1999 ISES Solar World Congress

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ISES 1999 Solar World Congress Jerusalem, Israel

July 4-9, 1999

Editor: G. GROSSMAN

Conference Proceedings Volume 111

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The 1999 ISES Solar World Congress was held under the theme Solar is Renewable, adequately representing a Congress on the threshold of the 21 st Century in Israel- a pioneer in solar energy utilisation. We would like to thank our International team of Focal Point Editors and the many reviewers who helped make this event a success. We also wish to express our gratitude and thanks for their support to:

Ben-Gurion University of the Negev Israel Electric Company Israel Ministry of National Infrastructures Israel Ministry of Science Ormat Industries Ltd. Technion- Israel Institute of Technology Tel Aviv University World Energy Council Weizmann Institute of Science

Y. Zvirin - Congress Chair G. Grossman - Chair, Scientific Committee D. Dvorjetski- Executive Secretary H. Tabor-Chair, ISES Israel

vi

Congress Committee Y. Zvirin, Chair

D. Dvorjetski, Executive Secretary Organizing Committe D. Dvorjetski A. Elazari M. Epstein D. Faiman G. Grossman A. Kribus E. Shaviv H. Tabor D. Weiner Y. Zvirin Scientific Committee

G. Grossman, Chair J. Appelbaum D. Dvorjetski D. Faiman A. Kribus E. Shaviv D. Weiner Y. Zvirin Professional Tours Committee

D. Weiner, Chair U. Fisher Exhibition Committee

A. Elazari, Chair D. Dvorjetski Finance Committee

M. Epstein, Chair A. Shavit

Vll

International Advisory Committe D. Aitken, USA W. A. Beckmann, USA T. Book, UK A. Goetzberger, Germany Y. Goswami, USA O. Headley, Trinidad 1. G. Hestnes, Norway K. G. T. Hollands, Canada L. Imre, Hungary L.F. Jesch, UK H. S. Jeon, Korea D. Lorriman, Canada D. Mills, Australia

M. Nicklas, USA M. Oliphant, Australia E. de Oliveira Fernandes, Portugal D. Serghides, Cyprus L. Sherwood, USA 2. Silvi, Italy T. Tani, Japan M. Vazquez, Spain

viii

International R e v i e w B o a r d - Focal Point Editors

Solar Energy Systems for Buildings, Solar Architecture and Daylighting: A. G. Hestnes, NORWAY Flat Plate and Non-Concentrating Solar Collectors: W. A. Beckmann, USA Solar Thermal and Photovoltaic Concentrating Collectors: J. J. O'Gallagher, USA Photovoltaic Cells and Modules: M. Konagai, JAPAN Solar Collector Optical Materials: R. E. Collins, AUSTRALIA Solar Hot Water and Thermal Energy Supply B. D. Wood, USA Solar Thermal Electricity. A. Kribus, ISRAEL Photovoltaic Electricity:. J. Appelbaum, ISRAEL Active Cooling, Refrigeration and Dehumification: H.-M. Henning, GERMANY Space Applications: K. P. Bogus, THE NETHERLANDS Wind Power Systems and Solar-Wind Hybrids: M. Hirsch, ISRAEL Biomass Energy Conversion: R. P. Overend, USA Sustainable Hydroelectricity and Ocean Energy Conversion: D. Bharathan, USA Thermal Storage: J. Rheinlander, GERMANY Electrical Storage: D. Weiner, ISRAEL Hydrogen, Chemical Energy Storage, and Fuels: A. Heinzel, GERMANY Solar Radiation Measurement and Analysis: P. Ineichen, SWITZERLAND Indirect Solar Resource Evaluations: H. G. Beyer, GERMANY Education and Information Exchange: L. F. Jesch, UK Marketing and Commercialization: T. Book, UK Policy and Programs: A. Rabl, FRANCE Developing Countries: A. Ramachandran, INDIA Environmental and Social Impacts of Energy Systems: E. de Oliveira Fernandes, PORTUGAL Special Topics: C. Silvi, ITALY

ix

FOREWORD These volumes of Proceedings are the record of the 1999 ISES Solar World Congress, held in Jerusalem, Israel on the 45 th Anniversary of the International Solar Energy Society. The Congress was held under the theme Solar is Renewable, adequately representing a meeting on the threshold of the 21 st Century. The event also marks the 20 th anniversary of the Israeli Section of ISES, founded in 1979 - the year ISES celebrated its Silver Jubilee. The tradition of the biennial congress of ISES has been established since 1973. This Congress followed meetings in Paris, France (1973), Los Angeles, California (1975), New Delhi, India (1977), Atlanta, Georgia (1979), Brighton, UK (1981), Perth, Australia (1983), Montreal, Canada (1985), Hamburg, Germany (1987), Kobe, Japan (1989), Denver, Colorado (1991), Budapest, Hungary (1993), Harare, Zimbabwe (1995) and Taejon, Korea (1997). Israel- a pioneer in solar energy with the highest per capita utilization in the world - has for a long time expressed its interest in hosting the Solar World Congress. The Israeli Section of ISES is happy and proud to have had the opportunity to organize the Congress in Jerusalem this year. The Congress organizers have made great efforts to assure the quality of papers. The Congress Scientific Committee, in consultation with ISES, has developed a review procedure by which to accept papers for presentation at the Congress and publication in the General Proceedings. Due to time limitations, it was decided to base the review on extended abstracts of at least 400 words and up to one page. The abstracts submitted were screened by the Scientific Committee and then referred to Focal Point Editors, depending on their technical category. The responsibility of each Focal Point Editor was to handle the review of the abstract by referring it to three qualified reviewers in the respective area, receiving their comments, and making the final recommendations to the Scientific Committee regarding acceptance/rejection of the paper and required revisions. We have recruited an excellent Review Board consisting of 26 Focal Point Editors from around the globe, covering the full range of ISES topics. Following the Call for Papers, 464 abstracts have been received. Of those, 192 papers were accepted as submitted, 1 2 5 - with recommended changes, 1 1 3 - with mandatory revisions and 34 were rejected. The Congress was attended by over 520 participants, representing 47 countries. The Program included 207 oral presentations that ran in six parallel sessions during the five days of the Congress, and 149 poster presentations in three main sessions. In addition, 10 plenary lectures and 14 keynote lectures were presented. A business track under the title Solar Means Business included presentations and discussions on market implementation of solar technology. The Congress further included two panel discussions and two workshops, dealing with "WIRE" (World-wide Information System for Renewable Energy) and with IPMVP (International Performance Measurement and Verification Protocol). An exhibition presented the latest in solar products. We wish to express our sincere thanks to the international team of Focal point Editors, who have done a remarkable job in handling the review of the papers in an expedient manner, and to the many reviewers who helped make the Congress program a success. I personally wish to express my pleasure of working as a team, on all aspects of the Congress organization, with my two colleagues: Yoram Zvirin - t h e Congress Chairman and Dubi Dvorjetski- the Congress Executive Secretary. It is my hope that the Congress participants as well as those who were unable to attend, will find these Proceedings a useful reference and resource material, describing the state-of-the-art in solar energy. We look forward to the next Congress to be held in Adelaide, Australia in 2001. Gershon Grossman Editor Scientific Program Chairman

Table of C o n t e n t s - Volume III

Flat Plate and Non-Concentrating Solar Collectors Hybrid Solar Collectors for Microclimate Forming System G. J. Basler, D. Kwiecien ............................................................................................................................................................. 3

Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical W. S. Duff, D. Hodgson ............................................................................................................................................................... 4

Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator A. J. Fasulo, J. Follari ................................................................................................................................................................ Solar Air Collectors - Investigations on Several Series-Produced Collectors H. Fechner, O. Bucek ................................................................................................................................................................ An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions K. G. T. Hollands, G. W. E. van Decker ..................................................................................................................................... A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions K. G. T. Hollands, S. J. Arulanandam, E. Brundrett .................................................................................................................... Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material

11 17 23 29

X. -rn. Jiang, H. Baba, K. Kanayama, N. Endoh ......................................................................................................................... 35

Research and Development of Solar Collectors Fabricated From Polymeric Material A. I. Kudish, E. G. Evseev, M. Romrnel, M. KOhl, G. Walter, T. Leukefeld .................................................................................. 40

Study of a Mixed (Water Or Air) Solar Collector S. Laiot ...................................................................................................................................................................................... 50

Uncertainty in Solar Collector Testing Results E. Mathioulakis, K. Voropoulos, V. Belessiotis ........................................................................................................................... 50

Optimized Finned Absorber Geometries for Solar Air Heating Collectors K. Pottier, C. M. Sippel, A. Back, J. Fricke ................................................................................................................................. 62

Inclination Dependency of Flat Plate Collector Heat Losses G. Rockendorf, B. Bartelsen, M. Kiermasch ............................................................................................................................... 72

PV-Hybrid and Thermo-Electric-Collectors G. Rockendorf, R. Sillmann, L. Podlowski, B. Litzenburger ........................................................................................................ 76

Elastomer-MetaI-Absorber - Development and Application G. Rockendorf, B. Bartelsen, N. Vennemann, R. Tepe, K. Lorenz, G. Purkarthofer ................................................................... 83

Solar Absorber System for Preheating Feeding Water District Heating Nets K. Vajen, M. Kr~mer, R. Orths, E. K. Boronbaev, A. Paizuldaeva .............................................................................................. 90

Statistical Analysis of Solar Collector Test Results in View of Future Certification K. Voropoulos, E. Mathioulakis, V. Balessiotis ........................................................................................................................... 92

Thermal and Electrical Yield of a Combipanel H. A. Zondag, D. W. de Vries, A. A. van Steenhoven, W. G. J. van Helden, R. J. C. van Zolingen ............................................. 96

A Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differernt Reflection Models Y. Zvinn, B. Aronov .................................................................................................................................................................. 102

Solar Hot Water and Thermal Energy Supply Thermal Destratiflcation in Small Standard Solar Tanks Due to Mixing During Tapping E. Andersen, S. Furbo ............................................................................................................................................................. 111

Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina J. R. Banal, P. D. Galimberti, A. Barone, M. A. Lara ................................................................................................................ 120

In Situ Short -Term Test for Large Solar Thermal Systems N. Benz, T. Beikircher, M. Gut, P. Kronthaler, C. Oberdorf, W. Sch~lkopf, H. DrOck ................................................................ 126

•

Solar Process Heat with Non-Concentrating Collectors for Food Industry N. Benz, M. Gut, T. Beikircher, W. Ru/~ ...................................................................................................................................

131

Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material M. Bosanac, J. E. Nielsen ........................................................................................................................................................

137

Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems S. Co/le, S. L. de Abreu, R. R ~ h e r ..........................................................................................................................................

141

Solar Pond as a Power Source for Desalination U. Fisher. .................................................................................................................................................................................

150

Multistage Still J. Franco, L. R. Saravia, S. Esteban ........................................................................................................................................ 155

Development of a Smart Solar Tank S. Furbo, E. Andersen .............................................................................................................................................................

Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying H. P. Garg, R. Kumar. ............................................................................................................................................................... Medium Scale Solar Crop Dryers for Agricultural Products 0. Headley, W. Hinds .............................................................................................................................................................. The Marstal Central Solar Heating Plant: Design and Evaluation A. Heller, J. Dahm ...................................................................................................................................................................

160 170 175 180

A Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy B. J. Huang, V. A. Petrenko .....................................................................................................................................................

188

A Solar Still with Minimum Inclination and Coupled to an Outside Condenser D. Inan, A. El-Bahi ...................................................................................................................................................................

191

Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank U. Jordan, K. Vajen, B. Knopf, A. Spieler, F. Hilmer. ................................................................................................................ 197

Performance of Transparently Insulated Solar Passive Hot Water Systems N. D. Kaushika, K. S. Reddy ....................................................................................................................................................

203

Thermodynamic Study of a Regenerative Water Distiller G. Koury Costa, N. Fraidenraich ..............................................................................................................................................

211

The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy A. I. Kudish, E. G. Evseev, L. Horvath, G. Mink ....................................................................................................................... 216

Performance and Analysis of a Multiple Effect Solar Still Utilizing an Intemal Multi - Tubular Heat Exchanger for Thermal Energy Recycle G. Mink, L. Horvarth, E. G. Evseev, A. L Kudish ...................................................................................................................... 226

Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers G. L. Morfison, G. Rosengarten, M. Behnia ............................................................................................................................. 236

Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards - Project Results D. Naron, M. Rolloos, M. J. Carvalho .......................................................................................................................................

245

Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater N. Zhu, Ho Zinian .....................................................................................................................................................................

253

Solar Process Heat: Distillation, Drying, Agricultural and Industrial Uses B. Norton .................................................................................................................................................................................

256

Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference L Odicino, J. Marchese, D. A. Perelld, G. Lesino .................................................................................................................... 261

Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency J. Rekstad, L Henden, A. G. Imenes, F. Ingebretsen, M. Meir, B. Bjerke, M. Peter ................................................................. 265

Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally H. Romero-Paredes, E. Torijano, A. V~zquez, A. TorTes, J. J. Ambriz, E. Torijano Jr. ............................................................. 271

Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters L. J. Shah, G. L. MorTison, M. Behnia ......................................................................................................................................

276

A System for Solar Process Heat for Decentralised Applications in Developing Countries F. Sp~te, B. Hafner, Ko Schwarzer ...........................................................................................................................................

286

Performance of a Cascade of Flat Plate Collectors T. Tomson ...............................................................................................................................................................................

292

A Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors H. Zinian, Z. Ning .....................................................................................................................................................................

297

~

Xll

Advanced Fuzzy Control of the Temperature in the Test Chamber B. Zupancic, I. Skrjanc, A. Krainer, B. Furlan ...........................................................................................................................

304

Solar, Thermal and Photovoltaic Concentrating Collectors Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation G. C. Bakos, D. Adamopoulos, N. F. Tsagas, M. Soursos ....................................................................................................... The Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues X. G. Casals, J. I. Ajona ........................................................................................................................................................... A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver C. A. Estrada, R. Dorantes, E. Rincon ..................................................................................................................................... Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors T. H. Fend, Jo Leon, P. Binner, R. Kemme, K. -J. Riffelmann, R. Pitz-Paal ............................................................................... Experimental Performance of a PV V-Trough System N. Fraidenraich, E. M. de Souza Barbosa ................................................................................................................................ Performance Analyses of a Combined Photovoltaic/Thermal (PV/T) Collector with Integrated CPC Throughs H. P. Garg, R. S. Adhikari ........................................................................................................................................................ An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping M. Lando, J. Kagan, B. Linyekin, L Sverdalov, G. Pecheny, Uo Achiam .................................................................................. Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications R. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi ......................................................................................................................... Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System G. Miron, S. Weis, I. Anteby, B. Ostteich, E. Taragan .............................................................................................................. Simulation and Analysis of the Performance of Low Concentration PV Modules M. Munschauer, K. Heumann .................................................................................................................................................. Practical Design Considerations for Secondary Concentrators at High Temperatures J. O'Gallagher, R. Winston ...................................................................................................................................................... Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) J. O'Gallagher, R. Winston, J. Muschaweck, A. R. Mahoney, V. Dudley .................................................................................. Double-Tailored Imaging Concentrators H. Ries, J. M. Gordon .............................................................................................................................................................. Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors K. -J. Riffelmann, M. B6hmer, T. Fend, R. Pitz-Paal, C. Spitta, J. Leon ................................................................................... Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors M. ROnnelid, B. Karlsson, P. Krohn, B. Peters ......................................................................................................................... A Solar Bowl in India S. Rousseau, G. Guigan, J. Harper ......................................................................................................................................... The Development and Testing of Small Concentrating PV Systems G. R. Whiffield, R. W. Bentley, C. K. Weatherby, A. Hunt, H. -D. Mohring, F. H. Klotz, P. Keuber, J. C. Minano, E. Alarte-Garvi .........................................................................................................................................................................

315 324

333 337 342

349 354 358

367 370 377

382 388

394 400 405

409

Active Cooling, Refrigeration and Dehumidification Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products H. D. Arias-Vatela, W. Soto Gomez, 0. Castillo-Lopez, R. Bast-Brown ................................................................................... 419 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System W. S. Duff, R. Winston, J. J. O'Gallagher, T. Henkel, J. Muschaweck, R. Christiansen, J. Bargquam ...................................... 424

xiii

Indirect Evaporative Cooling through a Concrete Ceiling B. Givoni, S. Nutalaya ..............................................................................................................................................................

428

Experimental Studies on a Hybrid Dryer S. Kumar, G. A. Mastekbayeva, P. C. Bhatta, M. A. Leon ........................................................................................................ 434

Combined Solar Heating and Radiative Cooling System M. Meir, H. Storas, J. Rekstad .................................................................................................................................................

441

Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico J. L. Wolpert, M. V. Nguyen, S. B. Riffat .................................................................................................................................. 447

Thermal Storage The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube Y. Changsoon, S. Taebeom, K. Jaeyoon ................................................................................................................................. 455

The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants E. Uobet, J. Sold, J. Pitarch, J. Prats .......................................................................................................................................

462

Solar District Heating with a Combined Pit and Duct Storage in the Underground M. Reuss, J. Po Mueller. ...........................................................................................................................................................

468

Solar Heating with Heat Pump and Ice Storage A. B. Schaap, J. M. Warmerdam, E. E. Gramsbergen .............................................................................................................. 475

An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store A. Trp, B. Frankovic, K. Lenic ..................................................................................................................................................

484

Modelling of Two - Layer Stratified Stores J. van Berkel, C. C. M. Rindt, A. A. van Steenhoven ................................................................................................................ 490

Full T a b l e o f C o n t e n t s ......................................................................................................................................... 499 I n d e x o f A u t h o r s ...................................................................................................................................................... 512 I n d e x o f P a p e r s ......................................................................................................................................................... 542

xiv


ISES Solar World Congress 1999, Volume III

XVIII.

Flat Plate and Non-Concentrating Solar Collectors




HYBRID SOLAR COLLECTORS FOR MICROCLIMATE FORMING SYSTEM

Dariusz Kwieciel, Gerard Jan Besler Department of Environmental Engineering, Wroc fiw University of Technology, ul. Norwida 4/6, 50-373 Wroc ~w, Poland, tel. 00487 /3226435, fax. 00487 /3203532, [email protected]

Contemporary and modem residential buildings most often are characterised by well thermal isolated walls and very tight doors and windows, which results in a very small infiltration of the outside air. Then gravitational ventilation cannot serve its purpose satisfactorily. Mechanical ventilation is thus indispensable. It makes solar energy and energy from the shallow depth of the ground possible to be used in microclimate forming. It comes out not only possible but also effective. Research on such solutions heve been being carried for a few years in Air Conditioning And District Heating Chair in Wroc ~w University of Technology [ ,2]. Conventional energy for heating and cooling air purposes is replaced by natural renewable energy. In the paper an analysis of the possibilities of the natural solar energy gain in hybrid (liquid- air) solar collectors is presented. The collectors co-operate with thermoventilation and domestic hot water systems. On the experimental hybrid solar collectors station, which was made in technical scale, the measurements were made in natural climate conditions. The hybrid solar collectors experimental characteristics were estimated for water and air. Based on the known mathematical models which describe heat exchange in solar collectors, theoretical characteristics of hybrid solar collectors were made. For the average climate conditions in Wroc aw, the hybrid solar collectors operation efficiency in the heat - ventilation and domestic hot water systems were analysed (fig. ). The solar collectors co-operation with ground heat and mass exchanger were also included in the conducted analysis. Results of the analysis showed that the heat gain for ventilation and thermoventilation is more effective in the cold period of the year then for domestic hot water (fig. 2). Co-operation of the solar collectors and membraneless ground exchanger in the ventilation and thermoventilation systems gives very good result because energy gain from the ground is the most effective in the months with the lowest insolation level (XI + I). At that time solar energy input in solar hybrid collectors has low efficiency. Two natural energy sources: solar energy from solar collectors and ground energy from the ground exchanger supplement each other and thus the total quantity of gained heat energy is kept at the same level during almost whole of the heat season, it means in months: XI -III. System with ground exchanger serves not only for heating and domestic hot water but also allows for good ventilation with

air cooling in the summer and in some cases is able to replace expensive conventional air-conditioning system.

F i g . . The system discussed: a,b-solar collectors; 2-grave heat container; 3-heater; 4-ventilator; 5-throttles; 6-outside air intake device; 7-suspension ceiling; 8-intake ventilating device; 9uptake ventilating device; 0-heat container for domestic hot water system; - ground heat and mass exchanger.

Fig. 2. Total energy gained from natural sources during whole heating period (X+IV) in various solar systems. [ ] Kwieciel D., Besler G.J.: Thermoventilationsystem for accommodations with use of solar energy. Report K070 , Wroc fiw University of Technoligy, 997. [2] Kwieciel D.: Hybrid solar collectors for microclimate forming system. Dissertation for Doctor of Engineering Degree, Wroc hw University of Technoligy, 998.

Keywords: hybrid collectors, microclimate, natural energy, solar energy, thermoventilation.


TESTING OF A FIAT PLATE COLLECTOR WITH SELECTIVE AND NONSELECTIVE ABSORBERS THAT ARE OTHERWISE IDENTICAL

William S. Duff and David Hodgson Department of Mechanical Engineering, Colorado State University, Fort Collins CO 80523, USA Telephone: 1-970-493-1321, FAX: 1-970-495-0657, E-mail: [email protected] Abstract- We tested a flat plate collector using two absorbers that were physically identical, except for the fact that one was nonselective and the other had a good selective coating. We experimented over a range of input temperatures, keeping the input temperature to the collector constant throughout any one day. During a full day test, data was recorded every five seconds. We used several measures to insure steady state operation had been achieved before we selected two half-hour reporting periods symmetrical about solar noon. We found 1) that there was no difference in the efficiency of the collector with the selective absorber before and after stagnation and 2) that the efficiency of the collector with the nonselective absorber was substantially lower than the collector with the selective absorber.

1. INTRODUCTION We tested a flat plate collector using two absorbers that were physically identical, except for the fact that one was nonselective and the other had a good selective coating. A selective surface will exhibit a high absorptance for radiation in the solar spectnm~ and a low emittance for the longer wavelengths produced at the collector operating temperatures. Radiative losses are significantly reduced while high fractions of solar radiation absorbed are attained by the choice of a good selective surface for the absorber of a solar collector. The better the selective absorber coating, the more efficiently a given solar collector will operate at high temperatures and under low solar radiation conditions. One often wants to know the actual performance of a selective surface in an application. Though direct measurement of the properties of a surface can be made by various optical means, use of such measurements to predict actual losses due to radiation from the surface in the application introduces a variety of errors. These errors include inaccuracies in the optical measurement process itself and in extrapolating it to the conditions of the application. In non-evacuated collectors, conductive and convective losses are a significant fraction of the total losses. Conductive and convective losses can be modeled, but the complex geometry of flat plate collectors usually require a finite element approach to get good predictions. Even so, errors are introduced in the approximations and assumptions and one can never be sure that all factors have be taken into account. In-situ calorimetric measurements along with accurate measurements of solar radiation and other environmental factors can provide accurate estimates of the performance of a collector. However, because there are still uncontrolled, unmeasured or unknown processes, separating out the performance of a component of the collector, such as the absorber, from the performance of the collector itself can be a significant, often daunting, task. One way to approach this issue is to eliminate as many uncontrolled, unmeasured and unknown processes as possible and look at relative differences in the performance of the collector when one component property is changed. In our experiment, we used the same collector to evaluate the performance of two physically identical absorbers, one

nonselective and the other with a good selective surface. In addition, we evaluated the performance of the selective absorber before and after stagnation. 2. T E S T I N G SET-UP The collector was oriented south with a slope equal to the latitude of 40 degrees. The water flow rate was maintained as close to 2.2 kg/min as possible. The experimental equipment consisted of 9 Data acquistion software: Labtech by Keithly 9 Data acquisition hardware: HP 75000 Series B 9 Flow meter: Micro Motion model d-40 9 Boiler: Argo Industries model AI20-240 9 Power variac: Powerstat 1296d 9 Thermocouples: t-type (tested before installation) 9 Pyranometer: Epley Precision (calibrated before test) 3. E X P E R I M E N T A L APPROACH Efficiencies over a representative range of collector temperatures were experimentally determined and a collector efficiency curve was created for each of the three test sets using ANSI/ASHRAE standard 93-1986 as a guide. 9 The first test set was performed by testing the collector with the selective absorber before the absorber had been allowed to stagnate. To ensure that the absorber did not reach high temperatures the collector was covered at all times when it was not being tested. 9 The second test set was performed by testing the collector with the selective absorber after it had been allowed to stagnate for thirty days. 9 The final test was performed by testing the collector with the non-selective absorber. To create the curves, five or six of the most stable full day tests were selected from a much greater number of full day tests for each of the three test sets. During a full day test, data was collected every five seconds. Five parameters were recorded: 9 ambient temperature 9 temperature of the water entering the collector 9 temperature of the water leaving the collector 9 the water flow rate


9 radiation incident on the plane of the collector. MATLAB was used to analyze the data and calculate a representative instantaneous efficiency. To minimize the effects of the thermal mass of the collector, the instantaneous efficiency was calculated by averaging a morning and an afternoon efficiency measurement. The morning and afternoon efficiencies were determined by first finding two symmetric stable fifteen-minute intervals of data on either side of solar noon. The data for these fit~een-minute periods were then averaged and the efficiency was calculated by mass flow rate * specific heat * temperature rise across collector radiation level * aperture area of collector These stable periods were selected by searching the calculated instantaneous efficiencies in the four hour period that had solar noon as its midpoint, excluding 15 minutes on either side of noon, and selecting the symmetrically placed set of fifteenminute periods having the smallest combined standard deviation of efficiencies. Once the fifteen-minute periods were selected, they were checked to make sure they conformed to Section 8.3 of the ANSI/ASHRAE 93-1986 standard. If not, the day was eliminated.

4. RESULTS Table 1 shows the results of all of the selected test days. The entries are the average of the selected fifteen-minute morning and afternoon test periods. Figure 2 shows the average efficiency for each day as a function of the difference between the collector temperature and the ambient temperature divided by the radiation level. The collector temperature used was the average of the incoming and outgoing water temperatures. From Table 1 and Figure 2, it is clear that the selective surface outperformed the non-selective surface. The selective surface's efficiency was about 7 percentage points higher than that of the non-selective surface when the collector temperature was close to the ambient temperature. At higher operating temperatures, the difference was closer to 12 percentage points. The collector with the selective absorber installed was stagnated for a thirty-day period at collector temperatures ranging up to and above 170C. As can be seen in Figure 2, the testing indicates that there was no significant degradation in the performance of the selective surface due to stagnation. 5. CONCLUSIONS

Figure 1 shows a sample plot of data collected. The top graph shows the measured temperatures and flow rates. The bottom graph shows the incident sunlight and a calculated instantaneous efficiency. The vertical lines on the top graph indicate the two fifteen-minute periods symmetric on either side of solar noon which were used to calculate the average instantaneous efficiency.

There was little or no difference in the efficiency of the collector with the selective absorber before and aider stagnation. However, the efficiency of the collector with the nonselective absorber was substantially lower than that of the collector with the selective absorber.


Figure I" Sample Daffy Data Collection.


0 = Selective A b s o r b e r b e f o r e S t a g n a t i o n x = Selective A b s o r b e r after S t a g n a t i o n * = Non-Selective Absorber

Figure 2: Efficiency Curves for Different Absorbers.


Table 1: Testing Results SET 1" Selective Absorber Before Stagnation

4-Jun

Radiation Tamb Flow Tin (W/m^2) (C) (kg/min) (c) AM 959 28.2 2.2 70.2 992 30.2 2.1 71.8 PM

Tout

Temp.Rise

(O

(O

Efficiency

81.0 84.1

10.8 12.3

0.46 0.5

12-Jun

AM PM

928 926

27.4 31.8

2.3 2.3

14.8 15.0

31.2 32.1

16.4 17.1

0.78 0.81

13-Jun

AM PM

932 976

31.4 32.1

2.2 2.2

41.9 42.4

56.8 58.0

15.0 15.6

0.67 0.67

29-Jun

AM PM

935 961

32.6 33.1

2.3 2.2

83.3 83.2

93.4 93.5

10.1 10.3

0.46 0.45

6-Jul

AM PM

915 925

32.4 32.9

2.3 2.3

59.6 59.8

72.3 72.8

12.7 13.0

0.59 0.61

7-Jul

AM PM

928 922

37.5 38.0

2.3 2.3

50.2 50.3

64.8 65.1

14.6 14.7

0.67 0.68


Table 1 (Continued)" Testing Results SET 2: Selective Absorber After Stagnation Flow Tamb (C) (k~min) 32.4 2.3 2.3 33.2

8-Aug

AM PM

Radiation (W/mA2) 980 997

12-Aug

AM PM

1005 1026

33.7 34.3

23-Aug

AM PM

1032 1019

24-Aug

AM PM

3-Sep

AM PM

Tin

Tout

Temp.Rise

(c)

(c)

(c)

Efficiency

18.2 18.4

36.3 36.8

18.1 18.4

0.81 0.81

2.3 2.2

30.9 31.2

48.6 49.5

17.7 18.3

0.75 0.75

31.7 36.8

2.2 2.1

72.3 75.0

85.0 88.4

12.8 13.4

0.51 0.52

1083 1071

33.9 35.2

2.1 2.1

58.9 59.1

74.2 75.0

15.3 15.8

0.57 0.60

1036 1014

33.6 35.3

2.2 2.1

31.1 30.8

49.2 49.2

18.1 18.3

0.71 0.73

ISES Solar World Congress 1999, Volume Ill

10

Table 1 (Continued): Testing Results SET 3: Non-Selective Absorber Radiation Tamb Flow (W/mA2) (c) (kg/min) 1051 29.7 2.3 1010 30.9 2.2

19-Sep

AM PM

25-Sep

AM PM

1072 1061

29.2 30.7

27-Sep

AM PM

1074 1074

8-Oct

AM PM

10-Oct

AM PM

Tin

Tout

Temp.Rise

(c)

(c)

(c)

Efficiency

19.1 19.4

37.0 37.1

17.9 17.6

0.72 0.73

2.0 2.0

77.3 77.8

88.2 88.8

10.9 11.0

0.39 0.40

25.7 26.6

2.3 2.3

40.1 40.1

54.6 54.6

14.5 14.4

0.59 0.58

1065 1061

28.4 24.6

2.3 2.3

25.8 25.7

41.6 40.9

15.9 15.2

0.66 0.63

1049 1055

27.1 27.7

2.3 2.3

58.9 59.0

69.9 70.6

11.0 11.6

0.46 0.47


11

COMPARISON BETWEEN A SIMPLE SOLAR COLLECTOR ACCUMULATOR AND A CONVENTIONAL ACCUMULATOR Amflcar Fasulo and Jorge Follari Universidad Nacional de San Luis Chacabuco y Pedernera ? 5700 San Luis ? Argentina Fax 054 2652 430224 - e-Mail [email protected]

Abstract- We have shown that, in dry regions with abundant solar radiation at a latitude lower than 40~ as the central-western part of Argentina it is possible to obtain domestic hot water by means of very simple collector accumulators less expensive than the current ones. The experimental assessment of a solar accumulator collector yielding daily 3001 of hot water is reported in this work. Therefore, the diurnal and daily global efficiencies and the nocturnal thermal losses have been systematically determined over a six-month period, from austral summer to austral winter. The results are compared with those obtained from two other systems tested at the same time. These systems are also designed to yield daily 3001 hot water. They are: A high quality solar system composed of a 4 m2 plane collector and an accumulator storage insulated by conventional material; an integrated plane and accumulator collector, IPAC, whose semitransparent thermal insulation has been reinforced. The new systems themselves provide hot water over 40~ during six months and reduce energy expenses the remaining six months, when installed in series with systems using conventional energy sources. Graphs and tables show the results obtained, such as diurnal and daily global efficiencies and nocturnal thermal losses of the systems.

1. INTRODUCTION The city of San Luis (Argentina), situated at 33.27 ~ South and 66.2 ~ West, with a temperate and dry climate, posseses an abundant wealth of sunshine over most of the year. A similar situation is there for the whole western region of the country from 40~ latitude northward. This favourable situation for development of solar energy esploitation allowed it to constitute itself as pioneer in the development and use of solar water heaters, Follari et al. (1998). The relatively high cost of solar devices relative to conventional ones however has limited their diffusion, and on arrival of the natural gas networks of low cost, use of them almost has disappeared, Fasulo et al. (1999). In spite of these circumstances, in isolated form there still subsists the use of solar water heaters, in most cases connected in series with conventional devices in order to assure provision with warm water at any hour and during all days of the year. On the other hand in the poor districts of the town we observe painting with black colouring of the domestic water reserve tanks. The residents argue that by this means they dispose of hot water during the summer period; what for sure they don't realise is that during the winter period energy consumption for warming water increases. Anyhow, this attitude gives evidence of the existence of a clear concience on the possibilities to

take advantage of by use of solar energy, and the wish to participate in their use. While new materials arrive on the market it is possible to introduce innovations in the solar devices, in the continuous search for reducing their cost, Torres et A1 (1997). Thus, a simple envelope of alveolar polycarbonate would not just allow improving the efficiency of those rudimentary collectors, but also would reduce the negative effects of the black paint during nighttime, particularly in the winter months. With these ideas in mind, in 1997 we began experimenting with an integrated collector storage, ICS, consistent in a tank of stainless steel of circular cross section 1 meter high and with 384 liters capacity. The literature schows the development of ICS, that are completely different, Schmidt et A1 (1988). These ICS" are similar to the plane collectos in that they have an inclined front surface facing the sun, with sides and back surface protected by opaque insulation. The cold water, in the ICS, enters by pressure from another reservoir situated at a higher altitude, and the warm water flows out through a pipe situated at the center of the tank's lid. In its interior and parallel to the base, at some 2 cm distance, there is a plate of the same material with perforations far from its center. This plate has the function of avoiding that the cold water current entering the tank at the center of the base might destroy the established thermal stratification. The tank is covered with matte black colour and enclosed in a

12

ISES Solar World. Congress 1999, Volume III

box of alveolar polycarbonate of 4 mm thickness. This device was compared to two solar water heaters designed to provide some 150 liters of hot water daily; one of them of low cost and craftmanshipmanufactured, T(100). The other one industrially produced, the one of best quality obtainable on the local market, T(160). The achieved results show us that this ICS is capable of producing - at least during six months of the year, from mid-spring till midautumn - 150 liters of warm water per day with temperatures above 40~ Fasulo et al. (1998). The efficiencies of the three compared devices measured by: 1)Extracted water, T(av) is 43~ for ICS Vs 48.5~ for the T(100), in the summer; 36~ for ICS Vs 44~ for T(100)and 51~ for T(160), in the fall (04/11 to 04/30). 22~ for ICS Vs 29~ for T(100) and 37~ for T(160), in the winter (06/02 to 06/19) 2) Efficiency of the ICS as solar collector, 0 in(av), determined by measuring the temperature increment of the water of the interior of the tank, relative to the periods of each of the extractions morningtime, evening, and nighttime; this last one permits us to determine 3) L(av), the nocturnal thermic losses, around of 7 MJ for ICS Vs 1.5 MJ for T(160). 4) The daily net efficiency of the devices, 0in-net(av), including in the former the nocturnal thermic losses. The table gives us account of the good prospects that ICS offers, as 0in(av) in all cases results superior to that of the other devices; at the same time it reveals the main defect it has: This lies in the high nocturnal thermic losses, by this having the effect that 0in-net(av) is inferior to that of the two devices equipped with solar collector with plane plaques. Finally, we show the estimated amortization time of the ICS for different conditions where the ICS is put in series with a heater functioning on some kind of conventional energy, as there are: Gas in tubes, electricity, or natural gas from provision network. We found that the amortization periods of the ICS were of 6 years, 3.5 and a half years, and 16 years, repectively. 2. INTEGRATED PLANE AND ACCUMULATING SOLAR COLLECTOR (IPAC). In a second stage, developed between end 1997 and 1998, we modified the device demanding it a higher rendering. For this purpose we combined it with a plane solar collector of 2 r~ surface, and positioned above it we put the accumulating solar collector that had been used in the former experiment. In the present case we reinforced the semitransparent coveting of the ICS adding a second layer of polycarbonate, separated some 3 cm from the original layer. We doubled the demand on the device as for the volume of warm water to be produced, so changing to 300 liters of water daily obtained in three extractions: 100 1 in the morning,

before sunrise, 100 1 at mid-day, and 100 1 in the evening, immediately after sunset. In the experiment the device was compared to a high quality commercial water heater designed por producing 300 1 of hot water daily, composed of two collector plaques of 2 ~ each, connected to a reserve tank of 270 1 protected by opaque thermic insulation, conventional system, CS. The results showed that the device works quite satisfactorily during the 6 months around summer, with ouput temperatures above or very near 40~ A. Fasulo et al. (1999). The background part of figure 1 shows us both devices on the test bench. The thermal losses in this last case are slightly higher in the winter period. These are consequences partly of the diverse climatic conditions, but mainly due to the circumstance that the IPAC operates at higher temperatures than the ICS. In this experiment, consisting in the systematic daily extraction of 300 1 of warm water the way indicated before, during four periods of no less than 15 days each and extending these periods until each of them included days of full sun, days partly clouded, and cloud-covered, we determined: 1) An average temperature of the IPAC compared to the CS, of: A) 46EC vs. 49EC for the period 13th to 31 st of January. B) 38EC vs. 43EC for the period 3ra to 27th of March. C) 34EC vs. 40EC for the period 29th of April to 22~a of May, and D) 31 ~ vs. 38E C for the period 16th to 30th of June. Thermal losses during each of the indicated periods and for each of the two devices were of: 4.35MJ vs. 0.26MJ, 5.38MJ vs. 0.95MJ, 7.46 MJ vs. 1.28 MJ y 8.37MJ vs. 1.33MJ, respectively.

3. A NEW ICS Keeping in mind from the former experiments: That one of the major advantages of those devices is their high accumulative capacity and their greater exploitation of diffuse solar radiation, allowing them to overcome cloudy days in the provision with warm water, as well as the major defect they reveal (high nocturnal thermal losses), a new ICS was designed meant for comparison with the IPAC and the conventional system, CS, of 4 n~ collecting area and thus produce 300 1 of warm water daily. This new ICS was constructed with a metallic cylinder made of stainless steel and of 768 1 capacity, 2 m high, covered by three envelopes of alveolar palycarbonate of 4 mm thickness. Thermic control of the device is done by means of five thermocouples installed in the length of the central axis of the cylinder and positioned at: 5 cm, 50 cm, 1 m, 1.5 m, and 1.98 m from the tank's base. On the exterior surface of the tank and on the inner sides of each of the polycarbonate coverings, as well as on the lid, we put thermocouples at the same altitudes as those positioned in the interior of the device.


13

Fig. 1. View of the system when being built. At the back the two other systems for comparison.

4. THE IMPROVED IPAC In this third stage, we introduced the following improvement to the IPAC: Given that its worst defect consists in the thermal losses, their reduction was sought introducing a third covering of alveolar polycarbonate, making the calculus that by this means we would increase the thermal gradient by some 5~ resulting a total difference of about 25~ between the interior surface of the ICS an the environment. Figure 1 shows us a photo of the three solar heaters installed on a test bench. As in the previous cases, the experiments cover a minimum of 15 consecutive days that must be extended according to the requirement of comprising sequences of completely sunny and completely clouded days. Each of those days water extactions are done: One in the morning before sunrise, one at mid-day, and one in the afternoon immediately after sunset, determining the entrance and exit temperatures of the water, as well as all the other environment variables, i.e. maximum and minimum temperatures of the day, humidity, velocity and direction of the wind. Thermal control of the device is complemented determining all the temperatures in its interior and on its coverings immediately before and after each water extraction. 5. DATA ANALYSIS Data analysis in the first line presents the difficulty of comparing completely different solar collectors. On one side we have the CS that has a collecting area of 4 m2 inclined 45 ~ and north-oriented. The ICS has a net area of 2.17 m2 , in its major proportion composed of a vertical cylindric surface, and a horizontal surface - the lid. As the last we have the IPAC that is composed of a combination of those two. In table 1 we show the

dissimilarity of the sun-exposed surfaces, as well as the quantities of total radiation that reaches each of the devices at two extreme seasons of the year. We also find there the water quantity we should extract from each device if this would be performed proportional to the surface, starting from the condition that we can extract 300 1 daily from the CS. The volume we shall extract from the second experiment will be proportional as well to the areas of the devices as well as to the radiation that reaches them. Collector Surface m2 CS 4 ICS 2.17 IPAC 3.35

january n (MJ) 115 83 108

june propr extracted H (MJ) liters 94 100 100 52 60 70 76 86 90

Table 1. Areas of each device, radiation reaching them in each of the two extreme seasons of the year (summer and winter) on clear days, water volumes proportional to both and the extracted. Figure 2 shows us the exit temperatures of the IPAC for the 180 extractions performed, between Julian days 16 and 131; here we can appreciate the vast dispersion these data exhibit; owing to the changing climatic conditions, lower value data almost every time correspond to the extractions done in the morning, before sunrise. This pattern of dispersion is present in all three devices, with a distribution slightly above the shown for the CS and slightly below for the ICS, so that the data will be presented as mean values for each of the different experiments we performed. In table 2 we present the obtained results: In each columns the significant data of one experiment. The

ISES Solar World

14

Congress 1999, Volume III

files agrouped to show the types of data for each of the experiments. Where: Tair(av) is the average daily temperature of air. Tout is the temperature (average) of the extracted water. )T = Tout - Tin, where Tin is the temperature (average) of the cold water that had entered the device during the preceding extraction. Qout is the amount of energy in Mj gained at each extraction. Qint is the amount of energy stored in the tank during each of the diurnal periods between extractions of water (between 5 and 6 hours). L(av) is the nocturnal thermal loss of the water quantity the tank contained. 0r-out is the relative efficiency of the device calculated departing from Qout, in Mj, and the horizontal solar radiation, in Mj, summed up for the hours passed between the subsequent extractions. 0r-int is the relative efficiency calculated on the base of Qin (MJ) and the horizontal radiation. I

Experience/da

Fig. 3. Tout vs. secuence of water extractions of IPAC.

HI

n

t

iTai r (oc) F I (MY) Tout

,i

,

( ~C ) 9

)Tout = Tout-Tin 9 ( ~C ) Qout = miCTout 9 (MJ) Qint = CMiTint ,, ( M J ) L = CMiTint 9 (MJ) Or out = Qout/SiFI '0rint= Qint/St"I ,

1

1/2 days ICS IPAC CS 9

ICS IPAC . C ICS IPAC . CS ICS IPAC . CS ICS IPAC . CS ICS IPAC CS ' ICS IPAC i CS

23.9 ,,12"04

41.3 44.2 47.1

24.6 11.82 45.5 45.5 46.3

ml

i48

I

,+

10.99 3i.3 38.9 44.4

IV 17.5 8.39 34.7 43.5 46.7

V

9

16.4 8.81 31.5 38.7 42.7

VI 11.4 4.48 24.8 31.6 34.8

9

,

,

12.i 18.8 22 3.55 7.10 9.21 ---

'

9

20.2 23.0 . 26.0 8.44 9.63 . 10.87 16.67 14.55 . 16.34 -7.04 -5.02 .-0.41 0.34 0.29 0.24 ' 0.66 0.23 . 0.24

24.3 24.4 25.2 7120 9.19 10.55 31.43 14.54 16.48 -7.29 -5.40 -0.71 0.34 0.29 0.24 ' 0.62 0.22 . 0.24

12.9 20.3 26.0

S.40 8.48 10.86 11.65 14.73 16.75 -6.65 -4.87 -1.03 0.32 0.29 0.27 0.53 0.44 0.27

'

k

.

17.1 25.8 29.1 5.0 9.7 12.7 12.30 14.65 17.58 -4.73 -4.83 -0.28 0.40 0.47 0.39 0.72 0.58 0.39

'

14.2 21.4 25.5 " 4.16 8.08 10.66 9.66 11.00 . 17.42 -5.99 -5.34 -0.85 0.30 0.36 ~ 0.35 0.49 0.30 0.38

"

"

9

9

,

'

,

,

i

9 ---

,

,

ii .

0.36 0.47 0.51 ---

Table 2 Experimental results. I :16 to 34 Julian Days, extraction mi =100 1. Each device; II: 35 to 43 J.d. extraction ml=100 1. of CS, m2 = 90 1. of IPAC and m3 = 70 1. of ICS; III: 95 to 109 J.d., mi = 100 1. of each device; IV: 110 to 118 J.d., ml=100 1. of CS, m2 = 90 1. of IPAC and m3 = 70 1. of ICS; V: Idem IV 121 to 131 J.d.; VI: 165 to 178 J.d., ml=100 1., m2 = 901. and m3 = 70 1. ; M I = 270 1. (CS); M2 = 384 1. (IPAC) and M3 = 768 1.(ICS).


Fig. 4. Gradual development of Tout and H for the three devices, having extractions of 100 1 of water in each operation.

6. RESULTS In table 2 first of all we observe that with the ICS we obtained values of Tout above 40~ for the summer period, extracting 3001 of water per day. In the column Qint we f'md a poor energy accumulation. This is an important aspect for this type of collector having in mind its accumulation capacity, as it gives account of one of the most important advantages compared to the CS, as it allows overcoming isolated clouded days. In figure 4 we can see this characteristic of the ICS when the exit temperatures of the three devices for three summer days are shown and after a sequence two clear days follow two days of low radiation. The ICS overcomes the first clouded day satisfactorily, begins declining at the second, and later when full radiation returns shows the effect of working at the limit of its reserves presenting lower temperatures of the mornig extractions. In consequence a second experiment was designed, this time extracting water in quantities proportional to a combination of the net collecting surfaces of each device and the amount of radiation they receive. The second line of table 2 shows us the obtained results. There first of all we can see that now the exit temperatures of all three collectors are similar, Qout of the ICS drops and slighty also that of the IPAC. A strong increase of the Qint of the ICS can be observed, as we anticipated, while the other two exhibit little variation with respect to the former experiment. Figure 5 shows us the Tout of the three heaters for a sequence of 11 extractions, the first 6 corresponding to two days of plain sun, following one day of low radiation (the 4 subsequent events), and finalizing with a fourth day with full sunshine. From comparison of the two graphs we can deduce some conclusions about the limits of warm water production each of those devices impose.

55

TO oC

50

50

45

45

40

40

35 I .3~

3o

25

cs (o~))

2o

20 15 10 5 0

9

~

i

~

9

i

62

9

I

6,

9

15

6'~

9

6'~

Secuence of estractions bewteen 35 and 38 Julian Days

9

70

55

55

50

50

45

45

4O

40

T o 35

-m-i

ut 30' ~ 252 2o-~ 15-' lO.~ 52 oSr162

(MJ) ISC

-A-

IPAC (~

I

i

60

62

35 1 3O

-0--

(~

- V - cc

9

,

MJ

(oc)

9

i

64

"

66

"

i

"

68

70

o f e s t r a c t i o n s b c w t e e n 35 a n d 38 J u l i a n D a y s

Fig. 5. Gradual development of Tout and H for the three devices, having water extractions of: 1001 for the CS, 90 1 for the IPAC, and 701 for the ICS. The third line of table 2 shows the results of the first experiment of the southern auamm period, performed between the 5~ of April and the 10~ of May. Here it is intended to work in a way similar to the summer experiment, i.e. one sequence of measurements extracting 3001 of water from each of the devices and a second sequence applying different extraction values to each heater according to the considerations before mentioned. During the fast part environment temperatures very below those normal for the season were registered, reaching minima of-4~ typical for the winter seasonl. This results in the noteworthy decreases of the Tout of both accumulator collector devices, in consequence only the CC hardly surpasses the working temperature of 40~ In the second experiment of this period, fourth line of table 2, with the mean environment temperature returning to its normal condition of +2.5~ compared to the former, but with a stronger cloudiness having the effect of a reduction of 5 MJ in the average daily H, we performed the second designed part, i.e. extracting 70 1 from the ICS, 90 1 from the IPAC, and 100 1 from CC, respectively. Compared to the previous series we see that the conventional device following the environment temperature increases Tout by 2.2~ whereas the ICS increases its Tout by 3.4~ and the IPAC by 4.5~ the increment in the )T is of 3~ 5~ and 5~ respectively. The autumn experiment is completed with

16


a series of 10 days, measured beginning with the 11th of May, fitth line of table 2. We can observe that 0 r - o u t as well as 0r-int of the CS, are lower in summer. This is a direct consequence of the fact that the plane collectors, inclined 45 ~ directed to the north, miss the first and the last hours of solar radiation, whereas the ICS receives radiation all the time. In the sixth line we show the results of the winter period, measured between the 14~ and 26th of June.

7. CONCLUSIONS 1) 2)

3)

The ICS takes better advantage of the solar radiation than the CS, in summer. The ICS can replace the CS profitly in technical and economic respect during the summer months, most of spring, and beginnings of autumn in regions temperate, dry and with abundant solar radiation. Search for improvements of the semitransparent thermal isolations must go ahead, in order to

succeed in making the ICS competitive for the complete annual period. REFERENCES Follari J y Fasulo A. (1998) Veinte afios con los calefones solares Argentinos. Energias Renovables y Medio Ambiente. 5, 1 - 6 Fasulo A., Perello D. And Follari J. (1998) World Renewable Energy Congress V. 4, 2307 - 2310 Torres M., Follari J. and Fasulo A. (1996) An/disis t6rmico comparativo entre colectores pianos con cubierta de vidrio y policarbonato. ASADES I , 5.17 Schmids Ch., Goetzberger A. and Schmid J. (1998) Test Results and Evaluation of integrated collector storage systems with transparents insulation. Solar Energy 41, 5, 487 Fasulo A., Perello D. And Follari J. Comparison Against collector accumulator with semi-transparent insulationand and conventional. EuroSun 98.2, III, 3, 4-1


17

SOLAR AIR COLLECTORS INVESTIGATIONS ON SEVERAL SERIES-PRODUCED COLLECTORS Hubert Fechner, Otto Bucek Dept. of Renewable Energies, Arsenal Research - Austrian Centre for Research and Testing, Faradaygasse 3, A-1030 Vienna, Austria, Phone: +43-1-79747-299 Fax: +43-1-79747-390, email: fechner.h@_,arsenal.ac.at

A b s t r a c t - Testing of solar liquid collectors is described in international standards (ISO 9806, prEN 12975). A standardized procedure for testing air collectors does not exist so far. Within Task 19 of the IEA-"Solar Heating and Cooling Program" tests of most of the few worldwide available types of solar air collectors were carried out. Collectors from Australia, Canada and several from Europe were tested at the Austrian Research Centre Arsenal. Development of testing conditions and appropriate presentation of results, as well as tests on efficiency, leakage, pressure drop, dependence on internal mass flow and wind effects was the aim of this project. Long-time proven products as well as promising prototypes have been tested. Various types of collectors were investigated: glazed modules with air flow below or on both sides of the absorber, different colored absorber and corrugated profiles, textile absorbers but also cheap site built collectors and uncovered perforated collectors. The choice of air flow pattern depends on the application. Generally, four distinct air flow patterns exist: A very simple air collector can be constructed with the air flow between the absorber and the glazing. Due to the high heat loss induced by the convective heat transfer to the glazing the efficiency will be quite low, especially, if a relative high temperature increase is needed. The air flow behind the absorber is probably the most common solution. Air flow on both sides of the absorber is used in medium temperature applications to increase the effective heat transfer area. The fourth air flow pattern - air flow passes through the (porous) absorber - offers the possibility of a cost effective solution for medium temperature collectors.

I. INTRODUCTION Active Solar Air Systems have been known for many decades. Although they are not very wide spread so far, solar air applications are a promising way of meeting the heating demand on ecological basis.

-

-

Starting with simple constructions at the end of the 19th century in the U.S.A., solar air systems are now in use for space heating, preconditioning of air as well as for cooling applications, for hay drying, for drying of tobacco, crops, fruits and timber. Air heating is tightly connected to architectural matters, but contrary to passive solar design, active air systems provide better heat distribution and regulation, which results in improved heat gains and finally more comfort. In order to pool the experience in designing air systems for space heating, the International Energy Agency (IEA) initiated a five year project: Within Task 19 "Solar Air Systems" of the "Solar Heating and Cooling Programme" more than twenty experts from nine countries, coordinated by the Operating Agent Arch.Robert S. Hastings, have worked out:

A book illustrating 33 exemplary buildings with diverse solar air systems. A catalog of manufactured components and guidelines for selecting them, A PC-based, easily used program to predict energy performance and comfort, A handbook for designing Air Heating systems.

2. GENERAL SPECIFICATIONS The central part of each solar system is the collector, where the energy radiated from the sun is collected and converted such that it can be easily distributed atterwards. Solar air collectors as active part of air heating systems are not really "common" products so far. Testing of various series produced air collectors in a reproducable and comparable way have not been done so far. Worldwide not even a dozen of companies are manufacturing air collectors, scarcely half a dozen have a relevant output. Beside manufactured collectors, site built collectors are important as well. Founded by the Austrian Ministry of Science the Austrian Centre for Research and Testing "Arsenal Research" has invited

18


manufacturers of solar air collectors to bring their products to Arsenal for being tested. Seven manufacturers from seven different countries, mainly from Europe but also from Canada and Australia, had taken up the offer to have their products tested. Long-time proven products have been tested as well as promising prototypes. The main topics of development, investigation and research during this project have been: -

-

-

-

-

Development of a steady state testing procedure for solar air collectors, suited for all types Discussion on physieaUy suitable efficiency presentations Development of different performance descriptions adequate for all common operation modes A comparison of available products Investigation of the technical behaviour of different types of air collectors Recommendations for an optimised utilisation of solar air collectors Recommendations for improvements of tested products Recommendations for a standardised testing procedure of air collectors, which can finally be integrated into an internationally starting standardisation process on ISO or CEN level. Adaptation of the existing solar-laboratory-facilities for testing solar air collectors

Testing requirements for solar-liquid collectors have been intensively developed since about 1980, now, beside an ISO (International Standardisation Organisation) standard there is also a new CEN (Comit6 Europden de Normalisation) testing standard under development. For solar-aircollector-testing there exists no standard. Compared to water collectors, the measuring procedure for solar aircollectors needs even more expenditure. Generally, measuring of air-temperatures and air-mass-flows requires higher effort for gaining comparable a c c ~ e s . Moreover, leakage, the air flow distribution inside the collector and the much lower heat transfer from the absorber to working fluid are further complex affects. Opposite liquid solar collectors the efficiency of solar air collectors is strongly influenced by the actual mass flow rate inside the collector due to the often rather low heat transfer coefficient between absorber and air. This heat transfer coefficient is highly dependent on the air speed. It is, therefore, often difficult or even impossible to extrapolate from tests of small modules of solar air collectors in test rigs to larger solar air collector arrays as the air flow pattern might be different.

The physical behavior of air collectors differs from liquid collectors mainly due to the much lower heat transfer and the lower heat capacity of air. Effects of uneven air flow pattern due to unproper installation and connection of the collectors with the manifolds in collector arrays makes the prediction of the solar gain difficult. The result of this work is a comparison of different collector types, further collector-improvement in collaboration with the manufacturers but also a detailed description on how to deal with the presentations of air collector efficiency.

3. EFFICIENCY PRESENTATIONS Optical features (absorptance and emittance of the absorber, transmittance of the cover), materials used (absorber material, cover material, frame, insulation) and constructing characteristics (mainly the airflow-principle and the effective heat transfer area) of the collector are of basic importance for the efficiency. However, the respective operation condition of the collector is decisive as well and the efficiency decreases with increasing temperatures within the collector because of the increasing heat losses. The efficiency of a solar-(air)-collector is defined as the ratio of useful gain of the collector (Qu) to the respective insolation Gr at the collector reference area A~.

111= Q u _

Qu Q sot AcGT

The useful gain is described by the massflow, the heat capacity and the temperature rise:

Qu =rn*c p * (To - Ti) As collector reference area can be considered: aperture area, absorber area or gross area. A general equation for solar collector performance is based on Hottel, Whillier and Bliss:

110= Fo['C(X-UL(To-T,)IQ,ol] where Ta is the ambient temperature and Fo is the collector heat removal factor in relation to To - the collector outlet temperature and TI0 is the efficiency when the outlet temperature is taken as reference. Fo accounts for the fact that the absorber temperature is not the same as the outlet air collector temperature neither in the horizontal, direction nor vertical.


It appears that the efficiency rl depends on the operation conditions of the collector. It decreases with increasing temperatures, because of the increasing heat losses. Important is, how to define the operation conditions of the collector given by the temperature difference between the "overall collector temperature" Tk and ambient Ta. An efficiency curve can be drawn in dependency of a certain reference temperature which corresponds to the collector temperature Tk. In a physically correct way one has to take a weighted mean temperature Tk of the whole collector box, but in the measuring practice this is not practicable. That is why three temperatures are for choice: The inlet temperature (Ti), the outlet temperature (To) and a so called ,,mean" collector temperature (Tm) which can be calculated as the arithmetic mean value between inlet and outlet temperature. Efficiency curves of a solar collector corresponding to the three possible reference temperatures for a constant mass-flowrate are shown below:

19

Measurements indicate that the physical mean temperature (Tk) of the collector is often much closer to the outlet temperature (To) than to the arithmetical mean temperature (Tm). Therefore the presentation of the collector efficiency curve using the outlet temperature (To) often seems to be the best solution. A proposal for doing this is also given by DuffleBeckmanr~ (Solar engineering of thermal processes, J.Wiley&Sons Inter-science, New York 1991)

4. AIR C O L I ~ C T O R TYPES For different applications of air heating systems (space heating, preconditioning, drying-processes.... ) different collector types are the best choice each. Moreover, the result of this project was a contribution to the handbook for designing solar air systems which is one main output of the IEA-Task 19 collaboration, as well as recommendations for standardising of air collector-testing procedures and an input for further discussions on evaluating the performance of solar air collectors. In principle there are 4 different construction modes of solar air collectors: -

-

Flow below absorber, "Underflow'' (mass-flow behind the absorber, the air gap between absorber and cover operates as insulation) Flow above absorber, "Overflow" (mass flow only between absorber and cover) Flow on both sides of the absorber Perforated absorber, the air flow penetrates through the absorber (black felt, porous metal .... )

Fig. 1 Efficiency related to different Reference Temperatures For solar liquid collector it is the custom to present the efficiency related to the "mean collector temperature" (Tm) representative for the heat losses of the collector. For liquid collectors, where the temperature difference between inlet and outlet is very small (normally less than 10 K) and the heat transmission from the absorber to the fluid is high, the arithmetical mean value (Tm) is in fact very close to the physical mean temperature (Tk) of the collector.

For air collectors the difference between inlet and outlet can be up to 30K or 40K dependent on the mass flow. Also important is the amount of heat transmission from the absorber to the fluid, which is for air collectors usually not that high. Due to these effects there will be no longer a linear increase of the fluid temperature along the absorber plate and the arithmetical mean value (Tm) is often not representative for the heat losses of the collector.

Fig. 2 Air flow principles in Solar Air Collectors

20


Influence of different air flow principles:

Advantages and disadvantages: Flow above absorber: + Simple construction - High losses, especially at a high difference between absorber and ambient temperature - Highly decreasing efficiency at high surrounding air velocities (mainly depending on the cover) - Only one surface of the absorber is used as effective heat transfer area - Double glazing reduces the losses but decreases the solar input Flow below absorber: + Air in the gap between absorber and glazing operates as insulation (few losses at high differences between absorber and ambient temperature) - Only one surface of the absorber is used as effective heat transfer area Flow on both sides of the absorber: + Double effective heat transfer area - At high differences between absorber and ambient temperature the heat losses due to the hot air directly under the cover increase and the dependency of surrounding air speed decrease Porous absorber: + High heat transfer-coefficient - High pressure drop - Depending on ambient air conditions (dust, pollution...) the absorber is often under a high technical stress.

3...unglazed perforated trapezoid absorber panel, aluminimn, anthrazit, strongly dependent on wind 4...glazed plane absorber, facade element, underflow 5...glazed, rippled absorber, air flow on both sides 6...glazed plane absorber, facade element, air flow on both sides 7...glazed site built collector, selective absorber, trapezoid profile, underflow 8...glazed plane absorber, facade element, underflow From liquid systems we know that judging a solar system is often to much concentrated on the thermal performance of the collector; other features of the system like control strategy, mounting of temperature sensors, connecting the modules, storages, insulation matters and many other questions should also be considered carefully. But this project focused on the assessment of the features of air collectors. As a result of these investigations a lot of hints can be derived how to built a solar air collector for a certain application. 5. OPERATING SOLAR AIR COLLECOTRS If nmning a solar aircollector usually 4 multiple combined effects must be considered: A) The higher the mass-flowrates the higher the efficiency (reason: at higher mass-flowrates two combined effects occur: the heat transfer from the absorber to air increases and the Outlet temperature- and therefore the heat losses decrease) B) The effect of air flow leakage increase with the air flow-rate C) The electrical power requirement for the fan increase with the mass-flowrate D) For heating purposes a certain temperature level is often needed, which further restricts the possible mass-flowrates

Fig. 3 Efficiency of Solar Air Collectors 1...glazed collector (low iron), aluminium absorber with uprofiles, selective coating, underfiow 2...glazed collector, black textile absorber

Fig.4 Temperature Rise and Efficiency Large effective heat transfer-areas are advantageous but constructions where the air is forced to flow in tight profiles

ISES Solar World Congress 1999, Volume !!!

causes high pressure drops. It is a challenge for the constructing engineer to find a compromise in high heat transfer and low pressure drop.

21

conditions. The leakage rate depending on the mean static pressure of the collector should be determined generally.

Air-flow pattern: Pressure drop:

The air flow pattern inside the collector is very important for a correct assessment of the performance. Normally, if you work with one collector only, the air flow pattern near inlet and outlet is often not satisfactory. To reach an even air flow pattern from inlet to outlet for each tested collector a special connection box was built.

The Pressure drop is important for the number of collectors in series and the electrical power of the fan. Pressure drop is about a square function of air velocity. Pressure drop increases about linear with air density.

Temperature measurements: During an engineering process for solar air systems, choosing an air collector should be based considering the following aspects: -

-

-

-

The desired temperature rise A low temperature rise (f.e. in stores, factory halls, sport halls, drying and preconditioning processes...) otten favours cheap collector constructions (simple absorber profile, cheap or even no glazing, minor insulation) A high temperature rise (for heating offices, living rooms .... ) mainly need high performing constructions The design and optical features Location, climate, orientation The costs Reasonable pressure drops; (Optimisation according to the temperature rise needed)

While the measurement of the inlet temperature is rather easy, to achieve a precise determination of the outlet temperature is difficult. Several layers of different air temperatures are often close adjacent, a ~ i f i c mixing device- optimized according to fluid dynamic experiences - at the outgoing duct just in front of the sensors and a sophisticated arrangement of temperaturesensors are necessary. Effects of condensation must be considered carefully.

Conditioning: For testing the collector with different air temperatures, the preparation of differently conditioned air is needed. An enclosed climatic chamber with 150 m 3 met these require-ments. Temperatures between some degrees below zero and up to 60 ~ could be reached at the collector inlet. Control devices and various control heating devices cared for stability.

7. C O N C L U S I O N 6. EXPERIENCES MADE CONCERNING THE TESTING PROCEDURE AIR COLLECTOR TESTING -

R

E

C

O

M

M

E

N

D

A

T

I

O

N

S

F O R

No standardised testing procedure exist for Solar Aircollectortesting so far. Starting a standardisation process for testing solar aircollectors has been already discussed in the Technical Committee 180 of the International Standardisation Organisation (ISO), but work is still resting. (Reuss et al, 1993) Generally, measuring of air-temperatures and air-mass-flows requires higher effort for gaining satisfactory accuracies. Moreover, leakage, the air flow pattern inside the collector and the much lower heat transfer from the absorber to the heattransfer-medium are further complex affects.

Leakage: For an accurate measuring process 2 fans are needed, one at the inlet and one at the outlet. For testing reasons the fans had to work such, that the mean static pressure at the collector is equal atmospheric pressure. Only by that you can manage the leakage rate to be minimised. It was also possible to carry out tests with only one fan near inlet or outlet the collector to simulate realistic

Solar air collectors are not wide spread so far. As main obstacle for a wide dissemination appears lacking information as well as lack of confidence on how these systems will work. Testing of the respective components is therefore essential. Test results from independent test institutes can serve as efficiency proof. Testing within this project showed a wide range of effectivity among the different products. Experiences how to optimise a collector as well as precise recommendations for a testing procedure of air collectors are result of this project. Further Investigations seem to be necessary in the general issue of presenting the thermal energy output of air collectors. The problem with reference a r e a - well known from liquid collectors as well as the problem of the reference temperature are open for further discussions.

-

22


REFERENCES 1. Duffle J. & W. Beckman. ,,Solar Engineering of Thermal Processes "' J . Wiley & Sons Interscience, New York 1991) 2. Morhenne J. & M. Fiebig. Entwicldung und Erprobung einer Baureihe yon optimierten, modularen Solarlufierhitzern .f'dr Heizung und Trocknung, Ruhr-Universit~t Bochum, 1990. 3. Dai, Hui and Li. Fully developed laminar flow and heat transfer in the passages of V-corrugated solar air heater, ISES "91, Denver, proceedings. 4. Lo, S. N. G., Deal, C. R. & B. Norton..4 School Building Reclad with Thermosyphoning air panels, Solar Energy Vol. 52, No. 1, pp. 49-58, 1994. 5. Biondi, P., Cicala L. & G. Farina. Performance analysis of solar air heaters of conventional design, Solar Energy, Vol. 41, No. 1, pp. 101-107, 1988. 6. ,41tfeld, K. Leiner, W. & M. Fiebig. Second Law optimisation of flat-plate solar air heaters, Part 1: The concept of net exergy flow and the modelling of solar air heaters. Solar Energy 41, 127-132, 1998. 7. ,41tfeld, K~ Leiner, W. & M. Fiebig. Second Law optimisation of flat-plate solar air heaters, Part 2: Results of optimisation of and analysis of sensibility to variations of operating conditions. Solar Energy 41, 309-317, 1998. 8. Reuss, M. Recommendations for standard procedures for testing of air heating solar collectors, Bayrische Landesanstalt fLandtechnik, ,4ugust 1993 (ISO TC 180/SC5/N53) 9. Corazza ,4., et. al. Design, development and performance studies of a large sized solar air heater in nonconventional mode of operation. Int. Conf. ,41ternative Energy Sources Today and for the 21st century. Brioni, oct. 5-8, 1988. 10. CE-Standard of solar collectors, Thermal solar systems and components - Collector- General requirements, CEN TC 312 N164, a draft paper by CE TC 312-PT1. 11. Gupta D., Solanki, S.C. and J.S. Saini, Thermohydraulic performance of solar air heaters with roughened absorber plates. Solar Energy Vol. 61, No. 1, pp. 33-42, 1997. 12. Keller, J., V. Kyburz and `4. K6lliker, Untersuchungen an Lufikollektoren zu Heiz- und Trocknungszwecken, 1988. Schlussbereicht des KWF-Projektes Nr. 1296. Paul Scherrer Institut, W'renlingen und Villingen, CH-5232 Vilh'ngen PSI. 13. ,4bbud, I..4.., G.O.G. L6f and D.C. Hittle, Simulation of solar air heating at constant temperature. Solar Energy Vol. 54, No. 2, pp. 75-83, 1995. 14. Matrawy, K.K., Theoretical analysis for an air heater with a box-type absorber. Solar Energy Vol. 63, No. 3, pp. 191-198, 1998. 15. Jensen, S.O., Roof Space Collector, Validations and simulations with EMGP2. 1987a. Institute for Energy and Building, Technical University of Denmark, Report No. 87-15. 16. Jensen, S.0., O. Olesen and F. Kristiansen. Lufi/vceskesolfangee. 1987b. Solar Energy Centre Denmark, DTI Energy. ISBN: 87-7756-470- 7.

17. Morck, O. and P. Kofod, Udvikling af luftsolfanger. 1993, Cenergia Energy Consultants, Denmark. ISBN 8 7-90314-02-6. 18. Muff, Christoph; Solarluft~steme- Vortrag Trisolar 98 Bregenz/,4 ustria 19. IEA Solar Heating and Cooling Programme, 1998 Annual report with a feature on Solar ,4ir Heating, Morse Ass. Inc. 1808 Corcoran Street, N.W. Washington, DC 20009, USA, March 1999.


