50% when the hot water temperature is about 125 °C. A CPC collector was developed and tested by Zauner et al. (2012) [94]. It featured a traditional mirror and absorber design, but make use of an inert gas filling as the box for additional insulation. This enabled the collector to deliver up to 230 °C pressurized hot water. A nearly 30 m2 CPC system was tested by Gudekar et al. [82] for the application of process steam generation. An acceptance angle of 6 degrees was achieved with a concentration ratio of 6.3X. For this design, a thermal efficiency of 25% was achieved for a steam generation process (~120 °C), and the performance analysis of the system showed potential to improve the thermal efficiency up to 71%.

2.3.3 Hybrid concentrators: utilizing both refractive and reflective elements

Some researchers have attempted to minimize the concentrator volume by shortening the focal length, adding a secondary concentrator, or using wave guides and other design concepts [95]. As a general strategy, this approach can be classified as hybrid (catadioptric) concentrators wherein the optical system contains both refractive and reflective elements. By its nature, there are many options in this category, but this section will review the main applications of multi-stage concentrators which can be used for medium temperature solar thermal systems. Thus, a summary of the available composite concentrators, light guides and fibre optics designs are shown in Table 2.6 – a table which compares their performance, technical challenges, and future outlook.

For multi-stage concentrators, the most common optical system consists of the lens as a "first stage" concentrator and a reflective mirror as a "second stage" concentrator. The combination of a secondary concentrator not only increases the concentration ratio and the acceptance angle, but can also decrease the focal length. Soriga and Neaga (2012) [12] designed and analyzed a collector using a linear Fresnel lens (i.e. one with an aperture width: 11cm and a focal length: 15cm) to concentrate solar radiation on a cylindrical cavity receiver. Experimental results showed that the 2-stage concentrating collector had an efficiency of 70%, which was shown to be higher than an evacuated CPC with an efficiency of 66% with an operation temperature of 70 °C under a clear sky summer conditions (905 W/m2). A Fresnel lens collector with a secondary compound parabolic concentrator developed by Pereiar, M. (1979) had a high efficiency

40% when the working temperature reached 255 °C [55]. Moreover, the acceptance half angle of the lens-mirror system was shown to be 4.5°, or double that of a lens alone (which has a very small acceptance angle 2.25°) [55, 61]. Chen (2003) patented a stationary solar photovoltaic array module design, which constituted three or four steps of optical concentration for a photovoltaic power generation system [96]. In this design, a compound parabolic concentrator (CPC) was mounted under the first or second Fresnel lens to further concentrate the sunlight to >20X. The combination of multi-stage Fresnel lenses and optical reflectors could, in theory, concentrate solar energy by 300 to 1000 times for a ~150 mm system height. Terao et al. [97] proposed a non-imaging optics design for a flat-plate CPV system which consisted of a conventional primary/secondary lens combination. They showed (via simulation) that a 250X concentration ratio and an acceptance angle of 2.6 were achievable.

Table 2.6 Summary the studies of representative hybrid solar concentrators Feature Performance Authors η and other Categories W or D th (year) f (m) Cr η θ (m) opt ac remarks Soriga and Neaga (2012) 0.11 0.15 20 / / 70% (To=70 ℃) Multi [12] Stages Primary/secon Terao et al. +/-2. / / 250 / dary lens [97] 6 combination Cylinder Zheng et al. up to Composite 2 1.9 33 / shaped solar [53] 84% concentrator Jason H. Karp / / 900 80% / High et al. [95] concentration Light guide 0.1(entra 0.019 Lin. S and C. nce pupil (thickne 625 / / ratio, low Liang [98] aspect ratio diameter) ss) Arnaoutakis et / / 2000 / / al. (2013) [99] High Feuermann concentration Fibre optic / / 3000 80% / (1999) [55] ratio, high Vu and Shin / / 100 84% ~+/-3 aspect ratio (2016) [100] One disadvantage of most concentrators is that they produce non-uniform light distributions. Fresnel lenses, for example, have a higher efficiency in their central section than near the edges. However, some of the designs proposed in literature can overcome this issue. Genequand and Stark (1980) [101], and Northrup (1977) [102] presented a few concepts to improve the transmission efficiency of Fresnel lenses. In

their study, a set of reflective slats were installed in a frame and supported on each side to better distribute the incident solar rays. Zheng et al. [53] presented a design of a cylindrical-shaped solar concentrator which comprised of an arched Fresnel lens and a Fresnel mirror and a secondary reflector. The cylindrical shape enabled low wind and snow loads on the top surface.

As another alternative, waveguide concentrators – which make use of total internal reflection – have also been proposed for solar thermal systems. In these designs light is effectively trapped inside a transparent dielectric material, whereby it can be directed and concentrated to the solar absorber. Karp et al. have developed a planar micro-optic concentrator with a 900X concentrator ratio and 85% optical efficiency [95]. The sunlight collected by this lens array was coupled into a microstructure waveguide surface, whereby the light was reflected by total internal reflection (TIR) onto the receiver which was mounted at the edges [95]. This system followed the sun's position using a two-axis tracking platform.

Several light guide solar panel collector designs have been patented – including those by Soleau (1981) [103], Morgan (2011) [104] and Murtha (2000) [105] [106]. All of these designs allow for very thin modules whose thickness is comparable to the height of the solar PV collector. Lin. S and C. Liang [98] proposed a concentrator design allowing the light beams from all parts of the concentric annular entrance pupil to be collected and transferred to the receiver with minimal loss. Their concentrator had an entrance pupil diameter of 100 mm which was designed to keep the total system thickness to be 19 mm for a solar cell size of 4 mm. Ultimately, the design had an impressive geometrical concentration ratio of 625X times, with a corresponding aspect ratio of 0.19 (is defined by concentrator system length divided the concentrator system diameter) [51, 98].

Rather than plate wave guides, solar energy can also be concentrated and efficiently transported by optical fibers, as was first proposed in 1980 by a group of French investigators [74]. By again trapping light in a dielectric material, fiber-optic technology provides a flexible way of transmitting solar flux directly to an application. These“ cables of light” provide a wide range of possibilities, such as: solar lighting, solar-pumped lasers, solar concentrating PV/Thermal, solar-powered photo-catalysis,

and many others [99]. As a solar lighting system, concentrated rays are directed by primary and secondary mirrors to a full-spectrum fiber optic bundle. Liang et al. (1998) reported on a flexible light guide which consisted of 19 optical fibers and was capable of transmitting up to 60W of optical power, with 60% efficiency. Arnaoutakis et al. (2013) [99] reported and tested a novel two-stage solar concentrator and optical fibre. The result showed that a maximum concentration ratio of 2,000X was achievable at the end of a single fibre system. Feuermann (1999) [55] proposed an efficient solar energy concentration and power delivery system. The system adopted a miniature (e.g. 0.2 m diameter) solar dish which concentrated sunlight into a single optical fiber. The solar field comprised of many modules, with all optical fibers transporting concentrated sunlight to a remote protected receiver. They estimated that collection efficiencies as high as 80% and flux concentrations in excess of 30,000 suns were achievable, even at a receiver distance of around 100 m from the concentrator. A cost-effective optical fiber composed of modified compound parabolic concentrators (M-CPC) coupled with plastic optical fibers (POFs) was proposed and analyzed by Vu and Shin (2016). This M-CPC was made by combining two conventional CPCs into one component. The optical simulation results demonstrated an optical efficiency of up to 84% when the concentration ratio of the M-CPC was fixed at 100 [100].

2.4 Thermal concentration

In order to achieve outputs suitable for commercial and industrial applications, optical concentrators are conventionally required to increase the temperature and efficiency of a solar thermal system’s output. In this section, we instead review the studies utilizing thermal concentrators to boost the performance of solar thermal collectors.

Thermal concentration, although rarely referred to as such, is a concept that is frequently employed in heat transfer systems. The use of thermal concentration to focus absorbed heat flux has been discussed in several studies. Kapadia, R. and Bandaru P. [107] proposed and demonstrated a composite thermal lens (heat flux concentrator) that enabled a five-fold enhancement of the heat flux by adjusting the thermal conductivity as well as parameters such as the thermal traversal length and cross-sectional area. Their ideas were borrowed from transformation optics, whereby light rays can be guided to their ultimate destination in a metamaterial lens by varying the refractive index

throughout the lens body [69]. Kraemer et al. (2011) reported a TEG which achieved a peak efficiency of 4.6% under AM1.5G (1kW/m2) conditions when the geometric thermal concentration was 299 [108]. Chen (2011) predicted that STEG efficiencies of approximately 12% could be achieved with a 10-fold optical concentration and a 200-fold thermal concentration [109, 110].

In solar thermal applications, most collectors are already providing some thermal concentration, but without a specific, intended design. A heat pipe collector with aluminium absorber plates was developed and tested by Azad (2008) [111]. The collector had a large absorber which exchanged heat at the much smaller condenser area of the heat pipe header. Additionally, both vapor compression and thermoelectric heat pumps could be regarded as energy concentrators since they pump energy ‘uphill’ from a cold side (the evaporator) to the hot side (the condenser) of a device. Various studies have been performed on these designs, including many for use in solar systems. A solar-assisted heat pump (SAHP) system with two flat-plate collectors, acting as evaporators for the refrigerant R-134a, was tested by Soldo, et al. (2004). COP values of up to 6.7 were achieved when the condenser and outlet temperatures were fixed at 50 °C and 48 °C, respectively.

2.5 Mitigating radiative loss in medium temperature collectors

2.5.1 Selective surface absorbers

The solar absorber needs to be spectrally selective in order to be optically and thermally efficient, which translates to a high solar absorption in the UV–vis–NIR solar spectrum and a low thermal emittance in the IR wavelength region. They are usually constructed with a selectively solar absorbing thin film coated on a highly infrared reflective metal substrate. The top absorbing layer is designed to absorb the solar radiation while the underlying metal substrate should reflect infrared light, i.e. minimise thermal emission. Highly reflective metals such as aluminum, copper and stainless steel are commonly used as substrates. For the absorbing layer, black chrome, black nickel and cermets are among the most widely used materials. Techniques like anodization, electroplating, sputtering, spin-coating, chemical and physical vapor deposition have been utilized to fabricate these absorbing coatings.

Selective absorber surface coatings can be categorized into a number of different types, including: a) intrinsic, b) semiconductor-metal tandems, c) multilayer absorbers, d) multi-dielectric composite coatings, e) textured surfaces, and f) selectively solar-transmitting coating on a blackbody-like absorber [112]. Intrinsic absorbers use a material having intrinsic properties that result in the desired spectral selectivity. Semiconductor-metal tandems absorb short wavelength radiation – or photons above their bandgap – but have low thermal emittance as a result of the metal layer. Multilayer absorbers use multiple reflections between layers to absorb light and can be tailored to be efficient selective absorbers. Metal-dielectric composites – cermets – consist of fine metal particles in a dielectric or ceramic host material. Textured surfaces can produce high solar absorptance by multiple reflections among needle-like, dendritic, or porous microstructures. Additionally, selectively solar-transmitting coatings on a blackbody-like absorber are used but are typically used in low-temperature applications.

A summary of the various types of selective surfaces that have been studied are shown in Table 2.7. It can be seen that both black chrome and TiNOx (TiNxOy cermet) have a good absorptance (>0.92) and a very low emittance (<0.1). More importantly for the thrust of this thesis, they are stable (in a vacuum) up to 400 °C.

Table 2.7 Properties of mid-temperature selective surfaces [112, 113] Emittance Stability (°C) Material Substrate Fabrication Absorptance (100 °C) Vacuum Air Electrodeposit Black chrome Ni-Cu 0.97 0.09 400 350 ion TiNOx Cu ARE 0.92 0.06 400 Electrodeposit Black nickel Steel 0.88-0.96 0.03-0.1 <200 ion Black copper Electrodeposit BlCu-Cu O: Cu 0.97-0.98 0.02 370 250 2 ion Cu Reactive Ni- NiOx Al 0.96 0.10 300 sputtering 300-4 Ni- Al O Al Anodization 0.85-0.97 0.08-0.21 2 3 00 Most of these solar thermal collectors use receivers with the ‘surface-based’ absorbers discussed above to efficiently convert solar radiation into thermal energy [1, 3]. It can be inferred that producing these selective films (e.g. TiNOx) via vapour deposition of high purity materials [112, 113] is somewhat complex and costly. From a thermal management point of view, heat from the outer surface must first conduct through the

solid plate/tube and then be transferred into a working fluid. A temperature drop is caused by conductive and convective resistances between the outer absorber surface and the working fluid. Additionally, from a heat loss perspective, applying the highest temperature on the outer absorber surface is not ideal since it ultimately drives the heat loss with the surroundings [114].

2.5.2 Volumetric absorbers

Several researchers have proposed that these issues can be addressed through alternative, volumetric absorbers where the working fluids themselves absorb solar energy [114-118]. This so-called ‘direct absorption collector (DASC)’ potentially benefits from suppressed heat loss on the outer absorbing surface due to a lower outer absorber surface temperature [115]. Since all common solar collector working fluids (water and oils) are transparent over the solar spectrum, they require additives (e.g. molecular dyes or nanoparticles) to ‘directly absorb’ solar energy. If nanoparticles are used, only a very small amount (< 0.01% by volume) of nanoparticles is needed to achieve high absorption for solar thermal receivers. Low nanoparticle concentration indicates a correspondingly low added cost from the base fluid [119] and, potentially, high stability due to less particle-particle interactions [118].

Recent studies (shown in Table 2.8) have demonstrated that direct volumetric absorbers [118, 120-122] and optical filters [117, 123] may result in solar collection efficiency enhancements of up to 35% [114, 115, 118, 124-127]. However, present direct absorption collector (DASC) experimental studies are limited to either: a) flat plate solar thermal collectors (operating temperature <100 °C) [126, 128-130] or, b) indoor laboratory-scale concentrated solar thermal collectors (operating temperature >100 °C) [115, 127]. In terms of the best nanoparticle materials to use, several studies have demonstrated excellent solar absorption with carbon-based materials, such as CNTs and graphite [114, 115, 118, 124, 125, 131]. Carbon-based nanofluids have broadband solar high absorption (close to 100%) at low volume fractions (~0.01 %) [115] and can be potentially produced at low cost (~ $1/L [131]). In addition, the co-authors have developed functionalization techniques for (specifically for carbon nanotubes) which can maintain stability to 250 °C [121, 131] – a technique which will be used in the experiments of the present study.

Thus, to our knowledge, no full-scale concentrated direct absorption collector has been tested ‘on-sun’. Since nanofluids appear to have some advantages in concentrated solar systems [118, 132], there is a clear and present need to conduct comparative experiments and simulations to evaluate their potential as compared to surface absorbers under real operational, high temperature conditions.

Table 2.8 Summary of DASC studies Category Author (year) Materials Remarks Karami et al. (2015) CuO + Water [126] Low-temperature Gupta et al. (2015) Experimental studies with Al O + Water (<100 °C) DASC [129] 2 3 flat-plate solar collectors Delfani et al. (2016) MWCNT + Water [130] Laboratory-scale dish Taylor et al. Graphite + collector (2011-2013) [114, Therminol VP-1 experimental/numerical 115] study High-temperature Carbon-coated Indoor test/numerical study Lenert et al. (2011) (>100 °C) DASC cobalt + Therminol of a cylinder [127] VP-1 central-receiver Laboratory-scale Khullar et al. (2013) Al/Therminol VP-1 concentrating parabolic [133] collector numerical study Optical filters Hjerrild et al. (2016) Ag–SiO nanodiscs Experimental study in a (Selectively 2 [123] and CNTs + water (PV/T) collector absorbing fluids) 2.6 Integrated storage

Even if everything performs well technically (section 2.1-2.5), the solar resource still may not meet the demand profile of industry process heating or commercial buildings (the demand may vary during any given day and from one day to next). Therefore, thermal energy storage (TES) systems can help balance energy demand and supply on a daily, weekly and even seasonal basis. They can also reduce peak demand, energy consumption, CO2 emissions and costs, while increasing overall efficiency of energy systems. Furthermore, the conversion and storage of variable renewable energy in the form of thermal energy can also help increase the share of renewables in the energy mix.

2.6.1 Thermal Storage Overview

Energy storage plays an important role in providing a more flexible and stable supply of energy. Solar thermal systems have one important potential advantage over photovoltaics – thermal energy storage (TES) [134]. Thermal energy storage can be cost-effectively employed to overcome the stochastic, diurnal, and (perhaps) even seasonal variations in solar radiation, enabling a closer match between energy supply and industrial heating demand [135]. Thermal energy storage at temperatures from 100-400 °C involves either sensible heat thermal energy storage (SHTES) (using either solid materials (pebbles, concrete, etc.) or liquids (salts or oils) or latent heat thermal energy storage (LHTES) which take advantage of the latent heat of solid-to-liquid phase change [136]. Several recent studies have been conducted to compare LHTES and SHTES which have shown that a significant reduction in storage volume can be achieved using LHTES due to high energy density (typically 5–14 times higher than a SHTES for the same working temperature range) [136, 137]. These advantages offered by LHTES such as heat storage capacity and a, small unit size are desirable for integrated collector/storage solar systems.

2.6.2 Solar collector/storage integration

There have been several recent developments in solar systems which integrate the collector and the TES into a single/compact package. Given the order of magnitude size reduction potential of LHTES systems, a number of the studies in the literature employ phase change materials (PCM) as a means to achieve performance enhancement and cost reduction.

For low temperature applications (<100 °C), as shown on Table 2.9, an integrated collector storage solar water heater (ICSSWH) with a cylindrical storage tank (containing PCM or water) properly mounted in a CPC has been investigated by Chaabane et al [138]. They found that the LHTES was more effective as it allowed better heat preservation [138]. An evacuated tube collector integrating PCMs in the inner glass tubes was proposed and tested by Papadimitratos et al [139]. They reported an efficiency improvement of 26% for normal operation and an improvement of up to 66% for the stagnation mode, as compared with standard solar water heaters without phase change materials [139].

For medium temperature applications (100-400 °C), ICS with LHTES system studies have been only found for solar cooker applications, where the LHTES is charged by direct illumination using a concentrator, the energy stored can then be used later for cooking [140, 141]. As shown in Table 2.9, these solar cooking systems incorporated with LHTES are able to store heat energy in latent form at 100–250 °C and are suitable for off-sun cooking during the evening (up to 3 hours after sunset) after a sunny day. The latent heat storage systems in this medium temperature range are preferred because of their high energy density and this relative high temperature could reduce cooking time and enable frying.

Table 2.9 Summary of ICS with built in LHTES system studies PCM(s) / Temperature Author (year) ICS type Melting Methods Range Temperature(s)

Air heater and Flat TES green house Bouadila et integrated with Capsule (SN27) Experimental heating (<40 al.(2016) [142] flat absorber in 27 °C study °C) flat plate collector Khalifa et al. (2015) [143]; Flat TES Theoretical Paraffin or Chen et al. integrated with and salt soda, (2010) [144]; flat absorber in experimental 33-80 °C Varol et al. flat plate collector study (2010) [145]; Cylindrical TES Domestic water Tritriacontane and Papadimitratos et integrated in the heating (<100 Erythritol, 72 °C al. (2016) [139]; inner glass tubes Experimental °C) and 118 °C; Naghavi et. al of study Paraffin wax, 64 (2015) [146] evacuated tube °C collector TES integrated Theoretical Chaabane et al. with the RT-42 graphite, and (2014) [138]; cylindrical myristic acid experimental absorber of CPC study Erythritol,118 °C; Lecuona et al. TES integrated Theoretical Galactitol, Solar cooker (2016) [140] with the receiver and 150-200 °C; (100~200 °C) Mussard et al. of concentrating experimental Nitrate salts, 220 (2013) [141] dish collector study °C Slocum et al. Theoretical Electricity TES integrated (2011) [147] Solar salts, ~222 and generation with the receiver Gil et al. (2015) °C experimental (>400°C) of CST tower [148] study

At the very high temperature side, a concentrated solar power on demand (abbreviated ‘CSPonD’) volumetric receiver/TES unit prototype was developed and analysed by Gil et al. [147, 148]. In this system, the sun light was directly concentrated onto the molten salt storage tank (cold salt in the bottom region, at 250 ºC and hot salt on the upper region, at 550 ºC), avoiding the necessity of pumping the salts to the top of a tower to two storage tanks. A total system performance enhancement was found due to this system simplification and reduction in heat loss and pumping power, resulting in a possible levelized cost of electricity of $0.07–0.33/kWh generated from this system.

Based on the available literature, it can be concluded that most studies to date are for low temperature phase change materials storage, with only a few studies covering medium and high temperature applications (>100 °C). Additionally, medium or high temperature rooftop collectors have not yet been integrated into TES systems.

2.7 Chapter Summary

It has been found that much of the fundamental science has been investigated in the area of optical, thermal, and materials development – topics which if brought together in innovative designs may lead to thin rooftop concentrated solar systems. Due to the vast market for the supply of 100-400 °C thermal energy for domestic and industrial applications, it is expected that these systems could see significant market penetration if high efficiency (>50% at 200 °C), low-profile (<15 cm) systems were to be developed which can be fit on standard PV racks. Ultimately, high efficiency, low profile, aesthetic, and easily integratable with rooftops will be the main features for the next generation of solar thermal collectors.

Current state-of-the-art rooftop solar concentrators for the temperature range 100-400 °C are mainly relegated to parabolic troughs – a technology which has changed little since their inception. Improvements have been made in vacuum design, maximum concentration ratio, better reflectors and selective surfaces, but there are many significantly different options for advancing the optical concentrating system with higher concentration ratios and lower focal length if we shift to alternative platforms. Although there are a lot of challenges, new technologies can be employed to create thin, high temperature solar concentrators. According to the literature discussed above, it can be identified that: 33

. Rooftop integrated solar thermal systems can, in theory, supply medium to high temperatures (100-400 °C) directly to industrial heating, air-conditioning and commercial steam applications. There are a number of technologies available which can deliver up to 400 °C and a high heat flux in the context of rooftop solar thermal collectors, including both innovative optical and thermal concentrating systems. . 6-8 times spectral concentrating ratio of band-focus Fresnel lens is achievable when the focal length ranges from 100-150 mm (the limit height of aimed package design). . It is a challenge for an optical concentrator to reach a high concentration ratio with short focal length, but novel designs such as metamaterials and, light guide concentrators may provide viable solutions if developed. In particular, a multi-stage concentrator seems to be the most feasible in near-term solar thermal applications. . Solar trackers add significant capital cost to conventional concentrating collectors. To reduce the complexity and cost of collectors, ‘stationary’ concentrators (e.g. where the module itself remains fixed but internal components can move) provide a promising and cost-effective approach to meet the needs of commercial and industrial applications. . Vacuum CPC or CPC with vacuum receiver design (half acceptance angle of 45 degrees) can reach a 30-44% thermal efficiency when operating at 200 °C (see section 2.3.2). Increasing the concentration ratio of vacuum insulated CPC collectors was demonstrated as a very promising and effective approach to enhance the instantaneous thermal efficiency further. However, improvement methods reported in the literature were either adding a lens or reducing the acceptance angle of the CPC, which restricts the advantage of CPC’s passive tracking as those design proposals require a rotating tracking system. . Most of today’s solar thermal collectors use receivers with ‘surface-based’ absorbers, however, alternative, volumetric absorbers where the working fluids themselves absorb solar energy, may provide some advantages. . A hidden benefit of developing new concentrating collectors is the potential to integrate storage units in ‘free’ space of the optical package. Storage, if designed

in, can help the transient solar resource to be more usefully dispatched to meet the energy demand of industrial and commercial customers.

Chapter 3

3. Innovative Optical Concentration

Solar energy, while being a highly abundant renewable energy resource (i.e. ~10,000 times the annual global energy demand is incident on the earth), has a relatively low energy flux density (around 1 kW/m2). Optical concentrators are designed to bring the maximum amount of sunlight to a smaller absorber. This is typically done to facilitate significant reductions in heat loss via a small heat loss area. To attain higher temperatures (>100 °C) for commercial and industrial applications, optical concentrators are usually required.

Optical concentrators can be categorized as being composed of either reflective or refractive optical components. In both cases optical concentrators can generally only utilise direct normal irradiance (DNI). This limitation not only reduces the amount of useful sunlight available, but also necessitates solar tracking as an essential component. Since a substantial amount of energy is consumed in tracking, researchers have developed several tracking options [53, 149], but, to date, most conventional optical concentrators are encumbered with the problem of being bulky and heavy (e.g. parabolic trough reflectors).

As concluded in section 1.2, the aim of the optical design is to provide a ‘stationary’ rooftop collector (e.g. with semi-active or passive tracking, as defined in section 2.2), which could reduce the complexity and cost of medium temperature collectors. At the same time, in order to integrate with building rooftops, and to avoid wind loading issues, it is desirable to have optical concentrator which have a low-profile (e.g. <10 cm). Hence, there is a clear design challenge to develop optical concentrators with a high concentration ratio, but minimal effective focal length.

The design presented in the following section takes up this challenge. By implementing a novel internal tracking mechanism, the final design provides an innovative concentration platform which can concentrate beam radiation to a tube receiver during the highest flux hours of a day without rotational tracking or any external moving parts. Thus, it can be mounted on standard PV racks, resulting in substantial system simplification and cost reductions.

3.1 Optical design and analysis

As stated above, it is desirable to achieve a low-profile design which achieves both a high concentration ratio and optical efficiency. As an initial design constraint, the total thickness of the optical concentrator system is limited to 10 cm (the approximate thickness of a PV module frame). The design process in the following sections will examine the achievable optical concentration ratio and optical efficiency based on this thickness (~10 cm).

As concluded in section 2.7, a multi-stage concentrator (e.g. a lens with a CPC) provides a feasible solution to increase both the concentration ratio (Cr) and the acceptance angle as compared to separated lenses or a CPC concentrator alone in terms of reducing the focal length. In this section, the design process for specifying the optical design of the lens / CPC combination is presented. As the proposed lens design can focus sunlight from a wide range of incident angles (up to 45 degrees) onto the same plane, the receivers need only to be moved linearly to stay within the focus band. Thus, this optical system eliminates the need for a conventional rotational tracking system, reducing the cost and the complexity of the system. Instead, a relatively simple linear tracking system is employed.

More specifically, the following design work will describe the procedure for determining the geometry the lens/CPC to realize an optical system with a ~10 mm focal length (e.g. inside the design objective as was stated in section 1.2), a concentration ratio of ~5X, and a 100 mm height concentrator. The optical system also has a ~45 degree half acceptance angle (to capture the peak sunny hours of a day). As a less critical constraint, the linear travel of the internal components should be limited to less than half the lens width.

3.1.1 Design of the Double Lens

The concentrator design was initiated by examining a single cylindrical lens. According to the Lensmaker's equation, a thin lens (e.g. where the thickness is very small compared to radii of curvature) can be characterized by Equation (3.1) [150]. A relatively high refractive index, n, and/or large radii of curvature, R2, is necessary to achieve a short focal length, f. As a baseline, a plano-concave lens was used for a preliminary design and analysis.

1 1 1 n1    (3.1) f R R 12 Due to the fact that most plastics and glasses have refractive indices in the range of 1.3 to 1.7, a typical optical glass BK7 with a refractive index of 1.52 and infinite R2 was chosen.

The width of lens was calculated to be ~150 mm by multiplying the perimeter of a ~10 mm diameter pipe receiver with effective concentration ratio (Cr~5 times). This can be calculated by Equation (3.2) [151].

W Cr  (3.2) D

The ray tracing results of a single lens with various incident angles is shown in Figure 3.1. This simulation shows that the focus line follows a parabolic path when incidence angle increases from 0 to 40 degrees. This highlights the need for conventional rotational tracking to maintain focus along the parabolic focusing path.

Figure 3.1 Ray tracing results of a single lens

By combining and optimising two single lenses with different radii of curvature (ranging from 50-500 mm) and refractive indices (ranging from 1 to 2), a double lens was designed which can focus sunlight from a wide range of incident angles (up to 45 degrees) onto the same plane. Figure 3.2(a) shows the detailed design parameters (radii of curvature and refractive indices chosen) of the lens. BK7 optical glass (nd=1.52) was chosen for the first lens and F2 optical glass (nd=1.62) was chosen for the second lens, with the difference in refractive indices enabling a nearly constant focal length of 100 mm. BK7 is used as a positive lens to concentrate the light and the F2 lens is used to reduce the optical aberration at large field angles.

(a) (b) Figure 3.2 Optical design (a) radii of curvature (unit: mm) and refractive indices chosen for

double lens (b) optical simulation of double lens Figure 3.2(b) shows a composite optical simulation (Zemax) of the final double lens design with different incident angles. To achieve concentration ratio of ~5, the width is set to 100 mm (~perimeter of the pipe receiver × Cr) based on the maximum thickness of the package (10 cm) and the standard aperture of the receiver (7.92 mm or 3/8 inch). Any further increase in the width will lead to additional optical losses due to edge effects [152]. The curvatures and thickness of the lenses are obtained by a Zemax optimization to ensure a constant focal line as the incident angle changes from 0 to 45 degrees. The advantage of using the double cylindrical lens design is that the focal line shifts nearly transversally, enabling a much-simplified linear tracking system.

3.1.2 Design of the Cylindrical Lens-CPC system

While the double lens approaches a constant focal line, the results are not perfect. Incident light approaches the absorber across a range of angles, which necessitates a CPC (one of the only optical components which can concentrate non-collimated light) as a secondary concentrator [45, 153].

Figure 3.3 Geometry of optical system (unit: mm)

Figure 3.3 shows the design of the CPC developed in this research. In order to ensure the concentrating system works effectively for the sunniest hours of the day, the

maximum half acceptance angle of CPC is set to 45°. [Note: Since the Earth turns at 15 degrees per hour, this system will work from 9 am to 3 pm solar time]. The aperture of the CPC (35.2 mm) and the working length of the CPC (29.5 mm) are determined based on the acceptance angle of CPC and the size of the receiver, which also satisfies the requirements of the total package thickness and the width of the double lens. Segment AB and A’B’ is parabola which captures all the light that is incident within 45 degrees. Segment BC and B’C is an involute of the receiver tube whose cross-section is circle. The involute parts (BC & BC’) connect to the parabola parts (AB & A’B’) smoothly at point B and B’.

0° 15° 30° 45° Figure 3.4 Simulation the performance of CPC with different incident angle Combing the double lens with a CPC gives a total concentration ratio of ~4.5 and a total system working length (to 9.8 cm in total). Figure 3.4 shows the CPC’s performance at various incident light angles from 0 to 45 degrees, while Figure 3.5 shows the performance of whole optical system.

0° 15° 30° 45° Figure 3.5 Simulation of Lens with CPC with different incident angle These ray tracing simulation demonstrate that the proposed double cylindrical lens with CPC can focus energy on the pipe receivers effectively during the sunniest six hours of a day (from 9am-3pm, solar time) by moving the receivers/lens left or right by ~50 mm. However, a large area cylindrical lens will likely add substantial cost and weight to a solar collector. Moreover, since photographic accuracy is not necessary for collection of the solar rays [154], an alternative (cheaper and lighter) non-imaging lens should suffice, as will be conducted in following (section 3.1.3).

3.1.3 Design of the Fresnel Lens-CPC system

A Fresnel lens is modified form of conventional lens, in which the contour profile of conventional lens is maintained and undesired material removed, thus absorption losses and the cost/mass of the material can be reduced substantially. A non-imaging Fresnel lens allows optimal transfer of light from source to target. For solar energy applications the fact that it does not form any image is acceptable. Also for this application, the widest possible acceptance angles and the highest tolerances in manufacturing and operation are desirable. This enables the system to work reasonably well with a less precise tracking system and some misalignment during assembly [155]. The use of PMMA in double Fresnel lens design can significantly reduce the weight of the system relatively to a glass lens. Also, plastic Fresnel lens is much easier and cheaper to manufacture by means of molding and extrusion [154-156].

However, during the optical design process it was found that it is difficult to shorten the focal length of a Fresnel lens to less than 10 cm while maintaining a relatively high concentration ratio. To solve this problem, a double linear Fresnel lens with curved structures on both top and bottom surface was proposed as shown in Figure 3.6. By optimizing the radius curvature of the two individual Fresnel lenses and the depth of grooves on surface, a focal length of less than 10 cm can be achieved with a relatively high concentration ratio. The material used for the Fresnel lens is PMMA (nd = 1.49).

Figure 3.6 Dimensional drawing of the optimised double linear Fresnel lens design (units: mm)

Similarly, as shown in Figure 3.7, with the help of the CPC, the double linear Fresnel lens is able to concentrate sunlight onto the receiver with an incident angle of up to 45o by moving CPC within a distance of ~60 mm.

(c) (d) Figure 3.7 Simulation of Lens with CPC with different incident angle As can be seen in Figure 3.5 and Figure 3.7, when the rays come in with a large incidence angle (e.g. >15 degree), the edges of the lens are a big source of loss due to total internal reflection. That is, at the interface light bounces back into the double cylindrical lens/Fresnel lens and occasionally escape at angles leading away from the receiver.

3.2 Optical Performance

The optical efficiency at different incident angles is defined by the following equation:

Q()r  o  (3.3) Q()t 

θ Where, Qr ( ) is the power received by tube at incident light angle θ, Q()t  is the total energy received by the concentration system at incident light angle θ. Ray tracing software (ZEMAX, version 12.EE) is employed to simulate the optical performance of the components at different light incident angles. From this analysis the total effective system optical efficiency can be calculated (see Figure 3.8).

Figure 3.8 Optical efficiency comparison as a function of incidence angle

It can be seen from Figure 3.8 that the maximum optical efficiency reaches around 73% for the double cylindrical lens design and 90% for the double Fresnel lens design (both at normal incidence, θ = 0o). This means that both concentrators work most effectively in the middle a day (as expected). As the incident light angle increases, the system efficiency drops slowly. However, the system still has an efficiency of around 52% for double cylindrical lens design and 45% for the double Fresnel lens design when the incident angle increases to 45o. This indicates both concentrators are most useful during the middle 6 hours of a sunny day. Options to improve the optical performance, particularly at large incidence angles, will be discussed and demonstrated (later) in chapters 8 and 9.

Another thing worth mentioning from Figure 3.8 is that within small incident light o o angle range (around 0 ~ 25 ), the optical efficiency of double Fresnel lens design is higher than double cylindrical lens design. However, the double cylindrical lens design o o has better performance across a larger incident light angle range (around 25 ~ 45 ).

The Incidence angle modifier (IAM) is defined as the efficiency at the given incidence angle divided by the efficiency at normal incidence, it can be determined by the following Equation (3.4) [71].

ηθo IAM (3.4) ηo θ 0

where ηθo represents the optical efficiency at an incident angle of θ, and

ηo θ 0 is optical efficiency at normal incident angle. The calculated results are given in Figure 3.9.

Figure 3.9 Transverse IAM comparison as a function of incidence angle

It should be noted that the individual incidence angle modifier can be estimated by considering it to be the product of the transversal and longitudinal modifiers, K(θ)T and K(θ)L, respectively [157]. [Note: The longitudinal incident angle modifier will be evaluated via experiment in chapter 7.] With accurate knowledge of the IAM and the instantaneous thermal efficiency, the performance of collector system can be predicted for all incidence angles over the year (also discussed in chapter 7).

Table 3.1 summarizes and compares the key parameters of the two proposed concentrator designs. As expected in section 2.6, a concentrating ratio of 6X was achieved in a limited height optical system (<10 cm). Compared to a single CPC (with a half acceptance angle 45˚), the average optical efficiency decreased by ~10-15%, but the concentration ratio is significantly enhanced by 3-4 times. Thus the lens combined with CPC designs proves advantageous over a conventional CPC design. It also worth noting that a vacuum insulation collector with a concentration ratio of 1.5X can reach

up to 44% thermal efficiency [41, 45, 81] when operating at ~200 °C. Thus, increasing the concentration ratio of a vacuum insulated CPC collector can enhance the instantaneous thermal efficiency further.

This result also differs from methods reported in the literature where either adding a lens or reducing the acceptance angle of the CPC was found to degrade the passive tracking of the CPC with/without a rotating tracking system. The design reported herein can both increase the concentration ratio and maintain the large acceptance angle.