23

AN EMPIRICAL HEAT TRANSFER EQUATION FOR THE TRANSPIRED SOLAR COLLECTORS, INCLUDING NO-WIND CONDITIONS Gerald W. E. Van Decker Active Solar Energy Technologies, Natural Resources Canada, 580 Booth St. Ottawa, Ont. Canada, K1A 0E6, 613 996-3648, 613 996-9416, [email protected] K. G. Terry. Hollands Department of Mechanical Engineering, University of Waterloo, Waterloo, Ont., Canada N2L 3G1, 519 888-4053, 519-746-0852, kholland @ solar 1.uwatedoo.ca

Abstract - The unglazed transpired solar collector is now a well-recognized solar air heater for heating outside air directly. Example applications include pre-heating ventilation air and heating air for crop drying. The outside air in question is drawn straight from ambient, uniformly through the whole surface of a perforated blackened plate (the absorber plate) exposed to the sun. An important parameter fixing the collector's efficiency is the heat exchange effectiveness, ~. Once e is known, finding the collector efficiency is straightforward. Recently, Van Decker et al. presented measurements of e under various wind speeds and suction velocities plates perforated with circular holes of various diameters and spacings, laid out on either a square or triangular layout. They also developed a predictive equations for ~, which contained various parameters adjusted to fit their measurements, but their equation did not cover wind speeds down to zero (still air). The present paper extends that earlier work so as to cover the zero wind speed case. This new model predicts the measured data of Van Decker et al. andKutscher with a root mean square error of about 5.5 %. (This degree of uncertainty would lead to a contribution to the error in the predicted efficiency of roughly 2.5%.) The model also gives the breakdown of the contribution to the total heat transfer from each of the plate regions: the front, the hole and the back. 1. INTRODUCTION Unglazed, transpired solar collectors (HoUick and Peter, 1990; Kutscher et al., 1991, 1993, 1997) have been the subject of a number of investigations. They are effective devices for applications where outside air is to be heated directly, such as in heating ventilation air for buildings and crop drying. The outside air in question is drawn straight from ambient, through the whole surface of a transpired, darkened plate (the absorber plate). Tests conducted on several installations indicate that the unglazed transpired collector (UTC) gives annual solar collection efficiencies reaching 72% (Carpenter and Kokko, 1991). Typical installations have simple paypack periods of 2-8 years, making the UTC an attractive investment. Over 70 large systems each having collector areas between 500 to 10,000 m 2 have been installed and are successfully operating for fresh-air heating in Canada, the United States, Germany, and Japan and heating process air for crop drying in countries throughout the world.

1.1 Prior Work on Heat Transfer Modeling A recent review of the heat transfer principles of the UTC has been given by Hollands (1998). Kutscher et al. (1991, 1993) determined that, as the air travels across the face of the collector (driven by the wind), the thermal and velocity boundary layer thicknesses reach an asymptotic value at a very short distance from the edge of the plate (about 0.1 m), so that almost all the plate is in the asymptotic region. This fact laid the foundation for their performance model. When one approximates the radiation loss by a linear equation (HoUands, 1998), the Kutscher et al. (1991) model for the collector efficiency r/reduces to

s/(1

h /(

CpV s ) )

r

where txs is the solar absorptivity of the plate, hr is the radiative heat loss coefficient, Vs is the superficial suction velocity (volumetric rate at which air is sucked through plate, per unit area of plate), eis a "heat exchange effectiveness" (see below), and p and Cp are the density and specific heat of the air, respectively.

Values of r/ranging from 50 to 80% are common in practice, and typically Vs ranges from 0.03 to 0.08 m/s. A key item in the efficiency equation is the quantity e, defined by -

ro-r. L-T

,

(1)

where To is the mean air temperature leaving the plate at the backside and p is the plate temperature, and T, is the ambient air tempearture. Effectiveness e has to be evaluated, say from experimental data, or from analysis, but once it has been evaluated, determining r/is straight forward. Based on his extensive measurements on relatively thin plates, Kutscher (1994) presented a predictive model for e, for UTCs with circular holes on a triangular layout. Cao et al. (1993) and Golneshan and Hollands (1998) reported numerical and experimental correlation equations for e for a plate with perforations consisting of an array of slits, with the wind flow assumed to be transverse to the slits. Using computational fluid dynamics (CFD), Arulanandam (1995) (see also Arulanandam et al. (1999)) obtained a correlation equation for e, applying to a plate with circular holes on a square layout, but only under no wind. Van Decker et al (1996) reported an extensive set of measurements on thin and thick plates with circular holes on a square or triangular layou, over a range of typical wind speeds, as well as a correlation equation for ~, fitted to their measured data. 1.2 Present Study The present paper extends the treatment of Van Decker et al, to cover the no-wind conditions. It presents a new model for e that is shown to predict both Kutscher's data and that of Van Decker et al., within an rms error of roughly 5.5 %. Like the Van Decker et al model, the new model gives a breakdown of the contribution to the heat transfer on each of the parts of the plate: the outside face, the hole, and the back of the plate. The breakdown given in the present paper is expected to be more accurate than that given


24

Table 1: Characteristics of the Plates Tested by Van Decker et al. (1996) Plate No. 1 2 3 4 5 6 7 8 9

Plate Material

Aluminum Aluminum Polyvinyl Chloride Stainless Steel Polyvinyl Chloride Polyvinyl Chloride Polyvinyl Chloride Stainless Steel Polyvinyl Chloride

Hole Pitch* P (ram) 16.89 16.89 16.89 13.33 13.33 8.00 24.00 24.00 6.67

Diam. D (ram)

Thickness t (mm) 1.60 1.60 1.60 1.60 0.79 1.20 3.60 3.60 0.93

Plate Thermal Conductivity, k, W/(mK) 0.86 0.86 1.69 1.57 3.11 6.51 1.60 0.57 1.97

186 186 0.149 15.12 0.149 0.149 0.149 15.12 0.149

*Shortest distance between two holes.

by Van Decker et al.

properties with one from Kutscher' s plates, good agreement was observed with the data of Kutscher.

2. REVIEW OF STUDY OF VAN DECKER ET AL.

2.1 Apparatus and Method A description of the apparatus and method used by Van Decker et al. is given in their paper (Van Decker et al., 1996).We give a only brief review of it here. Aside from the properties of air, which are more or less constant, the heat exchange effectiveness in the asymptotic range depends on five plate parameters: minimum distance between holes (called the pitch), hole diameter, plate thickness, and thermal conductivity, and two velocities: the suction velocity and wind velocity. The geometric properties of the nine 60cm by 60 cm plates they tested are listed in Table 1. The properties of Plate 1 were made identical to a plate in Kutscher' s study, in order to permit a direct comparison. The remaining plates had holes with the square geometry, both to provide experimental data on a layout other than the triangular hole layout and so that the CFD model of Arulanandam (1995) could be used in the interpretation. The asymptotic performance of each plate was measured on a test rig (Golneshan and Hollands, 1998) Primary components included: a solar simulator (or short wave radiant heat source), an air suction system, and a wind tunnel. Each plate was installed on a suction plenum, which divided the plate into 7 sections of equal area. By limiting the measurements to the downstream plenums, the asymptotic conditions were enforced. Temperatures of the plate, upstream air and outlet air were measured with thermocouples. Each plate (except Plate 2) was tested over

2.3 Model Development Preliminaries Van Decker et al. developed a model for their data, as follows. They assumed that the plate is isothermal, with a single temperature Tp. This was supported by direct measuremnts and by the fact that plate thermal conductivity was not found to be an important parameter. The air-heating by the plate takes place in three regions: the front-of-plate, the hole, and the back-of-plate. Each region was assigned an effectiveness (denoted ep eh, andeb, respectively), as follows:

to, -r Ef - T - T =

ro~ -fro, , ~h ~

ro -ro~ , % -= ~

T-Trol

L-To2

(2)

where (referring to Figure 1) To1 is the bulk mean temperature of the air as it enters the hole and To2 is the bulk mean temperature of the air as it exits the hole. They also defined a combined effectiveness ejh for the front of the plate plus the hole as

%-

to2 - r . T,,- r

"

(3)

the same set of wind and suction velocities: for the wind velocity this set ranged from 0.0 m/s (no wind) to 5.0 m/s; and for the suction velocity the set ranged from 0.028 m/s to 0.083 m/s. The error limits of e were estimated to be _+0.019.

2.2 Results The observed values of effectiveness e ranged from 0.32 to 0.91. The effectiveness was found to decrease with increasing Vs, P, and D, and to increase with increasing Uw and with t. Thermal conductivity, k, was found to have a very weak effect. For the plate (Plate 1) with the common

Figure 1. Sketch of the plate, showing the three temperatures, To1, I"o2,and To~.


geometry, as will be explained below. From Eqns (1), (2), and (3) it was easy to show that

% = 1 - (1 - ~ ) / ( 1

and

-c#).

e = 1 - (1 - ef) (1 - e h) (1 - %)

(4)

(5)

2.4 Back-of-Plate Model Arulanandam (1995) (see also Aralandam et al. (1999))had used a computational fluid dynamics (CFD) code to model the flow and heat transfer on the front face and in the hole for the no-wind situation, and correlated the results by an single equation for ejh. For the same conditions, the resulting model was found to predict values efh that are consistently less than the corresponding measured c, this was expected since the model did not include the additional temperature rise (To - I"2) associated with the back-of-plate heat transfer. The difference between the measured e and the computed elh was roughly constant, at about 0.18. For each of their no-wind data-point, Van Decker et al. determined the e~ as predicted by Arulanandam, and substituted this and their measured e into Eqn (5) to get a corresponding value for eb. These results were then correlated by a model for the back-of-plate heat transfer giving the following equation for eb

% - ~

1 + dRe~

'

(6)

where Reynolds number Reb is equal tO VhP/V where Vh is the velocity in the hole, (=Vs4p2/nrD2), and coefficients d and e being found to be equal to 0.144 and 1/3, respectively. The authors then assumed, for the purpose of model development, that eb is independent of U,. This meant that Eqn. (6) could be used for eb for all wind speeds. 2.5 Front-of-Plate Model For a plate perforated by long slots rather than circular holes, Golneshan (1994) has described a 2D momentumintegral analysis that predicts the heat transfer on the front of the plate under conditions of significant wind. He found that eI should depend upon only parameter, y, as defined by:

y =Re)lRe = V)P I Uwv ,

(7)

where the Reynolds numbers Rew and Res are given by Re,, =U,cP/v. and Res=V~P/v, respectively. In particular, he obtained eI = 1 - ( a + by-it2) -1 ,

(8)

where a and b are constants. Van Decker et al. adapted this model to the plate with circular holes, adjusting the values of the coefficients a and b, to values appropriate to this

2.6 Hole Model Since the hole Reynolds number Reh = VhD / v is found to be much less than 2000, Van Decker et al. assumed laminar flow in the hole. The flow in the hole is (hydrodynamically and thermally) developing flow in an isothermal circular tube, but with the fluid (air) entering the hole being non-isothermal: the temperature nearer the plate is hotter than that nearer the centre-line. The ample literature information on the heat transfer when the entering fluid is isothermal guided their model development. They assumed a linear fit for the Nusselt number as a function of the Graetz Gr, obtaining ~h = 1 - e x p { - 4 ( c ( P / D ) + 3 . 6 6 ( D ) Rehlpr-1)},(9)

in which c was a constant to be determined. 2.7 Correlation With appropriate values of a, b and c, Eqns (4), (9)constituted a model for e, which, was fit to all the "windy data" (i.e., all the data points for which the wind speed was not zero), the constants a, b and c being adjusted so that the root mean squared difference between the data and the model was minimized. The result gave a=0.8434, b--0.4867, and c=0.00665. This model was found to fit the windy data with an rms error of 5.2%. 3. N E W

MODEL

For the conditions of no-wind or low values of wind speed U, the above model of Van Decker et al. gives unreasonable, negative answers. To overcome this problem, Van Decker et al. suggested switching to the Aralanandam equations for the no-wind conditions, but such a strategy could not in itself make clear the range of U, where one uses one equation and the range where one would use the other. Also it is much more convenient to have just one equation to cover the full range of wind speeds of interest (as does the recommended equation of Kutscher). So we undertook to develop such an equation. 3.1 Model Development For the conditions of no wind, y is equal to infinity, and Eqn (8) gives negative values of %, which leads to unrealistic negative values for e.. Indeed, negative values for e at zero wind speed will always be predicted by the model whenever a is made to be less than unity. When a is put equal to unity, Eqn (8) yields ei= 0 for zero wind. In fitting coefficients a, b, and c to their data, Van Decker et al. had noted that forcing a =1 gives almost as small an rms error in the model as letting a be free to take on any value. On the other hand, letting a take on increasingly large values greater than unity gives greater and greater error. In subsequent developments, we have found that building a comprehensive model on one that predicts negative e under no-wind circumstances was unwise. So we fixed a at unity and proceeded with the no-wind model development from there. With a equal to unity Eqn. (8) reduces to

25


26

r = (1 + eylt2) -1 ,

(10)

where e = lib. It was felt that (with the possible exception of their respective parameters settings, which might need slight adjustments) the Van Decker et al. models for the hole and back-of-plate heat transfer are quite satisfactory. Thus the model that needs revision is the model for the front of plate heat transfer: that is, Eqn (10) for eI. Using the Van Decker et al. models for eh and eb, it is possible calculate what eI would have to be to get the values of e that was actually measured at no wind conditions, and we did this for every no wind data point. Then from dimensional arguments, we concluded that this eI should be mainly dependent on the suction velocity Reynolds number, Res=V~/v. So we correlated this eI against Re,. A strong correlation was found, and moreover it was one that could be fitted by an equation very similar to Eqn. (10): ey=(1 +fRe,) -1

(11)

with Res replacing yln, and a new coefficient f which was fitted to the no-wind data, giving f = 0.0654. When it is recalled that y = Re,2/Re, we see that we can express Eqn (10) as (~f =(1 + eRe,Rew-lt2) -1

Effectiveness Measured by Van Decker

Figure 2. Hot of both windy and no-wind data of Van Decker et al (1996) in the form of measured effectiveness vs. that predicted by the model of the present paper, namely Eq. (14). The Legend refers to the Plates listed in Table 1. values of 6. The result gave e=1.42, b=O.0400, c=0.00510, and d=0.294. So, in summary, the new model for effectiveness model for a UTC with a square hole geometry is given by:

(12) =1-

1-

max[Re -1/2 , 0.04]

1+1.412Re $

Eqn (12) applies to the windy situation and Eqn.(11) for the no-wind situation. We propose combining them as in the following equation

ex~_4(O.Oo51P + 3.66

D PrRehD el= 1 + eResmin[Re~lt2,b]

W

)HI_

1 ,] 1 +0.294Re;/~

(14)

(13)

where b is equal to fie, and the symbol "min[x,y]" means that one is supposed to take the minimum of x and y. Because of the nature ofthe min[x,y] function, one finds that this equation automatically shifts to Eqn. (11) under no wind (Rew = 0) conditions, while Eqn (12) is found to apply whenever Rew is greater than b 2, which was found to be about 700, and since almost all of the data points under windy conditions had Re, greater than 700, the model reduces to the V an Decker et al. (1996) model for the windy cases of their experiments. 3.2 Fitting the Model to the Square-pitched Data The entire model for e , as given by Eqns (5), (6), (9), and (13), was fitted to the full data-set containing measurements on plates with holes on a square layout, both windy and not, to obtain the values of the parameters e, b, c, and d that minimized the sum of the square of the deviations between the model for e and the measured

This model was found to fit all the data with a root mean square error of 4.3%. Figure 2 compares the data to the model, by plotting the prediction on one axis and the measured on the other. 3.3 Testing the Model on Triangular-pitched Plates Van Decker et al' s Hate 1 had the same values of t, D, k, and P (P being the shortest distance between two holes) as Plate 2, but the holes were laid out in a triangular arrangement, rather than square. For a given Vs and U~, e for Plate 1 was found to be about 0.05 greater than for Plate 2. That is, other things being equal, the triangular layout performs slightly better. Van Decker et al (1996) found that the same model could be used for both plates if one uses an appropriatelyadjusted value for the pitch. That is a triangular-plate model (like that developed by Kutscher) can be used for a square pitched plate if one uses for the pitch P a value that is ~" times the square layout pitch, where scaling factor ~ is equal to 1.6. Conversely, a square-plate model (like Eqn


placements It captures the effect of a range of variables: suction velocity Vs, wind velocity U~, hole pitch P, hole diameter D, and plate thickness t, having been tested over the following ranges of these variables: 0.028 m/s _< Vs -< 0.083 m/s; 0 m/s _< U| _O

',/

vs

the planes

(3) of

(4)

~T

at z = O or z = P/2,-~z = O, v = w = O -.~z = O, u > O (5)

For the front surface of the plate, we have t/'

,o,o,o, I/'i ,,,11

at x = t and y2 + z 2 > D 2 / 4 ,

,,,ll,"

/// i'N ,'/~ii'___

f /

k 0_~

V = (O,O,O), a G = - ~ ~ lt_

which air is sucked through the plate per unit plate area. At distance x**from the plate this air stream is assumed to be in uniform motion. 2.2 Governing Equations and Boundary Conditions The relevant governing equations for the air velocity V(u,v,w) and the temperature T are the equations for the conservation of mass, momentum, and energy, assuming laminar flow, steady conditions, constant properties, no external forces and no viscous heat dissipation. These are given in textbooks (e.g., Arpaei and Larsen, 1984) and will not be repeated here. For the region containing the solid, (i.e., the plate), the governing equation is simply the steady heat diffusion equation. The inlet boundary condition is the plane x = x., Ideally, this plane would be located an infinite distance from the plate, but any CFD representation has to place it at a finite distance of magnitude sufficiently large to realistically represent the infinite distance. The following boundary conditions were assumed to apply there:

P=Poo

(1)

where T** and P,. are the ambient air temperature and pressure, respectively. We modelled the back of the plate as adiabatic, so that ~T = 0 at x = x o, "~x

for

y2 + z 2 > D 2 / 4

(6)

where G is the solar irradiance, o~is the solar absorptivity of the plate, ks and k are the thermal conductivity of the plate and air, respectively, and h, is the radiative heat transfer from the plate to the radiant surrounds, which are assumed to be at the ambient air temperature, T.. At the interface inside the hole, there must be a balance of heat fluxes, leading to the boundary condition:

Fig.1 (a): Sketch of the representative elements of the absorber plate and (b) definition of the computational domain

V=(Vs,O,O ) T=T**

Ox lt+ + hr (T - T**)

~y

X

atx=x**,

_kb_~

(2)

at O < x < t and y2 + z2 = D 2 / 4 ,

v (o,o,o),o~-k,

a,- I(~,/~)+

(~/~)_

where r 2 = y2 + z 2" 2.3 Dimensional Analysis The governing equations and boundary conditions were transformed into non-dimensional equations by introducing certain dimensionless variables (Aralanandam et al., 1995). The following dimensionless groups arose: The plate porosity r = ~292/4P 2 , a Reynolds number defined by Re D = VhD[V

where

Vh = V~/cr

(7)

a non-dimensional approach distance x** = x**/D and a non-dimensional plate thickness t* = t / D , the "plate admittance" Ad, and radiative Nusselt number Nu, defined by Ad = k ~ t / ~

and

Nur = h,D/k

(8)

respectively and the Prandtl number of air, which is fixed at about 0. 7. The heat exchange effectiveness e is defined

by (r0 - r.)/(r, - ~.) where ~0 i~ tho b u ~ outlet temperature at x = 0 and Tp is the average plate temperature. When it is expressed in terms of dimensionless quantities, there results e=e

9

eD, G , t , A d , N u r

)

(9)


The heat transfer can also be expressed in terms of a Nusselt number, defined by Nu = ReD Pr ln(1 - 6) t7

(10)

From Equations (9) and (10) it follows that Nu=Nu(ReD, tr, t* ,Ad, Nur)

3.

COMPUTATIONAL MODEL

FLUID

(11)

MECHANICS

The governing equations were solved with the appropriate boundary conditions using TASCflow, a finite volume based CFD code. Using TASCflow, the computational domain shown in Figure 1(b) was divided into a finite set of control volumes. The TASCflow solver was used to solve the algebraic equations that result from integrating the governing equations over each control volume. The control volume formulation method is fully conservative, with the formulation guaranteeing conservation of mass, momentum, and energy over each control volume.

Fig. 2:

Plot of the grid lines used, at a constant x-plane

3.1 GridDesign The domain was broken down into a set of control volumes, with a node at the centre of each volume. The total number of nodes N in the resulting grid is limited by constraints on the available computer memory and also by CPU time. In breaking down the domain, a rectilinear grid with uniformly spaced nodes was tried, but there were several problems with this approach, the most significant being the large number of irregular control volumes located at the edge of the circular hole. Later a series of straight lines approximating circular arcs centred along the x-axis was used to create the grid in the region within D/2 of the centreline. A different series of straight lines was used to approximate the square cross-section of the domain at the edges of the domain. This choice of grid lines improved the grid in the vicinity of the hole, but it led to having several nodes co-located at the origin. So an additional grid was created for the region around the x-axis, and was then attached to the main grid. Thus, the final grid actually consisted of two grids: the main grid and the sub-

grid. A view of the final grid design is given in Figure 2. Preliminary investigations indicated that convergence was improved by increasing the number of nodes near the solid and inside the hole in order to fully capture the characteristics of the flow as it approaches the plate and enters the hole. Therefore, the nodes were distributed using an expansion ratio R, where R is the ratio (in a given direction) of the length of the last flux element to that of the first flux element. The expansion ratio used in the grid construction near the walls was 25, resulting in a higher density of grid points in regions near the walls, where they were most needed. 3.2 Grid Refinement and Other Preliminary Studies The approximations introduced in the discretization process become more precise as the gird is refined, i.e. as the number of nodal points N is increased. Grid refinement studies are used to study the sensitivity of the numerical solution size of the grid and then decide what value of N will be used for the bulk of the simulations. A grid refinement study is usually carried out by systematically increasing the number of nodes and comparing the results --in terms of global parameters, such as the effectiveness. For a three-dimensional problem, the number of flux elements is successively doubled in all of the three directions. Thus, if the first grid has NI nodes in total, the second will have 8N1 and the third 64N1 and so on. The solution as N approaches infinity can be extrapolated from the finite N grid solutions by using an extrapolation method such as Repeated Richardson (RR) Extrapolation (Zwillinger 1992). Comparing the global parameters from each finite grid with the extrapolated results can then be used to estimate the error induced by having a finite value of N. In the present study, the grid refinement study was completed using the following combination of parameters: ReD = 1375.5, tr = 0.0005, t* = 0.67, Ad = 4.93, and Nut = 0.26. This combination of parameters, having a high suction velocity, low conductivity and low porosity was the most difficult computationally, because it has high velocity gradients inside the hole and a large variation in temperature across the plate. Ideally the x***plane is located an infinite distance in front of the plate, but in practice a suitably large value must be used. Several runs were made in which x** was reduced in steps of about 10 from 168 to 42. Within round-off error, the same results were obtained at each setting. So, at x*** = 42, the inlet boundary face is far enough away from the plate to not affect the results. Thus, a value of x* > 42 was used for the grid refinement study and all subsequent simulation runs. An initial coarse grid, Grid A, was first constructed using a 6 x 12 x 12 grid with 6 flux elements in the xdirection, 12 flux elements in the y-direction, and 12 flux elements in the z-direction. (A flux element is a hexahedron defined by eight nodes, one at each vertex.) The number of flux elements in Grid A was doubled in the three directions to produce Grid B, a 12 x 12 x 24 grid. Similarly, Grid B was doubled to produce Grid D, a 24 x 48 x 48 grid. Computer memory limitations eliminated the possibility of running a fourth gird by

31


32

doubling Grid D. Instead, a fourth grid, between Grid B and Grid D, was constructed by using a refinement factor of 1.5; thus Grid C was an 18 x 36 x 36 grid. The effectiveness was calculated for the four different grids and plotted against the inverse of N. The Repeated Richardson (RR) Extrapolation method was used to estimate the value of the effectiveness and the efficiency as the number of nodes approached efficiency, and the results are presented in Table 1.

Grid Size Grid A (6 x 12 x 12) Grid a (12 x 24 x 24) Grid C (18 x 36 x 36) Grid D (24 x 48 x 48) RR Extrapolation

"Global Parameters E T1 0.1745 81.98 % ' 0.1473 66.78 % 0.1365 66.60 % 0.1362 66.45 % 0.1362 66.45 %

Table 1- Grid refinement study results

captured by the variables in Equation (9). 3.4 Comparison to Similar Studies Kutscher (1992) completed numerical studies for a transpired plate absorber with a hexagonal pitch under nowind conditions. While not identical to the square-pitch configuration investigated in the current study, the configurations were similar enough to merit a detailed comparison of the results. Kutscher's work covered high porosity plates and include heat transfer from the back of the plate, but in the current numerical model the back surface of the plate has been modelled as an adiabatic surface. This significant difference in the models was easily accounted for in the comparison of Kutscher's numerical results, since he broke down the total heat transfer into three sequential heat exchange processes: that occurring along the front surface of the plate, inside the hole, and along the back surface of the plate. To compare the present results, Kutscher's results for the front and hole regions were combined. The results of the two codes were found to agree to within about 1%.

Reo = 1375.5, G = 0.005, t* = 0 . 6 7 , A d = 4.93,Nu~= 0.26

4. Based on the results of this study, and considering the amount of CPU time available, Grid C was chosen as a suitable grid with acceptably small error and convergence time. Figure 3 gives a plot of the velocity vectors obtained using Grid C.

SIMULATION RESULTS

4.1 R a n g e o f P a r a m e t e r s The important ranges for the non-dimensional parameters were chosen based on a number of factors, including manufacturing constraints and solver limitations. Thus, the following ranges were specified for the dimensionless parameters: 150 50~ whereas the reference collector exhibited a maximum temperature < 40~ These results are also reflected in the values for the collector efficiencies, cf. Figs. (2) and (3). The average daily efficiencies for the reference and coaxial tubular

900

60 x o

X

50 x

e~

E

X

X

0

0

0

800 700

@ m

L_

~- 30

X

o

o~ 4 0

I-

KI = 0.148 W m l K "1 ~a = 0.090 W m l K "1 r~t = 0.024 Wm-lK"1 Kf = 0.63 W m l K "1

Cw= 4190 Jkg-lK"1 Mf = 10.9, 16.9 kg.h"1 p ~ = 1.165 kgm"3 pw = 958 kgm-3 Pa- 1100 kgm"3 ~:m - 0.027 Wm]K 1

('171~)1-" 0 . 0 4

O

D

o

m

D

m m

|

I

B m X Global radiatio

8

I

I

9

10

11

I

Houm 12

"O

- 400

-200

mTi (ref) DTo (ref) I

5 0 0 .__.

-300

OTo

10

'O

.,,,.

m m

eTi

20

600

a m

m .D O

- 100

I

I

13

14

15

Fig. 1 Inlet and outlet water temperatures for reference and coaxial tubular collectors and corresponding global radiation


45

data, May 26, 1999. The flow rates are 5.7 and 7.2 kgh-1 for the reference and coaxial tubular solar collectors, respectively.

50

12 -10

40

-

= 30

_0 ..... J

L_

i1) l:k

0-

_-0 .....

9 ...... n

E 20 I--.

=

10 .........................

8

J_ . . . . . . . . . . . . . . . . . . . . . . . . . . .

9

t ........

10

1

11

1

Houl

12

I

13

8

Ti (ref) expt'/ I Ti (ref) mod~ TO (ref) expt 1 To (ref) mod t Efficiency (%~

i

14

,I

15

v

u t--

1

-- 6 u --

4

--

2

uJ

,,I

16

17

Fig. 2 Comparison of experimental and simulation model values for inlet and outlet water temperatures for the reference tubular collector together with average hourly efficiency (%), May 26, 1999. Flow rate = 5.7 kgh"1.

A similar comparative analysis for the coaxial tubular solar collector is presented in Fig. (3) together with corresponding average hourly efficiency values. It is apparent that there is very good agreement between the values calculated by the simulation model and the experimental data. This simulation model will be utilized to help determine the optimum wall thickness dimensions for such a coaxial tubular solar collector design.

Selectively coated pol~'neric absorbers The selectively coated polymeric absorber plates all exhibited relatively high efficiencies aoutlet temperature during performance testing. In fact, the heat exchange fluid flow rates had to be increased in order to prevent damage to the prototype selectively coated absorber plates due to thermal stress. The flow rate during performance testing was 13.2 kgh1 for all selectively coated absorber plates. Results, typical for such performance testing, are shown in Fig. (4) and (5) for the 10 mm thick polypropylene-2 and 15 mm thick polycarbonate double-walled absorber plates, respectively.


46

40

60

35

50 -

30 IJ

30

= r

-20

Q.

_---

E I- 20

qt

9

...... El

ql

10

r

O

'"

8

I

9

,

, I

10

I ........

I

11 H o u r 1 2

Ti expt'l

,

U m

- 1 5 ~ 111

Ti model To expt'l To model Efficiency

-10 -5

'

i

,

i

13

14

15

16

O

17

Fig. 3 Comparison of experimental and simulation model values for inlet and outlet water temperatures for the coaxial tubular collector together with average hourly efficiency (%), May 26, 1999. Flow rate = 7.2 kgh 1.

In both cases the observed average daily absorber plate efficiencies were relatively high, viz., 58.6 and 67.4% for polypropylene and polycarbonate, respectively. The maximum outlet water temperatures were in the range of 60~ and the average daily temperature gradient between the inlet and outlet streams were ~ 14~ in both cases. The average daily efficiencies, based upon a series of tests performed during April and May 1999, on the selectively coated

absorber plates were as follows: 1] (polycarbonate-1) = 58.6%, 1] (polycarbonate-2) = 49.3%, I] (polycarbonate) = 58.6%, rl (polypropylene-1) = 64.3% and (polypropylene-2) = 60.3% (of., Table 2 for details on the absorber plates). The maximum outlet water temperatures and average daily temperature gradients between the inlet and outlet were similar to those reported in Figs. (4) and (5).

70

1100

v

-

>" t,J 60

1000

900 800 700 - 600 - 500 - 400 -300

t,"

~9 50

-

..,,,.

"' 4 0

30 ~- 20 ~-9

EIO i-0

To

10

11

12

13

t~

-~ .o O

(%) -

9

._~

-200

Efficiency

8

r

o

o.,.., ,i-,a

14

15

100

16

Houl

Fig.4 Inlet and outlet water temperatures and average hourly efficiency (%) for polypropylene -2 solar collector and corresponding global radiation data, May 13, 1999. Flow rate = 13.2 kgh q.


47

80

1100

>. 7 0

1 000 900

v

t,j t--

u

60

"I

El

800 - 700

4o = 30

r

o

...,.

- 600

"0

- 500

._~

er

- 400

L_

~

20

-300

.,-!. To

E ~10 I-0

"~ 0

200

Efficiency (%)

100 8

9

10

11

12

13

14

15

16

Hou=

Fig. 5 Inlet and outlet water temperatures and average hourly efficiency (%) for polycarbonate solar collector and corresponding global radiation data, May 19, 1999. Flow rate = 13.2 kgh1.

The results form the performance testing of the selectively coated absorber plates, to date, are very encouraging with respect to both efficiency and maximum outlet water temperature. The major design problem is to find a polymeric material that can be (a.) fabricated as a double-walled sheet, (b.) serve as a substrate for the selectively coated surface, and (c.) is stable under the thermal stresses to which a solar collector is subjected. Obviously, this is not a simple task; especially since the economics of the thermal energy conversion process is of utmost importance. We are also in the process of developing a simulation model for the selectively coated absorber plate solar collectors in order to optimize the design with regard to the double-walled sheet dimensions. 6. CONCLUSIONS The design and performance testing of two different solar collector design prototypes, fabricated from polymeric materials, has been presented. Such solar collectors have a number of inherent advantages vis-avis those fabricated from metals. These include the following: (a.) fabrication by techniques, such as extrusion and molding, which can, a priori, result in significant reduction in production costs; (b.) eliminate corrosion problems, especially with regard to their application in sea water desalination systems; (c.) reduced weight per unit area of solar collector which results in significant reductions in both shipping and labor/installation costs. These advantages are offset significantly by their inherently low thermal conductivity coefficients.

We have attempted to overcome this disadvantage by two very different approaches in the design. 1. Coaxial tubular solar collector: design concept that entails the use of an inner black tube as a solar absorber in intimate contact with an outer transparent tube as an insulator. In theory, the inner black tube should be of minimum wall thickness to compensate for its low conductivity and the outer transparent tube should have a wall thickness that optimizes the trade-off between the material insulation and its transmittance of the solar radiation. 2. Selectively coated polymeric absorbers fabricated from double-walled polymer sheets which function as the solar collector absorber plate. Their upper surface serves as the substrate for the selective coating and the heat exchange flows through the channels between the two walls. The coaxial tubular solar collector exhibited a significant performance enhancement relative to the reference prototype, a simple black tubular collector. The side-by-side performance testing of the tubular collectors gave an average daily efficiency of 27.7 and 9.6% for the coaxial and reference solar collectors, respectively. In addition, the coaxial collector achieved maximum outlet water temperatures > 50~ whereas the reference collector achieved a maximum outlet water temperatures < 40~ Though significant performance enhancement has been achieved, the absolute performance values, outlet temperatures and efficiency, are still relatively low and thereby limit such a design to applications such as the heating of swimming pools. We intend to continue this study to determine both the optimum and practical dimensions,


48

i.e.,wall thicknesses, for such a coaxial tubular solar collectordesign. A simulation model, which utilizes the measured average hourly solar radiation and ambient temperature values together with the initialwater temperature and calculates the inlet and outlet water temperatures at the end of each hour has been developed. It has been validated by comparing calculated to the measured inlet and outlet water temperature values. This simulation model will be utilizedto help determine the optimum wall thickness dimensions for such a coaxial tubular solar collectordesign The selectively coated polymeric absorber plates all exhibited relatively high efficiencies and outlet temperature during performance testing. The average daily absorber plate efficicncies for all those tested were relativelyhigh, viz.,in the range of 50 - 60. The m a x i m u m outlet water temperatures in the range of 60~ and the average daily temperature gradient between the inletand outlet water streams ~ 14~ were measured for all selectively coated absorber plates tested. The results from the performance testing of the selectively coated absorber plates, to date, are very encouraging with respect to both efficiency and maximum outlet water temperature. The major design problem being to find a polymeric material that can be (a.) fabricated as a double-walled sheet, (b.) serve as a substrate for the selectively coated surface, and (c.) is stable under the thermal stresses to which a solar collector is subjected. Obviously, this is not a simple task; especially since the economics of the thermal energy conversion process is of utmost importance. We are also in the process of developing a simulation model for the selectively coated absorber plate solar collectors in order to optimize the design with regard to the double-walled sheet dimensions.

Subscripts transparent tube 1 a absorber tube ambient amb bulk b c collector cax coaxial tubular collector conduction cond convection conv f fluid i inside/initial/inlet outside/output/outlet 0 ref refetubular collector

sky

sky

st

storage tank water

w

NOMENCLATURE A c

D G g h L M m N Nu Pr Q Ra Re T t

U

area (m2) heat capacity (JkglK ~) tube diameter (m) solar radiation ( W m "2) gravitationalconstant (ms 2) heat transfer c o e f f i c i e n t (Wm2K "1) length (m) mass flow rate (kgs "1) mass (kg) number of risers Nusselt number Prandtl number energy (W) Rayleigh number Reynolds number temperature (K) time (s) overall heat transfer coefficient (Wm2K 1)

Greek O~

B 8 s K

0

g P 0

solar altitude angle slope hour angle thickness (m) emissivity thermal conductivity (WmqK "1) latitude incident angle dynamic viscosity (kgm-ls"1) density (kgm"3) Stefan-Boltzmann constant (Wm2K -4) transmittance-absorptance product


Acknowledgment- This research was supported under Project No. GR.01463 E 1071, Joint German - Israel Research Program, Israel Ministry of Science Bundesministerium Rir Bildung Wissenschaft, Forschung und Technologie. REFERENCES Dickinson, W.C., Clark, A.F., Day, J.A. and Wouters, L.F. (1976) The shallow solar pond energy conversion system. Solar Energy 18, 3-10. Duffle, J.A. and Beckman, W.A. (1980) Solar Energy of Thermal Processes, 762 pp., Wiley Interscience, New York. Eckert, E.R.G. and Drake, R.M. (1972) Analysis of Heat and Mass Transfer, 675 pp., McGraw-Hill Book Co, New York. Garg, H.P., Chakravertty, S., Shukla, A.R., Agnihotri, R.C. and Indrajit (1983) Advanced tubular solar energy collector: A state of the art. Energy Convers. & Mgmt 23, 157-169. Gerich, J. (1977) An inflated cylindrical concentrator for producing industrial processing heat. J.Proc. FRDA Conf. Concentrating Solar Collectors, Atlanta, Georgia 2, 103-115. Grimmer, D.P. and Moore, S.W. (1975) Practical aspects of solar heating: A review of materials used in solar heating applications. Los Alamos Scientific Laboratory, University of California, LA-UR-1752, Los Alamos, New Mexico. Gopffarth, W.H., Davison, R.R., Harris, W.B. and Baird, M.J. (1968) Performance correlation of horizontal plastic solar water heaters. Solar Energy 12, 183-196. Harris, W.B., Davison, R.R. and Hood, D.W. (1965) An experimental solar water heater. Solar Energy 9, 193-196. Hedderich, C.P. (1982) Design and optimization of aircooled heat exchangers, ASME J. Heat Transfer 104, 683-690. Khanna, M.L. (1973) A potable-type solar water heater. Solar Energy 15, 269-272. Kudish, A.I. and Wolf, D. (1978) A compact shallow solar pond hot water heater. Solar Energy 21,317-322. Kudish, A. (1981) Sede Boqer shallow pond project. Energy 6, 277-292. Kuehn, T.H. and Goldstein, R.J. (1976) Correlating equations for natural convection heat transfer between horizontal circular cylinders, Int. J. Heat Mass Transfer 19, 1127-1134. Kuehn, T.H. and Goldstein, R.J. (1978) An experimental study of natural convection heat transfer in concentric and eccentric horizontal cylindrical annuli, ASME J. Heat Transfer 100, 635-640. Meinel, A.B. and Meinel, M.P. (1976) Applied Solar Energy, p. 13, Addison-Wesley, Reading, MA. (3zisik, M.N (1977) Basic Heat Transfer, 572 pp., McGraw-Hill Book Co, New York.

49

Rabl, A. (1985) Active Solar Collectors and Their Applications, 503 pp., Oxford University Press, New York. Rommel, M., K6hl, M., Graf, W., Wellens, C., Brucker, F., Lustig, K. and Bahr, P. (1997) Corrosionfree collectors with selectively coated plastic absorbers. Desalination 109, 149-155. Swinbank, W.C. (1963) Long-wave radiation for clear skies. J. Roy. Meteoro. Soc. 89. Tabor, H. and Zeimer, H. (1962) Low cost focusing collector for solar power units. Solar Energy 6, 55-59. Tsilingiris, P.T. (1997) Design, analysis and performance of low-cost plastic film large solar water heating systems. Solar Energy 60, 245-256. Whillier, A. (1963) Plastic covers for solar collectors. Solar Energy 7, 148-151.


50

STUDY OF A MIXED (WATER OR AIR) SOLAR COLLECTOR Sylvain LALOT M.E.T.I.E.R., E.I.P.C., Campus de la Malassise BP 39, 62967 Longuenesse Cedex, FRANCE, Telephone : +33 3 21 38 85 10, Fax : +33 3 21 38 85 05, E-mail : [email protected]

Abstract

-

paper presents a new concept of solar collectors. Usually, solar collectors are designed as water collectors or as air collectors. Here is presented a collector that can be used as a water collector as well as an air collector. First, the governing equations are given for steady state, explicitly including the ratio of the actual convective surface to the collector surface. Then it is shown that the product of the actual convection coefficient with the surface ratio plays a main role in the collector efficiency. Then it is shown that is possible to compensate the low value of the convection coefficient for air by a high convective surface. This can be achieved using standard heat exchangers components : finned tubes blocks. The idea is then to use the air side to build an air collector and the tube side to build a water collector. The feasibility of such a collector is shown and the performances of the proposed collector are given trough the efficiency curves. It is shown that the mixed collector has performances comparable to those of standard collectors; the performances of the air collector being higher than the performances of the water collector. Improvements are proposed such as the use of selective coatings and technical cover. Finally the dynamic behavior of the collector is briefly studied. It allows the determination of the thermal capacity of the collector. This

1. INTRODUCTION The scientific study of solar collectors has begun many years ago (Desautel, 1978) and many geometries has already been proposed : flat plate non-concentrating collectors, parabolic collectors... The fluid which is used may be water, air, oil, molten salt . . . . To increase the efficiency of the collectors, progress has been made on the coating of the absorber, on the geometry of the absorber (mainly for air collectors) , on the quality of the cover that acts as an infra-red barrier. Many devices has been introduced in the manufacturing of the collector to decrease thermal losses : anti-convection cells, transparent insulation (Aronov and Zvirin, 1997) . . . . To characterize accurately a collector and to manage efficiently a whole installation, the resource itself and its measurement are widely studied (Perrin de Brichambaut and Lamboley, 1974) : direct solar radiation and indirect solar radiation. In the last few years some new concepts have been developed. Hybrid collectors have been proposed : one part of the radiation is transformed by photovoltaic cells, another part is absorbed by a cooling medium (air for instance). But in the large majority of the cases, the solar collector is designed to heat one fluid. Here is presented a collector that can be used as a water collector as well as an air collector. In this case, the energy is used either by water or by air. It could be used by both at the same time, but this has not been already tested. First, the governing equations are presented for steady state, introducing explicitly the ratio of the actual convective surface to the eaptation surface. Then a description of the proposed collector is derived from the equation giving the efficiency of a

collector. Then, it is shown that the mixed collector has a good efficiency for water as well as for air. Then the governing equations of transient states are given. Experimental results allow the computation of the thermal capacity of the collector. Finally, the determination of the time response of the collector is achieved for both fluids, air and water. It should be noticed that the present work has been carried out for a company some times ago and that confidentiality has delayed the publication of the results

2. GOVERNING EQUATIONS In general, a water solar collector is built from "tubes" where the fluid flows, and from solid parts that receive the solar energy and transmit it to the fluid. For air solar collectors, it has been found that it is interesting to increase the convective surface. Figure 1 shows a schematic of a typical absorber. In this representation, the cover is characterized by its ability 2" to transmit the energy; the thermal losses are represented by a conductance g . In a first analysis, it is possible to group the characteristics of the cover and of the absorber in a single characteristic/70 = T E

E.

This study focuses on collectors in which the fluid has only one pass in the absorber. In this case the energy balance may be written as follows : E

_K

(1)


51

Fig. 1. Schematic of a solar collector and"

r: (x) = r:, +

dr: :x) = kt Aa (to : x ) - r: :x))

(2)

Ac

K,

(or+ K, )r~ c z L

(4) Introducing the reduced mass flow rate n ~ - - ~

r&

, the

wL global convection coefficient ~ = k t

loss conductance

K,

K -~, wL

Aa

Ac

, and the reduced

Equations (1) and (2) lead to

the governing differential equation of the evolution of the fluid temperature 9

dr:(x) ~o I~ - K , ( r : ( x ) - r . ) -

a (a+K,)~

c:

It is then easy to find the evolution of the fluid temperature"

From Equation (4) it is possible to calculate the efficiency of a collector. The efficiency is defined as the ratio of the actual energy taken by the fluid n~c: (T/o - T/,)to the normal incident radiation A c I s . It may be written as follows"

2 r ~ c / thF _ct_K~_ K, L2 (a + K, )n~ :

](

rio-

gr (Tf o + Tf i )/ 2 - Too"~ IS

(5) For given operating conditions, the efficiency of the collector depends on the quality of the heat exchange between the absorber and the fluid. This is shown in figure 2"


52

_

0.90.80.70.60.50.4-

0.5

0.3-

ot

0.2-

oJ + K r

0.7 A

0.9

..-o--1

0.1I

0

Kr

I

0.5

1

I

I

1.5

2

Fig 2" Influence of the convection coefficient on the efficiency of a solar collector For a typical water solar collector, it is possible to suppose 9 let = 2 0 0 0 W / m 2 K , K r = 1 0 W / m 2 K , ~ Ac

to ~ a+K,.

= 0.99"

0.5" This leads

The absorber surface is large enough to

This can be achieved using a well known technology used in the manufacturing of heat exchangers : the use of finned tubes. Figure 3 shows a detail of an uncoated finned tubes block. The geometrical characteristics are : - fin pitch : 1.7 ram, tube pitch : 35 ram, - fin height: 35 mm. -

assure a good efficiency. But the convection coefficient between air and an absorber is much lower, it can be considered that kt = 1 0 0 W / m 2 K is a large value. So, to get a good efficiency, the area ratio has to be at least 10.

Fig. 3. Detail of an uncoated finned tubes block

53


So, the area ratio is large"

-

A,, .

-

20.17

.

A~ The use of fins allows the use of both convective surfaces, as usual in a heat exchanger. So, this leads to the fact that the absorber could be used to heat air (along the fins) or water (in the tubes). To prove the feasibility of such an solar collector, a prototype has been manufactured. It has been built using standard finned tubes blocks (painted with a standard black paint, E = 0. 9), a polycarbonate cover ~" = 0. 83, and rock wool as bottom insulation. The fins are made of aluminum and the tubes are made of copper.

3.

PERFORMANCES COLLECTOR

OF

THE

From these results it is possible to calculate the characteristics of the water collector"

7"1o = O. 7 9 3 K r =8.55W/meK This shows that the efficiency of the water collector is close to the efficiency of standard collectors. The performances could be improved using a selective coating instead (Tal-tarlo I. and Zvirin Y., 1988) of standard black paint, and using a technical cover (Zvirin Y. and Avichai Y.,1989).

NEW It can be seen that the efficiency of the air collector is higher than the efficiency of the water collector in the first part of the efficiency curve. This can be explained by three facts. First, there is no need to take the efficiency of the fins into account :

The prototype has been tested at the French technical center "CETIAT" (CETIAT, 1984). For steady state, the results can summarized by the efficiency curves (Figure 4). For water, the variation of the mean temperature is obtained by the variation of the inlet temperature (using an external heater). So, equation (5) shows that the efficiency is a straight line. For air, as the inlet temperature is the ambient temperature, the variation of the mean temperature is obtained by the variation of the mass flow rate. In this case, to increase the mean temperature, one has to decrease the mass flow rate. This induces a decrease of the convection coefficient. Then equation (5) and figure 2 show that this leads to a decrease of the efficiency; then the efficiency curve is no longer a straight line.

/7o is higher. Secondly, the area ratio is very high and the first term i n equation (5) is higher for the air collector. Thirdly, the heat losses due to the convection over the absorber are null : the heat is used by the flowing air, so

K,

is reduced. Here again the performances could be

improved using a selective coating instead of standard black paint, and using a technical cover.

_

0.9

-

0.8

-

0.7

-

0.6

-

0.5

-

0.4

-

0.3 0.01

air .." w a t e r

r

00 s

i

002

1

(r o +

O Fig. 4. Efficiency curves of the mixed collector


54

4.

DYNAMIC THE NEW

CHARACTERISTICS

OF

COLLECTOR

The dynamic behavior of a solar collector is governed by the following equation :

rlo ls = K, (T/ - T= )+ a + K , Ln~ c/

+

~)x TCA + +

o~+K, cz

(6)

P.r c/V,

TCA n~ c: L ~ ~ T: ~

It can be seen that the time response of the water collector is about 5 minutes. As there is no analytical solution to equation (6), it is necessary to use a numerical method to deduce l~om the experimental data the value of the thermal capacity of the absorber. A standard finite-difference technique has been used, and it has been found that the value of the thermal capacity of the absorber is TCA=12 kJ/m2K. This value can be compared with the known values; the weight of aluminum in the absorber is about 5.2 kg/m2, the heat capacity of aluminum is about 880 J/kg K; so the thermal capacity of aluminum is about 4900 J/m2 K. The weight of copper in the absorber is about 12 kg/m2; the heat capacity of copper is about 400 J/kg K. So the thermal capacity of the copper tubes is about 4800 J/m2 K. This makes about 9700 J/m2 K for the absorber. We can deduce that the thermal capacity of the insulation is about 2300 J/m2 K.

+

a ~x~t p: c: V r TCA ~ 2 T/

Using the value of the thermal capacity of the absorber, it is possible to numerically calculate the time response of the air heater. It is found that the heater also needs 5 minutes to reach the stabilization.

~t 2

To determine the thermal capacity of the absorber, T C A , it is possible to use the response to a step of energy. This has been done for the water collector. The results given by the CETIAT are given in figure 5.

T:o (t ) - T: ,

_

0 0.9-

~

X X

.

T:o ( +Oo)- T:,

X

X

X

x

0.8-

0.70.60.50.40.30.2Time (minutes)

0.1I

0

1

I'

I

I

I

I

2

3

4

5

6

Fig. 5. Response of the water collector to step of irradiation


55

Greek symbols

5. C O N C L U S I O N S

O~

It has been shown that the fine study of the governing equation of a solar collector can lead to a new concept for a solar collector. Based on the geometry of finned tubes heat exchangers, the proposed collector is able to heat water or air and the experiments have shown that in both cases the collector has a good efficiency. It has also be shown that the time response is quite short and independent of the fluid used.

ACKNOWLEDGMENTS

global convection coefficient

= kt A a /A c

W/m2K

E T]

emissivity collector efficiency

dimensionless dimensionless

/7o

maximum efficiency =E~'E

dimensionless

pf

density of the fluid

kg/m 3

"/"

transmittance of the infrared barrier

dimensionless

REFERENCES

The author would like to thank the French subsidiary of GEA for its technical and financial support.

Desautel J. (1978). EDISUD, Paris

Les

capteurs h61iothermiques.

Aronov B. and Zvirin Y. (1997). Theoretical investigation of solar collectors with transparent insulation covers by a novel calculation algorithm. In Proceedings o f lSES World Congress, Taejon, Korea

NOMENCLATURE Aa

absorber convective area

m2

Ac

collector aperture area

me

cf

specific heat of the fluid

J / kg K

E

fm efficiency

dimensionless

Is

normal solar irradiance

W / me

CETIAT (1984). Proc6s-verbal d'essais n~

K

thermal losses conductance

W/ K

Kr

reduced loss conductance

Tal-tarlo I. and Zvirin Y. (1988). The effects of radiation properties of surfaces and coatings on the performance of solar collectors. In J. Solar Energy Eng., vol. 110, pp. 217-225

= K/(wL) L

length of the collector

n~

total mass flow rate

kg / s

n~

reduced mass flow rate

convection coefficient

= n~/(wL)

kg/s m2

t

time

s

Ta

absorber temperature

K

Tf i

inlet fluid temperature

K

Tfo

outlet fluid temperature

K

T**

ambient temperature

K

T C A thermal capacity of the absorber

Vr

reduced volume of fluid

J / K m2 3 m / me

w X

width of the collector abscissa

m m

per square meter

CETIAT (1984). Proc6s-verbal d' essais n~

W/m2 K W / me K m

kt

Perrin de Brichambaut C. and Lamboley G. (1974). Le rayonnement solaire au sol et ses mesures. Editions Europ6ennes Thermique et Industrie, Paris

Zvirin Y. and Avichai Y. (1989). Improving the efficiency of solar collectors by glass coatings. In Proc. ISES Solar World Congress, Kobe, Japan


56

UNCERTAINTY IN SOLAR COLLECTOR TESTING RESULTS Emmanouil Mathioulakis, Kostantinos Voropoulos and Vasilis Belessiotis Solar & other Energy Systems Laboratory, NCSR >, 15310 Ag. Paraskevi Attikis, Greece Tel. +301 6544592, Fax +301 6544592, E-mail: [email protected] Abstract - A systematic assessment of all experimental error is presented leading to the determination of the uncertainty in the solar collectors testing results. The use of specific statistical tools allows not only the evaluation of the reliability of the testing procedure itself, but also the quantification of the goodness of fit and the prediction of the uncertainty in the collector instantaneous efficiency.

1. INTRODUCTION The basic scope of solar collector testing is the determination of the collector efficiency by conducting measurements under specific conditions defined by international Standards. The experimental results of testing lead to determination of the parameters of a more or less complex model capable of satisfactorily describing the energy behavior of the collector. The equation derived is considered to express the specific collector and can subsequently be used to predict its output under any conditions. Although several more elaborated models and testing methods have been proposed by various authors (Perers, 1997), in the present study, the Standard ISO 9806-1 is examined, mainly due to its extensive use and international application (ISO, 1994). More specifically, it is assumed (Duffle and Beckman, 1991) that the behavior of the collector can be described by a 2 or 3-parameter single node, steady state model n=/( T~): n= no-U0 Ti* n=no-U1

Ti'- u2 G (T i• )2

(la) (lb)

The above equations 1a and lb as well as the whole analysis presented in this paper are also valid for reduced temperature difference T'm calculated with respect to the mean collector fluid temperature. In this case the variable ~, where it appears, must be replaced by T'm. During the experimental phase, the output, solar energy and the basic climatic quantities are measured. During analyzing the data, a least square fitting of the model equation is performed on the measured data, in order to the determine parameters no and U0 or no, U1 and U2. In practice, this procedure determines only the equation of the collector behavior without calculating the uncertainties in the determined parameters, and thus the suitability of the concerned model is not evaluated with statistical criteria. Despite the widespread use of testing and the great importance of the testing results, an objective and standardized method for the determination of uncertainty in test results is still lacking. The question of uncertainty is crucial if one wishes to investigate the efficiency of each model, to consider the experimental uncertainties and to determine the uncertainty in the model parameters. It is noted here that only a limited number of publications deal with the accuracy of test results of solar thermal devices, mainly in the context of test

procedures of solar water heaters (Burges et al. 1991a, 1991b). A corresponding analysis for solar collector testing methods has been proposed by Proctor (1984a, 1984b, 1984c). In this analysis only the uncertainties related to measuring device errors and the standard least square technique have been considered. However, as will be discussed later on, this approach is equivalent to assuming a good fit and prohibits an independent assessment of goodness-of-fit In this publication we develop the general rules of uncertainty analysis and their application in a typical case of a commercial collector tested according to ISO 9806-1. The test were conducted in the Solar & other Energy System Lab which operates under the EN45001 Quality Assurance System, with strict respect to the requirements of the testing standard. 2. UNCERTAINTIES EXPERIMENTAL DATA

ASSOCIATED

WITH

The terminology used in uncertainty calculation is often confused, leading to different interpretations. In area of testing performed according to commonly accepted standards, uncertainties in experimental data should be determined according to the recommendations of ISO VIM (1995), by taking into account Type A and Type B uncertainties. Type A uncertainties are those determined by statistical means while Type B uncertainties are determined by other means (ISO, 1995a; ISO, 1995b). The uncertainty which is associated with each measurement, is the accumulated result of the uncertainty of the measuring insmmaent (Type B uncertainty), the uncertainty which represents the deviation of the measured value during sampling of data (Type A uncertainty) and the uncertainty which derives from the fact that the measurement may not represent the true value of quantity. Generally, in cases where an attempt is made to describe the behavior of a certain system with an approximate model, a distinction on the following should be made: 9 On the one hand, the Type A and Type B uncertainties which characterize every measurement itself and which are related to the quality of the measuring instrument and the stability of the measurement. These uncertainties can be determined quantitatively. 9 On the other hand, the uncertainties which are related to the degree to which the measurement or the model is representative, and which characterize the quality of the methodology followed. These uncertainties cannot be


determined quantitatively and, after all, their determination has no meaning. Their influence is reflected on the ability of the methodology used (model and testing method) to describe the phenomenon. If, for example, it is proved that the certain experimental results are not represented satisfactorily by eqn (1 a) or (1 b), the whole methodology, or the suitability of the specific equation is in question. In our case, Type A uncertainties derive from the statistical analysis of the repeated measurements at each point of the steady-state operation of the collector. It should be brought in mind that, according to the Standard, N measurements are taken for 15 minutes (about 30 measurements), and the average value for each measured quantity is found. For every operation point of the collector, the best estimate of a quantity X is the arithmetic mean X of the N observations xj and its Type A uncertainty is the standard deviation of the mean (Fuller, 1987): N

(xj -~)~

57

provided by calibration certificates of sensors used for the measurements for this study, leads to the values of Table I. In most cases a measurand Y is determined indirectly from N other quantities X1, X2. . . . XN through a functional relationship Y=/(X1, X2.... XN). The standard uncertainty in the estimate y is given by the law of error propagation (ISO, 1995a; Fuller, 1987):

%=

dfl

2+2

~~u(xi,xj)l

~

(4)

where u(xi,xj) is the covariance associated to xi and xj. In our case, eqn (4) is used for the evaluation of combined uncertainty in the efficiency values n and of the reduced temperature difference Ti', which are calculated as a function of Tin, Tout, AT, Ta, m, G and Ar The calculation is conducted following the steps described in the flow chart of Figure 1.

0.5

(2)

j=l

N(N-~)

O-&x -

By nature, Type A uncertainties depend on the specific conditions of the test. Thus, they include the fluctuations in the measured quantities during the test which lie within the limits imposed by the Standard, and also the fluctuations in the testing conditions not considered by the model. Such fluctuations concern, for example, the air speed or the percentage of diffuse irradiance in global. Type B uncertainties derive from the calculation of uncertainties over the whole measurement, taking into account all available data, such as sensor uncertainty, data logger uncertainty etc. Although the Standard defines the upper limits of the accuracy of the measurements, the uncertainties that have to be taken account are the ones associated with the specific sensors used in the test. If there are more than one independent sources of uncertainty, (Type B or type A) ui, the final uncertainty is calculated according to the general law of uncertainties combination (Dietrich, 1991): (3) i

Figure 1: Propagation of uncertainties and synopsis of fitting procedure

Table I: Type B uncertainties in measurements The application of the above methodology for the calculation of Type B uncertainty based on the information

Figure 2 shows the expanded standard uncertainties ~n and ~T* of n and T: respectively, as calculated for a specific black-painted collector for each measurement point. The horizontal bars refer to ~T* and the vertical ones to an. In order to show the figure more clearly, only some of the 32 points (measured in the laboratory) are presented. The values of the uncertainties are presented in figure 2 as expanded uncertainty ax, as is the usual practice. The expanded uncertainty in an estimate x is obtained by multiplying the


58

combined standard uncertainty Ux by a coverage factor k=2, corresponding to a level of confidence of 95% (ISO, 1995b).

(6)

[Yi - y(xi;al, a2...a M)] 2 i=1

0.9

o.8

The problem with this approach is that, in reality, the typical deviation o is almost never constant and the same for all points, but that each data point (xi, Yt) has its own standard deviation 6i. Another very interesting alternative is the use of the weighted least square (g'ZS) method, which calculates, on the base of the measured values and their uncertainties, not only the model parameters but also their uncertainty. By this way a qualitative evaluation of fitting can be performed. In the case of WLS, the maximum likelihood estimate of the model parameters is obtained by minimising the chi-square function (Press et al., 1996):

,~

0.7

0.6 0.5 0.4

++

0.3 0.2

#

0.1

Z ~'= N ,yi-( v( xi;a~ a2...aM,/2 3~ 0 -0.01

i

/

0.01

i

i

0.03

i

i

i

0.05

(7)

i

0.07

i-~

u~

where 11/-is the variance of the difference yi-y(xi; al, a2 ...as): Figure 2: Values of n, T~ and combined standard uncertainties in n, In the case of the 3-parameter model the quantity G( T~')2 is treated as an independent variable, thus its uncertainty is calculated separately, by applying the law of propagation of errors on equation G(~ )2=(Ti-Ta)2/G. In fact a 2-dimensional linear fit is required, since a single variable n is modelled as a function of two variables ~" and G ( ~ )2. 3. THE FITTING PROCEDURE The general problem of fitting is to find a model with M parameters aj to represent a series of N observations (x. YO with the greatest accuracy: y(x)----y(x; al...a~

(5)

In the above equation a single variable y can be a function of either a single variable x or a vector x of more than one variables, in the case of a multidimensional model. The basic methodology is always the same (Press et al., 1996; Dietrich, 1991): a figure-of-merit function is selected, to give an indication of the difference between the real data and the model. After this, the model parameters are selected so that the value of the function is minimized. The deviations of the model from the real data can be attributed to experimental errors but also to model weaknesses. The least square (LS) method tries to give an answer to this question: given a set of parameters a:, a2...a~, what is the probability that this set is the desired one?. Assuming that every point of our data is associated to an error which follows a normal distribution around the "true" value with standard deviation o which is the same for all points, the maximalization of the probability that this is the correct set of parameters leads to the minimization of the ftmction:

u~ =Val~yi - y ( x i ;al,a2 ...aM) )

(8)

Since the parameters al...aM are to be calculated, not all the terms that appear in eqn (7) are statistically independent, for this the degrees of freedom are v=N-M. It emerges from eqn (8) that the quantity ~ depends on the experimental uncertainties ~ and ,~. With this consideration in mind, the chi-square merit function actually gives an idea about the relation between the model deviation from the experimental data and the uncertainties in the measurements. A relatively good model will be able to explain the deviations observed on the base of the experimental errors and the and the corresponding X2 function will have a value close to v. Among the advantages of the use of the weighted least square is the fact that the real experimental uncertainties are taken into account in determining the model parameters, the fact that it allows the calculation of the uncertainties in these parameters, and also that it gives a realistic estimation of goodness-of-fit. However, even in the case that a least square fitting is selected by neglecting the uncertainties ui in the phase of the calculation of parameters al...aM, the chi-square function and the goodness-of-fit can still be determined afterwards using eqn (7). From the values of ~ and v the probability Q(0.Sv, 0.5 Z~) that the data do not fit the model by chance can be calculated (Press et al., 1996, Bajpai et al., 1977):

e-tt~-~dt' a>O, r(a)=

Q(a,x)= F(a) x

t'-~e-tdt

(9)

0

The probability Q can be explained as a quantitative indication of goodness-of-fit for the specific model. Generally speaking, if Q is larger than 0.1, then the goodness-of-fit is believable. If it is larger than 0.001, then the fit may be acceptable, under certain conditions. If Q is less than 0.001, then the model (or the estimation procedure) can be called into question.


In the case of solar collectors, where a 2 or 3-parameter model is concerned, the denominator in eqn (7) is written as follows: 2-parameter model Y=a+bX: t~

--'U~i+ b 2 u,q2

(10a)

2 +c2 Ux2i 2 (10b) 3-param. Y=a+bXI+cX2: u~=uy2i + b 2 Ux~ So, the purpose is to minimize eqn (7) with respect to al...aM. Unfortunately, as can be seen from eqns (7), (10a) and (10b), the occurrence of b and c in the denominator makes the eqn (7) non linear. Its solution by analytical methods is possible only if the uncertainty in xi can be considered negligible (Press et al., 1996). Otherwise, the solution is possible by using numerical methods for minimisation of non-linear functions. Generally, the requirements for the acceptance of a good fitting can be reported as follows (ISO, 1995a; Press et al., 1996, Bajpai et al., 1977): I. The goodness-of-fit, i.e. the probability Q(0.5v, 0.5Z2) that the data do not fit the model by chance, should be high or, equivalently, the chi-square statistic should be about the number of degrees of freedom. II. The determined parameters al...aM should be independent, i.e. Covar(ai, aj), "- -100 t .......................................................... of energy content fins/ / . / / ""primary

l m

--o--fin thickness O. 1 mm, channel width optimized

--or-optimized fin thickness (channel width 30 mm)

2'0

4'0

8'o

do

-0.02

~6o o,oo

specific mass flow rate [kg/(m:h)]

Fig. 14: Optimized values for channel width and for fin thickness in dependence of specific mass flow rate. In both cases the fin spacing was not constrained in the optimization procedure. With increasing specific mass flow rate both the channel width and the fin thickness increase. In Fig. 15 the optimized fin spacing is given. The fin spacing for the higher flow rates shows only a small dependence on the two parameters, channel width and fin thickness.

Without fin spacing, fin thickness and channel width constraints we get the highest yearly net energy gain. This is not shown on the graph. If only the fin thickness is held constant, the gap width increases to large values as given in Fig. 14. In this case the heat transfer equations for the rectangular continuous fins are no longer valid and the equations for the smooth absorber are taken to compute the heat gain. This is the reason for the strong increase in energy gain and fin mass for flow rates over 60 kg/(m2h) in Fig. 16. Practical limitations such as constraining the channel width to the maximum value of 30 mm and the fin thickness to the minimum value of 0.1 mm do not lower the yearly net energy gain significantly. In this case an optimal fin spacing for a specific air mass flow rate of 70 kg/(m'h) would be about 6 mm.

Z 3 Absorber with offset stn'pfins Fig. 17 shows the energy gain for various offset strip fins in comparison to continuous fins. The specific air mass flow rate is 70 kg/(m2h). 250/

=

,

=

,

=

/

IllB1

,,o] |oo]

"=

,_..,

I= "64r

tn_ 111 rr

[

210

2--D-fin thickness O.1 mm, channel width 30 mm - o - fin thickness 0.1 mm, channel width optimized --o--fin thickness optimized, channel width 30 mm

1 o

o

2'0

io

6'0

8'o

specificmassflowrate[kg/ln~hl]

1;o

Fig. 15: Optimized fin spacing in dependence of specific mass flow rate, with fin thickness and channel width as parameters.

~'~ 190],

.... ~ =

>' 180 0T

- - - 100 mm long offset fins ~ continuous fins T

o

~

;o

1's

2'0

long offset fins

2'5

30

fin s p a c i n g [mm]

Fig. 17: Yearly net energy gain in dependence on fin spacing for continuous and offset strip fins.


70

The offset strip fins do not perform better than optimally spaced continuous fins. However, if the fin spacing is restricted to non-optimal values, offset strip fins will give better results as continuous fins. Fig. 18 explains why the performance of continuous fins is superior to that of offset strip firm. For a gap width of 30 mm and a fin thickness of 0.1 mm curves for the heat transfer coefficient based on the absorber area in dependence on the pressure drop in the air flow gap are depicted. Since in the interesting region (h > 100 W/(m2K)) the continuous fins give the highest heat transfer for any pressure drop, their performance is superior to all the offset strip fins. 300

i

~'

i

i

i

i

i

i

i

i

optimum for preheating ventilation air (fin spacing 5.9 mm)

250-

--'"

To show the efficiency potential of thin, closely spaced continuous fins for solar air heater applications Fig. 20 was derived. The two selected solar air heaters differ only by the geometry of their fins. One has 1 mm thick fins, spaced 27 mm apart. This (standard) fin geometry is found in a commercially available solar air heater. The other has 0.1 mm thick fins, spaced 5.9 mm apart (optimized geometry). The absorber is 1 m wide and 2.5 m long and has the optical and radiative performance of the selective coating "Black Crystal IF' ( a = 0.937, e = 0.065), see (Brunold 1999). The ambient temperature is 20 ~ the air inlet temperature is 80 ~ the specific air mass flow rate is 72 kg/(m2h) and the channel width is 28 mm. The back wall is made of a 60 mm thick insulation with a thermal conductivity of 0.O40 W/(mK). I

I

I

80.t

/Cf-'~

:

gB

q = 0.77 - 2.65 W/(m=K) - AT / G

~40-

~ continuous fins - - - offset strip fins - - - - smoolh absorber

=9

I

optimized geometry standard geometry

.5r~ 60-

50-

I

o o

0

o

o

~

1'0

1'5

2'o

2'5

3'o

3'5

,~

4'5

50

20-

q = 0.69 - 2.37 W/(rrFK) - AT / G

" " D ~ ' ~ ~

pressure drop [Pa]

Fig. 19: Heat transfer coefficient (based on the absorber area) in dependence on the pressure drop for a smooth absorber and for absorbers with continuous and offset strip fins. The specific air mass flow rate is 70 kg/(m2h). 8 CONCLUSIONS In this paper an optimization method for solar air heaters with flow behind the absorber plate has been introduced. It maximizes the net energy output of the system and takes the long term solar heat gain, the electrical pumping energy and the energy for manufacturing the fins into account. The method was applied to solar ventilation air preheating collectors for which important conclusions can be derived: a) Continuous fins provide the highest net energy gain if they are spaced close to each other. The optimal distance between the fins is about 5 to 10 mm. In the case of a highperformance collector nmning at much higher average insolation values the optimal spacing is generally smaller. b) Due to higher pressure losses offset strip fins show reduced net energy gains compared to optimally spaced continuous fins. However, they show good results generally for large fin spacings. c) The optimum flow regime is laminar, accompanied with low Nusselt numbers and large heat exchange areas. d) In contrast to the second law optimization which considers exergy instead of energy (Altfeld 1985), the obtained fin spacing in this work is much smaller and predicts higher thermal heat gains. To our knowledge there is no commercial solar air heater which incorporates the optimized geometry. They all seem to have notideal fin spacings. On the other hand most of the water to air heat exchangers used for air-condition purposes use thin and closely spaced fins, as recommended in our study.

0

o.oo

o.b~

o.~o (Tin -

o.:~

o.~,o

o.~,5

.-.

0.30

T : m b ) / G [Km=/W]

Fig. 20: Thermal efficiencies of two solar air heaters with different fin geometries for an ambient temperature of 20 ~ and an air inlet temperature of 80 ~ The solar air heater with the optimized fin geometry is about 13 % more efficient as the heater with standard geometry and, not seen in this graph, it has a much shorter thermal response time. Therefore, apart from the higher stationary efficiency, the collector utilizes short time insolation even better. The optimal continuous fin geometries derived in this paper are based on empirical equations and have not yet been validated. Therefore it is necessary to build and test collectors with the proposed optimal geometry. Because the pressure drop increases fast with decreasing fin spacing, the optimal fin spacing for real solar air heaters may be somewhat larger than calculated. ACKNOWLEDGEMENTS This work was supported by the Bavarian Research Foundation (BFS), Munich, within the project "SOLEG" which deals with solar assisted energy supply of buildings. We thank our industrial partners GlasKeil/Wfirzburg, Gebrtider Schneider/Stimpfach and Grammer Solar-Luft-Technik/Amberg, all in Germany as well as SIT EuropeNienna, Austria. REFERENCES

Altfeld K. (1985). Exergetische Optimierung flacher solarer Lufierhirzer. VDI-Fortschrittsberichte. Series 6, No. 175. VDIVerlag, Dfisseldorf.


Beavers G. S., Sparrow E. M., Lloyd J. R. (1971). Low Reynolds Number in Round Pipes and Infinite Channels and Heat Transfer in Transition Regions. J. Basic Eng. 93,296-299. Bhatti M. S., Shah R. K. (1987). Turbulent and Transition Flow Convective Heat Transfer in Ducts. In Handbook of Single Phase Convective Heat Transfer., Kakac S., Shah R. K., Aung W. (eds), Wiley-Interscience, New York. Brunold S. (1999) Qualification Tests of Thermafin Manufacturing, LLC (TML) "'Black Crystal 2 ""Solar Collector Absorber Coating with Respect to Thermal Stability and Resistance to Humidity Involving Condensation. Report, SPF, Hochschule Rapperswil. 5.5.1999. Corradini R. (1997) Ganzheitliche Bilanzierung von Metallen. Thesis, Lehrstuhl ftir Energiewirtschatt und Kraftwerkstechnik, TU-Miinchen, Germany. Duffle J. A., Beckman W. A. (1991) Solar Engineering of Thermal Processes. 2~dedn. Wiley-Interscience, New York. Blomberg T. (1996) Heat Conduct'on in Two and Three Dimensions. Report TVBH-1008, Lund University, Lund, Sweden. Diab M. R. (1981) Experimental and Analytical Study of Heat Transfer Characteristics of Solar Air Heater Incorporating a Finned Absorber. PhD-Thesis, Purdue University, West Lafayette, Indiana. Ebadian M. A., Dong Z. F. (1998) Forced Convection, Internal Flow in Ducts,. In Handbook of Heat Transfer. Rohsenow W. M., Hartnett J. P., Cho Y. I. (eds). 3ra edn. MacGraw-Hill, New York. FIE (1998) Die Bereitstellung von elektn'scher Energie in Deutschland (1996). Forschungsstelle ftir Energiewirtschaft, Am Bliitenanger 71, D-80995 Miinchen, Germany. Gnielinski V. (1976) New Equations for Heat and Mass Transfer in Turbulent Pipe and Channel Flow. Int. Chem. Eng. 16, 359-368. Gnielinski V. (1995). Forsch. im Ing.-Wes. 61, No 9, 240-248 Heaton H. S., Reynolds W. C., Kays W. M. (1964). Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int. J. Heat Mass Transfer 7, 763-781. Heibel B., Hauser, G. (1996). Durchstr~mte Vorhangfassaden zur gorwiirmung der Zuluft mechanischer Liiftungsanlagen. AbschluBbericht, DFG-Forschungsvorhaben HA 1896/1. Mai 1996. Universit[it Gesamthochschule Kassel Jones O. C. (1976). An Improvement in the Calculation of Turbulent Friction in Rectangular Ducts. J. Fluid Eng. 98, 173181. Kabeel E., Mec~'ik K. (1998). Shape optimization for absorber plates of solar air collectors. Renewable Energy 13, 121-131. Kreith F., Bohn M. S. (1986)Principles of Heat Transfer. 4th edn. Harper & Row, New York. Kutscher C. F., Christensen C. B. (1992). Unglazed Transpired Solar Collectors. In Advances in Solar Energy, An Annual Review of Research and Development. Boer K. W. (ed). Vol 7 pp. 283-307. ASES. Kinnebrock W. (1994). Optimierung mit genetischen und selelm'ven Algorithmen. R. Oldenbourg Verlag, Miinchen. Kuzay T. M., Malik M. A. S., B6er K. W. (1975). Solar Collectors of Solar One. In Proceedings of the Workshop on Solar

71

Collectors for Heating and Cooling of Buildings. Sargent S. L. (ed). May 1975, 99-108. Maryland University, College Park, New York City, Manglik R. M., Bergles A. E. (1995). Heat Transfer and Pressure Drop Correlations for the Rectangular Offset Strip Fin Compact Heat Exchanger. Exp. Therm. Fluid Sci. 10, 171-180 Mattox D. L. (1979) Evaluation of Heat Transfer Enhancement in Air-Heating Collectors. DOE-No. ALO-5352-T1, Northrop Services, Inc., Huntsville, Alabama. Merker G. P. (1987). Konvelm've Warmeiibertragung. Springer, Berlin, Heidelberg, New York. Mills A. F. (1962). Experimental Investigation of Turbulent Heat Transfer in the Entrance Region of a Circular Conduit. J. Mech. Eng. Sci. 4, 63-77. Morhenne J., Fiebig M., Barthel H. (1990). Entwicklung und Erprobung einer Baureihe von optimierten, modularen Solarlufterhitzern riD, Heizung und Trocknung. BMFT-Report-No. 0335003E6, Ruhr-Universit~it, Bochum. Piao Y., Hauptmann E. G., Iqbal M. (1994). Forced Convective Heat Transfer in Cross-Corrugated Solar Air Heaters. J. Sol. Energy Eng. 116, 212-214. Pottier K., Beck, A., Benz N. (1996). TestreferenzjahrUnstimmigkeiten in der Globalstrahlung. Sormenenergie 4/96, 22-23. Pottier K., Beck A., Fricke J., (1998a). Solarfassade zur Frischlutt-VorwErmung. In Proceedings of 11. Internationales Sonnenforum. 26.-30.07.1998.510-517. DGS, K61n, Germany. Pottier K., Beck A., Fricke J. (1998b). Dynamische Simulation und Optimierung einer Solarfassade. In Proceedings of 11. Internationales Sonnenforum. 26.-30.07.1998. 791-798. DGS, K61n, Germany. Pottier K., Sippel C. M., Beck A., Fricke J. (1999). Heat transfer and pressure drop correlations for offset strip fins usable for solar air heating collectors. In Proceedings of 15th European Conference of Thermophysical Properties (ECTP). 5.-9. September 1999, Wiirzburg, Germany. In Press. Prasad B. N., Saini J. S. (1988). Effect of arnficial Roughness on Heat Transfer and Friction Factor in a Solar Air Heater. Solar Energy 41,555-560. Shah R. K. (1978). A correlation for laminar hydrodynamic entry length solutions for circular and noncircular ducts. J. F1. Eng. 100, 177-179. Shah R. K., London A. L. (1978) Laminar flow forced convection in ducts. Academic Press, New York. Unsworth M. J., Montheith J. L. (1975). Long-wave radiation at the ground ~+II). Quart. J. R. Met. Soc. 101, 13-34. VDI (1997). VDI-Wiirmeatlas: Berechnungsbliitter fiir den Wiirmeiibergang. 8. edn. Springer, Berlin, Heidelberg. Weinl~ider H., Pottier K., Beck A., Fricke J. (1999). Angular dependent measurements of the thermal radiation of the sky. In Proceedings of 15th European Conference of Thermophysical Properties (ECTP). 5.-9. September 1999, Wiirzburg, Germany. In Press.


72

INCLINATION DEPENDENCY OF FLAT PLATE COLLECTOR HEAT LOSSES Bernd Bartelsen, Markus Kiermasch, Gunter Rockendorf Institut fiir Solarenergieforschung GmbH, Hameln/Emmerthal (ISFH), Am Ohrberg 1, D- 31860 Emmerthal, Germany, Tel. +49 5151/999-522, Fax +49 5151/999-500, e-mail [email protected] Abstract - For flat plate collectors the natural convection in the air gap between absorber and transparent cover is of major importance regarding the collector heat losses. The collector inclination angle affects the natural convection phenomenon and thus influences the collector heat loss coefficient. This has been investigated by experiments on four single glazed selective flat plate collectors, including an additional variation of the ambient air speed. For these collectors a reduction of the effective collector heat loss coefficient of around 0.1 W/m2K per 15~ increase of the inclination angle has been found in the angular range between 15 ~ and 90 ~ As a practical conclusion, this effect has to be taken into account during collector tests carried out on tracking devices. Furthermore, the higher efficiency at 90 ~ inclination angle is an advantage for fafade collectors.

1. I N T R O D U C T I O N For typical selective flat plate collectors the heat transfer by convection in the air gap between absorber and transparent cover is the dominant part of the collector heat losses. The effect of natural convection in inclined rectangular enclosures has been studied by different groups. A well known correlation describing the free convective heat transfer across an inclined air layer of large aspect ratio is given by the following equation (1) (Hollands et al., 1976). 1708 "]*. [1 _ sin(1.8 9T)1"6 91708] Nu = 1 + 1.44-[1 - R a - c"osyA Ra: co's~, +{FRa-eos~'] 1/3 }* / 5830 J - 1 The brackets with asterisk stand for:

(1) [~ *=(IXI + X ) / 2

Equation (1) quantifies the heat transfer by natural convection between two parallel plates, where the inferior one is heated. It is based on measurements with an isothermal temperature distribution on the heated plate and an even absorber surface. This equation is given a high accuracy in a range from 0 ~ (horizontal) to 60 ~ an extension up to 75 ~ is possible with a higher uncertainty. Experimental investigations carried out at the ISFH (Institut fiir Solarenergieforschung GmbH, Hameln/Emmerthal) in 1993 (Bartelsen et al., 1993) showed, that in a real flat plate collector, with a distinct temperature profile and an unevenness of up to 10 mm, the convective heat transfer increases significantly if compared to ideal surfaces. In particular these investigations on a test collector with variable gap size led to the following results: 9 higher heat transfer coefficient for real flat plate collectors than equation (1) gives, 9 only small reduction of the heat transfer coefficient with increasing gap sizes and no maximum resp. minimum at small gap distances, 9 similar curves of the heat transfer coefficient at different inclination angles (30 ~ - 60~ 9 no influence of the fluid mass flow rate and the orientation (horizontal or vertical) found, 9 higher heat transfer coefficient for an inverted flow direction with vertical orientation and 9 lower heat transfer coefficient at low (or zero) irradiance levels.

In numerous efficiency measurements on commercial collectors, carried out in the ISFH solar simulator test facility, the experience was made, that absorbers with even surfaces made out of one metal sheet in most cases lead to smaller heat loss coefficients than absorbers out of single stripes, especially if the stripe absorbers show a higher unevenness. Collectors with thick absorber sheets often showed smaller heat loss coefficients if compared to collectors with thin absorbers. These effects underline the main difference between equation (1) and real collectors: real absorbers in solar collectors always have a well marked temperature distribution, especially in vertical direction to the fluid pipe, and, if they are made out of stripes, they do not represent an even plate. Within the investigation discussed before the influence of the collector inclination on the heat loss coefficient has been considered only for a small angular variation (between 30 ~ to 60~ Especially the question of vertical collectors (90 ~) has not been considered there. With the increasing interest in fafade collectors, the question about the collector heat losses in vertical position arises. 2. E X P E R I M E N T A L INVESTIGATIONS In the following, the effect of the collector inclination on the natural convection in the air gap and thus on the collector performanee will be discussed by the presentation of experimentally determined heat loss coefficients for different flat plate collectors as a function of the inclination angle. For the measurements, two versions of a flat plate collector prototype with typical absorber and glass construction, one with normal opaque insulation, the second with an unusual high thickness of back and side insulation have been produced and investigated. The normally insulated prototype (A) has a thermal insulation of 5 cm thickness (back) resp. 2 em (side) and the highly insulated one (prototype B) a mineral wool thickness of 25 em resp. 5 cm. The insulation thickness of the second prototype is a typical value for the integration of collectors into the facade of advanced low energy houses, which is performed without thermal decoupling by a ventilation layer. In addition to these investigations, one flat plate collectors of commercial production has been measured at different inclination angles between 0 ~ and 90 ~


The tests have been performed in agreement with ISO 9806-1 with highly reproducible test conditions in the solar simulator test facility. The high reproducibility of the test facility is a necessary condition, as only small efficiency differences have to be identified. The collector parameters are referring to mean temperature of the collector heat transfer fluid Tm and the aperture as reference area. The measurements were carried out at an irradiance level of about 820 W/m 2, an ambient air temperature of around 22 ~ and an air speed of 3 m/s. Some efficiency curves have in addition been recorded without forced convection of the ambient air (air speed below 0.5 m/s). The results of the collector tests with the highly insulated prototype are shown in table 1. prototype B

1"10[-]

a 1 [W/Kmz] a2 tw/K~n~]

0 ~ inclined

0.764

3.86

0.013

15~ inclined

0.765

3.85

0.013

30 ~ inclined

0.767

3.75

0.013

45 ~ inclined

0.769

3.62

0.013

60 ~ inclined

0.770

3.52

0.013

90 ~ inclined

0.773

3.35

0.0i3

Tab. 1. Efficiency parameters for the highly insulated collector prototype B at various inclination angles Figure 1 displays graphically the efficiency curves of the highly insulated collector prototype B versus AT/G, where AT is the difference between the mean fluid temperature and the ambient air temperature Ta, with the inclination angle as parameter (numerical data from Tab. 1).

The test results of a fiat plate collector from series production with a selective absorber plate (single sheet, meandering pipe connection) are shown in table 2. commercial collector

110 [-]

,--., 0,7

t="

/3 0 ~ 15.

0,6

0,02

0,04

AT/G [Km'/W] Fig. 1: Efficiency curves of the highly insulated prototype at vari-

a 1 [W/Kmz] a2 [w/Ir

0 ~ inclined

0.778

3.86

0.011

45 ~ inclined

0.780

3.65

0.011

90 ~ reclined

0.783

3.28

0.011

~]

Tab. 2: Efficiency parameters for a commercial collector at various inclination angles These results show, that the heat loss coefficient decreases with increasing inclination angles. This effect is caused by the convective heat transfer in the air gap between absorber and glass pane. From table 1 and table 2 it may be derived, that a reduction of the linear heat loss coefficient a 1 of about 10% at an inclination of 90 ~ results if compared to the parameters determined at an inclination angle of 45 ~ (standard test conditions according to different standards like prEN 12975-2 or DIN 4757-4) 1). At horizontal installations or small inclination angles the heat loss coefficient raises for about 5%. The temperature dependent heat loss coefficient a2 showed a far-reaching independence of the inclination angle. In order to make the changes of the flat plate collector heat losses clear, figure 2 presents the measured effective heat loss coefficient as a function of the collector inclination angle. The effective heat loss coefficient is calculated for a temperature difference of 40 K between the ambient air temperature and the mean temperature of the heat transfer fluid Uloss,40K

0,8

73

=

a 1 + a 2 940K

(2)

For the presentation of figure 2, results from different collector tests have been taken into account: 9 standard flat plate collector from series production (tab.2) 9 normally insulated collector prototype A 9 normally insulated collector prototype A, measured with an ambient air speed below 0,5 m/s (free convection) 9 highly insulated collector prototype B (tab. 1) 9 highly insulated collector prototype B, measured with an ambient air speed below 0,5 m/s (free convection) 9 test collector with variable gap size from former investigations (1993) Figure 2 shows, that the effective heat loss coefficient of all collectors decreases with increasing inclination angles in a similar way, independently of the construction and the air speed conditions above the transparent cover. This can easily be recognized by the parallel shape of the curves.

ous inclination angles (irradiance level 800 W/m 2, air speed 3 m/s) Figure 1 shows, that the collector efficiency increases with increasing inclination angles. This leads especially at high temperatures to a significantly higher efficiency for a vertical collector if compared to a horizontally mounted collector. Due to the reduction of the heat loss coefficient, the collector efficiency factor and thus the conversion factor 110 of the collector increases slightly.

1.The standard ISO 9806-1 requires that the collector shall be mounted at an inclination angle equal to the latitude (+ 5% but not less than 30~


74

3. PRACTICAL RELEVANCE 4,75

I

3

r

n

prototype A

2,75

I

~

A

prototype B

prototype B

rut ~lm:lDr H

Scmmu~on~,econvect~n =Scrnmu=aar~freeoo.voceon re=r=19931 I -_. o r 0 r II

: ,

I

0

=

15

1

~

30

I

=

45

I

~

60

I 75

,

I ' I 90

Collector inclination angle [~

Fig. 2: Effective heat loss coefficient of different collectors at a temperature difference of 40 K for various inclination angles. The effect is very small below 15~ for angles between 15~ and 90 ~ it is significantly higher, whereas an approximately constant gradient may be stated. As a rule of thumb, it can be derived that the effective heat loss coefficient will be reduced by about 0.1 W/m2K (i-0.02) per each 15~ tilt angle increase in the tilt range between 15~ to 90 ~ For the different collectors presented in figure 2, simulation calculations with a theoretical collector model have been performed. In this model the natural convection in the air gap is described with the correlation from Hollands, equation (1). 4,75 4,5

J3,75

pnmtypeB 3,5

3,~

9

"

-~-

~ i cm i n ~ a ~ n

lamuhllJol~

I

0

prmotypeB

~ cm i n s u l i i o n

15

,

I

30

,

I

45

,

,

60

75

90

Collector inclination angle [*]

Fig. 3" Effective heat loss coefficient of the collector prototype B at a temperature difference of 40 K versus inclination angle, in comparison to model calculations. The results for an inclination angle below 60 ~ show a good correspondence of the curve shape. For an angle between 60 ~ and 75 ~ the simulation results indicate smaller heat loss coefficients of the calculated as compared to the measured coefficients. It is reminded, that the best accuracy of equation (1) has been specified in an angular range up to 60 ~.

The practical effect of the inclination dependency of flat plate collector heat losses will be discussed with special regard to the calculation of the yearly energy output of collector systems using simulation programmes and the measurement of collector efficiency parameters on outdoor tracking devices.

3.1 Energy output of collector systems Collector parameters, derived from efficiency measurements according with test standard ISO 9806-1 or other standards are normally valid for an inclination angle of 45 ~ The 45~ are used in general within simulation programmes to calculate the thermal collector output of solar systems and the solar fraction of the heat demand. For the following discussion, the collector parameters are used to determine the annual output of a typical domestic hot water system in Germany. If a collector is installed horizontally or with a small angle to the ground (0 ~ - 15~ the application of the efficiency parameters identified with the relevant inclination angle reduces the energy output of the collector by nearly 2% if compared to simulation results with collector parameters which have been identified at 45 ~. A vertically installed collector has a lower heat loss coefficient and therefore the energy output will increase by up to 4% if compared to the 45 ~ parameters. If the integration into a facade with thermal coupling (no backside ventilation) is considered, the higher rear insulation of the collector will additionally reduce the heat loss coefficient by about 0.3 W/m'K. This difference has been determined from the efficiency curves of prototype A and B. If this improvement is taken into account in addition to the influence of the vertical mounting, the energy output will increase by about 8%. Further positive effects of the fagade position like the lower wind velocity above the outer collector surface cannot be quantified and are therefore not considered in these discussion. With regard to the accuracy of collector efficiency measurements and comparability between different test laboratories, the parameters identified at an inclination angle of 45 ~ are sufficient for the calculation the annual energy output of roof mounted collectors. For vertical facade collectors the lower heat loss coefficient should be taken into account. 3.2 Discussion of collector tests on tracking devices For collector tests carried out on tracking devices the variable inclination angle during the tests will influence the measured efficiency parameters. The influence on the conversion factor % in the inclination range between 30 ~ and 70 ~ for typical fiat plate collectors may be ignored (less than 0.5 percentage points). On the other hand the heat loss coefficient can vary significantly. Depending on the test procedure of the performance measurements, the combination of collector temperature and inclination angle may lead to heat losses at a high collector temperature (80 ~ that differs by about ~- 6% from the value at constant 45 ~ In the regression analysis of the measured data this may lead to different impacts onto the two collector heat loss parameters a 1 and a2, thus creating a unrealistic modification of the curves shape. To avoid this error the efficiency analysis should be carried out only with data coming from a limited range of inclination angles.


As an alternative a selection of the measured data with a balanced proportion of different inclination angles below and above 45 ~ especially for high temperatures is also possible to attain a suitable analysis of the performance measurements. If one of these recommendations is taken into account, the inclination dependency of flat plate collector heat losses has no critical influence on the results of collector tests carried out on tracking devices. 4. CONCLUSION The collector inclination angle influences the heat transfer by natural convection in the air gap between absorber and the single glass cover and therefore the collector heat loss coefficient is affected. For an inclination angle range between 15 ~ and 90 ~ a reduction of the effective heat loss coefficient by about 0.1 W/m2K per 15 ~ inclination angle increase could be identified for each set of collector efficiency curves, where different constructions and ambient air speed conditions have been investigated. This behaviour may also be found by simulations using the equation of Hollands et al., if only an angular range of up to 60 ~ is regarded. The consequences of the inclination dependency for flat plate collector heat losses are of minor importance for the calculation of the energy output of roof mounted collector systems as well as for the analysis of collector efficiency tests on tracking devices. But for vertically mounted or facade integrated flat plate collectors the benefit of the lower loss coefficient should be considered. NOMENCI~TURE a1

linear collector heat loss coefficient, referred to T m (W/m2K) a2 temperature dependent collector heat loss coefficient, referred to T m (W/m2K Y inclination angle, between collector surface and ground AT temperature difference between mean fluid temperature and ambient air temperature (K) G Solar irradiance (W/m 2) rl collector thermal efficiency, referred to Tm (-) rl0 conversion factor (11 at AT = 0), referred to T m (-) Nu Nusselt number (-) Ra Raleigh number (-) Tm mean temperature of heat transfer fluid (~ UL effective heat loss coefficient of collector, referred to T m (W/m2K) UL,40K effect, heat loss coefficient (UL at AT = 40 K) (W/m2K) REFERENCES Hollands K.G.T, et al. (1976) Free Convective Heat Transfer Across Inclined Air Layer. ASME Journal of Heat Transfer, 98, pp. 189-193. Bartelsen B., Jard]en S., Rockendorf G. (1993) Heat Transfer by Natural Convection in the Air Gap of Flat Plate Collectors. In Proceedings of the ISES Solar World Congress, 23-27 August, Budapest, Hungary, pp. 267-272, Pergamon Press, New York.

75


76

PV-HYBRID AND THERMO- ELECTRIC- COLLECTORS Gunter Rockendorf and Roland Sillmann Institut ftir Solarenergieforschung GmbH, Hameln/Emmerthal (ISFI-I), Am Ohrberg 1, D- 31860 Emmerthal, Germany, Tel. +49 5151/999-521, Fax +49 5151/999-500, e-mail [email protected]

Lars Podlowski and Bernd Litzenburger SolarWerk GmbH, Iserstr. 8-10, D- 14153 Teltow, Germany, Tel. +49 3328/448-300, Fax +49 3328/448-301 Abstract Two different principles of thermoelectric cogeneration solar collectors have been realized and investigated. Concerning the first principle, the thermoelectric collector (TEC) delivers electricity indirectly by first producing heat and subsequently generating electricity by means of a thermoelectric generator. The second principle, the photovoltaic-hybrid collector (PVHC) uses photovoltaic cells, which are cooled by a liquid heat transfer medium. The characteristics of both collector types are described. Simulation modules have been developed and implemented in TRNSYS, in order to simulate the behaviour of typical domestic hot water systems. The discussion of the results shows, that the electric output of the PVhybrid-collector is significantly higher than that of the thermoelectric collector.

1. INTRODUCTION The aim of thermoelectric- hybrid- solar collectors is to cogenerate thermal and electric energy within the same module. In cooperation with the company SolarWerk, Teltow (Germany) two different types of thermoelectric- hybrid- collectors have been constructed and investigated at ISFH according to their corresponding physical principles. The first type is called thermoelectric collector (TEC). The principle is to combine a solar thermal collector with a thermoelectric generator (TEG), located between absorber and fluid pipe of the collector, delivering the electric energy. The second collector type is the photovoltaic-hybrid collector (PVHC). The idea of this collector is the combination of photovoltaic (PV) cells with a thermal collector. The PV cells are laminated on the surface of the solar absorber, which is cooled by a liquid heat transfer medium. For both collector types basic work has been carried out. Collector prototypes have been constructed and evaluated by experiments. Furthermore, mathematical collector models has been developed and validated, and thus system simulations could be carried out. Both collector constructions and the main results of the investigations will be described. 2. T H E R M O E L F . ~ C

TEG to its cold junction. This local concentration of heat may be obtained by a gravity assisted heat transfer processes like e.g. boiling-condensing process in heat pipes or thermosyphon cycles. For the following development a water filled heat pipe has been applied as appropriate solution. To generate a high amount of electric power, a high temperature difference at the TEG is necessary. This can only be achieved by a high thermal resistance of the TEG, which consequently leads to a high absorber temperature, if a significant amount of heat will be led over the TEG. The high absorber temperature however increases the thermal losses of the absorber and therefore reduces the solar heat production in the collector part. This results in a reduction of the thermal and electric gains. Therefore it is necessary to use high temperature collectors, what at least requires evacuated tubular collectors (ETC), or even better ETC with concentrating mirrors. Figure 1 shows a principle solution using an ETC with heat pipe, which has been investigated here.

COLLECTOR

2.1 Design principles and selected construction The thermoelectric collector (TEC) combines a solar thermal collector with a thermoelectric generator (TEG). The TEG, which delivers the electric energy, is located between absorber and fluid pipe of the collector. Peltier elements, which are normally taken for cooling purposes, were specially designed for electricity generation in order to use them as TEG. The thermal resistance of the TEG causes a temperature difference, which is proportional to the heat flux from the absorber to the fluid. Furthermore, this temperature difference is proportional to the electric power. Thus, for a high electric performance, all solar thermal heat has to be conducted over the TEG. Therefore a clear separation between the absorber and the fluid part of the collector is necessary, in order to concentrate the solar heat to one point, the hot junction of the TEG, and then to let it pass over the

Fig. 1: Scheme of thermoelectric collector Using a dry coupling, the condenser of the heat pipe heats the lower part of the heat exchanger to a high temperature, which will act as the hot junction of the TEG. The upper part of the heat exchanger is cooled by the heat transfer fluid of the solar loop, thus acting as the cold junction. The TEG is arranged between the hot and the cold junction. The heat passing the "lEG causes an electric power, which in its maximum power point (MPP) is proportional the temperature difference between the hot and cold junctions. The amount of heat transferred over the "lEG may be directed to

ISES Solar World Congress 1999, Volume I!!

an application like hot water preparation. 2.2 Investigations on thermoelectric generators A central objective of the development work is to investigate the behaviour of appropriate TEGs with regard to their electric and thermal properties. For this purpose, a heat exchanger test stand has been built up and the behaviour of different TEGs (area around 9 cm 9-, thickness 3 - 5 mm, manufacturers Kunze and TECOM) has been measured for varying boundary conditions (Giebel, 1997). The interactions of the electric and thermal properties depend on various parameters taking into account the different physical effects (mainly Seebeck- and Peltier- effect). E.g., the electric output is a function of mean TEG temperature, temperature difference and inner electric resistance of the element and is therefore coupled with the thermal resistance between hot and cold junction, which has been found is mainly depending on the mean temperature and the electric current generated by the dement. Thus, both the thermal and the electrical characteristics are depending on each other. Further complication is caused by the dependency of the inner resistance of the mechanical pressure, to which the element is exposed by the heat exchanger package. Finally, practical considerations like the heat transfer resistance between the TEG and the heat exchanger resp. the bypass heat flow caused by the clamping mechanism have also to be taken into account. For theses reasons, only simplified dependencies could be developed. The correlations, however, have been proved to be sufficiently precise for the description of the whole collector. A deviation between measured and calculated results of less than +/ - 3% concerning the electric output has been found during the collector investigations. For the behaviour of the electric output, the following simplified formula has been worked out:

Pel =

Rload Rl~

(Ri +

2 2 9[b I - ATTEG2 + b 2 9ATTE G 9Tavg (1)

2 + b 3 9ATTEG2- Tavg ]

77

Three vacuum-tubes with heat pipe (producer Thermomax, UK) with 0.l m 2 absorber area each and water as heat pipe medium have been connected via the specially designed heat exchanger to a fluid circuit. Figure 2 shows the construction of one heat exchanger element.

Fig. 2: Construction of the heat exchanger of the thermoelectric collector-prototype Special care was necessary in order to avoid additional thermal resistances between the heat exchanger and the hot resp. cold junction of the TEG. Furthermore, as only the heat passing via the TEG produces electricity, any bypass heat flow has to be minimized. This is important for the design and the selection of the clamp device and the surrounding insulation material. Finally, the temperature stability of the applied material has to be high enough to withstand the expected high temperature, especially in case of stagnation. The prototype collector has been tested in agreement with ISO 9806-1. To assess the influence of the TEG-integration, a modified collector without TEG has been investigated, too. Due to the small amount of electric output, the thermal and electric yield may be discussed separately. Figure 3 shows the thermal efficiency curves of both collector prototypes. i

The heat transfer capability between hot and cold junction may be described by

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UTE G- C1 9I -t- C2 9Tavg + c 3 9I. Tavg + c 4 9I. T2avg (2)

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+ UTEGO It has been shown during the TEG experiments and within the collector tests (section 2.3), that both equations describe the measured behaviour with a high accuracy. The measured performance of the TEGs is at 60 W input heat and 20 ~ fluid temperature between 1.3 and 2.0 W, i.e. an efficiency of around 2.3 to 3.2% has been achieved, while the TEG is operated in MPP. The thermal conductivity of one TEG is around 0.4 W/K, which causes an overtemperature at the hot junction of around 150 ~ (in MPP-operation, irradiance level approx. 900 W/m2). 2.3 Construction and assessment of the thermoelectric collector A prototype of a thermoelectric collector has been constructed.

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Fig. 3" Thermal efficiency curves of thermoelectric collector, compared to same collector without TEG, irradiance level approx. 800 W/m 2, air speed 3 m/s, referring to aperture area and mean fluid temperature The installation of the TEG with its high thermal resistance leads to a drastic decrease of the collector efficiency factor and thus reduces the conversion factor rl0 by around 45%, if compared

78


to the identical collector without TEG. The electrical efficiency came up to a maximum value of 1.1% of the incoming solar radiation, which is around 2.8% of the transferred heat. The integration of a TEG rises the absorber temperature and by this way the losses of the solar collector are increasing significantly, whereas the electric output remains rather small.

if the improved TEC would be operated with constant fluid inlet temperature (10 ~ over one year. With these improved elements, annual simulations of typical solar domestic hot water systems have been carded out, where the heat transfer capability of the TEG has been varied. Figure 4 shows, how the variation of the conductivity affects the output of electric and thermal energy.

2.4 Simulation of thermoelectric collector and system For the calculation of yearly energy gains, a dynamic simulation model was developed. As the thermal and electrical properties may not be isolated, an iterative calculation process is necessary. The model has been validated with the experimental results of the prototype collector tests (Sillmann, 1997). It has been transferred to a TRNSYS simulation tool, and thus, it could be implemented in a solar system simulation programme.

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The simulations lead to the following results: 9 The thermal connection between condenser and TEG and between TEG and fluid must be good. A minimum heat transfer capability between absorber and fluid except for the TEG itself of 20 W/m2K should be achieved for the used vacuum tube and heat exchanger, referred to the absorber area. The bypass heat flow should be minimized in order to come up to higher electric gains, whereas a higher bypass heat transfer increases the thermal output. 9 If the collector is operated during the whole year with a constant inlet temperature of 10 ~ (if irradiance is above 10 W/ m2), a thermal output of 660 kWh/m 2 and an electric output of 14 kWh/m 2 may be expected at Hanover (Germany). At 90 ~ inlet temperature, the output is 260 resp. 6 kWh/m 2 per year. 9 If a vacuum tube with a significantly lower loss coefficient as compared to the prototype (stagnation temperature around 70 K higher) would be used, the thermal output would nearly not be affected and the electric output would increase by around 5% at 10 ~ and 65% at 90~ fluid inlet temperature. That means, that a better insulated collector first of all promotes the electric gains. 9 An increase of the thermal conductivity of the TEG by a factor 2 would increase the thermal output to 750 kWh/(m2a) (constant fluid inlet temperature of 10 ~ and decrease the electric gain to approx. 50%. On the other hand, a reduction by 75% would lower the thermal output down to 320 kWh/(m2a), whereas the electric gain would rise to approx. 40 kWh/(m2a). The prototype of the TEC has shown the technology inherent disadvantage, that the high thermal performance of evacuated tubular collectors will be significantly decreased and the returned electric energy only comes up to small values. Therefore the conversion efficiency of the TEG has to be improved. The best laboratory elements attain effiency values, which are about 3 times higher than the used elements, they come up to around 30% of the Camot efficiency (Rowe et al., 1995), which is assumed to be at the upper technical limit. This means an efficiency increase of the TEG by a factor 3 if compared to those TEGs used in the prototypes. This theoretical TEG improvement would nearly not influence the thermal output of the improved thermoelectric collector 1), but it would enlarge the electric gain by a factor 2.5, 1. The improved TEC is a theoretical collector combination, in which the far better TEGs with a 3 times higher conversion efficiency have been applied together with the vacuum tubes of the prototype.

o

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Fig. 4: Annual energy yield of a thermoelectric collector with efficiency-improved TEG in solar domestic hot water system (Hanover) vs. conductivity factor of the TEG. The conductivity factor is a multiple of the measured heat transfer capability of the prototype (collector area 5 m 2, energy demand 2600 kWh/a). If 5 m 2 of the constructed prototype collector with the improxrA.TF,G would be installed in a typical hot water system, a yearly gain of 1650 kWh thermal and of 50 kWh electric energy could be achieved. Figure 4 shows, that a lower TEG conductivity increases the electric gain, but strongly reduces the thermal gain. If the conductivity is halved, the yearly electric output is nearly doubled to 100 kWh/m 2, whereas the thermal gain is reduced by around 500 kWh/(m2a) (i.e. 30%). The discussion shows, that a thermoelectric collector with acceptable technical properties needs high efficient vacuum collectors and high efficient TEGs, both at the upper technical limit. But even with these components, it seems to be unrealistic to develop a thermoelectric collector with economically promising prospects.

3. PHOTOVOLTAIC- HYBRID- COIJ,ECTOR

3.1 Design principles of a PV-Hybrid-Collector The idea of the photovoltaic hybrid collector (PVHC) is the combination of photovoltaic (PV) cells with a thermal flat plate collector. The PV cells are laminated onto the surface of the alum i n i m solar absorber. On its back side, a copper fluid pipe is clamped on. This PVH-absorber is cooled by a liquid heat transfer medium. The absorber is integrated in a standard ahminium frame, usual for solar thermal collectors with normal solar glass, an air gap of 3 cm and a backside insulation (mineral wool) of 5 era. Figure 5 shows the explosion drawing of the PVHC.


79

This model has been developed interactively with the experiments at the first prototypes. The development process has given, that it is sufficient to introduce only one thermal capacitance into the model; it is located in the fluid node. This inaccuracy is acceptable, as the thermal resistance between the fluid and cell node is rather low (this has not been the case at the TEC, compare chapter 2). The validation procedure of the f'mal simulation programme shows a good agreement with the measured data, also in the dynamic parts of the time series.

3.3 Performance of the PV-hybrid-collector

Fig. 5: Construction of the photovoltaic-hybrid collector (PVHC) Important for the effectiveness of the PVHC is a small thermal resistance between the PV cells and the fluid. Both the thermal and the electric performance decrease with rising fluid temperature. As the absorber temperature is strongly affected by the fluid temperature, the design and operation of the hot water system, e.g. the solar fraction, is very important, too.

The general thermal behaviour is similar to that of a nonselective flat plate collector. The reason is the high thermal emissivity of the laminate-cell-package (e around 0.90), which leads to a high thermal loss coefficient. The conversion factor rio of the PVH-collector is only somewhat smaller than that of a nonselective flat plate collector. The difference is caused by the absorption coefficient a, which comes up to 0.915. This is a few percent points lower than with a black solar colour. The electrical operation mode has to be taken into account. If the PV-part has no efficiency (open resp. short circuit), the heat generation will be higher if compared to maximum power point (MPP) operation.

3.2 Collector model In order to support the development work, a simulation model is necessary. The model is used for the component optimization and the simulation of a whole solar system. To use the universal properties of TRNSYS, the PVH-collector module was written as TRNSYS- type and implemented into a TRNSYS configurations of a solar domestic hot water system. The PVH-collector model has to describe both the electric and the thermal behaviour, as well as the interaction between these characteristics.

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J o

A T/G in K/(W/m a)

Fig. 7: Thermal efficiency curves of the PVH-collector, with zero or with maximum electric efficiency, (according to ISO 9806-1, referring to mean fluid temperature and aperture area)

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The efficiency curves have been measured at an irradiance level of about 820 W/m 2 and at an ambient air speed of 3 m/s. Figure 7 shows, that the curve in electric MPP-operation is moved almost in parallel downwards if compared to the open circuit curve by around 0.10. That means, that the extraction of electric energy directly affects the zero loss efficiency rio, but nearly not the heat loss coefficient. Consequently, the stagnation temperature of a PVH-collector decreases if the electric part is operated in the MPP- point (see below). thermal coefficients

. J

Fig. 6: Thermal and electrical model for the PVHC Figure 6 shows the node model as basis for the thermal considerations, including the integrated Two-Diode-Model for the description of the electric performance of the PV-cells. The interaction between both models is taken into account by the cell temperature, which on the one hand is affecting the cell performance and on the other hand is almost equivalent to the absorber temperature and therefore directly influencing the thermal gain and loss mechanisms.

electrical coefficients (STC)

rio [-]

0,726 (PV-opencircuit) 0,633 (PV-MPP)

Isc [A]

2,84

a1 [W/m2K]

5,88 (PV- opencircuit) 5,64 (PV- MPP)

Uoc [V]

107

a2 0,016 (pv- opencircuit) PMPP [W/m2K2] 0,015 (PV-MPP) [W] riel [-1

220 0,103 (aperturearea) 0,121 (cellarea)

Tab. 1: Performance parameters of the PVH-collector, prototype

80

I S E S S o l a r W o r l d C o n g r e s s 1999, V o l u m e III

The thermal parameters are determined according to ISO 98061 and they are referring to the mean fluid temperature and the aperture area (2.1 m2). The electric performance indicators are the aperture area and cell area related efficiency data, the electric power of the module (cell area 1.81 m2), the open circuit voltage and the short circuit current, all referred to standard test conditions (STC, irradiance 1000 W/m 2, cell temperature 25 ~ Table 1 directly shows, that the difference between the conversion factors rio in either MPP or open circuit is nearly equivalent to the electrical MPP-efficiency, what is a nice approval for the first law of thermodynamics in this special case. During the efficiency measurements with low inlet temperature, the mean cell temperature is only 11 K above the fluid temperature. For the implementation of the simulation programme, more detailed parameters than the ISO- coefficients and STC- parameters are required. For this purpose, the parameters of the Two-DiodeModel have been identified as well as the single resistances between the cell node and the fluid resp. the environment, as displayed in figure 6.

3.4 Simulation of the PVHC system yield As basis for the system simulations, a domestic hot water system with single drinking-water storage and internal heat exchanger has been placed at Wtirzburg, Germany, with an inclination angle of 35 ~, south. The annual heat demand is 2600 kWh/a, at a demand temperature of 45 ~ The collector area was enlarged from 1 module (2.1 m 2) in steps up to 5 modules. Figure 8 shows the results of these simulations. 5O0 4#5

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400

_= _= m

S00 " " " , ~ 2

3

combinations. 2 PVH & 1 SFP 2 PVH & 2 SFP 4 PVH & 1 SFP PVHth [kWh/m2]

291

232

193

SFPth [kWh/m2]

399

310

352

PVHel [kWh/m2]

90,8

88,5

87,0

SF

0,53

0,60

0,60

Tab. 2: Annual output of PVH and selective flat plate (SFP) collectors and solar fraction, for different module combinations (2.1 m 2 collector area for each module) Table 2 shows, that the combination of PVH-collector and selective flat plate collector leads to higher solar fraction values. The electric output of the PVH-collector is somewhat lower than that of standard PV-modules, which would come up to an annual yield of about 100 to 110 kWh/m 2. This difference is mainly caused by the higher reflection losses at the glass pane and a relatively low ratio of cell and aperture area. At the PVHC this ratio is 0.85, whereas around 0,90 for standard PV modules are typical. If the annual yield is related to the electric power (instead of module area), the electric output of the PVH-collector is nearly identical to that of standard PV-modules. The influence of the module temperature is rather small, which will be discussed in the foliowing. As the efficiency of PV-cells rises with decreasing temperatures, low module temperatures are desired. One of the main questions of PV-hybrid systems is, whether the mean operation temperature is higher if compared to standard PV-modules. For this purpose, the irradiation weighed mean cell temperature is defined as follows:

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200

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(3)

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Fig. 8: Annual thermal and electric energy gain versus the number of PVHC-modules, with additional data of the solar fraction of the hot water demand (SF) While the thermal gain decreases from 432 to 177 kWh/(m2a), the electric gain only decreases from 92 to 86 kWh/(m2a). The solar fraction of the thermal demand increases from 22 (1 module) to 50% with 5 modules. As thermal part of the collector shows the performance of a nonselective collector with additionally reduced thermal gains, the system output in central Europe is restricted to a solar fraction of around 50%. That means, that significantly higher solar fractions, corresponding to a 100% covering of the demand during summer may hardly be achieved. On the other hand, this collector should ideally be used in preheating systems, where both the thermal and the electric gains can benefit from the low fluid inlet temperature. In order to attain a higher solar fraction, a series installation of PVH-collector modules and standard selective flat plate collector modules is possible. Table 2 shows the annual output of different

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Fig. 9: Irradiation weighed mean cell temperature versus number of modules, in comparison to standard PV-modules Figure 9 shows, that the mean cell temperature is increasing with the number of modules, but even with 5 modules (10.5 m 2) it is still in the range of standard PV-modules, for which a wide variety exists, depending on wind exposure and integration technique. In a warmer climate like in southern Europe, the mean cell temperature of the PVHC could even be lower than that of standard PV-modules.


3.5 Reliability questions and possible potential of improvement The PVHC comes up to a stagnation temperature of 147 ~ at 1000 W/m 2, 30 ~ ambient temperature and calm wind conditions, if operated in open circuit conditions. This is a typical value for nonselective collectors. If the module is operated at the same time in its electrical maximum power point, the stagnation temperature decreases by around 12 ~ Special regard must be given to the laminate construction, the electric cables and the connecting boxes, which all have to withstand these extreme temperatures. The collector prototype already shows a good performance, which has only little potential of improvement for this construction type. The thermal contact between absorber and fluid may still be improved insignificantly, by which the thermal gain may be increased by about 2 to 4%. New high efficiency cells could lead to an enhancement of the electric gain.

4. CONCLUSIONS The principle of the TEC is to produce first heat, and then to transfer this heat over the thermal resistance of the TEG, where it will partly be transformed into electricity, the remaining heat has to be cooled away. It follows from this serial energy flow, that the absorber must be maintained at a high temperature as the electrical generator needs a high temperature difference. Therefore, even with high efficiency collectors, the thermal efficiency will decrease significantly. The first requirement is to use solar collectors with very low loss coefficients, e.g. by concentrating the irradiance. The further disadvantage is the low conversion factor of TEGs, where only a value of 30% of the Camot efficiency seems to be realistic. In contrary to the TEC, the principle of the PVH-collector is the direct electricity production, i.e. the efficient direct use of the high exergy content of the radiation, and only the remaining radiation energy will be transformed to heat. This heat will be used on a temperature level as requested by the solar system. Hence, the PVH-collector produces heat and electricity in parallel. The comparison of the solar system simulations between the existing PVH- collector prototype and the advanced extrapolated TEC shows the advantage of the PVHC-principle. The improved TEC (5 m 2 evacuated tubular collector) would lead to an electricity gain of only 50 kWh/a and meet the thermal demand with a solar fraction of 53%. The PVHC (10.5 m 2) delivers around 920 kWh/a electric energy and covers the thermal demand by 50%. Regarding the same collector area of 5 m 2, the PVH-collector comes up to an electricity production of 450 kWh/a, what at any rate is 9 times higher than the electricity production of the advanced extrapolated TEC. Precise cost statements or estimations are not available. It may however be assumed, that the production technology of the PVHcollector with its known processes from PV and thermal flat plate collector technique has a higher economic potential. It therefore may be concluded, that the TEC will only be of interest for special applications. In a direct comparison the PVHcollector technology shows many advantages. However, hybridcollectors often show a non optimal behaviour in comparison to the parallel operation of the basic technologies. But for specific applications and special purposes, the advantages of only one type of solar module for heat and electricity production may be so convincing, that to our opinion these collectors will occupy a place in future developments and a future market.

81

NOMENCLATURE A a1

area (m2) constant collector heat loss coefficient, referred to Tm (W/m2K) a2 temperature dependent collector heat loss coefficient, referred to T m (W/m2K2) b 1,b2,b3 coefficients to calculate the electric energy gain of the TEG (different dimensions) Cl,C2,C3,C4 coefficients to calculate the heat transfer capability of the TEG (different dimensions) Cfluid capacity of the PVHC referred to the fluid temperature

0d/K) D1,D2 G I Ise J Jph K Pel PlvtPP Quse Ri Rload Rs Rsh Tamb Tavg Teen Tfluid Tm AT

diodes of the Two-Diode-Model global solar irradiance (W/m2) current (A) short circuit current (A) current (A) photo-current (A) incident angle modifier coefficient (-) electric power (W) electric power in MPP operation (W) useable heat-flux (W) inner resistance (fl) load resistance (f~) serial resistance of the Two-Diode-Model (f2) shunt resistance of the Two-Diode-Model (f~) temperature of the ambient air (K) average temperature of the TEG (K) temperature of the PV- cells (K) temperature of the heat transfer fluid (K) mean temperature of heat transfer fluid (K) temperature difference between mean fluid temperature and ambient air temperature (K) ATTEG temperature difference between the hot junction and the cold junction of the TEG (K) U voltage (V) Uoe open circuit voltage UAlu heat transfer capability of the aluminium absorber (W/m2K) Ueonv fluid heat transfer capability between fluid pipe and fluid " (W/m2K) UEVA heat transfer capability of the laminate (W/m2K) Uloss effective heat loss coefficient between the absorber and the ambient (W/m2K) UTEG heat transfer capability of the TEG (W/K) UTEG0 heat transfer capability of the TEG, constant part (W/K) O~ absorption coefficient (-) conversion factor; i.e. thermal collector efficiency at AT rl0 = 0, referred to Tm (-) electrical efficiency (-) riel incident angle (o) O

82


REFERENCES

Giebel, Ulfert; Untersuchungen an thermoelektrischen Elementen zur Stromerzeugung in Sonnenkollektoren; diploma thesis at Institut fttr Solarenergieforschung; Emmerthal; 1997 Sillmann, Roland; Konstruktion, meBtechnische Bewertung und Simulation eines thermoelektrischen Kollektors; diploma thesis at Institut ftir Solarenergieforschung; Emmerthal; 1997 Rowe, D.M. et al.; CRC Handbook of Thermoelectrics; CRC Press, Inc.; 1995 TRNSYS 14.1; A transient system simulation program; Solar Energy Laboratory University of Winconsin; Madison, USA; January 1994 Litzenburger, B.; Podlowski, L.; Rockendorf, G.; Sillmann, R.; Entwickhmg eines PV- Hybrid- Kollektors; in Proceexlings of the 8. Symposium Thermische Sonnenenergie; Ostbayrisches Technologie Transfer Institut e.V., Regensburg; 1998; pp. 77-82


E L A S T O M E R - METAL- A B S O R B E R

83

- DEVELOPMENT AND APPLICATION

Bemd Bartelsen, Guntet Roekandorf Institut ftir Solarenergieforschung GmbH, Am Ohrberg 1, D-31860 Emmerthal, Germany, Tel. +49 (0)5151/999-522, Fax -500

Nortmrt I/ennemann Fachhochschule Osnabrtick, Albrechtstr. 30, D-49076 Osnabrtick, Germany, Tel. +49 (0)541/969-2940, Fax -2999

Rainer T e l l , Klaua Lorenz Solar Energy Research Centre - Dalarna University College, S-78188 Borl/inge, Sweden, Tel. +46 (0)23/778-703, Fax -701

Gottffi~t PurkarthoIer Arbeitsgemeinschaft Erneuerbare Energie, A-8200 Gleisdorf, Austria, Tel. +43 (0)3112/5886-16, Fax -18 Abstract - A new principle of a solar collector, that consists in appropriately shaped metal form plates as absorber and clipped in elastomer fluid pipes, the so called elastomer-metal-absorber, will be presented. The advantages are its freeze resistance, the seawater suitability and new possibilities for cost reducing collector installation and system techniques. The design parameters including a detailed analysis of the thermal resistance between absorber and fluid will be discussed, where special regard is given to the development of an appropriate elastomer material with high thermal conductivity as one of the key items. The first development steps have shown, that absorbers with a high thermal performance may be constructed. Finally, the idea to apply the principle of the elastomer-metalabsorber to metal roofs and faqades will be presented. This idea is followed up within a development project. 1. INTRODUCTION The idea of a combined absorber with a metal absorber sheet for the absorption of the solar radiation and a flexible elastomer fluid pipe for the transport of the solar heat has been developed.

Figure 1: elastomer-metal-absorber construction. As shown in figure 1, a round shaped clip profile is integrated into a metal plate, which has an absorption layer for solar thermal conversion. In this profile an elastomer tube for the heat removal is clipped in. The application of this elastomer-metal-absorber in solar thermal collectors offers the following potential advantages and essential possibilities: 9 Due to its inherent freeze resistance, operation without an antifreeze additive is possible. 9 System installation without heat exchanger in the solar loop may be discussed. 9 Operation with a corrosive fluid is possible, e.g. direct flow with sea or brackish water. 9 New and simplified techniques for the collector and system installation can be developed. The most promising application results from the new installation possibilities for the collector and the system. It is intended to integrate this new collector concept into roofs and faqades made out of metal form sheet elements. This idea will be presented in the following. Furthermore the elastomer-metal-absorber concept seems to be an attractive collector for the solar desalination of brackish and sea water, as the collector may be operated directly

with corrosive liquids without cost intensive corrosion protected heat exchangers. The desalination process should be designed to operate on a low temperature level (e.g. around 70 ~ The use of elastomer tubes in collectors requires appropriate absorber constructions. Different constructions have been developed and investigated with regard to the internal heat transfer resistance in combination with freeze-thaw-cycles. These absorbers have been integrated in solar collector prototypes, and thus the thermal performance and the reliability have been examined. A high thermal efficiency may only be achieved with a low heat transfer resistance of the complete construction. In order to minimize this thermal resistance between absorber and fluid, the low thermal conductivity of standard elastomer material has to be improved. Therefore different elastomer mixtures with significantly higher thermal conductivities and acceptable mechanical properties have been developed and investigated. In the following, the results of the development and analysis work will be presented, future applications will be discussed.

2. COLI~CTOR DESIGN It is evident, that the low thermal conductivity of normal elastomer material results in a high thermal resistance between absorber and fluid, which lowers the thermal performance of a collector with this design. For this reason, a theoretical study of the absorber heat transfer had to be performed first. The results of appropriate numerical calculations have led to the following conclusions: 9 A direct contact of the metal absorber fin with the elastomer tube is necessary, no additional adhesive or contact material should be used. 9 The thermal conductivity of standard black elastomer material (around 0.25 W/mK) should be increased to a value of around 0.7 to 1.0 W/mK. The contact area between the absorber fin and elastomer tube must be large, the wall thickness of the tube should be small and finally, the diameter of the tube should be large.

84


The last two requirements are in contradiction to the necessary strength of the elastomer tube at operation pressure. Of special importance for a high thermal performance is the contact between the metal absorber and the elastomer tube. Therefore, during the first development steps, different absorber stripe constructions as well as different collector prototypes have been investigated with special regard to the heat transfer characteristics, the thermal performance and the reliability. Figure 2 shows five different constructions of realized absorber shapes, which have been investigated up to now.

3. ANALYSIS O F I N F E R N A L T H E R M A L R F ~ I S T A N C E The efficiency of a solar collector mainly depends on the quality of the absorber. Beside the absorption and the emission of the coating, the capability to transfer the heat from the absorber to the fluid is important. Figure 3 shows the thermal resistance network of a typical absorber stripe.

Figure 3: Simplified thermal steady state model of absorber stripes.

Figure 2: Different design types of the elastomer-metal-absorber. Type "A" is a typically soldered or welded absorber construction which has been used for the first experiments. Type "B" is an absorber construction out of roll bended aluminium sheets. The clip profile, which embraces the elastomer tube, is integrated in the sheet, thus no welding or soldering is necessary. Type "C" and type "D" are aluminium roll shaped constructions, which are used as absorber in typical thermal collectors. Normally a copper fluid tube instead of the elastomer tube willbe used in the clip profile. Type "E" is a specially developed absorber construction out of roll bended aluminium sheets. This clip profile is an improvement of the types "A" to "D" and takes the capabilities of a roll-form machine for 1 mm thick aluminium sheets into account. These different constructions have been used as absorbers in collector prototypes for the measurement of the thermal performance and as single absorber stripes for the investigation of the internal heat transfer capability.

The heat has to pass four single resistances on its way to the fluid. These are the resistance of the absorber fin and of the base connection, the tube wall resistance and the convective resistance between tube and fluid. The serial connection of these single resistances is equivalent to the total resistance between the absorber and the fluid, (1/Uint). For the elastomer-metal-absorber the internal thermal resistance resp. the internal heat transfer capability depends on: 9 fin resistance - characterised by the tube distance W, the base diameter D, the fin thickness sf and the fin conductivity kf. 9 connection between fin and tube - characterized by the connection technique and its production quality. 9 tube resistance - characterised by the conductivity of the elastomer k t, the tube diameter d t, the wall thickness of the tube st and the contact area between the clip profile and the tube, i.e. the contact angle 9. 9 convection between the inner tube wall and the fluidcharacterized by the convective heat transfer coefficient txfluid, which is a function of the fluid, its flow velocity and its temperature and the inner tube diameter and surface. In metal absorbers, the resistance of the tube wall (1/Utube) is normally neglectable because of the high conductivity of the metal fluid tube. However, in the case of the elastomer-metalabsorber, the tube resistance is very important for the total resistance of the absorber construction. It may be summarized, that the internal heat transfer resistance depends on the construction parameters of the absorber sheet and the fluid tube, the connection technique between the absorber sheet and the fluid tube, its production quality and of the operation parameters. The dependency of the internal thermal resistance of these different construction parameters has been determined by calculations, for which the following base case parameters of the elastomer-metal-absorber have been used:


fin W [mm]

D [ram]

100

7

tube sf kf dt,i [ram] LW/mK] [mm] 1

200

9

conv.

st [ram]

tp [o]

r d L-W/m2]

2

~',7~)

2000

Table 1: Base case parameters for the calculation of the internal thermal resistance of the elastomer-metal-absorber. Figure 4 presents results of Uint-calculations carded out for different tube distances W, in figure 5 the tube diameter to wall thickness ratio is varied. In both figures, the parameter is the thermal conductivity of the elastomer material.

85

importance is the tube distance. Furthermore, the tube diameter resp. the thickness of the tube as well as the contact area between the clip profile and the tube have a clear influence on the internal heat transfer resistance of the construction and therefore on the efficiency of the collector. As the collector efficiency factor F' and thus the conversion factor 11o depend on the ratio Uint F'= Ulos s + Uint ,

(1)

the Uint-value should be maximized by optimization of the complete absorber construction. For a nonselective single glazed collector Uint should be higher than 50 W/m2K (Uloss = 5.5 W/m2K, F ' = 0.90) and for an unglazed absorber a value of more than 70 W/m2K is desired (Uloss = 15 W/m2K, F' = 0.82). These relatively high Uint-Values may only be achieved with a thermal conductivity of the elastomer of at least 0.6 W/InK, if a realistic tube distance of more than 80 mm is assumed. The other alternative, to reduce the tube wall thickness, has clear boundaries: A long term reliability requires a wall thickness of at least 1.5 ram. Therefore the thermal conductivity of standard black elastomer material (kt = 0.25 W/mK) has to be increased significantly.

4. DEVELOPMENT OF THE EI~STOMER-MATERIAL

Figure 4: Internal heat transfer capability Uint versus tube distance, parameter is the thermal conductivity of the elastomer material.

The department of material technology of the University of Applied Science in Osnabrtick has developed an elastomer material based on ethylene-propylene-dien-terpolymer (EPDM) for the application in the elastomer-metal-absorber. Main part of the development was to increase the low thermal conductivity of typical elastomer material with a simultaneous improvement of the mechanical strength. In addition to typically used well conductive f'dling materials like carbon black, particles out of aluminium and graphite have been applied during the development steps. The EPDM mixture is varied with different types of carbon black with high electrical conductivity, two kinds of aluminium particles and various kinds of graphite powder. The measured thermal conductivity and the tensile strength of some of the elastomer mixtures are presented in figure 6.

Figure 5: Internal heat transfer capability Uint versus the ratio of mean tube diameter to the tube wall thickness, parameter is the thermal conductivity of the elastomer material. As figure 4 and 5 show, the internal heat transfer capability of the elastomer-metal-absorber is mainly determined by the low conductivity of the elastomer material, which leads to a high thermal resistance of the tube. This tube resistance becomes even more important, if the amount of collected heat transported over this resistance increases. Therefore the second parameter of major

Figure 6: Thermal conductivity and tensile strength of different elastomer mixtures.

86


The mixture indication starting with a "V" labels the first series with only one single additive, the indication "M" is for laboratory mixtures with two conductive filling materials and the indication "D" stands for elastomer mixtures produced by using an industrial mixing device. For the elastomer mixtures in figure 6 the following filling materials have been used: "V-0" pure polymer "V-I" addition of 100 phr 1) aluminium particles "V-2" addition of 40 phr carbon black "M-4" addition of 40 phr carbon and 80 phr aluminium "M-7" addition of 30 phr carbon and 80 phr graphite "M-13" addition of 10 phr carbon and 90 phr graphite "D-4" addition of carbon and graphite The pure polymer without filling materials shows a thermal conductivity of about 0.2 W/mK. Aluminium as single filling material like in mixture V-1 improves only the thermal conductivity. If a conductive carbon black is added to the mixture (V-2), the tensile strength is raised more than six times and the thermal conductivity is doubled. The mixtures M-4 and M-7 contain two filling materials. Beside the conductive carbon black an aluminium or a graphite powder is added to the polymer. The thermal conductivity is raised up to around 0.8 W/mK, four times the value of the pure polymer, and the tensile strength is at a high level, too. For the efficiency measurement of the first improved prototypes, the elastomer mixtures M-5 (similar to M-4) and M-7, with a good thermal conductivity and a good tensile strength, have been used. From these new elastomer mixtures, tubes have been extruded and integrated into the test collectors. Due to the high carbon black content, the materials M-5, M-7 and M-13 have a very high viscosity during the mixing process and the extruded tubes show a high hardness and a low flexibility. Furthermore, the tube surface has a significant roughness. The conclusion of this first elastomer development step is, that high thermal conductivity and tensile strength values have been achieved, but the material is not appropriate for an industrial production process and does not result in the desired properties of elastomer tubes. The second EPDM development step therefore focuses on the improvement of the production parameters and the final elastomer material data like hardness, stress relaxation, torsion pendelum and ageing resistance. For this purpose, the content of carbon black has been reduced and the other components have been adjusted with regard to the special requirements. First result is the mixture D-4, the first sample produced in an industrial mixer, which shows a clear progress and already meets some of the requirements. However, further efforts are necessary for the optimization of the elastomer material for the use as fluid tube in the elastomer-metal-absorber, especially with regard to the production parameters, costs and long-term reliability. This work is going on. One problem is inherent with the application of EPDM as tube material. The temperature resistance is restricted to a short term maximum temperature below 160 ~ as the elastomer presents a

1. "phr" means "per hundred rubber", i.e. the number of weight parts of the Idling materialwhich will be added to hundred weightpart of the basis polymer material.

clearly decreasing strength with increasing temperatures and an accelerated degradation at such high temperatures. This has two consequences: 9 The stagnation temperature has to be reduced to a value below 160 ~ Therefore the heat loss coefficient must be higher than of commercial high performance flat plate collectors, which come up to more than 200 ~ at 1000 W/m 2, 30 ~ air temperature and low air speed. The heat loss coefficient a 1 (according to ISO 9806-1, referred to mean fluid temperature) must be higher than or equal to 4.5 W/m2K. 9 The system design has to avoid the simultaneous occurrence of high pressure and high temperature, which is the case for typical closed loop solar systems. 5. EXPERIMENTS O N THERMAL PROPERTIES

During the development steps of the elastomer-metal-absorber, the internal thermal conductivity between solar absorber and fluid, the Uint-value, has been determined by numerical calculations, measurements at single absorber stripes and measurements at complete solar absorbers during the performance test procedure of test collectors. Table 2 presents some of the most important results. profile

type

elastomer tube clip profile Oint q~ st i calcul. m e a s u r . [~ [mini [W/mK] [mm] ~V/m~] tW/m2K] 285

12,0

0,25

12,0

2,0

19,3

20,7

285

12,0

0,78

11,7

1,5

58,1

39,7

250

13,2

0,78

13,2

1,5

55,5

47,3

255

13,0

0,78

13,2

1,5

57,4

52,0

255

13,0

0,75

13,2

1,5

55,1

52,5

290

11,0

0,78

11,7

1,5

57,7

55,3

260

13,0

0,75

13,2

1,5

61,5

50,2

270

12,2 0,7- 1

12,2

1,5

51-61

58,7 ,

275

12,2

12,8

2,0

61,5

1,0

Table 2: Heat transfer capability Uint of different elastomer-metalabsorber constructions, measured and calculated values, tube distance is constant (W = 115 ram, except second line from bottom: W = 135 ram). The five types of absorber profiles presented in figure 2 have been investigated with different construction and material parameters. The calculated and measured internal heat transfer capability of the construction depends, like discussed in chapter 3, on the conductivity of the elastomer k t, the tube diameter dt, the thickness of the tube st and on the contact angle of the clip profile. The tube distance is the same for each construction (W = 115 ram) and the base diameter D is varied only in a small range. With the fin and tube construction parameters, the internal heat transfer capability has been calculated. These theoretical values may be compared with the experimental results. Up to now, the internal heat transfer capability has been increased from 20 W/m2K to 60 W/m2K, resulting in a collector efficiency factor which raised from 0.78 up to 0.92 for typical nonselective collectors (Uloss = 5.5 W/m2K). That means, that the realistic aim of 60 W/m2K has already been achieved, an objective for the future is 75 W/m2K.


If the construction type C (see table 2, Uint = 57.7 W/m2K) would have been equipped with a metal fluid tube instead of the elastomer tube, the collector efficiency factor would be 0.95 instead of 0.91. The difference of 0.04 is the price for an absorber construction with elastomer fluid tubes. Also with future optimized constructions (Uint = 75 W/m:K) this difference will be around 0.03. Table 2 shows, that some calculated values fit rather well to the measured ones, others show significantly lower measured values. The main reason is the thermal contact between the tube and the metal profile. As some of the tubes showed a low flexibility, the contact to the absorber has been reduced, as the uneven and hard tube wall does not touch the whole embracing metal area. Therefore, the flexibility is an important quantity. For this reason, the fluid pressure normally has a positive influence on Uint and it could furthermore be remarked, that a heating-up under pressure also improves the thermal contact. Another important influence may also be derived from table 2. If the outer diameter of the tube is too small in comparison to the clip profile, the measured value of Uint are significantly lower than the calculated ones. Therefore, the outer diameter of the tube should be around 0.5 mm larger than the profile circle. Here the production tolerance has to be taken into account. Up to now, five different test collectors with an integrated elastomer-metal-absorbers have been constructed and investigated. For the collector frame, insulation and cover components of a standard flat plate collector have been used. Figure 7 shows two of the test collectors in front of the institute's building.

87

The first test collector had a copper absorber plate with soldered clip profiles in form of the type A construction. With this collector, the base case investigations and the first measurements with the improved elastomer tubes have been performed. For the base case investigations, the absorber was equipped with a conventional rubber tube (low thermal conductivity of around 0.25 W/InK) and a black painted surface (EMA-1). For the second base case collector (EMA-2) an adhesive selective foil has been used instead of the black painted surface. The first improved test collector (EMA-3) contains the same absorber construction, but the rubber tube has been replaced by a tube made out of the improved elastomer (similar to mixture M-4, figure 5). Again an adhesive selective foil has been applied. The second improved test collector (EMA-4) was produced with a type B absorber construction (roll bended absorber sheets) and an improved elastomer tube with a larger diameter. The absorber coating is again the adhesive selective foil. The diagram in figure 8 presents the measured efficiency curves of the improved test collectors EMA-3 and EMA-4 in comparison to a typical selective flat plate collector as well as to the selectively coated base case test collector.

Figure 8: Efficiency curves of different test collectors compared with a typical selective flat plate collector, test conditions: irradiance level 800 W/m 2, ambient temperature 20 ~ air speed 3 m/s, according ISO 9806-1, referring to mean fluid temperature and aperture area. The resulting low conversion factor of 0.67 of the base case collector with a selective surface (EMA-2) is caused by the high thermal resistance of the rubber hose. For the first test collector with an improved elastomer tube (elastomer mixture M-5) and a selective absorber coating the conversion factor was raised up to 0.78. The conversion factor of the second improved test collector reached 0.81. The differences of this prototype EMA-4 is the use of a more flexible elastomer tube made out of mixture M-7 and the use of an other absorber construction, profile "B". These improvements during the first development steps have shown, that the proposed elastomer-metal-absorber construction gives a thermal performance close to that of typical flat plate collectors with selective metal absorbers. Figure 7: Test collectors in front of the institute's building.


88

However, the conversion factor of an absorber construction with an elastomer tube will remain at least 3 % smaller in comparison to the same construction with a metal fluid pipe. As the reliability of the absorber is the most important condition for any future applications, first reliability investigations have been carded out on the elastomer tube, the absorber construction and the collector prototypes. Burst pressure, long term stability at high temperature and pressure and the torsion vibration properties have been investigated on the tubes, freezethaw-cycles have been performed with different types of absorber stripes and exposition tests on a complete collector prototype have been carded out. The results showed, that the existing problems should be solvable.

6. APPLICATION AS FACADE AND ROOF ELEMENT Industrially produced roofs and faqades often consist of corrugated metal form sheets made out of steel or aluminium. These roof or faqade constructions are widely used for industrial, public or residential buildings. The elastomer-metal-absorber concept will transform these metal form sheets into uncovered or transparently covered roof and faqade absorbers by integrating an appropriate clip profile into the form sheets during the production process. The elastomer tube can then easily be clipped into these profiles after the installation of the roof or faqade. Figure 9 shows the conversion of a typical metal form sheet (presented here as insulated sandwich plate) into an unglazed or transparently covered solar collector.

" (

"

( 8

2( f'

( 8

2 ( "4

(

2(

" I

Figure 9: Steps from a metal roof and fafade element to a solar collector. The first step of the conversion is the integration of the clip profile into the metal form sheet during the roll form process. The form sheet is covered with a paint of high solar absorptivity, with or without selective properties. The sheet will be mounted on the roof or faqade by normal roofing or metal processing companies. The optical and technical properties will be the same like for normal metal roofs. The second step is the integration of the elastomer tubes into the form sheets. The elastomer tubes will be connected via the manifold tubes to the solar system. Thus, an uncovered absorber results with only little extra costs, where the technical properties of the metal roof or facade remain unchanged. As an additional option for systems with higher demand temperatures a transparent cover may be added, using single glass panes or transparent plastic covers. By this way also low cost glazed collectors may be produced, which are specially suited for large systems.

The idea of this building integrated collector type has the following advantages: 9 Metal form sheets are a common and well proved technology. 9 The transformation into the elastomer-metal-absorber does not affect the reliability of the original roof resp. faqade. 9 The additional effort to transform metal roofs into unglazed absorbers seems to be very low, on the other hand, the metal roof and faqade elements gain by their new property as active solar absorber further attractivity. 9 The extension to glazed collector roofs for higher demand temperature is possible. 9 The integration may be performed with a high aesthetical quality and architectural acceptance. Typical examples for a future application of this concept are buildings with a high demand of low temperature heat, e.g. swimming-bath and sports halls, hospitals etc. for glazed collector constructions and outdoor swimming-pools and heat-pump systems for the unglazed absorber type. Domestic hot water and residential room heating purposes may also be taken into account. Due to the very low additional costs expected for the transformation of the metal building envelope into a glazed or unglazed solar collector, this concept has the potential to result in new solar applications with a high economic benefit.

7. DEVELOPMENT PROJECT- STATF.,-OF-THE-ART Despite of the encouraging results of the first development steps, this absorber type is not available up to now. Open questions are mainly the production and installation technology, the long term reliability and the long term thermal performance. A research and development project, funded by the European Commission, has started to develop and investigate the integrated elastomer-metal-absorber in roof and facade metal form sheet elements and possible heat use applications. The main tasks within this cooperation between industrial partners coming from various activity fields and research institutions are: 9 further improvement of the elastomer material with special regard to heat conductivity, mechanical strength and durability, as well as the production of appropriate elastomer tubes, 9 development of absorber constructions with focus on production parameters, thermal performance and reliability, 9 construction and assessment of test collectors, determination of thermal performance and reliability characteristics, 9 development of different solar system concepts, assessment of collectors in test systems, comparison and extrapolation. The first results of this project are encouraging: 9 An improved elastomer mixture with special regard to the industrial produceability has been developed, from which 9 first prototypes of an appropriate elastomer tube have been produced in an industrial extrusion machine. 9 The construction of the form sheet elements for the integration into metal-roofs has been performed, the tools for the roll form machine are ready and the first absorber form sheet elements are in production. 9 A simulation tool for the elastomer-metal-absorber has been developed, first different heat use concepts have been worked out and simulated. However, still a couple of problems exist which have to be solved on the way to an industrial product with high performance and reliability.