Table 3.1 Comparison of two concentrator designs

A single CPC Double cylindrical Double Fresnel lens Characteristics (Half acceptance lens + CPC design + CPC design angle 45˚)

Concentration ratio 4.55 5.91 1.5

Average optical efficiency 66.5% 68.9% 80% (over 6 working hours)

Thickness of system 8.96cm 9.73cm 3-4 cm

Material Glasses (BK7 & F2) PMMA /

Internal moving ~50 mm ~60 mm / distance

As can be seen from Table 3.1, the double cylindrical lens design has a shorter system thickness, which can result in a more compact collector package. However, the double Fresnel lens design has advantages over double cylindrical lens design in terms of optical concentration ratio and efficiency. Moreover, while the proposed cylindrical lens was custom-made at a very high cost (>10,000 US $/m2), various Fresnel lens products (e.g. at a range of sizes, shapes, and focal types/length) are commercially available for low cost (~20-100 US $/m2). Thus, a Fresnel lens combined with a CPC represents a more cost-effective solution for further development and commercialization in the field of solar thermal collectors.

3.3 Chapter Summary

The design and analysis of a new low profile thermal concentrator and its collector package were presented in this chapter. Compared to other concentrating solar

collectors (which have rotational tracking systems), this represents a significant design simplification.

By integrating a specially designed concentrating lens, a CPC reflector, and a linear tracking mechanism, both designs (i.e. the Cylindrical Lens-CPC and Fresnel Lens-CPC systems) can achieve around a 5 times concentration ratio within a limited system thickness (<10 cm), so they can be easily mounted on industrial and commercial rooftops. Using the optical findings from this chapter, the next chapter will report on the thermal design of the collector.

Chapter 4

4. Collector/Package Design and Analysis

In the previous chapter, low profile (<10 cm height) optical concentrator designs based on a linearly-actuated, catadioptric optical system (using both reflective and refractive optical elements) were presented which can achieve a ~ 4X concentration ratio throughout the day. To reach a high thermal efficiency for the medium temperature range (100-250 °C) vacuum insulation was deemed necessary.

In this chapter, a holistic collector design is presented which brings together these components (i.e. the optical concentrator, internal tracking, and vacuum insulation). To analyse the performance of the proposed designs, Computational Fluid Dynamics (CFD) was used as an iterative approach to assess the impact of design changes. Note

that the optical results from the last section will be considered as ‘inputs’ to the thermal design and analysis.

4.1 Collector/Package design

Two general thermal designs are proposed in this research, both of which include the optics from chapter 3 and vacuum packaging to suppress internal convection and air conductance heat losses. In ‘Design 1’ a planar vacuum chamber is used whereby the CPC is exposed directly to the vacuum. In ’Design 2’, only the receiver is exposed to the vacuum via an evacuated glass tube. In both designs a stepper motor is used to move the receiver or lens relative to each other to keep the receiver in the focal line.

4.1.1 Design 1: Planar vacuum chamber

Figure 4.1 shows conceptual drawings of a design which incorporates the tracking system, and optical components mentioned in this chapter into a planar vacuum chamber. In this case, the collector utilizes a double lens because, regardless of incident angle, it has a nearly uniform focal length. A CPC is used as the secondary concentrator because it has a wide acceptance angle for focusing incident light onto the absorber. Together these allow for a simple linear actuator as a tracking mechanism. The absorber is a standard (5/16”) serpentine copper tube with a selective coating (emissivity = 0.05 - 0.10) on the outside surface [102].

(a) (b)

(c) Figure 4.1 Drawings of the proposed design (a) 3D view of collector; (b) exploded view of the collector; (c) sketch of the collector in the cross section For complete convective suppression, Beikircher et al. [103] suggest that the interior of chamber should be evacuated to a pressure below 0.1 Pa. This internal pressure can be achieved by a turbo molecular pump and maintained using getter flashing [158]. Due to the pressure difference between the ambient (1 bar) and the vacuum chamber, metal support pillars can be positioned within the collector to withstand atmospheric pressure forces , such as are found in the flat plate design of TVP-SolarTM [46]. To determine the optimum mechanical design of these supports, finite element modelling and experimental investigation have been used by researchers – e.g. Henshall et al. [159] who found that, the results indicated that 400 series stainless steel pillars are sufficient to support a sheet of 4 mm thick tempered glass as the cover. However, it is difficult to obtain good enough tolerance on the thin pillars to equally distribute the pressure load, so the pillars must be oversized in terms of their supporting contact area to achieve a more uniform load distribution and reduced stress for vacuum glazing structures [159]. In the proposed design, the upper surface of CPCs has large contact area, so the CPC itself could also be employed as a structural element to resist the forces induced in the cover by the vacuum, as is shown in Figure 4.1.

Figure 4.2 Ray tracing diagram for package with (a) 0o incident light angles; (b) 45o incident light angles For this design, a ray tracing analysis (simplified version depicted in Figure 4.2), the collector can easily track the sun by moving the receiver/lens horizontally. The total working distance of vacuum chamber of this tracking movement is ~10 cm, which matches well with the optical ray tracing found in section 3.1.2. The horizontal movement of planar vacuum chamber can be controlled by using an Arduino microcontroller controlling system. There are four wheels assembled at the corner of vacuum chamber, which allows the vacuum chamber to slide on a track directed by stepper motors.

4.1.2 Design 2: Evacuated glass tubes

Figure 4.3 (a) gives a 3D view of this collector design with evacuated glass tubes insulating only the absorbers. A cross-sectional view of the design is displayed in Figure 4.3 (b). The design consists of three optical units, each of which are comprised of a primary lens (a linear Fresnel lens, 1.5 m in length by 0.15 m in width) and a compound parabolic concentrator (CPC) [90, 160] (1.5 m in length and 0.043 m in width). Sunlight is concentrated through the primary lens onto the entrance aperture of CPC. The CPC pieces consist of custom shaped aluminium pieces which are coated

with a reflective film. The CPC then concentrates the incident light further onto the tube receivers.

(a)

(b) Figure 4.3 Collector design: (a) 3D end view of the collector; (b) transverse cross-sectional view

Characteristic results from this ray tracing analysis are shown on Figure 4.4. Stepper motors and a timing belt controlled by an Arduino microcontroller (not illustrated in the figure) are used to facilitate movement of the CPC carriages along horizontal rods either left or right by ~60 mm.

(a)

(b)

Figure 4.4 Ray tracing results for (a) 0o incident light angles; (b) 45o incident light angles

The tracking system has been implemented as an open loop control system as showed in Figure 4.5. The system uses an Arduino-UNO evaluation board and the stepper motors (manufacturer, 42BYGHM809).

Figure 4.5 Open Loop Control System: PC, Arduino microprocessor (μP) board, Motor Board and stepper motor

The Arduino UNO board is interfaced with a companion Adafruit Motor Shield V2 board, which drives the stepper motors through a Toshiba TB6612 MOSFET H4-bridge driver (see Figure 4.5). The motor has 4 input wires (two wires for each of the two motor coils) which are connected to the Motor Shield board through the terminal block connector. In this prototype configuration, the Arduino Board is powered by the DC power jack (7 - 12V) and a commercial power adaptor. In a commercial collector, an

AC-DC converter will be integrated in the system to supply power to the electronic devices.

A tracking code (attached as Appendix A) was developed for the Arduino Software 1.6.9 to realize the linear tracking of this collector. Table 4.1 shows the initialization parameters used in this custom control software. For instance, the motor resolution has been set to 1.8˚ per step. Thus, the stepper motors go through 637 steps (moving 0.03 cm each step) to cover a working distance of 12 cm. The PC communicates with the Arduino UNO board via an USB link. Once the initialization parameters have been downloaded on the embedded board, the system enters into run mode.

Table 4.1 Control software initialization parameters

Parameter Initialization Value Comment

Date Current date For sun tracking purpose

Resolution 1.8˚ Stepper motor accuracy positioning

Number of Number of steps required to move the receiver 637 steps over a distance of 12 cm

The motor are supposed to be at the rest position Direction Forward ([email protected]) and they move forward till they reach the stop position ([email protected])

When the optimal parameter combination has been identified, the embedded software will run autonomously. A built in microprocessor timer is configured as real time clock. Thus, for this system, at 9.00 am (solar time), the system starts tracking the sun, moving the receivers across the linear track until it reaches the stop time (3.00 pm in our experiments). After stopping, the microprocessor drives backward the motors to the starting position, ready for the next day’s tracking to commence.

A further improvement has been planned to implement a closed loop control system, with an angle position feedback, as depicted in Figure 4.6.

Figure 4.6 Closed Loop Control System with position measurement feedback.

Additionally, a WiFi communication system could be implemented to connect the microprocessor to an external Control Panel, which could be used to display in real time information of solar radiation, internal collector temperature, and ambient temperature, state of storage in the tank, the sun’s position in the sky, and the tracking position.

4.2 Performance Analysis

The performance analysis of this section uses a combined CFD and theoretical comparative study. The theory is presented first (in section 4.2.1) followed by the CFD simulation (in section 4.2.2).

4.2.1 Theoretical thermal performance analysis

During operation, the vacuum packaged absorber tube emits long wavelength thermal radiation to the glass cover and to the chamber. These components then exchange heat with the lens and the outer aluminium frame by radiation and convection. Finally, heat is dissipated to the environment via radiation and convection. It should be noted that a selective surface copper tube is used inside the vacuum packaging to minimize thermal losses from the receiver tube. While convection and conduction heat loss can be neglected in the vacuum space, heat conduction through the mechanical connection between tube and vacuum chamber (see Figure 4.7) must still be considered.

(a) (b) Figure 4.7 Heat loss schematic of the collector (a) modes of internal heat transfer (b) equivalent thermal resistance network

Figure 4.7(a) shows a heat loss schematic of the collector. Based on these heat losses, as thermal resistance network can be formed (depicted in Figure 4.7(b)), to determine the dynamic energy balance, with all equations shown in

Table 4.2. As such, it is clear that the energy balance equations for the lens (subscript ‘L’), aluminium frame (subscript ‘Al’), vacuum chamber (subscript ‘V’), copper receiver tube (subscript ‘r’), and the synthetic organic fluid (Therminol VP-1) (subscript ‘f’) are coupled. Radiative exchanges for these elements are defined by equations (4.7) to (4.9), and the convective heat transfer coefficients can be calculated by standard

correlations given in Table Table 4.3. If the fluid inlet temperature, Tin , is given, Equations (4.1) through (4.7) can be solved as a set of five nonlinear algebraic equations

with five unknown temperatures: lens TL , Al frame TAl , vacuum chamber TV ,

receiver tube surface Tr , and fluid outlet Tout .

Table 4.2 Dynamic energy balance formulas of all the components [161]

Parts Dynamic energy balance formulas

TL 44 mCLp,L   qq L  R,VL  hATThATT 3LV  L  7LL  a ε LLL σA T  T S  , Lens t (4.1)

qLLLL I  A  1 ρα 

TAl 44 mAl C p,Al  q Al  q R,V Al  h 5 A Al T Al  T a ε Al σA Al T Al  T S  , Al frame t (4.2)

qAl I  A Al α Al

TV Vacuum mCV p,G   qq V  R,rV  q k,rV   hATTq 3V V  L  R,VL   q, R,VAl  t (4.3) chamber qIV η o  A V  α V

Tr mr C p,r   q r  q R,rV  q k,rV  M  C p,f  T out  T in  , Cu tube t (4.4) q I  Cr  A  τα r beam r eff

hA  ff MC p ,f Fluid M Cp,f  T out  T in  h f A f e  1 T r T in  [58] (4.5)  

Conduction heat loss due to the mechanical connection between the tube and this vacuum chamber is 푞푘,푟−푉, as given by Equation (4.6).

TTrV qk ,r V  ln Rm 2 R m1  (4.6)

2π kmm W

Where δ Rm1 R m1 and k m are the thickness and thermal conductivity, respectively, of

the thermal insulation mounting ring. Wm is the total width of insulation mounting ring.

The thermal radiation from the receiver tube to vacuum chamber, qR ,r V , and thermal

radiation from vacuum chamber to the Lens and the Al frame, qR ,V Al ,qR,VL , respectively, are given by Equations (4.7) to (4.9):

44 σ TrV T  q  R ,r V 1 ε11 ε (4.7) r v ε A A F ε A r r r r v v v

44 σAv T V T Al  q  R ,V Al 11 (4.8) 1 εε V Al

44 σAv T V T L  q  R,VL 11 (4.9) 1 εε VL

Where Frv , is assumed to be 1 since all radiation leaving the surface of tube strikes the internal surface of the vacuum chamber. The rest of the input parameters for the theoretical model are listed in Table 4.3.

Table 4.3 Parameters of this collector

Symbol Value Unit Symbol Value Unit Symbol Value Unit

2 I 1,000 W/m L푚 0.1 (m) 훼푉 0.05

2 퐼푏푒푎푚 900 W/m 휌퐿 0.1 훼푟 0.95

2 퐴L 2 m 휌푐푝푐 0.95 휀푟 0.10

2 퐴Al 0.6 m 휏푉 0.95 휀퐿 0.8

2 퐴V 1.9 m 훼퐿 0.05 휀푉 0.8

2 퐴r 0.5 m 훼퐴푙 0.4 η표 73%

푅m1 8 Mm 휌퐿 0.1 푇푎 20 ℃

푅m2 11 Mm Cr 4 푇푆 25 ℃

푘푚 0.02 W/mK τα⎸eff 63% 퐶푝,푓 1.5-2.6 kJ/(kg·K)

The external convection heat loss coefficient (lens to ambient) and the internal

convection heat loss coefficient (vacuum chamber to lens) are h 7 and h 3 , respectively.

The variable h f represents the heat transfer coefficient from the tube wall to the synthetic organic fluid. These heat coefficients are given by the equations (4.10)-(4.12). It should be noted that the vapor–liquid critical point of Therminol VP-1 occurs at

around 400 °C and 12 bar, which means that liquid Therminol VP-1 can be used for this temperature range if the tube is pressurized [162]. By considering the variable thermal properties of pressurized Therminol VP-1 in the temperature range of 12-400 ℃, the

heat transfer coefficient from tube wall to fluid, h f , can be calculated via equation (4.12) and thermal properties of pressurized Therminol VP-1.

The ‘Solve’ function in Excel has been used for solving these 5 nonlinear equations. By setting the ambient wind velocity to v=0 m/s or v=10 m/s, and by controlling the inlet

fluid temperature Tin , the values of TL , TAl , TV , Tr and Tout can be obtained.

Table 4.4 Heat transfer coefficient correlations [163]

h 7 h7  2.8 3.0v (4.10)

1.6 1 Nuk 1708 sin 1.8θ  Ra cosθ 3 air  h3  , Nu  1  1.446 1   1      1 , L Ra cosθ  Ra cosθ  5830   h 3 (4.11) 2 gβρ Cp 3 Ra  TVL T L kμ

Nuk hf ,Nu  3.6, Re  2300, f D i (4.12) 0.0668RePr D / L 4 Nu 3.66 2/3 ,2300  Re  10 1 0.04 RePr D / L

The fluid properties, k, β , ρ , Cp and μ used in Table 4.4 represent the thermal conductivity, volume coefficient of expansion, density, heat capacity and viscous coefficient of air. Nu is the Nusselt number, Ra is the Rayleigh number, L is the distance from vacuum chamber to lens (designed to be 5.7cm) and θ is the collector

inclination, which, for Sydney 휃=34°. Re is the Reynolds number, Pr is the Prandtl number and D and L are the diameter and length of copper tube. Based on this analysis, the efficiency of collector increases only slightly when flow becomes turbulent. Therefore, the impact of variable flow is ignored and a low velocity of 0.25 m/s (mass flow rate: 0.0072 kg/s) is selected and analyzed throughout. Thermal efficiency can be

estimated based on the following simple relation, after the outlet temperature Tout is obtained by Equation (4.13) [58]:

mC T  P (4.13) A I L By this analysis method, the efficiency of the collector, and the impact of ambient wind velocity, sun incident angle, concentration ratio, and optical efficiency are evaluated independently. The results of these analyses are given in Figure 4.11 to Figure 4.13.

4.2.2 Collector thermal performance simulation

Computational fluid dynamics (CFD) has also been utilized to analyse fluid flow and heat transfer in the proposed solar collector. The main objective of this simulation is to determine the temperature distribution inside the solar collector under different inlet fluid temperatures and ambient wind speeds. Using this model, thermal efficiency can also be calculated based on equation (4.13).

The CFD model of this study is a three dimensional, multiple-fluid heat transfer model built in ANSYS Icepak, as shown in Figure 4.8. It combines the ANSYS FLUENT solver with robust meshing options to provide a more accurate thermal result than the analytical approach of the last section [105]. The module comprises a full scale prototype with a lens area of 2 m2 and a 10 cm height. In addition, the domain of vacuum space is defined by low pressure air (density: 1.29 × 10−9 kg/m3; thermal conductivity: 0.001 W/m℃).

Figure 4.8 Schematic of the three dimensional model built in ANSYS Icepak

The material, thermal conductivity and surface emissivity of the components are shown in Table 4.5. In order to reduce mesh complexity (and computational time), the shape of component parts have been simplified to an acceptable range. This maintains high quality mesh elements while producing a model which solves with a feasible number of elements for a single desktop computer. A multilevel ‘hexagonal unstructured’ mesh is used in this model since all complex geometric shapes were simplified. Moreover, the geometry is divided into zones: a background air zone, a solid zone, and a fluid zone, with a different mesh density applied in each domain. The total mesh number is >3.3 million elements. The resulting mesh size is approximately 0.5-1.0 mm in the domain of the collector, and a mesh size of 1-10 mm is used in the surrounding air area.

Table 4.5 Materials and characteristics of the model

Thermal Parts Material conductivity Emissivity (W/m℃)

Lens, vacuum glass Epoxy-glass 0.17 0.9

CPC Plastic(surface: 0.25 0.03 polished Al)

Cu Tube Copper (Surface: 380 0.1 Black Chrome and Nickel)

Frame AL extruded 220 0.8

Heat transfer fluid Therminol VP-1 0.07-0.14 [162] /

Vacuum space Low pressure air 0.001 /

According to the optical analysis (section 3.2), the collector optical efficiency is about 63% in the middle of the day. Therefore, under a global horizontal irradiance of 1,000 W/m2 (900 W/m2 direct beam radiation) and an aperture area of 2 m2, 1,134 W of heat is absorbed by the copper tube receiver. Considering the heat absorptance of components, 100.8 W, 64 W and 73 W of power are absorbed by top lens, Al frame and vacuum chamber, respectively. Since this is a multiple-fluid heat transfer model, the ambient air velocity and synthetic organic fluid flow rate determine their respective flow regimes. The air flow is turbulent and the fluid flow is laminar when air velocity is 10 m/s and fluid inlet velocity is 0.25 m/s – i.e. Reynolds numbers of 310,000 (air) and

95 (Therminol VP-1). The ‘Design Optimization’ function in Icepack is particularly suitable for this type of simulation work, since the boundary condition has two variables (air velocity varies from 0m/s to 10m/s and flow inlet temperature varies from 12 ℃, to 400 ℃ with 50℃ increments). The two equation, 푘 − 휀 model was used, and the surface to surface radiation model is selected. In addition, 1.0× 10−3 was selected for continuity and velocity convergence criteria, while 1.0× 10−6 was selected for energy minimum convergence criteria.

The temperature distribution of water inside the receiver tube can be simulated by selecting various inlet fluid temperatures (12 - 400 ℃) and ambient air velocities (0 and 10 m/s). The dependence of simulation result with adopted mesh has been examined by using coarse mesh, normal mesh and improved mesh. The mesh quality is critical for achieving convergence which reflects the stability and accuracy of results. The improved mesh quality with 3.3 million elements is employed in this CFD simulation.

Figure 4.9 Temperature distribution of fluid Therminol VP-1: Tin = 200 ℃, Tout= 255 ℃, 푇̅ = 228 ℃ (v_wind = 10 m/s, Ta = 20 ℃)

Figure 4.9 shows the temperature distribution of fluid when the inlet temperature is 200 ℃. The fluid temperature increased gradually as it progresses along copper tube. Figure 4.10 shows the temperature distribution on collector cross section when inlet

temperature is 200 ℃. The temperature gradient surrounding the tube is caused by the low thermal conductivity (0.001 W/m℃) of low pressure air (0.1 Pa) inside of chamber.

(a)

(b)

Figure 4.10 The temperature distribution on the collector cross section (a) V_wind = 0 m/s; (b) V_wind = 10 m/s

4.2.3 Thermal analysis summary

Collector efficiency and daily efficiency are predicted by both the theoretical and simulation models. In addition, as collector performance testing requires components to survive stagnation temperatures, the no-flow maximum achievable collector temperature, a function of stagnation temperature with an effective optical efficiency and concentration ratio can be evaluated with the theoretical model. The stagnation temperature analysis informs our design and suggestion room for further improvement.

The CFD simulation and theoretical curves for efficiency are compared in Figure 4.11. Theoretically, this collector can deliver heat at temperatures up to 385℃ under global 2 2 horizontal irradiance of 1,000 W/m (퐼푏푒푎푚= 900 W/m ), with the efficiency decreasing steadily from 56% with increased operation temperature. Moreover, it was found that the ambient wind speed has a negligible influence on collector efficiency collector by comparing the result under the condition of 0 m/s and 10 m/s wind speed. The CFD calculation of efficiency is slightly lower than theoretical result, mainly due to the ideal vacuum assumption (air thermal conductivity: 0 W/m℃) used in the theoretical model.

0.8 Theoretical result (V_wind = 0m/s) 0.7 Theoretical result (V_wind) = 10m/s Simulation result (V_wind = 0m/s) Simulation result (V_wind = 10m/s) 0.6 Chromasun MCT 0.5

0.4

0.3 Thermal Efficiency Thermal 0.2

0.1

0 35 85 135 185 235 285 335 385 435 Tm-Ta (℃)

Figure 4.11 Efficiency of the collector with different ambient wind speed (퐼= 1000 W/m2, 2 퐼푏푒푎푚= 900 W/m ) by theoretical and simulation analysis and performance comparison with Chromasun MCT

An efficiency comparison between our collector and the Chromasun MCT collector is also given in Figure 4.11. It should be noted that the maximum operation temperature of Chromasun MCT is 200 ℃, so the efficiency curve from 200 ℃ to 400 ℃ found with a regression analysis. If we compare the efficiency curves, it can be found that the performance of our collector is similar with Chromasun MCT, and that the efficiency is better than Chromasun MCT at high operation temperature region (200 - 400 ℃).

As Figure 4.12 shows, while the theoretical daily variation in efficiency (from 9 am to 3 pm) is 28-44%, the CFD simulation indicates efficiency is 22-41% throughout the day.

0.5

0.4

0.3

0.2 Theoretical(V_wind=0m/s)

Collector efficiency Collector Theoretical (V_wind=10m/s) 0.1 Simulation(V_wind=0m/s) Simulation(V_wind=10m/s) 0 0 5 10 15 20 25 30 35 40 45 Sun incident angle (deg)

Figure 4.12 Collector daily efficiency with variation of sun incident angle (퐼= 1000 W/m2, 2 퐼푏푒푎푚= 900 W/m , outlet temp.: 220 ℃)

The collector’s effective optical efficiency and concentration ratio are the most critical factors for this collector’s efficiency and stagnation temperature. Figure 4.13 was obtained by the theoretical model, and indicates the optimum concentration ratio and optical efficiency.

11 10 300℃ 400℃

9 500℃ 8 600℃ 7 6 5 4

Concentration ratio Concentration 3 2 1 20% 30% 40% 50% 60% 70% 80% 90% 100% Effective optical efficiency

As an example, the point in Figure 4.13 shows the current design with an effective optical efficiency of 63% and a concentration ratio of 4. With efficiency improvements (assuming we could achieve 85%), the stagnation temperature could be further increased to 500 degrees. Alternatively, by increasing the concentration ratio to 5.5, 500 ℃ is also achievable.

4.3 Chapter Summary

The design and analysis of a new low profile collector was presented in this chapter. By integration of a specially designed concentrating lens, a CPC reflector, and vacuum thermal insulation, this design can be easily mounted on rooftops and will supply high temperature thermal energy (100 – 385 ℃) for industrial applications. Based on thermal simulation, the overall thermal collector efficiency will range from 23% - 44% with a fluid average outlet temperature of 220 ℃ on a sunny day (e.g. under 1,000 W/m2 global horizontal irradiance).

In terms of the manufacturing cost, the proposed design is projected to be only slightly greater than a conventional thermal collector when mass produced as it uses common materials. The plastic lens, aluminium frame and CPC reflector can be fabricated by extrusion technology. Compared to other concentrating solar collectors (which have rotational tracking systems), this represents a significant design simplification. Lastly, the optical and thermal analyses show that there is scope to enhance the optical efficiency and concentration ratio, which, if achievable will significantly enhance the overall collector efficiency and/or operation temperature.

However, it was found that it may be possible to relax the total height constraint of this collector to achieve a higher concentration ratio (e.g. to accommodate a wider lens and a longer focal length). This is preferable, however, a compact package (< 15 cm height) was one of the major design factors. If the optics are limited, however, it may be fruitful to explore alternative concentration technologies, as is undertaken in the next chapter.

Chapter 5

5. Innovative Thermal Concentration

Solar energy only provides a flux density of around 1kW/m2, so to attain higher temperatures (>100 °C), optical concentrators are most commonly used, as was discussed in previous chapters. However, as stated in section 2.6, a band-focus Fresnel lens with a height of 10-15 cm is limited to a maximum of 6-8X optical concentration. Thermal concentration, although rarely explicitly designed for, is frequently employed in heat transfer systems. In the context of this work, thermal concentration was explored as an alternative approach to achieve even higher concentration of the solar source in a compact system.

In this chapter thermal concentration is defined as a method by which energy at low-flux density on one surface is delivered with a higher flux density to another surface. The effective thermal concentration ratio is expressed by Equation (5.1):

qQ'' A Cr out out in eff ''  (5.1) q Q A in in out

'' '' where q in and q out are the flux density of the energy receiver and output sides and

Ain and Aout denote the area of energy receiving and output sides, respectively.

The effective thermal concentration ratio also can be expressed by Equation (5.2):

Cr Cr (5.2) eff where  represents the working efficiency of concentration system which is given by

QQout/ in , Cr denotes the geometric concentration ratio and is given by Ain/Aout.

A general schematic of this concept is shown in Figure 5.1. This phenomenon, if designed for, may also provide an effective pathway for achieving high quality heat from low density sources.

Figure 5.1 General schematic of thermal concentration

However, to our knowledge, no systematic attempts have been made to categorize, compare, and estimate the potential values and limits of thermal concentration. In this chapter, thermal concentration methods were compared through an ‘effective concentration ratio’, Cr|eff, which we propose is an ideal metric for these systems as it gives the actual thermal flux concentration. The most readily available configurations for thermal concentration are presented and subsequently analyzed for their fundamental and realistic engineering limits. With a generalized, but comprehensive study of the limitations of the various thermal concentration methods, it is possible to explore this design idea to boost the performance and cost effectiveness of heat collection systems, such as solar thermal collectors, waste heat collection and other heat regeneration systems.

Thermal concentration techniques can be classed as either passive concentrators or active concentrators. Passive concentrators do not require external power while active concentrators do – these are shown in Figure 5.2 . The effectiveness of these techniques is strongly dependent on the mode of heat transfer, which happens via conduction, convection, phase change and radiation. Active concentrators can raise the temperature and thermal flux density from the heat absorber to the relatively smaller output side, while passive concentrators can only increase the thermal flux density. It should be noted that two or more of the above techniques may be utilized simultaneously to provide an enhancement that is usually larger than the individual techniques.

Energy Concentrators

Passive Active

Conduction- Heat Forced Heat pumps Radiative based systems pipes Convection

Thermoelectric Vapor Fan/pumps s compression

Figure 5.2 Categorization of thermal concentration techniques

5.1 Passive concentrators

In the case of solar thermal applications, high thermal conductivity materials, heat pipes and radiative surfaces are able to transfer heat from an absorber and deliver it with higher flux density to a fluid in solar collector applications. The ‘effective thermal concentration’ for these concentration systems is defined as the ratio of the actual thermal flux density from the solar energy relative to the delivered heat flux density on the heat output side.

5.1.1 Thermal conductive concentrator

In a conductive concentrator, a solid material serves as the thermal concentrator. The thermal concentration ratio in these systems is driven by the temperature difference between the solar absorber and the application. Figure 5.3 shows a general schematic of this type of system for a heating application. Heat can be absorbed from the source by the absorber. Part of the heat is lost to the ambient environment and the rest is transferred to the application. To maximize the heat loss resistance, Rloss, the proposed thermal concentrators are packaged in a vacuum environment. It should be noted that the thermal constriction resistance, Rc, is used in the thermal resistance network to describe the situation where heat flows to through a geometrical constriction. Many studies have focused on determining the constriction resistances [164], however, to

simplify the problem in our case, instead of solving for the constriction resistances, we evaluated the rate of energy gained at the contact region using the method developed by Duffie and Beckman [161]. The following analysis is conducted to determine the effective thermal concentration ratio of this general configuration.

(a) (b)

Figure 5.3 Thermal analysis of a conductive thermal concentrator in a vacuum: (a) schematic of conductive thermal concentration; (b) equivalent thermal resistance network

Based on the method developed by Duffie and Beckman [161] and the equivalent thermal resistance network, the rate of energy gained by the absorber can be expressed by Equation (5.3):

'' QWDFTgain  DUT  q in  g  abs  L b  a   (5.3)     Where the fin efficiency, F, can be calculated by Equation (5.4) [161]:

tan-1  m(W-D)/2 U F= ,m= L (5.4)

m(W-D)/2 tabs k

The heat conducted through the absorber to the application can be expressed by Equation (5.5), which equals to the rate of energy gained by the absorber.

Tb  Tu QoutQ gain (5.5) R u The thermal resistance between the thermal concentrator and the application is obtained from Equation (5.6):

t R  TIM (5.6) u kA TIM o where tTIM and kTIM are the thickness and thermal conductivity of the interface material. Arrays of aligned carbon nanotubes (CNT), when used as an interface material can achieve a thermal conductivity of 3,000 W/m-K – these may represent the best possible interface material [165]. Additionally, a thickness of 0.02 mm is recommended for application [166].

An approximate theoretical model can be developed to evaluate UL. To simplify the analysis and improve performance, vacuum insulation can be assumed to surround the absorber of Figure 5.3. It was suggested by others that when the pressure in an evacuated flat plate chamber decreases to below 10-2 Pa [41, 81], the convection and conduction heat losses are completely suppressed. By neglecting the convective heat transfer coefficient from the absorber to the glass, hconv,abs/g, the predominant heat loss is radiative heat transfer from the absorber to the glass cover.

With these assumptions, the total heat loss coefficient, UL, can be expressed by Equation (5.7):

1 U  L 11  (5.7) h h+ h rad,abs/g rad,g/a conv,g/a The radiation heat transfer coefficient from the absorber plate to the upper portion of the glass cover can be described by Equation (5.8):

22 σ Tabs T g T abs T g  h  rad,abs/g 11 (5.8) 1 εε abs g

Assuming a good selective surface like TiNOx with a temperature dependent absorbance 0.95, and an thermal emissivity which can be expressed as a function of temperature using Equation (5.9) [41]:

ε 0.0843  3.95  105 T  5.331  10 7 T 2 (5.9) abs abs abs

The radiation heat transfer coefficient from the upper portion of the glass cover to the ambient environment is expressed by Equation (5.10):

h ε σ T22  T T  T (5.10) rad,g/a g g a g a  The convection heat transfer coefficient from the upper portion of the glass cover to the ambient environment is given by Equation (5.11):

h 2.8 3.0v (5.11) conv,g/a where v is the ambient wind speed, which is assumed to be 0 in this case. To find UL, it is necessary to know the glass cover temperature, (Tg), which can be found by applying an iterative solution to the above linearized heat transfer coefficients and the following heat balance:

h T T  qα  (h  h )  T  T (5.12) rad,abs/g abs g ing rad,g/a conv,g/a g a  The heat flux density at the output side, and thus the effective thermal concentration ratio, can be solved via Equation (5.1). Several factors define the effective thermal concentration ratio. Salient factors are the overall heat loss coefficient, the material and thickness of the absorbers, the geometric concentration ratio, Cr, the application resistance, Ru, and the application temperature, Tu. A higher thermal conductivity and a larger absorber thickness result in higher thermal concentration ratio. Copper and aluminum are commonly used as solar absorber materials because they are good heat conductors while remaining cost effective. Carbon nanotube (CNT) absorbers are also considered in this study since they represent the best possible conductors, and therefore set a practical engineering limit to the thermal concentration for conductive systems. A thickness of 0.5 mm for the copper absorber was selected and analyzed here due to market availability, cost, efficiency and geometry optimization [167, 168], while a 0.5 mm thickness carbon nanotube absorber with a thermal conductivity of 3,000 W/m-K was assumed to be available. The parameters used in the analysis are shown in Table 1.

Figure 5.4 shows the effective thermal concentration ratio, Cr|eff, and working efficiency, , as they vary with the geometric concentration ratio, Cr, and application temperature, Tu. It can be seen that an 11-50X effective thermal concentration with an efficiency of 2%-80% can be reached when a copper conductive concentrator is used,

the thermal concentration ratio is limited to the range of 36-135X with an efficiency of 5%-83% for a CNT thermal conductive concentrator.

(a)

(b) Figure 5.4 The effective thermal concentration and working efficiency of conductive concentrators: (a) copper concentrator, kcu=380 W/m-K; (b) CNT concentrator, kCNT =3,000 W/m-K For solar heating applications, a higher geometric concentration ratio, Cr, indicates less material is required for transferring concentrated heat to the output region. Thus, having

a higher working efficiency and a larger absorber area (larger value of Cr) is usually preferable.

It should be noted that this analysis does not take costs into account. A more detailed cost analysis is required to find the exact optimum point for the specific application based on these results. For instance, assuming an ambient temperature of 25 °C, the working efficiency decreases slightly from 84% to 82% (75 °C application temperature), 71% to 66% (175 °C application temperature), and 35% to 30% (275 °C application temperature) when the Cr increases from 1 to 30 for a copper conductive concentrator. Consequently, an optimum Cr range of 15-30 is recommended according to the performance and economic factors (e.g. less expensive materials or less heat delivery area) for a detailed design. Along the same lines, an optimum Cr ratio of 40-60 is recommended for CNT conductive concentrator, but this may come with a cost trade off over other interface materials.

5.1.2 Heat pipe thermal concentrator

Heat pipes are characterized by a high heat transfer performance as they take advantage of the latent heat of phase change. In heat pipes, heat is transferred as latent heat energy by evaporating a working fluid over a heating zone and condensing the vapor in a cooling zone. Circulation is completed by return flow of the condensate to the heating zone through a capillary structure which lines the inner wall of the container. The working fluid is evaporated along the absorber area, and then transported to the smaller manifold area where condensation takes place. Heat transfer coefficients in the evaporator and condenser zones are in the range 103-105 W/m2K leading to heat pipe thermal resistances of 0.01–0.03 K/W [169-171].

A sintered powder metal wick structure was used in this analysis due to its better performance as compared with mesh and grooved wick structures [172]. The heat pipe thermal resistance may be mathematically modelled by a number of thermal resistances with the total thermal resistance given by Equation (5.13) [111]:

R R  R  R  R  R  R  R (5.13) hp ev,w ev,wick ev,i v cond,i cond,wick cond,wall

Table 5.1 Thermal concentration parameters

Parameter Symbol Values Parameter Symbol Values

Transmittance of glass  0.9 Liquid density (kg/m3) ρ 965 g l

Absorptance of glass  0.05 Vapour density (kg/m3) ρ 0.05 g v

Viscosity of the liquid Absorptance of absorber  0.95 u 0.0003 abs (kg/m-s) l

Thickness of thermal Latent heat of * t 0.02 h 2300 interface material (mm) TIM vaporisation (kJ/kg) fg

Thickness of absorber Effective capillary T 0.5 r 0.25 (mm) abs radius (mm) o

Thermal conductivity of Gravitational thermal interface material k 3,000 g 9.8 TIM acceleration (m/s2) W/m-K

The difference between External to internal d / d vapor saturation diameter ratio of pipe ex,p i,p 1.05 T 20 temperature to solid i wall surface (K)

External to internal dex,wick / d i,wick Vapour frictional 1.11 -1 -3 Fv 0.014 diameter ratio of wick coefficient (N.W .m )

External to internal d / d View factor between diameter ratio of interface ex,TIM i,TIM1.01 F 1 input and output plates 12 material

Thermal conductivity of Emissivity of input heat pipe material K 380 ε 0.95 cu plates in (W/m-K)

Thermal conductivity of Emissivity of output K 0.623 ε 0.95 liquid (W/m-K) l plates o

Heat exchanger overall Wick porosity ε 0.7 heat transfer coefficient U 7.5 wick ex (kW/m2K) where Rev,wall and Rev,wick are the thermal resistances across the thickness of the container wall and wick thickness, Rev,i is the thermal resistance that occurs at the vapor–liquid interfaces of the evaporator, Rv represents the vapor flow resistance, Rcond,i is the thermal resistance associated with the condensing process inside the heat pipe,

and Rcond,wick and Rcond,wall are the thermal resistances across the thickness of the wick and the heat pipe wall at the condensation side.