Some items to be worked out are in the fields of: 9 assessment of the absorptive surface and its stability, 9 design of the hydraulic system and the manifolds, 9 development of the connection technique between tubes and manifold, 9 and system operation and security technique, with special regard to the fact, that water will be used as heat transfer fluid. It has to be pointed out here, that at the moment it is planned to transform the whole roof area into an elastomer-metal-absorber, i.e. the normal application are large collector areas. For the specific problems arising from this aspect, the development has to go on over intermediate stages like medium sized pilot and demonstration plants. 8. C O N C L U S I O N AND O U T L O O K The principle of the elastomer-metal-absorber with its clip prof'lle contact opens up new possibilities with regard to the heat transfer fluid, the collector and system design and the architectural integration. The development steps have shown that the proposed elastomer-metal-absorber construction already has a thermal performance close to that of typical flat plate collectors, with only a slightly lower conversion factor. The essential results up to now are the increase of the thermal conductivity of the elastomer material from 0.25 W/mK up to 1.0 W/mK, which in combination with an optimized absorber construction leads to an internal heat transfer coefficient of at least 60 W/m2K, a value comparable to standard flat plate collectors. The existing reliability problems seem to be soluble, the first results of the actual development project are encouraging. The special attraction of this building integrated design is given for the following reasons: 9 high expected cost reduction for unglazed absorbers or glazed collectors, 9 significant reduction of energetic amortisation periods, 9 well suited solution for repair or recycling, 9 enlargement of the solar market by new manufacturers and solar systems, especially in large commercial and public buildings, as well as in residential buildings, 9 and improvement of architectural acceptance by the high degree of building integration.

Acknowledgements-The work is funded partially by the European Commission, within the project ,,Faqade and Roof Integrated Solar Collectors with a Combination of Elastomer Tubes and Metal Form Sheet Elements", contract no. JOE3-CT98-0236, organized in the framework of the Non-Nuclear Energy Research and Technological Development Programme JOULE Ill.

89

NOMENCIJ~TURE a1

linear collector heat loss coefficient, referred to Tm (W/m2K) 0t~luid convective heat transfer coefficient between the inner tube wall and the fluid (W/m 2) D base diameter (projection width of visible tube surface)

(mm) df dt dt,i dt,m F' G 11 kf kt tp sf st Tabs Tbase Tm Tt,out Tt,in Ufm Ubase Utube Ueonv

diameter of the clip profile (mm) tube diameter (ram) inner tube diameter (mm) arithmetic mean of outer and inner tube diameter (ram) collector efficiency factor (-) Solar irradiance (3br/m2) collector thermal efficiency, referred to T m (-) thermal conductivity of the fin material (W/mK) thermal conductivity of elastomer tube material (W/mK) contact angle of the clip profile (o) thickness of the fin (mm) wall thickness of the tube (ram) mean temperature on the absorber fin (~ temperature on the absorber base (~ mean temperature of heat transfer fluid (~ temperature on the outer surface of the tube (~ temperature on the inner surface of the tube (~ internal heat transfer conductivity of the fin (W/m2K) internal heat transfer conductivity of the base (W/m2K) internal heat transfer conductivity of the tube (W/m2K) convective heat transfer conductivity between tube wall and fluid (W/m2K) Uint internal heat transfer conductivity of absorber construction (W/m2K) Uloss overall heat loss coefficient of the collector, referred to T m (W/m2K) W tube distance (ram) REFERENCES Bartelsen B., Rockendorf G. and Vennemann N. (1996) Development of an Elastomer-Metal-Absorber for Thermal Solar Collectors. In Proceedings of the EuroSun '96, 16-20 September, Freiburg, Germany, pp. 495-499, DGS-Sonnenenergie Verlag GmbH, Mtinchen. Rockendorf G., Falk S.,Wetzel W. (1996) Bedeutung und Bestimmung des Kollektorwirkungsgradfaktors bei Sonnenkollektoren; 6. Symposium thermische Solarenergie, 08-10 May, Staffelstein, Germany, pp. 196-201, O T H e.V., Regensburg. Duffle J.A. and Beckmann W.A. (1991) Solar Engineering of Thermal Processes; 2n d edn, pp. 268-276, Wiley-interscience Publication; New York. B6kamp K., Vennemann N., Wallach J., Bartelsen B. and Rockendorf G. (1997) EPDM Compounds with Improved Thermal Conductivity for Thermal Solar Collectors. In Proceedings of the International Rubber Conference, 30 June - 3 July, Ntirnberg, Germany.

USES Solar World Congress 1999, Volume Ill

90

SOLAR ABSORBER SYSTEM FOR PREHEATING FEEDING WATER FOR DISTRICT HEATING NETS Klaus Vajen, Marcel Kdimer FBPhysik, Universit~itMarburg, D-35032Marburg, Germany phone -H-49/6421/28-4148, fax ++49/6421/28-6535, [email protected] Ralf Orths Wagner & Co Solartechnik, Ringstr. 14, D-35091 C61be, Germany

Erkin K. Boronbaev, Astra Paizuldaeva Kyrgyz State University of Construction, Transport and Architecture, 34 b Maldybaevstr., KS-720023 Bishkek, Kyrgyzstan

A b s t r a c t - EPDM-absorbers, made of artificial rubber and well-known in Central Europe for heating swimming pools, have been installed to preheat domestic water in a heat and power plant in Bishkek (Kyrgyzstan). Measurements were carried out during the summer 1998. The special construction of the district heating net and the climatic conditions of Central Asia lead to a favourable environment for the utilisation of solar thermal energy. Fluid temperatures nearly always far below ambient temperature result in convective heat gains instead of losses. Collector "efficiencies" far above 1 as well as nightly heat gains were measured. Calculations of solar energy prices lead to about 6 Euro/MWh useful energy.

1. INTRODUCTION The heat supply of cities in the former Soviet Union usually is provided by one or more district heating (and power) plants. So it is in Bishkek, the capital of Kyrgyzstan. 350.000 inhabitants receive domestic hot water and energy for room heating from the central Heat and Power Plant of Bishkek City. The district heating net (fig.l), however, shows some differences to common Central European technology. In Bishkek (as in many other cities of the CIS) one finds an open circle system: domestic hot water is taken by the consumers directly out of the net without any heat-exchanger coupling. Thus in Bishkek the amount of 3000..4000 ma/h water has to be refilled into the net. This is carried out at one central place. Cold water is taken from the ground and artesian sources and led to boilers which heat it to the required temperature of 60~ Due to Kyrgyzstan's climatic conditions (Central Asia, latitude 43 ~ north, comparable with Rome), altogether these are nearly ideal conditions for the implementation of solar thermal systems. So it stood to reason to preheat the cold water directly by uncovered solar collectors. In Central Europe they are wellknown for swimming-pool heating. 2. MEASUREMENTS In June 98 a test plant of a 50 m2 EPDM absorber field was installed on one roof of the District Heating Plant of Bishkek City. The absorber had an inclination of 4 ~ to west. The water was taken from the pressure pipe (behind the pumps) and led back to the pressure-less pipe, so no extra pump was necessary to force water circulation. The collector flow rate was varied by a hand valve and also unspecified altered due to pressure

changes in the net. The variation width was 10 l/m2h up to 1201/m2h. Measurements took place from June to October 1998. Apart from the flow rate, the global, diffuse and long wave radiation, Tin, Tout, Tamb, humidity and the wind-speed were measured, automated by a computer system Mean values of up to 4000 single measurements were stored on the harddisk minutely. The system worked nearly without problems during the summer.

Fig. 1. Simplified scheme o f the distn'ct heating net in mar absorber

,

~oreoe

mdim~

I1[111111111 ii

domestic hot wmeer

m II

c~a w ~ r 0 ~ ' Q

,.,._ r

:l

..g

In

(60~ C in summer)

an U

|

| ~

l

=

~ m

=

=

n

=

i

m ~ , ,

= u

=

=

m

= , m =

=

m m

=

m

(only in winter)

Bishkek. Huge cold- (not included in the figure) and hotwater storages lead to nearly constant cold water and heat demands. In order to use an existing pump, the back flow from the collector was connected before the turn-off o f the collector forward flow. The flow rates were about 3000 m3/h through the pump and 5 mS/h through the absorber. 3. RESULTS Data from June, 13 to August, 10 were taken for the following evaluation. The cold water inlet temperature was always about 12 to 13~ The ambient temperature, however, was nearly always higher, even at night. This leads to the


unusual behaviour, that the net energy balance of the absorber shows profits from the surroundings instead of losses, see fig. 2.

Fig. 2. Example o f the measured temperature courses. During the selected days the ambient temperature was always even higher than the outlet temperature. O f course, this temperature difference depends on the collector flow rate. I f the dew-point is above the inlet temperature, at least on a part o f the absorber condensation occurs.

91

independent of the flow rate. The average heat gains were 0.7 kWh/m2night or 80 W/m2, respectively (see fig. 4). With a further developed absorber model, taking into account also condensative heat gains, the measured results could be reproduced. More detailed results of the measurement and the modelling will be published later.

Fig. 4. Measured hourly mean values o f the nigh@ useful power gains (June to August 1998). 4. CONCLUSIONS

Fig. 3. Measured hourly mean values o f the collector efficiency during the day, which is nearly always > 1 at high flow reates (June to August 1998). Since the efficiency of an "'ordinary collector" depends on the wind speed an "'efficiency field'" is sketched in. In fig. 3 the collector efficiencies during the day (this means 5.30 to 20.30 h local time) are shown. Best results could be achieved at flow rates higher than 40 l/m2h. The high ambient temperatures lead to collector "efficiencies" more than 1. The highest values could be observed in the early morning and the late afternoon. Note, that unlike usual collector characteristics nearly all dots can be found in the 2~a quadrant of the coordinate system. In contrast to the figure in the "Book of Abstracts" fig. 3 does consider partly shading of the absorber field. In contrast to the collector efficiency at day, during the night (20.30 to 5.30 h local time) the power gains were nearly

The yearly heat gains of an uncovered collector connected with the district heating net in Bishkek can be estimated to be higher than 1100 kWh/m2. So the absorber heat gains can be expected to be more than twice as high as common for collector systems in Central Europe, furthermore the installation costs of the absorbers are very low. With the results measured at the test-plant solar energy prices of about 6 Euro/MWh useful energy can be expected for an absorber area > 1000 m 2. This is below the today's prices of fossil fuels on the world market. So the absorber system in Bishkek could be a solar thermal installation able to compete economically with all conventional energy sources. The estimated technical potential only in Bishkek is higher than 40000 m 2 absorber field. An installation (abt. 1000m 2) in Bishkek is under consideration. ACKNOWLEDGEMENTS The authors would like to express their sincere thanks to the following persons and institutions for financial and logistic support: Unversiti4t Marburg, Wagner Solarteclmik in Crlbe (Germany), International Bureau of the German Ministry of Education and Research, Heat and Power Plant of Bishkek City and to the Embassy of the Federal Republic of Germany in Bishkek.


92

STATISTICAL ANALYSIS OF SOLAR COLLECTOR TEST RESULTS IN VIEW OF FUTURE CERTIFICATION Emmanouil Mathioulakis, Kostantinos Voropoulos and Vasilis Belessiotis Solar & other Energy Systems Laboratory, NCSR >, 15310 Ag. Paraskevi Attikis, Greece Tel. +301 6544592, Fax +301 6544592, E-mail: [email protected] Abstract - This paper deals with the need to develop certification schemes for solar products as a means for a further promotion of solar energy applications. It also examines the ways in which these schemes could be implemented. More specifically, an analysis of the results of solar collector efficiency testing is presented, aimed at depicting the existing situation. A methodology for the exploitation of these results is proposed, leading to a realistic approach to the criteria that could be used in future certification schemes

1. INTRODUCTION Quality is well known to be the most essential factor for the survival of every commodity produced and offered for sale in today's strong competitive market. The quality level of a product can be proved with its certification and marking, based upon specific evaluation criteria and generally accepted procedures. Moreover, it is commonly accepted that the absence of objective and undisputed mechanisms for the assessment of solar collectors and the qualitative characteristics of the systems is one of the main obstacles to the further promotion of these products. These mechanisms should be based on results from tests undertaken in accordance with existing national, European or international testing standards, and should permit, within the framework of a certification scheme, the classification and marking of products depending on their performance. A certification scheme should not be based on abstract concepts. It can be realistic only if it takes into account the actual situation of the productive sector. In this regard, the exploitation of the existing test results could be an essential aid in the development of a certification scheme. In this paper, an attempt is made to relate an evaluation mechanism that could be used in a solar collector certification scheme with the experience gained from their testing to date, particularly concerning efficiency. 2. CERTIFICATION OF SOLAR ENERGY PRODUCTS A survey of not only the Greek but also the international experience from the existing applications of solar thermal energy systems leads to two, at a first sight, contradictory conclusions (EU, 1996). On one hand, the technologies used nowadays are mature, which means that there are no critical technological matters that could obstruct the dissemination of such systems. On the other hand, practical applications still present important efficiency and reliability problems, resulting in the fact that that there has been a reduction in the faith which the potential users have in these products (Mathioulakis and Belessiotis). Several explanations can be given to this situation, one being, without doubt, the quality of products. There is a distance between the technological know-how and the products available in the market. This is due, amongst others,

to the new reality of flee movement of goods and to the inadequate quality control. Because of this, the issues that concern the evaluation of performance and the marking for quality of solar thermal collectors are of great importance. Certification of solar energy products can become an essential tool for dealing with the problems mentioned above. It should not be forgotten that the quality and reliability of the proposed solutions are of key importance when trying to formulate a favorable legal and financial framework for the promotion of solar energy usage. The certified quality of the products is a motive for consumers who are now able to make their choices based on objective criteria. It also constitutes a motive for manufacturers to improve their products. It facilitates a more realistic approach to the issues of financing and economical effectiveness of solar energy exploitation systems. Finally it contributes to market control since it introduces more transparency in the evaluation of existing technologies. Confrontation of this problem varies from country to country. In some cases, certification schemes are applied in connection with support actions, such as subsidies. However, market internationalization within the last few years is leading in the direction of a common international practice, which is facilitated by the gradual development of common standards concerning testing methods. A typical case is the that of the new European Standards in the field of solar energy products. Their basic target is the technical harmonization of testing methods in all European countries and their use as a base for certification. The new European Standards adhere to the testing methods of ISO, with minor changes, and add new Standards of (CEN, 1998). The certification procedure that could be developed on the basis of the new European Standards is shown in figure 1. All certification schemes concerning solar energy products are optional. But in some cases they turn out to be compulsory in praxis when they are used as a criterion for subsidizing the installations. The gradual harmonization of the certification schemes used in several countries is now a realistic target. However it requires the harmonization of testing methods, which is still in progress, and of the way in which the products are classified with respect to their performance.


93

where no is the maximum collector efficiency and U0 the collector heat loss coefficient.. In some cases, a three-parameter model expresses the collector behavior more clearly, especially for collectors with non-selective absorbers. However, for the scope of the present paper and the homogeneity of the results, the model of equation 1 is used. : I n figure 2 the determined factors no and U0 of the instantaneous efficiency curve are plotted. On this figure the are distinguished, showing the separation between the technologies used and the direction of improved collector efficiency. It is shown from the graph that the collectors with selective absorbers are clearly separated from the rest with respect to their energy characteristics and especially the heat loss coefficient. Collectors with semiselective absorbers have improved performances compared with those which are black-painted but are till worse than the selective ones. Figure 1: General layout of certification procedure

The efficiency of the collector, which is a basic characteristic of quality, is examined in the next paragraph. The manner in which collector efficiency is introduced into the general evaluation scheme for certification is therefore of great importance. Furthermore the same methodology can be applied to the remaining qualitative characteristics of the collector or of the solar system. 3. COLLECTOR EFFICIENCY TEST RESULTS Evaluation can be reliable if the real situation of the products available in the market is taken into account. This situation is depicted in test results. Results from the testing of about 100 collectors, tested over the last 7 years in the Solar & other Energy Systems of NCSR "Demokritos', form the basis of the analysis presented. Tests were carried out according to the ISO 9806-1 Standard [ISO, 1994]. These collectors constitute a rather broad sample of the product, with regard to construction type, absorber surface treatment and year of production. Thus, they are representative of the majority of the products produced and used over the last ten years in the market. Through a statistical analysis of the determined factors of the instantaneous efficiency curve of the solar collectors, and also of the instantaneous values of efficiency in specific ordinary operating conditions, several important results can be derived. Three different types of collectors were separated, concerning the absorber surface treatment: black-painted, semi-selective and selective collectors. We should keep in mind that according to the ISO 9806-1 Standard, the energy characterization of the collector is achieved by determining, from testing data, the coefficients no ~:at U0 of the collector efficiency equation (Duffle and Beckman, 1991): n = no- U0 I m

(1)

Figure 2 - Maximum efficiency nO and heat loss coefficient U0 of tested collectors +: Black paint, x: semi-selective, o: selective It should be noted however, as shown in figure 2, that the separation between the three collector types is not complete, given the fact that the respective overlap. This conclusion is very important since it demonstrates not only that the final result surely depends on the technologies used, but also that the integration and the correct design of the collector play an important role as well. Moreover it shows that the only reliable method for collector energy characterization is testing, through which the whole technological, design and construction particularities of the specific product are incorporated in the final result. The same conclusion, i.e. that a product should not be seen as a > but as a whole (seen only by testing), is also depicted in figure 3. The results given in this figure concern systems of the same type (thermosiphon-type) and contain the values of their energy output calculated from testing according to ISO 9459-2 (ISO, 1995b), for the same climatic conditions. It is observed that the type of the collector used in the system is important, but it is not the only factor that plays important role. The advantages of the selective absorbers can be easily lost due to wrong choices

94


when integrating a collector with such an absorber into the system.

Figure 5: +: determined values of n, solid line: normal distribution

Figure 3 - Energy output for different systems as a function of the type of collector surface From the above, it can be concluded that the energy characterization of the collector, within the framework of a certification scheme, must be referred to test results, i.e. the coefficients no and U0. However, it may be better, for reasons of simplicity, for the evaluation to be based on the instantaneous efficiency n calculated under specific climatic and operating conditions. These conditions could be, for example, a global solar irradianee of 800 W/m2 and a temperature difference AT = Tm-Ta of 30 K. Figure 4 gives the results for the three types of collectors concerned.

From the graph of figure 5, the percentage of collectors whose performance lies above a certain range of values can easily be determined. Thus, it facilitates the correlation of the efficiency evaluation criteria in the framework of a certification scheme with the quality of products that are actually available in the market. It must be stressed that for a certification procedure to be objective, the physical quantities used for product evaluation should be accompanied by the confidence interval, within which the values of the characteristic quantities lie. Having this in mind, the introduction of procedures for calculating the uncertainties of results in testing standards, and especially those concerning efficiency, would be very useful (ISO, 1995a). 4. Conclusions

Figure 4: Instantaneous efficiency n for standard conditions (G=800 wm2, AT=30 K) Presentation of the instantaneous efficiency values of all collectors in a probability plot shows the distribution of the determined values over the range which they appeared (figure 5). This curve, with a tendency towards normal distribution, contrary to the similar curves of no and U0, results in the fact that a kind of collector rating could be based on the instantaneous efficiency. The classification criteria can be determined by the probability distribution curve and the quality level that is desired to be fulfilled.

Quality is well known to be the most essential factor for the survival of every commodity produced and offered for sale in the today's strong competitive market. The level of the quality of a product can be assessed by its certification and marking, by implementing specific evaluation criteria and commonly accepted procedures. This also applies to solar collectors, a product which uses the flee and abundant solar energy for a large range of applications, while simultaneously protecting the environment. A methodology according to which the marking of solar collectors will be made, based on real data, is the proper tool for the upkeep of a high quality level of solar collectors and also for the their continuous improvement. For example, the analysis of the determined factors of the instantaneous efficiency curve of the solar collectors, and also of the instantaneous values of efficiency in specific ordinary operating conditions, should be used for the evaluation of their performance. As a result, this could be the start, together with other quality indices, of the development of a certification scheme, in view of the subsequent marking of solar thermal products.


NOMECLATURE G m n

Ta Im

Tm Tm Tout AT

global incident solar irradiance, W m -2 mass flow rate through the collector, kg s1 collector instantaneous efficiency ambient air temperature, ~ reduced temperature difference I m =(Tm-Ta)/G, KW-1 m2 temperature of water in collector inlet, ~ mean temperature of water inside collector Tm---( Tin + Tout)/2, ~ temperature of water in collector outlet, ~ temperature difference AT=To~t-Tm,K

REFERENCUES CEN (1998), >, CEN ed., Brussels Duffle J.A. and Beckman W.A. (1991), Solar engineering of thermalprocesses, 2nd edn, Willey, New York. EU (1996), Sun in Action, Office for Official Publication of the European Communities, Luxembourg ISO (1994), Standard 9806-1. Test methods for solar collectors - Part 1: Thermal performance of liquid heating collectors including pressure drop, ISO ed., Switzerland. ISO (1995), Guide to the expression of uncertainty in measurements, ISO ed., Switzerland. ISO (1995), Standard 9459-2. Solar Heating- Domestic Water Heating Systems- Part 2: Outdoor Test Methods for System Performance Characterization, ISO ed., Switzerland. Mathioulakis E. and Belessiotis V., Active solar systems Review of technologies and applications in Greece, In Proceedings of NTUA National Congress >, 30 November- 2 December 1998, Athens, Greece (in Greek)

95


96

THERMAL AND ELECTRICAL YIELD OF A COMBI-PANEL Herman A. Zondag~ Douwe W. de Vries and Anton A. van Steenhoven Department of Mechanical Engineering, Eindhoven University of Technology, P.O.Box 513, 5600 MB Eindhoven, The Netherlands, Tel.: + 31-40-2472726, fax: + 31-40-2433445, [email protected] Wim G.J. van Helden, Netherlands Energy Research Foundation ECN, P.O. Box 1, 1755 ZG Petten, the Netherlands Ronald J.C. van Zolingen Shell Solar Energy BV, P.O. Box 849, 5700 AV Helmond, the Netherlands Abstract - A first, non-optimised prototype of a combi-panel was built of a PV-laminate and a sheet-and-tube absorber. The thermal efficiency at zero reduced temperature was found to be 54%, along with 8.5% electrical efficiency. The results of the measurements were used to verify the results of the simulations. It was concluded that the simulations and the measurements corresponded sufficiently well. Then, the simulations were used to find the annual efficiency of a PV/T-system that was used for hot-water production in a Dutch household, for which 33% thermal and 6.7% electrical efficiency was found. Finally, the simulations were used to quantify the contribution of the various loss terms to the reduction in thermal efficiency of a PV/T-system with respect to a thermal collector.

1

INTRODUCTION

A combi-panel consists of a PV-laminate that functions as the absorber of a thermal collector. In this way a device is created that converts solar energy into both electrical and thermal energy. The main advantages of combi-panels are: 1. An area covered with combi-panels produces more electrical and thermal energy than a corresponding area partially covered with conventional PV systems and partially filled with conventional thermal collectors. This is particularly useful when the amount of space on a roof is limited. In addition, installation costs are reduced. This will become increasingly important in the future when the price of PV will be reduced. 2. Combi-collectors provide architectural uniformity on a roof, in contrast to a combination of separate PV- and thermal systems. 2

The efficiency of the combi panel was measured and compared to the efficiencies of a conventional sheet-and-tube=type thermal collector and a multi=crystalline silicon PV=panel of the same length and width, which were positioned next to it in the test rig. A photograph of the test rig is shown in figure 2. The original collector surfaces were somewhat larger than the PV= laminate. In order to create similar areas for the PV laminate, the thermal collector and the combi=panel, the absorbing surfaces of the latter two were partly covered with insulation that had a reflective aluminium top layer. In figure 2 these covered parts appear as the white areas around the collector and the combi-panel. The uncovered parts have an area of 0.94 m 2 each.

SYSTEM

In order to quantify the efficiency of a PV/T-collector, an experimental prototype was built at the Eindhoven University of Technology. This was a non-optimised first prototype, that was built in order to be able to validate the simulations. The prototype was constructed by connecting a conventional PVlaminate, containing multi-crystalline silicon cells, to the absorber plate of a conventional glass-covered sheet-and-tube collector, as shown in figure 1. The panel was then integrated into a test rig on the roof of the department of Mechanical Engineering at the Eindhoven University of Technology.

Fig. 1. Cross-section of the combi panel

Fig. 2. The test rig. Left to right: a conventional thermal collector, the combi panel and a conventional PV-laminate. The system consisted further of a water tank of 130 litres. The water was drawn fi'om the tank into the thermal collector and the combi-collector by a NKF Verder ND 300 KT 18 diaphragm pump. The construction was such that the water that was heated by the system could either be returned to the tank or could be discharged on the sewage system in order to keep the water temperature in the tank at a constant value. In the latter case the water level in the tank was kept constant through a tap that was connected to the water supply of the building. The water flow through the combi-panel and the conventional thermal collector were measured independently with two rotary


piston KENT PSM-LT PL 10 water volume meters. The volume flow was measured by dividing the counted amount of litres by the measuring time. The wind speed was measured with an EKOPOWER MAXIMUM cup anemometer. The irradiation was measured with a Kipp & Zonen CM 11 pyranometer. The temperatures of the PV-laminate, the combi-panel laminate and the collector absorber as well as the in- and outflow temperatures of the collector and the combi-panel were measured with thermocouples type K which were calibrated to an accuracy of 0.2 K. The thermocouples, the pyranometer, the anemometer, the two water meters and the electrical output of the combi-laminate and the PV-laminate were read out by a DORIC digitrend 220 datalogger. The time between two measurements was typically 11 seconds. The PV laminate was a standard Shell Solar PV-laminate consisting of 72 10xl0 cm2 EVA encapsulated square multi-crystalline silicon cells with a low-iron glass front and an A1/tedlar film at the back. The cell efficiency under STC is typically 13%. The laminate efficiency at 25 ~ is 9.7%. 3

97

Both the electrical yield and the thermal yield are lower than found for the conventional collectors, as expected. However, the results show that two combi-panels together produce more energy per unit area than one PV-laminate and one thermal collector next to each other, which makes them interesting for solar energy production.

C O L L E C T O R EFFICIENCY CURVES

3.1 Measurements The thermal efficiency was measured as a function of reduced temperature. For these measurements, a mass flow of 76 kg/(m2 hr) was used. The conditions for the measurements were: 1. During a time span of 15 minutes the radiation is at least 750 W/m 2 and its value does not vary more than 100 W/m 2. 2. The fluid inlet temperature and outlet temperature do not vary more than 0.2 K during the measurement. 3. The flow rate for the collector and combi-panel is around 20xl 0 ~ m3/s and varies not more than 1.4xl 0 "~ m3/s. 4. The wind speed during the measurements does not exceed 2 m/s. In order to check if the restrictions mentioned above are sufficiently strict a set conditions of a somewhat less restrictive nature was applied. It was found that the results did not change significantly. The thermal efficiency is calculated from quantities averaged over 15 minutes. The thermal efficiency of the combi panel and the thermal efficiency of the conventional collector are presented in figure 3. The electrical efficiency of the combi panel and the PV panel are presented in figure 4.

Fig. 4. The electrical efficiency of the PV-laminate (,) and the combi-laminate (o).

3.2 Simulations The thermal efficiency is simulated assuming thermal equilibrium in the various layers of the combi-panel and the thermal collector. The relations describing the radiation and convection heat transfer are put in local heat balances for all layers in the combi panel. The sheet and tube type combi-panel is considered as one long straight tube, which implies that symmetry is assumed with respect to the centreline between two successive pipes. The heat flow in the direction perpendicular to the flow direction is calculated with a relation obtained from the well known Hottel-Whillier model for thermal collectors, that is based upon this same assumption of symmetry. It is described extensively by e.g. Duffle and Beckman (1991 ). The Hottel-Whillier model leads to a temperature distribution between two tubes in a sheet-and-tube collector that is given by

T ( x ) - Tarot, - V,~I / h / = T b -Ta,,a, - Z a I / h

I

cosh(mx)

(1)

cosh(m(W-D)/2)

in which x,~ is the transmission-absorption coefficient, I is the irradiation, hi is the heat loss coefficient, T ~ , is the ambient temperature and Tb is the temperature at the absorber surface directly above the tube, W is the distance between the tubes and D is the tube diameter. A typical temperature distribution between the collector tubes is shown by figure 5.

Fig. 3. The thermal efficiency of the thermal collector (x) and the combi-laminate (both with (+) and without (o) electricity production).

Fig. 5. The temperature profiles of the laminate and the absorber between the tubes of a sheet-and-tube collector.


98

The coefficient m in formula 1 determines the flatness of the temperature profile and is given by

m

2

= ~

F = tanh(m(W-D)/2) m ( W - D ) /2

(3)

For a conventional thermal collector, the heat collection efficiency factor is then defined as

F" =

l/hi

1

(4)

~bo.d

1

Lhl (D + (W - D)F) + Abo.db+ff.Dhf

q - WF" ('~al - hl (Zw - Za ) )

(5)

"~a,eff"="~ T-~elO-O.OO5(Tlam-25~

3.

4.

hi

1 "~

1

Whca

0.9 0.8 9tD -eo 0.7 tD

"•

0.6

"~ 0.5 I

0.4 0

I

I

I

0.005 0.01 0.015 0.02 reduced temperature in KmVW

I

0.025

(6)

For xa a value of 0.744 was found from a simulation of the optical characteristics of the PV laminate. For the transmission of the glass cover x a value of 0.92 was applied, based on general low-iron glass transmission data. In the equation for m an additional term appears, due to the fact that the silicon cells provide a parallel channel for heat conduction, along with the conduction sideways through the copper absorber plate.

m2 =

1

F is the fin efficiency factor and hca is the heat transfer coefficient between the cells and the absorber If hca is small, the temperature gradient between the laminate and the copper absorber will be large and a large heat loss to the ambient will occur, which will reduce the thermal efficiency. 5. Finally, a PV-laminate is not spectrally selective, so the emission coefficient was changed from 0.12, which is a typical value for a spectrally selective surface (e.g. Duffle and Beckman (1991)), to 0.9. The full set of equations provides a matrix, which was solved by a matrix solving procedure of MATLAB. This results in a set of efficiency curves. In figure 6 the calculated efficiency curves are presented together with a least-squares fit of the measurements that were presented previously in figure 3.

in which Ill is the heat loss at the top of the laminate and hf is the heat transfer coefficient to the water in the tubes. The useful energy gain per unit tube length is given by formula 5.

Tw is the temperature of the water. In order to account for the special characteristics of the combipanel, three equations had to be modified with respect to the equations for a conventional thermal collector. 1. Due to conservation of energy, the solar energy that is converted to electricity cannot be converted to thermal energy anymore. Therefore, in the heat equations, xa should be replaced by its effective value.

(8)

-i-~ Lh l ( D + ( W - 9 ) 1 7 ) + ~ff.Dhf

(2)

given by

2.

l/hi W(

h~

The temperature gradient across the absorber drives the conductive heat transfer to the collector tubes. At the same time, the high temperature between neighbouring tubes causes additional losses, which means that not all the heat can be collected. This is expressed in the fin efficiency factor F that is

W(

F" =

(7)

Only the effects of the silicon and the copper are expressed in the equation, since the heat conduction through the EVA and the glass are much smaller than these. If the heat conduction through the silicon is large, the temperature profile across the combi laminate will be flatter than if the heat conduction is small. In the equation for the heat collection efficiency, an additional term appears due to the heat resistance hca between cells and absorber. The bond conductance can be neglected due to the high silver content of the bond

Fig. 6. The simulations of the thermal efficiency (dashed) compared to the least squares fit of the measurements (solid) for a conventional thermal collector and a combi-panel either or not producing electricity. The figure shows a reasonably good agreement between the simulations and the measurements, although the difference between the curves is in the range 0%-4%, which is somewhat larger than the experimental inaccuracy, which was found to be around 1%. The differences still present might be due to a slight overestimation of the optical efficiency or to heat loss to the sides of the copper absorber (the parts which are covered by insulation in order to keep the area of the PV laminate and the eombi absorber-plate equal; see figure 2). In addition, the sky temperature was not measured. In the simulations, it was assumed to be equal to the value for a clear sky. This could also account for a part of the difference. The clear-sky temperature is calculated from the formula

T,ky = 0.0552Ta~

(9)

3.3 Estimating the loss terms Next, the simulations were used to obtain information about the loss mechanisms in a eombi-panel. Figure 7 shows the magnitude of the radiation loss, the convection loss and the back loss. Together with these losses, the thermal and the


electrical efficiencies are indicated. Finally, the straight line on top is the sum of all these terms, which is equal to the transmission-absorption coefficient of the combi-panel, as expected. The calculation was done by setting Ta~ = 20 ~ and I = 800 W/mEand increasing the inflow temperature, which is of some importance because these settings determine the PVlaminate temperature, which determines the electrical efficiency and which is by itself not a function of reduced temperature. The relative magnitude of the radiation and convection losses depends on the sky temperature. The calculation was done for a clear sky using formula 9.

99

seems to have a substantial effect on the slope of the thermal efficiency curve.

Fig. 8. The efficiency curves versus reduced temperature, successively removing the special features of the combi-panel. From low to high: (1) combi-panel, (2) optical efficiency enhanced, (3) heat transfer enhanced, (4) spectral selectivity enhanced, (5) heat conduction sideways through silicon removed, (6) no electricity produced. 4 Fig. 7. The loss mechanisms in the combi-panel as a function of reduced temperature (solid lines); (1) back loss, (3) convection loss and (4) radiation loss. The dashed lines indicate the electrical efficiency (2) and the thermal efficiency (5). The dash-dot line (6) represents the sum of all these terms, and is therefore equal to xa. With respect to the reduction in the thermal efficiency of the PV/T-system in comparison to the conventional thermal absorber, simulations were performed in which the special features of the PV/T collector were successively left out. These features are 1. A lower optical efficiency of 0.744 for the PV laminate applied in the first prototype, instead of 0.89 for a conventional thermal absorber. This is particularly important for long-wavelength irradiation 2. A smaller heat transfer between the absorber (the PV laminate) and the water, as indicated by formula 8 (see above). The value of laea that was found from measuring the temperature difference across the combi-absorber, was approximately 45 W/m2K. Due to this heat resistance, the absorber surface is relatively hot and therefore thermal losses are enhanced. 3. The PV-laminate is not spectrally selective, since glass has a high emission factor in the inflated. This changes the emission of the absorber from 0.12 (for a spectrally selective absorber) to 0.9. This strongly increases radiation losses from the absorbing surface. 4. Due to the additional heat transfer sideways through the silicon (which provides a thermal path parallel to the copper absorber), the heat loss from the collector surface was slightly reduced. However, this effect is very small. 5. Due to conservation of energy, electrical energy can only be produced at the expense of thermal energy. The effect of these features on the efficiency curves is indicated in figure 8. Particularly the spectrally selective layer

SYSTEM EFFICIENCY

4.1 Dailyyield The thermal yield was simulated as a function of reduced temperature by assuming that at each moment the panel is in thermal equilibrium. In these simulations, the top loss was calculated from the empirical formula found by Klein (Duffle and Beckman, 1991, p. 260). In order to test the software program, the daily yield was measured and subsequently simulated. The ambient conditions during the day were those presented in figure 9. The inlet temperature was kept constant.

Fig. 9. Ambient conditions on July 12, 1997, against the hour.


100

800

.

,

,

.

.

4.2 Annual Yield

.

Next, simulations were performed to find the thermal and electrical yield of the prototype combi-panel for the Dutch KNMI test reference year. The program was used to model the

600

.

~. 4oo

.

.

.

.

.

.

6O

0

5O

20o 0

I

10 600

,

12 ,

I

I

14 16 hour of day ,

|

I

18 .

20

[ 30 o 20

,

0

10

0

500

, 8

|

10

12

~: 400 0

.

,

.

14 16 hour of day ,

.

18 .

20 .

o 300 0

200

60

,~ 100

0

~ 40

I

8

I

10

12

14 16 hour of day

I

18

0

20

Fig. 10. Calculated (dashed) and measured (solid) thermal power for the conventional thermal collector (above) and the combi-panel (below), on July 12, 1997, against the hour. The results of the simulation are presented in figures 10-12. Figure 10 shows a good correspondence between the measurements and the simulations, although the calculated values tend to be slightly larger than the measured values, as found before. In addition, it can be observed that the simulations somewhat over predict the measured thermal efficiency in the morning and slightly under predict the measured thermal efficiency in the evening. It was found that this was due to the effect of the roof tiles, which effectively increased the heat capacity of the system. These results indicate that the assumption of thermal equilibrium in the simulations works quite well. On the basis of these results, it is concluded that hourly data are sufficiently accurate to give a good estimate of the annual thermal yield of the system, especially since the effect of the roof tiles largely cancels over a day. Figure 11 shows the electrical power of the system and figure 12 shows the temperature difference between the PV-panel and the combi-laminate. Figure 11 indicates that the electrical efficiency of the PV-laminate and the PV/T collector are almost the same. At the other hand, figure 12 indicates that in the PV/T-unit the temperature of the PV is much lower than the temperature of the conventional PV unit for the present case in which the inlet temperature was kept constant at approximately 18 ~ This implies that the electrical gains due to cooling of the PV by the water are of the same order as the optical loss of the PV/T-collector, that is due to the reflection at the glass cover.

0

0 i

I

8

10

12

14 16 hour of clay

18

20

Fig. 11. Calculated (dashed) and measured (solid) electrical power for the PV-panel (above) and for the combi panel (below), on July 12, 1997, against the hour.

60

50

L___

3o 20

I

8

10

12 14 hour of day

16

18

20

Fig. 12. Measured temperature of the PV panel (dashed) and the PV/T-laminate (solid), on July 12, 1997, against the hour. case in which two similar combi-panels with a joined area of 3.5 m 2 and a mass flow of 50 kg/(m2 hr) were used to heat a container of 175 litres of water l~om 10 ~ up to 60 ~ A boiler unit was assumed to do the remainder of the heating required if a temperature level of 60 ~ could not be reached by the PV/T unit. The tapping pattern was modelled atter the hot water withdrawal schedule of the ISSO (Institute for Study and Stimulation of Research in the area of heating and air conditioning), which is presented in table 1.


The thermal and electrical efficiencies were found to be 33% and 6.7% for the configuration used, as compared to 54% for the conventional thermal collector and 7.2% for the conventional PV-laminate under the same conditions. The electrical efficiency was calculated from an efficiency of the PV-laminate of 9.7% at 25 ~ corresponding to figure 4, and an inverter efficiency of typically 92%. Due to reflection of 8% of the incoming light at the glass cover, a thermal efficiency of 0.92 x 7.2% = 6.6% would be expected if the temperature effect on the PV could be ignored. This implies an increase in electrical efficiency of 0.1% of the yearly electrical efficiency due to the temperature effect. Clearly, for the yearly electrical efficiency, the effect of the glass cover is much more important than the effect of the temperature, which largely cancels out over the year. However, if the present collector would have been used for a low-temperature system instead of for the production of hot tap water, the increase in electrical efficiency due to the temperature effect would have been larger, as indicated by the results presented in figures 11 and 12. Finally, the annual thermal and electrical efficiencies were calculated when the special features of the combi-panel were successively removed. The results are summarised in table 2.

Annual thermal efficiency

Configuration

Annual electrical efficiency

Annual thermal efficiency 33.4% 6.7% of the combi-collector Optical efficiency increased 41.0% 6.5% Heat resistance removed 44.7% 6.7% Emission factor reduced 49.9% 6.6% Additional heat transfer sideways 49.6% 6.6% No electricity production 54.4% 0% TABLE 2. Contribution of the various loss mechanisms in the annual electrical efficiency of the combi panel. This table shows that the thermal loss due to production of energy is smaller than the electrical gain. This is due to the fact that the PV is effectively cooling the system by converting irradiation to electricity instead of heat. This implies a small reduction in thermal losses. The thermal loss due to production of electricity is only 5%, whereas the electrical energy produced amounts to 6.7% of the yearly irradiation.

Hour

1

2

3

4

5

6

7

8

9

5

101

CONCLUSIONS

A non-optimised first prototype of a combi-collector was built. From the measurements the thermal efficiency at zero reduced temperature in the absence of the production of electricity was found to be 59%, which is 25% less than the thermal efficiency found for the corresponding thermal collector. The electrical efficiency with electricity production was found to be 54% and the corresponding electrical yield is around 8.3%. From the KNMI test reference year and the ISSO tapping schedule, an annual efficiency of 33% thermal and 6.7% electrical was found if the collector was employed in a domestic water heating system. From the simulations, the magnitude of the factors limiting the performance of the combi-panel can be determined. The reduction in the annual electrical efficiency is mainly due to reflection at the insulating glass cover on top of the thermal collector (approximately 0.6% absolute). The reduction in thermal efficiency of the panel is mainly due to the fact that the glass on top of the PV-laminate is not spectrally selective which increases radiation losses (5% absolute) and the fact that the absorption of the PV-laminate is lower than the absorption of the thermal collector due to reflections in the PV-laminate (8% absolute). This model has proven to be an important tool for further optimisation of the eombi-panel. The results were used to build an improved prototype of the combi panel, which is presently under study.

REFERENCES Duffle J.A. and Beckman W.A. (1991) Solar Rngineering of Thermal Processes, 2 ~ edn, Wiley Interscience, New York. Vries D.W. de, Helden W.G.J. van, Smulders P.T., Steenhoven A.A. van, and Zolingen R.J.C. van (1997). Design of a Photovoltaic/Thermal combi panel momentary output model, outdoor experiment, ISES 1997 Solar World Congress, August

24-30 Taejon Korea. Vries D.W. de (1998), Design of a PV/Thermal Combi Panel,

PhD Thesis Eindhoven University of Technology. Vries D.W. de, Steenhoven A.A. van, Helden W.G.J. van, and Zolingen R.J.C. van, (1999) A panel-shaped, hybrid photovoltaic/thermal device, Dutch Patent 1006838.

1 1 1 1 1 1 1 1 1 1 2 2 2 2 0 1 2 3 4 5 6 7 8 9 0 1 2 3 Tapping . . . . . . . + . . . . + + - - - + + + - + + TABLE 1. ISSO warm water withdrawal schedule, (-) no withdrawal, (+) 175/8 litres withdrawal.

2 4 -


102

A COMPARATIVE I N V E S T I G A T I O N OF RADIATION HEAT T R A N S F E R IN TRANSPARENT INSULATION W I T H D I F F E R E N T R E F L E C T I O N M O D E L S

B. Aronov and Y. Zvirin Faculty of Mechanical Engineering. Technion, Israel Institute of Technology Haifa 32000, Israel Tel.: 972-4-8292070

Fax: 972-4-8324533, Email: [email protected]

- The present paper describes a comparative theoretical study of radiation heat transfer in TI structures, with three different models representing the reflection of the TI channel walls: specular, diffuse and a new anisotropic one. The latter assumes that the heat flux which impinges on the wall is reflected uniformly (isotropically), but only in the quarter sphere surrounding the specular direction, with nil reflection in the quarter sphere surrounding the incident direction (zero "back reflection"). For the specular reflection model a 1D ray tracing method is used and for the diffuse one a conventional Discrete Transfer Method is employed. For the third model, the 3D DTM has been modified to accommodate the anisotropie reflection mode described above. The radiation considered here is gray and the intensity reaching the TI structure is taken to be isotropic. The heat flux leaving it on the other side is calculated, i.e. the total transmittance of the TI is obtained. It is assumed, for convenience of comparison between the models, that the channel walls do not re-emit radiation (cold, at 0 K). As expected, the differences between the heat fluxes obtained with the specular and diffuse reflection models are enormous. The values obtained with the anisotropic model are in between, and closer to the specular case. The anisotropic reflection model was used by Aronov & Zvirin (1999) in a simulation algorithm for a solar collector with a TI separator placed between the cover and the absorber plate. The simulation results agree quite well with the experimental data ofRommel & Wagner (1992), which is an indication of the validity of our new anisotropic reflection model. Abstract

transmittance, 2", of the actual wall, and whose effective 1.

emissivity, 13e , is equal to the emittance, E, of the actual

INTRODUCTION

Transparent insulation (TI) has been used in various solar energy applications and systems both passive and active. The TI is made of a multitude of parallel capillary channels of rectangular, round or honeycomb cross section with transparent (glass or plastic) walls, see Fig. 1. Due to the rising spread of TI use, e.g. Goetzberger (1992), it has become apparent that more accurate modeling of the heat transfer inside it is needed, mainly because of quite significant differences between theoretical predictions and experimental data. One reason for this is the reflection characteristics employed in the available models, either specular or diffuse. It is well known that neither is completely correct, and in reality the reflection has components of both. A bundle of radiated thermal energy, when impinges on a surface, is partially absorbed, reflected and transmitted. In case of periodic structure, e.g. honeycomb transparent insulation (TI) as in Fig. l a, the transmission can be taken into account by effective reflection of an equivalent opaque wall. Hollands et al. (1984) developed an approximate model connected with a symmetry of the honeycomb according to the mirror-image technique of Eckert & Sparrow (1961): an opaque wall, whose reflectivity,

Pe , is equal to the sum of the reflectance,

/3, and

wall. The direction at which the effectively reflected bundle will travel from the surface is governed by an associated probability distribution. Surfaces that emit or reflect diffusely have a hemispherically uniform directional probability distribution. Other surfaces reflect specularly and there is also back-scatter reflection, meaning that all radiation is reflected back in the direction from which it arrived. Most surfaces, however, do not reflect in any of these ideal modes, but have an angular dependent reflection distribution in the hemisphere, which is bidirectional. A diffuse approximation is applied in many of the accurate multi-dimensional methods for solving the long-wave radiation problem. Schornhorst & Viskanta (1968), and Herring & Smith (1970) investigated experimentally the accuracy of the purely diffuse model for most common surfaces. These experiments indicated that the reflection of many surfaces of engineering importance was closer to the limit of specular reflection than to that of a perfectly diffuse one. However, applying an accurate specular reflection model is not practical: it requires complicated and time-consuming numerical calculations. This is not justified because the surfaces do not reflect perfectly according to an ideal specular model. Therefore, Aronov & Zvirin (1999) developed a new model, quite similar to the three dimensional Discrete Transfer Method (DTM), e.g. Lockwood & Shah (1981). The new model considers anisotropic reflection (the reflectivity includes transmissivity) of the cell side walls, so that the radiative flux from below is reflected upward only and vice versa. In the


"classical" DTM, the reflected heat flux is calculated only approximately, while in the method developed by Aronov & Zvirin (1999) with anisotropic reflection, it is computed accurately, by means of iterations until full relaxation is achieved. This model was applied to investigate heat transfer in solar collectors with TI, as a conjugate radiation - conduction convection problem. Comparisons with the experimental results of Rommel & Wagner (1992) showed good agreement. It was demonstrated that the results for the collector performance obtained by using the 'conventional' diffuse reflection (isotropic) model, tend to significantly overpredict the more realistic values obtained by the new anisotropic model.

~

qi

interior surface of the channel, see Fig. lb, while the rest of its surface is cold as before. The different orientations of the emitting surfaces lead to different results both quantitatively and qualitatively.

2. STATEMENT OF THE PROBLEM The transparent insulation (TI) structure is assumed to be of honeycomb type with the same cell geometry parameters in the transverse directions. The TI elements are square in cross section, uniform and equally spaced. This assumption allows to consider the TI element walls as opaque, with an effective reflectivity, ,Oe, given by the sum of the real reflectivity and transmissivity. Further assumptions are as follows. Direct absorption of heat radiation in the air within the TI structure is negligible. Convection in the channels is suppressed. The effective reflectivity of the TI channel walls is independent of the angle. The TI element walls are cold (at 0 K), i.e. not emitting. As mentioned above, three models of TI cell walls reflection are compared: specular, diffuse and anisotropic.

/1/1/1/1/1/ -'~

103

H

Output heat radiation fluxes, qo, transmitted through TI /

/ /

structures of different channel lengths, h, and with different wall absorptivities,

qo

Ew = I-De,

are calculated under

condition of a reference input heat radiation flux, q i, for the two cases mentioned above: aperture and hot thin rim radiation. For the former, q i is determined as an isotropic black radiation that corresponds to a black plate radiation at a given temperature

b"i ./

I "i~ I / e x i t p l a n e

Figure 1. Transparent insulation (TI) structure: a) general view, b) representative ray trajectory in a single channel, emerging from i on the hot rim and leaving at O, after multiple reflections at the walls.

In the present paper, the new model is used to compare pure radiation fluxes through TI structure calculated with three reflection models: for the two above-mentioned models (diffuse and anisotropic) and specular. The simplest way for the comparison is that under conditions of cold TI cell walls (absence of re-emitting). The comparison of radiation fluxes through the TI channel of the cold walls for the three reflection models is carried out for isotropic input heat radiation at one of the channel edges for two cases: 1) the radiation emerges from the whole channel aperture and 2) from a thin rim on the

Ti .

For the case of hot rim radiation, the

temperature of the rim at the upper channel edge is determined by the following condition. The downward rim radiation flux (into the channel) is set equal to the same value as in the previous case (on the cross section plane), and is isotropic too. Maximal absorbing plate temperatures have been calculated by the anisotropic and diffuse reflection models. These were made for a solar collector with a TI structure of glass capillaries, having the same parameters and the same climate conditions as in Rommel & Wagner (1992). In this case, the TI structure was considered approximately as a honeycomb of square cross section channels, with effective channel wall thickness taken based on "aperture conservation". As mentioned above, the results were obtained by means of conjugate heat transfer analysis within the collector. The radiation is considered as semi-gray, with two spectral ranges: short-wave solar radiation (beam and diffuse), and long-wave infrared radiation emitted diffusely by the collector elements. The TI channel walls reflect specularly the short-wave radiation, and diffusely, into a quarter-sphere (according to the anisotropic reflection model), the long-wave radiation: the upward one is reflected isotropically in the upward direction only, within the limits of the corresponding quarter-sphere, and vice versa for the downward radiation. The temperature varies only with the axial direction, z: it is assumed to be uniform across the thickness of the TI element wall and the enclosed air. The side walls of the TI structure are insulated.


104

2

3. G O V E R N I N G E Q U A T I O N S

J ; - i b (T,. ); J+- ( N ) - -- pe (N)q~c ( N )

The complete system of equations, which govern the nonlinear stationary problem of the conjugate heat transfer within an element of the TI structure, including the radiation, is presented in Aronov & Zvirin (1999). Thus only the radiation transfer equations for the three reflection models will be presented here. These equations correspond to the following boundary conditions: 1) Cold TI cell walls at 0 K, with the exception of the hot rim area in the corresponding case; 2) The channel under consideration ends in black covers at both of its edges; 3) In the aperture radiation case, the upper black cover temperature is specified corresponding to an isotropic heat flux; in the rim radiation case, the rim temperature is determined by the condition of having the same flux on the cross section plane and the upper black cover is cold; 4) The lower channel cover is cold in both of these cases. For the diffuse reflection model, the radiation transfer equations under these boundary conditions are as follows: 2~

dip d{~cos[g(M), ~]J(N)

q~.c(M) = 0

0

(1) 1 Jw ( N ) - - - Peq,,~

J, - i b (T~ );

point

M

on

the

channel

wall

are incident fluxes at point M,

obtained by integrating the radiosities:

J +(~)

over the lower

hemi- or quarter-sphere for the upper plate or the TI channel walls, and J - ( ~ )

over the upper hemi- or quarter-sphere for

the exit plane or walls, respectively. When N belongs to the upper plate or exit plane, the second of equations (4) includes

q~r

Thus J~(~)

for the direction

~

are

calculated under conditions of isotropic emission at the plate and anisotropic reflection (at the side walls) of the infrared radiation. A specular reflection case, under the above boundary conditions, can be solved in a way similar to that of Platzer (1992) for the insolation, since the heat radiation within a TI channel does not depend on the temperatures in the case of cold channel walls. His approach was used by Aronov & Zvirin (1999). In the case of isotropic 'aperture' input heat radiation, which is transmitted within the channel of cold walls, the above-mentioned calculation must be performed for every

the

~r2,and then the output ||

is incident flux at or

q+~(M), qL (M)

element of a discretized solid angle,

(2) q~,,~( M )

Here

U

(N);

Jo = 0 , where i b ( T i) - oTi 4 / z ,

(4)

exit

plane,

ffi, ffw ( N ) and fro are radiosities on the top plate, the TI channel wall and on the exit plane, respectively, q) and

nux, qo, can be obtained by umming of qo (fi)" For cross section channels, the corresponding equations for

qo(~)

are as follows:

qo ( ~ ) = q,~ ( ~ ) P :

costo

(5)

/

(6)

with n = n x + g/y ,

are the azimuth and polar angles and N in eqs. (1,2) is a point L/

on the other end of the beam ~r2, at its intersection with another channel wall or plate. Similar equations, but for two opposite directions (upward, downward), describe the radiation heat transfer in the case of the anisotropic reflection model:

cosco = sin ~c~

\sintp/

for x and y-oriented walls, respectively,

2g

q ~ (M) =

0

nx -

dip n/2 d~cos[~(M),~]J +(N) (3a)

Int(H tan ~ cos 9),

ny - Int(H tan r sin {p) (7) and H is the TI channel spacing.

2n

qi.r (M) = o

n/2

0

d{~ cos[g(M), ~ ] J - (N)

(3b)

In the case of 'hot rim' radiation, the algorithm for the calculation of qo is the following. At first, for every hot rim control element, N, and a channel wall surface element, M, the fluxes q'+mc( M , N , ~ ) , are calculated as in the anisotropic


reflection model, see the kernel of the integral in Eq. (3a). Then,

250

for every set of M, N, h , the output fluxes, qo ( M , N , h )

~x Specular reflection model [] Anisotropic model o Diffuse model Aperture input heat flux ----- Hot rim input heat flux

13=0.15

2001

are calculated as in the previous case, Eqs. (5 - 7), with distance z between M and the lower channel edge, and integrated over the angle space.

105

150 o 100

4. R E S U L T S AND D I S C U S S I O N 501[

The radiation heat flux emerging from the TI structure was computed with all three reflection models as a function of the channel aspect ratio, ~ - - h / H ,

and

the channel wall

0

absorptivity, E , for the two cases of emitting surfaces defined above: 'aperture' and thin hot rim. The results are presented in Figs. 2 - 4 and Tables 1 -3. As expected, the differences between the heat fluxes obtained with the specular and diffuse reflection models are enormous. The values obtained with the anisotropic model are in between, and closer to the specular case, especially for ~ < 8

and E ___0.3

for the 'aperture'

case. For the hot rim radiation case, the values associated with the diffuse reflection model are the lowest, again, and tend to those of the anisotropic model with increasing ~

Figure 2.

6

for

8

12

Exit radiation flux,

16

qo,

20

24

as a function of the TI

channel aspect ratio, ~, for the three reflection models and two input radiation cases: 'aperture' and 'rim'. Infrared emissivity of channel walls E -

0.1

5.

and E,

more moderately than in the previous case. In contrast with the aperture radiation case, the output fluxes for the anisotropic mode for the hot rim case exceed those of the specular one at

~:
~ k = 1

(AS) APPENDIX

B - CORRELATIONS m

The correlation of (Hay, 1979), between H T and H

is the expected value of X i . If the distribution of Eik = X i k - fli is normal, for N > 30 the

expressed in the dimensionless form as

estimator o f / t i is given by

m

-

1 = --

Xi

N

(rot) n HO

N X

k=l

ik

(~'a) n

m

(A6)

('t'a)g ~ T ( l _ c o S f l ) + (~"~)d ff(gT) + Pg (rct)n 2 ('ca)n

while cr/j is estimated by

m

1 - -

m

x { (K T - r 1

can be

T ))R b +

. _ -

m

(1 + cos fl)[1 - (K T - ~ ( K T) (B 1)

N

(Xik - ~ ' ) ( X j k - X i )

tr/j= N

(A7) B

k=l

m

where f0= H d / H o is expressed according to (Erbs et al., 1982) as follows

and cri is given by 1

cri = ~ -

N

- - -2

(Xik - X/) k=l

(A8)

e(KT) =

HO

[ 1.391KT-3.560KT 2 +4.189KT3-2.137K'r 4

In the case the same bias error B and precision index tS are assumed for all X i , it is easy to see from Equations (A1) and

=~

0.3 < K-'T - < 0.8, for cos < 81.4 ~ 1.311K'-r - 3.022~r 2 + 3.427Kr 3 - 1 . 8 2 1 K r 4

(A2) that

[ 0.3 < K-"T < 0.8,for cos > 81.4 ~ = ( n Of Of Pij) U2 Uf2 i,j=l ~Xi ~X j

(A9) m

m

n

where K r = H r / H o .

032)


qz'(r'"r ) = [1- q}'(X'-r )] (ra)6 R'b ('/'a) n

('t'a)g ( 1 - c o s f l ) + ('t'a) d tp,(~r) +P g ('t'a) n 2 ('ta) n x{[Kr - e ( K T )]Rb + 2(1 + cos fl)[1- (K r - tP(Kr))]} +

('t'tr)d e ( K r ){[I-e'(K T)]/~'b

(ra).

-

+

033)

(1 + c o s f l ) [ t p ' ( K r ) - l ] }

where (o'(KT) = d_._r . dKr The monthly solar fraction f expressed as

(Duffle and Beckman, 1991) is

f = 1.029Y-O.O65X-O.245Y 2

+ 0.0018X 2 + 0.0215Y 3

034)

The partial derivative of f with respect to Y is g ( Y ) = Of/~gY = 1.029 - 0.49Y + 0.0645Y 2

034)

149

150


SOLAR POND AS AN ENERGY SOURCE FOR DESALINATION Uri Fisher Head R& D Dept., Ormat Industries Ltd., P.O. Box 68, Yavne, 81100, Israel, Telephone Number 972-8-9433777, Fax Number 972-8-9439901, E-mail :[email protected]

Abstract - A shortage of potable water in many populated areas around the world has already reached the point at which desalination of seawater is the only solution. A large number of arid zones, near the seashore, have the requirements for seawater desalination and are situated in a high enough solar radiation zone suitable for large scale utilization for power or heat production. Compatibility between the heat source and the desalination plant utilizing it, is a major factor in the economics of such cogeneration plant. The development of solar pond technology was accelerated in the eighties mainly because of the threat of increasing oil prices that until now seem to be false alarm. The aim was electric energy production that was evaluated against electricity generated from gas, coal and oil. Because of the strong emphasize on energy, the discontinuing of energy projects due to the inability to compete in power generation, shadowed other applications such as desalination that lacked the sense of urgency which approaches us today.

The design, construction and most important, the continuous maintenance of the solar pond have been tested and evaluated. Cost of construction, construction materials and cost of maintenance work and materials has been up-dated. Information in regards with updated flash desalination techniques is available in the market. However, the simplicity in control and operation and the small temperature difference per effect, made us chose the LT-MED as a compatible system for cogeneration of heat and water. This paper discusses the combination of these two proven technologies: Salt Gradient Solar Pond (SGSP) and MultiEffect Distillation (MED), for large-scale solar desalination. The solar pond produces heat at temperatures ranging from 60~ to 95~ Assuming availability of sea water for desalination at 25~ to 35 ~ this same water act as heat sink and allow for large enough number of desalination effects between the heat source- the bottom layer of the pond and the heat sink- the sea water. Salt gradient solar pond needs regular supply of high concentrated brine to compensate for the natural diffusion of salt from the high concentrated bottom layer to the low concentrated top layer via the middle gradient zone. The combination of the MED plant with the pond eliminates the need for an evaporation system as it supplies the necessary concentrated brine by the continuous flashing of vapor from the brine in the flash chamber that acts as a salt generator. At the same time the continuous supply of concentrate enhances the stability of the solar pond that perfectly matches the operating temperature range of the MED and together provides a very inexpensive and most competitive seawater desalination systen~ It should also be specifically stressed that the combined system is environmentally benign. No brines are spilled on the ground and salt that may be accumulated on the bottom of the pond can be regularly collected and disposed of in sacks after sun drying or sent to other users. Water cost for 10,000 m3/day plant is about $1/m3. The product is high quality water of about 25 PPM or less. If mixed

with 3,000 PPM brackish water to give a final product of 800 PPM, production will rise by 35% and water cost will drop to $0.75/m3. Keywords: Solar Energy, Solar Pond, Desalination, High Concentration, Thermal Energy

Background The first practical research on the solar ponds was initiated in 1958 by Dr. Bloch who was the director of development of the Dead Sea Works in Israel. There was a slow but important development work Weinberger (1965) also by Tabor and Weinberger (1980) who took the initiative to summarize the technical achievements of those years. Then in the Seventies and Eighties came an era of enhanced development of solar pond by Ormat. The aim was to create an alternative energy source for power production as reported by Doron in 1986, Tabor in 1987. There were some demonstrations of utilization of the heat for green houses or industrial use as in the case of the El-Paso pond that was and is still used for power and heat supply as reported by Hightower, 1987. The continuous low price of oil kept industry away from this subject and most of the academic institutes do not have the financial resources to run such a system and further develop it. From time to time we find that another pond has been constructed, see Hassab (1992), Alagao (1994). Unfortunately in most cases it is only a demo that never develops into a full operating systerrL The possible use of the pond energy for desalination has already been mentioned by Tabor in 1975 as also by Doron et al 1991, Gluckstern 1991 and Hoffman 1992. Since the water situation in Israel is very sensitive to the annual precipitation with hardly any reserves, it frequently creates public discussion. The agreements between Israel, Jordan and the Palestinian Authority enhance the feeling that the increase of water supplies is a crucial matter. The use of recycled water for irrigation may postpone the exact date when large-scale desalination will be essential but it is probably in the near future.


The Eastern Mediterranean is not the only near crisis area. In many populated areas around the world shortage of potable water has already reached the point at which desalination of seawater is the only solution. Large number of arid zones, near the seashore, in North Africa, Greece, Israel, south Italy, the Persian Gulf etc. have already used their ground water potential and require seawater desalination. Those countries are situated in a high enough solar radiation zone suitable for large-scale utilization of solar energy. As mentioned, since the main aim resulting from the oil crisis in 1974 was electric energy production, the pond feasibility was always evaluated by comparing cost of electric energy from the pond against electricity generated from gas, coal and fossil fuels. As a result of the inability to compete in power generation and due to the strong emphasize on energy, the pond related projects were discontinued. This also shadowed other applications such as desalination that lacked the sense of urgency that approaches us today. Pond Technology Research on the Salt Gradient Solar Pond (SGSP) was initiated in Israel in 1958. A number of small demonstration ponds were built to test various operating regimes and parameters. In the late 50s a test pond reached temperature of 96~ This was a most encouraging achievement that contributed to the understanding of the hydrodynamics of the pond as reported by Weinberger 1964. Those findings served as basis for the Ormat pond in Ein Boquek that was the first to have been combined with power generation using an Organic Rankine cycle. The successful operation of this pond as reported by Doron and Tabor 1986, led to the construction of the largest pond ever built also by Ormat near the Dead Sea. Results of the operation of the 250,000 m2 SGSP and the 5MW power plant were reported at the first International Conference on solar ponds in Cuernovaca, Mexico 1987 by Tabor. The scheme of pond and power plant is given in Fig. 1. The major technical issues studied during the construction and operation of the large pond were: -Pond geometry and optimal size -Pond lining and leakage security -Brine leak detection -Heat Loss to the ground -Formation of gases under the pond -Initial pond filling method -Establishing the salinity gradient -Maintaining the salinity gradient -Limiting the upper mixed zone layer -Heat extraction method -Power generation -Wind protection -Pond clarity maintenance -Salt make-up

Intensive work was published by Zangrando1979, 1980 and others related to pond operation and control, see Swift1989. Other ponds built since then for example the El-Paso pond is still running and it serves as a study place for students from the University of Texas that share the operation with other consortium partners. See Swift, 1993.

151

Desalination technologies and the combination with SGSP. The solar pond can produce either heat or energy for desalination. In case of electric energy production the desalination processes suitable are the Vapor Compression (VC) or the reverse Osmosis (RO). Both need only electric energy for driving the process. Since we consider here largescale operation we will use only SWRO for the comparison. In case of heat processes, there is always need for pumping and the pond can supply both the heat and the power.

The solar pond produces heat at temperatures ranging from 60~ to 95~ Assuming availability of seawater for desalination at 25~ to 35 ~ this same water acts as heat sink and allows for large enough number of desalination effects

Figure 1. Solar Pond and Organic Rankine Cycle Power Plant. between the heat source- the bottom layer of the pond and the heat sink- the sea water. A lot of information in regards with updated desalination techniques is available in the market. However, the simplicity in control and operation that allows the MED plant to work between 40% to 120% of the load and the small temperature difference per effect, made us select the Multi Effect Distillation MED as a preferred system for water production. However, Multi Flash Distillation (MSF) and SeaWater Reverse Osmosis (SWRO) will be compared as well. Salt gradient solar pond needs regular supply of high concentrated brine to compensate for the natural diffusion of salt from the high concentrated bottom layer to the low concentrated top layer via the middle gradient zone. The combination with power production cycle has no influence on the salt concentration in the various layers of the pond however; the combination of the MED plant with the pond eliminates the need for an additional salt generation system. The continuous flashing of vapor from the brine in the flash chamber supplies the necessary concentrated brine to the pond.

152


ORC average thermodynamic efficiency: 6.15 % Pond electricity production: ................ 24.35 kWh/rn2/year Pond water production for the thermal processes is based on the average Economy Ratio (ER) Pond specific construction cost: 16 $/m2 Due to unknown financial source, the cost was calculated assuming change of interest rate between 2 % and 8%. In case of different numbers one can easily complete the table.

Table 1. Size and Cost Summary Process ER

Figure 2. Solar pond and MED desalination plant. |

. It should also be specifically stressed that the combined system is environmentally benign. No brines are spilled on the ground and salt that may be accumulated on the bottom of the pond can be regularly collected and disposed of in sacks after sun drying or sent to other users.

Comparison of water cost The three main desalination technologies are presented for comparison, i.e. MSF, LT-MED and SWRO. Heat is supplied by the pond for the MSF and LT-MED while electric power can be either supplied by the pond or by the grid. In case the pond supplies the electric energy, an additional pond area is considered. Since the pond is also storage of energy for non-operating plant hours, we can assume that the production will be recovered in later period and therefore the desalination plants operates 8760 hours per year. Numerous proposals for the construction of large-scale desalination plant based on solar pond were submitted in the last ten years to various government agencies in Israel. Costs were re-evaluated and the resultant figures were used for a renewed evaluation of water cost that follows here. It has also been decided to aim at a size that will combine the desalination plant with practical size and number of solar ponds and use it as a module for multiplication in case larger desalination plants will be considered.

General assumptions: Desalination Plant size: ... 10,000 m3/day Average global solar Radiation: . . . .

2,200 kWh/rn2/year

Solar pond thermal efficiency: .............. 18 % Power is produced by Organic Rankine Cycle (ORC).

MSF 5.5 |

Specific water . production Specific energy consumption Pond area for desalination Pond area for , power Total . pond area Specific cost of Desal. Plants Desal plant cost Pond cost +power ,production I Pond cost desal. , only Total cost=power , production Total cost desal only

L T - M E D S W R O units 10 |

|

3.382 , 4.5 . 1,079 674 , .

1,753 . 1,500

|

m3/m2/y

6.15 , 2.5 . 593 ! i374 , 968 . . 1,400

I 5.5 .

' kWh/m 3 i |

1,000xm 2

1,000m2

824 824

" 1,000xm 2

1,200

$/m3/day

15,000

14,000

12,000 $xl,O00

31,568

17,428

1 4 , 8 4 0 $xl,000

, 19,500 i

, 10,682

'

$xl,000 '

46,568

31,428

2 6 , 8 4 0 Sxl,000

34,500

24,682

14,840 $xl,O00

Table 2a- MSF water production cost

Plant costPower at $0.06/kWh O&M 4% of investment/yr Chemicals & consumables Total

Including P/P 46,568,000

Without P/P 34,500,000 0.275/m 3

0.515/m 3

0.3785/m 3

0.0505/m 3

0.050$/m 3

0.565/m 3

0.6985/m 3


Table 2b :MSF Investment & annual costs 25 years plant life-time

Including P/P

Without P/P

Interest %

Return Rate %

2 4 6 8 2 4 6 8

5.1 6.4 7.8 9.4 5.1 6.4 7.8 9.4

Fixed water cost $/m3 0.650 0.816 0.995 1.199 0.482 0.605 0.737 0.888

Total water cost $/m3 1.210 1.376 1.555 1.759 1.18 1.303 1.435 1.586

Table 3a: LT-MED water production cost

Plant costPower at $0.06/kWh O&M 4% of investment/year Chemicals and consumables Total

Including P/P 31,428,000

Without P/P 24,682,000 0.155/m 3

0.3445/m 3

0.2705/m 3

0.050$/m 3

0.0505/m 3

0.3945/m 3

0.4705/m 3

Table 4b SWRO Investment annual costs 25 years plant life-time Interest Return Fixed water cost % Rate % $/m3 0.375 2 5.1 Includin gP/P 0.470 4 6.4 0.573 6 7.8 0.691 8 9.4 0.207 2 5.1 Without P/P 0.260 4 6.4 0.317 6 7.8 0.382 8 9.4

153

Total cost $/m3 0.822 0.945 1.017 1.135 0.864 0.917 0.974 1.039

water

The results of tables 2,3,4 can be observed in Fig. 3. It shows that for a 10,000 m3/day plant the LT-MED system competes very well with SWRO (Desalination only), that is considered today as a most competitive plant for sea water desalination. Both desalination plants end up with water cost around 1 $/m3. The salinity of product water of the SWRO system is about 500ppm while the salinity of the product of the MED plant is about 25 PPM or less. This high quality product is actually tasteless for drinking. If mixed with 3,000 PPM brackish water to give a final product of 800 PPM, production will rise by 35% and water cost will drop to $0.75/m 3. Summary: A salt gradient solar pond can supply heat for the production of desalinated water at competitive price. The LTMED is the most suitable desalination process to be combined with a solar pond.