All the thermal resistances along the heat pipe can be determined using the formulations presented by Chi (1976) [173] and Azad (2008) [111]. Various parameters affect the heat transfer performance of the heat pipes, including the adiabatic length ratio, the working fluid fill ratio, and the properties of the working fluid. The parameters used in this analysis model are provided in

Table 5.1.

The thermal resistance of the thermal interface material between the condenser section and the application is obtained from Equation (5.14):

ln dex,TIM / d i,TIM  R  (5.14) u 2πk L TIM cond In practice, the heat from the increased heat pipe receiver area may exceed the heat transfer limitation of the heat pipe [174]. When using heat pipes for solar heating, only the entrainment limitation and the dry out limitation should be considered [170, 175, 176]. When the application involves the transfer of large heat loads over a long distance, the thermal performance of the heat pipe is substantially affected by entrainment losses. This negative effect can be reduced by combining heat pipes with a flat plate absorber as the thermal concentrator, as shown in Figure 5.5. Then, a relatively large absorber area can be achieved.

The concentrator exchanges heat from a large evaporator area to the much smaller condenser area of the heat pipe condenser, where the largest heat flux density exists. The thermal concentration ratio, then, is determined by the ratio of the absorber to condenser area and the efficiency of the system. Similar to the analysis in section 5.1.1, an energy balance equation for the vacuum glass cover can be defined by Equation (5.15).

h A  TTqA  '' α  (h  h)ATT    (5.15) rad,hp/g abs abs g in gg rad,g/a conv,g/a g  g a 

The rate of energy gained by the absorber, Qgain , can be modelled using Equation (5.1),as defined in section 5.1.1.

(a)

(b) Figure 5.5 Thermal analysis of a heat pipe with a flat plate absorber concentrator in a vacuum: (a) schematic of heat pipe thermal concentration; (b) equivalent thermal resistance network Heat conductance flux through the absorber to the application can be obtained from Equation (5.16):

TThp u QQ (5.16) gain out RR hp u where the heat pipe resistance, Rhp, can be calculated by Equation (5.13). The UL value was evaluated using the same approach detailed in section 2.2.1. A vacuum glass pipe -2 with a vacuum level of 10 Pa was employed in the model and TiNOx was chosen as the selective coating on the surface of the heat pipe. Consequently, the heat loss from absorber is minimized.

The effective thermal concentration ratio is defined by:

Q/A CR  out cond (5.17) eff Q/A in abs

(a)

(b) Figure 5.6 The effective thermal concentration ratio and working efficiency of heat pipe with absorber concentrator: (a) heat pipe with Cu absorber in a vacuum; (b) heat pipe with CNT absorber in a vacuum Azad (2008) [111] suggested that a ratio of evaporator length to condenser length of 8.25 gives a design which is best able to absorb and deliver useful heat. Thus, the area ratio of the absorber and condenser can be expressed by Equation (5.18):

A w abs  8.25 (5.18) A πD cond cond

If we define the total length of the collector, then by varying the width of the absorber, w, and the condenser temperature, Tu, it is possible to determine the thermal concentration ratio of this concentrator.

Figure 5.6 shows that a 30-95X thermal concentration ratio is achievable when the condensation to ambient temperature difference rises from 50 to 250 °C. For a heat pipe with a Cu absorber, concentration ratios of 50-175X can be achieved by combining a heat pipe and CNT absorber. For solar thermal applications, a Cr range from 50 to 100 should be possible for a heat pipe with flat plate absorber collector, but a Cr range from 50-150 is recommended for a heat pipe with CNT absorber collector.

5.1.3 Radiative concentrators

Radiative concentrators operate on the principle of collecting radiation over a large area, converting it into thermal radiation and then directing the radiation into a relatively small output target. As Figure 5.7 shows, parallel plates placed in an enclosure can be used to gather incoming radiation on the upper faces and emit radiation from the lower surface such that the energy can be concentrated and collected for the application.

In principle, if the heat input area is sufficiently large relative to the output area, the output area will be of a much higher heat flux, despite radiation and other heat losses internal to the system. The effective thermal concentration ratio depends on the temperature of the input and output plates, Tin and To, the emissivity of these surfaces,

εin and εo, the surface areas, Ain and Ao, and the reflectance of side wall.

The rate of energy gained by the absorber can be expressed by Equation (5.19) [161]:

'' Qgain Ain q ina g abs  ULi Tn  T  (5.19) 

(a) (b) Figure 5.7 Thermal analysis of a radiative concentrator in a vacuum: (a) schematic of radiative thermal concentration; (b) equivalent thermal resistance network

The UL value was evaluated by the same method described in the previous section. A vacuum chamber was assumed for this model, TiNOx was chosen as the selective coating on the surface of the absorber, and the enclosure was assumed to be adiabatic.

Assuming perfectly reflecting side wall, the view factor F12 was assumed to be 1, meaning all the radiation leaving the input surface is received by the smaller output surface. The emissivity of the absorber bottom surface and the output surface was assumed to be 0.95 in this analysis. The radiation heat transfer for these two surfaces can be determined using Equation (5.20):

44 σAin T in T o  QQ  out gain 1A1 ε  (5.20)  o in ε ε A in o o Similarly, heat conductance through the output area to the application can be obtained by Equation (5.21):

TT Q  o u (5.21) out R u The effective thermal concentration ratio can be expressed by Equation (5.1).

Figure 5.8 The effective thermal concentration of radiative thermal concentrators Figure 5.8 shows that the energy can be concentrated by around 122 times for a radiative concentrator with a 50 °C temperature difference between the applications and ambient.

Table 5.2 Analysis results summary for passive thermal concentrators Achievable Cr|eff Recommended Cr Working with with efficiency with Categorization ΔTTCT= ua ΔTTCT= ua ΔTTCT= ua 50 150 250 50 150 250 50 150 250 Copper absorber 0.83- 0.69- 0.33- 50 32 12 20-30 Thermal (currently 0.81 0.66 0.30 conductive available) CNT absorber 0.83- 0.70- 0.34- 136 95 34 40-60 (future) 0.82 0.68 0.32 Heat pipe with Copper 0.79- 0.65- 0.36- absorber 92 65 30 50-100 0.71 0.53 0.26 (currently

Passive Heat pipes available) Heat pipe with CNT 0.81- 0.67- 0.38- 176 115 52 50-150 absorber 0.71 0.53 0.26 (future) Coated absorber 0.7- 0.60- 0.36- Radiative 123 101 59 15-40 (currently 0.62 0.53 0.32 available) With increasing application temperature, the effective thermal concentration ratio decreases considerably. That is, only a 59X effective thermal concentration ratio is

possible when the difference between the application and ambient temperatures is 250 °C due to losses. A Cr range of 15-40 is recommended for solar heating applications based on these results. Table 5.2 summarizes the analysis results for the passive thermal concentrators examined here.

5.2 Active thermal concentrators

Forced convection and heap pumps are typical active techniques which can transport heat very efficiently from a solar absorber to a heat exchanger with high flux density. Heat pumps are of particular interest because they can transfer heat in the opposite direction of spontaneous heat flow by employing external power [177]. Subsequently, the heat flux can be increased on the output side of a system. Heat pump systems have been widely used in domestic hot water and space heating applications [177]. The integration of heat pumps with solar technology presents an opportunity significantly enhance ‘free’ heat income from solar energy [178].

5.2.1 Forced convection thermal concentrators

Forced convection thermal concentrators operate on the principle of collecting heat over a large area and directing it into a relatively small output heat exchanger. As shown in Figure 5.9, it typically consists of an absorber, a pump and a heat exchanger. The pump is used not only to enhance the convection heat exchange coefficient at the absorber and output heat exchanger, but is also able to circulate the working fluid in the system. Assuming the consumption of external power input is negligible and the pipe loops are adiabatic, and the heat loss can be minimized, the heat gained at absorber side can be concentrated and collected.

(a)

(b) Figure 5.9 Thermal analysis of a forced convection thermal concentrator: (a) schematic of forced convection thermal concentration; (b) equivalent thermal resistance network The rate of energy gained by the absorber can be expressed by Equation (5.22) [161]:

'' Qgain A in F q ingabs  U L  Tm Ta  (5.22)  Where F’is the collector efficiency factor for a conventional absorber which consists of pipes and can be evaluated using Equation (5.23):

1/ U F  L 11 (5.23) w   UWDFDL   πdh f   The collector efficiency factor, F , can be enhanced by geometric parameters of the pipes/channels and the heat transfer coefficient inside the pipes/channels. The collector efficiency factor can be enhanced by either decreasing the tube space, W, to be equal to the tube diameter, D, or employing a higher heat transfer coefficient, hf. Assuming a 0.99 collector efficiency factor is achievable in this case, the required heat transfer coefficient, hf, can be determined from Equations (5.23) and (5.24).

Assuming the loop between the absorber and heat exchanger is adiabatic, the rate of heat transfer at the absorber side is equal to the heat released to the application and can be expressed by Equation (5.24):

Q h A T  T  U A T  T (5.24) gain f in abs m ex out  mu where Uex denotes the overall heat transfer coefficient of the output heat exchanger. To transfer a higher density of heat flux through the heat exchanger, a compact heat exchanger which has comparatively large area density (the ratio of heat transfer surface to heat exchanger volume) needs to be employed. Their large area density, which translates to a small hydraulic diameter for fluid flow, results in a higher efficiency than a conventional shell-and-tube heat exchanger [179]. A plate heat exchanger was selected in this model which has an overall heat transfer coefficient, Uex, of around 7.5 kW/m2K – an assumption which is conservatively 3 to 4 times higher than shell-and-tube exchangers [179]. Note either heat transfer coefficient inside receiver tubes at the absorber side or heat exchanger at the collected side can be enhanced dramatically by employing some advanced technologies, for instance, For instance, the heat transfer coefficient inside tubes of 339 W/m2K for a conventional absorber with copper pipe can be increased to 1517.8 W/m2K if a micro channel based absorber employed according to the research result of Mansour [180], and even higher heat transfer coefficient up to 20 kW/m2K can be achieved by using impinging jets [181]. Additionally, micro-channel and multi-phase heat exchangers which can reach an overall heat transfer coefficient of 25 kW/m2K [182-184]. However, the extra pumping power is significant for these high heat transfer rates technologies, so these technologies are not studied in this case. The effective thermal concentration ratio can be expressed by Equation (5.1).

Figure 5.10 The effective thermal concentration of forced convection thermal concentrators, the 2 output heat exchanger, Uex = 7.5 kW/m K; Figure 5.10 depicts the change in effective thermal concentration ratio with geometric concentration ratio for different application temperatures. The analysis results reveal that the thermal flux density can be concentrated to around 2,200X using a forced convection concentrator with a 50 °C difference between the output to ambient temperatures (when the geometric concentration ratio is 25,000X). As the application temperature increases, the effective thermal concentration ratio decreases extensively. On a 70X effective thermal concentration ratio could be attained when the difference between the application and ambient temperatures was 350 °C.

5.2.2 Heat pump thermal concentrators

Solar-assisted vapor compression and thermoelectric heat pump systems may also be considered as a means to create thermal concentration. Compared to other technologies, the greatest advantage of heat pumps is that they can produce a temperature difference between the cold/heat-sink device side and the hot/heat exchanger side. In this case, the solar receiver could be employed as the ‘cold’ side of the device. Heat can then be pumped out of the absorber while at the same time increasing the temperature. This is advantageous since it reduces heat loss from the absorber, if the absorber kept at a lower temperature. However, raising the temperature require additional electrical power and

leads to lower COP. Consequently, the design of these systems is somewhat non-intuitive.

5.2.2.1 Vapor compression thermal concentrator

A thermal concentrator based on a vapor compression consists of a solar thermal collector, a refrigeration compressor, an expansion valve, and a heat exchanger. The system operates on an electrically driven vapor compression refrigeration cycle, as it pumps energy from the solar collector to intensify solar heat to produce hot water [177].

R-134a, a refrigerant with a low boiling point, was employed as the working fluid in the cycle shown schematically in Figure 5.11. In the operating cycle of this solar-assisted vapor compression refrigeration system, the solar collector operates as a heat-sink device, absorbing heat from solar irradiance at the evaporating temperature of refrigerant. The refrigerant leaving the evaporator at node 1 is compressed by the compressor up to node 2 and then passes through the heat exchanger. The heat

exchanger then transfers heat at the rate of Qcond to the application.

Figure 5.11 Schematic of solar assisted heat pump system A mathematical thermal analysis was carried out incorporating collector data, heat pump and load data. A MATLAB code (attached as Appendix B) was developed for calculation. Initially, the value of the collector operating temperature, Tev (which is equal to the evaporating saturated temperature of refrigerant T1), was assumed. The

thermodynamic properties of the refrigerant (R134a) were obtained from an in-house-developed software using data from the ASHARE Handbook [185].

During the operating cycle of this vapor compression refrigeration system the solar collector operates as a heat-sink device, absorbing heat the from heat source at the evaporating saturated temperature of refrigerant. The energy received by the solar thermal collector, which is given by Equation (5.25), can be assumed to equal the energy absorbed by the refrigerant, which is denoted by Equation (5.26):

'' (5.25) Qgain =Ain q in g abs  U L T1 T a   

Where T1 is the refrigerant evaporating temperature, assuming it is equal to the operating temperature of solar collector.

Q m h h (5.26) hs r 1 4

Where h1 and h 4 denote the enthalpy of refrigerant at node 1 and 4 of the loop, mr represents mass flow rate of refrigerant.

Denoting the compressor pressure ratio as Rp, the pressure of superheated vapor can be obtained by Equation (5.27):

PPR (5.27) 2 1 P

Assuming that the compressor work is an isentropic process, the entropy, S2, is equal to

S1, and the enthalpy of isentropic process at exit, h2s, can be found from the thermodynamic tables. The enthalpy of the actual process at the exit, h2, can be calculated using Equation (5.28), where η c is the isentropic efficiency of a compressor with the value assumed to be 0.8 [186]. As h2 and P2 are available, T2 and S2 of the superheated refrigerant can be determined.

hh η  2s 1 (5.28) c hh 21 The compressor power is given by Equation (5.29) [187]:

W m h h (5.29) comp r 2 1 It is anticipated that the superheated refrigerant is condensed to saturated liquid. Assuming the pressure drop along the condenser is 5% [188], the pressure of the

saturated liquid, P3, equals 95% of the pressure of the superheated vapor, P2, and the values of T3, S3 and H3 can be found from thermodynamic tables. The thermal energy delivered by the condenser side is given by Equation (5.30):

Q m h h (5.30) cond r 2 3 For a well-insulated plate heat exchanger, the rejected heat removed from the condenser is given by Equation (5.31):

Q mCT T UA LMTD (5.31) cond w p water,out water,inex cond

Where the log mean temperature difference, LMTD, is defined by Equation (5.32):

(T2 Twater,in )  T 3  T water,out  LMTD  (5.32)  ln T2 T water,in / T 3 Twater,out   As the thermal resistance of the heat transfer wall and the fouling factor are negligible, the overall heat transfer coefficient of the plate heat exchanger, Uex, is expressed by Equation (5.33):

1 U  ex 11  (5.33) hh r w where hr represents the condensation heat transfer coefficient between the warm

medium (R-134a) and the heat transfer surface, while h w is the heat transfer coefficient between the heat transfer surface and the cold medium (water). A maximum condensation heat transfer coefficient value of 5 kW/m2K was selected by assessing condensation heat transfer correlations in the modelling of plate heat exchangers [189].

Assuming that saturated liquid expansion occurs isothermally, the enthalpy, h4 (mixed liquid and vapor), is equal to h3. Additionally, P4 equals 1.05*P1 upon assuming a 5% pressure drop along the evaporator/heat-sink device [188]. Therefore, T4 and S4 can be evaluated from thermodynamic tables. As the refrigerant temperature and enthalpy at

each state can be solved, and the refrigerant mass flow rate, mr , can be evaluated, the

values of Qcond and Wcomp can be determined by mathematical modelling.

The COP of the Rankine refrigerant system is defined as the heat gain at the condenser

in relation to the power of the electrical motor ( Wcomp ) running the compressor, and can be expressed by Equation (5.34):

Q COP  cond (5.34) W comp While the COP of the Carnot refrigerant system is defined as Equation (5.35):

T COP  L (5.35) TT HL

Where TL denotes the refrigerant mean temperature at the evaporator, and is equal to

T1, TH is the refrigerant mean temperature at the condenser, which can be assumed using Equation (5.36):

TT23 TH  (5.36) 2 The system working efficiency can be calculated by Equation (5.37):

Q Eff cond (5.37) QW in comp Therefore, the effective thermal concentration ratio can be expressed by Equation (5.38):

Q/A cond cond (5.38) Cth eff  G The effective thermal concentration ratio and efficiency of this system was plotted against the heat-sink device temperature and geometric concentration ratio as shown in Figure 5.12(a). From Figure 5.12(a) 26 to 275 times thermal concentration ratio is achievable for the solar assisted vapor compression system. The drawback of a Rankine refrigeration cycle system is that the result does not represent the maximum COP values. Thus, a Carnot refrigeration cycle has been employed for the ultimate limit. The COP of both Rankine and Carnot refrigeration cycle systems is plotted with variable evaporating temperature and compression ratio in Figure 5.12(b).

(a)

(b) Figure 5.12 The effective thermal concentration, system efficiency and COP of a vapor compression refrigeration cycle: (a) Cr|eff and  variation with evaporating temperature and compressor pressure ratio; (b) COP comparison between Rankine and Carnot refrigeration cycle systems with varied evaporating temperature and compressor pressure ratio For a solar thermal application, the highest attainable condenser temperature of 112 °C

(superheated refrigerant temperature, T2) was observed for this vapour compression system, allowing it to deliver hot water at a temperature of around 110 °C. An evaporating/heat-sink device temperature from 25-50 °C coupled with a compressor compression ratio of 2-5 times is recommended. Consequently, a system working

efficiency of 80-95% with a 5-10 of COP can be considered as the maximum achievable system.

5.2.2.2 Thermoelectric heat pump thermal concentrator

A thermal concentrator employing a thermoelectric cooler (TEC) device and heat absorber is the final system considered in this chapter. Its schematic is given in Figure 5.13.

Figure 5.13 Schematic of thermoelectric heat pump thermal concentrator The system is capable of absorbing heat from the cold side and releasing heat at the hot side when an electric current is passed through the device [187]. Consequently, the heat flux density at the hot side can be significantly increased if the cold side of TEC is in contact with a solar absorber. The net heat concentrated at the hot end is the sum of the net heat absorbed at the cold end plus the applied electric power [190]. As the solar absorber is attached to the cold side of the TEC, assuming that the temperature at the base of absorber adjacent to the cold side of the TEC, Tb, is equal to the cold side temperature of the TEC, TC, the rate of energy gained by the absorber can be described by Equation (5.39) [161]:

'' Qgain WDFD)q    in  UTT L C  a    Q C (5.39)     Input electricity power is given by Equation (5.40)

2 Qnp αI TH  TC   IR (5.40)  Using a similar process as was detailed in section 2.1.1, the heat gained by the absorber can be determined. Assuming that the TEC is designed to pump all the absorbed heat

from the absorber to application, the amount of heat released from the hot side of the TEC can be determined from Equation (5.41):

1 2 Q Q  Q  n αIT  IR  k T  T (5.41) HCPHHC  2 The COP can be obtained by Equation (5.42), which is defined as the ratio between the thermal power pumped and the electrical power supplied [191]:

Q COP  C (5.42) Q P The optimum COP can be found using Equation (5.43):

T 1 ZT H TTm COP  CC (5.43) TT 1 ZT 1 HCm 

Where Tm is the mean temperature of TC and TH, and ZTm denotes the figure of merit.

Bi2Te3 thermoelectric coolers widely used and has a comparatively higher ZT between 0.8-1 at low to medium temperatures (-40 - 200°C) [190, 192] due to its high Seebeck coefficient, low electrical resistivity, and relatively low thermal conductivity [190].

The thermal concentration ratio between the hot side of the TEC and the solar absorber can be expressed by Equation (5.44):

QQ qout CpAin Cth eff   (5.44) q Q A inin o The system working efficiency can be calculated using Equation (5.45):

Q Eff cond (5.45) QW in p As Figure 5.14 demonstrates, if a copper absorber and an output heat exchanger (with a high overall heat transfer coefficient of 7.5 kW/m2K) are used, thermal concentration limits of 20-250X can be achieved by the TEC device when Tu-Ta rises from 50 to 250 °C and with a 10-70 °C temperature difference between the hot and cold sides. It is worth noting that a relatively high COP, ranging from 5 to 8.2, is found for a 10 °C

temperature difference between the hot and cold sides of the TEC, while a lower COP (0.25-0.7) is observed for a 70 °C temperature difference. The lower COP means a higher percentage of input electrical power is required.

(a)

(b) Figure 5.14 The effective thermal concentration, system efficiency and COP of an available thermoelectric heat pump thermal concentrator: (a) ZTm (figure of merit) = 1, TH-TC= 10 °C, kCu absorber=380 W/m-K, output heat exchanger Uex = 7.5 kW/m2K; (b) ZT = 1, TH-TC= 70 °C, kCu absorber=380 W/m-K, output heat exchanger Uex = 7.5 kW/m2K When using currently available materials to achieve the desired application temperature

(Tu-Ta= 50-250 °C), if a 20-60 times geometric concentration ratio, Cr, is employed, a 91

system efficiency of 0.86-0.3 and COP of 5-8 are obtainable for TH-TC= 10 °C, while a system efficiency of 0.96-0.66 and COP of 0.2-0.7 would be achievable for TH-TC= 70 °C. However, the lower COP means less "free energy" can be obtained, so it would be more economic to increase the COP by maintaining a lower temperature difference between the hot and cold sides, even though it will derate the total working efficiency.

(a)

(b) Figure 5.15 The effective thermal concentration, system efficiency and COP of advanced thermoelectric heat pump thermal concentrator: (a) ZT = 2, TH-TC= 10 °C, kCNT Absorber =3,000 W/m-K, output heat exchanger Uex = 7.5 kW/m2K; (b) ZT = 2, TH-TC= 70 °C, kCNT Absorber =3,000 W/m-K, Uex = 7.5 kW/m2K

Recent studies have suggested significantly increasing the figure of merit is possible [190]. Assuming a maximum ZT value of 2, and a CNT absorber is employed, a 58-340X thermal concentration ratio can be achieved for a 10-70 °C temperature difference between the hot and cold sides of the TEC (Figure 5.15). In this case, while a higher COP of 8-13 is found for a 10 °C temperature difference between the hot and cold sides of the TEC, a lower COP (0.63-1.4) is evident when the temperature difference between the hot and cold sides of the TEC is 70 °C.

When using active technologies to achieve the desired application temperature (Tu-Ta= 50-250 °C), we are able to predict the limits/possible range of effective thermal concentration ratio and efficiency. The analysis results are given in Table 5.3.

Table 5.3 Analysis results summary for active thermal concentrators

Achievable Recommended Working Efficiency with Cr|eff with Cr with

ΔTTCT= ua Categorization ΔTTCT= ua ΔTTCT= ua

50/150/250/350 50/150/250/350 50/150/250/350

U = 7.5 Forced ex 0.05 kW/m2K 2,20 1,50 77 7 0.81- 0.69- 0.45- convecti 200-400 -0.0 (currently 0 0 8 0 0.79 0.66 0.41 on available) 4 Rankine Cr=150-250 Cycle 80-90% (T -T =50-87 °C) R =2-5 u a Vapor (currently p COP=3-6.5 (Tu-Ta=50-87 °C) compres available) 275 (Tu-Ta=50-87 °C) sion Cr=150-250 Carnot Cycle 80-90% (T -T =50-87 °C) R =2-5 u a (future) p COP=4-12 (Tu-Ta=50-87 °C)

0.86-0 0.75-0 0.5-

Copper .76 .6 0.3 absorber, ZT (COP (COP (COP Active = 1, U = 7.5 25 =5) =6) =8) ex 70 130 / 20-60 / kW/m2K 0 0.96-0 0.92-0 0.81-0 (currently .94 .85 .66 available) (COP (COP (COP TEC =0.2) =0.5) =0.7) heat / 0.85-0 0.75-0 0.5- pump .82 .68 0.42 CNT (COP (COP (COP absorber, ZT 34 =8) =11) =13) = 2, U = 7.5 130 220 40-100 ex 0 0.94-0 0.89-0 0.76-0 kW/m2K .92 .85 .68 (future) (COP (COP (COP =0.6) =1) =1.3)

5.3 Summary

In this chapter, different methodologies for thermal concentration were examined and reported in detail. The generality and suitability of the developed approach was demonstrated for a variety of technologies. Passive concentrators were only capable of increasing the heat flux density, while active concentrators can also raise the temperature while concentrating the heat flux density. These passive and active concentrators were found to have effective thermal concentration ratios in the range of 1 to 2,200 times (based on the change in heat flux density, W/m2) for application temperatures over the range 75-375 °C. With increasing application temperature, the effective energy concentration ratio decreases considerably.

Among the passive techniques considered, the CNT conductive concentrator can be regarded as the most effective passive method to increase the effective thermal concentration ratio. Combining the heat pipe with a CNT conductor gave an even higher thermal concentration ratio of 50 to 175 times, whereby a 0.81-0.27 efficiency would be achievable (Tu-Ta=50-250°C) with a recommended geometric concentration ratio of 50-150.

Among the active thermal concentration methods, a forced convection concentrator was able to concentrate the heat flux density up to 2,200 times (Tu-Ta=50-350°C), while heat pump thermal concentrators were only able to achieve an effective thermal concentration ratio up to 340. However, heat pump thermal concentrators can provide a higher working efficiency. Despite a vapor compression system being a viable active concentrator, it would require a bulky compressor, condenser and all the related piping. Thus, a thermoelectric heat pump combined with a good conductive thermal concentrator may provide the best solution.

Overall, the proposed approach is beneficial for design and optimization in heating applications where high flux heat energy is required. Moreover, the solar absorber was found to play a critical role in both the optical and the thermal concentration system. Optimizing the concentrating systems is futile if the incoming solar radiation is not effectively absorbed, transferred, and preserved by the receiver. While high solar absorbance is relatively easy to achieve (e.g. by using a ‘black’ surface or fluid), achieving both a high heat transfer to the working fluid and low emissivity/convection 94

heat loss coefficients is more difficult and requires systematic investigation and optimization. Thus, the next chapter will zoom in on the solar absorber’s design and some potential options for innovation.

Chapter 6

6. Innovative solar absorbers

Conventional solar thermal collectors use ‘surface-based’ absorbers – e.g. an opaque metal surface coated with a selective thin film to efficiently convert solar radiation into thermal energy [1]. Although this configuration seems to be straight-forward, selective films (e.g. TiNOx) consist of multiple layers of vapour deposited, high purity materials [112]. In addition, heat from the outer surface must first conduct through the solid plate/tube and then be transferred into a working fluid. Depending on the thermal resistance between the outer absorber surface and the working fluid, there will be some temperature drop to create this heat transfer, which effectively lowers the potential working fluid temperature. Additionally, from a heat loss perspective, having the highest temperature on the outer absorber surface is not ideal since it ultimately drives heat loss into the surroundings [114]. Several researchers have proposed that these issues can be addressed through alternative, volumetric absorbers where the working fluids themselves absorb solar energy [114-118]. This so-called ‘direct absorption collector (DASC)’ potentially benefits from suppressed heat loss due to a lower outer absorber surface temperature relative to the inner working fluid [115].

However, as stated in section 2.4.2, only a select few high temperature DASC (operating temperature >100 °C) have been investigated, and these were either theoretic studies [114, 193] or indoor laboratory-scale tests. Benefits (e.g. efficiency enhancement) have been reported by these researchers, however, some challenges have also been revealed, especially when the operating temperature exceeds 100 °C. A major issue is high thermal radiative loss from nanofluid receivers (e.g. glass pipe/container) [114, 194]. The other problem is the stability of the resulting dispersion decreases with increasing temperature [194].

As a means to effectively deliver heat to industrial applications, this chapter explores the potential of DASC for use in the receiver/absorber design. As such, an ITO coated nanofluid receiver with vacuum insulation was developed for the first time here (i.e. to reduce the first challenge of radiative heat loss). Second, the proposed nanofluid

receiver comprises a glass tube containing chemically functionalized multi-walled carbon nanotubes (MWCNTs), which can remain stable to 250 °C [121, 131, 194] (i.e. to overcome high temperature stability issues). A prototype is tested ‘on-sun’ and used to validate a computational fluid dynamics (CFD) model to develop a fundamental understanding of the loss mechanisms and performance of both surface and volumetric absorbers in the designs described in previous Chapters.

6.1 Absorber design and analysis

6.1.1 Optical design

The absorbers designs were developed from the results of the previous optical designs (chapter 3), as is noted in the schematic drawing of Figure 6.1. Thus, the tube absorber tube has a diameter of 10 mm and is centered at the bottom of the CPC. The difference between a conventional receiver and nanofluid receiver is that the nanofluid receiver unavoidably requires a transparent glazing to contain the high temperature/pressure nanofluid [115].

Figure 6.1 Cross section of concentrator and receivers (glass tube or coated copper tube) Thus, while coated copper tubes have 10 mm outer diameter, the glass tubes have an inner diameter of 10 mm to ensure the relative absorber area matches with design parameter of CPC reflector. Ray tracing through both receivers was conducted for comparison, which can be seen in Figure 6.2.

(a) (b) Figure 6.2 Ray tracing results for the collector: (a) details of ray paths through the copper tube, (b) details of ray paths to the glass tube Surface absorbers (e.g. coated copper/aluminium/steel tubes or plates) utilize selective coatings to enhance their solar absorbance and to reduce long wavelength radiative emission losses. Electrodeposited black chrome was chosen for the surface absorber for comparison purposes in this study due to its good spectral selectivity, ease of manufacture, and long-term durability at the desired operating temperatures (up to 300 °C) [195].

Borosilicate glass tubes were chosen (over plastics) to contain the nanofluid because they are low cost, have high transmittance, are readily available, and have high chemical and UV stability. Borosilicate glasses also have very low thermal expansion coefficients and lower thermal emittance (~ 0.84) as compared to normal glass (~ 0.91) [196]. Thus, commercially available borosilicate glass tubes with 10 mm inner diameter were used for our nanofluid absorber. However, that the thermal emittance of the outer surface of the glass tubes was calculated to be 0.84, with a negligible change in the temperature range of 25-250 °C [197] – a fact that incurs significant loss at high temperature relative to black chrome. Thus, making the DASC ‘selective’ by adding a low emissivity coating to reduce the radiative heat loss, as investigated by Taylor et al. [5, 114]. Note that while it may be possible to achieve a better coating, the ITO coating was specifically designed for this application and was the best we could achieve from the manufactures available at the time of this study [114, 193].

Nanofluids make excellent absorbers for solar radiation if the particle material, size and volume fraction can be controlled precisely [114]. Several different types of nanoparticles in various base fluids have been modeled and tested experimentally for this purpose. These include metallic nanoparticles, carbon nanotubes (CNTs), graphite, carbon black, and carbon nanohorns (CNHs), all of these materials have been shown to be effective solar absorbers, but graphitic particles, such as MWCNTs fluids, are seen as

the most promising for low to medium temperature applications (<400 C) [114, 115, 118, 124, 125, 131] due to their high spectral absorptivity (close to 100%) [115] low cost (around $1/L [131]) and stability [131].

To determine the concentration of MWCNTs required to achieve good solar absorptance, one needs to first determine the spectral absorption and scattering coefficients of MWCNT-water nanofluids as a function of concentration. As the modeling of MWCNTs is tedious and time consuming, the combined-experimental method proposed by Dombrovsky et al. [198] is used in this work to experimentally determine the optical properties of water based MWCNT nanofluids.

First, the spectral directional-hemispherical reflectance (Rλ,dh) and transmittance (Tλ,dh) of water and three MWCNT-water nanofluids were measured using a UV-Vis-NIR spectrometer (Perkin Elmer Lambda 1050 with a 150mm integrating sphere). The concentrations of the three nanofluids tested were 9, 18 and 52mg/L (Note that water represents a 0mg/L nanofluid). The measurements were taken from 250 to 2500 nm at an interval of 2 nm. During the measurements, the nanofluid is contained in an IR grade quartz cell with a 10mm path length. As proposed by Dombrovsky et al. [198], the transport and modified two-flux approximations are used to solve the 1D radiative transfer equation (1D-RTE) in order to reverse calculate the spectral absorption and scattering coefficients of the three nanofluids. Here note that the reflective losses at the boundaries, including multiple reflections are accounted for in the boundary conditions of the 1D-RTE. The same method was used by Hewakuruppu et al. [199] to determine the spectral properties of a gold nanorods and silver nanospheres. There it was demonstrated that this combined experimental-theoretical method yields results that show good agreement with pure-theoretical calculations of nanofluid properties.

The spectral absorption and scattering coefficients of a nanofluid are linearly proportional to the nanofluid concentration As such, the absorption and scattering coefficients at any MWCNT concentration (that was not tested) can be calculated by interpolating the optical properties of the three nanofluids tested above (linear extrapolation is used to determine the properties of a nanofluid with a concentration larger than 52mg/L). Then using the same solution to the 1D-RTE, it is possible to calculate the Rλ,dh and Tλ,dh values for the considered MWCNT concentration and a

given fluid depth. Spectral absorbance (Aλ) can then be calculated as Aλ = Rλ,dh + Tλ,dh, The spectral absorption values can be weighed against the solar spectrum to calculate the solar weighted absorbance of the considered nanofluid.

Figure 6.3 Solar weighted absorbance of water-based MWCNT nanofluids at different concentrations and receiver depths Figure 6.3 plots the solar weighted absorbance of MWCNT-water based nanofluids at different concentrations and path lengths, calculated using the above method (here note here that, reflectance losses caused by refraction at boundaries are not considered to concentrate only on the radiative properties of the nanofluid). The design point marked on Fig. 2 shows that for an absorber with a mean fluid depth of around 7.5 mm, a SWA of 95% will require a concentration of 50 mg/L. Based on this result, a 50 mg/L nanofluid was used in this study. This concentration allows the MWCNT-water/oil nanofluid to compete with black chrome in terms of solar absorption capability.

Note that the effective specific heat of the nanofluid, Cp,nf , is estimated by Equation (6.1) [133]. Based on the calculation, the effective specific heat of nanofluid (with 50 mg/L) has a negligible difference from the base fluid. While the MWCNT-water nanofluid has an effective specific heat of ~4,120 J-Kg-1K-1 (~50 °C), oil -based nanofluid has an effective specific heat of 1,907-2,681 J-Kg-1K-1 at 25-250 °C. Note the pure Therminol 55 has a specific heat value of 1,930-2,720 J-Kg-1K-1 at the same temperature range.

100

Cp,nf f v C p,np (1 f v )C p,bf (6.1)

Overall, the absorber geometry was developed based upon the optical and thermal concentration studies defined in previous chapters. To compare against a conventional surface absorber (consisting of a black chrome-coated copper tube), a volumetric absorber (consisting of a MWCNT nanofluid contained within a glass tube) was designed which can achieve a high absorptance (~95%) for a low cost (around $1/L additional cost), with a negligible change in the thermos-physical properties as compared with base fluid (water or thermal oil).

6.1.2 Thermal design

As the collector’s thermal design was largely defined in chapter 4, the thermal design of absorber herein focuses on the detailed design (structure/material) of the absorbers.

To avoid a potential optical gap loss between the CPC and the receiver due to employing an evacuated glass tube [46], both the CPC and the absorber are housed inside the evacuated tube (Borosilicate glass), as shown in Figure 6.4. While the vacuum insulated surface absorber consists of a metal pipe, the vacuum insulated volumetric absorber consists of a glass pipe. Both receivers incorporate glass–to-metal seals in the experimental set-up use the same vacuum enclosure. Since the glass tube is very fragile it does not connect with conventional compression pipe fittings. Instead, two metal bellows are used, as shown in Figure 6.4(b) to accommodate the thermal expansion of the glass tube.

(a) (b) Figure 6.4 Design proposal of (a) a vacuum glass tube insulated surface absorber (b) a vacuum glass tube insulated volumetric absorber As the temperature difference between the receiver and the ambient will near 400oC at stagnation, avoiding/limiting thermal stresses in the seals is extremely important to 101

maintain seal integrity [200]. The metal material used in this design is Kovar – an iron-nickel-cobalt alloy which has very similar thermal expansion characteristics to Borosilicate glass (~5-10× 10−6 K−1, at 30-800 °C). This matched seal gives much lower residual stresses in the glass-to-metal seal [200, 201].

An additional improvement that can be made for the nanofluid absorber is to add a low emissivity coating on the outer surface of the glass tube [114, 202]. This coating should significantly reduce the radiative heat loss, particularly at a higher operating temperature. As such the borosilicate glass tubes were coated with a custom, commercially available 100 nm thick ITO coating. According to manufacturer’s specifications the solar weighted transmittance of the ITO-coated glass tube was determined to be 0.82 (manufacturer: Geomatic CO., Ltd), which is slightly less than bare glass with a solar weighted transmittance of 0.91. Meanwhile, the hemispherical emissivity of the glass tube was reduced to 0.35, which is much better than bare glass (0.84) at room temperature.

Table 6.1 [41, 203] shows the materials’ blackbody weighted emittances as function of surface temperature (50-250 °C), which were calculated using Kirchhoff’s law [204] and spectral emissivity data (obtained from manufactures’ specifications [158, 196]). Table 3 also gives the black body spectral energy distribution for these temperatures (using earlier work [41]) and Plank’s law [163]).

Table 6.1 Black body-weighted emittance of materials as a function of surface temperature Black body -weighted emittance Material Remarks 50 °C 100 °C 150 °C 200 °C 250 °C Borosilicate Glass 0.84 0.838 0.837 0.835 0.83 Measured result [203] ITO on 0.34 0.345 0.35 0.353 0.355 Calculated result Borosilicate Glass (this study) Black Chrome on 0.22 0.27 0.31 0.35 0.4 Calculated result Copper [41] While black chrome’s blackbody weighted emittance increases with temperature, the variation of the glass/ITO tube emittance is relatively constant across the absorber temperature range (50-250 °C) [205]. This comes from the fact that black chrome having a higher emissivity at short wavelengths, whereas glass/ITO has relatively constant optical properties for long wavelengths.

102

6.1.3 Optical efficiency

The collector’s optical efficiency represents the portion of solar flux that reaches the absorber, and this quantity can be calculated by Equation (6.2).

Gr ηopt (6.2) Cr G b

The amount of solar irradiance that is able to reach the receiver (Gr) can be calculated by using Equation (6.3) [55, 124]. It should be noted that the transmission of glass

receiver tube, τg,r , should be removed from the equation when calculating Gr for the surface absorber (i.e. the surface absorber does not have an outer layer of glass).

i Gr Cr G bη Lens τ g,v ρ CPC τ g,r K(θ) (6.3) where Cr, the geometric concentration ratio, can be obtained from Equation (6.4).

A (6.4) Cr lens Ar

Gb is the beam radiation and the value of ηLens was provided by the supplier (NTKJ Co.,

Ltd.). The parameter τg,v represents the transmittance of vacuum glass tube and ρCPC represents the reflectance of CPC. Superscript i is the average number of reflections in the CPC, which is assumed to be 0.8 [206]. All of these parameters are summarized in Table 6.2 using values from [81, 196, 206, 207]; The last parameter in Equation (6.3) accounts for non-normal incidence for this system. Thus, the parameter, K(θ) , is the incidence angle modifier (IAM), and it can be determined by Equation (3.4) [71]:

Based on these calculations, the MWCNTs nanofluid receiver without/with the vacuum glass packaging has an optical efficiency of ~76%/69%, while the BCCCT absorbers possess an optical efficiency of ~88%/79%. The variation in the optical efficiency can be explained by the optical reflective losses from outer glass surface of the nanofluid receiver, which yields only effective 90% transmittance through to the nanofluid.

103

Table 6.2 Details of the material and optical properties along with the geometric dimensions of the components used in this collector [81, 196, 207]

Symbol Value Unit Symbol Value Unit Symbol Value Unit

η 2 ρ -1 Lens 0.90 - Alens 0.675 m np ~2.1 gmL 0.94 ρ - A 0.1413 m2 f 0.02 - CPC [81] r v 1,907-2 ,681 0.91[1 800 -1 -1 -1 -1 τg - Cp,np JKg K Cp,nf (at JKg K 96] [207] 25-250 °C) 1,930- 2,720 K(θ 0 ) τ C -1 -1 g(ITO) 0.82 - p,b f (at JKg K 1 - 25-200 °C)

6.2 CFD analysis

To analyse the thermal performance of this collector with the proposed absorbers, computational fluid dynamics (CFD) was used. The computational model of the prototype was developed using ANSYS–CFX (version 14.5). For simplicity, this analysis has been separated into two steps. First, the total heat loss coefficient of the

tube receiver, UL , as a function of tube surface temperature, Tabs , was obtained via a 2D numerical simulation on a transverse cross-section of the collector. Note that in the 2D model any losses through the ends of collector are neglected. It should also be noted that in the 2D model it was assumed that the collector is mounted flat on the roof to obtain the convective and radiative heat loss coefficients for receiver tubes. A similar 2D UL determination was made (with reasonable agreement with experiments) in a previous study by the a colleague for the (somewhat similar) Chromasun collector [71].

Using this approach, the heat loss, Qloss , can be obtained by a 2D model. Thus, the overall

heat loss coefficient, , for different absorber temperatures, Tabs , was calculated using Equation (6.5) [71, 208].

(6.5) Qloss UL Aabs (T abs T a )

104

Second, the efficiency of the collector, ηth , as function of the working fluid temperature was determined from a 3D numerical simulation (using the 2D UL result), with the domain limited to the receiver tube. In the 3D simulation, then, the efficiency of the collector, , can be determined for various different working fluids and fluid

temperatures, Tin , using the Equation (6.6).

CP (T o T in ) m (6.6) ηth AGlens b

6.2.1 2-D Simulation

The absorber tubes were modelled as isothermal surfaces, and the tube surface temperatures were set to 50, 100, 150, 200, and 250 °C. The black chrome coating emissivity is a function of temperature and falls in the range of 0.2-0.4 for temperatures between 50-250 °C [41]. The emissivity was set to a constant of 0.84 for the glass tube at the same operation temperature [197]. The outside surface of the Fresnel lens (at the top and sides) was exposed to forced convection from wind at the boundaries. Thus, an 2 external heat loss coefficients of 15 W/m K was assumed [209], and the ambient temperature was set to a value of 25 °C. The lens has an internal emissivity of 0.9, and acrylic cover has an internal emissivity of 0.94. The bottom cover was also modelled with a convective boundary, but the external heat loss coefficient was assumed to be lower, 5 W/m2K [209]. The collector enclosure contains air, which can circulate due to buoyancy -driven natural convection. This was assumed to be laminar, and a steady state buoyancy model was used with Rayleigh numbers in the range of 2.2103-3.6 103. Radiation heat transfer was determined in this model with the Monte Carlo (surface to surface) simulation method [209, 210]. All discretization was carried out using second-order schemes and the air properties were calculated using an ideal gas model. A -3 -5 minimum convergence criterion of 10 was used for continuity and velocity, and 10 was used for energy. A grid dependency study was undertaken to ensure the adequacy of this mesh density. As shown on Figure 6.5, the final mesh size was approximately 0.5 mm, with ~76,100 mesh elements for the 2D model.

105

Figure 6.5 Meshing of the 2D transverse collector cross–section The temperature distribution of air inside the collector can be simulated by selecting various absorber temperatures (between 50 and 250 ℃). Figure 6.6 shows the temperature distribution on the collector cross-section with an absorber temperature of 100 ℃ (without vacuum insulation). The temperature contours show significant thermal gradients around the absorbers and CPC reflector, indicating a considerable amount of natural convection heat loss. Although there are clearly some turbulent natural convective flows in Figure 6.6, the time-averaged heat loss under steady operation conditions can be approximated from this model.

Figure 6.6 Temperature distribution for the collector cross section with coated copper tube receivers (without vacuum insulation) The heat loss values determined by the CFD model for different absorber temperatures are summarized in Figure 6.7. The results indicate that the average, overall heat loss increases from 25 to 350 Watts per meter of length for the surface absorber when it operates in the 50-250 °C temperature range. For the nanofluid absorber, the overall heat loss is was found to be in the range from 50 to 700 Watts per meter. This approximate doubling of the heat loss, particularly at higher temperatures, can be explained by the glass tube having a much higher emissivity (0.84) than the BCCCT (0.2-0.4).

106

(a) (b)

Figure 6.7 Absorber heat loss at different absorber temperatures: (a) black chrome-coated copper tube receiver (b) nanofluid volumetric absorber (contained within a glass tube) To demonstrate the effects of adding a vacuum around the receivers and a low emissivity coating on the glass tubes, the 2D simulations of overall heat loss were performed again with these improvements, as is shown in Figure 6.8(a).

(a) (b) Figure 6.8 Overall heat loss and overall heat loss coefficient for different absorber conditions (a) overall heat loss for absorber in a vacuum; (b) overall heat loss coefficient for absorbers with/without vacuum insulation It can be concluded from this analysis that the overall heat loss can be substantially reduced due to the convective heat loss suppression and that an ITO-coated glass tube also provides a benefit (albeit it is still higher than the BCCCT across the 50-200 °C range). In fact, the total heat loss of the ITO-coated glass tube is only lower than the BCCCT when the absorber temperature is above 200 °C.

107

6.2.2 3-D Simulation

A 3-D geometry was created to determine how much useful energy can be received by heat transfer fluid, water or Therminol 55, with various inlet temperatures in the range from 50-250 °C and an inlet mass flow rate of 0.01 kg/s (to achieve a temperature rise between inlet and outlet of at least 3 degrees, as suggested by the ISO 9806-1 standard [211]). The outer tube wall was modelled as a convection boundary with an overall 2 external heat loss coefficient of 2.5-30 W/m K (including convective heat loss and radiative heat loss) obtained from the 2D analysis. When simulating the nanofluid, only the spectral absorption coefficient was modified by mathematical expressions in CFX expression language, which can be generated by using the nanofluid spectral absorption data from test result [121]. The thermal properties were assumed to remain the same as the base fluid – a reasonable approximation for a particle loading of 50 mg/L (or 0.005 % by mass) [124]. The working fluid flow was modelled as turbulent, since the Reynolds number is > 3,000.

Radiation was modelled using the Monte Carlo simulation method [209, 210]. For the nanofluid, the ‘Participating Media’ mode with a ‘Multiband’ spectral model was used, which accounts for receiver absorption, emission and any reflection (or scattering). To

'' simplify the model, a uniform heat flux, q in , was applied to the surface of the receiver tubes. This heat flux can be calculated using Equation (6.7) for the copper tube receiver [58]:

'' qin G b Cr η opt τ g,v α abs (6.7)

2 In this equation the beam radiation, Gb, was assumed to be 800 W/m . The optical concentration ratio, Cr, and the optical efficiency were calculated using Equation (6.2)-(6.4).

For the nanofluid receiver tube, the heat flux density was found by using Equation (6.8), which also takes into account the reflectance/transmittance of the glass tube [133]:

'' qin G b Cr η opt τ g,v τ g α nf (6.8)

108

The transmittance of the vacuum glass tube, τg,v , and the clear glass tube receiver, τg are

given in Table 6.2, whereas the solar weighted absorbance of the nanofluid, αnf , was taken from Figure 6.3.

-4 In the 3D model, the minimum convergence criteria were set at 10 for continuity and -6 velocity while 10 was used for energy convergence. The resulting mesh size is approximately 0.5 mm, with ~656,000 mesh elements in total, as is shown in Figure 6.9. A grid dependency study was also undertaken which indicated the adequacy of this mesh density.

Figure 6.9 Meshing of the 3D receiver tube interior

The 3D simulation uses the overall heat loss coefficient, UL , from the 2D model shown in Figure 6.8(b). These are essentially treated as input parameters for the next step – aiming to determine the collector efficiency. This is accomplished with by focusing on the receiver tube domain and varying the inlet temperature (in 50 °C increments) from 50-250 °C. To systematically study these design alternatives, the collector efficiency was determined for both absorbers, with and without vacuum insulation.

Figure 6.10 (a) shows characteristic axial temperature distributions along a 1.5 meter length of BCCCT or nanofluid receiver without vacuum insulation at a 100 °C inlet temperature, respectively. It can be seen that the nanofluid receiver has a 4.7 °C fluid temperature rise over the length, while the BCCCT receiver has 6.4 °C temperature rise for the same conditions. Based on Equation (6.6), the thermal efficiencies for these cases can be calculated to be 65% for the copper tube and 46% for the nanofluid absorber.

109

(a)

(b) (c) Figure 6.10 Temperature distribution of the working fluid with an inlet temperature of 100 °C: (a) Therminol 55 and MWCNTs-Therminol 55 nanofluid on the axial cross sections of the absorber without vacuum; (b) Therminol 55 on the radial cross sections at middle of the BCCCT absorber without vacuum (at 0.75 meters); (c) MWCNTs-Therminol 55 nanofluid on the radial cross sections of nanofluid absorber without vacuum (at 0.75 meters)

To understand the differences in how heat is transferred radially, characteristic radial cross-sections can also be analysed. For this, we took a cross section at the middle the receiver length (at 0.75 meters). Figure 6.10(b) reveals that the wall of the coated copper tube absorber is about 0.6 °C higher than the middle of the fluid. This temperature drop is created by the thermal resistance of conduction and convection between the outer absorber surface and the working fluid. Figure 6.10(c) shows that the wall of the nanofluid receiver is about 0.3 °C lower than the middle of the nanofluid. As discussed in the introduction, and as is noted in the literature, lower surface temperatures (relative to the bulk mean temperature) were expected [115, 127]. However, under the conditions in this study, this temperature inversion was not large

110

enough so as significantly impact on the overall performance. Thus, the ‘volumetric’ benefit can be neglected for this collector application.

The CFD efficiency results of the vacuum insulated black chrome and ITO/glass absorbers at a 200 °C inlet temperature and 0.01 kg/s mass flow rate are shown in Figure 6.11. Based on Equation (6.6), the collector’s thermal efficiencies for these cases were found to be 51% for the vacuum-insulated copper tube and 36.5% for the vacuum-insulated nanofluid absorber.

(a)

(b) (c) Figure 6.11 Temperature distribution of the receivers with an inlet temperature of 200 °C (a) Therminol 55 and MWCNTs-Therminol 55 nanofluid on the axial cross-sections of the vacuum insulated BCCCT and nanofluid absorber; (b) Therminol 55 on the radial cross-sections at middle of the vacuum insulated BCCCT absorber (at 0.75 meters); (c) MWCNTs-Therminol 55 nanofluid on the radial cross-sections at middle of the vacuum insulated nanofluid absorber (at 0.75 meters)

111

The full CFD-derived efficiency curves for these configurations (from 25-250 °C) are shown in Figure 6.12. It reveals that without vacuum insulation, the black chrome-coated copper tube out performs the nanofluid absorber at all temperatures. For low temperature operation (60-80 °C), it can be seen that much is due to the nanofluid receiver’s lower optical efficiency. Without a vacuum, the black chrome receiver has an efficiency range of 67-75%, whereas the MWCNT nanofluid-based receiver has an efficiency range of 45-60%, across the low temperature operating range.

Figure 6.12 CFD simulation efficiency results At higher temperatures in the vacuum system, it appears that the ITO absorber begins to recover some of this ground by reducing the long wavelength emissive losses (although not enough to account for the initial optical loss). For the vacuum insulated absorber, the black chrome still out-performs the nanofluid absorber, with efficiencies of 45%-36% and 32%-27%, respectively, when operating at 200-250 °C.

Additionally, it should be noted that compared to the non-vacuum efficiency there is a large optical loss resulting from adding the vacuum glass (since the transmittance of an extra glass tube is 91%). This is in addition to the ITO coating where the transmittance of the ITO-coated glass tube is 82%, yielding a decrease in optical efficiency of nearly 25% in total. It can be seen that the optical efficiency loss of the vacuum glass is recovered when the nanofluid receiver operates above 60 °C and for when the coated copper tube receiver operates at 80 °C, due to heat loss reductions. However, even when an ITO coating is added to the glass tube, the evacuated nanofluid receiver efficiency is

112

still 10-15% lower than the copper tube due to the higher optical loss and larger thermal emissivity of the nanofluid absorber. There is room for improvement here, of course, by using anti-reflective coatings on the glass surfaces (although this was not incorporated into the present study).

6.3 Absorber prototyping

6.3.1 Nanofluid preparation

MWCNTs (average diameter of 6-13 nm and average length of 2.5-20 μm) were purchased from Sigma Aldrich and used without further purification. First, 40 mg of pristine MWCNTs was added to a flask of 50 ml of deionized water and dispersed using bath sonication for 60 minutes. To functionalize the MWCNTs in suspension, 0.45 g of potassium persulfate (KPS) was added to the flask and the pH of the reaction system was adjusted to 13 by adding concentrated potassium hydroxide (KOH) solution. The mixture was stirred for 3h with a magnetic stirrer at 85 °C. After cooling to room temperature, the solution was diluted with water and dispersed again, using bath sonication. After sonication, the solution was centrifuged at 3000 rpm for 20 minutes. The supernatant solution was then collected and filtered using a polyamide (PTFE) filter paper and washed with water to neutrality. The solids were then dried at 80°C overnight. After functionalization, 150 ± 0.1 mg of functionalized tubes were dispersed in 3l mL of Deionized Water (DI) with a probe sonicator (Misonix ultrasonic liquid processors) at 20 Watts in a water-ice bath for 20 minutes.

Figure 6.13 gives a transmission electron microscope (TEM) image of the prepared MWCNTs. Note that this TEM image does not depict MWCNTs in suspension, but of those deposited on a TEM grid for imaging.

113

Figure 6.13 TEM image of potassium persulfate functionalized MWCNTs The sample (10 mm thickness) was analyzed by UV-VIS-NIR absorption spectroscopy, Figure 6.14 shows the UV-VIS-NIR spectra of the MWCNT dispersed in DI water. The absorbance of nanofluid in the range of 1.5-4 has been tested across the wavelength range from 400-1400 nm.

4.0 MWCNT-DI

3.5

3.0

2.5

Absorbance(a.u.) 2.0

1.5

1.0 400 600 800 1000 1200 1400 Wavelength(nm) Figure 6.14 Absorbance of the prepared nanofluid (UV-vis-NIR spectra of KPS treated CNTs in DI) Since the absorbance is computed from transmission of the sample using the following Equation (6.9).

A = -Log10T (6.9) where T represents transmission of the nanofluid sample. The absorbance result indicates a very low transmittance (1-3%) of this nanofluid. It should be noted that the extinction coefficient if composed of the absorption coefficient and the scattering

114

coefficient. Due to the fact that the scattering is proportional to D4 based on Rayleigh scattering approach [120], and the average diameter of MWCNTs particle, D, is 6-13 nm. Consequently, the scattering coefficient is negligibly small [120]. Here note that no reflectance losses are considered at the boundaries, so the absorbance of this nanofluid sample is 97-99% across the wavelength range from 400-1400 nm (e.g. across the majority of the solar spectrum). This spectroscopy result is as predicted from Figure 6.3, and is similar to the solar absorption capacity of our black chrome coating (which has a SWA of ~95%).

6.3.2 Absorber assembly

The absorber prototype was constructed in a way that either a traditional surface-based absorption receiver or a nanofluid receiver can be place at the focal line. Figure 6.15 (a)-(d) give pictures of the real components integrated with the surface absorber (10 mm outer diameter) and the volumetric absorber (10 mm inner diameter) without vacuum insulation.

(a) (b)

(c) (d) Figure 6.15 Pictures of the solar absorbers: (a) copper tube receiver mounted in the CPC; (b) full optical assembly of the surface receiver; (c) MWCNT nanofluid receiver mounted in the CPC; (d) full optical assembly of the volumetric receiver;

115

To eliminate convective and conductive heat losses from air, a vacuum pressure below 10-3 mbar was required to be maintained inside the package. In order to ensure the absorber can be replaceable in the vacuum glass chamber, as shown in Figure 6.16, a vacuum chamber were designed and assembled with glass tubes (1.5 m length and 70 mm outer diameter) and vacuum flanges.

A cross section view of the vacuum flanges shows that 4 O-rings (silicone rubber) are used at the gaps between flanges and the glass tube to seal the both ends of glass tubes. Two O-rings (FFKM, Kalrez) are also used to seal the both ends of tube absorber.

(a) (b) Figure 6.16 Vacuum glass chamber prototype design: (a) 3D view of absorber inside the vacuum glass tube; (b) cross-sectional view of vacuum flange with O-rings The vacuum glass tube assembly is shown in Figure 6.17. Note that these vacuum chambers are used to connect with a vacuum pumping station, which can pump the air out of vacuum glass tubes continuously during the experiment. The full experimental test will be reported in next chapter.

116

(e) (f) Figure 6.17 Pictures of the solar absorbers in vacuum chamber: (a) surface absorber inside the vacuum glass tube; (b) both vacuum packaged receivers, side-by-side 6.4 Summary

In this chapter a combination of fundamental theory and numerical methods were used to investigate of the effect of modifying the absorber in the proposed industrial heat solar collector. In particular, a volumetric absorber (consisting of a MWCNT nanofluid contained within a glass tube) was designed and compared against a conventional surface absorber (consisting of a black chrome-coated copper tube).

The analysis revealed that the vacuum-packaged volumetric receiver can potentially achieve an efficiency of 54% and 32% operating at 80 °C and 200 °C, respectively. This is expected (from comparative simulations) to be lower than a vacuum-packaged black chrome-coated receiver, which had an efficiency of 68% and 45% in the same concentrator, operating at the same temperatures, respectivley. The inferior performance of the volumetric receiver was found to be due to higher reflective optical and radiative heat loss from the surface of the glass tube. Overall, these findings demonstrate that both receivers could work for industrial heating applications, but volumetric absorbers will require anti-reflective and good selective coatings to be competitive with surface absorbers. If these challenges can be overcome, nanofluid receivers may yet provide an effective and a low-cost approach to bring nanotechnology into industrial heating and air-conditioning applications since glass-to-glass vacuum sealing is easier to achieve than metal-to-glass.

117

Additionally, prototyping of both the surface and volumetric absorbers has been conducted from the CFD analysis in this chapter. Additional routes for improvement were also derived from the analysis – namely, anti-reflective coatings (optical loss) and increasing the nanofluid depth to get a true temperature inversion (thermal loss) should be incorporated into commercial designs. Without the expense of incorporating these, the next chapter will conduct a comparative experimental study to determine the real ‘on-sun’ performance of these absorbers and to validate the numerical analysis results.

118

Chapter 7

7. Collector Performance Testing

To establish the performance of the proposed design, a collector prototype was developed in a way that either a traditional surface-based absorption receiver or a nanofluid receiver can be tested in the same package. The performance of the prototype, in terms of the effective optical efficiency, incidence angle modifier and instantaneous thermal efficiency were evaluated with outdoor, steady-state efficiency experiments.

Details of the prototype (materials and costs), the experimental setup, and the procedure are described firstly, followed by the full thermal efficiency test as compared with the CFD results and selected commercialized collectors. Moreover, design improvements of this collector are discussed, such as employing a better selective coating (e.g. TiNOx) to enhance the thermal efficiency.

7.1 Experimental Setup

7.1.1 Non-vacuum prototype and loop

The preliminary experimental rig comprised a solar collector, a fluid loop, and a data acquisition system. As shown in Figure 7.1, the solar collector prototype had a dimensions of 2 x 0.7 x 0.15 m (length x width x height), with an adjustable height of up to 0.3 m. The collector prototype assembly drawings and parts list with price are given in Appendix C and D.

The working fluid loop consisted of one fluid tank, several manual valves, a mini DC water diaphragm pump, heat exchanger, tubes, and tube fittings to maintain constant flow rates in the system. However, the preliminary experimental fluid loop was not able to circulate the fluid temperature above 85 °C due to the operating temperature limits of the DC pump and an Acrylic tank. The efficiency testing was only performed with water and water-based nanofluids as the working fluid and for inlet temperatures in the range of 20-85 °C.

All of these components were mounted on a wooden board, which in turn, was clamped on the tracking plane of an external solar tracking rig. This tracking system allowed the

119

solar collector to maintain normal incidence with the sun throughout the test. Note that the internal linear tracking of collector was not operated in this case, since the aim of this experiment was to obtain the steady state efficiency of collector when it was directly facing the sun.

The mass flow rate of working fluid was controlled by DC power supply, and the set flow rate was measured by a digital flow meter (RS 508-2407) [212]. The temperature of inlet working fluid was controlled by using a heat exchanger. Three RTDs (resistance temperature detectors) (TC Direct PT100) [213, 214] were installed to measure the collector inlet and outlet fluid temperatures and the ambient temperature. The intensity of global and diffuse solar radiation was recorded using two pyranometers (Middleton EQ08 and Eppley radiometer Model GPP [215]). These measurements were collected and stored in a computer through a data acquisition module from National Instruments (NI cDAQ-9174). The flowmeter was calibrated by drawing off water from the system into a container and measuring the mass and time with a balance, which resulted in an uncertainty of about 2.5%. Temperature sensors were calibrated using a calibrated thermometer (Fluke 1521 [216]) to give an uncertainty less than 0.01 C. The pyranometers were calibrated with a standard pyranometers (Eppley PSP [217]) with valid calibration certificates to give an uncertainty less than 3%. The above mentioned accuracy values meet the requirements of the ISO 9806-1 standard [211].

The outdoor steady state efficiency test was conducted under satisfactory ambient weather conditions according to the test method in the ISO 9806-1 [157, 218-220]. Additionally, before every test run, the circulation loop was cleaned using distilled water and ethanol to remove any fouling residue that could affect the collector performance.

120

Pyranometer (Diffuse radiation)

Pyranometer (Global radiation) RTD (Inlet temperature) Heat exchanger

Collector Flowmeter

Data acquisition module

Tank

Pump

RTD (Outlet temperature)

Figure 7.1 View of the non-vacuum collector prototype undergoing testing 7.1.2 Vacuum glass insulated prototype and loop

According to our previous CFD analysis, adding vacuum insulation can increase the efficiency at this temperature range due to suppression of the convection and conductive heat losses [221]. Experimental work was also carried out to demonstrate the performance of this collector by employing vacuum insulation around the solar absorbers, as proposed in chapter 6.

Figure 7.2 Prototype with vacuum insulation 121

A view of vacuum insulated collector prototype is shown in Figure 7.2. The vacuum chambers were assembled using glass tubes (1.5 m length and 70 mm outer diameter) and vacuum flanges. To eliminate convective and conductive heat losses, a vacuum pressure below 10-3 mbar was maintained between the absorber and the outer glass tube. To achieve this, as is shown on Figure 7.3(a), the chamber was connected to a Pfeiffer HiCube 80 Eco Turbo pumping station. This system consisted of a turbopump (a HiPace 80) [222] and a specially matched backing diaphragm pump (a Pfeiffer MVP 051-2) [223]. A Pfeiffer Compact FullRange™ Gauge PKR 251 and a single channel measurement unit (a TPG 261) pressure gauge and display were connected to continuously monitor the vacuum pressure [224].

As shown in Figure 7.3(a), the experimental rig consisted of a full-scale solar collector prototype, a high temperature circulating fluid loop, a vacuum pumping station, several temperature, pressure, and flow sensors, and a data acquisition system (DAQ). Figure 7.3(b) shows a schematic of the high temperature circulating loop, which enables the working fluid to achieve a maximum inlet temperature of 250 C and a maximum pressure of 16 bars. To achieve at least a 3 degrees Celsius temperature rise between the inlet and outlet, a mass flow rate of 0.02 kg/s was maintained by controlling the rotational speed of the magnetic-driven pump (T ECO MAG-M1) and the aperture of the needle valves denoted as NV1 and NV2 in Figure 7.3(b). To ensure this was the case, a Coriolis mass flow meter (Micro Motion Coriolis F-Series sensor, Emerson Process F025) was used to measure the mass flow rate passing through the collector. As described in the section 7.1.1, the fluid and ambient temperature were measured by 4-wire resistance temperature detectors, and the global and diffuse solar irradiance were measured by two recently calibrated pyranometers (Middleton EQ08-S and EQ08-S, respectively). All of these measurement readings were recorded through a National Instruments data acquisition module cDAQ-9174 using the LabVIEW platform.

122

(a)

(b) Figure 7.3 Testing platform for the collector prototype: (a) Photo of the testing platform and (b) a flow diagram of the high temperature loop

The Coriolis flow meter was calibrated by the supplier, which gave mass flow rate uncertainty of less than 1%. RTDs were calibrated using a certified reference thermometer (Fluke Calibration 1521) and a precise temperature controlled water bath, which resulted in an uncertainty of less than 0.02 C. The pyranometers were calibrated by the supplier under normal irradiation using a reference pyranometer (EQ08-S,

123

S/No.4901) resulting in an uncertainty of less than 3% and 1.5% for the diffuse and global measurements, respectively. Taken together the accuracy of our experiment satisfies the test standard ISO 9806-1 [157, 211, 218-220].

According to EN 12975 and ISO 9806-1, performance data of a solar thermal collector can be used to find the coefficients of the following nonlinear efficiency equation [157, 218]:

2 (Tm T a ) (T m T a ) ηth η o a 0 a 1 (7.1) GGbb

By taking the efficiency of the collector, ηth , as the dependent variable ‘Y’, and x1=

2 (TT)ma (TT)ma and x2= as independent variables, the parameters ηo , a 0 , and a1 can Gb Gb be identified using multiple linear regression.

The combined uncertainty for evaluating collector efficiency, Uη , was obtained by the root sum square method (RSS), based on the following relation [128]:

m 2Gb 2 (T o T in ) 2 U()()()η (7.2) m Gb (T o T in )

The maximum combined uncertainty obtained in this study for evaluating collector efficiency was around 4.8% (including all sources of errors).

7.2 Experimental Results and Discussion

7.2.1 Thermal Efficiency Results

On-sun steady-state efficiency experiments were carried out to evaluate the performance of the prototype by using a BCCCT absorber with/without vacuum insulation, and a nanofluid receiver with/without vacuum insulation. Water or Therminol 55 oil -based working fluids were employed as for the non-vacuum (20-85oC) and vacuum (20-150oC) cases, respectively. Data was only taken at steady state conditions, according to the standard ISO 9806-1 [219]. In this case, ‘steady state’ means: less than 0.1K and 1K of variation in the inlet and ambient temperatures, respectively, and less

124

than 50 W/m2 of variation in the irradiation measurement, and less than 1% variation in the mass flow rate.

Additionally, thermal stagnation tests were undertaken to determine the maximum achievable receiver temperature. Thermocouple (type K) wires were attached to the surface of the receiver and connected to a NI-9721 thermocouple module to record the temperature change. Assuming the efficiency is zero when the receivers reach the stagnation temperature, this stagnation point can be used to generate a complete efficiency curve by regression analysis. Note that Therminol 55 was used for BCCCT absorber stagnation temperature test to prevent the working fluid from boiling, while Therminol 55 MWCNT-based nanofluid was employed for nanofluid receiver stagnation temperature test.

The experimental performance of the BCCCT absorber and the nanofluid absorber with/without vacuum insulation are compared in Figure 7.4. Without the vacuum insulation, the tested efficiency of this solar thermal collector employing the BCCCT (working fluid: water) can be described as Equation (7.3).

2 (Tm T a ) (T m T a ) ηth 0.87 1.56 0.013 (7.3) GGbb

Without the vacuum insulation, the tested efficiency of the solar collector employing the nanofluid absorber (working fluid: water/MWCNT -based nanofluid) can be described as Equation (7.4):

2 (Tm T a ) (T m T a ) (7.4) ηth 0.74 2.97 0.009 GbbG

With the vacuum insulation, the tested efficiency of this solar thermal collector employing the BCCCT (working fluid: Therminol 55) can be described as Equation (7.5).

2 (Tm T a ) (T m T a ) ηth 0.77 0.639 0.0045 (7.5) GGbb

125

With the vacuum glass insulation, the tested efficiency of this solar thermal collector employing the nanofluid receiver (working fluid: water/MWCNT -based nanofluid) can be described as Equation (7.6).

2 (Tm T a ) (T m T a ) ηth 0.58 0.683 0.0055 (7.6) GG bb

(a)

(b) Figure 7.4 Comparison of the experimental performance by the black chrome-coated copper tube and nanofluid receivers (a) tested without vacuum insulation (b) tested with vacuum insulation. 2 Note: (a) Weather conditions for the BCCCT tests on 6th-9th July, 2015 in Sydney: Ggl_m = 956 W/m , Gd_m = 241 2 2 W/m , Gb_m = 715 W/m , V= 3-5 m/s, Ta = 17 ± 1 °C. For the nanofluid test on the 16th Oct, 2015 in Sydney: Ggl_m = 2 2 2 1041 W/m , Gd_m = 258 W/m , Gb_m = 782 W/m , V= 3-5.5 m/s, Ta = 29 ± 1 °C; (b) For the BCCCT tests on 12 -13 2 2 2 May, 2016 in Sydney: Ggl_m = 942 W/m , Gd_m = 152 W/m , Gb_m = 789 W/m , V= 5-7 m/s, Ta = 26 ± 1 °C. For the 2 2 2 nanofluid efficiency test on the 24th Mar, 2016 in Sydney: Ggl_m = 978 W/m , Gd_m = 248 W/m , Gb_m = 730 W/m , V= 4-5.5 m/s, Ta = 26 ± 1 °C;

126

The experimental results also reveal that the prototype using a nanofluid receiver has a lower optical efficiency and higher heat loss coefficients. As the optical efficiency of

collector can be represented by the ηo parameter derived from the experimental results (0.87/0.78 for BCCCT absorber and 0.74/0.58 for nanofluid receiver with/without vacuum glass insulation.

The CFD and experimental performance results are compared in Figure 7.5. The experimental results (continuous solid curve) show good agreement with the CFD analysis results (dashed curve), indicating a reasonable accuracy of the CFD results. The profiles of Figure 7.5 show that both methods predict a crossing point between the without vacuum and vacuum cases (at temperature of between 75-85 °C for the BCCCT receiver and 70-80 °C for the nanofluid receiver).

Figure 7.5 Comparison between the experimental and CFD collector efficiency values for the black chrome-coated copper tube and MWCNT nanofluid receivers

Overall, it can be seen that the volumetric absorber used in this concentrated solar collector is inferior to the conventional surface absorber. However, the coefficients from the experimental tests and fundamental mechanisms behind this in the CFD, have indicated some possible methods to improve the real-world nanofluid collector performance.

127

First, it was found that the volumetric absorber employed here did not create a significant temperature inversion – i.e. the outer surface temperature might be designed to be lower that the bulk mean temperature in a better designed system. This temperature difference could be enlarged by considerably increasing the nanofluid depth or/and optical concentration ratio [115, 127]. However, the configuration of this concentrated solar collector has very limited room for improvement in increasing these factors. Second, an anti-reflective coating could be employed which would improve the optical efficiency of the nanofluid receiver by up to 9%. Lastly, it may be possible to obtain a better low-emissivity coating than the ITO featured here (0.35).

From an economic point of view, however, an evacuated ITO-coated nanofluid receiver still likely provides a cost advantage over an evacuated selectively-coated metal receiver. This can be argued on a pure materials cost basis (ITO versus selective coatings) and on the ease of sealing two glass tubes compared to a glass-metal seal for the vacuum insulation. Therefore, with a bit more performance improvement and a cost advantage, an evacuated nanofluid receiver (contained by ITO-coated glass tube) may still have the potential to compete with conventional evacuated selectively coated metal tubes receivers.

Solar thermal collector efficiencies are usually represented on a graph with the TT dependent variable, m a . However, only beam irradiance can be concentrated for Geff medium/high temperature solar collectors (e.g. PTCs). So, effective irradiation on flat plate and evacuated tube collectors corresponds to the total radiation (beam and diffuse). In order to compare the instantaneous efficiency of the tested collector with other medium temperature commercial collectors, the efficiency should be plotted against the same dependent variable on the x-axis. Therefore, the efficiencies for the PTC and the prototype collector should be modified to correspond to the total global irradiance.