Table 3b:Investment and annual costs. Assume plant lifetime 25 years.

Including P/P

Without P/P

Interest %

Return Rate %

2 4 6 8 2 4 6 8

5.1 6.4 7.8 9.4 5.1 6.4 7.8 9.4

Fixedwater cost $/m3 0.439 0.551 0.671 0.809 0.344 0.432 0.527 0.635

Table 4a :SWRO water production cost Including P/P 26,840,000 Plant costPower at $0.06/kWh 0.2945/m 3 of 4% O&M investment/year 0.100$/m 3 Membrane replacements 0.0505/m 3

Total cost $/m3 0.822 0.945 1.065 1.203 0.814 0.902 0.997 1.105

water

Without P/P 14,840,000 0.3305/m 3 0.162$/m 3 0.1005/m 3

0.0655/m a

Chemicals and consumables Total

0.444$/m 3

0.6575/m s

Fig. 3. Comparison of water cost for MSF, LT-MED and SWRO driven by solar pond.

154


References: Alagao F.B (1994) the design construction and initial operation of a closed cycle salt gradient solar pond. Solar Energy 53, (4) 343-351.

Doron B. (1986) Solar Ponds-Lessons learned kW power plant in Ein boqueque April Conference Anaheim. Doron B. (1991) Solar Pond as an actual desalination IDA conf. On desalination and Washington

from the 150 1986 ASME solution for water re-use,

Glukstern P. (1991) Potential use of solar energy for water desalination European seminar on new technologies for use of renewable energy sources in water desalination. Athens. 2628,September. Itassab M.A. (1989) Problems encountered in operating salt gradient solar ponds in the Arabian Gulf. Solar energy Vol. 43, No3, pp169-181 Bightower S. (1987) Installation and operation of the first 100 kW solar pond Power Plant in the US. International progress in solar ponds Cuemavaca, Mexico. Hoffman D. (1992) The application of solar energy for largescale seawater desalination. Desalination, 89, 115-184. Sargent, Stephen L. Solar Pond today, International solar ponds Vol.4, No 1, Feb. 1990

Swift A.H.P. (1989). Topics in gradient maintenance and salt recycling in an operational solar pond. ASME Mechanical Engineering Solar Energy division, pp391-400. Swift A.H.P. (1993) Final report Texas solar pond consortium project551, 1989-1993. UTEP Department of Mechanical Industrial Engineering, August. Tabor H. (1975), Solar Pond as heat source for low temperature Multi effect distillation plants. Desalinationl 7, 289-302 Tabor H. Weinberger H.Z. (1980) Non Convecting Solar Ponds. Solar Energy handbook, Chap.10. (Edited by Kreider) New York McGraw-Hill. Tabor H. (1987) The Beith Haarava 5MW solar pond power plant International conf. on solar ponds. Cuemavaca, Mexico. Weinberger, H. (1964) The Physics of solar pond. Solar Energy Vol.8, No 2. Zangrando F. (1979). Observation and Analysis of a full-scale experimental salt gradient solar pond. Ph.D. Thesis, University of New Mexico, Albuquerque. Zangrando F. (1980) A simple method to establish salt gradient solar ponds. Solar Energy 25,467-470.


155

MULTISTAGE STILL

Judith Franco, Luis R. Saravia, Sonia Esteban Instituto de Investigaci6n en Energias No Convencionales, INENCO, Universidad Nacional de Salta - CONICET Calle Buenos Aires 177, Salta- 4400- Argentina, E-mail: [email protected]

Abstract - A new design for a passive atmospheric multiple effect solar distillation unit is proposed. Inclined glass surfaces with a 4 ~ slope and placed one over the other in an isolated box are used. The cold salty water is fed only in the upper stage and flows along each surface, falling from one stage to the next by gravity and reaching finally a heated tray at the bottom. Vapour condenses below each surface and produces the water evaporation in the upper side of the same surface. The stilrs bottom is heated using a simple 1.3 m 2 solar collector with a fresnel type concentrator. The collector is separated from the still and heat is transported from one unit to the other using a 4 kg aluminium slab, which is placed in the absorber and it is heated at a temperature about 350 C. The slab is then placed below the tray in an isolated box. Four slabs are used and they are changed periodically when the slab temperature drops below 180 C approximately. Several slabs can be used for heat storage if several collectors are built, allowing the use of the still during some hours at night, improving its daily productivity. Experiments have been performed with a prototype and the results are discussed and compared with the values obtained with another electrically heated prototype.

1. INTRODUCTION In Argentina, as in most countries in the world, water is a priority. In many regions most of the water is salty, and no potable water is available. Passive solar desalination units can provide a solution in isolated rural areas for small group of persons. The simplest systems are the greenhouse solar stills. Small units with a daily production in the order of 4 litres per square meter are used. These productions are quite low. The use of passive multistage stills could provide an alternative if a simple and low cost design is available. In the past we have proposed the use of a still with a vertical disposition for the stages (1). Each one is made using an inclined stainless steel sheet with a 30 ~ slope and covered with a cotton fabric to improve the water distribution on the surface. Water is fed in the upper stage, runs along the surface and falls by gravity from one stage to the next until it reaches a tray in the lower position. Water is heated in the tray and the produced vapour condenses in the stage over it, which is cooled by the falling water. This process is repeated in each stage. In a recent work performed with greenhouse solar stills (2), it has been found that very small slopes can be used, in the order of 4 ~ A smooth water film is produced on the condensation surface when the glass is carefully cleaned with ammonia. In this paper the use of low slope glass surfaces is proposed for the multiple stage stills, reducing considerably the height of each stage and allowing the elimination of the cotton fabric, which is always a source of maintenance problems. 2. N E W DESALINATION SYSTEM

2.1 The Still Module and The Experimental Arrangement The still body is a rectangular box (50 cm x 50 cm x 36 cm) entirely built with glass and insulated externally with 50 mm thickness polystyrene foam. Stainless steel is used for the lower tray where the temperatures are higher. The system has a tray in the lower position, being heated from below. Five stages using glass surfaces with a 4 ~ slope are used, as it is shown in a cross section in Fig. 1.

Fig. 1: Multistage still cross section. The cold saline water is fed from above and falls from one stage to the next by gravity until it reaches the tray, which is heated from below. The water excess, with a high salt content, is eliminated from the system in the tray. In the upper part of each glass surface the cotton fabric is substituted by small glass dikes 1 cm high and placed as shown in Fig. 1. Water is fed in the upper parts and each dike forms a small lake with and average 1/2 cm depth covering the entire surface, since the slope is quite low and a very small amount of water is necessary to cover the surface. On the upper stage the water runs freely to maintain the temperature as cold as possible; the water supply for this tray being independent from the others. The lower tray is heated from below with four aluminium slabs that are placed in an isolated camera; the dimensions of the slabs are (27.5 x 27.5 x 2) cm 3. Figure 2 shows a general view of the still without the glass box insulation. One of the storage slabs is seen in an intermediate position.

156

ISES Solar World Congress 1999, Volume ill

Figure 2.- General view of the still without the upper insulation. 2.2 The Solar Heater Multistage stills work with good productivity if the tray temperature is higher than the one used in greenhouse stills, typically around 90 C. In the high slope prototype the wood heated the system. In the new one simple 1.3 m2 solar collector with a fi'esnel type concentrator is used. The collector is separated fi'om the still and heat is transported fi'om one unit to the other using a 4 kg aluminium slab, which is placed in the absorber and it is heated at a temperature about 350 C. The slab is then placed in an isolated box below the tray. Four slabs are used and they are changed periodically. Several slabs can be heated us storage elements if a larger

collector is built allowing the use of the still during some hours at night, improving its daily productivity. The reflector concentrator is of the fi'esnel type. It is made with small height cone trunks placed concentrically above a plane. Each cone trunk is manufactured with a highly reflective aluminium sheet. The reflector is placed on an equatorial mount so that a single axis is moved during the day. A second axis normal to the first one is adjusted biweekly as the sun declination changes along the year. Figure 3 shows a picture of the concentrator prototype and one of the slabs

Figure 3.- Shows a view of the concentrator and one o f the slabs


This disposition has several advantages: a)

the cones are built very simply from a flat sheet in comparison with other forms as the parabolic one.

b)

The whole mirror is flat simplifying its transportation.

c)

between cone and cone there are grooves allowing a better control of the forces produced by strong winds.

The two axes are placed in the centre of the mirror, where the incoming solar radiation is blocked by the absorber. Vertical steel column fastened to the floor supports the mirror structure. The angle of the equatorial axis and the floor can be adjusted since it should be equal to the latitude in the place where the concentrator is used. Soldered steel pipes with a rectangular cross section are used to make the mirror fiat structure. Steel wires circles with a diameter equal to the final inner cone diameters are fixed to the structure. The cones cut from fiat aluminium sheets are fixed against the wire circles as shown in figure 4 and they adopt the conical form with the angle needed to concentrate the radiation on the absorber placed 76 cm above the fiat surface. The external diameter of the whole mirror is 1.50 m and the effective reflecting surface measures 1.3 m 2.

157

Temperatures inside the still were measured every five minutes with small thermistors thermometers connected to a computer The temperature measured in the tray is not the real water temperature since the thermistor was placed in contact with the base of the metallic tray and this temperature is a little higher. Slabs temperature were measures with K type thermocouples placed inside a small hole made in the slab. Fig. 5-a and 5-b shows water's tray and slabs temperatures vs. time for two different days.

500 450 400 350 9300 0 Q.

250

E 200 150

50 0

10:00

,

~

,

,

,

,

,

11:00

12:00

13:00

14:00

15:00

16:00

17:00

18:00

hour

500 450 400 35O L)

Aluminum m

i

r

r

o

~

J

300

4,-'

m 250 0 o.

E 9200

I

I---

150

Rectangular p i p e /

100 50 -

0 10:30

Figure 4.- Scheme o f the procedure fo fix the conical mirror to the structure. Sheets of aluminium 0.5 mm thick with a 0.86 reflectivity are used to build the reflectors. The aluminium heat storage slab rear surface is insulated usin~ a high temperature low-density ceramic blanket. A 0.3 x 0.3 m" glass fixed to the mirror structure protects the front of the slab and decreases the heat losses.

3.

EXPERIMENTAL RESULTS

The measured experiments were carried out heating each slabs with an electric heater up to 400 C to obtain uniform conditions, allowing the comparison of the results obtained at different times. The slabs are changed every one and half-hours approximately.

11:30

12:30

13:30

14:30

_

iii

15:30

....

16:30

17:30

Hour

Fig. 5-a and 5-b: Tray and slabs temperature during the experimental time interval for two different days. The upper curve with a serrated shape is the slab temperature, the vertical lines indicates when the slabs are changed. The middle curve is the water temperature in the lower tray, and the other is the temperature in the water of the upper tray that is maintained at ambient temperature It takes almost three hours to reach the operative temperature at 95 C. When this temperature is reached it is necessary to put an aluminium screen between the slabs and the tray to decrease the heat transfer and maintain the temperature constant. This behaviour can be seen in Fig. 5 between 13:00 and 15:00 hours and in 5 b between 12.30 and 15:00 hours, it takes a larger time before changing slabs. The distilled water production was obtained manually. Fig. 6 shows the distillate production for different average


158

Table 1: Different values ofrp

temperatures that was measured in the lower tray. The temperature in the upper Way was kept constant around 18 C.

~

AT

Qs~bs

md

Qdestillate

(c)

(MJ)

(kg)

(MJ)

rp

Tavera ~e

1,8

236

3.57

2,25

5.40

1,51

97

1,6

236

3.57

1,8

4.32

1,21

97

245

3.70

1,85

4.44

1,20

98

185

2.80

1,35

3.24

1,16

97

209

3.16

2,7

6.48

2,05

96

1,4

~" 1,2

i ~0,8

In previous works (1) we have reported a value for rp around 2.7 for the same still heated electrically or with natural gas. This value did not considered the losses in the insulated box below the Way. The smaller rp measured here indicate that some losses in the isolated chamber are being produced.

Q 0,6 0,4 0,2 0 80

82

84

86

88 90 92 94 Average Temperature C

96

98

100

Rp vs Tmedia 2,50

Fig 6: Distilled water vs. average temperature 2,00

The best measured production is 1.8 It/hour at 96 C, the values for points on the right are smaller, probably due to observed vapour leaks from the Way. A performance ratio, rp, giving the relation between the heat needed to evaporate the produced water and the consumed energy was calculated from the experimental results. The results are shown in figure 7.

rp =

Qdistillate Q lab

1,50

i

AA 1,00

&

A

0,50

0,00 80

,

,

,

85

90

95

100

Tempemtura C

The amounts of heat were calculated as

Fig 7: The Performance ratio rp vs average temperature of the lower Way

Qdistillate - m d f~ md -

distilled water mass

1-2.4~

4. CONCLUSIONS

MJ

Kg

Qslabs = 4 m C p s A T m = slabmass = 4kg J

Cps = 9 4 5 ~

AT = Yfinal - Tinitia I Table 1, shows the different parameters used for the calculation of rv they were obtained during different days

The new still is quite compact and maintenance problems are kept to a minimum since it is completely built in glass, no cotton fabric is used and it is fed with salty water in a single point. The still starts to produce distilled water very quickly at the beginning of the day since its thermal inertia is quite low due to the small amount of water in the stages and the Way. The productivity of the still, about 1.8 kg/hour at the higher temperatures, is good. Values of rp lower than 2 indicate that heat losses in the insulated box should be better controlled. The experiments have shown that the aluminium slabs are too heavy for an easy manipulation at high temperature during long time intervals. A new design of the slabs is being considered to solve the problem.


5. ACKNOWLEDGEMENTS This work was partially supported by the CONICET (Consejo de Investigaciones Cientificas y T6cnicas). J. Franco and L. Saravia are researchers from the CONICET. The authors appreciate the collaboration of R. Caso and C. Fernfindez from the Universidad National de Salta to build the prototype.

6. REFERENCES

(1) Franco, J., Saravia, L., A New Design For Passive Atmospheric Multistage Still, Renewable Energy, Vol. 4, N ~ 1, pp 119- 122, 1994. (2) Franco, J., Destilador De Baja Pendiente, Avances en Eenergia Renovables y Medio Ambiente Vol 1, No. 1, pp. 6568, 1997.

159


160

DEVELOPMENT OF A SMART SOLAR TANK Simon Furbo and Elsa Andersen Department of Buildings and Energy, Technical University of Denmark, Building 118, DK-2800 Lyngby, Denmark

Abstract - Theoretical and experimental investigations of small SDHW systems based on so-called smart solar tanks are presented. A smart solar tank is a hot-water tank in which the domestic water can both be heated by solar collectors and by an auxiliary energy supply system. The auxiliary energy supply system heats up the hot-water tank from the top and the water volume heated by the auxiliary energy supply system is fitted to the hot-water consumption and consumption pattern. In periods with a large hot-water demand the volume is large, in periods with a small hot-water demand the volume is small. The investigations showed that the yearly thermal performance of small SDHW systems can be increased by up to about 30% if a smart solar tank is used instead of a traditional solar combi tank. The thermal increase is strongly influenced by the hot-water consumption and consumption pattern. Recommendations for future development of smart solar tanks are given.

1. INTRODUCTION Almost all small solar heating systems for domestic hot water supply, SDHW systems, for single-family houses in Denmark are single-tank systems based on a combi hot-water tank. The domestic water in the eombi hot-water tank can be heated both by the solar collectors and by means of an auxiliary energy supply system. The water at the top of the combi hot-water tank is heated to a required temperature by means of the auxiliary energy supply system. In this way the top of the tank is always kept at a high temperature level. The volume of the water at the top of the tank heated by the auxiliary energy supply system is determined by the design of the tank. In the marketed Danish combi tanks this volume is sufficiently large for families with relatively large hot-water consumption. For marketed solar tanks with total volumes between 155 1 and 390 1 the top volume is situated in the interval from 601 to 1601. Measurements by Otto et al. (1997) have shown that most Danish families today have a relatively small hot-water consumption of about 60-150 l/day. The average hot-water consumption for a family is about 100 l/day. Measurements have also shown that the hot-water consumption and the consumption pattern vary strongly from family to family, and that the hot-water consumption and consumption pattern are not the same for a specific family during all periods of life. Further, increased water price and water saving equipment will most likely result in decreased hot-water consumption in the future. Furthermore, the hot-water consumption is normally not known before solar heating systems are installed. Obviously it is very difficult to choose the volume of the combi hot-water tank and the top volume of the tank in the right way. It is also obvious that the marketed solar tanks are oversized for typical hot-water consumption. Theoretical investigations have shown, Furbo and Shah (1996) and Shariah and L6f (1997), that the thermal performance of typical solar heating systems based on combi tanks can be strongly influenced by the hot-water consumption pattern and that the thermal performances of combi tank systems are much smaller than the thermal performances of preheating solar

heating systems with tanks which can only be heated by solar collectors. The ideal solar heating system from a thermal and energysaving point of view is therefore based on a preheating tank which can only be heated by the solar collectors and an auxiliary energy supply system built into the hot-water pipe from the tank close to the tapping locations. The auxiliary energy supply system heats up the domestic water instantaneously to the required hot-water temperature during tappings. In this way the thermal performance of the solar heating system is maximized since the operation temperature of the solar collectors is reduced to a minimum and the heat loss from the hot-water pipe and from the auxiliary energy supply system is minimized. However, a large power supply from the auxiliary energy supply system is required in order to maintain a reasonable hot-water comfort. For instance, a power supply of 20 kW is needed to heat cold water from 213~ (cold water temperature in March in Denmark) to 5013~ for a tapping flow rate of 6 l/rain. With a tapping flow rate of 12 1/min. 40 kW is needed. The maximum power supply from typical oil-fired boilers or natural gas burners for one-family houses in Denmark is about 20 kW. Therefore the hot-water comfort will not be sufficiently high for typical boilers/burners. Consequently the pure preheating system is not attractive in most houses. 2. SMART SOLAR TANK 0 PRINCIPLE The advantages of the pure preheating system O the large thermal performance of the solar collectors and the small heat loss from the auxiliary energy system n are to a certain extent also obtained in systems making use of a so-called smart solar tank. Investigations have thus indicated, Furbo and Shah (1997) that the thermal performance of small SDHW systems can be increased if a smart solar tank is used as the heat storage instead of a marketed combi tank. Fig. 1 shows schematic illustrations of the auxiliary energy supply system of a typical marketed combi tank and of a smart solar tank.

161


marketed solar tank

smart solar tank

3. TESTED SYSTEMS Three small low flow SDHW systems have been tested under the same realistic conditions side-by-side in a laboratory test facility. Two of the systems are based on differently designed smart solar tanks and one system is a traditional system. Fig. 2 shows schematic illustrations of the three systems. All three systems are based on inexpensive vertical mantle tanks used in many domestic hot water systems in Denmark. The solar collector fluid from the solar collector enters the top of the mantle and returns from the bottom of the mantle to the solar collector. Electric heating elements are used as the auxiliary energy supply systems.

J Fig. 1. Schematic illustration of the auxiliary energy supply system of a typical marketed solar tank and of a smart solar tank. In the marketed solar tank the constant top volume of the tank is always heated to a required temperature by the auxiliary energy supply system. In the smart solar tank the auxiliary energy supply system can be built into a side-arm from the middle to the top of the tank. In periods with energy supply from the auxiliary energy supply system heat is transferred from the auxiliary energy supply system to the domestic water in the side-arm. By means of thermosyphoning in the side-arm/tank loop the hot water is transferred to the top of the tank. In that way the tank is heated from the top. The energy supply from the auxiliary energy supply system can be controlled in such a way that the energy content in the top of the tank during all hours can have a predetermined (variable) minimum quantity. In periods with a large hot-water demand the energy content can be large and in periods with a small hot-water demand the energy content can be small. That is: The water volume heated by the auxiliary energy supply system is fitted to the hot-water consumption and the consumption pattern. In most periods the hot top volume is much smaller in the smart solar tank than in the marketed solar tank. The heat loss of the smart solar tank is therefore smaller than the heat loss of the marketed solar tank. Further, the solar volume of the smart solar tank is greater than the solar volume of the marketed solar tank and the thermal performance of the solar heating system is increased if a smart solar tank is used instead of a marketed solar tank.

For the smart solar tanks the electric heating elements are built into a side-arm, which in one system, by means of a plastic pipe from the middle to the bottom of the hot-water tank, connects the middle of the hot-water tank to the top of the hot-water tank. In the other system the side-arm connects the middle of the mantle to the top of the mantle. In periods with the electric heating element in operation heat is transferred from the electric heating element to the domestic water/solar collector fluid in the side-arm. By means of thermosyphoning in the side-arm/hotwater tank loop or in the side-arm/mantle loop the heat is transferred to the domestic water located at the top of the hotwater tank. For increasing duration of the operating time of the electric heating element the volume of the water at the top of the tank heated is increasing. The data of the solar collector used in each of the three systems are given in Table 1 and the data of the tested systems are given in Tables 2 and 3.

Electrlc heatlnc element

/

a r m

electric heating

J

] I: ,.~

heating : ~ element

Cold water

Hot water

Cold water ~

Trad/tional system

Fig. 2. Schematic illustrations of the three tested systems.

J

-

/

A

.~

~

-"

Hot ware

Thermosyphovd~,_m in the eide-erm/tenk loop

Col

Hot water

Thermosypho~tn_m in the eide erm/menUe loop

162


Table 1. Data of the solar collector used in the tested systems. Area

3.00 m 2

Efficiency for small incidence angles

[] Fq0.756 []4.37 FqTm []Ta [30.010 (Tm []Ta) 2 E E

Incidence angle modifier

1-tg3s ([3/2)

Heat capacity

5000 J/K m 2

Tilt

45E]

Orientation

south

Table 2. Data of the solar collector loop and control system used in the tested systems.

Traditional system

Smart solar tank system


Thermosyphoning in the sidearm/tank loop

Thermosyphoning in the sidearm/mantle loop

Solar collector loop Pipe material

Copper

Copper

Copper

Diameter

10/8 mm

10/8 mm

10/8 mm

Length of pipe from solar collector to storage, outdoor

10.0m

10.0m

10.0m

Length of pipe from storage to solar collector, outdoor

13.3 m

13.3 m

13.4 m

Length of pipe from solar collector to storage, indoor

5.1 m

5.1 m

5.1 m

Length of pipe from storage to solar collector, indoor

4.6 m

4.6 m

4.5 m

40% (weight) propylene glycol/water mixture



Volume flow rate in solar collector

10"Sm3/s

10Sm3/s

10Sm3/s

Power of circulation pump

35 W

50 W

65 W

6 K/2 K

6 K/2 K

6 K/2 K

Solar collector fluid

Control system Differential thermostat control with one sensor in the solar collector and one at the bottom of the mantle start/stop difference


163

Table 3. Data of the solar tanks in the tested systems.

Traditional system



Thermosyphoning in the sidearm/tank loop

Thermosyphoning in the sidearm/mantle loop

Solar tank Tank material

Steel St 37-2

Steel St 37-2

Steel St 37-2

0.175 m 3

0.175 m 3

0.175 m 3

1.484 m/0.394 m

1.484 m/0.394 m

1.484 m/0.394 m

0.003 m

0.003 m

0.003 m

0.009 m 3

0.029 m 3

0.058 m 3

0.700 m/0.425 m

0.700 m/0.473 m

1.285/0.473 m

0.002 m

0.002 m

0.002 m

Hot-water tank Domestic water volume Height/diameter Material thickness

Mantle Volume Height/diameter Material thickness Location

The mantle surrounds the hot-water tank. The upper 0.081 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle

The mantle surrounds the hotwater tank. The upper 0.081 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle

The mantle surrounds the hotwater tank. The upper 0.009 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle

Auxiliary energy supply system

Upper 0.072 m 3 of the hotwater tank is heated to 50[X~ by the electric heating element

Side-arm from middle to top of hot-water tank. Electric heating element built into side-arm

Side-arm from middle to top of mantle. Electric heating element built into side-arm

Volume in side-ann: about 0.4 1/min.

Volume in side-arm: about 0.3 l/min.

The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 50[X~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 45[X~ and if the difference between the time and the predetermined tapping hours is smaller than 2 89hours

The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 5 0 ~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 45[X~ and if the difference between the time and the predetermined tapping hours is smaller than 2 89hours

1060 W

1120 W

1140 W

Mineral wool

Mineral wool

Mineral wool

Top

0.25 m

0.25 m

0.25 m

Side

0.05 m

0.05 m

0.05 m

Bottom

0.05 m

0.00 m

0.00 m

Power supply of electric heating element

Insulation Material Thickness

164


4. TEST RESULTS

The three systems were tested side-by-side under the same realistic conditions: The solar irradiance on the collectors and the daily hot-water consumption of 183 1 is the same for all three systems. An energy quantity of 2.44 kWh, corresponding to 61 1 of hot water heated from 10IX2 to 45EE, is tapped from each system three times each day: 7 am, 12 am and 7 pm. The tests were carried out during April-June 1999. So far 8 test periods of 4-5 days' duration have been carried out. The measurements showed that the thermal performance of the traditional solar heating system is always very close to the thermal performance of the smart solar tank system with thermosyphoning in the side-ann/mantle loop. In some periods the traditional system performs best, in some periods the smart solar tank system performs best. The measured net utilized solar energies for the 8 test periods are given in Fig. 3 for the traditional solar heating system and for the smart solar tank system with thermosyphoning in the side-arm/tank loop. The net utilized solar energy is defined as the tapped energy from the solar tank minus the energy supply to the electric heating element.

The smart solar tank system with thermosyphoning in the sideann/tank loop has always a higher thermal performance than the traditional solar heating system. For the whole test period of 35 days' duration the net utilized solar energy was 9% higher for the smart solar tank system than for the traditional solar heating system.

5. CALCULATIONS

A simulation model for the smart solar tank system with thermosyphoning in the side-arm/tank loop has been built up and validated by means of measurements. The yearly thermal performance of small SDHW systems based on differently designed and controlled smart solar tanks were calculated with the model. Also calculations for traditional solar heating systems based on marketed solar tanks were carried out. All the systems taken into calculation are identical except for the solar tank. The data of the system taken into calculation is given in Table 4. The data of the two marketed solar tanks are given in Table 5 and the data of the smart solar tank are given in Table 6. Different daily quantities of hot-water consumption are assumed. The water is heated from 10~ to 50~ A third of the daily hot-water consumption is tapped three times each day. The weather data of the Danish Test Reference Year is used in the calculations. Fig. 4 shows the calculated net utilized solar energy and performance ratio of the systems. The net utilized solar energy is the tapped energy minus the energy supply from the auxiliary energy supply system. The performance ratio is the ratio between the net utilized solar energy for the system with the heat storage in question and the net utilized solar energy for the system with the Danlager 1000 heat storage. The thermal performance of the system is increasing for increasing hot-water consumption. For small hot-water consumption the thermal performance is not strongly influenced by the storage volume of the marketed solar tanks and by the consumption pattern. For increasing hot-water consumption the influence of both the storage volume and the consumption pattern on the thermal performance is increasing. If, for instance, hot water is only tapped during evenings the thermal performance of the systems with the marketed solar tanks is relatively small and the thermal performance is relatively strongly increased by increasing the storage volume.

Fig. 3. Measured net utilized solar energy for the traditional solar heating system and for the smart solar tank system with thermosyphoning in the side-arm/tank loop for 8 periods of 4-5 days' duration.

The thermal performance of the solar heating system with the smart solar tank is higher than the thermal performance of the system with the marketed solar tank with the same total tank volume. The smart tank system can have a thermal performance up to about 30% greater than the thermal performance of the traditional system with the same total tank volume.


The smart solar tank is especially attractive if hot water is only tapped in the evenings. The thermal advantage is somewhat smaller if hot water is only tapped in the mornings and the thermal advantage is smallest if water is tapped in the morning, at noon and in the evening. Fig. 5 shows calculated net utilized solar energies of the system with different smart solar tank designs. The tank is designed as indicated in Table 6 with only one parameter changed at a time. The daily hot-water consumption is 1601 and hot water is tapped at 7 am, 12 am and 7 pm.

165

From the figure it is obvious that the side-arm outlet should be placed as high in the tank as possible with regard to the hotwater comfort and that the control system should stop the supply from the auxiliary energy supply system when the energy content at the top of the tank is as low as possible, of course also with regard to the required hot-water comfort. A variable flow rate in the side-arm resulting in a constant inlet temperature to the tank a little higher than 50[Z; or a high constant flow rate results in the highest thermal performance.

Table 4. Data of the SDHW-system taken into calculation. Solar collector Area

3m 2

Efficiency for small incidence angles

[3 = 0.75 - 5.40. (Tin- Ta)/E

Heat capacity

7000 J/Kin2

Tilt

45 ~

Orientation

South

Solar collector loop Pipe material

Copper

Diameter

12/10 mm

Length of pipe from solar collector to storage, outdoor

1.5 m

Length of pipe from storage to solar collector, outdoor

1.5 m

Length of pipe from solar collector to storage, indoor

3.5 m

Length of pipe from storage to solar collector, indoor

3.5 m

Heat loss coefficient of pipe

0.25 W/mK

Solar collector fluid


Volume flow rate in solar collector loop

7.5-10.6 m3/s

Power of circulation pump

3O W

Control system Differential thermostat control with one sensor in the solar collector and one at the bottom of the mantle Start/Stop difference

10K/2 K


166

Table 5. Data of two marketed solar tanks taken into calculation. Name Type Tank material

Danlager 1000

Daulager 2000

Mantle tank Steel St 37-2

Mantle tank Steel St 37-2

0.189 m 3 1.082/0.500 m 0.003 m Upper 0.080 m 3 of the hot-water tank is heated to 50.5~ by the auxiliary energy supply system

0.265 m 3 1.450/0.500 m 0.003 m Upper 0.089 m 3 of the hot-water tank is heated to 50.5~ by the auxiliary energy supply system

0.007 m a 0.395/0.525 m 0.003 m The mantle surrounds the bottom part of the hot-water tank. The upper 0.094 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle

0.012 m 3 0.715/0.525 m 0.003 m The mantle surrounds the bottom part of the hot-water tank. The upper 0.109 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle

PUR-foam 1.7 W/K

PUR-foam 2.3 W/K

Hot-water tank Volume Height/diameter Material thickness Auxiliary energy supply system

Mantle Volume Height/diameter Material thickness Location

Insulation Insulation material Heat loss coefficient

Table 6. Data of the smart solar tank taken into calculation Mantle tank Steel St 37-2

Type Tank material

Hot-water tank 0.189 m 3 1.082/0.500 m 0.003 m

Volume Height/diameter Material thickness

Mantle 0.007 m 3 0.395/0.525 m 0.003 m The mantle surrounds the vertical sides of the hot-water tank. The upper 0.094 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle.

Volume Height/diameter Material thickness Location

Side-arm and auxiliary energy supply system Side-arm location Auxiliary energy supply system Power of electric heating element Volume flow rate in side-arm in periods with energy supply Control system

801 located above the side-arm's outlet pipe. Side-arm inlet connected to top of hot-water tank Electric heating element built into the side-arm 1200W 6.67[:106m3/s The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 50~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 50[]2, and if the difference between the time and the predetermined tapping hours is smaller than 2.5 h.

Insulation Insulation material Insulation thickness

Top Side Bottom

PUR foam 0.10m 0.05 m 0.05 m


167

Hot water tapped at 7am, noon and 7pm

Hot water tapped at 7am, noon and 7pm 1.3

-.- o a . ~ e ,

1300 @

1200

C9

1100

,-'C'

lOOOl

_o .m

_~ ~1ooo

1.2

1.15

o

>,

o

-.4-Danlager 1000 -4,--Danlager2000 Smart tank

1.25 -*- Smart tank

~

1.1 1.05

_m Q

9

O.

600

z

0.95 4oo 0

0.9

,

~

,

50

100

150

0

200

50

150

200

Hot water tapped at 5am, 6am and 7am

Hot w a t e r tapped at 5am, 6am and 7am 1.3

14oo

-o- Danlager 1000 Danlager 2000 -*-" Smart tank

1300 1200

0

100

Hot water consumption [I/day]

Hot water consumption [i/day]

-~- Danlager 1000 -,,- Danlager2000 -*- Smarttank

1.25

_o .m

r ,~,11oo (g ~ 1 0 0 0

1.2

1.15

Q

~

1.1

m

[E 1.115 700 9 Z

o

600

0.95

5OO 4oo 0

,

,

,

50

100

150

0.9 200

0

,

i

,

50

100

150


200


Hot water tapped at 7pm, 8pm and 9pm

Hot w a t e r tapped at 7pm, 8pm and 9pm 1.3 Danlager 1000 -'- Danlager 20001 -dr- Smart tank

1300 1200

_~..~~

Z

O

r Q

--=-Danlager I000 --- Danlager2000 -,-- Smarttank

1.25

O

11oo

,.,~, (g =1000 O >~

~

1.2

1.15

o

~

1.1

m 9

800

~ 1.05

9 Z

600

a.

a

0.95

500 40o 0

,

l

,

50

100

150


0.9 200

0

50

100

150

200


Fig. 4. Yearly net utilized solar energy and performance ratio as a function of the hot-water consumption for different consumption patterns and heat storage.


168

1350

>,

1300

cO Q

1250

e==

1200

m : 4,,

1100

Z

4O5O

0

1000 0.4

1300 1250

~'~

1200

=

1100

0 Z

0.45

0.5

0.55 0.6 0.65 0.7 M a n t l e h e i g h t [m]

0.75

1350 / ..._.-~-----, ,_.____..__ ,._ c9 o

9 0

K,7

0.8

1050 1000

0.65

0.2 0.4 0.6 0.8 1 Side-arm outlet pipe's distance from t o p o f tank [m]

Constant hl~t temperatureto the hot water tank fn:xn the side-arm

o

1250

i: 4)

1250

,Iz

ID

m "~

1100

Z

=

1100

o Z

4O50

1050

1000 50 2 4 6 8 10 Power of auxiliary energy s u p p l y system [kW]

65

70

75

80

85

90

95

1350

>~

1300

ro Q

1250

.ram m

1200

m ~

1100

u

Z

50

Inlet t e m p e r a t u r e t o t h e h o t w a t e r t a n k f r o m t h e s i d e - a r m [ ' C ]

1350

=3

55

1200

J

7

1100

Q

1050

Z

1000 1 2 3 4 5 Lower limit of energy content of consumption water with t e m p e r a t u r e s h i g h e r t h a n 50=C [ k W h ]

6

1050 1000 0.3

0.5

0.7

0.9

1.1

1.3

1350

~)

1250

0 C Q

~'~'U

1200

~'~

1150

m 1050 z

lOO0 95o 0

50

1.5

1.7

V o l u m e flow rate in side-arm [I/min]

100

150 Tank Volume

200

250

300

[I]

Fig. 5. Yearly net utilized solar energy of the solar heating system with differently designed smart solar tanks.

1.9

2.1


6. DISCUSSION AND CONCLUSION

Investigations of smart solar tanks based on a mantle tank with a side-arm in which an auxiliary energy supply system is built in have been carried out. Both tanks with the side-arm connected to the hot-water tank and to the mantle have been investigated. The investigations showed that the tank with the side-arm connected to the hot-water tank is the best tank from a thermal point of view. Further, the investigations showed that the thermal performance of solar heating systems can be improved by up to about 30% by making use of such a smart solar tank. The thermal advantage of smart solar tanks is largest if hot water is not tapped during the light hours of the day. It is estimated that the costs of a typical small SDHW system will be increased by about 3% by making use of a smart solar tank. Consequently the performance/cost ratio can be improved by up to about 25% by making use of a smart solar tank. So far, detailed investigations have only been carried out for smart solar tanks with the auxiliary energy supply system built into a side-arm. However, smart tanks can be designed differently. Preliminary investigations indicate that the auxiliary energy supply system can be built into the hot-water tank in such a way that thermal stratification is built up as good as or even more efficiently than in tanks with the side-arm connected to the hot-water tank. These investigations will be finished during the summer of 1999. From the autumn of 1999 the two most promising designs of smart solar tanks will be tested in two small SDHW systems in practice. The thermal performance of the systems will be measured during the first year of operation. In this way possible operation or durability problems connected to the designs of smart solar tanks will be elucidated. It will also be elucidated if the hot-water comfort can be accepted by the consumers and if the thermal performances of the systems in practice are as good as expected. Based on the promising results it is recommended to start work to develop smart solar tanks based on other auxiliary energy supply systems than electric heating elements. In order to further improve smart solar tank systems it is also recommended to start work to develop a smart control system both for the energy supply from the auxiliary energy supply system and for the pump in the solar collector loop. Most likely, the system performance can be somewhat increased if the flow rate in the solar collector loop is controlled in such a way that water in the top of the hot-water tank is heated by the solar collector to a directly usable temperature. The flow rate in the solar collector loop will therefore vary from one period to another.

169

REFERENCES

Furbo S. and Shah L.J. (1996) Optimum solar collector fluid flow rates. EuroSun '96. 10. Internationales Sonnenforum Proceedings, Freiburg, Germany. Book 1, 189-193.

Furbo S. and Shah L.J. (1997) Smart Solar Tanks [3 Heat Storage of the Future? Proceedings of ISES 1997 Solar World Congress, Taejon, Korea.

Otto W, Nielsen J.E. and Dalsgaard Jacobsen T. (1997) Ydelsesstatistik for mindre brugsvandsanl~eg [3 erfaringer fra det femte ~ s rrfilinger 1996. Danish Solar Energy Testing Laboratory.

Shariah A.M. and L6f G.O.G. (1997) Effects of auxiliary heater on annual performance of thermosyphon solar water heater simulated under variable operation conditions. Solar Energy 60, 119-126.

170


THERMAL

MODELLING

AND PERFORMANCE

PROCESSES

UNDER

PREDICTION

OF DRYING

OPEN-SUN-DRYING

H.P. Gar.q and Rakesh Kumar Centre for Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi - 110 016, India Tel.: +91-11-6861977, Fax: +91-11-6862037, E-mail: [email protected] Abstract - An analytical model has been developed for the drying characteristics of any product under open-sundrying(OSD). The model is based on the theory of 'Generalised drying curve'(GDC) in terms of receding front. The developed model can be used to ascertain the drying characteristics of any product under OSD. The model has included all the climatic and product parameters explicitly. The main concern of the present study is to estimate the effect of product thickness and climatic conditions on the drying rate and drying time. The numerical calculations have been made for the climate of Delhi. For the calculations, the chosen crop is grapes(initial moisture: 82%). The results have been plotted for both s u m m e ~ a y ) and winter(December) conditions of Delhi. It is noticed that under open-sun-drying conditions, the drying rate depends significantly on the product thickness and climatic conditions. The thickness of the product is taken(or kept) as small as possible for fast and quality drying under OSD. Also, the developed model for drying characteristic under OSD has been validated with published experimental observations on typical crop under Delhi climatic conditions.

1. INTRODUCTION

final moisture:

12%). The results are plotted for both

summer(May) and winter(December) conditions of Delhi. It is The most widely used method for crop drying is open-sundrying(OSD). This is simplest method for product dehydration. In OSD, the product is spread in a thin layer on the horizontal ground and exposed directly to the solar radiation, wind and other atmospheric conditions. In this type of drying, heat is transferred from the surrounding air and the sun to the exposed stnface of the product. A part of this heat is travelled to the product interior to rise its temperature and remove the moisture from the product interior to its surface. The remaining heat is utilized to evaporate the moisture from the product surface to the SUl'rounding air. This process of heat and mass transfer has occurred simultaneously in OSD. The rate of drying depends on the number of external parameters(solar radiation, ambient temperature, wind velocity and relative humidity) and the internal parameters(initial moisture content, type of the crop, mass of the product per unit exposed

seen that under OSD conditions, the drying rate depends on the product thickness and climatic conditions. The thickness of the product is taken(or kept) as small as possible( from 2-4 cm) for fast and quality drying. The present model on OSD has been validated with the experimental results of typical crop(mango) under Delhi climatic conditions. 2. MATHEMATICAL ANALYSIS

area, etc.). Some theoretical and experimental study on OSD are reported in the literature (Garg, 1987). However, the available study is not sufficient for full understanding of the drying processes in OSD. The principle of drying in OSD is different from solar drying. There are enough literatures available on solar drying (Muhlbauer, 1981, 1986; Garg and Kumar, 1998; Cfarg et al, 1998) whereas, very few work has been done on OSD. Sodha et al (1985) has developed an analytical model for OSD. The developed model has considered a variation in the product temperature along its thickness and has applied thin layer drying equation for moisture evaporation. The analysis has not taken into account the heat capacity effect of the product, product thickness and the quantity of the product per unit exposed area to the sm'rounding air. In this paper, the analysis and the numerical results for a typical crop are presented under OSD. The developed model has used the receding front phenomena in which product can be considered into two parts, viz wet part and dry part. Earlier, the study on 'Generalised drying curve'(GDC) was proposed by Ratti and Crapiste (1992) and extended by Chou et al (1997). The present study is the extension of the work of Chou et al. Also, the model has included all the climatic and product parameters explicitly. The developed model can be used to identify the drying characteristic of any product under OSD. The numerical calculations have been made for the Delhi climate. For the calculations the chosen crop is grapes(initial moisture: 82% and

Product 9

-

L .....

~

r,-,

lffftt

Lo..

; :~

(a) Fig. l(a) Schematic of open-sun-drying procedure The schematic of open-sun-drying(OSD) is shown in Fig. 1.. Following assumptions are made in the present analysis: (1) Thermal properties of dry product is constant. (2) The product is considered as a material of uniform thickness. (3) There is no heat conduction in the product slab. (4) No volume shrinkage of dried product. (5) No temperature and moisture gradient along the product thickness. (6) The temperature in the ground at 4 m depth is constant and taken as 24 ~ for Delhi conditions. The heat and mass balance in the OSD is expressed in terms of the following equations.


171

2.1 Equation for moisture evaporation The present study is based on the concept of receding front phenomena. Ratti and Crapiste has developed this concept on the basis of 'Generalized Drying Curve'(GDC). In this model, the GDC is independent on the drying conditions, type of the product to be dried and is a function of the moisture content in the product. This model is fuaher simplified by Chou et al and used in the present study. In this model, the product has been considered into two parts viz., dry part and wet part. These two parts are separated by an arbitrary curve known as receding front. During the drying process, the receding front moves from dry part to wet part. The position of the receding front at any moment of the drying is shown in Fig. l(b).

1=11+12+13+14

(6)

in which I is solar radiation absorbed by the product, I1 heat losses to the product due to convection and radiation, 12 heat stored as a sensible heat in the product, 13 heat flux losses to the ambient due to evaporation and !4 heat flux conducted into the ground. The above Eq. in terms of different temperatures on simplification is written as,

Tp ( ~TL) ( "-~-) ~M . mpcp'-'~'-=Qp-hg(Tp-Ta)-hfg t)TG/y'aPo -KG'--~- =

h

2.3 Equation for ground temperature The temperature variation inside the ground To is characteristic of the heat conduction Eq. as,

Ms -x ----_____

zl

(7)

_~_x

Wet part

/

O2TG igTG K G ~)y2 = P G CG Ot

_=Wsat

(8)

Mo This conduction Eq. is applicable with the following boundary conditions,

(b) Fig. l(b) Schematic of receding front model

Tp/y=o= TG/y=O ,

The moisture profile in the product at any time is represented by the following relations,

To get the instantaneous values of various parameters (moisture content, product temperature, ground temperature), the developed heat, mass and temperature balance equations are solved by using finite difference technique. Equations for moisture evaporation, product temperature and ground temperature are re-written as, For moisture content,

M=Mo

O> Vstore, and p , v

(1)

~ = ~(q~,~stor,d,...),

199

Fig. 8. Measurement of the DHW-temperature, the angle of the valve, and the DHW-flow rate. The time resolution is 5 s.

On the left hand side, the driving forces for the storage water flow in the discharge unit are expressed by the density differences of the water columns shown in figure 1. The fi'ictional forces on the right hand side are described by the kinetic pressure drop of the storage water. The pressure drop coefficient is a function of the geometry of the discharge unit, the angle of the valve, as well as the flow rate of the storage water. Due to the fact that the velocity of the storage water in the vertical pipe is much higher than the upwards velocity of the storage water in the tank, the second term on the fight hand side in Eq. (1) can be neglected. For stationary conditions there is a linear correlation between the angle of the valve and the temperature of the expansible material (fig. 6a). The time delay T occuring for the reaction of the valve to a change in the domestic water outlet temperature, caused by the heat capacity and conductivity of the expansible material, can be described by Eq. (2). "Cq~+ ~ = k ( Tdom,out - T b )

(2)

The value for T was determined experimentally by the response of the angle of the valve to step changes of Tao,,~o.t(fig. 6b). The heat exchanger is modeled with the energy balance (Eq. (3a) and (3b)). Whereas the differentiation of the time dependent variable was carded out by the simulation tool SMILE, a discretization in the space coordinates was necessary. A one-dimensional approach with 20 nodes for each fluid, using backward differences was applied. In this way, the heat exchanger was treated as a counter flow heat exchanger.

(3)

and ~" : density, velocity, and pressure drop

coefficient, respectively, of the storage water outlet temperature below the heat exchanger in the vertical pipe (index: pipe) and of the surrounding storage water in the tank (index: store).

C ~om .Ord~.

bt

i

~TsLr

C stor .

Ot

. UA. ( r~o m

. = v A . ( rAom.

i ) - J~dom Cp "( rdom - r~om) i-1 rstor

r;Lr . )+

,,orC .

(r;'to

200


Fig. 9. Calculated DHW-temperature during a discharge. Initial storage temperature: 60~ domestic water flow rate of 5, 10, and 201/min.

In the model, the vertical pipe was divided into four nodes. Since the valve causes the storage water flow rate to approximate the domestic water flow rate, the heat capacity rate (UA) was described as a function of the domestic water flow rate, solely. Validation of the model is shown in figure 7. The DHWtemperature and the storage water temperature measured during a discharge at the top and at the bottom of the vertical tube (at the positions a, b, and c in figure 3) are shown. The time delay of the temperature drop at the top and at the bottom of the pipe as well as the values of the temperatures show good correspondence. 4--

PROTOTYPE CONTROL DEVICE OF THE STORAGE WATER FLOW RATE

The time delays, arising due to the heat capacities of the system's components, cause oscillations of the DHWtemperature at the beginning of a discharge. A measurement made with a prototype of the valve is shown in figure 8. The flow rate of the domestic water was about 15 l/rain. For the measurements, no mixing devices at the DHW-outlet were employed, which are usually applied in solar water heating systems (figure 2c). At the be~nning of the draw-off, the domestic water outlet temperature has nearly the same value as the storage water temperature in the containment, as long as the water content of the heat exchanger is taken. Meanwhile, the storage water inside the containment is cooled, since there is only a small flow rate of the storage water due to a high pressure drop caused by the valve. This leads to a minimum of the outlet temperature, 5 K below the value in equilibrium. In the following, the influence of parameters concerning the valve's operation will be discussed.

4.1 Domestic waterflow rate The influence of the DHW-flow rate on the minimum DHWtemperature T=~, is shown in figure 9: The higher the DHWflow rate for equal initial storage temperature distributions, the higher the amplitude of the oscillation. The DHW-temperature under stationary conditions T ~ increases, if

Fig. 10. Calculated DHW-temperature with and without mixer. Without the mixer, the flow rate in the heat exchanger is constant, whereas with use of the mixer, the total DHW-flow rate discharged is constant. In this case a variing part of the water is bypassing the heat exchanger.

the DHW-flow rate decreases. This also means, that the higher the flow rate, the less energy is inserted into the bottom of the storage tank during a discharge.

4.2 Storage water temperature dis~'bution Although there is a high amplitude of the oscillation (T~,Train) for high storage temperatures and high flow rates, the absolute value of T~. is smaller for low than for high storage temperatures (for fully mixed tank): T~.(95~

> T~.(60~

for I ~ m < 22 1 / min

However, if two or more different temperature layers are taken into account, no uniform tendency can be formulated concerning the behaviour of the minimum value of the DHWtemperature in dependence of the mean storage temperature. At a homogeneous storage temperature of 95~ and a flow rate of l0 1/min the curve is critically damped, for smaller flow rates the curves are overdamped. Due to the influence of the storage temperatures on the oscillations, the set temperature for the auxiliary heater always has to be considered, when the behaviour of the valve is investigated.

Fig. 11. Calculated DHW-temperature oscillations for different time constants: x = 20 s, 10 s, and 0 s (theoretical case) and with a differential component (PD). Initial temperature of the storage: 60~ DHW-flow rate: 15 l/min.


4.3 Mixer Measurements done with a standard mixing device show that temperatures above the DHW-set temperature can be smoothed, temperatures below the DHW-set temperatures, however, can not be avoided. Simulations done with two different flow rates show, that Tmi, only increases slightly, whereas the time, at which the minimum occurs, is shifted (fig. 10). 4.4 Time constant The influence of the time delay of the regulation loop caused by the heat capacity of the expansible material is shown in figure 11. The simulations show, that the values for Tmi, can be decreased distinctly by a reduction of the time constant and an improvement of the heat transfer beween the DHW and the expansible material. Other than that, the oscillations could be avoided with an ideal regulation device, containing a differential (e. g. electrical)component. 4.5 Pressure drop characteristic Furthermore, the pressure drop characteristic, resulting from the geometry of the vertical pipe and the containment, influences the oscillations. Simulations show, that with a change of the pressure drop characteristic, the oscillations can be avoided (fig. 12). However when the oscillations are reduced, the values of the storage water flow rate, the storage water outlet temperature, and T~q= rise. Therefore, there is a conflict of aims between the efficiency of the valve, influenced by the storage water outlet temperature and the comfort for the user due to a well damped course of the DHW-temperature.

5.

CONCLUSIONS

A regulation device for the density driven storage water flow of the discharge unit of a combistore was investigated. Measurements showed, that the stratification can be improved decisively with the regulation device during a discharge cycle. The investigations presented were focused on DHWtemperature oscillations caused by the regulation device. The oscillations can be described well by a model, composed of the one-dimensional pressure drop equation, a differential equation to decribe the dynamic behaviour of the valve, and a 20 node heat exchanger model. The validation of the model showed good accuracy between measured and calculated values. The influence of the domestic water flow rate and the temperature distribution on the oscillations were shown. The temperatures at the top of the storage came out to play an important role for the oscillations. With a mixer, which is usually applied in a solar system, the DHW-minimum temperature could not be increased perceptibly. Values which directly influence the oscillations are the time constant, describing the time delay of the valve due to the heat capacity and conductivity of the expansible material, and the pressure drop characteristic of the storage water flow in the vertical pipe. The time constant of the valve's reaction could be reduced by more than 50 % compared to the prototype. By changing the pressure drop charcteristic, a sufficiently damped course of the DHW-temperature can now be reached and the oscillations can be avoided. Furthermore, the simulations showed that there is a conflict of aims between the energy efficiency reached by the system

0 o

=3

57

----------

55

............................

201

53

t~ L_ ID

51

E

49

............................

(D "*7'

47

............................

I

45

prototype

121 43

~

0

30

60

90

120

150

180

210

time / s

Fig. 12. Calculated DHW-temperature for different pressure drop characteristics of the valve. Initial storage temperature: 60~ DHW-flow rate: 10 l/min.

and an extensive reduction of oscillations. Latter can be reached by a higher auxiliary temperature or by changes of the pressure drop characteristic. NOMENCLATURE % C d g h k

specific heat capacity, J/gK heat capacity, J/K diameter, m gravity constant, rn/s2 storage height, m transfer coefficient, 1/K

~: T UA v

mass flow rate, kg/s temperature, K heat transfer coefficient area, W/K velocity, m/s

I&

volume flow rate, mVs

Greek P (o T

Indices b dom, out dom equ min pipe stor store

density, kg/m3 angle of the valve, o time constant of the expansible material, s pressure drop coefficient

begin of the regulator temperature interval of the valve domestic water outlet domestic water in the heat exchanger nodes equillibrium at minimum of oscillation inside the vertical pipe storage water in the heat exchanger nodes storage water layers

202


REFERENCES

Dahl, S. D., Davidson, J. H. (1998). Mixed Convection Heat Transfer and Pressure Drop Correlations for Tube-In-Shell Thermosyphon Heat Exchangers with Uniform Heat Flux. Journal of Solar Energy Engineering. Vol. 120, pp. 260269. Dehmel, K.-H., Klein-Robbehaar, C. (1997) Universit~rer Forschungsschwerpunkt 4: Entwicklung rechnergestiitzter Simulationshilfsmittel zur Beschreibung des Betriebsverhaltens komplexer energiewandelnder Systeme, Berlin. Driick, H., Hahne, E. (1998) Test and Comparison of Hot Water Stores for Solar Combisystems. In Proceedings of EuroSun ISES Europe Solar Congress, 14-17 Sept. 98, Portoroz, Slovenia, Vol. 2, III.3.3. Dahm, J., Bales, Ch., Lorenz, IC (1998). Evaluation of Storage Configurations with Internal Heat Exchangers, Solar Energy, Vol. 62, No. 6, pp. 407-417. Leibfried, U. (1998) Kombispeicher mit ThermosiphonWErmetauschem ftir Warmwasser. In Proceedings, 8. Symposium thermische Solarenergie, 13-15 May 98, Staffelstein, Germany, pp. 39-44. Phillips, W. F., Dave, R. N. (1982). Effects of Stratification on the Performance of Liquid-Based Solar Heating Systems. Solar Energy. Vol. 29, No. 2, pp. 111-120. Sharp, M. K., Loehrke, R. I. (1979). Stratified Thermal Storage in Residential Solar Energy Applications. Solar Energy. Vol. 3, No. 2, pp. 106-113.


203

PERFORMANCE OF TRANSPARENTLY INSULATED SOLAR PASSIVE HOT WATER SYSTEMS N a r f n d r a D. Kaushika and Kalvala S. Reddy Centre fi)r Energy Studies, Indian Institute of Technology Delhi, Hauz Khas, New Delhi - 110 016, India, Ph. +91 -i 1-686 ! 977 Ex. 5006. Fax" 9 l- i 1-6862037, [email protected]

Ab~lract - This paper presents the design and performance characteristics of TIM insulated solar ICS water heaters with water/ground/ .,,and and concrete as storage materials; it uses computer simulation approach based on accurate determinations of the criteria of convection suppression, solar transmittance and thermal loss reduction characteristic.,, of TIM device. The simulation model is validated with prototype field experimental ob.~rvations and is used to evolve the optimum system design and trade-off characteristic.,,. The TIM cover system characteristics significantly influence the overall system perii~rmance. The effectiveness of several configurations of TIM cover system as a comparative study, has therefore been investigated. The results seem to favour absorber perpendicular (honeycomb) configuration over others. The system performance tends to level at a honeycomb cover depth of 7.5 cm. Compounding of honeycomb w~th an air layer 112 mm thicknes.,,) tends to improve the performance, the honeycomb cover depth of 5 cm is near optimum in this c~mliguration. The absorber parallel configuration is simple in practical realisation: it may be recommended for application in passive .~olar water preheaters. TIM insulated ground integrated collector storage water heating system has also been investigated.The system consists of a network of pipes embedded in a concrete slab whose top surface is blackened and covered with TIM device and bottom is insulated by the ground. Solar gain (solar collection efficiency of 30-50 ch corresponding to temperature of 40-60 ~ and diurnal heat storage characteristics of the system are found to be of the right order of magnitude for solar air/water heating application.,, 1. INTRODUCTION For many years now air-filled honeycomb devices have been considered for use as transparent insulation in solar ponds ( Lin. 1982; Ortabasi et al.. 1983; Kaushika et al.. 1983:Sharma and Kaushika. 1987:Schaefer and Lowre). 1992). hot water systems (Kaushika and Banerjee. 1983; Goetzberger. 1984; Rommel et al.. 1987). buildings (Goetzberger. 1984: Kaushika et al.. 1992) and other integrated collector-storage systems (Gordon. 1987:Kaushika et al. 1990). These applications have opened up a .,,ignificant area of developing transparent insulation matenals in parallel with the conventional(opaque) insulation materials.The configurations based on the storage water tank seem yen.' suitable for domestic and industrial applications. Kaushika and Banerjee (1983) suggested and analysed the configuration, which consists ,)t storage tank 'Cul'rt)id' in shape, transparently insulated at the surface and covered with opaque insulation at all other sides. Subsequently, Goetzberger and Rommel(1987) examined the prospects of such a system for application in central Europe. A cubic storage water tank using transparent insulation at its surface as well as side walls (Kaushika and Sharma. 1994) and a simulated well stratilied tank made of tubular subunits (Schmidt et al. 1988: Schmidt and Goetzberger. 1990) have also been considered and solar energy gain up to 40% has been reported. The ICS solar water heater having water tank of triangular cross-section and with transparent insulation such as Methyl Methacrylate(MMA) on the top and sides has been studied by Prakash et al. (1994). The above thermal analyses/evaluations of the performance of transparently insulated ICS solar water heaters are in general based on discrete measurements of solar transmittance and heat losses across the TIM cover system. In simulation models very little attention has been paid to the tbrmulations of ~lar beam and diffuse radiation transmittance and thermal loss reduction characteristics of the TIM cover system. This paper is Intended t() present design and perlormance data ~dTIM insulated solar ICS ~ater heater.,, with water, ground. ~and and concrele a.,, ,,forage material.,,: it u.~s c()mputer simulation approach ba.,,ed on accurate deternunatlons of the criteria ol convection suppresslt)n, solar

transmittance and thermal loss reduction characteristic.,, of TIM device. 2. T H E S I M U L A T I O N M O D E L Consider a TIM insulated integrated-collector-storage solar water heater which involves the solar heating of storage water tank. cuboid in shape, having TIM cover on the top surface and opaque insulation on all other surfaces. The solar radiation, after transmission through the TIM cover is absorbed by the top surface ()f tile tank. Part of the absorbed energy is used to heat the water and the remaining energy being lost to the surroundings by conduction, convection and radiation. The energy balance Ior water can be written as:

dTw(t) M

-

S '(t)- Qt.(t)

(1)

dt or

dT(t)

(2)

+ E T ,(t)= F(t)

dt The total heat loss(QL(t) ) from the storage water is given as:

Qc(t)

=

U L[T(t)-

T(t)]

(3)

Where U L is overall heat loss coefficient and is expressed as: U,_ = UT+UR+U.~ (..I)

u,.

E:--, M and

F(t)=

s :(t). U,.T(:t)

M-MC

M .M,C,

Eq..(2) is linear differential equation with integration factor eEr. Applying initial condition T.(t) = T . , at t = O. At small interval of

ISES Solar World Congress 1999, Volume Iii

204

time(t). F(t) may be regarded as constant (F). The solution is obtained as : l

T(t):

~-F[I-c

-El']

- T

e

.wr

(5)

This expression may be used to calculate mean water temperature as a function of time. The radiant energy(S'(t)) reaching the absorber plane at time t is given by:

S '(t) - Ib(t) 1~ (z a)b 9 I,(t) R, (, a). 9 [Ib(t) - l,(t)] R (x a ) .

(6)

The fiwmulations of transmittance-absorbtance products corresponding to beam radiation (I: a)h. sky diffuse radiation ( 1:a )a, and ground diffuse radiation ( z a ) ~ for the cover system embodying the TIM device were discussed in Kausika and Arulanantham ( 1996)and Kaushika and Reddy (1999) and have been adopted in the present work. The UL values may be computed by using the concept of thermal network.The steady-state energy transfer between the absorber surface at 1", and the bottom cover of TIM at T, invoives the heat transfer (a) through TIM (compound honeycomb), (b) between top of the TIM and tempered glass cover and (c) between tempered glass cover and ambient air.The collection efficiency of the system ts the ratio of energy collected by water mass to solar radiation received on absorber plane during time t and, is given by :

the bolt on cross-bar according to the requirement of placing the absorber plane perpendicular to solar rays at noon. The top surface of the tank is painted with black paint to absorb the .q)lar radiation and covered with TIM to reduce the heat losses. The temperature of water inside the tank is measured by Copper-Constantan(typeT, Copper(+)/Constantanr thermocouples. The CopperConstantan thermocouple wires are arranged in the form of a multichannel probe, wherein the junctions are placed at a separation of 10cm to measure the vertical temperature distribution in the water tank. The thermocouple probe is connected to a SC-7501 multi logger (IWATSU ELECTRIC Ltd., Japan). The water temperatures have been recorded at an interval of one hour. The water was not drained out from the tank for two days. The time-history of temperature development in the tank at various heights was recorded.The experiments were carried out tn December ( Dec. 11-13, 1996, a winter month) at New Delhi. The temperature gradient builds up during the day and tends to diffuse during the night. The system exhibits significant retention of heat dunng off-sunshine hours. The experimental variation of mean temperature of water in the tank ts compared with simulation results which indicate an excellent agreement between the experimental observations and simulation model results (Fig. ! ).

t

fQ.(t)dt

q_

(7)

o

t

A fs(t)dt 0

3. EXPERIMENTAl, VALIDATION O F SIMULATION MODEl,

3. ! Fabrication of TIM Cover System The TIM cover system for the proposed water heater is fabricated from extruded cellular strips supplied by ArEI Energy Lid, Israel. The product is in the form of cellular strips of (70 x i.6 x 5 cm) and (70 x !.6 x 10 cm) sizes. The square cells are of widths 3 mm and 4 ram. The strips were glued to form a cellular array. The gluing process was carried out manually using pure liquid Chloroform (CHCI~, which is a solvent for lexan. Agarwal Chemical Industries, New Delhi) as adhesive. The cellular matrix was finally encapsulated in a tray made of transparent polycarbonate sheet of 0.5ram thickness. In the covet" system, an air layer of 12ram is maintained by placing the structured sheet ribs in the bottom of the tray.The cover system is easy to handle and has sufficient built-in strength to maintain rigidity 3.2 Experimental Set-Up A prototype field experiment system has a tank of 25 Iitrcs (66 x 45 x 8.5 cm) with a rectangular cross-section; it is made from an 18 gauge galvaniscd iron sheet, it is covered with mineral wool insulation at its sides and bottom and encased in a w(mdcn box. It ts kept in an inclined position and facing due south. The inclination of the system can be adjusted manually by changing

Fig. 1 4. SYSTEM O P T I M I S A T I O N AND TRADE-OFFS Experimentally validated simulation model may be used for the derivation of optimum design parameters and trade off characteristics. For a given solar absorber area, the temperature in the tank will vary with the capacity of the tank. These variations of the storage water temperature as a function of capacity of the tank, selective and black absorber characteristics, cell width and thickness of TIM configuration are portrayed in Figs.2 (a, b). The results may be used for trade off between the system efficiency and required hot water temperature. The effectiveness of a honeycomb cover system on the development of mean water temperature in the tank has also been investigated. Two configurations of the cover system are considered: (i) honeycomb cover system which consists of the encapsulated cellular array (ii) compound honeycomb cover


205

5. PERFORMANCE COMPARISONS

Fig. 2 (a)

The ~lar transmittance and heat loss reduction characteristics of TIM cover system considerably influence the overall thermal performance of ICS solar water heating system. We therefore investigate the effectiveness of various kinds of TIM cover systems. The cover system consists of TIM device compounded with mr layers of near critical Rayleigh regime at its top and/or bottom. Following configurations of TIM device placed between the top tempered glass ( S m m thickness) cover and the absorber have been considered. I. Absorber-Parallel Configuration of T I M Device (CI) Air Layer (near criticalRaylcigh regime) - single cover (C2) Glass Sheet (2 m m thickness) (C3) Polycarbonate Sheet (0.5mm thickness) (C4) Double Wall Structured Polycarbonate Material (6ram thickness, G E Plastics) (C5) Double Wall Structured Polycarbonate Material (10ram thickness. G E Plastics) 2. Absorber- Perpendicular Configuration of T I M Device (C6) Cellular Array(Honeycomb) of 5 cm thickness ~C7) Encapsulated Cellular Array of 5 c m thickness compounded with 12 m m air layer (C8) Encapsulated Cellular Array of 10 cm thickne.,,.,, compounded with 12 mm air layer All the configurations have been experimentally tested as well as analysed by the simulation model. The mean temperature of water in the storage tank as measured experimentally as well as predicted from the simulation model is considered. A comparison of performance characteristics of absorber parallel configurations of TIM in terms of solar gain efficiency defined by equation (7) is presented in Table-!. Results indicate an excellent agreement between the experimental ob~rvations and simulation model results. The performance of solar ICS water heater with cover system embodying a double walled structured sheet of 10ram (GE plastic product) as TIM excels over others. The comparison of performance characteristics of ~lar ICS water heater with absorber- perpendicular structures of TIM cover systems is summarised in Table-2. The cover system C7. made of comp()und honeycomb of 5 cm thickness and having 12 mm air laycl ~ts top and bottom corresponds to relatively higher solar collection-storage efficiency of the water heater. Table-i Performancecomparison of transparently insulated (absorberparallel structure) solar ICS water heater. r

Fig. 2 (b)

syslem which consists of an encapsulated air layer (12mm thickness placed in the bottom region) and the cellular array; in this geomeffy the air layer remains in near critical Rayleigh regime and provides additional insulation without affcclfing the solar transmittance of the cover system. The system performance tends to level at a honeycomb cover depth of 7.5cm. Compounding of honeycomb with an air layer tends to improvc the ped'ormance: the honeycomb cover depth of 5cm is near optimum in this configuration.

Dine of Exp.

Insolation kWh/day-

TIM

P e r f ~ c c Characteristics

m"

28-I 1-96 01-12-96 30-I 1-96 02-12-% 13-05-98

5.28 5.96 5.91 6.32 6.15

CI C2 C3 C4 C5

T,,

T,,,~

T,,

qc %

14.2 14.0 15.3 15.5 29.0

47.7 52.8 51.2 50.5 56.0

23.1 29.5 31 7 31.4 47.0

14.33 22. II 23.59 21.40 30.97


206

Table-2 Performance comparison of Tmnspareutl), insulated (absorber- perpendicular structure) solar ICS water heater.

/ Dme of Exp.

Insolation kWh/dayI112

TIM

Performance Charmm27 May | 1999sties J Storage water temp. 'C

I 1-12-9~

5.49

C6

09-12-~

5.72

C7

05-,2-9(

5.25

C8

T.,

r.~

T.~

9.2 9.0 li.0

47.6 ~,5., 39.3

31.8 33.1 Y).7

(qc) qt

on a winter day at New Delhi. The adjustment of the inclination is often considered as a liability by the users, in the cuboid shape, the tanks of large capacity may also be used: they may be segmented or made from tubular subunits to simulate well-stratified tanks of good structural stability. In fixed tilt configurations several design variations are possible. For example one could use a triangular shaped tank wherein the honeycomb and absorber plane are inclined say perpendicular to sun rays at winter noon.

35.35 35.82 31.10

in Fig.3 the diurnal variations of storage water temperature as obtained from the validated simulation model for all the configurations arc portrayed. All the variations correspond to same radiation and atmospheric air data shown therein. These results further support the inferences of preceding sections. The relattve merits of various TIM cover systems can also be judged l?om their ~)lar transmittance (': a) and heat loss reduction ~t.',.) characteristics. For this purpose, the effective transmittanceabsorptance product(t a h, of TIM cover systems is experimentally measured at normal solar angle of incidence. The Ut is theoretically calculated from thermal network analysis and experimentally evaluated from mght-time cooling of water temperature in storage tank as follows: During the oil-sunshine ~eriod. the solar intensity term S'(t) is lave Ter'it;' :q;'o'i:f'n: ' _ : : i w: '

,T)]

(8)

From eq.(4, i ). night time Ut. may be expressed as: [

Ut

=

_M T

T(t)- T ( t ) lr~ [ T,,(t-1)- T(t)

(9)

Where M = M,, C,, + M, C, The theoretical and experimental values of overall heat loss coefficient and transmittance-absorptance product for TIM cover systems are summarised in Table-3.The multiple wall structured sheet provide good heat loss reduction but simultaneously cut the solar radiation whereas the TIM devices made of absorber perpendicular structure provide good heat loss reduction as well as high .solar transmittance resulting in relatively higher solar gain of the hot water systems.The results on system performance charactenstics as well as UL and (a :)dr values seem to favour absorber perpendicular configurations over absmlmr parallel configurations. The integrated-collector-storage units have often been used as solar water preheaters. Our expenments were perfmmed during the December (winter month at New Delhi). The indinmion of the tank was adjusted such that the absorber plane was peqmndicular to sun rays at noon. It was found that the hot water lemperatures of 50-60'C could bc attained with solar energy alone. It has also t~en found that the unit using encapsulated compound lameycomb ;L',TIM holds promise of use as 100% solar fraction water heater

Fig. 3

C i- Air layer : C2- Glass sheet: C3- Polycarbonate sheet C4-Structured shee (6mm); C5- Structured sheet (10 ram) C6-Ceilular array (5cm): CT-Encapsulated TIM (Scm) C8- Encapsulated TIM ( I 0cm) Ib(t)- Beam radiation Id(t)- Diffuse radiation Ta(t)- Ambient temperature

6.TIM INSULATED GROUND ICS WATER HEATER ~

6. i System Configuration And Approach The concrete-ground integrated-collector-storage system consists of a concrete slab. a pipe network and is placed in the ground. The PVC pipe-network is of 20ram outer diameter and l.Smm wall thickness "and embedded inside the concrete slab. The top surface of the slab is painted with black (c = 0.9) paint to absorb solar radiation and covered with Transparent Insulation Materials (TIM) device to reduce the top heat loses. The TIM cover system


consists of a compound honeycomb. The ground b assumed as semi-infinite media and has a low thermal conductivity to provide adequate insulation from the bottom. The side surfaces of the collector are perfectly insulated. Solar radiation passes through TIM cover and is absorbed at the absorber plane. The absod)cd energy is transmitted to the fluid(water) flowing Ihrough the pipe network. ,An exploded view of ground TIM -ICS solar water heating system is illustrated in Fig.4. The temperature distribution in the ground ICS system may be estimated by solving the Fourier heat conduction equation with appropnate initial and boundary conditions characteristics of system geometry.The ground ICS is assumed to have constant therrr~)physical properties. In this numerical method, the temperature ts calculated at certain discrete points of space and time. The space and time derivatives are converted into finite diflcrcnces and a set or linear simultaneous algebraic equations arc obtained which may bc solved by matrix algebra (Reddy c1 al.,

207

20cm; Ihe variation of heat gain efficiency as a function of slab thickness al.~) supports the infercnoc. It is advL~blc to consider a slab thickness of 10-25cm In this regard, wc have. therefore.