Thus, if the initial expression for efficiency depends on the effective irradiance, expressed according to Equation (7.7):

2 TT TTm  a  η a  am a  a (7.7) 0 1GG 2 eff eff

128

Then the thermal efficiency correlation given above is also equivalent to:

TT 2 GTTGtm  a t  m a  η a0  a 1  a 2 (7.8) GGGGefftt eff And the thermal efficiency with respect to the total irradiation can be expressed by Equation (7.9):

TT 2 GeffTTm  a  m a  η  a0 a 1 a 2 (7.9) GGGt t t The instantaneous efficiency of the tested collector as compared with other medium temperature commercial collectors is given in Table 7.1. The efficiency correlations are also plotted in Figure 7.6.

Table 7.1 The coefficients of tested collector (with vacuum insulation) and other compared collectors

Geff Collector a a1 a 2 0 G (W/m².K) (W/m².K²) t Proposed collector with vacuum 0.624 0.640 0.0045 insulated BCCCT (tested) Proposed collector with vacuum insulated TiNOX 0.624 0.55 0.003 coated absorber (CFD) Evacuated flat plate 0.759 0.508 0.007 collector (EFP) Evacuated tube 0.687 1.505 0.011 collector (ETC) Parabolic trough 0.544 0.360 0.001 collector (PTC) Regarding future improvements to the collector, the efficiency of collector can be enhanced by employing a lower emissivity coated receiver instead of using the black chrome-coated copper tube. According to validated theoretical and CFD models, with the vacuum insulation, the efficiency of this solar thermal collector employing the

TiNOX absorber can be expressed as Equation (7.10).

2 (Tm T a ) (T m T a ) ηth 0.77 0.55 0.003 (7.10) GGbb

129

Figure 7.6 Instantaneous solar collector efficiency comparison As shown in Figure 7.6, the evacuated tube collector (ETC) provides high efficiency (>50%) at low temperatures (<100 °C), indicating it is the most efficient collector for low temperature industrial heating application. For higher temperature applications, the parabolic trough collector (PTC) has the highest efficiency, and – if there is sufficient space and/or structural provisions – represents the most efficient collector for medium-high temperature heating applications (250-400 °C).

Both of the evacuated flat plate collector and proposed collector show good efficiency (30-60%) in medium temperature range (100-250 °C). Since the beam and diffuse radiation can be received by evacuated flat plate collector (EFP), it has higher efficiency (>30%) at 0-250 °C. The proposed collector presents a much higher efficiency than the ETC at operating temperature ranges from 100-250 °C. However, the current design can only compete with evacuated flat plate collector (EFP) when operating at temperatures higher than 250 °C. By employing a TiNOX coated absorber, however, the proposed collector could deliver up to 8% of efficiency higher than EFP collector at 200-250 °C operating temperature range (as is denoted by the “Proposed Collector – Future” curve in Figure 7.6).

130

7.2.2 Incident Angle Modifier Test

The incidence angle modifier (IAM) is a numeric value that refers to the amount of available solar radiation striking the absorber of the collector. A value of 1 is achieved when the collector is perpendicular to the sun’s rays. To account for off-normal incidence angles, the detailed collector efficiency equation can be expressed as Equation

(7.11) [157]. The incidence angle modifier (IAM), K θ , is essential to be determined for this application since it will significantly influence the annual performance of integrated storage collector.

2 (Tm T a ) (T m T a ) ηCol K θ a 0 aa12 (7.11) GGbb

Measurements of the IAM were carried out for selected angles along the symmetry planes of the collector. The individual incidence angle modifier can be estimated by considering it to be the product of the transversal and longitudinal and incident angle modifiers, K(θ)T and K(θ)L respectively [157].

K θ K θ K θ TL (7.12)

K θ K θ Where T and L denotes the transversal and longitudinal IAM of this collector.

According to EN 12975 [157], the IAM can be evaluated by thermal performance of collector when operating the collector equals to ambient temperature (Tf=Ta) [157]:

ηθ th T T 0 K θ ma (7.13) ' F τα eff ,θ0

As the proposed collector would have unusual optical performance characteristics compared to flat plate collectors, it is necessary to measure the incidence angle effects for more than one direction (longitudinal and transversal) and at different angles (e.g. 10°, 20°, 30°, etc.) to fully characterize the IAM. Figure 7.7 shows the photos of prototype tested on-sun with 40° longitudinal incidence angle and 45° transverse incidence angle.

131

(a) (b) Figure 7.7 Incidence angle modifier (IAM) outdoor test (a) prototype tested with longitudinal incidence angle of 40°; b) prototype tested with transverse incidence angle of 45° The transversal and longitudinal IAM of this collector were determined by testing thermal performance at different angles from 0° to 45° as they are presented on Figure 7.8.

Figure 7.8 Transverse and longitudinal incident angle modifier

Figure 7.8 shows that the IAM in both orientation decreases from 1 to around 0.5 when incidence angle increases from 0° to 45°, and the longitudinal IAM of this collector is relatively higher than transverses IAM. This indicates that the design can account for seasonal variation in an angle of the sun, or for daily variation in the angle of the sun. In

132

this study, the collector was titled at the latitude angle of the location, and deploy with longitudinal axis along the north-south direction.

7.3 Summary

In this chapter the low profile concentrating solar thermal collector was constructed and investigated experimentally by comparing the thermal efficiency of a volumetric receiver versus a surface solar absorber. Initial testing revealed that the black chrome-coated copper tube operating at 80 °C in this collector had an efficiency of 68%, whereas the MWCNT nanofluid-based receiver operating at 80 °C had an efficiency of 54%. Further experimentation revealed that with vacuum packaging, these would operate at 47% and 26% efficiencies, respectively, at a temperature of 200 °C. Since commercial concentrated systems operating at 200 °C have efficiencies in the range of 44-57%, the proposed collector with a considerably lower profile (<15 cm height), can be considered as a viable alternative for supplying thermal energy in the 100-250 °C range.

However, the instantaneous efficiencies presented in this chapter are given for a normal incidence on the collectors, which does not give much indication as to their performance over time. The following chapters will assess and compare the collectors’ efficiency over one year of operation for different working temperatures.

Furthermore, improvements of the IAM of collector in the transverse orientation are essential and possible. Some passive or active components could be added to the design to extend the acceptance angle (read: hours of operation), but these will add complexity to the system.

Additionally, inspection of the experimental prototype revealed that there was a substantial volume of ‘free’ space inside the collector’s package which could be used for integrating latent heat storage tanks to also extend the working hours of this collector. The design and details of this design with integrated thermal storage will be reported in the next chapter.

133

Chapter 8

8. Innovative ICS solar system

Recent studies have demonstrated that integrated collector/storage (ICS) solar systems are energy-efficient, environmentally friendly, and economically viable [138]. However, most studies to date focus on low temperature phase change materials storage, as shown in Table 2.9, with only a few studies covering medium and high temperature applications (>100 °C).

A hidden benefit of the proposed class of ‘stationary’ rooftop collectors is that there is free space (where sunlight does not pass through the volume) in the package. This can be seen in the ray tracing of Figure 4.4, and enables compact rooftop systems which can follow demand profiles which do not necessarily match with solar irradiance profiles. As such, the ‘free’ space inside concentrating collectors can filled (to some extent) with thermal energy storage.

In this chapter, the proposed concentrating solar collector integrated with a shell-and-tube LHTES system is presented and analysed. The proposed system can integrate with rooftops and is suitable for 100-200˚C temperature range applications (e.g. industrial heating process). The operation of an integrated system was analysed numerically via a validated shell-and-tube LHTES model developed by a fellow PhD student [225-227]. The best configuration of the integrated system in terms of the optimum design of LHTES units and collector thermal insulation methods are reported in below, based on both technical and economic metrics.

8.1 Integrated collector/storage design and analysis

8.1.1 Design overview

Figure 8.1(a) illustrates the 3-dimensional (3D) view of the integrated collector/storage (ICS) design, and cross-sectional views of the design are displayed in Figure 8.1(b). Note that Figure 8.1 shows a half scale (3 x lens/receiver units) prototype with an aperture area of 0.9 m2, whereas the full ICS module is expected to consist of six lens/receiver units and seven cylindrical LHTES system, located in the optical ‘dead

134

space’ beside the receivers. Each collector module, then, has a 1.8 m2 aperture area and 0.05 m3 of available space for thermal energy storage. The concentrated solar energy is absorbed by selectively-coated absorber tubes and the excess energy can either be used directly or stored in shell-and-tube LHTES units. Additionally, the pump, pipe fittings and control valves were integrated within the ISC package, resulting in a very compact, simplified system which is suitable for rooftops.

(a)

(b) Figure 8.1 ICS module design: (a) End view of the internal components (b) Cross-sectional view

Figure 8.1(b) shows an end-on cross-sectional view of how the tube receivers are connected in parallel with the LHTES system, which are also connected in parallel. The receiver tubes and inner heat transfer tubes of the LHTES tanks are then connected with a pump and valves in series. Control valves are used to switch between i) direct solar

135

heating mode (LHTES bypass), ii) combined direct use and LHTES charging mode (a closed collector loop), and iii) discharging mode (receiver bypass). Note that a back-up auxiliary heater can also be implemented if solar-derived heat is insufficient to meet the thermal request.

The collector prototype (without the LHTES) was built and tested in chapter 7. To finalize the integrated design, an optimization process was required to achieve the best configuration of an integrated system which is the main target of this study. Accordingly, in the following sections, the integrated performance of the whole system with LHTES units was evaluated numerically.

8.1.2 Optical and thermal performance

Ray tracing software (ZEMAX, version 12.EE) was employed to simulate different light incident angles to help determine the system’s daily performance. Figure 8.2 depicts that the optical concentrator can achieve a linear focus onto the acceptable region of CPCs with incident angles increasing from 0 to 45 degrees, without shading/blocking by the LHTES units.

Additionally, the ray tracing shows that the LHTES tanks potentially can catch some scattered and diffuse radiation, especially at large incident angles. However, the following analysis will conservatively neglect any direct solar inputs to the LHTES tanks. It should be noted, though, that this design could be modified to use evacuated glass tubes for the LHTES tanks to capture sunlight if the inner tube employs a selective coating to suppress heat loss (similar to the main receiver tubes) [139, 146].

136

(a)

(b) Figure 8.2 Ray tracing results for (a) 0˚ incident light angles and (b)45˚ incident light angles The efficiency curves of the collector (with/without the vacuum insulation, employing a black chrome -coated tube receiver) are given in Figure 8.3. It is worth noting that for the purpose of integrated simulations and analysis, the IAM (see Figure 7.8) must be applied on these correlations as shown by Equation (7.11).

137

Figure 8.3 Collector thermal efficiency with and without vacuum insulation Figure 8.3 illustrates that the collector has an efficiency of around 60-70% for the 100-150 ºC operating temperature range required for the dairy process heating application proposed here. Note that although the proposed system is suitable for a 100-250 ºC temperature range, the analyses in this study are limited to the case study of the dairy industry that needs the collector to operate between 120-150 ˚C.

8.1.3 Shell and Tube LHTES

As shown in Figure 8.1 and Figure 8.4, the cylindrical LHTES units proposed for this study rely on a shell and tube design to segregate the heat transfer fluid (HTF) and PCM. The HTF flows through the inner tubes and exchanges heat with the PCM in the surrounding region (i.e. in the surrounding shell). During the charging process (melting), the hot HTF flows from top to bottom, whilst the flow direction is reversed during the discharging process (solidification). Since the length is variable in this analysis, the top of the tank can be referred to as Z/L=1 and the bottom of the tank can be referred to as Z/L=0. In the charging process, the flow direction is from Z/L=1 to Z/L=0, and vice-versa for the discharging process. The theoretical treatment of this

138

problem is similar to that described by Esen [228] and Visser [229]. Hence, to simplify the physical and mathematical models, the following assumptions have been made [225]:

1. The PCM is isotropic (having identical thermophysical property values in all directions);

2. The thermal resistance of the inner tube is negligible;

3. The thermophysical properties of the HTF and the PCM are independent of temperature; however, the thermal conductivity and the specific heat capacity of the PCM in solid and liquid phase are different;

4. Axial conduction in the HTF is negligible;

5. Any small void spaces between LHTES pipes are neglected;

6. The latent heat storage unit is well insulated, so there are no heat losses to the environment.

(a)

139

(b) Figure 8.4 Generalized cylindrical latent heat thermal storage system (a) 3D view of cylindrical tank with 5 tubes in shell; (b) cross section view of cylindrical tank with 1 tube (D0) and 7 tubes (D2) in shell

Therefore, the mathematical model for the HTF can be written as shown in Equation (8.1) [225]:

2 TTff ρCP πr o  mC P  λ T r r  T f 2πr o (8.1) ff o  tz

Where 휌, 퐶푃, 푟표, λ and T, represent the density, specific heat, inner radius of pipe, convective heat transfer coefficient and temperature, respectively. Subscripts f, t, and z represent fluid (or heat transfer fluid), total time of simulation, and axial direction.

The two-dimensional (z, r) energy equation for the PCM is presented in Equation (8.2) [225]:

140

T (8.2) ρCPP  div  k grad T P t

Where the subscript P represents phase change material and the k gives the thermal conductivity.

The left-hand side of Equation (8.3) [225] represents an increase with time of the energy content of an arbitrary volume, V. This is equal to the net heat transfer into V through its surrounding surface area, A. Therefore, the enthalpy equation describing this is:

d ρh dV  kP grad T n dA (8.3) t where 푛̅ is in the outward normal direction to A (area). The enthalpy method allows for easy determination of a PCM over its phase change range. By using the effective heat capacity method [230, 231], the enthalpy, h, can be broken down by temperature range (relative to the phase change temperature) as shown in Equation (8.4) [225].

CTp,s TT m1   ΔH T Tm1  TTTm1 m2 h T  Cp,L T (8.4) ΔT  m TT  CT ΔH m2  p,L Where L and s represent latent and sensible heat capacity. The lower and higher melting

temperature of PCM are represented by 푇푚1 and 푇푚2 and where ΔTm T m2 T m1

For the problem involved, the initial and boundary conditions are as follows [225]:

t 0,T  120  C (8.5)

t 0,Tf z  0,r  T f ,in (8.6) TT z 0, roo to R  0,  z  L, r to R  0 (8.7) zz T k z, r ro λ T r r  T f (8.8)  o  r

141

T z, r R 0 (8.9) r The 1st order implicit, backward, finite difference method was applied to discretize the above HTF and PCM energy equations. The details of the numerical discretization have been reported by Esen and Ayhan (1996), Visser (1986) [228, 229]. At every time step, an implicit set of simultaneous non-linear equations was solved. A matrix inversion in MATLAB combined with the ‘Sparse’ function was used to solve the equations. This option yielded a numerical code up to 20 times faster than the commonly used Gauss-Seidel method. Note, again, that this numerical model was built and validated by Tehrani et al. [225].

8.2 Integrated collector and LHTES system

As discussed in the preceding sections, the efficiency correlations of the collector were evaluated based on experimental analysis (reported in chapter 7). These efficiency correlations can then be used in a MATLAB environment along with the LHTES model so that the integrated performance of the system can be analysed for any arbitrary application. The purpose of this study is to find the best configuration for the ISC by evaluating various design alternatives by changing the number of LHTES units, total receiver area and considering vacuum and non-vacuum collectors.

8.2.1 Description of case study

A case study here is a dairy industry that requires heat energy between 120-150 ºC. This temperature range of heat energy can meet the demands of all of the diary processes, such as multiple stage evaporation (~70 ºC under a vacuum) and spray drying (conducted at ~ 120 to 125 ºC inlet air temperature) [232, 233]. Note that around 500 kWthh of heat energy (in range of 50-150 ºC) is required for producing 1 tonne of milk powder [234].

In this study, it has been assumed that the diary is located in Cobar, NSW, Australia (31°29′0″S 145°48′0″E). The annual direct normal irradiation and the mean dry bulb

142

temperature for this location are 1,995.1 kWh/m2 and 18 ºC, respectively [235]. Note that this location can represent the east coast of Australia which receives between six to eight peak hours of sunshine a day and has an annual solar exposure between 1,200 and 2,400 kWh/m2 [6]. This analysis will also assume that the thermal demand is in the 120-150 ºC temperature range and is constant throughout the day, at 50 kW for 13 hours from 6 a.m. to 7 p.m [234].

To maximize the sun captured throughout the year, the slope angle of collectors was set at the latitude angle of the location (31o), limiting the maximum longitudinal incident angle to about 23 degrees in summer or winter. According to the longitudinal IAM of the collector, shown in Figure 7.8, there will be only a maximum of ~5% loss of collector efficiency at the solstices. Thus, only transverse IAM are considered in the following annual analysis.

The thermophysical properties of the HTF and PCM used in the numerical simulations are given in Table 8.1. Since the system is above 100 °C a synthetic oil HTF, Therminol 55 [236], was used and the PCM selected herein is a eutectic mixture of

KNO3–NaNO2–NaNO3 (53–40–7) [237]. It should be noted that both of these materials were chosen since they are commercially available, cost effective, and have feasible operation temperatures.

Table 8.1 Physical parameters of the HTF and PCMs HTF and PCMs properties Unit HTF [236] PCM [237] Dynamic viscosity kg/m s 0. 00199 - Melting point ˚C - 142 Solid density kg/m3 - 1,980 Liquid density kg/m3 770 1,980 Solid Specific heat capacity J/kgK - 1,300 Liquid Specific heat capacity J/kgK 2,025 1,570 Solid thermal conductivity W/mK - 0.5 Liquid thermal conductivity W/mK 0.11 0.5 Latent heat of fusion J/kg - 80,000 Lower melting point ˚C - 139 Upper melting point ˚C - 145

143

8.2.2 Design alternatives

The primary design of the integrated storage collector consists of 6 receivers and 7 storage units, which has been called D0 in this study, as shown in Figure 8.4 (b). According to a study by Tehrani et al. [225], the number of storage units has a substantial effect on the overall performance of shell-and-tube LHTES. To include this effect, this study analyses 4 design alternatives for the LHTES unit as shown in Table 8.2. The available volume for the LHTES unit is 0.05 m3 per collector. Keeping this total volume constant and considering that the outer radius of each storage unit should not exceed 0.035 m (to avoid shading/blocking), four LHTES design alternatives are introduced in Table 8.2. In each design, the number of storage pipes and the total heat transfer surface area are different. However, the number of receiver pipes is still 6 in all design options (e.g. total receiver surface area is constant - 1.8 m2). It is expected that as the LHTES total heat transfer surface area increases, the storage unit performance can be improved, but this also adds complexity and cost to the system. Regarding the solar collector options, both vacuum and non-vacuum collectors will be analyzed. Hence, 8 design alternatives were analyzed in this study, e.g. D0-D3 for both vacuum and non-vacuum collectors.

Table 8.2 Various design alternatives for LHTES unit in one solar collector VLHTES 3 2 Design L (m) (m ) R (m) ro (m) R/ro (-) L/2ro (-) Np (-) S (m ) D0 2 0.051 0.033 0.025 1.3 40 7 2.41 D1 2 0.051 0.015 0.010 1.5 100 35 4.52 D2 2 0.051 0.01 0.0067 1.5 150 49 6.77 D3 2 0.051 0.0075 0.005 1.5 200 70 9.03 Finally, to meet the constant thermal request dictated by the dairy industry, the total number of solar collectors required should also be identified through techno-economic analysis. All combinations will be assessed for various total receiver area to find the best-case scenario. The performance analysis of this ICS system with several design configurations will be reported in the ensuring sections.

144

8.2.3 Annual performance analysis and system size

8.2.3.1 System Sizing via technical metrics

Apart from the configuration of each collector that was discussed in section 8.2.2, the total number of collectors is another design variable. The number of ICS modules determines the amount of energy that is available for storage and load. Increasing the number of ICS modules will increase both the solar fraction and the amount of available energy for the storage, but due to a number of reasons such as limited storage capacity, this number must also be optimized both technically and economically for any particular application. For the dairy application studied here (see section 8.2.1), the number of ISC modules should be varied between 100 and 180 for all design alternatives introduced in section 8.2.2. These boundaries were chosen since solar fractions below 30% (for <100 modules) were considered to be unacceptable and at >180 modules the total available collected energy often exceeds the thermal demand.

To perform the annual simulations and analysis, the experimental efficiency correlations (for vacuum and non-vacuum cases) and the IAM test results are integrated with the shell-and-tube LHTES model in the MATLAB environment.

As an illustration, the result for a typical day in January with 100 ICS modules (employing LHTES with design D0) is shown in Figure 8.5 and Figure 8.6. In Figure 8.5(a), the difference between the collector output and thermal demand (50 kW) gives the amount of available energy for storage. As can be seen, the vacuum collector provides more energy for the LHTES compared to the non-vacuum collector as is shown in Figure 8.5(b).

145

(a) (b) Figure 8.5 Comparison of vacuum and non-vacuum collector performance for a typical day in January with D0: (a) collector thermal output, (b) available energy for storage The excess energy during peak solar hours for a typical day in January will be stored and discharged later when there is a shortage of thermal energy. However, depending on the configuration of the LHTES (e.g. D0-D3 as shown in Table 8.2), the amount of stored energy is different in each design alternative. Moreover, because of space limitations for LHTES, not all of the available energy can be stored by LHTES modules as it will be shown in the remainder of this section. The amount of extra energy that has not been stored in LHTES must be dumped, a situation which should be minimized in the design. For instance, as displayed in Figure 8.6, the transient performance of the system with 100 vacuum collectors over a sunny day in January (LHTES with D0) is presented. Specifically, between hour 54 and 57 (early morning), the gas boiler supplies the heat energy to meet the demand. From hour 57 to 63, around 565 kWh of solar energy could be received by the solar collectors, where part of the absorbed energy (~350 kWh) can be directly delivered to load, and the excess energy (214 kWh) will be stored in the LHTES. However, it is apparent that only 71% of the excess energy can be stored due to the limited storage capacity of LHTES, indicating around 61.3 kWh of energy could be dumped. After hour 66, the gas boiler provides the heat energy to load again when the energy from LHTES was full discharged.

146

Figure 8.6 Transient performance of the system with 100 vacuum collectors over a typical day in January (LHTES with D0) To analyze the feasibility of the integrated system, the annual simulation of the whole system was performed for each design alternative with the number of collectors varying from 100 to 180. Figure 8.7 shows that as the total number of collectors increases, the total amount of stored energy increases. It is also clear that the overall system performance is improved when the LHTES design changed from D0 to D3. With less collector modules (e.g. 100) the difference between design alternatives is not substantial, but as the number of modules increases, the gap between design alternatives becomes wider. As expected, the annual amount of stored energy in vacuum collectors is always higher than that of non-vacuum collectors. Further analysis is required to analyze whether higher capital investment on vacuum collectors has is justified or not.

147

Figure 8.7 Total amount of annual stored energy in various design alternatives for different number of collectors One of the best metrics to evaluate the solar thermal systems is the solar fraction – a quantity which is defined as the total annual delivered energy by the solar system divided by the total annual thermal demand. For most designs (e.g. D0 to D3), a 40% increase (from 100 to 140) in the number of modules resulted in only a 5-10% increase in solar fraction, as is depicted by Figure 8.8 for both vacuum and non-vacuum collectors. The minimum solar fraction is around 35% – a design which corresponds to non-vacuum collectors (with any LHTES configuration). The vacuum collectors with LHTES that are designed based on D1-D3 accounts for the maximum solar fraction, around 65%. While a design with highest solar fraction is desirable environmentally, it may not be suitable due to the higher capital cost or other technical issues, as will be considered later.

148

Figure 8.8 Solar fraction versus the number of collectors for different design alternatives

Figure 8.9 shows the share of LHTES in terms of the total amount of delivered energy by the solar thermal system. As was expected, increasing the number of solar collectors increases the solar fraction or LHTES fraction. The minimum LHTES fraction is around 10% that corresponds to non-vacuum collectors (with any LHTES configuration). The vacuum collectors with LHTES that are designed based on D1-D3 accounts for the maximum LHTES fraction, around 40%. These results are in line with earlier results. In fact, a higher solar fraction always corresponds to higher LHTES fractions. Although the trend in both figures is linear (Figure 8.8 and Figure 8.9), the effect of number of modules is slightly sharper on the LHTES fraction, especially when it comes to non-vacuum collectors. When the number of modules was increased by 40% (from 100 to 140), the LHTES fraction was increased by 5-15 %, depending on the LHTES design.

149

Figure 8.9 LHTES fraction versus total number of collectors for various design alternatives To realize the real potential of a LHTES unit in an ISC system, it is vital to evaluate the annual charging efficiency of the LHTES system which is the real amount of stored energy divided by the total energy available for storage. Charging efficiencies below unity indicate that either the heat exchange between the PCM and HTF was not perfect or the amount of available energy was much higher than the capacity of the LHTES. The annual charging efficiency for each design alternative with different numbers of collectors is displayed in Figure 8.10. It shows that the annual charging efficiency decreases when the number of collectors increases. In fact, by increasing the number of collectors, the gap between the available capacity for storage and the amount of available energy for storage become wider. This implies that less solar collectors (e.g. 100) is desirable for this application and location. This is in clear contrast with previous results, where higher number of collectors resulted in higher solar fractions. However, this result implies that solar fraction alone is not sufficient for studying the technical feasibility of such a system. Despite the number of collectors above 100 leading to higher solar fractions, more energy will be dumped which is not desirable. For the vacuum collectors, it is evident that designs D1-D3 for the LHTES unit are better than D0, and that the performance difference between D1-D3 is negligible. Thus, D1

150

provides the best solution in this case. However, when the number of collectors is set to 100, the LHTES design has little influence on the LHTES performance for the non-vacuum collectors, so D0 is the best technical solution in this case. To make the final decision and find the best case scenario, an economic analysis will be performed in the next section.

Figure 8.10 Annual LHTES charging efficiency versus total number of collectors for various design alternatives

8.2.3.2 System sizing via economic analysis

In this section, a preliminary economic and environmental analysis of the ISC is presented to evaluate the ISC’s capability for yearly production of process heat and the relevant economic parameters of such an integrated system. More importantly, the optimum design of this ISC (the heat transfer area in LHTES modules and whether to use an evacuated absorber) is expected to be determined by combining the technical and economic performance analyses. For this economic study, fuel prices are averaged based on the energy prices of various distributors across Australia. The operation and maintenance costs and project life time are assumed according to other studies [47]. The capital cost of collectors (with/without vacuum insulation) was obtained from industrial

151

companies, are given in Appendix E. The investment cost of LHTES is estimated based on the shell and tube surface area as suggested by literature [226, 227]. All economic assumptions are tabulated in Table 8.3.

The total capital cost (CC) of ISC system can be estimated by accounting the cost of collectors (including fluid loop, e.g. pipes/pump/fittings), cylindrical LHTES tanks, PCM material and maintenance cost. The energy delivered directly by collectors ( ECOL ) and the energy delivered from LHTES tanks ( ELHTES ) can be derived from the annual simulation of whole system, as shown in Table 8.4. Moreover, the annual fuel consumption saving (FCS) can be computed as Equation (8.10) [226, 227].

3 (8.10) m (ECOL + ELHTES) η boiler FCS  year LHV

Table 8.3 Economic assumptions

Item Unit Value

Mass production (>10,000 units) cost for collector (with vacuum $/m2 589 insulated absorber) per aperture area Mass production (>10,000 units) cost for collector (with $/m2 407 non-vacuum insulated absorber) per aperture area Cylindrical LHTES tanks per heat transfer area [225] $/m2 100 PCM [225] $/ton 1200 System economic lifetime (year) [47] year 20 Fixed operation and maintenance (O&M) cost per installed kW th $/kW 20 [47] th Interest rate [6] - 6% 3 Average low heating value of natural gas (LHV) [226, 227] kWthh /m 10.5

Gas boiler efficiency ( ηboiler ) [226, 227] - 0.9 Fuel saving revenue [226, 227] $/m3 0.25 3 CO2 generation rate [226, 227] tonne/m 0.0019 CO2 abatement revenue [226, 227] $/tonne 15

Additionally, since CO2 generation rate has a linear relationship with the fuel consumption rate, the annual carbon dioxide abatement (CDA) can also be found as Equation (8.11)[226, 227].

152

tonCO CDA 2 FCS 0.0019 (8.11) year By using the data from the Table 8.3, the total capital cost (CC) of the ICS system, the annual energy produced by the collectors and LHTES, and the carbon dioxide abatement for various configurations are summarized in Table 8.4.

Table 8.4 Prediction of capital cost (CC), annual energy produced for various configurations of proposed ICS Capital cost E (kWh) E (kWh) Design Number ($) COL LHTES category of ICSs With Non- With Non- With Non- vacuum vacuum vacuum vacuum vacuum vacuum 100 151,640 117,260 85,996 70,537 16,555 5885 D0 140 204,296 156,164 93,368 83,591 30,000 17,847 180 256,952 195,068 96,966 90,449 42,222 28,875 100 172,640 138,260 85,996 70,537 21,111 6,077 D1 140 233,696 185,564 93,368 83,591 40,277 20,597 180 294,752 232,868 96,966 90,449 58,333 35,750 100 194,640 160,260 85,996 70,537 21,666 6,022 D2 140 264,496 216,364 93,368 83,591 43,055 21,422 180 334,352 272,468 96,966 90,449 62,222 37,400 100 217,640 183,260 85,996 70,537 21,944 5,940 D3 140 296,696 248,564 93,368 83,591 43,888 21,587 180 375,752 313,868 96,966 90,449 62,722 41,250 The levelised cost of heat energy (LCOH) [11], indicating the minimum price at which heat must be sold for a heat generating system to break even, is calculated for economic evaluation of the ISC system. Typically the LCOH is calculated over a 20 year lifetime, and for easy comparison with electricity rates, it is given in units of currency per kilowatt-hour (i.e. $US/kWh). Hereby, a simplified LCOH formula is used as shown in Equation (8.12)[11]:

CC CRF OMC LCOH  (8.12) Eth

153

Where CRF denotes the capital recovery factor (CRF) which is defined by [6]:

i (1 i)n CRF  (8.13) (1 i)n 1 where i is the average annual interest (discount) rate and n is the plant lifetime.

The simplified payback period (PBP) of these ISC modules can be calculated by Equation (8.14) [6].

CC PBP  (8.14) ARFCS  AR CDA

where ARFCS represents the annual revenue of fuel consumption saving, and ARFCS donates the annual revenue of carbon dioxide abatement.

Figure 8.11 shows the LCOH values for various design configurations of this ISC module. The optimum design configurations for this ISC can be found: i) 100 ISC modules with vacuum insulated receiver and D0 LHTES units; ii) 100 ISC modules with non-vacuum insulated receiver and D0 LHTES units; iii) 140 ISC modules with non-vacuum insulated receiver and D0 LHTES units; iv) 100 ISC modules with vacuum insulated receiver and D1 LHTES units. Considering the fact that using the non-vacuum collectors result in higher charging efficiencies (shown in Figure 8.10), 100 ISC modules with a non-vacuum insulated receiver and D0 LHTES units seems to be the optimum design both technically and economically.

154

Figure 8.11 Levelized cost of heat energy for various design configurations The results in Figure 8.11 shows 0.12 $/kWh of LCOH can be obtained from the current design, which can compete with solar PV and electricity heaters with LCOH ranges from 0.11-0.25 $/kWh [238, 239]. However, the LCOH results indicate that the current ISC proposals cannot compete with natural gas boilers as their LCOH is in the range 0.03-0.07 $/kWh [10, 11].

From an economic point of view, therefore, cost reductions of the proposed ISC will be required. Assuming that the main specific cost of ICS ($/m2) can be reduced by 25%, 50% and 75% though mass production and design optimization, Figure 8.12 shows the corresponding LCOH and payback time of the four optimum design configurations obtained from the techno-economic analysis. It can be seen that the LCOE of ISC can be reduced to 0.065-0.07 $/kWh through a 50% cost reduction. In this case, the industrial process heat from this ISC system can compete with natural gas without government subsidy.

155

Figure 8.12 Levelized cost of heat energy and payback time for various design configurations and specific cost of ICS module It also indicates that specific costs of 250 $/m2 (non-vacuum design) and 400 $/m2 (vacuum insulated design) could result in ~0.04 $/kWh of LCOE and ~10 years payback time from this ISC system.

8.3 Conclusion

To promote the utilization of solar thermal in industrial rooftop applications, a convenient (semi- plug-and-play) and cost-effective integrated ICS was proposed. The proposed system could represent a pathway towards providing high quality heat to a large potential industrial process heat market. Due to its compact (< 15 cm height) integrated package, the ‘stationary’ module has been optimised techno-economically for integration with shell and tube LHTES units herein. The results show that:

(1) Increasing the number of pipes improved the storage performance to some extent, but no significant economic results were observed. Hence, the LHTES design with 7 LHTES units with design D0 was identified as a preferable design;

156

(2) For a characteristic industrial heating process (i.e. a dairy which requires 500

kWhth per day at 120-150 ˚C), a solar fraction and annual charging efficiency of ~35-60% and up to 100%, respectively, are achievable with the proposed ICS system;

(3) A LCOH of 0.12 $/kWh can be achievable from the current ICS proposals, which is competitive with solar photovoltaic powered electric heaters. If the main capital cost of the ISC (collector and LHTES tank) can be reduced by 50% through mass production and/or design optimization, industrial process heat with a LCOE price of ~0.065 $/kWh can be derived from this ISC system. This price would be competitive with natural gas in many of today’s markets, even without any government subsidy.

Overall, this design represents an effective and, potentially, lower-cost approach to bringing rooftop solar thermal into industrial heating applications. The current design can be further improved by considering a PCM with better thermophysical properties, a LHTES with embedded fins to achieve a higher rate of heat exchanges, and improvements to the optical design. For instance, the transverse IAM at a large incidence angle (>30o) could be enhanced significantly by employing other optical active components, which will be detailed in the next chapter.

The economic study undertaken in this chapter was mainly for system sizing. As such, it only utilized a simple economic model which takes into account the equipment cost and operating and maintenance costs, without taking into account the installation cost of the system. Another uncertainty in this study was that the heating load was assumed to be constant for a low-to-medium temperature industrial heating application. To have a more accurate assessment of real-world economic viability, a more detailed techno-economic analysis will be performed in next chapter for a solar assisted air-conditioning application.

157

Chapter 9

9. System Techno-Economic Analysis

This chapter presents a techno-economic analysis of solar heating and cooling (SHC) system based on coupling the proposed collector with a double-effect LiBr-H2O absorption chiller. A gas-fired burner was used to supply back-up heat to the system when the solar input was not sufficient. A dynamic simulation model of the proposed configuration was developed in TRNSYS software to study its performance and economic feasibility. The annual solar fraction and economic metrics (e.g. total levelized costs) were used as selection criteria among the design options (e.g. varying the solar array, storage tank sizes).

Moreover, future improvements to proposed collector (adding prism arrays and an improved selective coating, TiNOx) were considered by performing an annual system analysis via TRNSYS.

9.1 Collector future improvements

The analysis results of chapter 7 and Figure 7.6 and Table 7.1, in particular, illustrated that the collector efficiency can be enhanced by employing a lower emissivity coated receiver (i.e. TiNOx) instead of the black chrome-coated copper tube. A TRNSYS model was developed to analyze the annul efficiency of the collector with the analysis result, as given in Figure 9.5, indicating up to 5% of annual efficiency can be enhanced by employing a TiNOx coated absorber.

Moreover, the optical analysis reported in section 3.1.3 showed that a considerable optical loss occurred at the edges of the lens, particularly when rays come in at large incidence angles (e.g. >30o), while most of light can be received when rays arrive at a small incidence angles (e.g. <15 o).

158

To enhance the daily performance of previous design, inspired by solar ‘beam steering’ technologies reported in the literature, solid prism arrays could be used to redirect light in the morning and afternoon, so that the optical efficiency of the lens / CPC combination can be improved as high incidence angles. The amount of ‘beam steering’ depends on the face angles and refractive indices of the prism, as shown in Figure 9.1.

(a) (b) Figure 9.1 Prism designed to redirect beam from right side of normal of prism normal by (a) 15 degrees and (b) 45 degrees A prism (made from low-cost PMMA) with face angles of 29o can redirect a beam with an incidence angle of 15o by 15 degrees, while a prism with face angles of 66.2o can redirect a beam by 45 degrees, as is shown in Figure 9.1 (a) and (b), respectively. Note that the face angles of prism were determined by other known parameters (e.g. incidence angles, refractive indices of the prism and redirected beam angles) according to Snell's law.

Figure 9.2 (a) shows the ray tracing for a collector operating during the middle of the day with an incidence angle of 0o (note no prism arrays are required), while Figure 9.2 (b-c) shows the ray tracing for a collector with a prism array (with face angles of 66.2o) under an incidence angle of 45o and 52.5o. The ray tracing indicates during the middle of the day, it is better to have no prism arrays (i.e. a blank section was placed above the lens). Furthermore, since the CPC can accept light from a range of angles, only a small

159

number of prism arrays is needed, with an array change for each ~ 15o step as the sun moves through the sky. To optimize the design factors (e.g. the number of prism arrays required, overall performance, and system complexity/cost), a detailed optical analysis was conducted.

(a)

(b)

(c)

Figure 9.2 Ray tracing results for (a) 0o incident light angle; (b-c) 45o and 52.5o incident light angle with the same prism array

The collector’s optical efficiency represents the portion of solar flux that reaches the absorber, and this quantity can be calculated by Equation (9.1).

i ηopt η prism η Lens ρ CPC τ g,v K(θ) (9.1)

160

It should be noted that the efficiency of prism arrays, ηprism , should be removed from the equation when calculating the optical efficiency of this system when prism arrays are not employed during the middle of the day. The combined efficiency product,

i ηprism η Lens ρ CPC , can be simulated via the Zemax software, with the required parameters

summarized in Table 9.1. Note that the value of ηLens (at normal incidence angle) was

provided by the supplier (NTKJ Co., Ltd.); τg,v represents the transmittance of vacuum

glass tube; ρCPC represents the reflectance of CPC; Superscript i is the average number of reflections in the CPC, which is assumed to be 0.8.

Table 9.1 Details of the material and optical properties along with the geometric dimensions of the components used in this collector

Symbol Value Unit Symbol Value Unit

ηlens,(θ 0 ) 0.91 (-) φprism1,θ 10 19.8 (°)

ρCPC 0.94 (-) φprism2,θ 15 28.9 (°)

τg 0.91 (-) φprism3,θ 20 37.2 (°)

τPMMA 0.92 (-) φprism4,θ 25 44.6 (°)

ηopt,(θ 0 ) 0.78 (-) φprism5,θ 30 51.1 (°)

n air 1 (-) φprism6,θ 35 56.8 (°)

n prism 1.48 (-) φprism7,θ 40 61.8 (°)

n len 1.48 (-) φprism8,θ 50 66.2 (°)

Note that the value of ηopt (θ=0) is given in Table 9.1, and that this value is the same as the optical efficiency of our previous validated collector (using lens combined with CPC). The optical efficiency at any incident angle (θ) was obtained from optical simulation carried out in Zemax software. Similar to the previous experimentally-validated optical model, the optical efficiency was defined as the fraction of power received by tube receiver divided by the total incidence solar power.

The theoretical IAM results (from ray tracing) for a collector with a variable number of prism arrays (from 1 to 8) were calculated. For instance, the results for the collector

161

employing three prism arrays (throughout half a day) are displayed in Figure 9.3, as well as a comparison with the transverse IAM of the previous design (with no prisms).

Figure 9.3 The transverse IAM of the proposed designs

These optical simulations revealed that adding 3 prism arrays (when the incidence angle is between 15-55o) can increase the optical efficiency from ~50% to ~75% when the incidence angle is 45o. Additionally, when the incident angle increases up to 55o an optical efficiency of ~64% can be achieved in the presence of this prism array (as compared to 0% previously). Figure 9.3 indicates that the half acceptance angle can be extended from 45 to 55 degrees, which should increase daily IAM significantly. The IAM results combined with the instantaneous efficiency will be used as input for further system techno-economic analysis.

Figure 9.4 (a) illustrates a 3D interior view of the proposed semi-passive tracking concentrating solar thermal collector design, with a collector cross-sectional view illustrated in Figure 9.4 (b).

162

(a)

(b)

Figure 9.4 Collector design: (a) 3D end view; (b) transverse cross-sectional view

The optical system of the design consists of prism arrays, Fresnel lens, and a CPC, which was built upon on a previous lens / CPC combination design. In this paper, a set of prism arrays was introduced which are broken up into small sections/elements (e.g. 1-10 mm widths), for mounting on a roller. As shown in Figure 9.4, the prism arrays can be scrolled over the top of the collector (driven by a stepper motor) as the solar incidence angle changes from 15o-55o. To facilitate smooth movement, two sets of guiding rollers are installed underneath the prism arrays. During the sunniest two hours

163

of a day (from 11:00 am to 1 pm, solar time), no prism array (i.e. a blank section) are placed above the collector, since sunlight is already near normal incidence during this time. When sunlight arrives at large incidence angles (e.g. >15o), prism arrays are scrolled onto the top of the collector where they can redirect the beam component more normally to the lens and (subsequently) to the CPC and the tube receiver. According to the optical analysis, it is envisioned that each prism array can operate for ~1 hour to deal with a ~ 15o change in incidence angle. This allows most of concentrated light (after passing through the lens) to be within the acceptance angle of the CPC and, thus, captured by the receiver. Consequently, five prism arrays are required (noting that due to symmetry a single prism array can be used for both the morning and afternoon). Accordingly, each prism array is replaced every hour. Stepper motors and a timing belt controlled by an Arduino microcontroller are used to facilitate this change via prism rollers, as is shown in Figure 9.4. An Arduino-UNO evaluation board and stepper motors are interfaced with a companion Adafruit Motor Shield V2 board, which drives the stepper motors through a Toshiba TB6612 MOSFET H4-bridge driver. The motor has four input wires (two wires for each of the two motor coils) which are connected to the Motor Shield board through the terminal block connector.

Using these optical and thermal efficiencies, a TRNSYS model was developed to analyze the annual efficiency of the collector employing various design configurations and operating temperatures for the meteorological data of Sydney. In this TRNSYS model, the main components included a collector, an iterative feedback controller, an auxiliary heater, a pump, and a tank. The controller activates the variable speed pump to keep the outlet temperature of the solar collectors at a fixed temperature.

The annual transient simulation results are shown in Figure 9.5, illustrating the annual thermal efficiency of various design configurations as a function of collector operating temperature. The annual thermal efficiency is defined by the fraction of total energy gained by collector to the total irradiation on the collector over a year. The simulation results were obtained by annual simulation in TRNSYS environment with different

164

input parameters, such as the outlet temperature (across 25-400 °C), efficiency

(coefficients, a0, a1 and a2) and IAM files for each configuration.

These parametric simulations revealed that a 5% annual efficiency enhancement can be achieved by employing a TiNOx-coated versus a black chrome-coated absorber. Moreover, a substantial improvement (>10%) in annual efficiency can be derived by adding two prism arrays (for the TiNOx receiver), whereas adding more than three prism arrays has diminishing returns. Thus, adding up to three prism arrays provides the best overall optical efficiency, as it covers incidence angles from 15o to 55o, which harvests most of the daily solar potential.

Figure 9.5 Collector efficiency over the year for different prism configurations as a function of operating temperature

Table 9.2 summarizes and compares the key parameters of the improved collector to our previous experimentally-validated collector. Specifically, while the peak optical

efficiency ( ηopt,(θ=0°) ) and concentration ratio can be maintained while extending the

165

working hours of collector from 6 hours to 7.3 hours per day by adding prism arrays. As a result of implementing prism arrays and the TiNOx coating, the annual thermal efficiency could be enhanced by 10%. Moreover, the thickness of collector only increased slightly, and the tracking method becomes more easy and effective. Thus, significant improvements have been achieved by adding prism arrays and employing a

TiNOx receiver.

Table 9.2 Summary of comparison to our previous experimentally-validated collector design Previous collector (Lens Improved design + CPC and employing (adding Prism and Characteristics Black Chrome coated employing TiNOx receiver) receiver) Effective concentration ratio ~4.3 ~4.3 Half acceptance angle 45 55 Average optical efficiency 65% (7.3 working 58% (6 working hours) (over the working hours) hours) Instantaneous thermal efficiency (when 46% 54% working at 200 °C) Annual thermal efficiency (when 12% 22% working at 200 °C) Thickness of collector 12 cm 13.2 cm Moving receiver/lens Rotating prism roller ~8 Tracking method ~120 mm per day times per day 9.2 Solar Absorption Cooling

In this section, a case study is developed to demonstrate the capabilities of this collector (comparing design options) for solar-assisted air-conditioning applications, using models developed by collaborators at UNSW-Australia [5, 6, 36].

Solar-assisted absorption systems – e.g. double-effect LiBr–H2O absorption chillers – can be driven by the heat output from the proposed collector. As was shown in [5, 6, 36] a medium temperature solar collector in parallel with an auxiliary gas burner represents a feasible solar thermal cooling (and heating) solution for a commercial building. For this type of system, the rated cooling capacity of the absorption chiller is selected to satisfy the maximum cooling load of the building. Figure 9.6 shows a general layout of a solar heating and cooling (SHC) absorption chiller system. Several major pieces of

166

capital equipment are needed to make this system work, including: solar thermal collectors, a storage tank, an auxiliary gas burner, an absorption chiller equipped with a wet cooling tower, cooling and heating coils, pumps, valves and a well-designed control system [5, 6, 36]. The previous and improved collectors can be used to deliver medium-temperature around ~180 °C heat to a double-effect absorption chiller, and a gas burner is employed as the auxiliary device. The solar collector arrays are connected to the stratified storage tank where excess collected solar thermal energy is stored. A variable speed drive is placed on the solar pump to control the water flow rate through the collectors to achieve a constant outlet temperature, just above the nominal driving temperature of the absorption chiller or the heating coils. Once the tank bottom temperature is equal to or greater than the collector outlet temperature, the controller turns off the pump. The hot water stored in the tank can feed either to the absorption chiller or directly through a heating coil unit for heating in the building.

During the cooling season, when the temperature of the top 75% of the storage tank is higher than the chiller’s required hot water temperature, hot water from the top of the tank is drawn by pump 2 (P2) to drive the absorption chiller. This temperature is required to boil off the refrigerant (water vapor) from the lithium bromide–water solution in the chiller’s generator(s). Once the temperature at the top of the tank drops below the required value, the auxiliary heater is switched on (bypassing the storage tank) to meet the entire energy requirement of the absorption chiller. When space heating is required, the harvested solar energy is used to supply hot water to the heating coil unit. Similarly, if the temperature of the tank drops below the required temperature, the auxiliary heater is activated to satisfy the entire heating demand. In the following, the mathematical model of the main system components is described in detail.

The collector has been modeled using ‘Type 1288’ of the TRNSYS TESS library. The thermal efficiency of the collector was given by Equation (7.11). The storage tank has been modeled using ‘Type534’ from the TRNSYS TESS library. The tank is divided into N fully-mixed isothermal segments of the equal volume where each segment

167

interacts thermally with the nodes above and below through fluid conduction and fluid movement. An energy balance of the fixed nodes in the tank is given by Equation (9.2.

dT MCi mCT T mCT T UATT (9.2) ipdt Cpi1 i Lpi1 i Tii a

th where Mi is the mass of the fluid in i section, mC and mL are the heat source

(collector loop) and load mass flow rates, and UT is the overall heat loss coefficient between the tank and environment.

The auxiliary heater is used to supply heat into either the absorption chiller or the heating coils when adequate solar energy is not available. To model the auxiliary heater, a new TRNSYS type (labeled ‘Type 223’) was developed by the co-author in C++. The heat supplied by the heater depends on the inlet temperature of the fluid and the burner set-point temperature. Applying the energy balance equation on the gas-fired burner, the water temperature at the burner outlet can be obtained as follows:

QQAH loss,AH TTAH,out AH,in (9.3) mCw ,AH p,w

where Qloss,AH represents the heat losses within the auxiliary gas-fired heater, which was determined by its overall heat loss coefficient and the thermal efficiency.

In this study, a hot water-fired double-effect absorption chiller from the Broad BDH series of chillers was selected to provide chilled water to the building. To simulate the chiller performance, a new TRNSYS type (labeled ‘Type 220’) was developed based on the adapted characteristic equation method. In this method, an adapted characteristic temperature function ( ΔΔT ), a function of the average temperature of the external water circuits, is defined as follows:

ΔΔT TG a T AC e T E (9.4)

168

where TG, TAC, and TE represent the external arithmetic mean temperatures at the generator (hot water loop), absorber condenser (cooling water loop), and evaporator (chilled water loop), respectively. The characteristic equations for evaporator and generator heat transfer rates as functions of ΔΔT can be calculated by:

QsEEEΔΔT r (9.5)

(9.6) QsGGGΔΔT r

The terms a, e, sE, rE, sG, and rG in Equation (9.5(9.6) are the characteristic coefficients which are determined using multiple regression algorithms applied to manufacturers performance data.

Applying the first law of thermodynamics to the chiller, the heat transfer rate removed

from the absorber and condenser ( QAC ) is obtained as follows:

QQQWAC E G P (9.7)

where WP is the power consumption of the solution pump within the chiller. Therefore, the COP of the chiller can be obtained by:

Q COP E (9.8)

QQG aux

The developed model requires only three average temperatures and constant flow rates of external heat carriers passing through evaporator, generator, and absorber-condenser to predict the performance of the absorption chiller.

TRNSYS ‘Type 51b’ was used to simulate the performance of a counter-flow wet (open circuit) cooling tower to provide the cooling water required to remove heat from the absorber and condenser of the absorption chiller.

169

1n m NTU c w (9.9)

ma where c and n are empirical coefficients specific to a particular tower design. These parameters were calculated based on the data obtained from the manufacturer's catalog.

The hotel building modeled in this study comes from an ASHRAE and consists of six floors and a basement with total floor area of 11,346 m2. The building has a 36.7% glazing fraction for the north side, 24.5% for the east side, 24.5% for the west side, and 26% for the south side. (Note: the total window-to-wall ratio is 30.2%). A detailed thermal analysis of the building was carried out using a pre-defined TRNSYS code ‘Type 56’, which is a multi-zone building model to determine the cooling and heating requirements.

Figure 9.6 General layout of a solar-assisted absorption chiller system for air-conditioning applications [36] The proposed SHC configurations in the present work have been modeled in the TRNSYS 17 environment. Details on the mathematical model of the main system components can be found in the work of Shirazi et al. [5, 6, 36]. The set of fixed parameters and manufacturer's data used for simulating the SHC plant are summarized in Table 9.3.

170

Table 9.3 Input parameters used for simulation of the solar -assisted heating and cooling system (SHC) Parameter Unit Value Solar collector Aperture area per collector unit m2 1.8 Mass flow rate at test condition L/hr m2 60 Optical efficiency (a0) -Current 0.78 Optical efficiency (a0) -Future 0.78 2 Loss coefficient a1 –Current/Future W/m K 0.639/0.55 2 2 Loss coefficient a2 –Current/Future W/m K 0.0045/0.003 Storage tank Aspect ratio - 3.5 Node number (N) - 10 Heat loss coefficient W/m2 K 0.83 Fluid density Kg/m3 1000 Fluid specific heat kJ/kg K 4.19 Auxiliary gas-fired heater Thermal efficiency - 0.9 Heat loss coefficient kJ/h K 30 Absorption chiller CHW temperature (inlet/outlet) °C 14/7 CW temperature (inlet/outlet) °C 30/37 HW temperature (inlet/outlet) °C 180/165 CHW mass flow rate m3/hr 143 CW mass flow rate m3/hr 244 HW mass flow rate m3/hr 51 Cooling capacity kW 1163 Pump electricity consumption kW 6.8 COP - 1.41 Cooling tower Sump volume m3 0.2 Air flow rate (rated) m3 /hr 207250 Fan electricity consumption (rated) kW 15 c coefficient - 1.62 n coefficient - −0.62

9.3 SHC System Analysis

The solar heating and cooling absorption chiller system modeled in this section is applied to the reference hotel building in Sydney, Australia. Figure 9.7 shows the annual total cooling and heating demand of the building. The maximum cooling and heating demand of the building is 965 kW and 520 kW, respectively. Accordingly, an

171

absorption chiller with a nominal cooling capacity of 1163 kW has been selected, and a nominal heating capacity of the auxiliary heater of 920 kW was selected.

Figure 9.7 The load profile of the hotel building throughout a year in Sydney In this section, a detailed parametric study is performed for the proposed SHC plants to evaluate annual solar fraction with respect to the collector area and storage tank volume and to determine an optimum size of each configuration. It should be noted that the units for collector area and storage tank volume were normalized to ensure the analysis is more generally applicable.

The solar field size was varied from 1 to 7 m2/kWc of the absorption chiller rated capacity, while the storage tank volume was varied from 10 L/m2 to 175 L/m2 of collector aperture area. The total solar fraction calculated for the SHC1 (i.e. using the current collector) and SHC2 (i.e. using the improved, future collector) layouts as a function of storage tank specific volume ratio for different collector specific areas (A) are given in Figure 9.8.

172

(a)

(b) Figure 9.8 Solar fraction variation with specific storage tank volume ratio for different collector specific areas for: (a) SHC1 and (b) SHC2 layouts

173

It is evident from these plots that the solar fraction increases as the collector area increases. However, at high solar fraction values, increasing the collector area results in diminishing returns since the demand for heat is limited by the cooling and heating demand of the building which can no longer use the extra heat. Moreover, as shown in Figure 9.8, the solar fraction improves as the size of the tank increases to an optimum point, indicating that there is no advantage in going beyond a certain storage volume. It should be mentioned that the solar fraction of these configurations is also less sensitive to the volume of the tank for lower collector areas.

Thus, in order to achieve an ~40-50% solar fraction, as given in Table 9.4, the specific collector area and the specific tank volume were identified to be 7 m2/kW and ~30 L/m2 for the prototype collector (i.e. the current design), respectively. For the improved (future) collector, the ‘best’ system design parameters were identified to be 3 m2/kW and ~40 L/m2 of storage. The system sizing results show that the improved collector is likely to be much cheaper since it provides the same solar fraction for a significantly reduced number of collectors and storage tank volume. This saving is explained by the superior annual efficiency of the improved collector design as shown in Figure 9.5.

Table 9.4 System sizing results SHC1 (Current SHC2 (Future Parameter Unit collector) improved collector) Solar fraction - 0.42 0.51 Solar collector specific m2/kW 7 3 area Solar collector area m2 8,141 3,489 Specific tank volume L/m2 30 40 Tank volume m3 244.2 139.56 However, although the improved collector design (adding prism arrays) can significantly reduce number of collectors, they add to the system’s complexity and cost. An economic analysis will be performed in the next section to determine the more profitable system configuration

174

9.4 Economic Analysis

In order to evaluate the economic feasibility of the solar double-effect absorption chillers compared to an electric chiller, an economic analysis was conducted on the proposed SHC configurations – one using the current collector and the other employing the future improved collector. To perform an economic analysis on the proposed configurations, various financial assumptions were made, as summarized in Table 9.5 [5, 6, 36]. Note that some parameters (e.g. auxiliary heater heating capacity, cooling capacity of cooling tower, heat transfer area of cooling/heating coil) were determined by calculations, while some results (e.g. annual electricity/gas consumption) in this table were obtained from the TRNSYS simulation by using the system sizing results from Table 9.4.

While a simplified levelized cost of heat energy (LCOE) formula was defined in Equation (8.12), a more detailed model was applied in this system analysis, as given by Equation (9.10) [11], which takes into account the capital cost of the entire system (CC), operating and maintenance costs (OMC), and the fuel cost (FC, including electricity and gas consumption) [6].

CC CRF  OMC  FC LCOE  (9.10) EECH

Where CRF denotes the capital recovery factor (CRF) which was defined by Equation (8.13).

175

Table 9.5 Economic analysis - assumptions and justifications Item Unit Value Justifications Mass production (>10,000 units) cost for collector per aperture area—Current USD/m2 589 Appendix E collector Mass production (>10,000 units) cost for collector per aperture area—Improved USD /m2 709 Appendix E future collector (adding prism arrays) Equipment cost for low pressure storage USD/m3 708 [6] tank (3 bar) per unit Equipment cost for auxiliary heater Capital cost function (heating capacity is shown in the following USD/ kW 77 th can be found in [6] row) Auxiliary heater heating capacity, QAH kW 920 Calculated result [6] Double-effect absorption chiller USD/ kWth 536 [6] Cooling tower (capacity is shown in the Capital cost function USD 177,226 following row) can be found in [6] Calculated result based Cooling capacity of cooling tower, Q kW 1,990 CT on equation in [36] Capital cost function Cooling/Heating coil (Heat transfer area) USD 46,275 can be found in [6] Heat transfer area of cooling/heating coil, 455.1 m2 Calculated result ACC and AHC and 81.9 Capital cost function Pump and controllers USD 52,516 can be found in [6] System lifetime year 20 [6, 47] %(of Maintenance cost of the SHC plant 1.25 [6] capital cost) Installation, integration, and piping cost of %(of 150 [6, 47] the SHC plant capital cost) Interest rate - 6% [6] Average business electricity price USD /kWh 0.2 [6] Annual electricity consumption MWh 138.5 Simulation result Average business natural gas price USD /GJ 14.4 [6] Annual gas consumption MWh 845.07 Simulation result The capital cost of the entire system (CC) can be calculated by accounting for all system components. Moreover, the installation, integration, and piping cost of the SHC plant, as shown in Figure 9.9, should be taken into account as part of the capital cost calculation.

176

Figure 9.9 Comparison of the capital cost of the proposed SHC systems (employing both the current and improved collector designs) The LCOE described in Equation (9.10) enables economic returns on both heating and cooling energy. However, it can also be separated by the levelized cost of heating energy (LCOH) and levelized cost of cooling energy (LCOC).

As such, the levelized cost of heating energy (LCOH) can be calculated by Equation (9.11) and the levelized cost of cooling energy (LCOC) can be calculated by Equation (9.12):

CC CRF  OMC  FC LCOH  (9.11) ECH COP  E

CC CRF  OMC  FC LCOC  (9.12) EC

177

The results in Figure 9.10 show 0.99 $/kWh of LCOC can be obtained from the current design, and 0.61 $/kWh of LCOC can be achieved by employing the improved collector. Note that 0.34-0.5 $/kWh and 0.68-0.96 $/kWh were reported by Cabrera et al. [241] for a hotel in Madrid and Copenhagen by employing solar -assisted double-effect absorption chillers for ETC and PTC systems, respectively.

The LCOC results indicate that if the specific cost of a SHC system ($/kW) can be reduced by 25% more, the LCOC of the SHC system employing the improved collectors can reach 0.41 $/kWh, which would just compete with standard electric chillers (LCOC in the range 0.15-0.45 $/kWh [242]).

Figure 9.10 SHC system LCOC and LCOE for the current and improved collector designs using various specific capital costs 9.5 Conclusion

In this chapter, improvements to the proposed collector were reported and analysed, with respect to an annual solar heating and cooling system performance using TRNSYS.

The results revealed that adding prism arrays and a TiNOx coating can enhance the

178

annual performance of proposed collector by ~100% and, consequently, yield significantly better economics.

The solar field was sized to cover ~50% of both the cooling and heating demand of a large hotel building (in Sydney) through solar energy. A specific collector area of 7 m2/kW or 3 m2/kW was required for the current and improved collector designs, respectively. Moreover, the study revealed that while the current collector could supply cooling and heating energy with a LCOE of 0.81-0.99 $/kWh, the improved collector could achieve a LCOE of 0.5-0.6 $/kWh. This is a big improvement, but a further cost reduction of around 25-50% would still be needed to be competitive with electric chillers, which have an LCOE of 0.15-0.42 $/kWh.

Overall, it was found that with selected improvements and cost reductions, the proposed collector is nearing the range of economic viability for solar heating and cooling applications. Additionally, the practical application of the proposed tracking/concentrating platform could be extended to many other existed stationary or full tracking concentrated solar systems. For instance, by implementing this tracking/concentrating platform, a stationary rooftop concentrated PV can be developed, and the acceptance angle of a CPC collector can be enlarged and the seasonal IAM effect on a daily tracking PTC system can be reduced significantly.

179

Chapter 10

10. Conclusion and Recommendations for Future Work

Medium temperature rooftop integrated solar thermal systems (100-400 °C) can open up a huge – as yet untapped – market to meet the demand of high quality heat for industrial heating, air-conditioning and commercial steam applications. However, compared to other energy sources (i.e. solar PV), they are under-developed and their commercial value is less recognized. In this thesis, a number of innovative technologies, including optical/thermal concentration, semi-active tracking, direct absorption solar absorbers, and latent heat storage/collector integration have been explored to develop an efficient, low-profile rooftop solar thermal collector. Detailed numerical and experimental investigations were carried out to assess the performance of these designs. The analysis revealed that the final design of the proposed collector can achieve ~50% instantaneous thermal efficiency and ~20% annual efficiency operating at 200 °C in Sydney, Australia. A system-level techno-economic analysis also demonstrated its feasibility to deliver low cost heating/cooling energy with a LCOE of 0.12-0.6 $/kWh for industrial processes and commercial HVAC systems. Thus, the proposed collector can be considered as one of the only viable alternatives for integrating into rooftops and for supplying thermal energy to these applications (in the 100-250 °C range).

Among the technologies that were explored and developed in this thesis, non-imaging optics are likely the key component for achieving non-tracking rooftop concentration. The optical designs reported in this thesis represent the first time a non-imaging Fresnel lens combined with a CPC has been used to create a compact solar concentrator (<10 cm) with an internal tracking mechanism (realized by moving the lens or CPC horizontally without rotational tracking). Thus, the design can both increase the concentration ratio and maintain the large half acceptance angle (~45°, ~6 hours per

180

day). Upon this design, further improvements were explored to extend the operating hours to 8 hours per day by adding prism arrays (i.e. ~3-4 arrays) on the top of the lens to capture a further 15° or 30° of incidence (i.e. more sunlight hours). The combination of prism arrays, Fresnel lenses, and CPCs in the final design represents a truly innovative tracking/concentrating platform, which could significantly increase the daily/annual efficiency of solar thermal collectors.

In this work, thermal concentration was also explored as an alternative approach to boost the performance and cost effectiveness of solar energy collection systems. All possible thermal concentration configurations and their fundamental and realistic engineering limits were classified and subsequently analyzed. The findings revealed that while a heat pipe with a carbon nanotube (CNT) absorber was the most promising ‘passive’ concentrator, thermoelectric solar thermal collectors showed the highest development potential as ‘active’ thermal concentrators. In the proposed thermoelectric solar thermal system, the electricity and solar inputs can both be concentrated and harvested while minimizing the heat loss on the cold side of the system (e.g. a solar absorber).

This research also uncovered and quantified the non-negligible optical/heat loss mechanisms involved in a real system by investigating/comparing volumetric and surface absorbers. As such, the study provides guidance and highlights the challenges in the field of volumetric solar absorption systems for providing high temperature, industrial heat. Thus, a vacuum insulated volumetric absorber consisting of a MWCNT nanofluid contained within an ITO-coated glass tube was developed for the first time in this thesis (i.e. to reduce the radiative heat loss from these types of absorbers). However, an inferior performance of this volumetric receiver (e.g. an efficiency of 32% at an operation temperature of 200 oC) was observed when compared with a conventional selective surface absorber (e.g. an efficiency of 47% at 200 oC), due to higher reflective optical losses and higher radiative heat losses from the surface of the glass tube. Although it was not seen in our prototype, a considerable inverse

181

temperature (Tf < Ts) difference might also be designed into these absorbers to close the efficiency gap between the volumetric and surface absorbers. These findings demonstrate that both receivers could work for industrial heating applications, but volumetric absorbers will require anti-reflective and good selective coatings to be competitive with surface absorbers. If these challenges can be overcome, nanofluid receivers may yet provide an effective and low-cost approach to bring nanotechnology into industrial heating and air-conditioning applications since glass-to-glass vacuum sealing is easier to achieve than metal-to-glass.

Furthermore, the optical simulations revealed that this class of ‘stationary’ rooftop collectors has ‘free space’ which could be beneficially employed as thermal storage. Thus, the proposed concentrating solar collector was simulated with an integrated shell-and-tube Latent Heat Thermal Energy Storage (LHTES) system, to understand how this may assist industrial heating applications. It was found via a techno-economic analysis that compared to separated, sensible tanks (i.e. conventional storage technology) this convenient (semi-plug-and-play) integrated package could significantly reduce the complexity, installation cost/challenges, piping energy loss, and maintenance costs. Based on the economic analysis, a LCOH of 0.12 $/kWh can be achieved from the ICS, which is competitive with solar photovoltaic powered electric heaters. Though mass production/design optimization, the industrial process heat from this integrated collector/storage system can, ideally, compete with natural gas even without any government subsidies.

Lastly, the annual performance and economic feasibility of the proposed collector was simulated for solar assisted heating and cooling applications in the TRNSYS (Transient System Simulation Tool) software environment. An economic analysis of the system indicated that an LCOE price of ~0.60 $/kWh can be derived from this collector when used to drive a double-effect absorption chiller on a typical hotel building in Sydney’s climate. It was found that with a few improvements and cost reductions, it may be possible to produce a system based on the proposed collector which is competitive with

182

electric compression chiller systems in Sydney (0.15-0.42 $/kWh). It should be noted that cost will likely change considerably (although somewhat proportionally to conventional HVAC costs) in different countries and climates. The capital cost of equipment would be 30-50% cheaper in countries such as China due to the fact that equipment and labor cost in China are much cheaper than in Australia.

In conclusion, the research objective of developing innovative solar thermal harvesting technology for industrial needs has been met. Moreover, this thesis provides several significant advances towards effective and, potentially, lower-cost rooftop solar thermal harvesting technology for industrial heating and air-conditioning applications.

As was briefly mentioned in many of the Chapters above, there is still room for further optimisation in several of the concepts developed here, before implementing this technology in commercial applications. Thus, recommendations for future work based on the findings of the project are listed below:

1) Optimization and analysis of the collector design upon implementing integrating prism arrays. Feasible structural solutions to employ these prism arrays will need to be finalized. Either mounting the prism sections/elements (i.e. 1-5 mm width) on a roller chain or making a micro size prism array on a flexible film depends on the achievable optical performance and manufacturing costs;

2) Even though the current volumetric absorbers have shown inferior performance, the cost advantage of avoiding the metal-glass vacuum seal (see Appendix E) may make these absorbers viable (if further developments on selective and anti-reflective coatings can be incorporated). In this case, developing a low-to-medium temperature (<200 °C) glass-glass vacuum sealed surface absorber (high absorptivity coating on inner glass outer surface) would be an alternative and cost-effective approach;

3) Further research work on developing a low profile thermoelectric solar thermal collector which can ‘pump’ medium temperatures up to 250 °C may also be

183

desirable;

4) Installation, integration, and piping costs of the solar heating and/or cooling plant are clearly a big part of the total capital cost (see Table 9.5). Integrating latent heat thermal storage in the proposed collector will bring a significant cost reduction to the total plant, but further techno-economic analyses to compare a built-in Shell and Tube Latent Heat Thermal Energy Storage (LHTES) with a separate latent storage system should be investigated to fully quantify the benefit;

5) If the built-in Shell and Tube LHTES has significant cost-effectiveness, the collector design will need to be upgraded by employing both prism arrays and built-in LHTES.

6) As a final step, noting all of the above, a demonstration plant could be built for industrial heating or solar thermal cooling applications to assess all the practical issues of collector design and plant system integration before commercialization and mass production.

184

Appendix A ---Tracking code

#include byte Minute; #include byte Second; //* This is a test sketch for the Adafruit assembled byte Day; Motor Shield for Arduino v2 */ byte Month; // Jan is month 1 #include byte Year; #include #include "utility/Adafruit_PWMServoDriver.h" //------// #include boolean StopFlag=1; #include boolean Restart=0; //#include "DS1307RTC.h" //to be added when the //------// external IC is integrated on the board #include //functions prototypes #include void parser(void); void updateSpeed_M1(int m1_speed); // create the motor shield object with the default void updateSpeed_M2(int m2_speed); I2C address void updateMotorSpeed(void); Adafruit_MotorShield AFMS = void Adafruit_MotorShield(); SetMotorSpeed(int*m1_Speed,int*m2_Speed); // Connect a stepper motor with 200 steps per void updateDirection_M1(int* m1_Direction); revolution (1.8 degree) to:// void updateDirection_M2(int* m2_Direction); motor port #1 (M1 and M2) Adafruit_StepperMotor *myMotor_1 = void readInitialData(void); AFMS.getStepper(200, 1); void readInputString(void); motor port #2 (M3 and M4) void Adafruit_StepperMotor *myMotor_2 = SetMotorDirection(int*m1_Direction,int*m2_Dir AFMS.getStepper(200, 2); ection); void Initialization(void); //-----Defines debug------// void SetMotorStep(int* pm1_Step,int* pm2_Step); #define xxxDebugSerialInput void MoveToStartPos(void); //-----Defines Timer------// void TimeDateManagement(void); #define TIME_MSG_LEN 28 void digitalClockDisplay(void); #define TIME_HEADER 'T' void printDigits(int digits); // Header tag for serial time sync message void TimerFunction(void); //Global variables initialization void TimerManagement(void); int initialPosition =0; boolean CheckEndTime(void); int tilt =0; void returnInitialPosition(void); byte StartHour=8; void OpticalSensorFunction(void); boolean runTest=true; //Set Up int m1_Speed=100; void setup() { int m2_Speed=100; // set up Serial library at 9600 bps int m1_Step=200; Serial.begin(9600); int m2_Step=200; Serial.println("Set Up System"); int m1_Direction=FORWARD; //create with the default frequency 1.6KHz int m2_Direction=FORWARD; AFMS.begin(); //------TIME------// time_t time, timeNow, updateTime; //set the motor speed tmElements_t tm; //myMotor_1->setSpeed(200); // 10 rpm //if it byte Hour; is included, motor_1 vibrates but doesn't rotate

185

myMotor_2->setSpeed(200); // myMotor_2->step(*m2_Step, m2_Direction, DOUBLE); //read initial data } readInitialData(); //------MOVE to Start Position ------// //read initial settings from SerialMonitor void MoveToStartPos(){ Initialization(); //local variables byte LHour; //initialize time and date byte LMinute; TimeDateManagement(); byte LSecond; //Move motor to initial position int TotalSeconds=0; MoveToStartPos(); int m2_Step_Start; //Initialize timer for interrupt routine //get current time and convert in total number of TimerManagement(); seconds //Alarm.delay(1000); LHour=hour(); //Serial.println(); LMinute=minute(); //Serial.println("System in RunMode, use parser to LSecond=second(); update step/direction"); //total number of seconds elapsed from StartHour //Alarm.timerRepeat(10, TimerFunction); TotalSeconds=(LHour-StartHour)*3600+LMinut } e*60+LSecond; void loop() { //Serial.println("Double coil steps"); m2_Step_Start=round(TotalSeconds/100); myMotor_1->step(m1_Step, m1_Direction, Serial.print("current hour="); DOUBLE); Serial.println(LHour); //delay (1000); Serial.print("TotalSeconds="); //myMotor_2->step(100, FORWARD, DOUBLE); Serial.println(TotalSeconds); parser(); Serial.print("Number Start Steps = "); //Serial.println("Double coil steps"); Serial.println(m2_Step_Start); //myMotor_2->step(100, FORWARD, DOUBLE); myMotor_2->step(m2_Step_Start, FORWARD, //delay (3000); DOUBLE); if (CheckEndTime()){ } returnInitialPosition(); //------DIRECTION------// Restart=1; void updateDirection_M1(int* m1_Direction){ } myMotor_1->step(m1_Step, *m1_Direction, if(Restart){ DOUBLE); TimeDateManagement(); } //STOP TIMER void updateDirection_M2(int* m2_Direction){ MoveToStartPos(); myMotor_2->step(m2_Step, *m2_Direction, //start timer DOUBLE); TimerManagement(); } Restart=0; //------SPEED------// } void SetMotorStep(int* pm1_Step,int* pm2_Step){ //checkTemperature(); //to be implemented //define Alarm.delay(1); #define MOTOTR_1 '1' } #define MOTOTR_2 '2' //------MOTOR FUNTIONS------// //Local Var //------POSITION------// String MotorSel; int MotorStep; void updateStep_M1(int *m1_Step){ //Begin myMotor_1->step(*m1_Step, m1_Direction, Serial.println(); DOUBLE); Serial.println("Select Motor please M1,M2 (S to } skip)"); void updateStep_M2(int *m2_Step){ //wait until new data has been entered from user while (!Serial.available()){}