1998). 6.2 R.'.~uhs and Discu.vsion The above mathematical model may be used to evaluate the thermal performance of TIM insulated ground ICS solar w a l ~ Table-3 Thermal and optical characterktics of v i r i o ~ TIM cover systems for ICS solar water heater.

Tl~'oretical (W/re" "C)

Exp.

TIM

(Who" "C)

Covcr

UT Ci

C2 C3 C4 C5 C6 C7

C8

6.23 3.73 3.43 2.63 2. I I 1.93 1.73 !.33

UI,

U~

Uc

0..S9 0.59 0.59 059

0.38 0.38 0.38 0.38

7.20 4.68 4.40 3.60

059

0.38

3.08

0.59 0.59 0.59

0.38 0.38 0.38 ,,

2.90 2.70 2.30

,,,

(T=),~ at 8=O

UL

7.9:t.0.25 5.6:t.0.10

0.7:59 0.658

4.5r

0.696 0.565 0..S65

4.4:t.0. I I 3.8:t,0.12 3.4:t0.13 2.8r 2.6~-0. I$

0.625 0.588 0.430

l:ig. 4

heating system. The values of thermophysical parammers used in the analysis are as follows: (a) Cover system:Compound Honeycomb (b) Concrete:K, = 1.75 W/m K, C, = 880 J/kg"C and p, = 2242 kgJm' ( ~ Ground: K~ = 0.56 W/m K. C~ = 1840 J/kg"C and Pc = 20:50

considered concrete slab of 25cm thickness for further computer runs. The final water temperature corresponding to concrete slab thickness of 25cm and the pipe network at different depths are sho~) in Fig.@ It is observed that the rise in water temperature is higher at lower depths but the thermal storage effect is better at

k~m'

larger depths: middle of the slab is considered a good compromise.

(d) Common Brick:K,, = 0.72 W/m K. C, = 1884 J / k ~ and p, = 1922 kg/m ~ (d) Iron impregnated sand:K, = 2.41 WIm K, C. = 2860 J/kg"C and p,, = 2466 kg/m -~ (e) Mass flow rate of fluid:0.001 kg/s The solar radiation data and ambient temperature data used m simulation model are corresponding to the month of December 1996. The diurnal variations of final water tempcf'~urc for the surface slab of 4(km thickness and Ih pipe network placed at various depths for dillcrcnt collection-storage materials arc tlluslralcd in Fig.5 (a.b.c). The u.~ of highly condm.live storage material n)provcs the thermal performance. For exan~le, iron impregnated sand (ratio I: 19) as collection.storage malerial of Md,lr ICS system cxc'cls over ground, c o l l l n l o n brk.'k and c[mc'rcle. ll~ u.~ of such material incrca.~s the total system r The rt.~tn waler tcmwraturc ts negligible for the slab thickness ove[

208


Fig, 5(a) Concrete

Fig, 5(b) Iron im1~n~gn~i~d

Fig 5 (c} Total ground

Fig~ 6


NOMENCLATURE A,

C, C. l~(t) I,j( t )

M M, M.: Qi-ill Qt.~t I Rr, R,

Rr Sit~ S'<tp "|"

"l"

1"., Till T.,II i ['i

I!, ['I (:ix it, ( ": ix ),t~

(~

Area of the solar absorber surface of the tank (m:) Specific heat of tank material (i/kg "C) Swcific heal of water (J/kg "C) Solar beam radiation at time t (W/m:) Solar diffu.~ radiation at time t (W/m') Number of grid lines along width of concrete slab Mass of the tank (kg) Mass of the water in the tank (kg) Retrieved heat flux of heater at time t (W/m: I "l',ual heat loss from the system at time t (W/m:) Tilt factor (or beam r',Kliatnon Tilt lacl()r for diffu.~ radiation Tih factor fi)r reflected radiation Solar irradiatmn at time t (W/m:) Solar irr',gtiation reaching the absorber at time t (W/m:) l"nn~ duration between two successive observations (sect Average ambient air temperature ("C) Average water temperature in the tank CC) lnntnal water temperature in the tank ('~) Ambient air temperature at time t r W:iter temperature at time t CCI Bottom heat hiss coefficient W/m-"C ()~erall heal loss coefficient (W/m: K) Side heat loss coefficient (W/m-" K) Top heat loss coefficient (W/m-" K) Tran.~mittance-absorptance product for beam radiation "rransminancc-absorptance product for ground diffuse radiation "I'ransmlnancc-absorptance pr(~luct for sky dill'use radiation

REFERENCES

209

Kaushika, N.D. and Reddy, K.S. (I 998). Thermal design and field experiment of transparesnt honeycomb insulated integratedcollector-storage solar water heater. Applied Thermal Engineering. Vol.19, pp. 145-161. Kaushika, N.D. and Arulanantham, M.(1996). Transminanceabsorptancc product of Solar glazing with transparent insulation matenals. Solar Energy Mat. and Solur Cell.~. 44, 383. Kaushika. N.D. and Avanti. P. 11996). Temperature distribution in the ground ICS system with TIM. In Proceedings of National Solar Energy Convention, 96, 80. Kaushika, N.D. and Sharma, P.P. (1994). Transparent honeycomb insulated solar thermal systems for energy conservation. Heat Recovery Systems & CHP, 14. I. 37. Kaushika, N.D., Sharma, P.K. and Padmmapriya, R.(1992). Solar thermal analysis of honeycomb r(~f cover system for energy conservation in an air-conditioned building. Energy & Building. 18. 45-49.

Kaushika. N.D., Ray. R.A.. and Padmaprnya, R.~i990). A honeycomb ~ l a r collector and storage system. Energy Corn's. Magml. 30, 127. Kaushika, N.D. and Banerjce. M.B. ( 1983 ). Honeycomb solar pond; evaluation of applications, hi Proceedings of ISES Solar World Congress, Perth, Australia, 246. Kaushika, N.D., Banerjee, M.B. and Yojana Kant.(1983). Honeycomb solar pond collector storage system. Ener~tv. 8.883. Lin E.I.H. (1982). A saltlcss solar pond. in Prr~'eedings of

ISES(Amerwan section of ISES).225. Arulanantham. M. and Kaushika. N.D. (1996). Coupled radiative and condu, thermal transfers across transparent honeycomb insulation lli.,;Cllals. AI,plied Thermal l='ngg. 16, 3,209.

Onabasi, U.. Dyksterhuis, F.H. and Kaushika, N.D. (1983). Honeycomb stabilised sahlcss solar pond. Solar Energy. 31,229.

Avann. P., Arulananiham. M. and Kaushika, N.D. (1996). Solar lhcnnai analysis of ground integrated collector/storage system with transparent insulation. Applied Thermal Engg. 16, ! !. 863.

Prakash, J., Kaushika, S.C., Kumar, R. and Garg, H.P. (1994). Performance prediction for a triangular built-in-storage-type solar water heater with transparent insulation. Energy. 19, 8. 869.

Duffle. J A. :rod Beckman, W.A. (1991). Solar Engineering of Tiu,ruud Proc'es.,~c's. 2n ~ edn, pp. 54-59. Wiley inierss

Reddy, K.S., Avanti, P. and Kaushika, N.D. [1998). Finite-time thermal analysis of ground integrated-collector-storage solar water heater with transparent insulation cover. Inlernalional Journal of Energy Research. in Press.

New York Gucizh

0.40

--------~ I

.,.,..

"~

0.30

ja.~a---

"0

t~ II

0.20

0 Q.

0.10

||

0.00

t

0.00

1.00

2.00

3.00

4.00

5.00

6.00

7.00

8.00

Air flow rate (kg m "2 h~) Fig. 4. Productivity as a function of air flow rate when utilizing only waste thermal energy and placing an insulating cover on the still glazing. Ambient temperature = 23.7 - 25.9~ feedstock flow rate = 9.3 kgm'2h"1 and feedstock inlet temperature = 86 - 90~ Table 7. Temperature profiles at steady-state as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 - 34.4~ feedstock flow rate = 5.7 kgm2hl; feedstock inlet temperature = 86 - 90~ and irradiation = 650-&10 Wm 2. Air Flow Rate

UPPER CHAMBER T1

(kgm'2h-1) (%)

LOWER CHAMBER

Tto=

T~.I

Tx.in

T~o~

T~

T2

T3

T4

T5

T6

T7

T8

T9

T10

Tll

T12

T13

T14

T15

(~

r

(~

r

(~

r

(~

(~

r

r

r

(~

(~

r

1.33

72.2

95.(i

97.6

97.5

99.i)

97.4

97.1

48.5

94.1

84.7

87.6

86.1

84.9

51.5

73.4

2.03

62.5

88.5

91.7

91.6

94.3

92.1

91.2

42.4

87.1

78.2

80.5

79.3

78.3

49.3

67.4

3.52

56.9

80.8

85.5

85.4

89.5

86.7

84.8

36.3

80.1

73.3

74.2

72.9

71.6

48.9

60.6

4.02

56.6

79.4

83.8

83.8

90.9

86.3

83.4

34.9

78.5

72.5

73.3

72.3

71.1

51.2

60.7

223


Table 8. Productivity rates and calculated waste energy input, qwme, at steady-state as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 - 34.4~ feedstock flow rate = 5.7 kgm-2h-1; feedstock inlet temperature = 86 - 90~ and irradiation = 650!-_10 Wm -2.

II

Irradiation (Wm -2)

qwaste

I

(Wm:)

(kgm-2h -1)

II (kgm-2h-1)

E (kgm-Eh-1)

1.33

650

251

0.95

0.53

1.48

2.03

650

290

0.98

0.60

1.57

3.52

650

330

0.93

0.63

1.57

4.02

650

331

0.87

0.64

1.51

Air Flow Rate (kgm-2h-1)

1.8 1.6 -

iT_, t--

E

1.41.2-

1.0-

,L

0.8 "O

2 n

0.6-

,-II |l

l

0.40.2

0.0

-

-

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

Air Flow Rate (kgm2h ~)

Fig. 5. Productivity as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 34.4~ feedstock flow rate = 5.7 kgm2hl; feedstock inlet temperature = 86 - 90~ and irradiation = 650-2-_10Wm -2.

Inter-comparison of the modes of operation The results of this experimental study are summarized in Fig. 6 with respect to the total distillate rate as a function of the air flow rate for all the modes of operation investigated. It is not recommended to operate this prototype solar still at relatively high air flow, since this will result in an increase in the parasitic electric energy required to drive the air pump. The optimum air preheating the feedstock in the lower chamber (i.e., flow through the serpentine tube) even in the hybrid mode of operation. 4. CONCLUSIONS The performance of an air-blown multiple-effect solar still consisting of an upper evaporation chamber and a lower condensation chamber has been analyzed in three modes of operation, i.e., driving forces: (i) solar energy; (ii) waste thermal energy and (iii) hybrid, both solar and waste thermal energy.

flow rate for both the hybrid and the nocturnal distillation modes is observed to be in the range of 2 - 3 kgm'eh1. We believe that a productivity in excess of 20 kgmeday "1 may be achieved when operating the still in arid zones under these conditions, i.e., hybrid mode with nocturnal distillation. It may be possible to further enhance the productivity, based upon the results of this study, by In all modes of operation the performance of the still was determined as a function of the flow rate of the entering air stream and the optimum range of the air flow rates were determined experimentally. The optimum air flow rate for both the hybrid (during the daytime) and the nocturnal distillation modes was in the range of 2-3 k g m 2h'l. Based upon the experimental results of this study, a productivity in excess of 20 kgm2day "1 may be achieved when operating the still in arid zones under these modes, i.e., hybrid mode with nocturnal distillation. It may be possible to further enhance the productivity, based upon the results of this study, by preheating the


224

feedstock in the lower chamber (i.e., flow through the serpentine tube) even in the hybrid mode of operation. 1.8

-

1.6

-

Such a mode of operation will be studied in the future.

Sol~tr +wa ~te 1.4

4r

)

-

"7

"=

W a s t ( ; + in,.;ulate,]

~

J

0.8

91,=,l

Sohr

/

1.2

t'q

o

0.60.4

-

0.2

-

c~

.__ ~ - - J k - -

~

~

------

~ ' - - - -

"-'-

...-'~v

waste

~-j

_

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

Air Flow Rate (kgm2h "1) Fig. 6 Inter-comparison of the total productivity as a function of the air flow rate for all modes of operation investigated. The experimental conditions for each mode of operation are defined in the text. MENCLATURE c heat capacity ( J k g l K "1) m mass flow r a t e ( k g m 2 s "1) q thermal energy normalized to unit still area (Wrn2) qw~ waste energy input = q l ~ normalized to unit still area (Wm2) q~l=soa thermal energy released by the preheated feedstock in the upper c h a m b e r (Win"2) T temperature (~

Subscripts a brine f I II in out sep u w X

ambient brine drain-off feedstock primary distillate secondary distillate stream entering chamber stream exiting chamber vapor/liquid separator upper chamber water external heat exchanger/condenser

Acknowledgment- This research was supported under

Agriculture. Parker. B.F. (Ed). Vol. 4. pp. 255-

Grant No. TA-MOU-95-C15-050. US-Israel Cooperative Development Research Program. Office of Agriculture & Food Security. Center for Economic Growth. Bureau for Global Programs. Field Support and Research. USAID. One of the authors L. Horv t h acknowledges also the support of the Hungarian OTKA Fund, project No F-025 342.

294. Elsevier Science Publishers B.V. Amsterdam. Kudish, A.I., Evseev, E.G., Aboabboud, M.M., Horv~ith, L. and Mink, G. (1997) Heat transfer processes in an air-blown multiple-effect solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6,. Taejon, Korea, pp. 158-167. Malik, M.A.S., Tiwari, G.N., Kumar, A. and Sodha, M.S. (1982) Solar Distillation. Pergamon Press. Oxford, pp. 175.

REFERENCES Kudish. A.I. (1990)

Energy

in

Water Desalination.

Agriculture-Energy

in

In Solar World

ISES Solar World Congress 1999, Volume III Mink, G., Aboabboud. M.M., Horv~ith., L., Evseev, E.G. and Kudish, A.I. (1997) Design and performance of an air-blown solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 135-144. Mink, G., Horv~ith, L., Evseev, E.G. and Kudish, A.I. (1998) Design parameters, performance testing and analysis of a double-glazed, air-blown solar still with thermal energy recycle. Solar Energy 64, 265-277. Talbert, S.G., Eibling, J.A. and L f., G.O.G. (1970) Manual on Solar Distillation of Saline Water. R&D Progress Report No. 546. US Department of Interior.

225

226


PERFORMANCE AND ANALYSIS OF A MULTIPLE-EFFECT SOLAR STILL UTILIZING AN INTERNAL MULTI-TUBULAR HEAT EXCHANGER FOR THERMAL ENERGY RECYCLE G. MINK*, L. HORV/~TH*, E.G. EVSEEV ~ and A.I. KUDISH** *Research Laboratory of Materials and Environmental Chemistry, Chemical Research Center, Hungarian Academy of Sciences, 1525 Budapest Pf 17, HUNGARY; Tel: +36 1 325 5992, Fax: +36 1 335 7892, E-mail: [email protected] **Solar Energy Laboratory, Institutes for Applied Research, Ben-Gurion University of the Negev, Beer Sheva 84105, ISRAEL.; Tel: +972 7 6461488, Fax: +972 7 6472916, E-mail: [email protected] Abstract - To achieve a relatively high productivity at reduced investment costs, an innovative, air-blown, multi-tubular

solar still module was fabricated l~om readily available, corrosion resistant materials. Performance studies were made as a function of the air flow rate (in the range of 0.36 - 5.29 kgm2hq) utilizing a solar simulator providing a constant irradiation of 630-!-_10 Wrn2 and a constant feedstock flow rate of 2.5 kgm2h1. A simulation model has also been developed to describe the heat and mass transfer processes occurring in this prototype solar still and it was validated by the experimental data. It has been found, both experimentally and by the simulation model, that the total amount of thermal energy recycled (to preheat the feedstock and to directly heat the evaporating surface) is a maximum at an air flow rate of-- 1.28 kgm-2hl; corresponding to a maximum still productivity of 0.97 kgm2hq. 1. INTRODUCTION The consumption of water unfit for drinking is a major health hazard in rural areas. To supply these rural people with water free of salinity and/or pathogens is an urgent task to be solved. Solar distillation could be an ideal source of fresh water production, however, the crucial problem is that the productivity per unit area of the traditional solar stills is low. In addition, the fixed capital investment cost of a solar desalination plant is roughly proportional to the still area. Consequently, there are two possible approaches to overcome these limitations; to either increase solar still productivity per unit area and/or decrease the fixed capital investment per unit area. In our previous papers 13 the performance of an airblown, multiple-effect solar still that offers a significant increase of the still productivity per unit area at marginal incremental costs has been presented and analyzed, viz., the first approach. In this still, consisting of an upper evaporation chamber and lower condensation chamber, a large fraction of the heat of condensation of the distillate is successfully recycled both to the evaporation plate and to preheat the feedstock. The aim of the present work was not to enhance the productivity of the above still but to simplify its construction and thereby make it a more economically viable alternative, viz., the second approach, reducing the fixed capital investment cost per unit area. This was achieved by utilizing relatively inexpensive and corrosion resistant materials in the construction of the solar still. The new solar still design, though significantly different in appearance from the original, utilizes the same heat transfer processes to obtain the final distillate product. In spite of its simple construction, the heat and mass transfer processes occurring within the still are numerous and mutually interrelated. Another aim of

this work was to develop a mathematical model, utilizing non-linear differential equations, which is capable of simulating still performance under both transitional and steady-state conditions. 2. SOLAR STILL DESIGN The solar still module under investigation is shown schematically in Figs. 1.a and lb. It consists of a bottom and edge insulated, thin, rectangular plastic tray (L = 1.84 m; b = 0.54 m; H = 12 mm; area = 1 m2), 40 glass tubes (Dolt = 7 mm; D~, = 5 mm; L = 1,8 m) covered with a black wick and a plastic serpentine tube (Dout = 4 mm; Dr, = 3 mm; L =20 m). The still glazing is a solar grade polycarbonate double-walled sheet (10 mm thick), a transparent insulation material (TIM). A low pressure variable speed air pump was used for the air, the mass and thermal energy carrier, and a peristaltic pump for the feedstock. The still shown schematically in Fig. 1, operates in the following manner: (i.) ambient air at temperature To enters at the lower extremity of the still; (ii.) evaporation occurs from the wick, which is also wets the external surface of the glass tubes; (iii.) the air stream achieves both its highest temperature and vapor content at the upper extremity of the still; (iv.) the air stream is directed into the longitudinal glass tubes at the top of the still and then flows downwards, i.e., it reverses direction, and most of its vapor content condenses on the inner surface of the glass tubes and the thermal energy of condensation is conducted via the tube wall to the wet wick to enhance the rate of evaporation from the wick; (v.) the enthalpy of air stream, at temperature Tr, entering the lower compartment, which unifies the air streams exiting the 40 glass tubes, is utilized to preheat the feedstock, prior to its entering the evaporation chamber. The feedstock flows through a 5 m long black tube packed within this compartment to facilitate the heat exchange process.


(vi.) in the evaporation chamber the feedstock is further heated as it flows through another 15 m of the black serpentine tube, positioned above the glass tubes, before exiting the serpentine tube onto the black wick at the uppermost part of the evaporation chamber. (vii) the air stream and distillate exiting the solar still enter a gas-liquid separator (Sep. 1) to collect the primary

distillate (I), in order to determine the amount of distillate which condenses within the still and is a measure of the efficiency of the thermal energy recycle process; (viii) the air stream exiting Sep. 1, still containing vapors, then enters an external heat exchanger and a second separator (Sep. 2), where the secondary distillate (II) is collected.

Fig. 1.a Schematic diagram of the multi-tubular solar still and the experimental setup. temperature probes are indicated by numbers 0 to 10.

Fig. 1.b

227

The location of the

Simplified scheme of the top view of the multi-tubular solar still. The actual number of glass tubes is 40.


228

3. EXPERIMENTAL The solar still was tilted at an angle of 20 ~. To facilitate the parametric sensitivity studies (viz., the dependence on air flow rate) a solar simulator, also tilted at 20 ~ was utilized. It provided a constant solar radiation intensity of 630-2_10 Wm 2. A 50 mm thick polyurethane foam was used to insulate the edge and bottom of the solar still (not shown in Figs. 1.a and 1.b). The differential and cumulative yields from Separator 1 and 2 were measured automatically with a type PT 6 Satorius electric balance with an accuracy of +1 g. The temperatures were measured at 11 locations with an accuracy of +1 ~ using calibrated temperature sensors of the silicon base type KTY l l-2A. The absolute humidity of air entering the glass tubes was calculated from material and energy balances on the still. The feedstock was pumped into the still at ambient temperature, which varied between 23 and 27~ The feedstock flow rate was kept constant at 2.5 kgm2h "1 throughout the experiments and the still performance was investigated in an air flow range between 0.36 and 5.29 kgm-2h1.

A data acquisition system served both to monitor and store the temperature data from the eleven thermistors and to calculate differential and cumulative yields at variable time intervals. It consisted of a PC with an A/D-D/A converter card, electronic measuring and magnetic valve control unit, temperature sensors and a digital balance with a RS232C serial interface. The data acquisition, control and analysis software were developed for the study. 4. RESULTS AND DISCUSSION Transitional stage of the distillation Athe mass and energy flows were normalized to a 1 m 2 still area. Due to the low thermal mass of the still, the steady-state yields and temperatures were achieved within one hour of start-up in all experiments. A set of experimental result, typical of those obtained, are shown Figs. 2 and 3. The transitional and steady-state temperatures corresponding to air temperatures T1, T2 and T5 in the evaporation chamber (cf., Fig. 1.a) are shown in Fig. 2.

90 _AAAAAAAAAAAAAAAAAAAAAAAA 80 --

70-A 60 -0.

All I

50

;

iii IIIIIIIIIIIIIIIIIIIIIIIIIIII

&I I

iI

IT1 i T2

40 ~-i~I~l!

AT5

30 20 0

I

I

I

i

i

I

20

40

60

80

100

120

140

Time (min) Fig.2.

Variation of the air temperatures T1, T2 and T5 with the time of irradiation. Feedstock flow rate = 2.5 kgm2h -1, irradiation = 630s Wm "2.

The approach to steady-state distillation rates, primary (I), secondary (II) and total (Z), as a function of the time lapse since start-up is presented in Fig. 3. The primary distillation rate refers to that condensed within the glass tubes and in the feedstock preheating compartment of the still. The secondary distillation rate is that obtained by passing the saturated air stream exiting the solar still through an external heat exchanger prior to venting to the ambient.

It is observed in Fig. 3 that there is a time delay of 27 minutes before the first breakout of distillate from the still. This is caused by the fact that when starting with a dry still, distillate first appears in the separator only after all the inside surfaces have been wetted and the small but not negligible dead volumes of the system are filled with the distillate. Once steady state is achieved, approximately after one hour, a random variation in the measured yields is observed. This is due to the high surface tension water which flows into


separator 1 both continuously and periodically, in the form of rivulets. Therefore, steady state yields have been defined as an average of the data measured during

Fig. 3

229

at least one hour of operation under steady state conditions.

Still productivity as function of time since start-up. Feedstock flow rate = 2.5 kgm-2h1, irradiation = 630-2_10 Wrn-\ r~

Still performance under steady state conditions The temperature profile of the still at different air flow rates expressed in kg bone dry air per m 2 still area per hour is reported in Table 1, where T~mb is the ambient temperature; TVREnEATthe temperature of the

preheating compartment; TSEPa the temperature of the first separator; and Tx,m and Tx,o~t the air temperatures entering and exiting the external heat exchanger, respectively.

Table 1. Temperature profile of the multi-tubular solar still at steady-state as a function of air flow rate for a constant feedstock flow rate of 2.5 kgm2h 1 and irradiation = 630-2_10 Wm 2. The position of the thermistors are shown in Fig. 1.a. Air Flow

Tamb

CHAMBER

Rate (kgrnEh1)

To (~

T1 (~

T2 (~

T3 (~

T4 (~

T5 (~

(~

(~

(~

(~

0.34

23.5

80.3

97.1

99.1

97.8

98.3

80.1

57.7

51.9

25.8

73.6

65.9

63.6

27.3

TpREHEhT TSEP.I T6

T7

Tx,t~

Tx,out

T8

T9

1.28

24.7

76.2

91.1

93.6

90.2

92.8

2.53

25.0

65.5

75.7

82.3

84.0

84.6

65.6

57.8

56.3

26.8

5.29

27.1

57.2

62.3

70.9

73.4

75.2

60.1

53.3

52.8

28.8

The still productivity, as a function of air flow rates in the range from 0.34 to 5.29 kgm-2h-1 is reported both in Table 2 and in Fig. 4 . It is observed that for the primary distillate rate, which is a direct measure of the thermal energy recycle efficiency, there exists an optimum range of air flow rates in the vicinity of 1.3 kgm2h 1. It is also observed that the ratio of secondary to primary product increases with increasing flow rate, i.e., the primary decreases and the secondary increases with increasing air flow rate (cf., Fig. 4). The reason

for an optimum air flow rate has been discussed in our previous papers 2"4 and it will be also shown later in section 5 that the simulation model also predicts a maximum for the primary distillation at the same air flow rate. The heat and mass balances on the still were calculated from the experimental data assuming that the air stream, after its temperature in the glass tubes dropped to the dew point of the air that enters into the glass tubes, was always saturated while passing


230

through the glass tube to the ambient. Consequently, at T9. The dew point temperature of the air and the saturated air exits the glass tubes at temperature T6, mass flow of vapor as calculated from the mass and enters separator 1 at TT, the external heat exchanger at energy balances on the still are reported in Table 3. Ts, the separator 2 at T9 and vented to the ambient also Table 2. Still productivity as a function of air flow rate. Feedstock flow rate = 2.5 kgrn2h-1; irradiation = 630s Wm 2.

, (kgm2h Flow-1)te II(kgm-2h I -1)

II

Z

(kgm2h "1)

(kgm'2h'l) '

0.56

0.05

0.61

1.28

0.82

0.16

0.97

2.53

0.68

0.19

0.87

5.29

0.66

0.29

0.95

0.34

~

Fehler! Keine gfiltige Verkniipfung.

Fig. 4.

Still productivity as a function of air flow. Feedstock flow rate = 2.5 kgm2h "1, irradiation = 630-~_10 Wm 2.

Table 3. The mass flow of vapor at steady-state as a function of air flow rate. Feedstock flow rate = 2.5 kgm2hl; irradiation = 630s Wm "2. Air Flow Rate (k~l:l2h -1)

TlxmX

TDw

(Ts)

(~

(~

m v,m .B

mv,m~xD

mv~

mv~ r

(k~rm'2h"1) (kgm-2h-1) 0cgm-2h-1) (kgm-2h-1) (kgm-2h-1)

0.614 0.617 0.003 91.9 98.3 1.000 0.013 0.987 84.3 92.8 0.901 0.927 0.026 74.2 84.6 2.53 1.084 1.022 0.062 65.0 74.2 5.29 A _ the calculated dew point of air stream turning into the glass tubes at T.~; B _ the mass flow of vapor carried by the entering air stream at Ta; c _ evaporaion rate in the evaporaion chamber; D _ the mass flow of vapor at Tm~; F _ the mass flow of vapor entering the exchanger at Tx:,; F _ the mass flow of vapor exiting the exchanger at Tx,out and vented to the ambient. 0.34 1.28

The following comments are based upon the above analysis presented in Table 3" 1. The calculated dew point temperatures are below the observed fluid temperature Tm~x,, which suggests that the mass transport between the wet wick and the air stream is not intense enough. 2. The operational optimum is observed in the vicinity of an air flow rate of 1.28 kgm-2h"1 since (i) the total productivity is high (0.97 kgm2hl); (ii) the load on the heat exchanger is moderate (0.18 kgm2h "1 vapor at a relatively high fluid temperature (63.6 ~ (iii) the amount of vapor vented to the ambient is marginal (0.03 kgm2h q) and (iv) because the air flow rate is low, the parasitic electrical energy required to drive the air pump is marginal (< 1 W per m 2 still area), which in rural areas might be provided by PV panels.

.

0.057 0.180 0.247 0.414

0.007 0.030 0.057 0.134

Increasing air flow rate results in a higher evaporation rate but lower condenser efficiency and thereby, greater thermal energy and water vapor losses which are vented to the ambient. Additionally, at higher flow rates the parasitic electrical energy requirement increases.

5. SIMULATION MODEL OF THE MULTITUBULAR SOLAR STILL A simulation model of the multi-tubular, air-blown solar still was developed. The temperatures of the air stream and feedstock at each node in the thermal analysis of the solar still system were modeled by the appropriate energy and mass balance relationships. This resulted in a model consisting of a set of nonlinear energy transfer equations. The model was solved numerically, using an explicit predictor-corrector difference scheme assuming "steady-state conditions"


by using sufficiently small time intervals during which the feedstock flow rate and ambient temperature are assumed to be constant. The following assumptions have been made in this analysis: 1. The water film is replenished locally on a continuous basis, is stagnant and very thin; thus its thickness and temperature are assumed to be that of the wetted wick. 2. The system is considered to be uni-dimensional in the flow direction of the individual streams. 3. The Lewis number for the air-water mixtures in the operational temperature range is assumed to be equal to 1.0, yielding l~m-----Cs 9 4. In theevaporator, as well as in the internal heat exchanger in the lower compartment, a single-row serpentine geometry is assumed. Therefore, the moist

231

air stream is assumed to pass in a cross-flow pattern over the feedstock tube. 5. In order to simplify the model by eliminating the j-th surface temperatures Tj (j=g, t or w) and to preserve at the same time a reasonable accuracy for the longwave radiative exchange processes (Duffle and Beckman, 1980), the net flux containing the above temperature is expressed by Qrj = Ej(~.( T 4- Tj 4) = hrj "( T - Tj), where hrj = 413jO-[( Tj,in + 2T + Tj,out )/4] 3 and T is the corresponding air stream temperature, either above or below the surface node.

Mathematical model The total thermal energy transferred through any j-th surface temperature Tj (j=g, t or w) is given by (2) O0) + U0),i~'( Tin- Tj) = Aout/Ain-U0),out-( T]- Tout), where Q0) is the sensible thermal energy transfer due to either the condensation/evaporation process or the incident solar radiation heat flux.. The total temperature difference above and below any surface is given ATtot- Tin- Tout = ATe, + ATout; AT~n= Tin- T3, ATout = Tj- Tout.(3) The differences from Eq. (3) may be calculated formally in term of ATtot, by substituting for the corresponding surface temperature as ATin = (O(j),,in+Aout/Ai~-U0),out )-1 [Aout/Ain.O0),out ATtot- Q0)], (4) ATout = (U0),,i~+Aout/Ai~'U0),out)-1 [U0),i~ ATtot + Q0)](5) The glazing, tube wall and wetted absorber temperatures may be substituted from the corresponding mass and energy balances (Veza et al., 1993; Kudish et a1.,1997). Consequently, the energy transfer equations in the evaporator are obtained as follows: (pc/5)~OTa~/& + ma/bw~" O((Ca+ W2 c, )Ta,2)/Ox2) - Qev,2 - Qoo~t~ - Qcon,g + Ag/Aw,2-O2 (Uamb + 02 )'l[Uamb(Tamb - Ta.2) + Ocon.g + Gg] + UE(U1 + Din,1/Dout,l'UE)'l'[Din,1/Dout.l'U](Ta.1- Ta,2) + Qev,2+ Din,1/Dout3-Qcon.1 + Gw] + At/Aw,2" UE(U2 + Oin,t/Oout,t'Uf)'l'[oin,t/Oout,t'Uf(Tf,2 -Ta.2) -Qcon.t,2 - Gt], (6) (pc Din t2/4Dout,t)f.2OTf.E/& + me Cy//1;Dout, t'OTf,E/OXt= Qeon,t~ + Gt + lJr (Din,t~o:~tUf ~-U2)"1 [U2(T2-Tf) - Qoon,t~- Gt], (7) ma ()W2/t) X2 "- hm,2b2" (Ws,2 ((Ta,2+Tamb)/2) - W 2 ) . ( 8 ) Similarly, the energy transfer equations within the glass tube are given by the following, where X 1 -- -X 2 to account for the reverse in flow direction: (pc Di~ 12/4Dour 1)a,lOTa,1/& +ma,1/~Doult,l" O((ca+W1 Cs )Ta,1)/~Xl) = - Qcon,1 + [.Jl(Din,1/Dout,lOl + U2) [ W2(Ta,2 - Ta,1) + Oeva + Q0o~,1+Gw],(9) ma,1 OW1/~Xl- h~l- ~Do,~l" (Ws, l((Ta, l+Tamb)/2)- W1). (l 0) The energy transfer equations for the air stream, feedstock and humidity in the internal heat exchanger are analogous to Eqs. (6) - (8) but with a view to being concise they are not detailed in this manuscript. Equations (6)-(10) were solved by imposing the following initial and boundary conditions: Ta,l(X,0) = Ya,2(x,0) - Tamb, Yf,2(x,0)= Ywag0, W2(x=0) - Wamb, (11) Ta~lx=O= Zamb, Tf,2Jx=O= Twat,i~, Ta,llx=L= Ta,2lx--L, Wllx=L- W2]x=L-(12) The definitions for the heat and mass flux and corresponding transfer coefficients applying to Eqs. (6)- (10) are: Qev = ~, (Ta)M~v, Q~o.= ~(Ta)Mr (13) U2 = hr + 0.9-a-4-[(Ta,1 + 3"Ta.2)/4]3, (14)


232

Uamb= hamb+ 0.9"6"4"[(3"T arab+ Ta,2)/4]3, 1/U1 = 1/he,1 +Din, l-log(Dout,1/Din,1)/2Kw + 1/heon, l, (16) 1/Uy = 1/hf + Din,t'log(Dout,t/Din,t)/2~:w, 1/hamb= 1/hg + 1/ha + 2-~ig/rg, (18) hv,d = 5.7 + 3.8Vamb; (19)

(15) (17)

where hg (cf., Eq. (18)) is determined from the Nusselt number correlation for natural convection between two parallel planes as proposed by Buchberg et al. (1976): Nug = 1 + 1.446(1 - 1708/Ra*), for 1708 < Ra*< 5900, (20) and Ra* = 2gp 2(Ta~ - Tamb) ~a3cOs0"Pr/(3T~2 + Tamb)Ba2. (21) The heat transfer coefficient for forced convection in the laminar flow regime is determined from the following correlation (Heaton et al., 1964), Nu~j = 5.4 + 0.0019[RejPrDrq/L]l71/(1 + 0.00563[RejPrDnj/L]L17), j=1,2, (22) which is valid for Rej< 2.3"103. The values for l~j determined by Eq. (22) were corrected for the effect of simultaneous heat and mass transfer by applying the Ackerman correction (Treybal, 1980). The forced convective heat transfer coefficient hf for laminar flow is estimated from the Nusselt number (Kreith, 1976), Nuf = [3.65 + (0.0668RefPrDm, t/Lt)/(1 + 0.04(RefPrDm,t~) 2/3. (23) The overall heat transfer coefficient through the bottom of the still to the ambient is given by Oloss

= 1/(Sim/lq~s + ~)b/Kb). (24)

The condensation heat transfer coefficient at low mass fluxes inside smooth horizontal tubes is given by Chato (1962) with the correlated coefficients proposed by Singh et al.., (1996): 3 lko~,~ = 0.0925-[p~pw- Ps)g~* ~:w/D~B~Ta,1Ta~)]~/4, ~*= ~, +0.68 Cw(Ta,1- Ta~). (25a)

(25)

The heat transfer coefficient for condensation from the saturated air stream on the inner glazing surface is given by (Kreith, 1976) as h~

= 0.725[ps(Ps- Pa)g sin0~3/2l-aBs(T~- T arab)]v4 .

(26)

The heat transfer coefficient for condensation from the moist air stream onto the surface of the feedstock tubes is given by (Kreith, 1976) as

heon,t = 0.943[pw(Pw- ps)g~:ff/2L1B~(Ta,1-Tamb)] 1/4.

(27)

The mass transfer coefficient hm is determined utilizing assumption 3. The rates of evaporation and condensation are calculated using the following relationships: M ~ = hm,co,,[W~2 - Ws (0.5(T~2 + Tamb))], (28) Moo~,tj = hm,~[W~j - W, (0.5(T~j+ Tfj))], j = 2,3, (29)

M~= h~ov[Ws (0.5(T~2+T,,0)- W~2],

(30)

where the value of Ws at the given temperature Ta is calculated from Ws = 0.622- V/(V- Ps). (31) The saturation pressure of the water vapor P~(Ta,) is evaluated from an empirical formula which was derived by a least square analysis of the data taken from the steam tables (see Elsayed, 1983). Numerical solution We have used the "predictor-corrector" difference scheme 0Vlarchuk, 1975) along each x-direction for the numerical solution of Eqs. (6) - (10), in the following format: (Ti n+l/2- Tin)/0.5At + f (W") (Tin+'/2- Ti-1 n+l/2)/AX = F(Tin), (32) 0.5-(Tin+l _ Tin + Ti 1n+l - Ti.I n )/At + f (W n+l/2)(Tin+l/2 . Ti-1 n+l/2)/Ax = F(Ti n+l/2), (33) where i=l,...,I and I=L/N is the number of mesh points.


233

function of air flow rate are shown in Fig. 5 and the dependence of the primary distillate on the air flow rate is shown in Fig. 6. In both cases, there is good agreement between the measured and calculated values. It is important to note, that the simulated primary distillate rate exhibits a maximum value, corresponding a maximum in the thermal energy recycle efficiency, at the same location as that measured.

At first, we utilize Eq. (32), a linear but strongly stable scheme, to obtain intermediate values of Tn+1/2 and apply Eq. (3) to refine the calculation by correcting for the whole step. Convergence was achieved for the initial air and fluid temperatures with various gridsteps. The results shown in Figs. (5) and (6) were obtained with grid-steps Ax = 0.1m and At = 0.3s. The measured and calculated air temperature T5 and T6 as a

120 100

G"

~

T5meas

80c

e

600) [-.,

T5calc T6meas

40

e

T6calc

20 I

I

I

I

I

1

2

3

4

5

Air flow rate (kgm2h "1)

Fig. 5

Measured and simulated values for T5 and T6 as a function of air flow rate.

6. CONCLUSIONS An innovative, air-blown, multi-tubular solar still fabricated from readily available, corrosion resistant materials has been studied experimentally and a simulation model describing it was developed and validated by the experimental data. The performance testing was done using a constant feedstock flow rate of 2.5 kgrn-2h1 (sufficient to maintain the wick completely wetted). A maximum in the steady-state productivity ( 0.97 kgrn2h-1) was

observed as a function of the air flow rate at 1.28 k g m 213-1, both experimentally and predicted by the simulation model. These optimum operating conditions correspond to the highest thermal energy recycle efficiency of the system, a relatively low thermal load on the external heat exchanger and relatively low parasitic electrical energy requirements to drive the air pump.


234

1 I

0.9 0.8

~,= '~ ..~ ~9

0.7 0.6 0.5 0.4 o 0.3 0.2 0.1 0

--4-- Imeas "

0

I

I

I

I

I

1

2

3

4

5

Icalc

Air flow rate (kgm-2h1) Fig. 6

Measured and simulated values of the primary distillate rate as a function of air flow rate.

NOMENCLATURE

arnb

A b c D(Dn) G g H h hm L M m N P Q T U V W~

surface area (m2) width (m) heat capacity (Jkg'lK "1) diameter (hydraulic diameter) (m) solar radiation (Wm"2) gravitational constant (ms "E) height (m) heat transfer coefficient ~m-2K "1) mass transfer coefficient (kgs-lm-2K1) length (m) mass flow rate (kg m2s 1) air/water flow (kg sl) number of mesh points in the chamber pressure (Nm-2) thermal energy (Wm2) temperature (K) overall heat transfer coefficient (Wm2K1) linear velocity (ms 1) saturated humidity for air at T (kg v-(kg dry

b c

W

air humidity (kg v.(kg dry air)"1)

air) "1)

Greek thickness (m) surface tilt angle thermal conductivity (WmqK1) latent heat of vaporization (Jkg-1) dynamic viscosity (kgm-ls-1) density (kgm-3) Stefan-Boltzmann constant (Wm2K "4)

Subscripts a

air

con ev f g in ins out r s t w wd 1 2 3

ambient bottom convection condensation evaporation feedstock glazing inlet/inside insulation outlet/outside radiation saturated tube wick/water wind internal heat exchanger evaporator external heat exchanger

Acknowledgment- This research was supported under Grant No. TA-MOU-95-C 15-050, US-Israel Cooperative Development Research Program, Office of Agriculture & Food Security, Center for Economic Growth, Bureau for Global Programs, Field Support and Research, USAID. One of the authors Mr. L. Horvfith acknowledges also the support of the Hungarian OTKA Fund, Project No. F-025 342. REFERENCES Aboabboud, M.M. and Mink, G. (1993) Solar still of improved efficiency. Proceedings of lSES solar World Congress, Vol. 4, Budapest, Hungary, pp.319324. Kudish, A.I., Evseev, E.G., Aboabboud, M.M., Horvfith, L. and Mink, G.(1997) Heat transfer processes in an


air-blown, multiple-effect solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 158-167. Mink, G., Horv~ith, L., Evseev, E.G. and Kudish, A.I. (1998) Design parameters, performance testing and analysis of a double-glazed, air-blown solar still with thermal energy recycle. Solar Energy 64, 265277. Mink, G., Aboabboud, M.M., Horwith, L., Evseev, E.G. and Kudish, A.I.(1997) Design and performance of an air-blown solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 135-144. Buchberg, H., Catton, I. and Edwards, D.K. (1976) Natural convection in enclosed spaces: A review of application to solar energy collection. Trans. of ASME, Jr. Heat Transfer 98, 182-188. Duffle, J.A. and Beckman, W.A. (1980) Solar Energy of Thermal Processes, Wiley Interscience, New York, 762 pp. Veza, J.M.,and Ruiz.,V. (1993) Solar distillation in forced convection. Simulation and experience. Renewable Energy 3, 691-699.. Heaton, H.S., Reynolds, W.C. and Kays, W.M. (1964) Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int'l. J. Heat & Mass Transfer 7, 763-781. Kreith, F. (1976) Principles of Heat Transfer, 3rd edn, Harper & Row, Publishers, New York, 656 pp. Marchuk, G.I.(1975) Methods of Numerical Mathematics, Springer-Verlag, New York, pp. 316. Treybal, R.W. (1980) Mass Transfer Operations, 2nd edn, McGraw-Hill, New York, 717 pp.. Chato, J.C. (1962) Laminar condensation inside horizontal and inclined tubes. ASHRAE Journal 4, 52-60. Elsayed, M.M.(1983) Comparison of transient performance predictions of a solar-operated diffusiontype still with a roof-type still, Journal of Solar Energy Engineering 105, 23-28. Singh, A., Ohadi, M.M., and Dessiatoun, S.V., (1996) Empirical modeling of stratified-wavy flow condensation heat transfer in smooth horizontal tubes, ASHRAE Transactions 102, 596-603.

235


236

Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers Graham L. Morrison, Gary Rosengarten and Masud Behnia Department of Mechanical and Manufacturing Engineering, University of New South Wales Sydney, Australia 2052 [email protected] Abstract- This paper describes the characteristics of horizontal mantle heat exchangers for application in thermosyphon solar water heaters. A new correlation for heat transfer in horizontal mantle heat exchangers with bottom entry and exit ports was used to predict the overall heat transfer and stratification conditions in horizontal mantle tanks. The model of the mantle heat exchanger tank was combined with the thermosyphon solar collector loop model in TRNSYS to develop a model of thermosyphon solar water heaters with collector loop heat exchangers. Predictions of stratification conditions in a horizontal mantle tank are compared with transient charging tests in a laboratory test rig. Predictions of daily energy gain in solar preheaters and systems with in-tank auxiliary boosters are compared with extensive outdoor measurements and the model is found to give reliable results for daily and long term performance analysis.

1. INTRODUCTION The thermosyphon solar water heater is the principal product concept in most major solar water heater markets. Thermosyphon systems with open loop connection between the tank and the solar collector have been widely adopted for both low pressure and pressurised water supply systems in climates that do not experience freezing conditions. Without freeze protection these systems are limited to tropical climates or locations that never experience frost. Freeze protection in open loop thermosyphon systems can be provided by water dump valves, electric heating in the collector header and by tapered riser tubes that control the growth of ice so that a rigid and expanding ice plug is not developed. Although all these techniques have been used in commercial products, they are not suitable for widely traded products that may be installed in any climate zone. The only inherently freeze tolerant systems are drain-down systems or closedloop collector circuits with a heat exchanger between the collector and the tank. The drain down concept is difficult to implement in a thermosyphon system if a pressurised water tank is used, however, drain-down thermosyphon systems are widely used in China for systems that provide hot water only in the evening. There have been numerous studies of the performance of closed loop thermosyphon systems based on tube heat exchangers inside vertical and horizontal storage tanks (Mertol et al. 1981, Webster et al. 1987). Tube heat exchangers can provide adequate heat transfer between the collector loop and the tank, however internal tube heat exchangers for thermosyphon operation are difficult to construct. The heat exchanger configuration that has been widely adopted for horizontal tank thermosyphon systems is the mantle or annular concept shown in Fig 1. A mantle heat exchanger is easy to construct, provides large heat transfer area and with appropriate design can promote thermal stratification in the storage tank. The primary advantage of mantle heat exchangers for thermosyphon systems is that they provide little resistance to thermosyphon circulation and maintain the simplicity of the thermosyphon concept. Mantle heat exchangers are also used for vertical tank, pumped circulation systems (Furbo, 1993, Baur et al, 1993, Shah and

Furbo, 1997). If a mantle heat exchanger had been used in the study by Webster et al (1987) instead of the eight immersed copper tubes, the heat transfer area would have been two and a half times larger and the heat exchanger penalty would have been significantly reduced. Furbo (1993) compared low flow solar water heating systems with a range of heat exchanger configurations and found that mantle heat exchangers outperformed immersed coil and external shell and tube heat exchangers in vertical tank systems. Although many manufacturers of thermosyphon solar water heaters use horizontal mantle heat exchangers there is very little information on the performance characteristics of this type of heat exchanger. Manufacturers of horizontal tank systems usually take a conservative approach to sizing and use the largest possible mantle (full c~umference and full length of the storage tank). A wide range of mantle widths and positions of the hot inlet pipe have been used. Systems designed to operate as solar pre-heaters typically have the hot-inlet pipe mounted near the top of the annulus. Systems with in-tank electric boosting have the hot-inlet pipes mounted below the level of the electric heater or have both the inlet and outlet mounted in the bottom of the annulus, Fig 1.

Fig 1. Full circumference mantle heat exchanger on a horizontal tank, showing alternative collector return points.


Baur et al (1993) also studied vertical mantle heat exchangers for pumped circulation systems using an empirical heat transfer correlation for laminar flow in the mantle gap developed by Mercer et al (1967) and a finite difference solution for the energy exchange between the mantle fluid and the tank contents. Based on experimental assessment, Baur concluded that there was little difference in the performance of pumped circulation solar water heaters incorporating vertical tanks with mantle heat exchangers or external heat exchangers. Inlet connections to the upper level in the mantle can result in substantial heat loss due to reverse circulation at night unless the input pipe has insulation equivalent to the tank insulation. High level connections may also result in poor circulation under low radiation conditions due to opposing buoyancy in hotter upper levels of the annulus. If a single tank mantle system with an upper level electric booster is used, the input level to the mantle must also be below the level of the boost element to avoid heat dumping at night and low circulation during the day. Due to the combination of these effects in many systems a low level input point is typically used for horizontal tank mantle systems. For a low level connection to the annulus (Fig. 1) the flow in the annulus is a combination of forced circulation from the collector loop and internal natural convection within the annulus. This paper presents a model of horizontal mantle heat exchangers and the integration of this model into the TRNSYS dynamic simulation package.

237

tank conditions. The flow structure in a bottom entry mantle depends on the temperature of the inlet flow relative to the thermal stratification in the inner tank. Numerical simulation of the flow structure in the mantle is shown in Figs 2 and 3 for mixed and stratified inner tank conditions. In Figs 2 and 3 half of the curved mantle has been unwrapped so that the flow field can be represented on a 2D projection. For all operating conditions there is a complex flow structure near the impinging inlet which results in 10 to 15 % of the total heat transfer taking place near the inlet port. For a uniform inner tank and mantle inlet temperature higher than the top of the inner tank (Fig 2) the mantle flow covers the full circumference of the heat exchange surface. For stratified inner tank conditions and a mantle inlet temperature less than the top of the tank, the mantle flow only rises to its thermal equilibrium level, Fig 3. For stratified conditions and a mantle inlet temperature colder than the top of the tank a recirculation zone develops above the mantle inlet as shown in the bottom right comer of Fig 3, where the inlet is displaced slightly from the end wall.

2. MODELLING THERMOSYPHON SYSTEM PERFORMANCE Simulation models for mantle heat exchangers on vertical tanks have been developed by and Baur et al (1993), Furbo (1993) and Shah and Furbo (1997). The model proposed by Baur has been implemented in the TRNSYS simulation package, and a heat exchanger model based on turbulent heat transfer correlations has been implemented in the WATSUN simulation package. The modelling procedure is the same for top entry mantles on either vertical or horizontal tanks. However, for horizontal tanks with low level or bottom entry into the mantle there is a significantly different flow structure due to mixed forced and free convection processes inside the mantle. Horizontal mantle tanks usually have a much narrower annular gap than vertical systems, typically 5 mm for horizontal systems and 20 mm or more for vertical systems. The flow entry port normal to the heat exchange surface in a horizontal mantle heat exchanger also results in higher local heat flux levels near the entry, due to impingement effects. The flow structure in a horizontal mantle with bottom entry and exit points has been investigated experimentally and numerically by Nasr et al (1997, 1998) and Morrison et al (1998). Due to the large area of a full circumference mantle and low flow rates inherent in thermosyphon solar water heaters the flow in horizontal mantle heat exchangers is usually in the developing laminar flow regime. For flow rates corresponding to the high end of thermosyphon circulation, mixing induced by the impinging inlet flow has been observed by Rosengarten et al (1997). The flow structure in a horizontal mantle has been studied by Morrison et al (1998) for both mixed and stratified inner

Fig 2. Simulated flow structure in a horizontal mantle heat exchanger with bottom level entry and exit for mixed inner tank temperature 30~ and a mantle inlet temperature of 50~ (half of the horizontal mantle circumference unwrapped).

Fig 3. Simulated flow structure in a horizontal mantle heat exchanger with bottom level entry and exit for stratified inner tank conditions 20~ - 40~ and a mantle inlet temperature of 30~ (half of the horizontal mantle circumference unwrapped).


238

3. HEAT TRANSFER CORRELATIONS Heat transfer in vertical mantle heat exchangers in solar water heaters has been modelled by Baur et al (1993) using a correlation (eqn 1) for developing laminar flow between parallel flat plates developed by Mercer et al (1967).

0.!6 Rew Pr N u = 4 ~ 6 § .............................................................................................

(])

Nu =Nusselt numbeT based on .logmean temperature difference hw =

............... k

given in eqn (1) is used in a thermosyphon model the additional iteration loop to find the log mean temperature for the mantle slows the convergence process. Although the log mean temperature difference form of the Nusselt number has the attraction of a constant value for fully developed flow, it is difficult to apply in a thermosyphon loop simulation model. Mercer at al (1967) also suggested a more direct solution for heat transfer using a Nusselt number based on a heat transfer coefficient defined in terms of the temperature difference between the inlet fluid and the wall. This modified form of Nusselt number is referred to as a non-dimensional heat flux to differentiate it from the customary log-mean temperature difference form of the Nusselt number. The modified Nusselt number N~ based on inlet temperature difference is given by

h = average convective heat transfer coefficient

q w Twall k

Nu w D

and QfhHLATIAr

(2)

Tin I-I

where q = average heat flux over the heat transfer surface.

R % = Reynolds number based on mantle gap width ~w

Hw# Pr = Prandtlnumber = :PeP k

A

=

(

(T. - r.a )

ln (To-T~mu)

The form of correlation function that is applicable to this definition of nondimensional heat flux can be derived as follows. The variation of fluid temperature in a heat exchanger with a constant wall temperature is given by

T.o-T~a ATo ~ 1 HL~I , r_..-

= at.

=

(3) )

(r. -

T,,~u = temvemure of heat ~ansf= surface T., = fluid inlet temlmmture To ffi 9

outlet temlmratme from the mantle

= flow rate through the ~ e

A = heat transfer ~ w ffi ~

area

o f mantle gap

H

= width of mantle- Imrpendicularto flow d ~ o n k = thermal conductivity of fluid ] . / = viscosity of fluid

Cp= specific heat of fluid The results of the simulation model were compared with experimental data reported by Furbo and Berg (1992). For vertical entry into the top of the mantle Baur found that a correction factor of 1.8 had to be applied to Mercer's heat transfer correlation for flow between flat plates. The heat transfer coefficient in eqn (1) is based on the log mean temperature difference and hence requires knowledge of both the inlet and outlet temperatures for the flow through the mantle. Although this type of correlation has been successfully implemented for pumped circulation systems by Baur et al (1993) it presents numerical difficulties if used in models of thermosyphon solar water heaters. Numerical solution of thermosyphon collector-loop circulation through a mantle heat exchanger requires an iteration for the collectorloop flow linked to an iteration for the tank temperature stratification. If a mantle heat transfer correlation of the type

whcrc L = length of mantle if flow dim~on.

The ovcralt heat transfr is given by

Q-mp(ro - r.) =

(4)

Aro)

The modified n o n d ~ i o n a l heat franker coef~i,nt t h ~ becomes ~. Q I w

u~= HLAr.-~

Although the wall temperature in a mantle heat exchanger will not be constant, the form of the correlation given in eqn (5) could be expected to apply to other boundary conditions if the mean wall temperature is used. Conventional heat exchanger effectiveness functions cannot be applied to mantle heat exchangers because the free convection flow rate on the tank contents side of the heat exchanger wall is not usually known. Heat transfer in horizontal mantles with the entry and exit points at the bottom of the mantle 9(Fig 1) has been studied by Rosengarten et al (1998,1999b). Rosengarten measured heat transfer in a vertical-slot mantle heat exchanger with bottom entry and exit ports, for a range of flow rates, mantle widths and stratification conditions on the heat exchanger surface (set by stratification in the inner tank). Rosengarten also used


239

The variation of the modified Nusselt number N u w with Reynolds number Rew in a typical horizontal mantle is compared in Fig 5 with the Nusselt number for developing flow between parallel fiat plates with one plate heated. This comparison indicates that the heat transfer in a horizontal mantle with mixed convection conditions is significantly higher than for heat transfer from flow between parallel flat plates. It is interesting to note that the increase of heat transfer in a horizontal mantle is similar to the increase of 80% observed by Baur et al (1993) for vertical mantles with top entry. Although the geometry of these two cases is very different the flow in the vertical mantle is also strongly influenced by buoyancy effects, Shah et al (1999), hence a similar heat transfer correlation may apply for both orientations of the mantle.

manUe z

2.5

Fig 4. Comparison of measured heat transfer in a horizontal mantle heat exchanger and the modified Nusselt number correlation, eqn (6). a detailed CFD model of the mantle and the inner tank to determine the distribution of heat flux in a horizontal mantle. The experimental and numerical data for overall heat transfer was correlated using eqn(5) with a value of h selected to give the best fit to a wide range of measured and simulated data. For typical conditions in thermosyphon solar water heater mantle heat exchangers ( 1 [] Rew D 1 O0 , w / L [2 0.02 and w / H [2 0.07 ) Rosengarten found that a value of h =392 W/m2K gave the best fit to the measured data, as shown in Fig 4.

Rosengarten et al (1999b) also developed a stratification correlation factor (St) to account for non-uniform wall temperatures due to stratification in the inner tank. The form of the correlation for bottom entry and exit mantles is

(6) whr162162

St = 0.93- 0.050+ 0.1202 o = (T= -

-

T~ao,, , = temperature aI the bottom of the mantle wall

Tm,~. = mean wall t e m ~ Tjn= fluid inlet t e m ~

to the mantle

z0

. . . . . . . .

1.5 z c

"1.0

0.5

0.0

T

0

2O

40

60

80

Re,adds N~ Re.

Fig 5. Comparison of heat transfer in a horizontal mantle heat exchanger and in a parallel channel heat exchanger with heat transfer through one isothermal boundary.

4. MODELLING OF STRATIFICATION IN A SOLAR PREHEATER The local heat transfer rate around the circumference of a horizontal mantle heat exchanger depends on the flow structure in the mantle and the variation of heat-transfer area with height. Due to the circular cross section of a horizontal tank a large proportion of the heat transfer area is in the top and bottom sections of the tank. If a horizontal tank divided horizontally into 20 equal mass segments then 20% of the heat transfer area will be in each of the top and bottom segments, for a 40 element tank approximately 15% of the area is in each of these elements. The development of stratification in a horizontal tank with a mantle heat exchanger was measured in a scaled 57 litre tank instrumented with thermocouple grids in the inner tank (Fig 6).

100

240


and the effect of jet impingement opposite the bottom inlet results in a relatively uniform temperature zone in the bottom of the tank. 4.1 Simulation model

Fig 6. Laboratory test rig used to measure mantle heat exchanger characteristics.

Fig 7 Measured and simulated heating of a horizontal tank with a mantle heat exchanger. Mantle inlet temperature 41.7~

The collector loop hot water flow into the mantle was supplied from a temperature-regulated source. A three-hour heat-up test was monitored for steady mantle flow conditions. The tank was started with a slight temperature stratification (25.3~ to 27.5~ and the mantle flow was set at 0.582 litres/min and a temperature of 41.7~ The tank temperature stratification was monitored over a 3 hour period as shown in Fig 7. As a result of the relatively large wall area in the top of the tank a hot stratified layer develops in the top of the tank as soon as the mantle heating starts. The mantle effectively acts as a stratification promoter and directs a significant part of the heat input to the top of the tank. High heat transfer into the bottom of the tank, due to the large bottom contact area

A thermal model of a ho"rtzontal mantle tank was developed using the correlation proposed by Rosengarten et al (1999b) for heat transfer in the mantle passageway. Convection heat transfer inside the tank was modelled as free convection along a vertical plate. Flow along the inner surface of the tank wall was assumed to be laminar and to stratify without mixing the contents of the inner tank. The heat exchanger model was developed in the TRNSYS package (Kline et al (1996)) as part of the TYPE38 stratified tank routine. The flow in the mantle gap was divided into horizontal segments in line with the segments used to model the inner tank. The TYPE38 tank model in TRNSYS uses a plug flow concept to follow the fluid into and out of the tank, however when a collector loop heat exchanger is included the only flow in the tank is the load flow. The TYPE38 model for the tank was retained for consistency with the well-established open-loop thermosyphon model in TRNSYS. Energy transfer between the mantle and the inner tank was modelled by dividing the tank into equal-mass elements of 2% of the tank contents. The predictions of this model for the steady heat-up tests are shown with the test data in Fig 7. The model required solution time steps of less than 0.2 hours before the predictions were independent of time step. This is less than the 0.5 hour time step commonly used for TRNSYS modelling of solar water heaters. The agreement between the model predictions and the measured data was generally good, although the temperature gradients above the bottom mixed layer differed from the observed data. This error is due to the simplified stratification assumption that is used to model the natural convection process over the inner surface of the mantle wall. The model correctly predicted the initial rapid temperature rise in the top of the tank due to the direction of the hot mantle inlet stream to the top of the mantle during the initial hour of heating. Once the temperature in the top of the tank approached the mantle inlet temperature (after one hour of heating), the flow in the mantle gap could not reach the top of the heat exchanger surface and more heat was directed to the bottom of the tank.

5. THERMOSYHON CIRCULATION IN A SOLAR COLLECTOR LOOP WITH A MANTLE HEAT EXCHANGER A common form of solar water heater is the horizontal-tank close-coupled thermosyphon system shown in Fig 8. The mantle heat exchanger is formed as part of the tank and then covered with insulation. The flow connection to the mantle in Fig 8 is a bottom entry configuration (see Fig 1).


241

Fig 8. Close-coupled thermosyphon solar water heater with a mantle heat exchanger in the collector loop. The advantage of bottom connections to the mantle is that heat loss due to reverse flow at night are suppressed compared to systems with top level input to the mantle. A model of thermosyphon solar water heaters with collector loop heat exchangers was developed by Morrison (1994) from the detailed open-loop model in TRNSYS. Bickford and Hittle (1995) compared this model of a solar water heater with measured data and showed that the model over predicted collector energy gain by up to 10%. The model also over estimated the degree of stratification in the storage tank, when it was operated as a preheater. This model has now been extended to include the heat transfer correlation developed by Rosengarten et al (1999b) and improved analysis of conduction down the shallow depth of the tank contents and through the tank walls. In the original TRNSYS horizontal tank model the equivalent conductivity of the tank walls and tank contents was quantified by a single value for the effective conductivity. This is correct for a vertical tank however, for a horizontal tank the effect of wall conduction increases significantly for the top and bottom sections of the tank. Heat conduction in the walls of a tall vertical tank has only a minor effect on thermal stratification however, for a horizontal tank the wall heat conduction has a substantial impact on thermal stratification in the top and section of the tank. The variation of effective thermal conductivity with depth for a 3 mm wall thickness steel tank is shown in Fig 9. As a result of the high wall effect in a horizontal tank stratified conditions cannot be maintained for extended periods in the top 10 to 20% of a horizontal tank. A new model of a thermosyphon solar water heater with a collector-loop heat exchanger was developed in TRNSYS, using the TRNSYS solar collector model (TYPE1), the TRNSYS stratified tank model (TYPE38) and a new heat exchanger routine integrated within the TRNSYS thermosyphon loop model (TYPE45). The heat exchanger model and the TYPE45 thermosyphon collector loop model both allow for the temperature dependent properties of propylene glycol that would usually be used in the collector loop for freeze protection.

Fig 9. Effective thermal conductivity as a function of depth in a horizontal tank. 6. COMPARISON WITH OUTDOOR MEASURED PERFORMANCE The performance of two mantle-tank thermosyphon solar water heaters installed on an outdoor test rig, were monitored over an extended period and compared with the new TRNSYS model. The water heaters were installed on a roof with a 34 ~ slope (latitude angle) and were operated either as a solar preheater or as an integrated system with in-tank electric auxiliary boosting. During the preheater tests the systems were filled with cold water at the start of each day and allowed to operate throughout the day without any draw off. In the early evening the tanks were discharged and the net useful energy collected over the day determined by integrating the heat discharge rate:

, cp30

F sss SS

.~

re -

sssJ Sr

IsSSSJJ~SS

s ~ sss JSS

0.4) will lead to higher annual performance predictions. 9 Simulations show that for r < 0.4, the errors were satisfactory. 9 DST give systematic errors for SDHW systems with strong IAM (r > 0.4). However, when the irradiance is corrected for the IAM during data processing and performance prediction, these systematic errors disappear completely. Therefore the scope of the DST procedure can be broadened to include systems with 'strong' IAM, provided that their IAM is determined and used for correction of the irradiance during both the parameter identification and performance prediction.

Fig. 3: Correction for incident angle dependency in DST


248

Overheating during testing Sensitivity to extrapolations *

9 9

Extrapolations of the (design load) heat demand, ranging from 50% up to 200%, sometimes lead to increasing errors. These errors stay always precise to within • at the 95% confidence limits. Extrapolation to other climates than the test climate leads to no effect on the prediction error. Extrapolations of DST predictions to performance predictions of identical, but differently sized (collector, store) systems seem to be precise.

Conductivity effects Several simulations have been performed on stores with different conductivity behaviour from the atvdliary part downward to the solar heated part. DST does not explicitly take this effect into account, but adapts the value of the parameter f ~ to simulate a larger part of the store being heated by the atvdliary heating.

3.4 Recommendations for improvement of DST The following changes are recommended in order to improve the accuracy of the DST procedure.

Negative system outputs In the DST DIS [2], it is said that during performance prediction any negative system output (cooling instead of heating of the water) is to be ignored. However for systems where the store is located outdoors, this is not physical: cold winter nights will allow the store temperatme to go down (= negative system output). According to TNO calculations for ICS systems, ignoring these negative system outputs leads to larger prediction errors, than when the negative system outputs are included. Therefore it is recommended that these negative outputs are no_.~tignored and that the standard text is changed to reflect this.

Wind dependency of heat losses The heat losses of a collector depend on the air velocity around the collector. Not taking into account this wind dependency (although the DST model describes a special, optional parameter uv quantifying this dependency), this sometimes lead to underpredictions in cases for strongly wind dependent collectors. With respect to the treatment of the air velocity (wind) surrounding the collector, it is recommended to use the option Wf,,~ (of the DST software) for SDHW systems with glazed collectors. Investigations on the wind speed during testing showed that for collectors with spectral selective absorbers a difference between mean wind velocity during testing and yearly mean wind velocity for the location of the performance prediction of +_2 m/s can be accepted. For collectors with black absorbers either a maximum difference of +1 m/s either the option W~ should be required. Note: This recommendation implies that all SDHW systems should be tested with a forced air velocity (fans) around the collector.

When the overheating protection mechanisms of the SDHW system are activated during the test, this destroys the precision of the test. Therefore this must be avoided at all times. Small improvements in the testto achieve this are suggested.

Conductivity effects; H/D Ratio of Store It is recommended to limit DST testing of SDHW-systems with auxiliary heating to a certain minimum ratio Store Height / Store Diameter (H/D) I. Note 1: It is the (solar) system configuration which requires this; this is not a limitation of the DST method. Note 2: This recommendation implies that thermosyphon and ICS systems with a horizontal store should be tested as 'solar-onlysystems'.

Auxiliary set-point temperature during testing For systems with integrated auxiliary, the present procedure requires the auxiliary set-point temperature to be set at 45 ~ for Ssol,b days and at 80 ~ for Ssto~ and Saux sequences. It is probably better to require the same set-point temperature in all sequences (close to the temperature used in real operation). Calculations suggest that switching on the auxiliary is only required in the Saux sequence, and is not necessary for the Ssotb sequence.

Changing the 12 MJ/day requirement for 'valid' test days For systems with high heat losses and test periods with irradiation just over 12 M J / m 2 for valid Sso~b days, a (certain) threshold temperature is sometimes not reached. This may lead to insufficient variability in the test data and thus to precise results. These high uncertainties can be prevented by testing according to the following extra requirement: "If the temperature of the water withdrawn in a S, ol.b sequence is always below a threshold temperature (to be specified later), the sequence shall be extended until two consecutive days with higher irradiation (for instance 15 MJ/m e) are included'. This will guarantee sufficiently that the system reaches higher temperatures and the heat loss parameters are fitted with sufficient precision. Note: This will occur only occasionally, so normal testing duration is not affected. Very cold test periods Tests performed during very cold circumstances (average outdoor temperature below or around O~ during testing) may result in higher prediction errors, especially for systems with high heat losses (systems with non-selective absorbers, uncovered collectors). R is recommended that for such systems, a lower limit is required on the outdoor temperature during testing.

1TO be specified later in this project


Sensitivity with respect to sensor errors It has been extensively investigated how sensor errors would affect the DST results. 9 The DST method is insensitive to random errors in the sensor readings. 9 DST is sensitive to systematic errors in sensor readings 2. It appears that the systematic errors allowed in the DST/DIS 9459-5 [2] still can lead to deviations up to +5-6 % (absolute) at 95% confidence limits. Therefore it is recommended that the allowed sensor deviations in the DST/DIS should be reviewed to see if they could be formulated more strictly. Reducing DST parameters The DST model seems to be overdetermined. Testing (some) SDHW systems lead to parameter values which are either negligible, e.g. Sc=0, DL=0, or highly cross-correlated, e.g. Uc* with Us. It might be necessary to reduce the amount of parameters.

3.5 Conclusions Work Package 1 Work Package 1 has been very succesfull, because it has resulted into improvements and a clear demarcation of the test metho& The accmacy has been found mostly within :L5% and always within +10% (if the improvements have been implemented). 4.

WORK PACKAGE 2: COMPARISON W I T H CSTG M E T H O D

The objectives of WP2 are: Comparison of the DST method with the CSTG method, which is also used in the CEN, leading to correspondence factors which enable comparison of DST results with CSTG results.

4.1

Overview of work done

The four participants 1NETI (Portugal), CSTB (France), NCSR (Greece) and UWCC (UK) have been carried out both DST tests as well as tests using the CSTG method on SDHW systems [6]. 9 INETI tested two thermosyphon systems, each with nonselective fiat plate collector. During the tests fans have been used in order to take care of wind influences. 9 CSTB tested one ICS system with double glazing and tubular tank (surrounded by minors). 9 NCSR tested two forced circulation systems, one with a fiat plat collector, one with an evacuated tube collector. 9 UWCC tested one ICS system; only a CSTG method, both indoor and outdoor. TNO has carried out simulations according to the 'simulated test data approach'. The data analysis has been carried out by TNO and INETI.

2 This recommendation applies not only to the DST method but also to similar methods that use short-time measurements to predict yearly energy gain: ISO 9459 part 2 and ISO 9459 part 3.

249

4.2 CSTG test method The SDHW systems, described above, have been tested according to the CSTG method, described in [3]. Also the system characterisation has been done using the CSTG method in order to determine the three parameters and their associated uncertainties of the model:

Q = a 0 + a 1 +a2(tamb(av ) +tmains) Where: Q [MJ] H [MJ] t [~

(2)

Daily energy gain from SDHW system Daily solar irradiation on collector Temperature (ambient / mains)

In ISO 9459-2 (1995) a calculation method for estimating the system long term performance has been formulated. The calculation procedure includes two load patterns: 9 Load pattern 1: Load determined by the volume of daily hot water consumption 9 Load pattern 2: Load determined by a minimum useful temperature limit for the hot water consumption; when the outlet store temperature is lower then this minimum value no water is extracted from the store. Using these load patterns it is not possible to compare the CSTG outcome with the DST outcome. In order to be able to compare the CSTG result with the DST result, a load pattern 3 must be defined: 9 Load pattern 3: Load determined by maximum energy needed (see also [7]). Load pattern 3 is in agreement with the Reference conditions of prEN 12976-2 (1997) [1]. One will use load pattern 3 comparing DST with CSTG

4.3 DST test method Because the CSTG test method cannot deal with auxiliary energy, the DST test must be carried out with (possible) auxiliary heater switched off. The systems were tested according to DST test method, i.e., ISO/DIS 9459-5. The test sequences performed for comparison with CSTG test method were: 9

Ssol, a

9

S~o1,bauxiliary off

9

Sstor c

The parameter identification as well as the LTPP has been carried out using the Insitu Software Package - version 2.7a.

4.4 Comparison of test results Based on the values of LTPP done according to DST and CSTG test methods, a comparison was possible. The percentage difference between the results of the two test methods is calculated according to:

Difference = Qosr - Qcs~ x 1oo (%)

Q~

(3)

In the case of thermosyphon and ICS systems, the DST values are almost always higher than the CSTG values. The differences are not higher than 15%. Values between 5% and 13% are observed for one of the (2) thermosyphon systems and ICS system. For the other thermosyphon system differences are lower than 5%.


250

Figure 4 gives a graphical representation of the predicted annual thermal performances (using CSTG and DST) of those systems mentioned.

Parameter adoption of CSTG method For the determination of the model's parameters al, a2 and a3, the SM&T group recommends to use Weighted Least Squares Regression in order to take into account the uncertainty of the measurements of Q, t~y) and tm,~.

The determined uncertainty of parameter ~ gives guidance for the adoption of a simplified model: Q = alH + a2(t~,d~y tmains) i.e.

If a3 < t (V,95%) ~a3 ~

Fig. 4: Yearly thermal performance of thermosyphon and ICS systems predicted according to the CSTG and DST method. The results obtained for forced circulation systems will not be reported here; more details can be found in w Simulations.

4.5 Simulations Simulations have been carried out on an ICS and a forced circulation system, using the 'simulated test data approach'. These simulations have led to the following findings: 9 The discrepancies (differences between CSTG and DST) are small (within 5%). 9 Also CSTG does not take incident angle dependency into account. In order to have a proper CSTG-LTPP one has to correct for strong incident angle dependency (r>0.4 in equation (1)). 9 DST predicts a slightly higher thermal performance of a SDHW system than CSTG.

4.6 Recommendations to CSTG (1SO 9459-2) The following adaptations concerning CSTG method are recommended: Procedure of LTP prediction of CSTG method The group recommends a change in the LTP calculation procedure on point 9.4 of ISO 9459-2 standard. The energy delivered by the system must not exceed the energy demand given by:

Energy demand = VioadPw Cr~ (tload- tmai~ instead o.o.o~thepresent defined draw ~ in ISO 9459-2 Validation of the test method done earlier already considered this type of load. Simplification of the CSTG method Tests show that there is no need to perform test for determination of the 'mixing draw-off profile' The normalised draw-off profile is recommended to be used in all corresponding equations in LTP calculation.

the simplified model can be adopted. t (v,95%) is the 'student distribution' for v degrees of freedom and 95% confidence level. The uncertainty of parameter a9 (determined from either the complete model or simplified model) gives guidance to the need of more test days for a good determination of this parameter. i.e. 9 If a2 < t (v,95%) 6~2 ~ more test days are needed Resulting in a2 > t (v,95%) 6a ;this will correspond to a larger value range for (t~day)- tmm~). 5.

WORK PACKAGE 3: EXPERIMENTAL VALIDATION

The WP3 objective is: Experimental validation of the DST method by means of intercomparison tests in a number of ten recognised laboratories throughout Europe.

5.1

Overview of work done

The participants of WP3 have carried out DST tests in order to fia~er validate the DST test method. A second goal op the tests in WP3 is a field check on the results coming from WP1. The tests have been carried out on the below mentioned systems. 9 1NETI (Portugal) have been carried out DST tests on a thermosyphon system (with non selective flat plate collectors) 9 NCSR (Greece) tested a thermosyphon system and a forced circulation system; both flat plate collectors; both stores have a electrical auxiliary heater installed 9 SPF (Switzerland) carried out three DST tests. Two tests on one same ICS in order to detect (possible) seasonal influences. One other DST test have been carried out on another ICS (tested by CSTB (France) in WP2). These two ICS systems were preheat systems 9 DTI (Denmark) has been carried out a DST test also on an ICS; the same ICS SPF has tested earlier. 9 FGH (Germany) has been tested one thermosyphon (preheat) system. 9 SP (Sweden) has tested a forced circulation system (which has been tested by NCSR before) and a thermosyphon system. The thermosyphon system has been tested by FGH before. 9 Infa Solar (Germany) has tested a forced circulation system. This is a low-flow system with a external load side


heatexchanger. An auxiliary heater could be installed directly to the storage tank. ITW (Germany) has carried out four DST tests. Two tests on one same thermosyphon system (again) in order to detect possible seasonal influences. One test on a thermosyphon system tested before by 1NETI and one test on a forced circulation system earlier being tested by TNO (The Netherlands). TNO has also carried out four DST tests. One test on the thermosyphon system earlier being tested by INETI (for WP2). One test on the forced circulation low-flow system (Infa Solar) One test on the forced circulation system with evacuated tubes earlier tested in WP2 by NCSR (see figure 5). one test on a forced circulation drain-back system; this system has been shipped to ITW afterwards Another two systems, a thermosyphon system and a forced circulation system have been validated back in 1996 by TNO and ITW. These result will also be used for this Work Package.

9

9

9 9 9

5.3

251

Because of correlation between the Uc*-parameter and the Us parameter, the collector loss parameter (Uc*) can (but does not need to) be omitted. Thermosyphon systems being tested as solar only systems, show a good mutual agreement tested at the two different laboratories. One can detect a significant Sc parameter only when thermosyphon systems are concerned. LTPP predicted for cold climates might be questionable because of possible freezing problems. Thermosyphon systems (with electrical auxiliary) as well as store in which the electrical auxiliary take care of (extreme) mixing of the store will decrease the solar gain enormously. DST can detect this effect very well. Recommendations

Mounting influence The mounting of a solar system is not stated clearly (enough) in CEN (or ISO). This can affect the test result and therefore the long term performance prediction.

Load Side Heat Exchanger It is recommended to exclude Solar Domestic Hot Water Systems with an external load side heat exchanger from the scope of ISO 9459-5 until sufficient experience is available for these system types.

Auxiliary power for performance prediction It is recommended to amend the specifications for the auxilia~-heater-thermal-performance to be used for Long Term Performance Prediction in the reference conditions of prEN 12976 and 12977 (table B 1): A d d ~ change at r e m a r k s : . . , i f not specified otherwise by the manufacturer. Change: '100 + 30 Wattper litre. . . ' into "150 • 50 Wattper litter... " Fig. 5 : W P 3 test on an f o r c e d circulation system with an evacuated tubes collector

5.2

Preliminary WP3 Results

Although not all participants have finalised their reports testing at the time of writing this paper, some preliminary results can be formulated. 9 The DST test procedure give good results testing preheat systems (ICS, thermosyphon as well as forced circulation systems). 9 DST is good in predicting a SDHW system with internal auxiliary heatexchanger. 9 As stated in WP1, DST has difficulties to predict the performance of an ICS with a high heat loss coefficient of the store. 9 The method to correct for incident angle dependency (e.g. for ICS systems) works out well. Assumed that the incident angle dependency is well know, this is however at this stage a potential for errors; it is recommended to integrate this into the DST software.

Energetic Performance Representation The (DST-) data processing software uses Watt as the unit for energetic performance representation per given time interval (e.g. one year). The respective prEN specifies that MJ/year should be used (to report) the energetic performance. It is suggested that in the future, either one of these units will be used.

Component Testing Experimental investigations within Work Package 3 of this SM&T project showed a very promising agreement between the component test method (CTSS), prEN 12977 and the DST method. 6.

CONCLUSIONS

This project 'Bridging the Gap' has been divided into three work packages. Three general conclusions are:

252


WPI: Definition of the scope The accuracy and the reproducibility of the DST method is considered as good. A statistical analysis of the results has led to an accuracy of • an in critical cases up to • of the solar fiacfon for different testing and prediction climates, hot water demands and for the system types common on the Etwo~an market. Those cases where this accmacy was not reached, have been isolated and analysed which has led to several conclusions, leading to suggestions for improvements in the DST procedure. WP2: Comparison with the CSTG method Based on the values of Long Term Performance prediction done according to the DST and CSTG test methods, the comparison between the two methods showed, for thermosyphon and ICS systems, differences on an average 78%; those differences lead up to 15% for more critical cases. Simulations lead to promising results, showing that CSTG and DST appear to be comparable. WP3: Experimental validation The comparison of results within the validation programme confirms the accuracy figures obtained by theoretical investigations in Work Package 1. This accuracy is on the average 5% and leads up to 10% for problematic cases. However, a series of tests have also led to significant and unacceptable discrepancies between test results. The analysis of these cases is not finished so far. Therefore the SM&T-group can only give an preliminary assessment of the DST method.

REFERENCES

[1] [2]

[3]

[4]

prEN 12976-2 (1997), Thermal solar systems and components. Factory made systems - Test Methods. ISO 9459/DIS-5 (1996), 'Solar heating - Domestic water heating systems - Part 5: System performance characterization by means of whole-system tests and computer simulation'. ISO 9459-2 (1995), 'Solar heating- Domestic water heating systems - Part 2: Outdoor test methods for system performance charaeterisation and yearly performance prediction of solar only systems'. Naron, D. J. Van der Ree B. G. C. M. Rolloos (1998) Bridging the Gap: Reasearch and Validation of the DST Performance Test Method for CEN and ISO StandardsProgress and Preliminary Results -. In Proceedings of EuroSun 98, September 14-17, Portoro~ Slovenia, Goetzberger A. and Krainer A. (Eds), pp. 111.3.10-1 111.3.10-7, The Franklin Company Consultants,

Birmingham. [5] [6]

[7]

Naron, D. J., Van der Ree, B., (1999) 'Bridging the Gap' Final Report of Work Package 1: Definition of Scope, TNO, Delft, The Netherlands. Carvalho M. J.,Busearlet C., Marshall R., Mathioulakis E. Van der Ree B. (1998) Factory Made Systems Thermal Performance: Comparison of Test Methods. In Proceedings of EuroSun 98, September 1417, Portoro~, Slovenia, Goetzbe rger A. and Krainer A. (Eds), pp. 1/1.3.2-1 - 111.3.2-6, The Franklin Company Consultants, Birmingham. Carvalho M. J., (1998) Interim Report: W P 2 Comparison with CSTG test method, INETI, Lisbon, Portugal.


253

R E S E A R C H ON A N E W TYPE OF HEAT PIPE VACUUM TUBE S O L A R WATER HEATER Zhu Ning and He Zinian Beijing Solar Energy Research Institute, No.3 Huayuan Road, 100083 Beijing, China Tel: 86(10) 62001022, Fax: 86(10) 62012880, E-mail: [email protected]

ABSTRACT Anew type of heat pipe vacuum tube solar water heater that can be placed at a very small tilt angle is introduced. A series of experiments was made to test the new system, which indicates that the new system has the same performance as conventional one. A further theoretical analysis of this phenomenon as also made in this paper.

1. INTRODUCTION

2. SOLAR WATER HEATER DESCR/PTION

It is well known that heat pipe vacuum tubes have been used in solar water heater for many years owning to their advantages of anti-freezing, anti-corrosion and low heat loss. However, the tubes are always placed with a tilt angle of more the 15 ~ in north-south direction. This is because it is regarded that the gravity-assisted heat pipe can only work with a minimum tilt angle of 15~ and the condenser must be higher than the evaporator, which obviously limits the application of this type of solar water heater. A new type of heat pipe vacuum tube solar water heater is investigated which can easily installed both in south wall or in the outside surface of balcony and can work well with a very small tilt angle (o._1 of produced energies in the cascade and in the conventional system during a certain time interval (hour, day or season).

Solar collectors

m, 60 ~

m, 60 ~

Ac2 = A c l = Aco /2

Solar collector

T (h) + 50 K

Storage tank M

/r( I

T(h) + 50 K

It(h) ~

I

Storage tank M T(h) = 10 ~

T(h) = 10*C

FPC1 ,~

Ac1

m, 1 0 o c

m, 10~ Fig. 1. Layout of the cascade.

2. THEORETICAL BASIS The properties of the cascade will be compared with the properties of the conventional DHW system by the simulation of the useful energy produced for a day and for a current hour during the daytime. As our task is to

" ISES MEMBER

compare the cascade with the conventional DHW system, we shall consider the conventional DHW system to be a "cascade" of two equal FPCs with the properties of FPC0 and calculate its useful energy by the same method.


The useful energy converted to heat from the solar radiation on the specific surface area during a certain time interval (hour) is accordingly the basic law Q~ = s - u~ . ( r~ - Ta) = b " F~ " ( , a ) - F~ " U~ " ( T , -

ra).

(la)

Here FR'(za)=r/0 and instead o f UL'Tp=FR'UL" 7',. in Eq. (la), we shall consider FR'UL" Ti=kl " Tm as a simplification. Correspondingly

Qum lT']~o- kl "(Tm- Ta)

293

absorber plate temperature Tp describes precisely the instantaneous value of the FPC efficiency. Calculations based on Tm cause some error, but it is small for the FPCs with a long hydraulic duet where (Tp-T,~)25~

(which

is

ever

satisfied

T~2-- T i + AT1 > 25 ~

A cascade of both FPCs connected in the hydraulic duct in series makes it possible to use each FPC ("a" and "e") in their preferred range and provides the maximum of specific yield of the system, compared with the same collector surface area of the FPC0 with the medium characteristic '~:" (a compromise between the light and advanced optical cover). The said positive effect can be provided by the calculation of the daily or/and the current energy yield. 5. INFLUENCE OF THE FPC PROPERITIES ON THE DAILY YIELD Two different regimes for the cascade were simulated and compared on the basis of daily energy yield simulation:


1. The case with constant flow rate mf=const (and nonconstant output temperature To= var). 2. The case with constant output temperature To=const (and nonconstant flow rate my= var). When the average values of output temperatures for both cases are kept equal, the daily useful energy is equal, too. The influence of the FPC characteristics on the calculated daily yield at Irm~=0.6kW/rd 2, Ta= 15 ~ is given in Table 2.

Table 2. Comparison of the FPC data to the energy yield. Conditions FPCI: FPC2:

efficiency factor ~ ("gain"), which is the ratio of the produced energy by a l m2 of the collector surface area during a current time interval (an hour, for instance) in the cascade (AQ,1+AQ,2) to the same in the FPC with the medium characteristic: = (aQ.~ + aQu~) / aQuo >1. (7) The simulation example (Fig. 5) is calculated for FPCs with the formal (but realistic) data: FPCI: 'b", FPC0: '}" and FPC2: "a".

1.4

modest advanced conventional cascade cascade "c . . . . d . . . . e" "c . . . . b . . . . a"

Q., kWhrn "2 if To = const

2.13

2.28

2.42

Q,, kWhm2 if mr= const

2.15

2.26

2.42

average gain (e

1

1.07

1.13

T~(z-

1.21 1.1

.I

10

f

~

~\

60.5

4 "1-

-*--m~2

~- 59.5

2 -[

----rooo

I

8.30

10.30

12.30

14.30

h

Fig. 4. Current output temperature and flow rate in the cascade of similar "00" and diferent "12" FPCs.

6. INFLUENCE OF THE AMBIENT CLIMATE ON THE ENERGY YIEI~ The said positive effect can also be provided by the

: @0.6/15 \\

I\

~

- - - : @0.45/15 - - - : @0.75/15///

\\\ \.~

1/

/

"

." I

8.30

12

\

1.3 ~ - k

1.o

Due to the variable solar irradiation It(h) and ambient temperature Ta(h) during the daytime 8 < h < 16 in the first ease To(h) will vary and in the second case my(h) will vary (Fig. 4). The energy gain of the cascade depends on the inversely proportional data of FPCs, used in the cascade (Table 1). The efficiencies and the particular temperature rise steps are not equal for FPC 1 and FPC2 in the cascade and in case similar FPCs are used in the cascade, the "hot" one works ever ineffectively.

295

I

10.30

I

I

12.30

I

I

14.30

h

Fig. 5. Oltrent efficiency factor of the cascade in a day.

The said efficiency factor ~r increases at the low value of solar radiation being important for the Nordic countries and for the utilisation of solar energy in the morning and afternoon hours as demonstrated in Fig. 5 for the daily performance at the irradiance values I r , , ~ {0.75, 0.6, 0A5} kWrli 2 and ambient temperature Ta= 15~ In Fig. 5 the outlet temperature of the cascade is constant 60~ The dependence of the daily average value of the efficiency factor ~ is presented in Fig. 6 as function of daily average temperature Ta and the maximum value of the solar irradiance Irmax. The character of the functions may to mean that at the lower latitudes a cascade is out of importance, but it's an option to help the utilisation of solar energy in Nordic conditions. An optimisation of the cascade (and its flow conditions) has to be made to find the possible limit of the cascade gain for the selectively matched fiat plate collectors in a DHW system.


296

30

To, ~

25

I r ~ , kWm "2

"~.4P

20

00, 001, 002

0.9 0.g

15

0.7

10 ~

"~o

9Ir,,~ t

1.1

1.05

0.5

t

1.15

0.6

gain ~'e 1.2

Fig. 6. Average efficiency factor of the cascade depending on climate conditions. CONCLUSIONS In the DHW system performing in the single-pass regime, a cascade of two flat plate collectors with different data is feasible. The "cold" fiat plate collector has to be with light and the "hot" one with the advanced optical cover. The gain of the cascade increases with the difference of parameters of the said collectors increasing. The cascade can be recommended just for the modest solar conditions, available in Nordic countries. The positive effect does not depend on the flow conditions of the heat carder, but the regulated flow rate allows to supply hot water with the constant temperature and the latter could be preferred.

initial value of the

efficiency

~e

ratio of the produced energy: cascade versus conventional DHW system AQ~o,AQ,1, AQ,2 the amount of energy generated during a time interval, kWhm2 temperature rise step of a FPC, temperature aT, aT1, aT2 difference, K Ao Acu, Acl, Ac2 collector area, m2 specific heat, klkgqK1 c. FR collector heat removal factor instantaneous time, hour h (hourly) irradiation on the tilted surface, It, It(h) kWhm-2 irradiation amplitude on the tilted surface, Ir.~ kWhm-2 slope of the linearised efficiency curve, kl, kll, k12 kWm-2K-1 heated water mass, kg m g, g~, g2 storage tank, stored water mass, kg specific flow rate,kgslm "2, l ~ ' l m "2 mf(h), my useful heat energy, kV~qlm"2 Qu useful absorbed energy, kWhm2 S fluid temperature, ~ r(h) ambient air temperature, ~ Ta(h) input temperature of the fluid, ~ T,, T~I,T~ (fluid) mean temperature, ~ T. output temperature of the fluid, ~ To, To(h) absorber plate (mean) temperature, ~ r, uL coefficient of the collector overall thermal losses, kWrn2K1

AcknowledgementsmWe thank the Estonian Scientific Fund for supporting these investigations with the grant G-3133 in 1997 and 1998.

NOMENCLATURE za

transmittance-absorbtance product

REFERENCES Tomson T. (1995) Comparative analysis of conventional and splitted solar domestic hot water systems. Proc. Estonian Acad. Sci. Engin. 1, 2, 183- 198.


A SOLAR A B S O R P T I O N AIR-CONDITIONING PLANT USING HEAT-PIPE EVACUATED TUBULAR COLLECTORS HE Zinian and ZHU Ning Beijing Solar Energy Research Institute, Beijing 100083, China

Abstract.A solar-powered absorption air-conditioning plant with 100 kW cooling capacity has been successfully designed and constructed in Shandong Province, China. The plant consists of heat-pipe evacuated tubular collector array, LiBr-H20 absorption chiller, cooling tower, water storage tanks, circulating pumps, fan-coil units, auxiliary oil-burned boiler and control system. The solar collector

array using 2160 heat-pipe evacuated tubular has a total aperture area of 540 m 2 .This paper introduced design characteristics and measuring performance of the plant which has a multifunction of space cooling in summer, space heating in winter and domestic water heating in other seasons. Thermal efficiencies of the collector array are respectively 40% for space cooling, 35% for space heating and 50% for domestic water heating. It was found that the cooling efficiency for the entire system is around 20%.

1. INTRODUCTION Electric power required for providing airconditioning takes a very large portion of total electric power consumption in the world. For this reason, various solar-powered air-conditioning systems have been investigated (Bong et al.,1992.Yeung et a1.,1992. George Lof,1993. Back et a1.,1997) Among them, absorption air-conditioning systems were commonly utilized. The solar cooling has an obvious advantage that the most cooling demand is matched with the strongest sunshine in summer. Besides, solar absorption cooling can be combined with solar space heating and solar water heating so that this comprehensive system will increase the economic benefit of solar air-conditioning. The LiBr-H20 absorption chillers have been widely commercialized. The chiller requires moderately high inlet temperature. In view of thermal performance of the system, a higher inlet temperature will result in a higher COP value. Fortunately, heat-pipe evacuated tubular collectors, developed by Beijing Solar Energy Research Institute (BSERI) and recently produced by Beijing SUNDA Solar Energy Technology Co., Ltd. in China, can meet this requirement ( He,1997 ). They have been used for a solar absorption air-conditioning plant incorporated with a LiBr-H20 absorption chiller at an operating temperature about 88~ This paper described the design characteristics of the completed solar air-conditioning plant, and the primary measuring results under conditions of space cooling,

space heating and domestic water heating.

2. SITE OF THE PLANT The solar absorption air-conditioning plant has been successfully designed and constructed in Yintan Tourist & Holiday-Spending Zone of Rushan, Shandong Province, China. Rushan is located at the southeast end of Shandong Peninsula, which is 360north latitude and 1210east longitude. It is 100 km from Weihai in the east and 150 km from Qingdao in the west, bordering on the Huanghai Sea in the south. In this area, the annual average daily solar radiation is around 17.3 MJ/m 2, the annual average air temperature is around 12.3~ maximum air temperature in summer is 32.1~ the minimum air temperature in winter is-7.8~ Under the local climate, both space cooling and space heating are required for comfort in summer and winter. The solar air-conditioning plant was installed in the Solar Energy Hall of the Chinese Renewable Energy Popular Science Park in this Zone. The Hall is a two-storey building with a construction area over 1000m2 and was architecturally designed to meet requirements of solar collector placement, as shown in Figure 1.

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3. SYSTEM DESIGN

3.1 Layout

Fig. 1. Full view of the solar air-conditioning plant in Rushan, Shandong Province, China The solar air-conditioning plant consists of evacuated tubular solar collector array, LiBr-H20 absorption chiller, cooling tower, water storage tanks, circulating

pumps, fan-coil units, auxiliary oil-burned boiler and control system. Figure 2 shows a layout of the plant.

Fig.2. Layout of the solar absorption air-conditioning plant

ISES Solar World Congress 1999, Volume Ill 3.2 Solar collector array In order to supply the absorption chiller with higher inlet temperature, heat-pipe evacuated tubular collectors have been used. Each evacuated tube has an outer diameter 100mm and a length 2000mm.

The evacuated tube is mainly composed of heat-pipe absorber plate, glass envelope tube, metal sealing cover, getter and others, as shown in Figure 3. The heat-pipe consists of an evaporator section and a condenser section.

To obtain more solar irradiation over a day, evacuated tubes with a semicylindric absorber plate were selected. Theoretical calculation and measuring data show that

the semicylindric absorber plate increases energy gain 10-14% more than the flat absorber plate (He et al., 1997).

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2. metal sealing cover 4. absorber plate 6. getter Fig.3. Configuration of heat-pipe evacuated

As heat-pipe technology is applied to evacuated tubes as well as "dry connection" between evacuated tubes and manifolds is utilized to modules, the heat-pipe evacuated tubular collector has many advantages, such as freeze resistance, fast start-up, high pressure bearing, thermal shock endurance, etc. The solar collector array using 2160 heat-pipe evacuated tubes has a total aperture area of 540 m 2 and a total absorber area of 364 m 2. The collectors were arranged in 9 rows, in which 7 rows were installed on an inclined south-facing roof and 2 rows were installed on a fiat roof with a tilt angle 350. To reduce flow resistance within the system, front 4 rows and back 5 rows were respectively connected in parallel. Then these two parts were connected in series. The solar system is driven by the circulating pump P1.

3.3 Cooling chiller The air-conditioning plant applies a Model LCC-03 lithium-bromide absorption chiller that is made by Dalian SANYO Refrigeration Co., Ltd in China. The chiller has a maximum cooling capacity of 176 kW ( 50 USRT ). The solar collector away supplies hot water at 88. to an inlet of the chiller and the water leaves the chiller at 83.. The chiller produces chilled water at 8. and the water returns to the chiller at 13.. The cooling water temperature through the chiller is 31. and 37. successively. The hot water, chilled water and cooling water through the chiller is respectively driven by circulating pumps P2, P3 and P5. 3.4 Storage tanks There are totally 4 storage tanks in the plant. Volumes of tank 1, tank 2, tank 3 and tank 4 are successively 8 m 3, 4 m 3, 6 m3and 10m 3.

The tank 1 and tank 2 are called as hot water storage tanks and used to store the hot water produced by the solar collector array. The smaller tank 2 aims to reach a specified temperature for the chiller in the early morning. The tank 3 is called as a chilled-water storage tank and is used to store the chilled water to reduce heat losses of the storage tank because temperature difference between the ambient temperature and the chilled water temperature is much smaller than that between the hot water temperature and the ambient temperature. The chilled water between the tank 3 and fan-coil units is driven by the circulating pump P4. These three tanks are also applied to store the hot water for space heating in winter. The largest tank 4 is called as a domestic hot water storage tank. It supplies hot water by means of a heat exchanger inside the storage tank. It also can be used to rescue the plant from overheating when necessary in summertime.

3.5 Auxiliary boiler To ensure an all-weather operation for the plant, the auxiliary oil-burned boiler was installed. The boiler has a rated thermal power of 350 kW with a rated outlet water temperature of 95~ A designed thermal efficiency of the boiler is approximately 88%. 3.6 Controlsystem The control system consists of temperature sensors, electro-actuating valves, network control modules and operator working station. There are totally 9 temperature sensors, 2 three-way valves and 14 two-way valves in the system. Positions indicated by the temperature sensors are as follows: T1 outlet from collector array

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ISES Solar World Congress 1999, Volume !11

300 T2 T3 T4 T5 T6 T7 T8 T9

outlet from tank 1 inlet to collector array water in tank 2 water in tank 1 inlet to chiller outlet from tank 3 inlet to tank 3 water in tank 4

The control system has three main functions: First, it serves to open / close valves and switch on / off circulating pumps to operate the plant for space cooling or space heating or domestic water heating as well as to arrange storage tanks for minimizing heat losses. Second, it serves to light up / off the auxiliary boiler to compensate the thermal power provided by the solar collector army when water temperature is below the specified level. Third, it also serves to prevent the plant from overheating when necessary in summertime and from freezing when required in wintertime.

4. MEASURING PARAMETERS AND INSTRUMENTS In order to determine performances of the plant, following parameters must be directly measured: solar irradiance, inlet / outlet temperatures and water flowrates for the solar collector army, cooling chiller and fan-coil units, water temperatures within storage tanks and ambient temperature. Solar irradianee is measured by an EPPLEY Model PSP pyranometer on the plane of the solar collector army. The pyranometer has linearity of ~0.5% and temperature dependence of fi1%. Inlet / outlet water temperatures for the solar collector army, the cooling chiller and fan -coil units are measured by Pt 100 resistance thermometers. The accuracy of the thermometer is ~0.1~ Water flowrates for the solar collector array, cooling

chiller and fan-coil units are measured by vortex flowmeters. The flow-meter has an accuracy of ~1%. A c o m p u t e r program has been developed for data acquisition and processing so that not only instantaneous flowrate but also accumulated flowrate can be simultaneously read out. Furthermore, instantaneous heat flux and accumulated heat flux can be immediately calculated, incorporating the instantaneous flowrate with the corresponding temperature difference between inlet and outlet of each facility. Data acquisition for solar irradianee, water temperatures within storage tanks and ambient temperature are conducted by a DATATAKER Model DT 600 data acquisition device.

5. RESULTS AND DISCUSSION The solar absorption air-conditioning plant has been operated since November 1998. Its performances for different purposes were measured: space cooling in June 1999, space heating from January to March 1999, and domestic water heating in May 1999.

5.1 Spacecoolingperformance In June 1999, the measuring data indicated that the outlet temperature was generally over 88.on a sunny day, which is necessary for operation of the chiller. Figure 4 shows the variation of solar irradiance with time on June 25. The solar irradiance ranged from 380W/m 2 to 1000

W/m 2 during the day.

The

maximum value at noon was 1047 W/m2. The solar isolation from 8:30 to 16:30 was 20.3 MJ/m2. The variation of solar power gain with time on the same day is shown in Figure 5. The total solar energy gain was 3264.9 MJ.

Fig. 4. Variation of solar irradiance with time.


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Fig.5. Variation of solar power gain with time As a definition, thermal efficiency of the solar collector array is the total solar energy gain divided by the product of the solar isolation and the total absorber area. Using all above measuring data, the collector array thermal efficiency was 44.5%. Figure 7 shows the cooling power in the whole day. It varied from 20kW to 90kW. This implies that the

The variation of the collector exit temperature with time is given in Figure 6. As expected, the collector exit temperature was kept at around 85~ during most of the day on June25. coefficient of performance (COP) of the chiller ranged from 0.4 to 0.7.

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Fig. 7. Variation of cooling power with time The cooling efficiency will be the cooling energy divided by the solar isolation. On June 25, the cooling efficiency of the entire system was 22%. The fact that the cooling efficiency of the entire system is higher than that published in previous literatures, is owe to the higher thermal efficiency of the collectors and the higher COP of the chiller.

Figure 8 is the temperature variation in the tank 2 on June 25, It can be seen that hot water temperature within the take 2 rose very fast and reached 88~ at 9:40without supply of the auxiliary boiler, which met the requirements of the chiller.


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5.2 Spaceheatingperformance The measurements of solar irradiance, solar power gain, collector exit temperature and temperature in the tank2 for space heating, with the same methods as that for space cooling, were conducted from January to March 1999. Similarly, thermal efficiency of the

collector array can be calculated. Table 1 gives the solar isolation, solar energy gain and collector thermal efficiency on Jan.15, Jan.16, Feb.25, Feb.26, Feb.27 and Mar. 1. It indicates that the thermal efficiency of the collector array is approximately 35% for space heating.

Table 1 Solar isolation, solar energy gain and collector thermal efficiency Solar energy gain

Solar isolation (MJ/m2)

Date

1859.9 1840.7 953.3 1251.7 2966.6 2175.3

16.41 15.93 7.85 9.75 24.45 17.13

Jan.15 Jan.16 Feb.25 Feb.26 Feb.27 Mar.01

Collector efficiency (%)

(MJ) 31.2 31.8 33.5 35.3 33.2 35.0

Lowest ambient Temperature (*C)

-2.9 -2.1 4.8 7.2 0.5 6.9

within the tank 2 on Feb. 27. Hot water temperature in the tank 2 also increased quickly Fig. 9. Temperature increase in tank 2 (Feb. 27) without supply of the auxiliary boiler.

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5.3 Domesticwater heatingperformance

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The similar measurements and calculations have been done for domestic water heating purpose. On May 25, water temperature inside the tank 4 changed from 18.2~ at 8:00 to 44.2~ at 10:30, witch was suitable for domestic use. The water temperature increment in the tank 4 along with solar isolation is shown in Table 2.

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10:04

At that moment, the solar isolation was 5.83 MJ/m2, the solar energy gain was 1088.6 MJ, and thus the collector thermal efficiency was 51.3%. According to these data, it is evident that the plant can supply approximately 32 m 3 of domestic hot water above 440C per day.


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Table 2 Water temperature in tank 4 and solar isolation (May 25 )

Time 8:00 8:30 9:00 9:30 10:00 10:30

Upper 19.8 25.6 30.3 35.1 40.1 45.7

Water temperature(~ Lower Middle 16.4 18.5 21.8 24.3 26.6 29.3 31.7 34.2 36.8 39.2 42.2 44.7

6. CONCLUSION The solar absorption air-conditioning plant with 100 kW cooling capacity using heat-pipe evacuated tubular collectors has been designed, constructed and operated in Shandong Province, China. The performance of the plant has been primarily measured and analyzed. The operating results show that the solar absorption air-conditioning plant can be used for space cooling in summer, space heating in winter and domestic water heating in other seasons. This multifunctional plant increases the economic benefit of the solar airconditioning system. The measured thermal efficiencies of the solar collector array are approximately 40% for space cooling, 35% for space heating and over 50% for domestic water heating. It indicates that heat-pipe evacuated tubular solar collectors are suitable for applications at relatively high operating temperature and quite low ambient temperature. As a smaller storage tank is specially adopted, hot water temperature can be raised from 70~ to 88~ in the early morning in summer to meet the requirements of the chiller, and from 40~ to 55~ in the early morning in winter for space heating. The chilled-water storage tank is obviously useful for reducing heat losses of the tank because temperature

Average 18.2 23.9 28.7 33.7 38.7 44.2

Isolation (MJ/m 21 0.00 0.94 1.94 3.02 4.31 5.83

difference between the ambient temperature and chilled water temperature is much smaller than that between the hot water temperature and the ambient temperature. The measured COP of the chiller is approximately 0.70 under solar-powered conditions. The cooling efficiency of the entire system is around 20% under local circumstance REFERENCE Back N.C., Shin U.C. and Jeung S.H. (1997) Study on the Solar Absorption Cooling and Heating System, Proceeding of lSES Solar World Congress 4, Taejon, Korea. Bong T.Y., Ng K.C. and Tay A.O. (1987) Performance Study of a Solar-powered Air-conditioning system, Solar Energy 39, 173-182. George Lof (1993) Active Solar System, The MIF Press, Cambridge, London, England. He Z.N. (1997) Development and Application of Heat Pipe Evacuated Tubular Solar Collectors in China, Proceeding of lSES Solar World congress 2, Taejon, Korea. He Z.N., Ge H.C., Jiang EL. and Li W. (1997) A Comparison of Optical Performance between Evacuated Collector Tubes with Flat and Semicylindric Absorbers, Solar Energy 60, 109-117. Yeung M.R.,Yuen P.K.,Dunn A.and Cornish L.S.(1992) Performance of a Solar-powered AirConditioning system in Hong Kong, Solar Energy 48, 309-319.


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ADVANCED FUZZY CONTROL OF THE TEMPERATURE IN THE TEST CHAMBER

Borut Zupan6i6 and Igor ~;krjanc Faculty of Electrical Engineering, University ofLjubljana, T1-2a~ka25, 1000 Ljubljana, Slovenia, Tel. No.: +386 61 1768 306, Fax No.: +386 61 126 46 31, Email: [email protected]

Aie.~ Krainer and Bo.~tjan Furlan Faculty of Civil Engineering, University ofLjubljana, Jamova 2, 1000 Ljubljana, Slovenia, Tel. No.: +386 61 1768 604, Fax No.: +386 61 125 06 88, Email: [email protected]

Abstract-The paper deals with a systematic approach to the control system design with a final goal to efficiently control some living conditions in a test chamber. The first step of each systematic approach is mathematical modelling. It was based on theoretical approach, which means that the model was developed on the basis of physical laws, energy equilibrium laws etc. The mathematical model was implemented in MATLAB Simulink environment as a simulator entitled KAMRA. The structure is modular, the robust numerical algorithms give accurate results and fast simulation runs. As simulator is implemented in MATLAB environment, it is very easy to implement arbitrary pre simulation and post simulation processing. The validation was made by real measurements as test chamber is equipped with all needed sensors. The developed simulator gives an ideal environment for the design and validation of different control structures. Four fiuzy logic controllers were proposed for efficient control of indoor temperature: for heating, for cooling, for coordination of both subsystems and for roller blind positioning. However a good dynamic response to the reference changes and appropriate disturbances elimination was only one requirement. Energy consumption and some other comfort living conditions (e.g. daylight illumination, ...) were also taken into consideration. Simulation gave applicable information for control system implementation with industrial hardware.

1. INTRODUCTION From prehistoric times bioclimatic conditions in buildings were of extreme importance for pleasant and healthy feeling. As such they represent a process with inexhaustible possibilities for the studying of new control design approaches. The development of new information technology enables new artificial heating and cooling solutions and realisations and harmonized combination with natural resources. So the gab between natural and artificial environment is becoming smaller and smaller. Recent outcomes of the control engineering area are more and more applicable on different areas due an incredible development of software and hardware technologies. Beside traditional PID algorithms which are implemented in industrial hardware for many years, it is also possible to use more advanced techniques, e.g. fuzzy logic, neural nets, expert systems, genetic algorithms, adaptive and multivariable control etc.

There are several important steps in the so called control system life cycle. In the design phase mathematical modelling is an unavoidable phase at least for moderate and severe process complexity. It is also very important that mathematical model is developed with particular aims. Of course models must be verified and validated by extensive number of process measurements. Only well tested models can be transformed into user friendly simulators and efficiently used in control design procedures. So the first stage of control system design is performed by the aid of process simulator. Different control strategies from traditional PID controllers, ON/OFF controllers, lead-lag compensators to more sophisticated approaches which originate

from artificial intelligence area e.g. fuzzy logic or neural net controllers can be tested in simulation environment. The solution which gives the convenient results with regard to design specification (e.g. fast response, zero steady state error, efficient disturbance rejection, ...) is finally implemented with industrial hardware. In this stage many additional functions must be included beside by simulation developed control algorithm. This are functions for operator's interaction, for safe and reliable operations, for monitoring and some more complex supervision functions, for fault detection and diagnosis etc. Unfortunately appropriate simulation results do not guarantee that the control system will satisfactory operate with industrial hardware and real process. Most frequently the reason is a bad mathematical model. However the reason can be also in significant differences between control algori~m implementations in simulation and implementation environments (Zupan6i6, 1998b). 2. MODELLING AND SIMULATION 2.1 Principles of theoretical modelling The designed building simulator (entitled A M R ) is based on theoretical mathematical modelling approach. From many reasons mathematical models are the most suitable and the most widely used category of models. They are concise, unambiguous and unique interpretable, while their manipulation and the evaluation of alternatives are relatively inexpensive. Mathematical model can be defined as a mapping of relationship between physical variables of a system to be modelled into corresponding mathematical structures. The essence to theoretical mathematical modelling lies in the


decomposition of the studied system into particular subsystems, which must be as simple as possible. The corresponding relations between chosen subsystems must then be determined on the basis of different balance equations and physical laws for the area under investigation. In the case of technical systems modelling, the known mass, energy and momentum balances are most frequently used, which gives the overall model expressed by difference or differential equations (Matko et al, 1992). The theoretical modelling of heat dynamics of a room was based on the analyse of thermal conduction, thermal convection and solar radiation and on appropriate energy balance equations (Kladnik, 1987, Krainer and Kladnik, 1995, Zupan6i6 et al, 1998a, Furlan et al, 1998). The properties of the envelope are treated as time-varying parameters as they are variable by their own nature. The variable nature is especially worth for the openings in the building envelope. Frequently openings are equipped with different less or more sophisticated shading systems (e.g. roller blinds), what enables at least changing the shading ratio of opening or even their geometry. In some cases high-tech glazing (electrochromic, fotochromic .... ) are used, where optical characteristics of glass could be changed. The automatic adaptation of such envelope properties appears as a new great opportunity of indirect controlling of the indoor living space parameters according to the current outdoor conditions. 2.2. Important features of the simulator KAMRA Simulator KAMRA can be used for different purposes. In this case it was used for control system design purpose, so only input/output relations will be presented in more detail. The inputs of the simulation model are the outside conditions as well as dynamical parameters of envelope: Variable outdoor (weather) conditions: 9 the outdoor air temperature, 9 the temperature of the terrain, 9 global solar radiation, 9 level of cloudiness and 9 ratio of diffuse/direct radiation. Changeable properties of the building's envelope are: 9 the opaque elements: thermal capacity and resistance of these elements can be changeable, 9 the transparent elements (windows): geometry of openings, optical characteristics of glass and resistance of fill between glass panes are variable, 9 interior properties: absorption, emission coefficients of walls and thermal capacity of furnishing are variable, 9 other characteristic: changeable orientation, Additional heating and cooling: the power of heater and ventilator. The outputs of the simulation model are: 9 the indoor air temperature and interior heat flow 9 the walls, windows and surface temperatures There are following important features of the simulator: 9 It is possible to simulate rectangular building with arbitrary walls, floor and ceiling composition. 9 The opaque elements of the building envelope are floor, ceiling, walls and they are composed of 5 layers, which enables adequate thermal description of different envelope structures.

9

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In each wall one window of rectangular shape could be placed. All windows in the model are supposed to be double-glazed and filled with different gases. 9 The inner space of the building can contain furniture and equipment. The ratio of furnishing/surface of the envelope is flexible, the material properties of furniture are optional. 9 The solar radiation is composed of direct radiation and diffuse solar radiation. The ratio direct/diffuse radiation (DDR) in the model is flexible. 9 The level (CLD) of cloudiness is also an attribute of the outside conditions, as it affects the final amount of direct and diffuse radiation. 9 The orientation of the building is optional parameter and it is defined by the declination angle between real (geographic) south and the direction of building' s axes. The following suppositions are considered in the mathematical model: 9 The whole mass of the inner air is supposed to have uniform temperature. In reality the temperature of the air in the inner space is position dependent function, but the temperature used in calculation is an average temperature of the whole air mass. 9 Temperature changing in directions along the wall or window surfaces are neglected, thus the conduction problems through the envelope elements can be treated as one-dimensional crosswise through them. 9 Whole mass of furnishing/equipment is heated only by the surrounding air. ARer the development of the mathematical model and the simulator concept the appropriate programming tool had to be selected. The most important requirements for the appropriate selection were as follows: 9 Modular and transparent syntax. Model can easily be understand and modified as well. 9 Modern graphic user interface should enable that modelling and simulation unskilled users can efficiently experiment with the model. Users must concentrate to thermal problems instead of problems with modelling, simulation, programming etc. 9 High numerical accuracy and robustness. 9 Fast simulation. 9 Portable models. The selected environment should be a widely spread one, used not only on academic institutions but also in industry. So developed models can be easily transferred between different computers, groups or institutions. 9 With regard to the developed simulation concept the capability for the inclusion of continuous and discrete submodels into the simulation model must be presented. 9 In the chosen environment different toolboxes must give powerful possibilities not only for simulation but also for analysis, design, graphical results presentation etc. 9 If possible, control structures obtained by off line simulation and design can be automatically coded for appropriate target hardware giving efficient real time implementations (Zupan/~it, 1998b). 2.3. Implementation in MATLAB-Simulink environment

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Having in mind the itemised requirements the mathematical programming environment MATLAB-Simulink with appropriate Toolboxes (for control, fuzzy logic .... ) was used. Fig. 1 shows the highest hierarchical level of the modularly constructed simulation model, which is already prepared to serve as a test cell for control system development and validation.

of Ljubljana. Fig. 2 shows the scheme of the test chamber with basic sensors and actuators used in the verification and validation of the simulator KAMR .

Fig. 2. Scheme of the test chamber with basic sensors and actuators used in the verification and validation of the simulator KAMRA.

Fig. 1. Simulation scheme of the simulator in MATLABSimulink environment. In the block Initialization all the parameters about the materials, geometry of window, orientation, geographic location and starting simulation time are given. So the simulation of the behaviour in the case of different materials, orientations, geographic location, position and number of windows and period of the year can be performed. Outdoor temperature and Global solar radiation are defined with appropriate data files obtained from real measurements. They can easily be defined by some other signals from Simulink library. Temperature of terrain, Direct radiation and Cloudiness are prepared as constants or step signals. However in the presented model Direct radiation and Cloudiness are not independent (e.g. if the parameter of direct radiation changes from 0 to 1, the parameter of cloudiness changes from 1 to 0). Up to now described signals are treated as disturbance inputs. The last three inputs signed with Heater, Ventilator and Blind are control inputs as they will be fed by controller signals in order to assure the appropriate indoor temperature. Indoor temperature is the model output or from the point of control system the controlled variable. Of course the simulator can be easily modified so that also other variables of the model can be influenced or monitored. 2.4. Simulator validation The verification and validation of the simulator is one of the most important tasks in each modelling cycle. It is mainly based on the comparison of the measured and simulated results. The verification and validation of the simulator KAMRA was performed with a real system (chamber)- a testing cell built on the roof platform of the Faculty of Civil Engineering, University

The test chamber is a box with all dimensions lm. The south wall is completely glazed, double-glazing is composed of two layers of standard clear glass and air fill, the thickness of wooden frame is 5cn~ The roller blind is as external PVC blind and the alternating window geometry was realised by moving the blind to desired position. Walls, floor and ceiling are composed of dry wall panel lcm, mineral wool 8cm dry wall panel 2cm (from outside). Internal walls are painted in light grey colour. The box is shifted off the ground and the roof is ventilated in order to avoid overheating caused by direct radiation on the roof. Measured values for outdoor conditions were global and reflected solar radiation and outdoor air temperature. Pyranometer CM-6B (Kipp&Zonnen delft BV) was used for measuring solar direct and reflected radiation. Termocouples type T was used for measuring temperature. The temperature of indoor air defines thermal response of the object and it was also measured with termocouples type T. Window size was expressed as ratio of shaded area and whole glazing area. For the purpose of collecting of different samples of the outdoor environment conditions, some series of measurements were executed in different seasons of the year. Position of blind was changed randomly in different time intervals independent of the outdoor conditions. Several measurements were used for appropriate final parameters tuning of the theoretical model of the test chamber. Another set of measurements was used for simulator validation. Simulations were obtained with the measured outdoor temperature and global solar radiation as input variables taken from the experiments as well as with the signal for blind moving r e , m e (see Fig. 3). The comparison of the simulated indoor temperature and the measured one is presented in Fig. 4. The error between calculated and measured values is acceptable in the range of 5-20%. Mainly it is caused by unexpected ventilation heat-losses through some cracks in the dry wall panels and by the influence of wind.


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temperature, radiant flows . . . . ). On the other hand feedback control can also be implemented on numerous ways. From simplest ON/OFF control algorithms to widely used PID algorithms, compensators, cascade controlers, fuzzy controllers to most sophisticated adaptive, predictive and multivariable controllers. The latter are probably not justifiable as the process is not enough complicated and the solutions would be too expensive. Among other possibilities preliminary studies (Zupan~i~, 1998a) indicated that conventional approaches (e.g. PID, ON/OFF . . . . ) give acceptable results in conjunction with particular working regimes. But due to very different regimes, nonlinear behaviour, time varying characteristics, several conventional controllers with appropriate switching mechanisms should be developed. Our preliminary investigations confirmed that fuzzy control approach could represent more appropriate solution. 3.2. Fundamentals of fuzzy logic Fig. 3. Blind position, global solar radiation and outdoor temperature.

Fig. 4. Comparison of the measured and simulated indoor temperature.

The development of fuzzy logic in the late 70s has provoked many applications and enlargements of a simple fuzzy controller (Babu[ka, Verbruggen, 1996, ~krjanc et al, 1996, ~krjanc et al, 1997a, ~krjanc et al, 1997b, Kav~ek et al, 1997). Further investigations pointed out certain advantages over traditional (e.g. PID) approaches. Beside satisfactory robustness in case of more complicated nonlinear and time varying systems probably the most important advantage refers to design principle, which is very similar to human reasoning. This principle is based on simple conditional rules. The design and tuning of fuzzy controllers consists of membership function design and definition of fuzzy rules. However no general tuning methods exist so the design approach is based on real or simulation trial and error experiments. So the control system can be designed also without precise mathematical model (with direct experiments on real process). Using Fuzzy control toolbox ANFIS in MATLAB environment the trial and error experimentation approach is very easy. The designer must select appropriate membership functions, fuzzy rules and some other parameters. With this the controller is automatically prepared for Simulink simulation environment. So it can be easily combined with the simulator KAMRA.

3. CONTROL SYSTEM DESIGN AND VALIDATION

3.3. Proposed control scheme

As mentioned the final aim was to develop a control system for pleasant and comfort living conditions but also for economic energy consumption. The comfort was mainly determined with indoor temperature. However the appropriate daylight was also important as it can significantly influence comfort behaviour and energy consumption as well. It is well known that optimization of each technical system starts with different and contradictory demands so the most important step in the design procedure is to choose appropriate criterion functions.

Fig. 5 depicts the Simulink simulation scheme of heat dynamic control system. The heater and ventilator are controlled by two fuzzy systems, each of them consists of two fuzzy controllers. The inputs of both fuzzy systems are control error and error derivative. With such approach a fuzzy logic controller can be treated as a dynamic controller. Obviously heating and cooling (ventilation) are not simultaneously in operation. This operation is harmonized by the third fuzzy feedforward controller entitled fuzzy logic coordinator. The output of this system is the third input of the fuzzy controllers for heating and ventilation. The blind positioning is controlled by the fourth fuzzy logic feedforward controller with the following inputs: solar radiation, direct solar radiation and day index (indication of summer and winter period respectively).

3.1. Possible control strategies Different control strategies can be taken into account: from feed forward open loop control schemes to closed loop regulation (feedback control). Feed forward is usually a cost effective solution but in our case unsuitable as the system is exposed to significant disturbances (time varying external

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For the clearness and simplicity only two inputs, which are important for dynamic characteristics - control error and its derivative are presented. Appropriate membership functions and fuzzy rules were obtained with many experiments. Fig. 7 shows the membership functions for Fuzzy PD part for control error, for its derivative and for controller output (heating power).

Fig. 7. Membership functions for PD controller.

Fig. 5. MATLAB Simulink simulation scheme of the heat dynamics control system. 3.4. Fuzzy logic controller for heating Fuzzy logic controller for heating consists of a parallel structure of proportional derivative (I'D) and proportional integral (PI) part. It is a kind of fuzzy PID controller. PD part enables fast response and appropriate damping and PI part eliminates the steady state error. The appropriate simulation scheme is presented in Fig. 6.

Fig. 6. Simulation scheme of the fuzzy logic controller for heating.

The mnemonics in conjunction with membership functions have the following meaning: N negative error ZE zero error P positive error DEC decreasing error INC increasing error ZE zero heating LH low heating HH high heating After the inputs and output are appropriately determined with fuzzy sets, the appropriate rules between inputs and output should be defined. For Fuzzy PD part 6 rules were used: 1. If (E is N) and (ED is DEC) then (HPD is ZE) 2. If (E is N) and (ED is INC) then (HPD is ZE) 3. If (E is ZE) and (ED is DEC) then (I-IPD is ZE) 4. If (E is ZE) and (ED is INC) then (HPD is ZE) 5. If (E is P) and (ED is DEC) then (I-IPD is LH) 6. If (E is P) and (ED is INC) then (HPD is H ~ E means error, El) means error derivative and HPD means heating (contribution of PD part). Beside membership functions and fuzzy rules, the following methods in fuzzy approach were selected: Decision method for fuzzy logic operators AND: MIN Decision method for fuzzy logic operators OR: MAX Implication: MIN Agregation: MAX Deification: CENTROID (centre of gravity) The same methods were used in all other fuzzy controllers.


Fig. 8 shows the membership functions for Fuzzy PI part for error, error derivative and for output respectively. NI-G

::/ o.sl-

|/

=~ o/

1

018

0 i6

0.4 .

. 0.2 . . .0

0.2

0.4

Error IN

~

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3.5. Fuzzy logic controller for cooling Cooling is in simulator KAMRA modelled as a negative heat flux from the ventilator. The structure of fuzzy logic controller for cooling is very similar to the heating one. It also consists of a parallel structure of PD and PI part. The membership functions are mirror images of membership functions for heating.

Thefuzzymembership 1i

309

0 i6

o18 |

3.6. Simulation tests for heating and cooling

'

The described fuzzy controllers were validated with several simulation studies. Heating was tested with the following parameters: ~ o| I I I I I I I I I winter (January 31) 0.05 Season: 0.03 0.02 0.01 0 0.01 0.02 0.03 0.04 0.05 0.04 Error derivative Outdoor temperature: 0~ i 1~ ~ J ~ i~ I I Temperature of the terrain: 0~ Global solar radiation: 400 W/m2 ~ 0.5 Parameter for direct radiation: 1 I Parameter for cloudiness: 0 0 I I I I I i I 3 2.5 2 1.5 1 0.5 0 0.5 Ventilator power: 0W Output Roller blind: shut Fig. 8. Membership functions for PI controller. Fig. 9 shows the transient response of the indoor temperature when the reference temperature is changed from 18 ~ to 20 ~ The mnemonics in conjunction with membership functions have the following meaning: 2O NEG negative error SN small negative error ZE zero error SP small positive error POZ positive error IN increasing of indoor temperature Z no change in error 18i DE decreasing indoor temperature 2.5 2.55 2.6 2.65 2.7 D decreasing of output Time [s] xlO~ SD slow decreasing of output NA no action Fig. 9. Indoor temperature as the response to the reference SI slow increasing of output change. I increasing of output The relations between inputs and output are determined with It can be noticed, that the control system has appropriate the following fuzzy rules: performance. The response to the step change is fast, the 1. If (E is Z) and (ED is DE) then (I-IPI is NA) overshoot is small (about 0.04 ~ The steady state error 2. If(E is Z) and (ED is Z) then 0-IPI is NA) vanishes in reasonable time. This confinm the integral action of 3. If (E is Z) and (ED is IN) then 0-1PI is NA) the controller. 4. If (E is SP) and (ED is DE) then 0-IPI is SI) Cooling was tested with the following parameters: 5. If (E is SP) and (ED is Z) then (HPI is SI) Season: summer (July 15) 6. If (E is SP) and (ED is IN) then 0-IPI is NA) Outdoor temperature: 30 ~ 7. If (E is POZ) and (ED is DE) then 0-IPI is I) Temperature of the terrain: 20 ~ 8. If (E is POZ) and (ED is Z) then (HPI is I) Global solar radiation: 500 W/m 2 9. If (E is POZ) and (ED is IN) then (I-IPI is SI) Parameter for direct radiation: 1 10. If (E is SN) and (ED is DE) then 0-IPI is NA) Parameter for cloudiness: 0 11. If (E is SN) and (ED is Z) then 0-IPI is SD) Heating power: 0W 12. If (E is SN) and (ED is IN) then 0-IPI is SD) Roller blind: shut 13. If(E is NEG) and (ED is DE) then 0-IPI is SD) Fig. 10 depicts the indoor temperature and the cooling power 14. If (E is NEG) and (ED is Z) then 0-IPI is D) of the ventilator when reference temperature changes from 18 15. If (E is NEG) and (ED is IN) then (HPI is D) ~ to 20 ~ and back to 18 ~ E means error, ED means error derivative and I-IPI means heating (contribution of PI part). ~o.sF

[/


310

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The output of the fuzzy logic coordination system is connected to the third input of the heating and cooling fuzzy controllers.

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3.8. Simulation results of fuzzy logic coordination

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i

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Fuzzy logic coordination was tested with the following parameters: Season: April 30 Outdoor temperature: 19 ~ Temperature of the terrain: 10 ~ Global solar radiation: 400 W/m e Parameter for direct radiation: 1 Parameter for cloudiness: 0 Fig 11 shows the indoor temperature in the presence of reference change.

8

xl0'

20

Fig. 10. Indoor temperature and cooling power as responses to reference temperature change. Cooling power is signed by negative values. At the first reference change the ventilator completely stops for a short time. In this period the indoor temperature increases due to the high outdoor temperature. It can be noticed that lower reference temperature needs higher cooling power. The control system again assures zero steady state error. For the sake of simplicity heating and cooling systems were tested independently. It is obvious however that the appropriate switching coordination is needed in order to prevent simultaneous heating and cooling actions and to assures bumpless transfers from cooling to heating phase and minimal energy costs. 3.7. Harmonisation of heating and cooling with fuzzy logic coordination It is very important to appropriately define the season for heating and the season for cooling. Unfortunately theirs starting and ending dates are not known in advance or fixed. The heating period is sometimes extended in May and the cooling period in October. The best additional information for economic heating and cooling is the difference between the reference and outdoor temperature. If this difference is positive, the cooling should be suppressed and if it is negative, the heating should be suppressed. So the fuzzy logic coordination seemed to be a right solution for the mentioned problem. It is implemented as a feedforward control as the action of the controller does not depend on the indoor temperature. The first input is the day index. January 1 means day index 0 and December 31 day index 365. The one year range was split to three periods: Jan. 1 -May 15, May 15Oct. 1, Oct. 1 - Dec. 31. All three ranges are covered with three trapezoidal membership functions with very small overlapping periods. The second input is the difference between the reference and outdoor temperature. It is fuzzyfied with three symmetrical membership functions in the range -30 ~ to 40~ The output variable is described with three symmetrical membership functions (active heating, inactive heating and cooling, active cooling). The fuzzy inputs and outputs are linked with nine fuzzy rules.

!

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19 18

.

.

.

.

.

i

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2.5

2.6

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Fig. 11. Indoor temperature in the presence of reference change - fuzzy logic coordination. Until the outdoor temperature is higher than the reference temperature, the heater is switched off. When the reference temperature exceeds the outdoor temperature (which is 19 ~ the heater is switched on and the indoor temperature reaches quickly the reference one. Again the overshoot is very low and the steady state error is zero. 3.9. Fuzzy logic controller for roller blind positioning The roller blind is installed in the south window of the tested chamber. As one of the important aim of the control system is also to assure the economic energy consumption, the appropriate control of the roller blind positioning is very important. The fuzzy control of the roller blind is implemented as a feedforward control as the action of the controller does not depend on the indoor temperature. The fuzzy controller has three inputs: day index, global solar radiation and direct solar radiation. The latter one is again connected with cloudiness. The first input- day index is fuzzyfied in the range 0-365 days with three membership functions as already described in fuzzy logic coordination system. The second i n p u t - global solar radiation has the range 0-500 W/m2. It is fuzzyfied with five triangular equally spaced membership functions for very weak, weak, moderate, strong and very strong radiation. The third input - direct solar radiation has the range 0-1 and two membership functions connected with terms: sun shines (no clouds), sun does not shine (clouds on the sky). The output of the controller is the position of the blind. This variable is fuzzyfied on the interval 0-1 with five triangular membership

311


functions for different roller blind shading areas. The input and output signals are linked with sixteen fuzzy rules. Ten rules are used in conjunction with cooling season. Here it is possible to achieve considerable energy savings as roller blind suppresses the warming due to moderate and intensive global or direct solar radiation. Beside energy consumption criterion the daylight illumination is also taken into account. During heating season however the priority was given to living conditions in comparison with energy consumption. In the case of clouded sky the roller blind remains opened. In sunny weather however the fuzzy logic shades the window at the most for 50%.

5OO

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3

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5

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Time [s]

x104

Time [s]

x104

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0.8 0.6 04 oo

3.10. Simulation tests for roller blind Different simulations were performed for cooling and heating periods. In the example presented in Figs. 12-14 the influence of global radiation disturbance to indoor temperature in the heating period was studied. The following experiment parameters were used: Season: summer (July 15) Outdoor temperature: 25 ~ Temperature of the terrain: 20 ~ Global solar radiation: 200 W/m 2 Parameter for direct radiation: 1 Parameter for cloudiness: 0 During observation the reference temperature is changed from 18 ~ to 20 ~ and back to 18. ~ In the moment t=-50000s the global solar radiation is changed from 200 W/m 2 to 500 W/m 2. Fig.12 depicts the reference and indoor temperature. Fig. 13 shows the global solar radiation disturbance and the appropriate reaction of roller blind. At the beginning the roller blind is already lowered to app. 0.6 due to direct solar radiation. At the moment when the disturbance occurs the reaction of the roller blind is instantaneous as it is controlled with feed forward approach. Both diagrams i n Fig. 14 illustrates however the cooling power, which is needed for appropriate control. It can be noticed, that the cooling is completely switched off for a short period after the reference changes to 20 ~ At t=50000s the fuzzy controller in the cooling system intensifies the cooling power. All these actions cause that the radiation disturbance does not effect the indoor temperature (see Fig. 12).

40o

o

Fig. 13. Global solar radiation disturbance and the response of the roller blind. o

-10 -20

..~ -40 0

o -50

r~ _76~0

1

2

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4 ; Time [s]

6

7

8 xl0 4

-5

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i

4

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Fig. 14. Cooling power. In the presented example the disturbance was eliminated by the appropriate roller blind positioning and by increased cooling. Other simulations show that the roller blind is more effective in ease of direct solar radiation disturbance. It is also worth to emphasise that criteria for good leaving conditions (e.g. good illumination) and low energy consumption are contradictory, so appropriate compromises must be considered.

o [..,

4. CONTROL SYSTEM I M P L E M E N T A T I O N 17.5

o

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S

xlO4

The advanced fuzzy control system, which was designed with simulation, was finally implemented with industrial programmable logical controller MITSUBISHI. The control schemes are defined in a special progranmaing environment IDR BLOK on PC computer. This is a graphical block oriented editor, so the programming is very similar as in Simulink. IDR block library contains many elementary arithmetic blocks (e.g. summer, gain .... ), as well as more control oriented blocks (e.g. analog and digital inputs and outputs, P ID controller, fuzzy controller . . . . ). This low level control functions were improved with supervision level which was realized by Factory link. As

312


this application is linked to the Microsoft Access database a very complex and flexible experimenting environment was obtained. The described implementation will be systematically tested in near future. 5. CONCLUSIONS

Krainer A., Kladnik R. and Perdan R. (1997). Light and Thermal Energy Coordination in Building, PLEA 1997 Book of Proceedings, Kushiro, Japan. Matko D., Karba R. and Zupancic B. (1992). Simulation and Modelling of Continuous Systems: A Case Study Approach, Prentice Hall Int., New York.

The life cycle of complex control systems consists of many complicated phases. In this paper some of them, which demand high level of knowledge about modelling, simulation and control system design approaches as well as about modem software and hardware technology, were discussed and implemented. The development of mathematical modelling is probably the most crucial part. As the validation of the model was successfully accomplished, it was later used for the design of control system based on fuzzy logic. The proposed control scheme which consists of two feedback and two feedforward controllers is only the first attempt with very promising results. A lot of work has to be done with more detailed evaluation of the proposed control system in the simulation and implementation environment. The experiments must accurately evaluate all influences to energy consumption and living conditions as well. There are many new possibilities for new modelling and control system design approaches. Experimental modelling or the combination of theoretical and experimental modelling (e.g. with neural nets) can result in more accurate model for control design purpose. There are of course many other possibilities for control system design. It is expected that the best results will be achieved with the combination of traditional methods (e.g. PID ) and methods that originate in artificial intelligence (e.g. fuzzy logic, neural nets, genetic algorithms, expert systems .... ).

~krjanc I., Kavgek-Biasizzo K. and Matko D. (1996a). Fuzzy Predictive Control Based on Fuzzy Model, Proceedings of 4~

European Congress on Intelligent Techniques and Soft Computing, EUFIT-96, Vol.3, pp. 1864-1869, Aachen. ~krjanc I., Kav~ek-Biasizzo K. and Matko D. (1996b). Fuzzy Predictive Control based on Relational Matrix Models, Computers chem. Engng, Vol.20, pp. $931-$936, Elsevier Science Ltd. ~krjanc I., Kav~ek-Biasizzo K. and Matko D (1997a). RealTime Fuzzy Adaptive Control, Engineering, Applications of Artificial Intelligence, Vol.10, No.l, pp.53-61, Elsevier Science Ltd. ~krjanc I. and Matko D. (1997b). Fuzzy Adaptive Control versus Model-reference Adaptive Control of Mutable Processes, Methods and Applications of Intelligent Control, Edited by Spyros G. Tzafestas, Kluwer Academic Publisher, Dordrecht.

Zupan6i6 B. and Klop6i6 M. (1995). Environment for the Simulation and Design of Control Systems, Proceedings of the Session "Software Tools and Products", Eurosim Congress '95, Ed. F. Breitenecker, I. Husinsky, Vienna, pp. 147-150.

REFERENCES

Autar R. and Bakker L. G. (1998). Smart integrated control of lighting and solar shading for oJ~ces, Solar control, EU-Joule III Project, Contract JOR3CT960113, TNO Building and construction research, Delft.

Zupan6i6 B., Krainer A. and ~krjanc I. (1998a). Modelling, Simulation and Temperature Control Design of a Test "Chamber", Proceedings of 1998 Summer Computer Simulation Conference, Ed. M. S. Obaidad, F. Davoli, D. DeMartinis, Reno, USA, pp. 173-178.

Babu~ka P,. and Verbruggen H. B. (1996). An Overview of Fuzzy Modelling for Control, Control Eng.Practice, Vol.4, No.ll, pp. 1593-1606.

Zupan6i6 B. (1998b). Extension software for real-time control system design and implementation with MATLAB- Simulink, Simulation Practice And Theory (6)8, pp. 703-719.

Duffle J. A.and Beckmann W. A. (1991). Solar Engineering of Thermal Processes- 2ndedition, John Willey&Sons, Inc.

ACKNOWLEDGMENTS

SOLAR CONTROL-integrated approach to solar techniques, Furlan B., Krainer A. and Perdan R. (1998). Measurements of thermal Response of Test Object with variable Geometry of Openings, Comparison to Computer Simulation, EuroSun98 Book of Proceedings Vol. 1, Portoro~, Slovenia.

CE DGXII JOULE-THERMIE, JOR3CT960113 (1996-1999)

Intelligent Control System for efficient use of Energy and indoor climate Conditions in the Building, Ministry of Science ofrep. Slovenia, L2-7695

Kladnik R. (1987). Theory of KAMRA, publication, Faculty of Civil Engineering, University ofLjubljana, Slovenia. Krainer A. (1994). Toward Smart Buildings, TEMPUS Joint European Project JEP 1802, Building Science and Environment-Conscious Design, Module 1: Design Principles, London.

The Interaction between dynamical Openings and Building Envelope, Ministry of Science ofrep. Slovenia, J2-9080


XX. Solar, Thermal and Photovoltaic Concentrating Collectors

313

314




315

DESIGN AND CONSTRUCTION OF A LINE - FOCUS PARABOLIC TROUGH SOLAR CONCENTRATOR FOR ELECTRICITY GENERATION

Bakos G.C., Adamopoulos D. and Tsagas N.F. Democritus University of Thrace, Department of Electrical Engineering and Electronics, Laboratory of Energy Economy, 1 Vas. Sofias Str.- Tel. 30 054179725,Fax. 30 054179734, 67100 Xanthi - GREECE, E-mail: [email protected] [email protected]

Soursos M. SENERS Energy Systems, 16 Kleovoulou Str., 11744 Athens - GREECE

Abstract - The design of a parabolic trough concentrator (PTC) used for electricity generation is presented in this paper. PTCs are the preferred type of collectors used for steam generation due to their ability to work at high temperatures with a good efficiency. Parabolic trough collectors are frequently used for solar - thermal applications because temperatures of about 300 oc can be obtained without any serious degradation in the collector efficiency. An important issue about solar concentrators of this kind is the thermal behavior analysis. Moreover, the conversion efficiency is determined. The different parts of the parabolic trough concentrator such as the metal frame, the parabolic mirrors, the solar energy absorption system are described. The sampling and control unit is tested on a similar small - scale PTC model which is developed and installed inside the Energy Economy Laboratory. A simulation program is developed which combines the level of solar irradiance, meteorological data, orientation and dimensions of the parabolic mirrors aiming to a better evaluation of the system efficiency. The advancement in this project is the use of parallel processing (real time processing) using OCCAM language for the system control and analysis of various parameters. Suggestions for system performance improvement are given concerning the choice of heat transfer fluid. The possibilities about combined operation of such a solar system using an MHD generator rather than conventional generators are examined.