186

MotorSel=Serial.readString(); //feedback on monitor if(MotorSel.equals("S")){} Serial.print("MotorSel="); else{ Serial.println(MotorSel); Serial.println("Enter motor step "); Serial.print("Motor Direction="); Serial.println(); Serial.println(MotorDirection); //wait until a valid character is available from console if (MotorSel=="M1"){ while (!Serial.available()){} *m1_Direction = MotorDirection; //speedVal=Serial.read(); }else{ MotorStep=Serial.parseInt(); *m2_Direction = MotorDirection; //feedback on monitor } Serial.print("MotorSel="); } Serial.println(MotorSel); } Serial.print("MotorStep="); //------Test Speed------// Serial.println(MotorStep); void SetMotorSpeed(int*m1_Speed,int*m2_Speed){ if (MotorSel=="M1"){ //define *pm1_Step = MotorStep; #define MOTOTR_1 '1' updateStep_M1(pm1_Step); #define MOTOTR_2 '2' }else{ //Local Var *pm2_Step = MotorStep; String MotorSel; updateStep_M2(pm2_Step); int MotorSpeed; } //Begin } Serial.println(); } Serial.println("Select Motor please M1, M2 (S to void skip)"); SetMotorDirection(int*m1_Direction,int*m2_Dir //wait until new data has been entered from user ection){ while (!Serial.available()){} //define MotorSel=Serial.readString(); #define MOTOTR_1 '1' if(MotorSel.equals("S")){} #define MOTOTR_2 '2' else{ //Local Var Serial.println("Enter desired motor speed, String MotorSel; please"); int MotorDirection; Serial.println(); //Begin //wait until a valid character is available from Serial.println(); console Serial.println("Select Motor please M1, M2 (S to while (!Serial.available()){} skip)"); MotorSpeed=Serial.parseInt(); //wait until new data has been entered from user while (!Serial.available()){} //feedback on monitor MotorSel=Serial.readString(); Serial.print("MotorSel="); if(MotorSel.equals("S")){} Serial.println(MotorSel); else{ Serial.print("Motor Speed="); Serial.println("Enter motor direction Serial.println(MotorSpeed); please"); Serial.println(); if (MotorSel=="M1"){ *m1_Speed = MotorSpeed; //wait until a valid character is available from updateSpeed_M1(m1_Speed); console }else{ while (!Serial.available()){} *m2_Speed = MotorSpeed; MotorDirection=Serial.parseInt(); updateSpeed_M2(m2_Speed); }

187

} case MOTOTR_2: } Serial.println("debug_case2"); //------SPEED------// updateSpeed_M2(&speedVal); void updateSpeed_M1(int* m1_speed){ Serial.print("Motor2 speed updated = myMotor_1->setSpeed(*m1_speed); "); } Serial.println(speedVal); break; void updateSpeed_M2(int* m2_speed){ myMotor_2->setSpeed(*m2_speed); default: } Serial.println("debug_default"); //------// break; void updateMotorSpeed(){ } //define } #define MOTOTR_1 '1' } #define MOTOTR_2 '2' //Local Var //END MOTOR FUNCTIONS------// String MotorSel; void parser(){ char MotorSelSwitch; //used to update Motor settings int speedVal; #define M1SPEED 1 #define M2SPEED 2 //check if new data has been entered from user #define M1STEP 3 if (Serial.available()){ #define M2STEP 4 //MotorSel= Serial.read(); #define M1DIR 5 MotorSel=Serial.readString(); #define M2DIR 6 Serial.println("Enter new motor speed, please"); //Local Var Serial.println(); String MotorCommand; int MotorParser; //wait until a valid character is available from int value; console //check if new data has been entered from user while (!Serial.available()){} if (Serial.available()){ //speedVal=Serial.read(); MotorCommand=Serial.readString(); speedVal=Serial.parseInt(); Serial.println(MotorCommand); Serial.println("speed entered"); Serial.print("MotorSel="); Serial.println("Enter value, please"); Serial.println(MotorSel); Serial.println();

if (MotorSel=="M1"){ //------// MotorSelSwitch = MOTOTR_1; }else{ if(MotorCommand.equals("V1")){MotorParser= MotorSelSwitch = MOTOTR_2; M1SPEED;} } else if(MotorCommand.equals("V2")){MotorParser= switch (MotorSelSwitch){ M2SPEED;} case MOTOTR_1: else Serial.println("debug_case1"); if(MotorCommand.equals("S1")){MotorParser= updateSpeed_M1(&speedVal); M1STEP;} Serial.print("Motor1 speed updated = else "); if(MotorCommand.equals("S2")){MotorParser= Serial.println(speedVal); M2STEP;} break;

188

else if Serial.println("debug_case S2"); (MotorCommand.equals("D1")){MotorParser=M m2_Step=value; 1DIR;} break; else if default: (MotorCommand.equals("D2")){MotorParser=M Serial.println("debug_default"); 2DIR;} break; else {Serial.println("Wrong command } entered");} } } //wait until a valid character is available from //------USER INTERFACE------// console void readInputString(){ while (!Serial.available()){} String stringIn; while (!Serial.available()){} value=Serial.parseInt(); stringIn=Serial.readString(); Serial.println(stringIn); switch (MotorParser){ } case M1SPEED: Serial.println("debug_case V1"); void readInitialData(){ m1_Speed=value; /* updateSpeed_M1(&m1_Speed); while (!Serial.available()){}//wait until a valid break; character is available from console initialPosition = Serial.parseInt(); case M2SPEED: Serial.print("initialPosition = "); Serial.println("debug_case V2"); Serial.println(initialPosition); m2_Speed=value; */Serial.println("set the tilt angle (in degree) "); updateSpeed_M2(&m2_Speed); while (!Serial.available()){/*wait until a valid break; character is available from console*/} case M1DIR: tilt = Serial.parseInt(); Serial.println("debug_case D1"); Serial.print("tilt = "); Serial.println(tilt); if(value==1){m1_Direction=FORWARD;}else Serial.println("set the start hour"); {m1_Direction=BACKWARD;} while (!Serial.available()){/*wait until a valid Serial.println(m1_Direction); character is available from console*/} //direction will be changed when StartHour =Serial.parseInt(); timer expires Serial.print("StartHour = "); break; Serial.println(StartHour); } case M2DIR: Serial.println("debug_case D2"); //------END USER INTERFACE------// void Initialization(){ if(value==1){m2_Direction=FORWARD;}else //define {m2_Direction=BACKWARD;} #define PROCEEDE_TEST 'Y' //m2_Direction=value;FORWARD //Local Var Serial.println(m2_Direction); char proceed; //direction will be changed when timer expires //initialize Speed with default values break; //updateSpeed_M1(&m1_Speed); ////if it is case M1STEP: included, motor_1 vibrates but doesn't rotate Serial.println("debug_case S1"); updateSpeed_M2(&m2_Speed); m1_Step=value; Serial.println("Default Setting:"); //step value will be issued when timer expires Serial.println("Motor M1,M2 resolution=1.8d break; (#Resol=200)"); case M2STEP:

189

Serial.println("Motor M1,M2: Speed=100; //Serial.println("here"); Direction= Forward; #Step=200"); time_t inputTime = 0; Serial.println(); for(int i=0; i < (TIME_MSG_LEN -1); i++){ Serial.println("Direction Setting"); //Serial.print("read c = "); Serial.print("FORWARD= "); c = Serial.read(); Serial.println(FORWARD); //Serial.println(c); Serial.print("BACKWARD= "); delay(10); Serial.println(BACKWARD); if( c >= '0' && c <= '9'){ //proceed changing setting? Y/N timeArray[j] = (c - '0') ; // convert Serial.println("Do you want change settings? char-digits to a number Y/N"); //Serial.println(timeArray[j]); while (!Serial.available()){} j++; proceed=Serial.read(); } } if(proceed==PROCEEDE_TEST){runTest=true;} } else{runTest=false;} //-----update tm var------// while (runTest){ //Serial.println("Resolution"); tm.Hour //Set Motor Resolution =(uint8_t)(timeArray[0]*10+timeArray[1]); (&m1_Resolution,&m2_Resolution); //Serial.print("tm.Hour="); //getStepper() rather than in the object declaration //Serial.println(tm.Hour); //Set Motor Speed tm.Minute = SetMotorSpeed(&m1_Speed,&m2_Speed); (uint8_t)(timeArray[2]*10+timeArray[3]); //Set motor direction //Serial.print("tm.Minute="); SetMotorDirection(&m1_Direction,&m2_Directi //Serial.println(tm.Minute); on); tm.Second =(uint8_t) //Set motor steps (how many steps to do each time (timeArray[4]*10+timeArray[5]); 1min elapses //Serial.print("tm.Second="); SetMotorStep(&m1_Step,&m2_Step); //Serial.println(tm.Second); //proceed with test? Y/N tm.Day Serial.println("Do you want change settings? =(uint8_t)(timeArray[6]*10+timeArray[7]); Y/N"); //Serial.print("tm.Day="); while (!Serial.available()){} //Serial.println(tm.Day); proceed=Serial.read(); tm.Month = if(proceed==PROCEEDE_TEST){runTest=true;} (uint8_t)(timeArray[8]*10+timeArray[9]); else{runTest=false;} //Serial.print("tm.Month="); } //Serial.println(tm.Month); } tm.Year = void TimeDateManagement(){ (uint8_t)(2000+(timeArray[10]*10+timeArray[1 //local variables 1]) - 1970); tmElements_t tm; //Serial.print("tm.Year="); byte timeArray[12]={0}; //Serial.println(tm.Year); int j,k=0; //synchronize Arduino time //read time from serial updateTime = makeTime(tm); Serial.println("Set Time: hh:mm:ss:dd:mm:yy"); setTime(updateTime); //Time: 18:20:00:19:07:15 //display time on Monitor // if time sync available from serial port, update digitalClockDisplay(); time and return true while(!Serial.available()){} //Alarm.timerRepeat(1, TimerFunction); char c = Serial.read() ; Serial.println("Time and Date initialized"); //Serial.println(c); } if( c == TIME_HEADER ) {

190

void TimerManagement(){//Alarm.delay(10); Serial.print(" "); //Timer elapses each 60sec Serial.print(year()); Alarm.timerRepeat(60, TimerFunction); Serial.println(); //Alarm.timerRepeat(seconds, function); } Serial.println("Timer initialized"); void printDigits(int digits) } { //----INTERRUPT SERVICE ROUTINE------// Serial.print(":"); void TimerFunction(void){ if(digits < 10) Serial.println("1 min elapsed"); Serial.print('0'); //myMotor_1->step(m1_Step, m1_Direction, Serial.print(digits); DOUBLE); } //digitalClockDisplay(); boolean CheckEndTime(){ //myMotor_2->step(100, FORWARD, byte currentHour; DOUBLE); currentHour=hour(); myMotor_2->step(m2_Step, m2_Direction, if (currentHour==18){ DOUBLE); Serial.print("current Hour is ="); } Serial.print(currentHour); void OpticalSensorFunction(void){ Serial.println(" time to stop tracking!"); StopFlag=0; return 1; } } void digitalClockDisplay() else { return 0; // digital clock display of the time } Serial.print(hour()); void returnInitialPosition(){ printDigits(minute()); m2_Direction=BACKWARD; printDigits(second()); Serial.println("Moving Backward"); Serial.println(); while (StopFlag){ myMotor_2->step(m2_Step, m2_Direction, Serial.print(" "); DOUBLE); Serial.print(dayStr(weekday())); } Serial.print(" "); Serial.println("Finish Loop"); Serial.print(day()); StopFlag=1; Serial.print(" "); m2_Direction=FORWARD; Serial.print(monthShortStr(month())); }

191

Appendix B--- A MATLAB code for analysis of heat pump

% Vapour Compression Refrigeration Cycle for Thermal Concentration Purposes % in Solar Thermal Collectors clear all clc % Input parameters: % Collector: G = 900; % Solar irradiation on a titlted surface [W/m2] a0 = 0.81; a1 = 1.77; % [W/m2 K] a2 = 0.013; % [W/m2 K2] A_a = 50; % Collector aperture area [m2] T_amb = 25; % ambient temperature [C] T_coll = 70; % Collector inlet temperature [C] % Vapour Compression Refrigeration System: r_p =2; % compressor pressure ration eta_comp = 0.8; % compressor isentropic efficiency dp_evap = 0.05; % pressure drop along the condenser [-] dp_cond = 0.05; % pressure drop along the condenser [-] U_cond = 4166; % Overall heat transfer coefficient [W/m2 K] % storage tank: T_tank_bott = 75; % the return DHW temperature from the load [C] V_tank = 0.5; % The tank volume [m3] RT = 30; % residence time [minute] FR_w = (V_tank*1000)/((RT/60)*3600); % water flow rate [kg/s] cp_w = 4.19; % water specific heat at constant pressure [kJ/kg K] T_coll_mean = T_coll; % [C] eta_coll = a0 - a1*(T_coll_mean-T_amb)/G-a2*((T_coll_mean-T_amb)^2)/G; Q_coll = eta_coll*G*A_a/1000 % collector useful heat [kW] t_evap = T_coll; % refrigerant saturation temperature [C] % state 1: saturated vapoor t_1 = t_evap; % Refrigerant temperature at the evaporator outlet [C] T_1 = t_1 + 273.15; % [K] x_1 = 1; % quality of the refirgerant (state: saturated vapour) [kg/kg] p_1 = refpropm('P','T',T_1,'Q',x_1,'R134a'); %pressure %[kPa] h_1 = refpropm('H','T',T_1,'Q',x_1,'R134a')/1000; %enthalpy %[kJ/kg] s_1 = refpropm('S','T',T_1,'Q',x_1,'R134a')/1000; %entropy %[kJ/kg K] % state 2: superheated vapour p_2 = p_1*r_p; %[kPa] s_2_i = s_1; % ideal entropy at state 2 (isentropic process) %[kJ/kg K] h_2_i = refpropm('H','P',p_2,'S',s_2_i*1000,'R134a')/1000; %[kJ/kg K] h_2 = h_1 + (h_2_i-h_1)/eta_comp; % Actual enthalpy at state 2 %[kJ/kg K] s_2 = refpropm('S','P',p_2,'H',h_2*1000,'R134a')/1000; %[kJ/kg K] T_2 = refpropm('T','P',p_2,'H',h_2*1000,'R134a'); %[K] % state 3: saturated liquid p_3 = (1-dp_cond)*p_2; %[kPa] x_3 = 0; % quality of the refirgerant (state: saturated liquid) %[kg/kg] T_3 = refpropm('T','P',p_3,'Q',x_3,'R134a') %[K] h_3 = refpropm('H','P',p_3,'Q',x_3,'R134a')/1000; %[kJ/kg] s_3 = refpropm('S','P',p_3,'Q',x_3,'R134a')/1000; %[kJ/kg K]

192

% state 4: 2 phase (liquid and vapour): h_4 = h_3; %[kJ/kg] p_4 = p_1*(1+dp_evap); %[kPa] x_4 = refpropm('Q','H',h_4*1000,'P',p_4,'R134a'); %[kJ/kg K] s_4 = refpropm('S','P',p_4,'Q',x_4,'R134a')/1000; %[kJ/kg K] T_4 = refpropm('T','P',p_4,'Q',x_4,'R134a') %[K] FR_R=Q_coll/(h_1-h_4); % Refrigerant flow rate %[kg/s] % Compressor work: W_comp = FR_R*(h_2-h_1) %[kW] % Rejected heat at condenser: Q_cond = FR_R*(h_2-h_3) %[kW] T_tank_top = Q_cond/(FR_w*cp_w)+T_tank_bott %[C] % System COP: COP=Q_cond/W_comp % Thermal concentration: TC = Q_cond/(G*A_a/1000); % Condenser area: LMTD = ((T_2-(T_tank_top+273.15))-(T_3-(T_tank_bott+273.15)))/(log (((T_2-(T_tank_top+273.15)))/((T_3-(T_tank_bott+273.15))))) % [k] UA_cond = Q_cond/LMTD % [kW/K] A_cond = UA_cond/(U_cond/1000); % [m2] TC_specific = (Q_cond/A_cond)/(G/1000) % Carnot COP: T_H = (T_2+T_3)/2 % The refrigerant mean temperature at condenser %[K] T_L = T_1 % The refrigerant mean temperature at evaporator %[K] COP_carnot = T_L/(T_H - T_L) % efficiency: Effi = eta_coll %total efficiency Eff_total=Q_cond/((G*50/1000)+W_comp % geometrical concentration ratio: Cr=A_a/A_cond

193

Appendix C--- Prototype assembly drawings

194

Appendix D--- Prototype parts and price list

Total Price Part name Size Material Supplier QTY Price /unit (AUS $) Frame ($409)

6040mm length

30x30 4x1950mm Aluminium Extruded Australis 4x605mm $89.63 2 $179.26 Extrusion Aluminium Engineering (parts 122) 4x100mm 4x40mm 4x300mm Australis Cuts/Taps n/a $3.00 16 $48.00 Engineering Stainless Australis Quarry Nut M5 6mm slot $0.25 125 $31.25 Steel Engineering M5 Hex Cap Stainless Australis Hex Screw $0.15 125 $18.75 Bolt Steel Engineering Stainless Australis Square Washer $0.40 25 $10.00 Steel Engineering Clamping Cast Australis 29x29mm $2.80 12 $33.60 angle(part 124) Aluminium Engineering 304 Stainless Coventry Hex Screw M6x12mm $0.16 100 $16.43 Steel Fasteners Frame Brace 5mm n/a n/a 605x30mm Finch 2 Bars Aluminium Frame Brace n/a n/a 37.5x30x25mm Aluminium Finch 4 Brackets Lens support n/a n/a 60x30x5 Aluminium Finch 4 brackets Lens Support 6x12x1950 Steel Edcon Steel $30.00 2 $60.00 M5 bolts / / / / 100 $12

Frame Cladding ($524) End Panel 665x300x4.5m Clear Acrylic Australian $40.00 2 $80.0 (part 126) m Plastic Side Panel 2019x300x4.5 Fabricators Clear Acrylic $75.00 2 $150.0 (part 126) mm PTY LTD

195

2010x77x4.5m $37.5 2 $75.0 Top Panel m Clear Acrylic (part 128) 510 x 300 x 4.5 $10.0 2 $20.0 mm Bottom Panel 2019x674x4.5 $160.0 (part 120) Clear Acrylic 1 $160.0 mm 0

Lens supporter 2000x7x4.5mm $2 6 $12.0 (part 130) Rubber Seal 1.5x50x1000m Natural Clark Rubber $3.25 3 $9.75 (part 132) m Rubber 304 Stainless Coventry Hex Screw M5x16 $0.14 100 $13.75 Steel Fasteners 304 Stainless Coventry Hex Nut M5 $0.02 200 $3.00 Steel Fasteners Tracking System ($2,193) Timing Belt Adafruit 1m Neoprene $16.95 3 $50.85 (part 108) Industries 5mm bore, 16 Adafruit Timing Pulley teeth, 6mm belt Aluminium $38.49 2 $76.98 Industries width 8mm, 2x605, Linear Rods Hardened Makershop 2x625, $17.73 4 $70.92 (part 112) Steel NZ 2x20mm Linear Rod cutting 762mm Makershop $1.33 6 $7.98 Cutting rod to size NZ Stepper Motor Adafruit NEMA 17 Steel $10.33 2 $20.66 Mount Industries Frame n/a n/a Mounting 5mm Aluminium Finch 4

Plate Carriage Aluminium PanAsia Assembly 1 100x70x20mm $400 2 $800.0 CNC Aluminium (part 114) Carriage 420*81.5*9 Aluminium PanAsia $260.0 4 $1040.0 Assembly 2 mm CNC Aluminium 316 Stainless Coventry Hex Screw M4x25mm $0.21 100 $21.29 Steel Fasteners Countersunk 304 Stainless Coventry M4x25mm $0.07 100 $7.72 Hex Screw Steel Fasteners 304 Stainless Coventry Hex Screw M3x10mm $0.05 100 $4.93 Steel Fasteners

196

304 Stainless Coventry Hex Nut M4 $0.02 200 $4.70 Steel Fasteners 304 Stainless Coventry Hex Nut M3 $0.01 200 $2.60 Steel Fasteners Stepper Motor Adafruit NEMA 16 n/a $16.95 2 $33.90 (part 110) Industries

Adafruit Motor Adafruit MotorShield n/a $19.95 1 $19.95 Controller Industries V2 Adafruit Arduino Uno n/a $24.95 1 $24.95 Industries Optics System ($1,856) Fresnel Lenses 150x500mm NTKJ $77.32 9 $695.88 (part 102) Aluminium PanAsia CPC (part 104) 500x43x35 mm $117 6 $702.00 CNC Aluminium CPC Aluminium PanAsia $8.5 18 $152 connectors CNC Aluminium 3M Reflective Adhesive Tain $307.8 4.57m 1 $307.80 Film film Industrial 0 Fluid Transfer System $6,843 Reducing 3/8"mm Stainless T-Fitting connector, 1/4in Swagelok $56.40 2 $112.80 Steel (Male, Run) NPT thread 1/4" connector, Male Elbow Stainless 1/4" NPT Swagelok $31.90 4 $127.60 Connector Steel thread 609mm live PTFE with Flexible Tube length, 1/4" Stainless Swagelok $84.20 2 $168.40 connectors fittings 1/4" to 1/4" Stainless Tube Adapter Swagelok $14.20 2 $28.40 NPT Steel Echain cable E2 094 Series PVC Igus $31.79 2 $63.58 Carrier Hard Drawn Copper Tube 3/8" Tradelink $50.00 1 $50.00 Copper Black chrome GK coated copper 1500*10 mm Copper $75 6 $452.0 International tube (part 106) ITO coated Geomatec 1500*10 mm Glass tube $1,099 $5,499 pipe (part 106) Co., Ltd

197

3mm sheath, Stainless $172.3 RTD Sensors TC Direct 2 $344.60 50mm long Steel 0 Vacuum glass chamber $3,550 Huamei Quartz glass Glass tube 1560*70 Quartz $260 3 $780 tube Electrical Australian Glass tube 1500*70 $150 3 $450 Lab supplier NanYang Vacuum XinYu / Steel $559 3 $1,677 flanges furnaces Co.,Ltd Type K Mini John Morris Electrical Plug–KF / Scientific Pty $643 1 $643 feedthrough Flange Ltd Total $15,375

198

Appendix E--- Collector mass production cost estimation

Mass production cost (10,000X) Part name (USD/Collector with 1.8 (USD/Piece) m2 Aperture area) Fresnel Lens 10.5 189 CPC (Al extrusion, Length=1.5 m) 7 42 Optical System 3M Reflective film 2.1 12.6 Subtotal: 243.6 Metal tube (Copper tube, O.D.=10 mm, Length=1.8 m) 1.82 10.92 Coated Black chrome coating (O.D.=10 mm, 12.67 76.02 Copper Tube Length=1.8 m) Sealed in Evacuated Glass Tube Receiver 54.67 328.02 Therminol 55 21 63 Subtotal: 477.96 Carriage 3.5 42 Linear rod 2.1 8.4 Linear bearings 1.4 33.6 Stepper motor 8.4 16.8 Tracking Stepper Motor Mount 1.4 2.8 System Timing belt 2.1 4.2 Timing Pulley 0.7 2.8 Arduino kit 16.2 16.2 Switches 0.7 1.4 Subtotal: 91.8 30x30 Aluminium Extrusion (6,040mm length) 3.5 56 Clamping angle 0.35 4.2 Hex Screw 0.07 2.1 Frame and Square Washer 0.07 2.1 cover Hex Screw 0.07 2.1 Quarry Nut 0.07 2.1 ACRYLIC Cover 14 42 Subtotal: 110.6 Reducing T-Fitting 3.5 14 Fluid Male Elbow Connector 2.1 4.2 transfer Flexible Tube 10.5 21 system Tube Adapter 3.5 7

Subtotal: 46.2 Adding Prism array 54 162 prism arrays-future Roller chain and additional 54 54 improved components collector Parts Total (collector with coated tube receiver): 970.16

199

Assembly ,ship and test: 84.58 Total (Current collector with vacuum insulation): 1061.97 ($589/m2) Total (Current collector without vacuum insulation): 732.98 ($407/m2) Total (Future improved collector with vacuum insulation): 1277.97 (709 $/m2)

200

References

[1] Q. Li, A. Shirazi, C. Zheng, G. Rosengarten, J.A. Scott, R.A. Taylor, Energy concentration limits in solar thermal heating applications, Energy, 96 (2016) 253-267. [2] Renewables 2015 global status report, Renewable Energy Policy Network For The 21th Century, 2015, http://www.ren21.net/wp-content/uploads/2015/07/REN12-GSR2015_Onlinebook_low_nolinks.pd f. [3] Q. Li, C. Zheng, S. Mesgari, Y.L. Hewkuruppu, N. Hjerrild, F. Crisostomo, G. Rosengarten, J.A. Scott, R.A. Taylor, Experimental and numerical investigation of volumetric versus surface solar absorbers for a concentrated solar thermal collector, Solar Energy, 136 (2016) 349-364. [4] J.-O. Dalenbäck, European Solar Thermal Technology Platform, in: NorthSun 2007 Conference Proceedings, 2007. [5] A. Shirazi, S. Pintaldi, S.D. White, G.L. Morrison, G. Rosengarten, R.A. Taylor, Solar-assisted absorption air-conditioning systems in buildings: Control strategies and operational modes, Applied Thermal Engineering, 92 (2016) 246-260. [6] A. Shirazi, R.A. Taylor, S.D. White, G.L. Morrison, Transient simulation and parametric study of solar-assisted heating and cooling absorption systems: An energetic, economic and environmental (3E) assessment, Renewable Energy, 86 (2016) 955-971. [7] R. Taylor, Q. Li, G. Rosengarten, Optical Design & Engineering Bright ideas for the future of industrial heating. [8] T. Otanicar, R.A. Taylor, P.E. Phelan, Prospects for solar cooling – An economic and environmental assessment, Solar Energy, 86 (2012) 1287-1299. [9] R.J. Fuller, Solar industrial process heating in Australia–Past and current status, Renewable energy, 36 (2011) 216-221. [10] J. Stern, International gas pricing in Europe and Asia: A crisis of fundamentals, Energy Policy, 64 (2014) 43-48. [11] R. Pitchumani, R. Gabbrielli, P. Castrataro, F. Del Medico, M. Di Palo, B. Lenzo, Proceedings of the SolarPACES 2013 International ConferenceLevelized Cost of Heat for Linear Fresnel Concentrated Solar Systems, Energy Procedia, 49 (2014) 1340-1349. [12] I. Soriga, C. Neaga, Thermal analysis of a linear Fresnel lens solar collector with black body cavity receiver, UPB Scientific Bulletin, Series D: Mechanical Engineering, 74 (2012) 105-116. [13] R.A. Taylor, P.E. Phelan, T. Otanicar, R.S. Prasher, B.E. Phelan, Socioeconomic impacts of heat transfer research, International Communications in Heat and Mass Transfer, 39 (2012) 1467-1473. [14] S. Kalogirou, The potential of solar industrial process heat applications, Applied Energy, 76 (2003) 337-361. [15] U. Al-mulali, S.A. Solarin, L. Sheau-Ting, I. Ozturk, Does moving towards renewable energy causes water and land inefficiency? An empirical investigation, Energy Policy, 93 (2016) 303-314.

201

[16] S. Kalogirou, The potential of solar industrial process heat applications, Applied Energy, 76 (2002) 337-361. [17] S. Mekhilef, R. Saidur, A. Safari, A review on solar energy use in industries, Renewable and Sustainable Energy Reviews, 15 (2011) 1777-1790. [18] A.K. Sharma, C. Sharma, S.C. Mullick, T.C. Kandpal, Carbon mitigation potential of solar industrial process heating: paper industry in India, Journal of Cleaner Production, 112 (2016) 1683-1691. [19] C. Lauterbach, B. Schmitt, U. Jordan, K. Vajen, The potential of solar heat for industrial processes in Germany, Renewable and Sustainable Energy Reviews, 16 (2012) 5121-5130. [20] D. Pietruschka, R. Fedrizzi, F. Orioli, R. Söll, R. Stauss, Demonstration of Three Large Scale Solar Process Heat Applications with Different Solar Thermal Collector Technologies, Energy Procedia, 30 (2012) 755-764. [21] W. Weiss, SOLAR HEAT FOR INDUSTRIAL APPLICATIONS, (2006). [22] W. Weiss, M. Rommel, Solar heat for industrial processes: medium temperature collectors, AEE e Institute for Sustainable Technologies, Gleisdorf, Austria. Published online at: energytech. at/pdf/medium_temperature_ collectors_task33. pdf, (2005). [23] A. Häberle, P. Frey, S. Fischer, H. Drück, K. Jakob, International Conference on Solar Heating and Cooling for Buildings and Industry, SHC 2014Monitoring Results of a Solar Process Heat System Installed at a Textile Company in Southern Germany, Energy Procedia, 70 (2015) 615-620. [24] Z. Sayadi, N. Ben Thameur, M. Bourouis, A. Bellagi, Performance optimization of solar driven small-cooled absorption–diffusion chiller working with light hydrocarbons, Energy Conversion and Management, 74 (2013) 299-307. [25] N. Kalkan, E.A. Young, A. Celiktas, Solar thermal air conditioning technology reducing the footprint of solar thermal air conditioning, Renewable and Sustainable Energy Reviews, 16 (2012) 6352-6383. [26] X.Q. Zhai, M. Qu, Y. Li, R.Z. Wang, A review for research and new design options of solar absorption cooling systems, Renewable and Sustainable Energy Reviews, 15 (2011) 4416-4423. [27] D.S. Kim, C.A. Infante Ferreira, Solar refrigeration options – a state-of-the-art review, International Journal of Refrigeration, 31 (2008) 3-15. [28] I. Sarbu, C. Sebarchievici, General review of solar-powered closed sorption refrigeration systems, Energy Conversion and Management, 105 (2015) 403-422. [29] F. Calise, M. Dentice d'Accadia, R.D. Figaj, L. Vanoli, A novel solar-assisted heat pump driven by photovoltaic/thermal collectors: Dynamic simulation and thermoeconomic optimization, Energy, 95 (2016) 346-366. [30] R. Gomri, Investigation of the potential of application of single effect and multiple effect absorption cooling systems, Energy Conversion and Management, 51 (2010) 1629-1636. [31] F. Calise, A. Palombo, L. Vanoli, Design and dynamic simulation of a novel polygeneration system fed by vegetable oil and by solar energy, Energy Conversion and Management, 60 (2012) 204-213.

202

[32] M.J. Tierney, Options for solar-assisted refrigeration—Trough collectors and double-effect chillers, Renewable Energy, 32 (2007) 183-199. [33] D. Chemisana, J. López-Villada, A. Coronas, J.I. Rosell, C. Lodi, Building integration of concentrating systems for solar cooling applications, Applied Thermal Engineering, 50 (2013) 1472-1479. [34] O. Ayadi, M. Aprile, M. Motta, Solar Cooling Systems Utilizing Concentrating Solar Collectors - An Overview, Energy Procedia, 30 (2012) 875-883. [35] Y.L. Liu, R.Z. Wang, Performance prediction of a solar/gas driving double effect LiBr–H2O absorption system, Renewable Energy, 29 (2004) 1677-1695. [36] A. Shirazi, R.A. Taylor, S.D. White, G.L. Morrison, A systematic parametric study and feasibility assessment of solar-assisted single-effect, double-effect, and triple-effect absorption chillers for heating and cooling applications, Energy Conversion and Management, 114 (2016) 258-277. [37] A. Al-Alili, Y. Hwang, R. Radermacher, I. Kubo, Optimization of a solar powered absorption cycle under Abu Dhabi’s weather conditions, Solar Energy, 84 (2010) 2034-2040. [38] M.S. Buker, S.B. Riffat, Building integrated solar thermal collectors–A review, Renewable and Sustainable Energy Reviews, 51 (2015) 327-346. [39] Rabl, A. "Optical and thermal properties of Compound Parobolic Concentraters," Solar Energy, Vol. 18, pp. 497-511. 1976. [40] W. Platzer, Status quo and new developments in medium temperature collectors, in: 2nd annual conference of the European technology platform on renewable heating and cooling. Budapest, Hungary, 2011. [41] X. Gu, R.A. Taylor, Q. Li, J.A. Scott, G. Rosengarten, Thermal analysis of a micro solar thermal collector designed for methanol reforming, Solar Energy, 113 (2015) 189-198. [42] A. Fernandez-Garcia, E. Zarza, L. Valenzuela, M. Pérez, Parabolic-trough solar collectors and their applications, Renewable and Sustainable Energy Reviews, 14 (2010) 1695-1721. [43] H. Apostoleris, M. Stefancich, M. Chiesa, Tracking-integrated systems for concentrating photovoltaics, Nature Energy, 1 (2016) 16018. [44] K. Alam, B.B. Saha, A. Akisawa, Adsorption cooling driven by solar collector: a case study for Tokyo solar data, Applied Thermal Engineering, 50 (2013) 1603-1609. [45] R. Winston, L. Jiang, B. Widyolar, Performance of a 23KW Solar Thermal Cooling System Employing a Double Effect Absorption Chiller and Thermodynamically Efficient Non-tracking Concentrators, Energy Procedia, 48 (2014) 1036-1046. [46] F. Calise, M.D. d’Accadia, M. Vicidomini, M. Scarpellino, Design and simulation of a prototype of a small-scale solar CHP system based on evacuated flat-plate solar collectors and Organic Rankine Cycle, Energy Conversion and Management, 90 (2015) 347-363. [47] S. Cohen, G. Grossman, Development of a solar collector with a stationary spherical reflector/tracking absorber for industrial process heat, Solar Energy, 128 (2016) 31-40. [48] W. Weiss, M. Rommel, Solar heat for industrial processes, IEA SC–task, 33 (2003).

203

[49] K.H. Solangi, M.R. Islam, R. Saidur, N.A. Rahim, H. Fayaz, A review on global solar energy policy, Renewable and Sustainable Energy Reviews, 15 (2011) 2149-2163. [50] S.V. Boriskina, M.A. Green, K. Catchpole, E. Yablonovitch, M.C. Beard, Y. Okada, S. Lany, T. Gershon, A. Zakutayev, M.H. Tahersima, Roadmap on optical energy conversion, Journal of Optics, 18 (2016) 073004. [51] J.-S. Lin, C.-W. Liang, Thin solar concentrator with high concentration ratio, in, 2013, pp. 88210H-88210H-88215. [52] S.A. Kalogirou, Low Concentration Ratio Solar Collectors, (2012) 149-163. [53] H. Zheng, C. Feng, Y. Su, J. Dai, X. Ma, Design and experimental analysis of a cylindrical compound Fresnel solar concentrator, Solar Energy, 107 (2014) 26-37. [54] W.T. Xie, Y.J. Dai, R.Z. Wang, K. Sumathy, Concentrated solar energy applications using Fresnel lenses: A review, Renewable and Sustainable Energy Reviews, 15 (2011) 2588-2606. [55] D. Feuermann, J.M. Gordon, Solar fiber-optic mini-dishes: a new approach to the efficient collection of sunlight, Solar Energy, 65 (1999) 159-170. [56] K.E.J. Al-Jumaily, M.K.A. Al-Kaysi, The study of the performance and efficiency of flat linear Fresnel lens collector with sun tracking system in Iraq, Renewable Energy, 14 (1998) 41-48. [57] F. Franc, V. Jirka, M. Malý, B. Nábělek, Concentrating collectors with flat linear fresnel lenses, Solar & Wind Technology, 3 (1986) 77-84. [58] H. Zhai, Y.J. Dai, J.Y. Wu, R.Z. Wang, L.Y. Zhang, Experimental investigation and analysis on a concentrating solar collector using linear Fresnel lens, Energy Conversion and Management, 51 (2010) 48-55. [59] W.T. Xie, Y.J. Dai, R.Z. Wang, Theoretical and experimental analysis on efficiency factors and heat removal factors of Fresnel lens solar collector using different cavity receivers, Solar Energy, 86 (2012) 2458-2471. [60] W.T. Xie, Y.J. Dai, R.Z. Wang, Numerical and experimental analysis of a point focus solar collector using high concentration imaging PMMA Fresnel lens, Energy Conversion and Management, 52 (2011) 2417-2426. [61] D. Chemisana, M. Ibáñez, J.I. Rosell, Characterization of a photovoltaic-thermal module for Fresnel linear concentrator, Energy Conversion and Management, 52 (2011) 3234-3240. [62] D. Erickson, D. Sinton, D. Psaltis, Optofluidics for energy applications, Nat Photon, 5 (2011) 583-590. [63] C. Liu, Liquid prism for beam tracking and steering, Optical Engineering, 51 (2012) 114002. [64] N. Yeh, P. Yeh, Analysis of point-focused, non-imaging Fresnel lenses' concentration profile and manufacture parameters, Renewable Energy, 85 (2016) 514-523. [65] Y.-S. Tsou, K.-H. Chang, Y.-H. Lin, A droplet manipulation on a liquid crystal and polymer composite film as a concentrator and a sun tracker for a concentrating photovoltaic system, Journal of Applied Physics, 113 (2013) 244504. [66] P. Kotsidas, V. Modi, J.M. Gordon, Gradient-index lenses for near-ideal imaging and concentration with realistic materials, Opt. Express, 19 (2011) 15584-15595.

204

[67] H. Chen, C. Chan, P. Sheng, Transformation optics and metamaterials, Nature materials, 9 (2010) 387-396. [68] A. Boltasseva, V.M. Shalaev, Fabrication of optical negative-index metamaterials: Recent advances and outlook, Metamaterials, 2 (2008) 1-17. [69] N. Kundtz, D.R. Smith, Extreme-angle broadband metamaterial lens, Nat Mater, 9 (2010) 129-132. [70] N.E. Gharbi, H. Derbal, S. Bouaichaoui, N. Said, A comparative study between parabolic trough collector and linear Fresnel reflector technologies, Energy Procedia, 6 (2011) 565-572. [71] T. Sultana, G.L. Morrison, G. Rosengarten, Thermal performance of a novel rooftop solar micro-concentrating collector, Solar Energy, 86 (2012) 1992-2000. [72] M. Vivar, V. Everett, M. Fuentes, E. Thomsen, J. Harvey, M. Ebert, P. le Lievre, M. Greaves, A. Tanner, A. Blakers, Results from the first ANU-chromasun CPV-T microconcentrator prototype in Canberra, (2012) 114-117. [73] M. Lin, K. Sumathy, Y. Dai, R. Wang, Y. Chen, Experimental and theoretical analysis on a linear Fresnel reflector solar collector prototype with V-shaped cavity receiver, Applied Thermal Engineering, 51 (2013) 963-972. [74] D. Liang, L. Fraser Monteiro, M. Ribau Teixeira, M. Fraser Monteiro, M. Collares-Pereira, Fiber-optic solar energy transmission and concentration, Solar Energy Materials and Solar Cells, 54 (1998) 323-331. [75] W. Weiss, M. Rommel, Process heat collectors, State of the Art within Task, 33 (2008). [76] A. Fernández-García, E. Zarza, L. Valenzuela, M. Pérez, Parabolic-trough solar collectors and their applications, Renewable and Sustainable Energy Reviews, 14 (2010) 1695-1721. [77] A. Lokurlu, F. Richarts, D. Kruger, High efficient utilisation of solar energy with newly developed parabolic trough collectors (SOLITEM PTC) for chilling and steam production in a hotel at the Mediterranean coast of Turkey, International journal of energy technology and policy, 3 (2005) 137-146. [78] J. Blanco, D. Alarcón, B. Sánchez, S. Malato, M.I. Maldonado, A. Hublitz, M. Spinnler, Technical comparison of different solar-assisted heat supply systems for a multi-effect seawater distillation unit, in: ISES Solar World Congress, 2003, pp. 14-19. [79] F. Buttinger, T. Beikircher, M. Pröll, W. Schölkopf, Development of a new flat stationary evacuated CPC-collector for process heat applications, Solar Energy, 84 (2010) 1166-1174. [80] G. Zhu, T. Wendelin, M.J. Wagner, C. Kutscher, History, current state, and future of linear Fresnel concentrating solar collectors, Solar Energy, (2013). [81] X. Gu, R.A. Taylor, G. Morrison, G. Rosengarten, Theoretical analysis of a novel, portable, CPC-based solar thermal collector for methanol reforming, Applied Energy, 119 (2014) 467-475. [82] A.S. Gudekar, A.S. Jadhav, S.V. Panse, J.B. Joshi, A.B. Pandit, Cost effective design of compound parabolic collector for steam generation, Solar Energy, 90 (2013) 43-50. [83] J. Facão, A.C. Oliveira, Numerical simulation of a trapezoidal cavity receiver for a linear Fresnel solar collector concentrator, Renewable Energy, 36 (2011) 90-96.

205

[84] G. Morin, W. Platzer, M. Eck, R. Uhlig, A. Häberle, M. Berger, E. Zarza, Road map towards the demonstration of a linear Fresnel collector using a single tube receiver, in: 13th International Symposium on Concentrated Solar Power and Chemical Energy Technologies, Seville (Spain), http://elib. dlr. de/44052/(15.5. 2012), 2006. [85] R. Bernhard, H. Laabs, J. de Lalaing, M. Eck, M. Eickhoff, K. Pottler, G. Morin, A. Heimsath, A. Georg, A. Häberle, Linear Fresnel collector demonstration on the PSA, Part I–design, construction and quality control, Proc. 14th Biennial SolarPACES Concentrating Solar, (2008). [86] P. Singh, S. Ganesan, G. Yadav, Technical note: performance study of a linear Fresnel concentrating solar device, Renewable energy, 18 (1999) 409-416. [87] R. Abbas, J.M. Martínez-Val, Analytic optical design of linear Fresnel collectors with variable widths and shifts of mirrors, Renewable Energy, 75 (2015) 81-92. [88] D. Chemisana, J.I. Rosell, Design and optical performance of a nonimaging Fresnel transmissive concentrator for building integration applications, Energy Conversion and Management, 52 (2011) 3241-3248. [89] J.S. Coventry, Performance of a concentrating photovoltaic/thermal solar collector, Solar Energy, 78 (2005) 211-222. [90] R. Winston, Light collection within the framework of geometrical optics, JOSA, 60 (1970) 245-247. [91] N. Sellami, T.K. Mallick, Optical efficiency study of PV Crossed Compound Parabolic Concentrator, Applied Energy, 102 (2013) 868-876. [92] K.A. Snail, J.J. O'Gallagher, R. Winston, A stationary evacuated collector with integrated concentrator, Solar Energy, 33 (1984) 441-449. [93] Z. Lu, R. Wang, Z. Xia, X. Lu, C. Yang, Y. Ma, G. Ma, Study of a novel solar adsorption cooling system and a solar absorption cooling system with new CPC collectors, Renewable Energy, 50 (2013) 299-306. [94] C. Zauner, F. Hengstberger, W. Hohenauer, C. Reichl, A. Simetzberger, G. Gleiss, Methods for Medium Temperature Collector Development Applied to a CPC Collector, Energy Procedia, 30 (2012) 187-197. [95] J.H. Karp, E.J. Tremblay, J.M. Hallas, J.E. Ford, Orthogonal and secondary concentration in planar micro-optic solar collectors, Opt. Express, 19 (2011) A673-A685. [96] L.L. Chen, Stationary photovoltaic array module design for solar electric power generation systems, in, Google Patents, 2003. [97] A. Terao, W.P. Mulligan, S.G. Daroczi, O.C. Pujol, P.J. Verlinden, R.M. Swanson, J. Miñano, P. Benitz, J.L. Alvarez, A mirror-less design for micro-concentrator modules, in: Photovoltaic Specialists Conference, 2000. Conference Record of the Twenty-Eighth IEEE, IEEE, 2000, pp. 1416-1419. [98] A.P. Plesniak, J.-S. Lin, C.-W. Liang, Thin solar concentrator with high concentration ratio, 8821 (2013) 88210H.

206

[99] G.E. Arnaoutakis, J. Marques-Hueso, T.K. Mallick, B.S. Richards, Coupling of sunlight into optical fibres and spectral dependence for solar energy applications, Solar energy, 93 (2013) 235-243. [100] N.-H. Vu, S. Shin, Cost-effective optical fiber daylighting system using modified compound parabolic concentrators, Solar Energy, 136 (2016) 145-152. [101] P. Genequand, V. Stark, Solar energy system with composite concentrating lenses, in, Google Patents, 1980. [102] L.L. Northrup Jr, Compound lens solar energy system, in, Google Patents, 1977. [103] B.S. Soleau, Thin-line collectors, in, Google Patents, 1981. [104] J.P. Morgan, Light-guide solar panel and method of fabrication thereof, in, Google Patents, 2011. [105] R.M. Murtha, Side-collecting lightguide, in, Google Patents, 2000. [106] R.M. Murtha, Solar concentrating liquid lightguide, in, Google Patents, 2003. [107] R. Kapadia, P. Bandaru, Heat flux concentration through polymeric thermal lenses, Applied Physics Letters, 105 (2014) 233903. [108] D. Kraemer, B. Poudel, H.-P. Feng, J.C. Caylor, B. Yu, X. Yan, Y. Ma, X. Wang, D. Wang, A. Muto, K. McEnaney, M. Chiesa, Z. Ren, G. Chen, High-performance flat-panel solar thermoelectric generators with high thermal concentration, Nat Mater, 10 (2011) 532-538. [109] L.L. Baranowski, G.J. Snyder, E.S. Toberer, Concentrated solar thermoelectric generators, Energy & Environmental Science, 5 (2012) 9055. [110] G. Chen, Theoretical efficiency of solar thermoelectric energy generators, Journal of Applied Physics, 109 (2011) 104908. [111] E. Azad, Theoretical and experimental investigation of heat pipe solar collector, Experimental Thermal and Fluid Science, 32 (2008) 1666-1672. [112] N. Selvakumar, H.C. Barshilia, Review of physical vapor deposited (PVD) spectrally selective coatings for mid-and high-temperature solar thermal applications, Solar Energy Materials and Solar Cells, 98 (2012) 1-23. [113] C. Atkinson, C.L. Sansom, H.J. Almond, C.P. Shaw, Coatings for concentrating solar systems – A review, Renewable and Sustainable Energy Reviews, 45 (2015) 113-122. [114] Y.L. Hewakuruppu, R.A. Taylor, H. Tyagi, V. Khullar, T. Otanicar, S. Coulombe, N. Hordy, Limits of selectivity of direct volumetric solar absorption, Solar Energy, 114 (2015) 206-216. [115] R.A. Taylor, P.E. Phelan, T.P. Otanicar, C.A. Walker, M. Nguyen, S. Trimble, R. Prasher, Applicability of nanofluids in high flux solar collectors, Journal of Renewable and Sustainable Energy, 3 (2011) 023104. [116] R.A. Taylor, P.E. Phelan, R.J. Adrian, A. Gunawan, T.P. Otanicar, Characterization of light-induced, volumetric steam generation in nanofluids, International Journal of Thermal Sciences, 56 (2012) 1-11. [117] R.A. Taylor, T. Otanicar, G. Rosengarten, Nanofluid-based optical filter optimization for PV/T systems, Light: Science & Applications, 1 (2012) e34. 207

[118] R. Taylor, S. Coulombe, T. Otanicar, P. Phelan, A. Gunawan, W. Lv, G. Rosengarten, R. Prasher, H. Tyagi, Small particles, big impacts: a review of the diverse applications of nanofluids, Journal of Applied Physics, 113 (2013) 011301. [119] R.A. Taylor, T.P. Otanicar, Y. Herukerrupu, F. Bremond, G. Rosengarten, E.R. Hawkes, X. Jiang, S. Coulombe, Feasibility of nanofluid-based optical filters, Applied optics, 52 (2013) 1413-1422. [120] R.A. Taylor, P.E. Phelan, T.P. Otanicar, R. Adrian, R. Prasher, Nanofluid optical property characterization: towards efficient direct absorption solar collectors, Nanoscale research letters, 6 (2011) 1-11. [121] S. Mesgari, S. Coulombe, N. Hordy, R. Taylor, Thermal stability of carbon nanotube-based nanofluids for solar thermal collectors, Materials Research Innovations, 19 (2015) S5-650-S655-653. [122] Y. Xuan, H. Duan, Q. Li, Enhancement of solar energy absorption using a plasmonic nanofluid based on TiO 2/Ag composite nanoparticles, RSC Advances, 4 (2014) 16206-16213. [123] N.E. Hjerrild, S. Mesgari, F. Crisostomo, J.A. Scott, R. Amal, R.A. Taylor, Hybrid PV/T enhancement using selectively absorbing Ag–SiO 2/carbon nanofluids, Solar Energy Materials and Solar Cells, 147 (2016) 281-287. [124] V. Khullar, H. Tyagi, N. Hordy, T.P. Otanicar, Y. Hewakuruppu, P. Modi, R.A. Taylor, Harvesting solar thermal energy through nanofluid-based volumetric absorption systems, International Journal of Heat and Mass Transfer, 77 (2014) 377-384. [125] T.P. Otanicar, P.E. Phelan, R.S. Prasher, G. Rosengarten, R.A. Taylor, Nanofluid-based direct absorption solar collector, Journal of renewable and sustainable energy, 2 (2010) 033102. [126] M. Karami, M. Akhavan-Bahabadi, S. Delfani, M. Raisee, Experimental investigation of CuO nanofluid-based Direct Absorption Solar Collector for residential applications, Renewable and Sustainable Energy Reviews, 52 (2015) 793-801. [127] A. Lenert, E.N. Wang, Optimization of nanofluid volumetric receivers for solar thermal energy conversion, Solar Energy, 86 (2012) 253-265. [128] H.K. Gupta, G.D. Agrawal, J. Mathur, Investigations for effect of Al2O3–H2O nanofluid flow rate on the efficiency of direct absorption solar collector, Case Studies in Thermal Engineering, 5 (2015) 70-78. [129] H.K. Gupta, G.D. Agrawal, J. Mathur, An experimental investigation of a low temperature Al2O3-H2O nanofluid based direct absorption solar collector, Solar Energy, 118 (2015) 390-396. [130] S. Delfani, M. Karami, M. Akhavan-Behabadi, Performance characteristics of a residential-type direct absorption solar collector using MWCNT nanofluid, Renewable Energy, 87 (2016) 754-764. [131] N. Hordy, D. Rabilloud, J.-L. Meunier, S. Coulombe, High temperature and long-term stability of carbon nanotube nanofluids for direct absorption solar thermal collectors, Solar Energy, 105 (2014) 82-90. [132] S. Suman, M.K. Khan, M. Pathak, Performance enhancement of solar collectors—A review, Renewable and Sustainable Energy Reviews, 49 (2015) 192-210.

208

[133] V. Khullar, Solar Energy Harvesting Using Nanofluids-Based Concentrating Solar Collector, Journal of Nanotechnology in Engineering and Medicine, 3 (2013) 031003. [134] I. Dincer, M. Rosen, Thermal energy storage: systems and applications, John Wiley & Sons, 2002. [135] P.V. Bhale, M.K. Rathod, L. Sahoo, Thermal Analysis of a Solar Concentrating System Integrated with Sensible and Latent Heat Storage, Energy Procedia, 75 (2015) 2157-2162. [136] B. Xu, P. Li, C. Chan, Application of phase change materials for thermal energy storage in concentrated solar thermal power plants: a review to recent developments, Applied Energy, 160 (2015) 286-307. [137] R.A. Taylor, C.-Y. Chung, K. Morrison, E.R. Hawkes, Analysis and testing of a portable thermal battery, Journal of Thermal Science and Engineering Applications, 6 (2014) 031004. [138] M. Chaabane, H. Mhiri, P. Bournot, Thermal performance of an integrated collector storage solar water heater (ICSSWH) with phase change materials (PCM), Energy Conversion and Management, 78 (2014) 897-903. [139] A. Papadimitratos, S. Sobhansarbandi, V. Pozdin, A. Zakhidov, F. Hassanipour, Evacuated tube solar collectors integrated with phase change materials, Solar Energy, 129 (2016) 10-19. [140] A. Lecuona, J.-I. Nogueira, R. Ventas, M. Legrand, Solar cooker of the portable parabolic type incorporating heat storage based on PCM, Applied Energy, 111 (2013) 1136-1146. [141] M. Mussard, A. Gueno, O.J. Nydal, Experimental study of solar cooking using heat storage in comparison with direct heating, Solar Energy, 98, Part C (2013) 375-383. [142] S. Bouadila, S. Kooli, M. Lazaar, S. Skouri, A. Farhat, Performance of a new solar air heater with packed-bed latent storage energy for nocturnal use, Applied Energy, 110 (2013) 267-275. [143] M.A. Sharif, A. Al-Abidi, S. Mat, K. Sopian, M.H. Ruslan, M.Y. Sulaiman, M. Rosli, Review of the application of phase change material for heating and domestic hot water systems, Renewable and Sustainable Energy Reviews, 42 (2015) 557-568. [144] Z. Chen, M. Gu, D. Peng, Heat transfer performance analysis of a solar flat-plate collector with an integrated metal foam porous structure filled with paraffin, Applied Thermal Engineering, 30 (2010) 1967-1973. [145] Y. Varol, A. Koca, H.F. Oztop, E. Avci, Forecasting of thermal energy storage performance of Phase Change Material in a solar collector using soft computing techniques, Expert Systems with Applications, 37 (2010) 2724-2732. [146] M. Naghavi, K. Ong, I. Badruddin, M. Mehrali, M. Silakhori, H. Metselaar, Theoretical model of an evacuated tube heat pipe solar collector integrated with phase change material, Energy, 91 (2015) 911-924. [147] A.H. Slocum, D.S. Codd, J. Buongiorno, C. Forsberg, T. McKrell, J.-C. Nave, C.N. Papanicolas, A. Ghobeity, C.J. Noone, S. Passerini, Concentrated solar power on demand, Solar Energy, 85 (2011) 1519-1529. [148] A. Gil, D.S. Codd, L. Zhou, D. Trumper, R.B. Campbell, B. Grange, N. Calvet, P. Armstrong, A.H. Slocum, Design of a 100 kW Concentrated Solar Power on Demand Volumetric Receiver With Integral Thermal Energy Storage Prototype, in: ASME 2015 Power Conference collocated

209

with the ASME 2015 9th International Conference on Energy Sustainability, the ASME 2015 13th International Conference on Fuel Cell Science, Engineering and Technology, and the ASME 2015 Nuclear Forum, American Society of Mechanical Engineers, 2015, pp. V001T001A011-V001T001A011. [149] H. Mousazadeh, A. Keyhani, A. Javadi, H. Mobli, K. Abrinia, A. Sharifi, A review of principle and sun-tracking methods for maximizing solar systems output, Renewable and sustainable energy reviews, 13 (2009) 1800-1818. [150] A.L. Vessey, Design and evaluation of polycarbonate microlens arrays for hybrid concentration photovoltaic cells, in, University of Delaware, 2015. [151] C. Zheng, Q. Li, G. Rosengarten, E. Hawkes, R.A. Taylor, A new optical concentrator design and analysis for rooftop solar applications, in: SPIE Optical Engineering+ Applications, International Society for Optics and Photonics, 2015, pp. 957209-957209-957211. [152] N. Yeh, Analysis of spectrum distribution and optical losses under Fresnel lenses, Renewable and Sustainable Energy Reviews, 14 (2010) 2926-2935. [153] R. Winston, Principles of solar concentrators of a novel design, Solar Energy, 16 (1974) 89-95. [154] R. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi, Design of a nonimaging Fresnel lens for solar concentrators, Solar energy, 65 (1999) 379-387. [155] V. Kumar, R.L. Shrivastava, S.P. Untawale, Fresnel lens: A promising alternative of reflectors in concentrated solar power, Renewable and Sustainable Energy Reviews, 44 (2015) 376-390. [156] C.-F.J. Kuo, C.-C. Huang, Y.-L. Kuo, Analysis of processing parameters in fabrication of Fresnel lens solar collector, Energy Conversion and Management, 57 (2012) 33-41. [157] P. Kovacs, A guide to the standard EN 12975, Technical Research Institute of, (2012). [158] T.P.V.O.O. POGOJEV, Z.P. GORIV, Flame temperature as a function of the combustion conditions of gaseous fuels, Materiali in Tehnologije, 40 (2006) 89. [159] P. Henshall, P. Eames, F. Arya, T. Hyde, R. Moss, S. Shire, Constant temperature induced stresses in evacuated enclosures for high performance flat plate solar thermal collectors, Solar Energy, 127 (2016) 250-261. [160] Y. Wang, Y. Zhu, H. Chen, X. Zhang, L. Yang, C. Liao, Performance analysis of a novel sun-tracking CPC heat pipe evacuated tubular collector, Applied Thermal Engineering, 87 (2015) 381-388. [161] J.A. Duffie, W.A. Beckman, Solar engineering of thermal processes, John Wiley & Sons, 2013. [162] K.R. Varma, M.P. Mengüç, Effects of particulate concentrations on temperature and heat flux distributions in a pulverized-coal fired furnace, International Journal of Energy Research, 13 (1989) 555-572. [163] S. Orsino, R. Weber, U. Bollettini, Numerical simulation of combustion of natural gas with high-temperature air, Combustion Science and Technology, 170 (2001) 1-34.

210

[164] S. Song, V. Au, K.P. Moran, Constriction/spreading resistance model for electronics packaging, in: Proceedings of the 4th ASME/JSME thermal engineering joint conference, 1995, pp. 199-206. [165] T. Tong, Y. Zhao, L. Delzeit, A. Kashani, M. Meyyappan, A. Majumdar, Dense vertically aligned multiwalled carbon nanotube arrays as thermal interface materials, Components and Packaging Technologies, IEEE Transactions on, 30 (2007) 92-100. [166] X. Luo, D. Chung, Effect of the thickness of a thermal interface material (solder) on heat transfer between copper surfaces, International microelectronics and packaging society, The international journal of microcircuits and electronic packaging, 24 (2001) 141-147. [167] R. Lates, I. Visa, C. Lapusan, Mathematical optimization of solar thermal collectors efficiency function using MATLAB, in: Proceedings of The 4th IASME/WSEAS International Conference on Energy, Environment, Ecosystems and Sustainable Development (EEESD'08), 2008, pp. 42-46. [168] R.W. Bliss Jr, The derivations of several “plate-efficiency factors” useful in the design of flat-plate solar heat collectors, Solar Energy, 3 (1959) 55-64. [169] L.L. Vasiliev, Heat pipes in modern heat exchangers, Applied Thermal Engineering, 25 (2005) 1-19. [170] P. Nemec, A. Čaja, M. Malcho, Mathematical model for heat transfer limitations of heat pipe, Mathematical and Computer Modelling, 57 (2013) 126-136. [171] S.B. Riffat, X. Zhao, P.S. Doherty, Developing a theoretical model to investigate thermal performance of a thin membrane heat-pipe solar collector, Applied Thermal Engineering, 25 (2005) 899-915. [172] C. Loh, E. Harris, D. Chou, Comparative study of heat pipes performances in different orientations, in: Semiconductor Thermal Measurement and Management Symposium, 2005 IEEE Twenty First Annual IEEE, IEEE, 2005, pp. 191-195. [173] S. Chi, Heat pipe theory and practice: a sourcebook, (1976). [174] K. Ismail, M. Abogderah, Performance of a heat pipe solar collector, Journal of solar energy engineering, 120 (1998) 51-59. [175] H. Yang, S. Khandekar, M. Groll, Operational limit of closed loop pulsating heat pipes, Applied Thermal Engineering, 28 (2008) 49-59. [176] S. Jack, J. Parzefall, T. Luttmann, P. Janßen, F. Giovannetti, Flat Plate Aluminum Heat Pipe Collector with Inherently Limited Stagnation Temperature, Energy Procedia, 48 (2014) 105-113. [177] A. Hepbasli, Y. Kalinci, A review of heat pump water heating systems, Renewable and Sustainable Energy Reviews, 13 (2009) 1211-1229. [178] K.J. Chua, S.K. Chou, W.M. Yang, Advances in heat pump systems: A review, Applied Energy, 87 (2010) 3611-3624. [179] Q. Li, G. Flamant, X. Yuan, P. Neveu, L. Luo, Compact heat exchangers: A review and future applications for a new generation of high temperature solar receivers, Renewable and Sustainable Energy Reviews, 15 (2011) 4855-4875.

211

[180] M. Khamis Mansour, Thermal analysis of novel minichannel-based solar flat-plate collector, Energy, 60 (2013) 333-343. [181] C. Glynn, T. O’Donovan, D.B. Murray, Jet impingement cooling, in: Proceedings of the 9th UK National Heat Transfer Conference, Manchester, England, September, 2005, pp. 5-6. [182] C. Yunus, J. Afshin, Heat and Mass Transfer: Fundamentals and Applications, in, Tata Mcgraw Hill, New Delhi. ISBN, 2011. [183] H. Mohammed, G. Bhaskaran, N. Shuaib, R. Saidur, Heat transfer and fluid flow characteristics in microchannels heat exchanger using nanofluids: a review, Renewable and Sustainable Energy Reviews, 15 (2011) 1502-1512. [184] S.G. Kandlikar, S. Colin, Y. Peles, S. Garimella, R.F. Pease, J.J. Brandner, D.B. Tuckerman, Heat transfer in microchannels—2012 status and research needs, Journal of Heat Transfer, 135 (2013) 091001. [185] A. Handbook, ASHRAE Handbook–Fundamentals, Atlanta, GA, (2009). [186] S. Sanaye, A. Shirazi, Four E analysis and multi-objective optimization of an ice thermal energy storage for air-conditioning applications, international journal of refrigeration, 36 (2013) 828-841. [187] S.K. Chaturvedi, V.D. Gagrani, T.M. Abdel-Salam, Solar-assisted heat pump – A sustainable system for low-temperature water heating applications, Energy Conversion and Management, 77 (2014) 550-557. [188] S. Ghosh, I. Dincer, Development and performance assessment of a new integrated system for HVAC&R applications, Energy, 80 (2015) 159-167. [189] J. García-Cascales, F. Vera-García, J. Corberán-Salvador, J. Gonzálvez-Maciá, Assessment of boiling and condensation heat transfer correlations in the modelling of plate heat exchangers, International Journal of Refrigeration, 30 (2007) 1029-1041. [190] S.B. Riffat, X. Ma, Improving the coefficient of performance of thermoelectric cooling systems: a review, International Journal of Energy Research, 28 (2004) 753-768. [191] B. David, J. Ramousse, L. Luo, Optimization of thermoelectric heat pumps by operating condition management and heat exchanger design, Energy Conversion and Management, 60 (2012) 125-133. [192] R.A. Taylor, G.L. Solbrekken, Comprehensive system-level optimization of thermoelectric devices for electronic cooling applications, Components and Packaging Technologies, IEEE Transactions on, 31 (2008) 23-31. [193] R.A. Taylor, Y. Hewakuruppu, D. DeJarnette, T.P. Otanicar, Comparison of selective transmitters for solar thermal applications, Applied optics, 55 (2016) 3829-3839. [194] S. Mesgari, R.A. Taylor, N.E. Hjerrild, F. Crisostomo, Q. Li, J. Scott, An investigation of thermal stability of carbon nanofluids for solar thermal applications, Solar Energy Materials and Solar Cells, 157 (2016) 652-659. [195] H. Wang, V.P. Sivan, A. Mitchell, G. Rosengarten, P. Phelan, L. Wang, Highly efficient selective metamaterial absorber for high-temperature solar thermal energy harvesting, Solar Energy Materials and Solar Cells, 137 (2015) 235-242.

212

[196] A. Schott, Borofloat® 33 product information, , in, 2011. [197] R. Gardon, The emissivity of transparent materials, Journal of the American Ceramic Society, 39 (1956) 278-285. [198] L. Dombrovsky, J. Randrianalisoa, D. Baillis, Modified two-flux approximation for identification of radiative properties of absorbing and scattering media from directional-hemispherical measurements, JOSA A, 23 (2006) 91-98. [199] Y.L. Hewakuruppu, L.A. Dombrovsky, C. Chen, V. Timchenko, X. Jiang, S. Baek, R.A. Taylor, Plasmonic “pump–probe” method to study semi-transparent nanofluids, Applied optics, 52 (2013) 6041-6050. [200] D. Lei, Z. Wang, J. Li, J. Li, Z. Wang, Experimental study of glass to metal seals for parabolic trough receivers, Renewable energy, 48 (2012) 85-91. [201] O. Shekoofa, J. Wang, J. Qi, J. Zhang, Z. Yin, Analysis of residual stress for mismatch metal–glass seals in solar evacuated tubes, Solar Energy Materials and Solar Cells, 128 (2014) 421-426. [202] F. Giovannetti, S. Föste, N. Ehrmann, G. Rockendorf, High transmittance, low emissivity glass covers for flat plate collectors: applications and performance, Energy Procedia, 30 (2012) 106-115. [203] S. Sasaki, H. Kou, H. Masuda, H. Kiyohashi, Total hemispherical emissivity of glass sheets with different thicknesses measured by a transient calorimetric, (2004). [204] D. Ren, H. Tan, Y. Xuan, Y. Han, Q. Li, Apparatus for Measuring Spectral Emissivity of Solid Materials at Elevated Temperatures, International Journal of Thermophysics, 37 (2016) 1-20. [205] K. Sun, W. Zhou, X. Tang, Z. Huang, F. Lou, D. Zhu, Effect of the heat treatment on the infrared emissivity of indium tin oxide (ITO) films, Applied Surface Science, 257 (2011) 9639-9642. [206] R. Viskanta, M. Mengüç, Radiation heat transfer in combustion systems, Progress in Energy and Combustion Science, 13 (1987) 97-160. [207] N. Pradhan, H. Duan, J. Liang, G. Iannacchione, The specific heat and effective thermal conductivity of composites containing single-wall and multi-wall carbon nanotubes, Nanotechnology, 20 (2009) 245705. [208] T. Sultana, G.L. Morrison, R.A. Taylor, G. Rosengarten, Numerical and experimental study of a solar micro concentrating collector, Solar Energy, 112 (2015) 20-29. [209] T. Sultana, G. Morrison, R. Taylor, G. Rosengarten, Performance of a Linear Fresnel Rooftop Mounted Concentrating Solar Collector, in: 50th Australian Solar Energy Society Annual Conference (AuSE), Melbourne, Australia, Dec, 2012, pp. 6-7. [210] J.R. Howell, Application of Monte Carlo to heat transfer problems, Advances in heat transfer, 5 (1968) 1-54. [211] V. Sabatelli, D. Marano, G. Braccio, V.K. Sharma, Efficiency test of solar collectors: uncertainty in the estimation of regression parameters and sensitivity analysis, Energy Conversion and Management, 43 (2002) 2287-2295.

213

[212] R.C. International, 508-2704 Instruction Sheet, , in, May, 2005. [213] G. FREEmER, Measuring Temperature by Direct Contact, Chemical Engineering Progress, 107 (2011) 26-30. [214] E.O. Doebelin, Measurement systems: application and design, McGraw-Hill New York, 1990. [215] M. Islam, I. Kubo, M. Ohadi, A. Alili, Measurement of solar energy radiation in Abu Dhabi, UAE, Applied Energy, 86 (2009) 511-515. [216] C. Austerlitz, D. Campos, R. Allison, C. Sheng, C. Bonnerup, C. Sibata, Short and long term stability of the Diomed 630PDT laser evaluated with integrating sphere, power meter, and calorimeter, in: 12th World Congress of the International Photodynamic Association, International Society for Optics and Photonics, 2009, pp. 73804L-73804L-73808. [217] D.L. King, J.A. Kratochvil, W.E. Boyson, Measuring solar spectral and angle-of-incidence effects on photovoltaic modules and solar irradiance sensors, in: Photovoltaic Specialists Conference, 1997., Conference Record of the Twenty-Sixth IEEE, IEEE, 1997, pp. 1113-1116. [218] S. Fischer, W. Heidemann, H. Müller-Steinhagen, B. Perers, P. Bergquist, B. Hellström, Collector test method under quasi-dynamic conditions according to the European Standard EN 12975-2, Solar Energy, 76 (2004) 117-123. [219] I. Standard, 9806-1 (1994), Test Methods for Solar Collectors, Part, 1. [220] L. Xu, Z. Wang, X. Li, G. Yuan, F. Sun, D. Lei, S. Li, A comparison of three test methods for determining the thermal performance of parabolic trough solar collectors, Solar Energy, 99 (2014) 11-27. [221] Q. Li, C. Zheng, X. Gu, A. Woffenden, G. Rosengarten, E. Hawkes, M. Yang, R.A. Taylor, DESIGN AND ANALYSIS OF A LOW-PROFILE, CONCENTRATING SOLAR THERMAL COLLECTOR, in: Proceedings of the 15th International Heat Transfer Conference, Kyoto, Japan, 2014. [222] M. Wiemer, Pfeiffer HI_Pace80Turbo_OperatingInstructions, in, file:///C:/Users/z3407130/Desktop/THESIS/Literature/PfeifferHI_Pace80Turbo_OperatingInstructi ons.pdf, 2010. [223] M. Wiemer, Pfeiffer_MVP020_AC_3_Diaphragm_Operating_Instructions, in, file:///C:/Users/z3407130/Desktop/THESIS/Literature/Pfeiffer_MVP020_AC_3_Diaphragm_Oper ating_Instructions.pdf, 2010. [224] P. Vacuum, Pfeiffer_PKR_251_Pirani_ColdCathode.pdf, in, file:///C:/Users/z3407130/Desktop/THESIS/Literature/Pfeiffer_PKR_251_Pirani_ColdCathode.pdf . [225] S.S.M. Tehrani, R.A. Taylor, P. Saberi, G. Diarce, Design and feasibility of high temperature shell and tube latent heat thermal energy storage system for solar thermal power plants, Renewable Energy, 96 (2016) 120-136.

214

[226] S.S.M. Tehrani, M. Saffar-Avval, Z. Mansoori, S.B. Kalhori, A. Abbassi, B. Dabir, M. Sharif, Development of a CHP/DH system for the new town of Parand: An opportunity to mitigate global warming in Middle East, Applied Thermal Engineering, 59 (2013) 298-308. [227] S.S.M. Tehrani, M. Saffar-Avval, S.B. Kalhori, Z. Mansoori, M. Sharif, Hourly energy analysis and feasibility study of employing a thermocline TES system for an integrated CHP and DH network, Energy Conversion and Management, 68 (2013) 281-292. [228] M. Esen, T. Ayhan, Development of a model compatible with solar assisted cylindrical energy storage tank and variation of stored energy with time for different phase change materials, Energy Conversion and Management, 37 (1996) 1775-1785. [229] H. Visser, Energy Storage in Phase-change Materials: Development of a Component Model Compatible with the" TRNSYS" Transient Simulation Program: Final Report, Office for Official Publications of the European Communities, 1986. [230] Y. Dutil, D.R. Rousse, N.B. Salah, S. Lassue, L. Zalewski, A review on phase-change materials: mathematical modeling and simulations, Renewable and sustainable Energy reviews, 15 (2011) 112-130. [231] H. Badenhorst, Performance comparison of three models for thermal property determination from experimental phase change data, Thermochimica Acta, 616 (2015) 69-78. [232] V.V. Mistry, Manufacture and application of high milk protein powder, Le Lait, 82 (2002) 515-522. [233] C. Ramirez, M. Patel, K. Blok, From fluid milk to milk powder: Energy use and energy efficiency in the European dairy industry, Energy, 31 (2006) 1984-2004. [234] M. Karagiorgas, A. Botzios-Valaskakis, Solar systems applications in the dairy industry, CRES’leaflet.(unpublished). [235] M. Lalwani, M. Singh, Investigation of solar photovoltaic simulation softwares, International journal of applied engineering research, 1 (2010) 585. [236] G. Kumaresan, R. Sridhar, R. Velraj, Performance studies of a solar parabolic trough collector with a thermal energy storage system, Energy, 47 (2012) 395-402. [237] R. Tamme, T. Bauer, J. Buschle, D. Laing, H. Müller‐Steinhagen, W.D. Steinmann, Latent heat storage above 120 C for applications in the industrial process heat sector and solar power generation, International Journal of energy research, 32 (2008) 264-271. [238] R. Hayes, S. Brewster, B. Webb, M. McQuay, A. Huber, Crown incident radiant heat flux measurements in an industrial, regenerative, gas-fired, flat-glass furnace, Experimental thermal and fluid science, 24 (2001) 35-46. [239] V. Siva Reddy, S.C. Kaushik, K.R. Ranjan, S.K. Tyagi, State-of-the-art of solar thermal power plants—A review, Renewable and Sustainable Energy Reviews, 27 (2013) 258-273. [240] C. Zheng, Q. Li, G. Rosengarten, E. Hawkes, R.A. Taylor, A novel semi-passive compact concentrator design and analysis with prism arrays, in, Applied Optics, 2016, pp. Unpublised. [241] F.J. Cabrera, A. Fernández-García, R.M.P. Silva, M. Pérez-García, Use of parabolic trough solar collectors for solar refrigeration and air-conditioning applications, Renewable and Sustainable Energy Reviews, 20 (2013) 103-118.

215

[242] E.T.a.P.J.-K. (IRENA), Renewable energy opportunities-for island tourism, in, International Renewable Energy Agency, http://www.irena.org/DocumentDownloads/Publications/IRENA_RE_Island_Tourism_report_201 4.pdf, 2014.

216