I. INTRODUCTION The design of a parabolic trough concentrator (PTC) used for electricity generation is presented in this paper. PTCs are the preferred type of collectors used for steam generation due to their ability to work at high temperatures with a good efficiency. Parabolic trough collectors are frequently used for solar - thermal applications because temperatures of about 300~ can be obtained without any serious degradation in the collector efficiency. A typical application of this type is the Southern California power plants known as Solar Electric Generating Systems (SEGS). About 2500000 m 2 collectors have supplied more than 4000 GWh of electricity into the Californian grid [Kearny and Price, 1992, Jesch, 1998]. Current commercial PTC thermal electricity generation uses collectors which comprise an evacuated - annulus receiver consisting of an inner stainless steel tube mounted in a concentric evacuated cylindrical glass envelope which serves to minimize convective and conductive losses [ Norton, 1992, Kalogirou, Lloyd and Ward, 1997 ]. An important issue about solar concentrators of this kind is the thermal behavior analysis. Moreover, the conversion efficiency should be determined. The trough concentrator was designed following appropriate strength and pressure criteria. The different parts of the

parabolic trough concentrator such as the metal fimne, the parabolic mirrors, the solar energy absorption system are described. Furthermore, the sampling and control unit is tested on a similar small - scale PTC model developed and installed inside the Energy Economy Laboratory. The peripheral electronics, including the interface to PC and the sampling software, were developed and tested. Theoretical results, produced by an appropriate simulation program, are compared with experimental results taken from the experimental facility which is installed outdoors and is available at a specially formed area nearby Democritus University of Thrace. The system performance, under the specific environmental conditions and ground configuration noticed in the area of Xanthi in Northern Greece, was tested. A new simulation program is developed which combines the level of solar irradiance, meteorological data, orientation and dimensions of the parabolic mirrors aiming to a better evaluation of the system efficiency. Another important issue is related to the thermal loss of these facilities. The reduced performance of the experimental facility in comparison to the theoretical system is due to the existence of irregularities in the experimental solar receiving system and the quality of the heat transfer fluid which, in the case of experimental application, is tap water. Suggestions for system performance improvement are given. For instance, a good choice of heat transfer fluid should be

316


made. The transfer fluid could be steam, to be used directly in a steam turbine, or a proper oil. Although it seems to be an expensive solution, the use of a thermochemical storage medium is under discussion. Another important issue related to these systems is their cost especially when they are intended to be used in a commercial scale. The overall cost of the PTC system described in this paper is estimated at approximately $10000 and is a unique application for Greece. It is designed to be used only for research purposes. The research work in our Laboratory is also focused on the possibilities of a combined operation of such a solar system with a non - conventional electricity generation system such as MHD (magnetohydrodynamic) generator [Mankins, 1996]. Some commercial companies and local authorities showed great interest to sponsor this effort since it is the first solar application of this type in Greece. 2. DESCRIPTION AND FUNDAMENTALS OF THE PTC SOLAR SYSTEM The experimental facility is installed at a specially formed area nearby Democritus University of Thrace. The sampling and control unit is tested on a similar small - scale PTC model developed and installed at Energy Economy Laboratory. Concentrating collectors of that kind present operational difficulties compared to fiat - plate collectors [Duffle and Beckman, 1991]. Except at the very low end of the concentration ratio scale, they have to be oriented to 'track' the sun so that beam radiation will be directed onto the absorbing surface. There is also need for maintenance in order to retain the optical system operation in a satisfactory level for long temporal periods in the presence of dust, nasty meteorological conditions, oxidizing or other corrosive atmospheric components. The total facility of the PTC consists of the following parts: a) metal support frame of the PTC b) parabolic mirrors c) solar radiation absorbing system (pipes) d) solar orbit "tracking" system e) sampling system- control panel 2.1 Metal support frame of the PTC A schematic diagram Of the metal frame is given in Fig. 1. It consists of two fixed base parts and three movable parts. The first movable part is a turret which is resided to the low base by a couple ofjoints. This gives the possibility for the turret to move (spherical coordinates - v movement). There is a small carriage arranged onto the turret which slips from the top to the bottom point of the turret. One of the two edges of the parabolic mirrors rotation axis is mounted onto this carriage. The other edge is supported from an horizontal barrier which is arranged through a couple of joints to the large base of the syster~ The combination of the turret - carriage system is used as a "tracking" system of the seasonal solar orbit. When the small carriage moves onto the turret, the last one starts to incline from the vertical position leading to the inclination of the parabolic mirror rotation axis from the horizontal position. This happens in order to achieve a perpendicular direction, for the rotation axis, to the incident solar in~adiance. This is very important in order to achieve maximum efficiency of the system.

Fig 1. A schematic diagram of the metal frame. The various movable and immovableparts are depicted. It is noticed that the seasonal level of solar orbit is different regarding to the horizontal level of the site. The turret height is the appropriate one, so when the small carriage reaches the top point, the achieved inclination of the turret at that moment, brings the rotation axis of the parabolic mirrors to the appropriate slope from the horizontal level (ground). This leads to the conservation of the verticality for the incident solar radiation. This boundary case appears for the lowest observed solar orbit. There is a parabolic metal lattice supported to the rotation axis of the parabolic mirrors. The metal lattice's size is corresponding to the total area of the parabolic collector surface. It is used for the rotation of the parabolic mirrors (daily tracking of the sun). The metal lattice is constructed covering the appropriate mechanical criteria of strength. When the rotation axis of the mirrors is horizontal, then a rotation within 180 degrees is achievable. Tracking of the daily solar orbit is achieved by rotation of the horizontal metal lattice around the PTC axis independently from the vertical movement of the system used for the seasonal solar orbit tracking. A natural view of the facility is given in Fig. 2. 2.2 Parabolic mirrors The total collector surface consists of 4 parabolic mirrors of 3m2 (1640 m m x 1700 ram) each. The incident solar radiation is focused to a light line (focus point) which is parallel to the rotation axis, at a distance of 1700 ram. Mirror thickness is 4 mm, which is small compared to the total magnitude of its surface. These mirrors where imported from the German company Flachglas Flagsol GMBH. They have high reflectance (98 %) and their efficiency was tested at companies' laboratory.


317

and the mirrors can be modified for a few centimeters in such a way that the tubes are positioned to the focus line. This kind of regulation takes place only once when the parabolic mirrors are installed and is not repeated during the measuring procedure. The order of magnitude for the absorbing pipes cross section was chosen according to the broadness of the area which the collected solar radiation occupies on the focus line (magnitude of sun's image).

Fig. 2. A natural view of the metal frame in the facility of the PTC systc~n. The points through which the parabolic mirrors are mounted onto the metal frame allow micro - scale adjustments to achieve an exact focusing of the incident solar radiation. 2.3 Solar radiation absorption system Solar radiation absorption system consists of metal pipes (tubes) which are placed parallel to the rotation axis. Their placement coincides with the focus point. The pipe cross section is 4 cm and the wall thickness is small so that quick transfer of energy from solar radiation to the working fluid is achieved. The outer pipe surface is covered with a special black paint (selective surface) which increases the absorbance of the incident solar irradiance and reduces, simultaneously the reflectance. The most important issue about the absorption system of that pipes, is that they are covered externally with glass tubes of equal length, concentrically placed. These tubes have a 9 cm cross - section. The area between the glass tubes and the absorbing pipes is hermetically closed. The internal pressure is less than the atmospheric pressure. The usefulness of this section is significant since it reduces the convection heat losses. The external glass tube reduces the heat losses due to radiation, because the produced heat from the metal surface of the absorption pipe is "trapped". At the edge of the absorbing pipes, a flexible tube is used for the conveyance of the heat transfer fluid. A proper oil is planned to be used for the large scale facility, but for the experimental operation of the small scale facility, tap water was used. For the circulation of the fluid, a regulating flux pump is used ( provided with a fluxmeter). The absorption pipes are supported to the focusing line by a number of trapezoidal metal light - weighted barriers, whose axis of symmetry is perpendicular to the rotation axis. At the support points of the tubes, there are joints which allow two axis regulation. This is very helpful in order to improve the focus quality. In other words, the distance between the tubes

2.4 The solar orbit tracking system The support system of the parabolic mirrors is constructed in order to track the sun. The orientation of the parabolic trough collector need to be the appropriate one, so that the incident solar beam will fall always perpendicularly. The orientation is achieved by two independent movements of the whole system, as it was referred to the previous paragraph. The first movement is for "tracking" the solar daily orbit. It is achieved by the use of a DC motor of 1,5 KW. The power of the motor is transferred to the rotation of the PTC axis through a gear system with a significant reduction ratio (1: 6000). This is necessary to accurately orient the optical system, since the mirrors should continuously change their position according to the slow movement of the sun in real time. The second movement is also achieved by a motor of the same nominal power. This movement corresponds to the seasonal solar orbit "tracking". The philosophy of the second movement of the PTC system was given in the previous paragraph. The motor movement is achieved trough the Pulse Width Modulation ( PWM ) technique. 2.5 Sampling system - Control panel The sampling system monitors three parameters: temperature, solar radiation density and heat - transfer fluid flux. The light intensity measurement is also used for the positioning of the optical system [ Bakos, 1991 ]. There were developed appropriate programs and interface systems using sensors and amplifiers for the experimental results. The position checking is being done using an absolute rotary encoder. The control panel provides the possibility to move the system manually or automatically through the computer. Normally, the system operation is taking place through the computer. There is continuous evaluation of the pre - referred parameters together with continuous storing of data through the interface system to the computer. 3. SYSTEM M O D E L L I N G 3.1 Thermal analysis In this part of the report, the thermal analysis of the PTC system is described. Also a description of the developed simulation program used for the calculation of the efficiency and the fluid outlet temperature as a function of solar radiation intensity and characteristic parameters of the PTC is given. The energy absorbed by the heat transfer fluid per unit area is given as follows: q = h(Tp - Tf) where h = heat transfer coefficient

( 1)


318

dT

Tp =pipe temperature Tf --fluidtemperature The useful energy absorbed by the fluid can be calculated from the difference between the incoming solar radiation and the total heat loss:

q=

Ief - ( F t + Fp )-(Tp - T a )

Fp = passive area thermal loss factor T a = ambient temperature We notice that the actual absorbed solar radiation Ief is evaluated from the incident solar radiation Io using the appropriate equations: Ief = I o 9cosCO where V&is the hour angle evaluated for the current relative position of the sun to the earth :

h - [ I e f - ( T f - T a)-(F t + Fp)]

h+Vt +Vp

4

d . u . Cp ~ - ~ q dy def

(4)

d = fluid density u = fluid speed Cp = specific heat

where D is the diameter of the absorber pipe. The conversion efficiency n is given from the following equation: Q

n = ~

3.20n'ginal system simulation According to the thermal analysis described in the previous section, an appropriate simulation program was developed. It is a conventional sequential (serial) computer program The input data to the simulation program is shown in Table 1. The results were taken for a line - focus PTC (small - scale facility) of an active pipe length of 1.7 m. There were many graphs produced and conclusions derived concerning the PTC operation. In this paper some of them are presented. These results were taken for an indicative summer day, with flux rate of 50 g/s. Table I

1

h

collector's width W absorber's pipe diameter D1

[Ief - ( T e f - T a ) ( F t + Fp)]

def ducp h+F t+Fp

(5) If we place: -

-

-

4 ~

def

---

+Fp

Parameter

d ef = collector's active diameter From equations ( 3 ) and ( 4 ):

C

h+Ft

(7)

S c = collector's active surface (m 2)

where

~

Q =V(DL. h .[Ief - ( T f - T a)-(F t + Fp)]

(3)

The temperature increase d Tf of the heat transfer fluid in a

B

(6)

I o = incoming solar radiation ( W / m 2)

length dy of the absorber pipe when energy q is absorbed, is given as follows:

A

e_BL

where

where t is the solar hour. From equations ( 1 ) and ( 2 ) we take :

dy

C

I o "S c

co =15~ (t-12)

4

C

To =-~+1 Ti - ~ l .

absorber pipe length L. The values of B and C can be calculated from the pipe and heat - transfer fluid characteristics. The useful energy Q absorbed from the pipe is given from the following equation:

F t = active area thermal loss factor

dT

and after numerical integration:

Equation ( 6 ) gives the outlet temperature To of the heat transfer fluid as a function of the inlet temperature T i and the

I ef = absorbed solar radiation

dTf

=C-B-Tf

(2)

where

q=

dy

~

1 ducp

A-h (Ft 9 +Fp) h+Ft +Fp A-h

h+Ft +Fp

[Ief 9 + T a .(F t + Fp)]

then equation ( 5 ) can be written as follows:

Value 1.64m 0.04m

shield's diameter D2 transmission coefficient 12 absorption coefficient 1. reflectance r active area thermal loss factor F t

optional (0.3 m)

passive area thermal loss factor Fp

0.2 W/m 2 K ~

ambient temperature pipe length L flux rate u solar radiation I o

Ta

0.85 0.92 0.9 2 W / m 2K~

290 K 1.7m 50 g/s selection

In the simulation program there is an option to put a proper parabolic piece of metal upwards the absorber pipe which is often called shield. The shield is aimed to reduce the heat

319


losses from the absorber. The absorber loses radiation only in directions unprotected by the shield [Twidell and Weir, 1990]. The PTC facility is not provided with such a shield, but there is a possibility to evaluate the concentration ratio of the collector with or without shield and find out the difference from derived output of the simulation program. Since the shield use increases the concentration ratio, the PTC energy conversion efficiency is also increased. The mathematical formulas for the evaluation of the concentration ratio for both of these cases are: W-D

With a protecting shield: C r =

2

3.3 Current system simulation The advancement in this project is the use of parallel processing (real time processing) using OCCAM language for the system control and analysis of various parameters. The simulation software is being written also in OCCAM language, running on a T800 transputer of the 1NMOS company. 1NMOS Transputer is a high performance single chip computer whose design facilitates the construction of parallel processing systems. The Transputer executes OCCAM programs more or less directly [Pountain D., 1987].

(8)

BEGIN

vC]DlZ~ 3600

i

where

INPUT Ti Io I

C r 9concentration ratio D 1" absorber's pipe diameter D 2" shield's diameter W" collector's width z angle representing the uncovered absorber part

:

W Without shield" C r = ~ (9) V(lD1 The output of the simulation program is the concentration

I

INPUT PTC PARAMETERS DATA .J FOR EACH HOUR OF THE DAY FIND ACTUAL SOLAR RADIATION FOR COLLECTOR ORIENTATION

ratio C r , thermal conversion efficiency n, output temperature T o and produced thermal power P. In Fig. 3 there is a schematic diagram for a PTC provided with protecting shield. In the simulation program there is also derived an evaluation of the maximum attainable output temperature Tmax for ideal quality of the PTC operation. For instance, when the PTC is provided with a protecting shield, Tmax could be achieved when the shield allows radiation to move towards the mirror side. Of course, this temperature cannot be practically achieved because parabolic trough collectors are not perfectly parabolic and part of the useful heat is removed as long as the heat - transfer fluid pass through the absorber pipe. The flowchart of the original simulation program is shown in Fig .4.

Y

o

FIND CONCENTRATION RATIO (SHIELD UNPROTECTED) I

L~ FINDTo ~-J FIND Tmax FIND ENERGY ABSORBED FIND CONVERSION EFFICIENCY

~sTE

THE~ ULTS /

GO TO THE NEXT HOUR

Fig. 3. Scheme of a shield protected PTC where W is the PTC width and z is the angle representing the uncovered absorber part.

Fig. 4. The original simulation program flowchart. The angle mentioned into that flowchart, concerns the z angle for the uncovered part of the absorber when a protecting shield is applied to the PTC.

320


The advantage of this approach is that the PC can control and analyze the system input parameters (such as collector positioning, weather parameters etc.) simultaneously leading to a better system efficiency evaluation. It can be used as an input data, information from measurement elements, concerning ambient temperature, sky clearness index, wind speed etc. Variations of these parameters would entrance continuously into the OCCAM language program attaining real time processing evaluation. The main idea is represented in Fig. 5.

line expresses the same factor for a PTC system without shield protection (concentration ratio value is 13).

Fig. 6: Output temperatureTo versus incident solar radiation Io. The high solar radiation gives an increased efficiency of the PTC system. It is also expected that a concentration ratio increase leads to the conversion efficiency increase. The difference is obvious for low values of Io. When we have high values of Io the efficiency difference for these two values of the PTC concentration ratio is getting smaller but is still significant (a 10% difference is not considered negligible). The conclusion is that using a well designed protection shield in a proper position, results an increased efficiency for the PTC system.

Fig. 5: The current simulation program flowchart. Input data is given continuously. 3.4 Simulation results In this paper, four graphs concerning the small scale PTC system operation are presented. In Figure 6, the output temperature To variation, as a function of incident solar radiation Io is shown. Continuous line represents the simulation result while the dotted line corresponds to the output temperature experimentally attained, for the small scale PTC system. As it was normally expected, the output temperature increases with the incident solar radiation. We also observe a linearity between these two magnitudes for high values of Io. For Io = 800 W/m2, the simulated output temperature is To = 440 ~ while the actual value is To = 416 ~ Given that the working fluid was tap water this result i s considered encouraging. In normal conditions, steam production is the output of the system for high values of Io (at the middle of the day or early alternoon). In Fig. 7 the thermal conversion efficiency variation n, as a function of the incident solar radiation is given. The dotted line shows the variation of the factor n, for a shield - protected PTC (concentration ratio value is 32) while the continuous

Fig. 7: Thermal conversion efficiency n versus incident solar radiation Io. In the following graph (Fig. 8) the absorbed thermal power P variation is shown as a function of the incident solar radiation Io. The dotted line represents the variation of the power P, for a shield- protected PTC (concentration ratio value is 32) while the continuous line expresses the same magnitude for a PTC system without shield protection (concentration ratio value is 13). The absorbed thermal power is higher when concentration ratio is increased, as it was expected, because the conversion efficiency factor is higher. Considered that the simulation program concerns a small - scale PTC facility using tap water, the derived level of the absorbed thermal power is encouraging.


In Fig. 9 there is a comparison between simulation and experimental (actual) results for the PTC system operation. The thermal conversion efficiency n as a function of the incident solar radiation Io is shown.

321

is still under experimental operation and improvements should be applied. 4.2 Magnetohydrodynamic and solar energy combinan'on Based on the existing work which has been done so far, there is a strong intention to form a PTC system used as a heat source for a magnetohydrodynamic (MHD) generator rather than a conventional generator. A schematic diagram of an MHD generator is given in Fig. 10. One of the most promising plasma applications is the magnetohydrodynamic (MHD) generators. At this kind of generators, there is an overheated gas flow with high velocity through a proper magnetic field B , leading to the production of an electromotive force E. The magnetic field is developed into a duct which is called MHD channel (Fig. 10). The

Fig. 8: Thermal power absorbed (P) versus incident solar radiation Io. The continuous line represents the numerical results committed by the simulation program. The dotted line represents the experimental results of the small - scale PTC system for the corresponding solar radiance values Io. It is expected that the actual performance of the PTC is reduced compared to the simulation procedure.

intensity B of the magnetic field should be significantly high. The magnetic field source is a superconducting magnet or field coils which is preferred for experimental applications (Fig. 10). During the MHD power production procedure, conversion of thermal energy into electricity is taking place. The applied magnetic field direction is perpendicular to the direction of the plasma flow. As a result, we have reduction of the total hot - plasma energy during its movement into the MHD channel. This energy reduction leads to the electricity production. The derived amount of electric energy can be driven to a proper load. In figure 10, the incoming plasma to the MHD channel has velocity u directed along the X axis, the magnetic field direction is along the Z axis while the direction of the produced electric field is along the Y axis (see figure 10). The MHD duct (channel) is rectangular shaped. The electrodes used for the electric field development produced by the plasma movement are the MHD channel walls. The operation principle of the MHD generator is shown in Fig. 10. The incoming hot - plasma is partially or fully ionized so the main idea is to drive the corresponding charges to the electrodes. That is achieved because of the properly directed applied magnetic field. If the electrodes are not attached to an output load there is no closed electric circuit neither electric current. For the case of the closed circuit, the general form of Ohm's law gives:

where J Fig. 9: Thermal conversion efficiency n versus incident solar radiation Io. Simulatedand experimentalcomparison. 4. ELECTRICITY PRODUCTION 4.1 General aspect Having a maximum output temperature To= 440 ~ is quite difficult to produce superheated steam directly and drive it to a conventional steam turbine for electricity production. The alternative solution is to use the produced thermal power into a preheat cycle aiming to the reduction of the fuel quantity needed for steam superheating. It is our intention to extend the PTC facility by purchasing additional mirror and absorber components. The whole system

is the current density, 11 the specific plasma

conductivity, E and B the electric and magnetic field intensity which influence the moving charge carriers having velocity u . It is easily derived that the developed electrode voltage is: AV=u.B-z Plasma velocity ( u ) is reduced into the MHD channel. Part of the plasma kinetic energy is transformed to electric energy. Half of the total power derived from the gas plasma flow is consumed for the internal plasma heating. The rest of it, is provided to the output load. Basic demands for the system operation are high plasma flow rate and high conductivity. It is also important to have a high plasma density in order to assure a sufficient level of input power. When the plasma ionization degree is high, the demand for high conductivity is fulfilled. Therefore, it is intentional to introduce a plasma seeding with low ionization potential, like Cesium.

322


type of generator, at the inlet of MHD channel the nonequilibrium plasma clots are initiated by high power electron beam. After that, a supersonic gas flow drags this layers through the cross magnetic field of high intensity and produce electric energy. The numerical simulation has shown that enthalpy extraction ratio near 42% and isentropic efficiency 85% could be achieved [Slavin et al, 1996]. For the creation of a solar MHD generator on the basis of solar concentrator there is a fundamental scientific problem which must be solved: the development of heat source which consists of solar concentrator and heat accumulator for space conditions at working body temperature (probably noble gas) of about 2000 ~ 5. CONCLUSIONS

Fig. 10: Schematic diagram of a magnetohydrodynamic generator (MHD channel) and its operation principle. M H generators drive significant amount of power through a small resistor. That means high current intensity. Therefore, use of conversion devices is necessary to control that power. MHD power production is an open research field. According to many researches which have been taken place worldwide, it is estimated that MHD generators could give a reliable solution to the energy problem which will appear during the 21 st century. Currently, there is cooperation on this issue with foreign institutes. MHD electricity production using solar power seems very promising especially for space applications where the incident solar radiation is significant compared to the incoming irradiance which reaches the earth. The project of space power plant using solar radiation as a source of thermal energy is very attractive. However, there is a problem of cooling at the working cycle of a heat engine in space conditions. Removal of heat is possible only through thermal radiation. Emission of thermal energy from the surfaces of radiators to outer space, is effective only at the temperature about 600 ~ for power higher than 1 MW, that determines the temperature of heater in about 2000 ~ at efficiency level near 30%. Unique opportunities occur by using an M I - I generator combined to a concentrator of solar energy aboard of a space power plant, because MHD generator can effectively operate working body temperature of about 2000 ~ Until now, the effective closed cycle MHD generator has not been created. Experiments conducted in Tokyo Institute of Technology on the MHD facility of a Hall type, the best result have been obtained: the enthalpy extraction ratio has reached 38%. This result satisfies the requirements for perspective industrial MHD power plants. However in Hall generator it is difficult to obtain the satisfactory results on the second major parameter which is isentropic efficiency [Okamura, 1994]. This should not be less than 70%. It is easy to fulfill this condition in Faraday- type MHD generator. The new idea for realization of Faraday type closed MHD generator using stratified recombinated plasma flows of the noble gas has been suggested [Slavin et al, 1996]. In a given

In this paper, there was a description of a parabolic trough concentrator (PTC) design and construction. There was also a comparison between simulation and experimental results for a small- scale model of the PTC system. During the experimental procedure, the PTC was moving in two directions. In one direction the axis of PTC was vertical to the incident solar radiation following the seasonal solar orbit variation and in the other direction, it follows the solar daily orbit variation. These movements are almost impossible to be realized simultaneously in a large scale PTC solar system. In this case, the absence of one of these movements costs an additional heat loss part of incident solar energy. Similar behavior is observed between numerical and experimental results. The difference noticed is very logical and depends on technical reasons. The absorbing pipes used in the experimental measurements were provided only with glass covers without vacuum and their quality was poor. As it was mentioned earlier, the heat - transfer fluid used in the experiment was tap water with low anti-corrosion protection. For the numerical calculations high quality vacuum pipes (Farnell - Philips) and better quality heattransfer fluid than water was considered. The possible uses of a similar larger system could be hot water production for industrial use and electric energy production for combined use with natural gas systems. There is a strong intention to form a PTC system used as a heat source for a magnetohydrodynamic (MHD) generator rather than a conventional generator. This research field seems very interesting especially for space applications. The described PTC system can reach a conversion efficiency level of 35 - 45 %. This could probably increased even more by applying proper improvement techniques. In the experimental procedure there was also a serious disadvantage ; the reduced dimensions of the PTC system introduces difficulties towards the precise behavior estimation of large scale facilities. REFERENCES

Bakos G.C. (1991). Three dimensional (3D) acoustic and vision systems. Ph.D. Thesis, University of Liverpool, UK. Duffie J.A. and Beckman W.A. (1991). Solar Engineering of thermal processes. 2~dEd., J. Wiley & sons, New York.


Jesch L. Solar thermal power . Renewable Energy World, 1998, 1, p. 52-53. Kalogirou S., Lloyd S. and Ward J. (1997). Modeling, optimization and performance evaluation of a parabolic trough solar collector steam generation system. Solar energy, Vol.60, No 1, p. 49 - 59. Keamey D.W. and Price H.W. (1992) Solar thermal plantsLUZ concept. In Proceedings of 2*d Renewable Energy Congress, Reading UK, Vol.2 ,p. 582-589. Mankins J.C. (1996). A flesh look at space solar power. In Proceedings of 31st Intersociety Energy Conversion Engineering Conference (IECEC - 96), Washington D.C., v.3, p. 451 - 457. Norton B. Solar energy thermal technology (1992). Springer Verlag, Heidelberg, Germany. Okamura T. et al (1994). Review and new results of high enthalpy extraction experiments at Tokyo Institute of Technology. In Proceedings of SEAM - 32, Pittsburgh, USA. Pountain D. (1987). A tutorial introduction to OCCAM programming. 1NMOS Company, UK. Slavin V.S., Danilov V.V. and Sokolov V.S. (1996). Closed cycle MHD generator with non - uniform gas plasma flow driving recombined plasma clots. In Proceedings of 31 ~t Intersociety Energy Conversion Engineering Conference (IECEC - 96), v.2, p. 836-841, Washington D.C., USA. Slavin V.S., Lobasova M.S., Finnikov K.A., Danilov V.V. and Sokolov V.S. (1996). Numerical simulation of MHD process in the planned experimental facility with non uniform gas plasma flow driving recombined plasma clots. In Proceedings of 12ta International Conference on MHD electric power generation, Yokohama, Japan. Twidell J.W. and Weir A.D. (1990). Renewable energy resources. 2=dEd., University Press, Cambridge, p. 130- 133.

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324

THE DUCT SELECTIVE VOLUMETRIC RECEIVER: POTENTIAL FOR DIFFERENT SELECTIVITY STRATEGIES AND STABILITY ISSUES Xavier G. Casals Departamento de Fluidos y Calor, Universidad Pontificia Comillas-ICAI, C/Alberto Aguilera 23, Madrid, 28015, Spain, Tel./915422800 ext 2366, e-mail: [email protected] Jose Ignacio Ajona Departamento de Energia Solar, Viessmann, c/Volta 4, Poligono Industrial San Marcos, Getafe, 28906, Spain, Tel./91-6820911, email: [email protected] Abstract - Recently much theoretical and experimental work has been conducted on volumetric receivers, but not much attention has been paid to the possibilities of different selectivity mechanisms to be used in order to minimize radiation thermal losses which are the main ones at high operating temperature. In this paper we present a duct volumetric receiver model and its results, which allow the evaluation of different selectivity strategies such as: conventional e/a, geometry, frontal absorption and diffuse/specular reflection. In recent work on volumetric receivers based on simplified models, it has been concluded that the duct volumetric receiver is inherently unstable when working with high solar flux. We didn't find any unstable receiver behaviour even at very high solar fluxes, and conclude that a substantial potential for efficiency improvement exists if selectivity mechanisms are properly combined.

1. INTRODUCTION In order to achieve high thermodynamic conversion efficiencies, modem combined power cycles or advanced gas turbine cycles demand high quality thermal energy input, with maximum working fluid temperatures in the range of 1700 K. When trying to solarize or hybridize such a power cycle, high temperature solar thermal collector-receiver technologies are required if no limitations are to be imposed on the maximum conversion efficiency or solar fraction. Therefore, we need a solar thermal concentrating technology able to work at high temperatures with a high receiver thermal efficiency. Amongst the different technologies available for electricity generation with solar thermal power plants, the central receiver technology is one that seems to have a high technical potential for power generation in the 100-200 MWe range, because of introducing the least limitations on the temperature level at which solar energy can be introduced in modem thermodynamic power cycles. From the central receiver technologies available, volumetric receivers seem to be the most appropriate to reach these objectives. Different types of volumetric solar receivers have been tested since 1981. However, meama'ed receiver efficiencies and air outlet temperatures have not been as high as one would e ~ t from the volumetric concept. Early results from Odeillo (Menigault, Flamant and Rivoire, 1991) were ofhth = 66 % at Tso = 500 ~ for packed bed receivers in 1981 and lath= 73 % at Tgo = 600 ~ for honeycomb structure receivers in 1983, being unable to work at higher air outlet temperatures because of overheating at the front surface. This problem was overcome with fluidized bed receivers, but with lower effieieneies: lath= 47 % at Tso = 800 *C. From 1987 on, tests on different volumetric receivers have been done in the PSA (Almeria-Spain). With the Sulzer metal wire receiver (1987) one of the highest effieieneies was measured giving lath= 75 % at Tgo = 700 ~ (Becker, Cordes and Bfihemer, 1992). The ceramic foil receiver (B6hemer and Chaza, 1991) operated at Tgo = 782 ~ with lath= 59 % , and the ceramic foam

receiver at Tgo = 730 0(3 with lath = 54 % (Chavez and Chaza, 1991). Although these first kind volumetric receivers were not optimized, and operating solar irradiations were under 0,8 MW/m 2 (for tests in PSA), these efficiencies look rather low. Operation with modem gas turbines would demand higher air outlet temperatures, and therefore lower receiver efficiencies, in order to reach a significant solar fraction, but overall eonvertion efficiency would be very low. Higher gas outlet temperatures have been measured in the last years for both open and dosed volumetric receivers. In (Pitz-Paal, Morhenne, and Fiebig, 1991) results are given for several open volumetric receivers tested in the solar fumance of DLR (Cologne-~y) with non-homogeneous solar irradiation averaging and peacking up to 1,3 and 1,9 MW/m2. The highest air outlet experimental results where of lath = 82 % at Tso = 800 0(2 with the corrugated foil receiver, 1~ = 38 % at Tgo = 1100 ~ with the SiC honeycomb receiver, l~h = 7 1 % at Tso = 820 *(2 with the wire mesh receiver and hth = 58 % at Tso = 810 ~ for the ceramic foam receiver. In (Karni et al, 1996) the highest air outlet temperatures are reported for the DIAPR closed volumetric receiver with secondary concentration, attaining lath= 74 % at Tso = 1200 *C operating at 4 MW/m2 solar irradiation. To overcome the limitations of low operating tempertaures and efficiencies, one thinks of introducing selective strategies. In (Becker, Cordes and B6hemer, 1992) it was suggested that a high potential existed for receiver efficiency improvement by using selective coatings with a high a/e ratio. However, the first try to benefit from selectivity effects was proposed in (Flamant, Menigault and Olalde, 1987), with a different strategy: a two-slab closed selective volumetric receiver, based on disposing a semitransparent volumetric material (best results were obtained with a silica honeycomb) above a packed bed of absorbing particles. After theoretically and experimentally optimizing the receiver, they measured lath= 70 % at Tgo = 814 ~ operating with a 0,67 M W / m 2 almost uniform solar flux in a solar fumance (Variot, Menigault and Flamant, 1992). The same selective effect was proposed in (Pitz-Paal, Morhenne and Fiebig, 1991) for an


open receiver, using the geometry of a duct volumetric receiver. The first experimental results from this receiver (Pitz-Paal and Fiebig, 1992) were ofhth = 60 % at Tgo = 750 ~ with 0,6 MW/m2 average solar irradiance. Recently much theoretical and numerical work has been conducted on volumetric receivers, but no attention has been paid to the possibilities of different selectivity mechanisms to be used in order to minimize radiation thermal losses, which are the main ones at this high operating temperature. The numerical model we have developed accounts for the main radiative, convective and conductive heat transfer fenomena, and allows for the analysis of different selectivity approaches, such as conventional a/e selectivity, geometry, frontal absorption and diffuse/specular reflection, allowing the axial variation of thermophysical and thermo-optical receiver properties to explore further selectivity mechanisms that could arise from it. We conclude that there exists a high potential for receiver efficiency improvement, mainly by introducing solar especular behaviour in the duct walls. Conventional a/e selectivity applied to the inner tube walls doesn't have a high effect on receiver efficiency, but it does have a high effect in the frontal surface exposed by the receiver. Minimising this frontal surface by sharpening the tube walls at the inlet is one of the most important issues to be able to benefit from the full potential of the other selectivity strategies. In recent work on volumetric receivers (Kribus, Ries and SpirE, 1995), (Pitz-Paal et al, 1997) it has been concluded that the duct volumetric receiver is inherently unstable when working with high solar flux (above aprox. 1 MW/m2). The models on which these conclusions have been drawn are ot~en based on simplified radiation exchange formulations and do not include all flow inertial effects, non developed flow entrance friction and convection phenomena, temperature dependence of therrnophysical properties other than viscosity and radiation selectivity mechanisms. Many of these assmnptions are not appropriate for the duct receiver, and can si~ificantly modify the temperature distribution along the duct, and thus its performance. In the model we developed we have tried to overcome most of these limitations, and when exploring stability issues, we didn't find any unstable behaviour even at very high solar fluxes ( 10 MW/m2).

2. THE DUCT RECEIVER MODEL The conjugate heat transfer problem in the receiver may be formulated by an integro-differential second order equation if convection effects are determined through Newton's law of cooling with a heat transfer convection coefficient based on experimental correlations. This equation, together with the proper integral energy and momentum balances in the air flow, permit computation of wall and air temperaturs along the duct. Spectral effects are considered by allowing different surface behaviour for the solar and infrared radiation exchanges. The infrared radiative terms are formulated with the enclosure theory (Siegel and Howell, 1992), on such a way that allows for convenient linealization in the iterative resolution procedure. Solar irradiation terms are formulated using conventional configuration factors (Siegel and Howell, 1992) when duct walls have a solar diffuse behaviour, and specular configuration factors (Lin and Sparrow, 1965), 0Labl, 1977) when duct walls have solar-specular

325

behavior. Dependence of all thermophysical and thermodynamical air properties (r, ca, , m , k) with temperature is retained in the formulation, because, due to the high air temperature increments when flowing through the receiver its properties can experiment variations up to 400 % for conductivity, 200 % for dynamic viscosity, 80% for density and 60 % for specific heat. Air flow will be laminar due to the small tube diameters considered (2-8 mm), and depending on the Reynolds and Rayleigh numbers, forced or mixed convection heat tansfer may be found (Shah and London, 1978). Rayleigh number will be higher at tube inlet, and secondary flows could develop due to bouyancy forces. Also, due to the small L/D ratios considered, significant portions of tube lenght will have a thermally and hydrodynamically developing flow with important effects on heat and momentum transfer. All of this is taken into account by using the appropriate convection and friction correlation to evaluate the Nusselt number and friction factor (Holman, 1997), (Shah and London, 1978). The integro-differential equation is linealized and iteratively numerically solved with a second order finite difference scheme. Appropriate relaxation factors are included to reach convergence of the linealized iterative scheme due to the non linealities introduced by radiative infrared terms, and convergence studies performed to reach proper wall and gas ten-q3emturedistributions. The results presented here are obtained with a one-chnanel model for an open volumetric receiver, and are therefore strictly valid for uniform irradiation at receiver inlet, which is approximated by the use of secondary concentrators. When solar irradiation at receiver aperture is not uniform, the fundamentals of the receiver behaviour are still in the one chanel model, but the results for parallel channels have to be combined in order to get a precise description of receiver performance as in (Spirkl, Ries and Kribus, 1997) and (Pitz-Paal et al, 1997). Nevertheless, as our focus is on finding potential improvements by introducing selectivity strategies, and all this information is included in the one channel model there is no point on combining parallel chanel results, which moreover depend a lot on solar flux distribution in receiver aperture, and therefore on concentrator geometry and configuration. To discuss the receiver performance and the effect of the different selecivity stategies, we find it convenient to introduce the equivalent emissivity defined by Eq. (1), which is the ratio of the radiative receiver losses to the losses that would exist to mantain the inner cavity at the end of the tube with its temperature without tube wall. Therefore this parameter permits to evaluate the selective capabilities of the duct receiver.

=

--eq

q rad,loss

(i)

a . jr~ . D2 //4 . (y4+2_ T1 )

3. DIFFUSE SOLAR ABSORPTION We'll begin by considering the duct solar receiver with a diffuse reflection of incident solar irradiation. This is the case of the volumetric receivers tested up till now and therefore will allow us to evaluate the ability of the model to predict receiver performances as well as the possibilities for performance improvement when introducing selectivity strategies in these

326


receivers. When solar radiation undergoes diffuse reflection in the duct walls, in order to minimize reflection losses it is necessary to have a high solar absorption, and therefore most of the radiation is absorbed in the inlet tube region. This leads to a negative axial temperature gradient in almost all the tube with a decreasing absolute value as one moves towards the end of the tube. This is so as well with and without frontal absorption. As a consequence, temperatures in the inlet region and thus radiative losses are very high, and the selectivity potential of the duct geometry completely diluted. The selectivity strategies for these receivers involve appropriate choosing of emissivity and absorptivity in the tube surface as well as on the tube front, and thermal conductivity. High conductivities will improve receiver performance because of allowing a conductive transport of thermal energy in the front towards the tube end, reducing therefore wall temperatures in tube entrance (radiative losses), and allowing a higher air temperature outlet. In fact, wall conductivity is one of the factors that most affects receiver equivalent emissivity, but its effect saturates at values in the order of 50 W/m-K. High solar absorptivities will improve receiver performance because of reducing reflection losses, in spite of generating higher inlet tube temperatures and eoq. The reduction in reflection losses is so important with diffuse solar behaviour that the effect of high solar absorptivity is more important than the effect of low infrared emisivities. Low infrared emissivities will improve receiver performance because of reducing e, a. The tube inlet region is the one that has the higher axial temperature gradients, solar absorption, and view factor of the aperture, therefore the influence of a , e and k is mainly restricted to this region. When the tube leading edge is not sharp, frontal absorption becomes very important, increasing the wall inlet temperature and axial ~ t u r e gradient, and therefore radiation and reflection losses, with a very important reduction in receiver efficiency. In this situation, thenno-optieal properties of the tube edge become very important and dilute the effect of thereto-optical properties in the tube wall. High solar frontal absorption keeps on being very ~ r t a n t because of its reduction in reflexive losses in spite of its increase in radiative losses. Low frontal infrared emisivity is also very important because of reducing era. Tube wall a, e, and conductivity follow the same trends as before, but they are less important than ale and ele. In Table-1 we can find the effect of the different parameters on a diffuse tube without frontal absorption, while in Table-2 we present these results in the ease of frontal absorption. The base ease referred at these tables, as well as in Fig.l, is given by IJ i l l 5 ; D=2mm ; e--0,8 mm; R~=150 ; a---ale=0,8, e=ele=0,2 ; k=-50 W/inK. Solar irradiation is 4 MW/m 2 in the case without frontal absorption and 2,4 i W / m 2 with frontal absorption giving air outlet temperatures in the order of 1800 K in both cases.

T~ (K) 9

Base Case

i~

eel

0,79

1,04

!Dhth

De~l

|

1808

9

I

Modification (%) D T s o m

m

a = 0,95

1,73

mm

.9

m

a = 0,2

~

- 16,49

I

a = 0,2 - 0,8

0,00 -3,70 m

-3,70 m

-9,77 m

1,38 9

k =2-50 W/mK

0,00

-5,77

48,53

-5,77

48,53

- 15,11

166,62

2,16

-16,26

-.0,81

7,74

m

k = 100 W/mK mm

0,00

m

k = 2 W/mK mm

-10,44

!

e = 0,2 - 0,8 mm

-25,31 m

e = 0,8 m 9

0,90

I

,d

mm

2,73 mmM

m

-0,52

Table-l: Solar diffuse without frontal absorption.

In Fig.1 we present some wall and gas temperatme distributions for a diffuse receiver with frontal absorption and different design parameters in order to illustrate the above-mentioned comments. The receiver performance in the different cases from Fig. 1 may be found in Table-2. In Fig.2 & 3 we present the effects of conventional e/a selectivity on hth and e, a for a diffuse receiver with frontal absorption working at different air outlet temperatures (solar fluxes). As we may see, for increasing gas outlet temperature, the effect of e increases, and the one of a decreases. Receiver efficiencies on Fig2 are in the order of those obtained in volumetric receiver tests. Ts, (K) Base Case m

1841 m

Modific. (%) 99

hth 0,69 m

Tle (K)

1,32

2118

Deeq

DTle

9

) DTp m

%q

Dhth 9

9

I

i

ale - 0,2 mm

-17,94

9

ale = 0,95

3,85

ele = 0,8 9

mm

-4,19

m

9

0,73

-9,30

4,18

-6,50

72,00

9

, -8,26

-19,4

9

6,05

-6,01

m 9

a = 0,2

-27,3 9

9

-12,8

m

mn

-9,63

-5,05

9

9 J

a = 0,95 ii

0,91 i

82

e = 0,8 mm

i 1,43

-3,32 9

m

i

-5,16 9

1,08

0,51

ml

-3,90

36,19 9

1

m|

I

k = 2 W/mK mm

- 10,64 9

k = 100W/mK m 9

- 16,3 m

1,56 9

2,44 9

20,45

155,7 l

m

In

-3,18

- 15,3 9

Table..2: Solar diffuse with frontal absorption

9

iron


qdA,, ~i~ I d l ~

; I~

327

design desicions. In Fig.4 we present such a plot to evaluate the appropriate tube diameter in a solar diffuse with frontal absorption duct volumetric receiver. Such a figure should be complemented with Fig.5 which provides air outlet temperature.

n-tO0;~K

26OO

IJD~8 D -'2imi

O - 0,8 man

Z;k,,k 4,,,O0WbIK;T

qu/A,,2Mmtn

~,,MItK

o

[---o--- T. I I b " 0,2

tm

S"

----~-- a ,, O~ - - - B - - k m3 l l ~ K

400

in"

400

o- % ' U

l i b " OJI

180 UD,, t8

300 (phg 0

~ " 0'8 imi

M

~) aM

o

u

o,4

u

u

t.2

1

--o--D" 8 n

1010

--x--D,, 8 i m

]d.

Fig.l: Diffuse with front absorption. Temperature distributions. 0 k,,kb,,B)W/al.K;Ri

D,,180;T

al,,li~lK;

Ib,~8;

1100

800

1800 Tpl~

- o - - p,0,0; --~e-0~;

a=0.O 000,3

-x-

,,..0,3

~ 0,o ;

21100

Fig.2: Diffuse with front absorption. Effect o f conventional ;electivity at different solar fluxes on receiver efficiency.

Till " 8 4 3 K ; k1r

80 ; iilo ,, 0.8 ; elo ,, 0,2 ; RED,, 100 ;

1,0 k " 80 W/m.K

1.8 1.7

UD,, 18

1.8

D . . 2 mm

e,, 0.O mm

qr I 1.8 1,4

I --o--e- 0,O ;

1.3 1.2

",,0il

a - 0,2

-•

I~ ;

~o,,

0,2 : - - 0il

I

Y

1.1

1

On Fig.4 we may see that there is a very pronounced bend on the curves, wich means that once this point is reached, for a 5ttle increment in receiver efficiency, we'll get a huge increment in fan power consmnption to have the air flowing through the receiver. Clearly there will be a point from which it won't be worthwhile to pursue higher receiver efficiencies because extra fan power consumption will exceed extra energy generation, and will therefore have a negative impact on the overall plant efficiency. For two receivers working at the same efficiency, the best choice will be the one of lower pressure loss, and therefore, the envelop of curves in Fig.4 will define the most apropriate diameter to work at each receiver efficiency. This means that for the range of receiver efficiencies between 60% and 80 % the best choice seems to be D = 4 - 6 rnm. However, when solar receiver is coupled to a thermodynamic cycle, outlet air temperature has also influence on overall efficiency, and therefore the information on Fig.5 is also ing)ortant. Here we see how for a given air outlet temperature the smaller the receiver diameter, the higher the receiver efficiency, but what is not seen in this figure is that pressure loss for a given outlet temperature becomes much higher as receiver diameter is decreased.

I---~o" U ; - " U qo#~,,2MIm

1 ?00

U

md mass flow.

o ,,0.8 mm

O,4

U

Ng.4: Solar diffuse with frontal absorption. Effect of diame~

UD,, t8

"~'~,,~

O,4 ha,

u

0,6

U

eb,,0,2

1200

1700

2200

Tp00

U

S;k"k

b"O01id(;T

a"E48K

-

an" II b - O,il

U-

Fig.3: Diffuse with front absorption. Effect of conventional ~electivity on equivalent emissivity.

0,7-

i"

U-

%=0,2

I J D - 18

When designing a duct volumetric receiver decisions have to be made on the receiver geometry for a process optimization. As well as in the case of conventional compact heat exchangers (Kays and London, 1964), two of the main factors to be considered are heat transfer capabilities and pressure loss, which will mean fan power consumption. A good receiver design means a balance between these two factors. Since the most relevant effect of good heat tansfer to the working fluid is receiver efficiency, we propose a plot of pressure loss vs. thermal efficiency as a good tool to take

0.4-

e-'0~n

---G--D -, 2 r i

U-

--Z~-D,, 4 mE

U-

- - X - - D " e nan

0.1 -

--O---D,, II i i

0O00

1000

1000

30OO

mOO

1",,00

Fig.5: Solar diffuse with front absorption. Effect of receive] diameter.

328


4. IDEAL SPECULAR SOLAR REFLECTION We have seen that the main reason that spoils receiver efficiency are the high radiation losses that exist when avoiding excessive reflection losses. All of this is strongly associated to the diffuse behavior of duct wall to solar radiation. Indeed, a diffuse reflection is good for infiared radiation because part of the radiation emission goes towards the inside of the duct, allowing for a reabsorption of this energy. But the diffuse behaviour with solar radiation is responsible for the high dominating reflection losses that appear when increasing solar reflectivity with the aim of reducing font wall temperatures. This problem would dissapear with a specular solar reflection. Indeed, specular reflection of solar radiation will strongly increase the duct capability to transport solar radiation towards the inside of the duct. However, if a mnqace is specular for solar radiation, it also will be specular for the infrared longer wavelength radiation, and the benefit of diffuse infrared reflection would be lost, with an increase in radiation losses. We propose to overcome this limitation by using a composed surface consisting of a solar specular coating deposited on a substrate, and a semitansparent coating over it, which allows solar radiation to go throuh, but which is opaque and diffuse to infiared radiation. Such surfaces are familiar to solar concentrating technologies, since they are being developed to irro.lement on heliostats or other solar reflectors. In such applications the function of the semitransparent coating is to protect the specular coating, without bothering at all how it behaves with infiared radiation, because heliostats don't deal with such radiation. With this concept, the top performance would be expected from the perfect specular reflector (r~ = 1), in which all the solar energy incident on the aperture would be transported through the duct without any loss for being absorbed in the cavity at the end of the duct, that behaves almost like a black body if the ratio of its inner surface to the cross duct section is high enough. Collected solar energy would be transported through the duct by infrared radiation and conduction mechanisms, and transfered convectively to the air flowing inside. In such a duct volumetric receiver axial temperature gradients are always positive, whatever the duct wall infi-ared emissivity and conductivity, with the lower wall temperatures located at the tube inlet. Thermal efficiencies are incredibly high, and equivalent emissivities incredibly low. With ideal specular reflection all the selective potential of the duct geometry comes out. The geometry selectivity is so good that the best value for the rest of thermo-physical and thenno-optical properties will be the ones that less spoil geometry selectivity. The lower the thermal conductivity, the better the efficiency, because now a positive axial temperature gradient exists, and therefore, conduction heat transfer will tend to increase inlet wall temperatures with the associated increase in equivalent emissivity. In fact is enough to keep low conduetivities in the inlet duct region. Thermal conductivity is the parameter with the highest effect on this receiver performance, by being the one that most effectively can destroy the selective effect of the non-isothermal cavity. The effect of infrared emissivity on this receiver is less important, and cualitatiely depends on the value of thermal conductivity. With low conductivities, receiver performance improves when increasing infrared emissivity, which is due to the higher infrarred absorption in the inner part of the duct, decreasing the amount of infrared radiation that reaches the front

duct side, and therefore emission losses. However, with high conductivity, temperature in the front duct wall is higher (conduction is the main heat transfer mechanism in this case), and therefore high inflated emisivity increases radiative losses. However, the effect of infrared emissivity is less in'~rtant as conductivity increases. Therefore, we conclude that with the perfect solar specular duct receiver there exists a high potential for efficiency improvement, and if conductivity is low conventional selectivity mechanisms with low e are negative for receiver performance, while if conductivity is high the potential improvement with low e is very small. To illustrate all of this, in Table-3 we present the effect of different design parameters on receiver performance. The base ease is without frontal absorption and given by L/D=l 5 ; D = 2 mm ; e = 0,8 mm ; R ~ = 150 ; k = 2 W/mK ; e = 0,8, and the solar flux is 3,14 MW/m 2 , in order to achieve 1800 K air outlet temperature. We further insist on the high receiver efficiency (lath= 99,4 % at Tgo = 1800 K), showing the high potential for the specular solar reflection-infiared diffuse reflection concept. Like before, when two values appear on a parameter the first one is associated to the inner duct half, while the second one to the outer duet half.

Tg,00

e,q

Base Case

1800

0,994

0,0044

Modific. (%)

V~

D~

De,~l

k = 50 W/mK

-2,89

-4,33

2621

k= 100W/inK

-4,61

-6,98

6078

k=-50-2W/mK

-0,26

-0,10

196

e = 0,2

-0,70

-0,99

142

e = 0,2 - 0,8

-0,09

-0,12

14

e = 0,8 - 0,2

-0,52

-0,62

96

||

, l|

: ||

l|

,i

Table-3: Ideal solar specular reflector.

5. SPECULAR SOLAR REFLECTION When we consider non ideal specular solar reflection, most of the incident solar energy will be absorbed in the inlet duet side because of multiple reflections ocuning to the rays that come at high angles. Therefore, the most important effect of specular reflection is limited to this part of the duet, having less effect on receiver performance if for manufacturing or materials reasons we have to limit the specular behaviour at the inlet region. If the cavity at the end of the duct does have a finite solar reflectivity, as r, from the tube walls is increased, reflection losses will increase, but with L/D high enough these losses can be reduced to very small values. Increasing r, improves solar transport, reducing wall temperature in the inlet region and generally increasing air outlet temperature. However, this depends on ~ a t i o n value: For high irradiations higher r, gives higher receiver efficiencies because radiative losses dominate over reflective losses, but at lower solar irradiations this is otherwise. With r~ above r, = 0,8 there are


positive axial wall temperature gradients in all the duct lenght, and therefore receiver performance is favoured with low wall conductivities. However, for smaller values of solar reflectivity there are negative axial wall temperature gradients, increasing receiver efficiency with high conductivity for the same reason as in the solar diffuse duct. For rs = 0,8 axial wall temperature gradients are almost zero, and therefore axial conductivity has little effect. Infrared emissivity has less effect on receiver performance, which increases slightly if low infiared emisivities are used in the duct inlet region. As the L/D ratio increases, reflection losses and solar transmission decrease, with an increasing equivalent emissivity. All this produces an optimum value of L/D for each solar irradiation and mass flux. For rs = 0,8, ReD=150, k=5 W/ml~ we get (L/D)opt = 15 for a wide range of solar irradiations, but receiver efficiency is almost the same for higher values of L/D. Small effects come from an axial variation of these properties. When solar absorption occurs in the front duct edge, a negative axial wall temperature gradient exists near the tube inlet, which becomes positive luther on if r~ is high enough. The effects of frontal absorption are the same as for the solar diffuse receiver. In Table-4 we quantify the effects of different parameters on the solar specular without frontal absorption (q~/A = 3,6 MW/m:), while in Table-5 we do the same for the solar specular with frontal absorption (q~/A = 2,2 MW/m2),being the reference case given by L/D = 15 ; D = 2 mm ; e = 0,8 mm ; ReD = 150 ; e = ele = 0,2 ; r~ = 0,85 ; k = 50 W/mK ; ale = 0,8. while in Figures 6 and 7 we see the effect of solar reflectivity at different air outlet temperatures (solar ~ a t i o n s ) , for the solar specular without frontal absorption, on receiver efficiency and equivalent emissivity,

329

in the front duct part, and diffuse solar behaviour in the inner duct region, generally reflection losses and equivalent emissivity increase, and therefore receiver efficiency decreases, although if we keep high solar absorptivity in the inner region, the efficiency reduction is not very important. Nevertheless, with high r~, receiver efficiency improves when introducing low solar reflectivities (diffuse or specular) in the duct's inner region because of reducing reflection losses, which are significative in this case. This opens the way to reduce the temperature operation limitations imposed by the solar-specular/infiared-diffuse material on gas outlet temperatures. Still these limitations are rather important. With r,,e=0,8 in the outer duct region and rs,d = 0,2 in the inner duct region, if the maximum operating temperature of the specular material is 900 ~ solar irradiation has to be kept below 1,5 MW/m2 , and air outlet temperatures under 1140 IC The higher the specular solar reflectivity, the less important the material temperature limitation. |_

Tie (K)

.. -

nu

Base Case

nn

1852

0,76

n

Modific. (%) " e = 0,8 ii

.. k = 2 W/mK k=-100W/mK ,. i rs = 0,5

i

DTt~ n

-2,36 1 ! -2,29

~

a

n

TgoOK)

hth

eeq

0,07 -3,96

Base Case

1800

0,901

0,46

Modif. (%)

DTgo

Dhth

Deeq

II

-1,94

-3,02

36,76

k = 2 W/mK

1,36

2,14

-49,02

k=-IO0 W/mK

-0,50

-0,78

14,75

r, = 0,5

-3,19

-4,96

88,84

10,25

1,99

-0,49

45,69

5,78

-35,56

-5,48

-17,11

-21,94

3,21

4,78

73,96

-6,18

- 11,22

1,15

-0,30

-0,34

l

-27,47 9

ale = 0,95

3,89 l

9

6,12 l

ele = 0,9

-5,07 9

ele = 0,1

0,94 9

kle=100W/mK

0,024 l

I"

9

0,037 l

kle= 2 W/mK

ll

9

1,46

l

ii

l

-7,84

l

.. ii

ii

-18,06

-0,12

1

,|

e=0,8

3,54

l

-3,09

l

l

ale = 0,2

ii

||

2,26

31,33

l

9

l

DTle

n

l

rs = 0,95

i

Dem

' -3,56

-2,55

1902

m

-3,66

0,05 l

0,81 a

Dhth

m

l |_

nn

-1,25 9

l

-1,95 |

13,86

16,27 9

n

nI

Table-5: Solar specular with front absorption.

r~ = 0,95 2,50 3,93 -54,05 n' Table-4: Solar specular without front absorption.

6. STABILITY ISSUES As it may be seen in these tables without solar frontal absorption, the real solar specular duct volumetric receiver has very high efficiencies in spite of a considerable incease in equivalent emissivity. However, frontal absorption reduces a lot the receiver efficiency, and therefore will be the main limiting factor to benefit from this concept. Tube inlet edges should be sharpened with the minimum manufacturable and structurally admissible slope in order to tend towards the no front absorption limit. When exploring the posibilities of combining specular reflection

In the last years analytical (Kribus, Ries and Spirkl, 1995) and numerical (Pitz-Paal and Hoffschmidt et al, 1997) studies have claimed that duct volumetric receivers have inherent limitations on the maximum solar flux they can manage due to unstable behaviour, although this has not been experimentally confirmed. The hot points that have been observed experimetally are likely to be due to the synergic effects that anise when exposing parallel receiver channels to non-uniform solar irradiation, and problably have nothing to do with unstabilities.


330

First of all we have to point out that the unstabilities which have been mentioned are only relative, and don't mean that one couldn't work with such receivers without any problem Indeed, the unstable behaviour depends on the characteristic pressure lossmass flow curve of both the fan and the receiver, and the condition for such unstability to affect the system is that in the working point the fan curve slope is higher than the receiver curve slope. One speaks about unstable branch in the receiver curve when the slope is negative, and about unstable branch on a fan curve when the slope is positive. Most fans do have an unstable branch and they are used without any unstable behaviour, so one shouldn't panic if a receiver curve would come out to have an unstable branch in its characteristic curve. A bit of engineering care (mostly in the selection and operation of the fan) should be enough to guarantee a stable behaviour of such receiver in all working conditions. T~,, 643 K ; I ~

o" le0 ;

e-. 0~ ;

r,o,, 0,2 k - E0 W/InK UD,, 16

US

#-

D - - 2 mm e,, U mm

us --G-- r m , , U

U

- - ~ - r , . - US -o--r,.- U - x - t u - 0,2

0,rS 0,7 800

1800

1000

2000

21500

Tp(K)

Figure~: Solar specular without front absorption. SolaJ -effectivity and solar flux effects. T#,,E43K;Re

a,,150;

e,,0,2;

r,o,,0,2

1.1. k , , 80 W/inK

-

/

I.A) ,, 18 D , , 2 mm

11,7

-

o m 0mS IIIIll

r u , , 0,8

..-b--r u , , 0,118 -•

11,1

,~"

0,2

D

tO00

11110

2000

2500

1"1,00 Rgure-7: Solar specular without front absorption. Effect ol ~olar reflectivity on equivalent emissivity. ,

However, the hypothesis on which are based the models that were used to predict such unstabilities have important simplifications of the heat and flow processes that occur in a duct volumetric receiver. As the unstability has its origin on the dynamical viscosity dependence on temperature, and the hypothesis in this models may affect the temperature distribution inside the duct, one can not conclude that such unstable branches exist without a more detailed analysis. On the other hand, in ( P i t z -

Paal and Hoffschmidt et al, 1997) they only find the unstable branch in the duct receiver after applying the model to several volumetric receivers, being this the one for which the model is less appropriate, and justified the fact of experimentally not finding the unstable branch by the radial conduction effects. With our model, we didn't find any unstable behaviour even at much higher solar fluxes, and without introducing radial conduction effects. More specifically, the model simplifications in (Kribus, Ries and Spirld, 1995) that could affect the stability are: - No inertial effects in momentum equation (asymptotic limit for Reynods ~ 0). - Temperature effects only on density and viscosity. - No radiative exchange calculation inside the receiver, and therefore no reabsorption mechanisms for infrared radiation inside the receiver matix. - Uniform solar absorption in the receive matrix, while most of the absorption is in the font part, even for a solar specular duct. - Infinite convection coefficient. - No conduction in receiver matrix. The numerical receiver model used in (Pitz-Paal and Hoffschmidt et al., 1997) is more sophisticated. It accounts for solar absorption, radiative, convective and conductive heat transfer in a one channel model, and then goes further by combining these results to evaluate the effect of non uniform solar irradiation and radial conductivity. However, it is still based on some simplifications that are not appropriate for the analysis of the duct volumetric receiver. - lnfi'ared radiative exchange modeled with a simplified version of participating media with isotropic scatering and specular reflection at the duct end, which is good, and in fact one of the few things one can do, to model porous media, but not a duct receiver. - No spectral dependence ofthenno-optical properties. - It doesn't include the entrance length effects on heat transfer, friction and fluid acceleration. These effects are equivalent to the inertial effects that in this same reference are identified as stabilizing factors. - It doesn't include temperature dependency of other fluid properties than density and viscosity. - It doesn't include inertial effects in the momentum equation except for the shape pressure loss. - It also doesn't consider the possibility of secondary flows and mixed convection transfer to occur, although with small duct diameters Rayleigh numbers won't be high enough for it to happen at the solar ~ a t i o n they used. We applied our models to evaluate pressure losses under different circumstances, and never found an unstable branch in the receiver characteristic curve, in spite of calculating with much higher solar fluxes (up to 10 MW/m2) than in these references. However, most receiver characteristic curves show an inflection point that reminds of the mechanism that seemed to generate the unstable branch in the other models. There are several factors that increase the tendency of this inflection point towards unstability, but without ever reaching it. We observed that the tendency towards unstability grows as air inlet temperature is lowered, L/D is increased, or duct diameter is reduce& The first of this factors affects through the higher variation of air viscosity through the duct, and the two last ones affect because the less influence of the entry lengh effects. In Figure-8 we present some results of receiver


characteristic curves for a diffuse duct receiver with fontal absorption working at different solar fluxes. The pressure loss is presented on a non dimensionalized way to better appreciate how tendency towards unstability grows as higher solar irradiations are used. But still at 2,5 MW/m2 the inflection point is far from creating an unstablre branch. In Figure-9 we present the results for a specular receiver with front absorption and different L/D ratios working at 2 MW/m 2. k - k ssuS0'filnl(;

om osou0,2; itse-0,8 ;

/

100

00

r u -0B

-x--qo/A 9U IdW/m t --n--qoJ~- 1 MVlln " -u--q,~-U~ 9

(SL)

00 i

L/D~6

Id

Do|l e"Um

6

21)

331

conventional e/a selectivity doesn't have a very important effect. For both specular or diffuse solar reflection, frontal solar absorption reduces a lot receiver efficiency, diluting other selectivity mechanisms. Sharpening of the leading edge with such a small angle as structurally posible, and conventional e/a selectivity stategies in the frontal edge have to be used to get higher receiver efficiencies. A plot of receiver presure loss vs. receiver efficiency is introduced as a useful tool to take design decisions as appropriate values o fL/D and D. When using a receiver model incorporating the effects that were left out in former studies, we didn't find any unstable branch on the receiver characteristic curve up till 10 MW/m 2 solar irradiation, which is much higher of what is needed to work with modem gas turbines. Therefore we conclude that there is no inherent limitation on the maximum solar irradiation a volumetric receiver can manage without presenting possibilities of unstable beheaviour. NOMENCLATURE

0

2

4

8

8

ae ~ t) Figure..8: Solar diffuse with frontal :urves at different solar fluxes. q,/h-2~ lOO

t ; T 6,,648K;

absorption. Characteristic

ro-0,O;k,,Ido,,O01WmK

-

-o--Lq) - S0 --x--LA~ - 80 ~LJD

~)

-" 18

--I~--IJD - 8 M.

D-'Rmm

I

o " O,ll mm

/

B

e-

el,-U

11)

0

2

4

8

a : Duct wall solar absorptivity. Leading edge solar absorptivity. Dl~t: Total air pressure loss through receiver D : Duct diameter. e: Duct wall thikness. e : Infrarred wall emissivity. elo : Leading edge infrared emissivity. e~q :Equivalent emissivity. hth : Receiver thenml efficiency. k: Thermal wall conductivity. kle : Thermal wall conductivity on the between leading edge and first wall node. L : Duct length. qs/A : Solar irradiation on receiver aperture. rs: Solar reflectivity. r~e: Specular solar reflectivity. rs,d" Difi~e solar reflectivity. Tgi : Inlet air temperature. "['go : Outlet air temperature. ale:

8

Figure-9: Solar specular with front absorption. Effect of L/E "atio on characteristic receiver curve.

7. CONCLUSIONS In this paper we have explored the potential of different selectivity strategies to improve solar receiver efficiencies. For solar diffuse duct receivers, wall conductivity plays an in'gmrtant role and conventional e/a selecivity works although can't bring a very in~rtant efficiency improvement. Important effects are restricted to the entrance region of the duet receiver. A very high potential for efficiency improvement is associated to the new concept of solar-specular/infiared-diffuse wall behaviour. Again the main effect of thennophysical properties is located in the duct entrance region. Small wall conductivities also have a significant effect on receiver efficiency improvement, but

REFERENCES Becker M., Cordes S., B6hemer M. (1992) The Development of Volumetric Solar Receivers. In Proceedings of the 6th Int. Syrup.

on Solar Thermal Concentrating Technologies, MAdrid, Spain, pp.945-952 B6hemer M., Chaza C. (1991). The Ceramic Foil Volumetric Receiver. Solar Energy Materials 24, pp.182-191 Chavez J.M., Chaza C. (1991). Testing of a Ceramic Absorber for a Volumetric Air Receiver. Solar Energy Materials 24, pp. 172-

181 Holman J.P. (1997). Heat Transfer. 8th ed., McGraw-Hill Flamant G., Meningault T., Olade G. (1987). Nouveau Dispositif d'Abnsorption S61ective de l~Energie Solaire Concentr6e par des Lits de Particules. C.R.Acad.Sci.Paris t 304, SbT"e 11, 3, pp.689-

332


694. KalTli J., Rubin IL, Kribus A., Doron P., Sagie D. (1996). Test Results with the Direct-Irrdaiated Annular Pressurized Receiver. In Proceedings of the 8th Int. Symp. on Solar Thermal Concentrating Technologies, 6-11 October, Cologne, Germany, M.Becker, M.BiJhmer (Eds), pp.607-621, C.F. Mailer Kays W.M., London A.L. (1964). Compact Heat Exchangers. 2nd ed., McGHraw-Hill Kribus A., Ries H., Spire W. (1995). Inherent Limitations of Volumetric Solar Receivers. Solar Engineering-VoL1, ASME 1995, pp.649-655 Lin S.H., Sparrow E.M. (1965). Radiant Interchange Among Curved Specularly Reflecting Surfaces-Application to Cylindrical and Conical Cavities. Journal of Heat Transfer, May 1995, pp.299-307 Meningault T., Flamant G., Rivoire B. (1991). Advanced HighTemperature Two Slab Selective Volumetric Receiver. Solar Energy Materials 24, pp 192-203 Pitz-Paal R., Morhenne J., Fiebig M. (1991). A New Concept of a Selective Solar Receiver for High Temperature Applications. Solar Energy Materials 24, pp.293-306 Pitz-Paal IL, Fiebig M. (1992). First Experimental Results from the Test of a Selective Volumetric Air Receiver. In Proceedings of the 6th Int. Syrup. on SOlar Thermal Concentrating Technologies, Madrid, Spain, pp.277-289 Pitz-Paal IL, Hoffschmidt B., B6hmer M., Becker M. (1997). Experimental and Numerical Evaluation of the Performance and Flow Stability of Different Types of Open Volumetric Absorbers Under Non-Homogeneous Irradiation. Solar Energy, vol.60, Nos. 3/4, pp.135-150. Rabl A.. (1977). Radiation Transfer Through Specular Passages-A Simple Aproximafion. lnt. J.Heat Mass Transfer, vol.20, pp.323330 Shah ILK., London A.L. (1978). Laminar Flow Forced Convection. Acadmic Press Siegel IL, Howell J.IL (1992). Thermal Radiation Heat Transfer. 3rd ed., Taylor & Francis Spirkl W., Ries H., Kribus A. (1997). Performance of Surface and Volumetric Solar Thermal Absorbers. Journal of Solar Energy Engineering, May 1997, VoL119,pp.152-154 Variot B., Meningault T., Flamant G. (1992). Modelling and Optimization of a Two-Slabs Selective Volumetric Solar Receiver. In Proceedings of the 6th Int. Syrup. on Solar Thermal Concentrating Technologies, Madrid, Spain, pp.325-345


333

A PARABOLIC DISH CONCENTRATOR FROM A TELECOMUNICATION ANTENNA: OPTICAL AND THERMAL STUDY OF THE RECEIVER

Claudio A. Estrada Centro de Investigaci6n en Energia, UNAM, A.P. 34, Temixco, Morelos, 62580, M6xico. Tel. (052 73) 25 00 48, 25 00 44, e-mail: [email protected], [email protected]. Rubtn Dorantes Departamento de Energia, UAM-A, Av. San Pablo No. 180, Col. Reynosa, Tamaulipas, 02200 Mtxico, D.F.

Eduardo Rinc6n Facultad de Ingenieria, UAEM, Cerro de Coatepec s/n, Toluca, M6xico.

Abstract - In Mtxico, Telcom, a big telecommunication company has been removing, due to the installation of new systems, hundreds of parabolic antennas made by aluminum and other light materials without further future use. Now, it is being investigated the adaptation of those parabolic antennas as parabolic dishes (see fig. 1) to be used as solar concentrators for thermal conversion of solar energy. This paper, which is a continuation of Estrada et al (1), presents theoretical advances of the research focused on the determination of the optimal shape of the receiver and the thermal losses associated. It is intended to have low cost, parabolic dishes with high surface's reflectance and high receiver's efficiencies. The antenna's dimensions are 3.32 m in diameter, with focal length of 0.83 m, which gives an aperture angle of 90*. For the theoretical analysis, a software package for facilitated optical analysis of 3-D distributed solar energy concentrators called CIRCE2 (2) was used. The flux distributions for a cylindrical and a conical shape receivers were analyzed and the error function of the concentrator was determined giving a value of a = 2 mrad. The peak local concentration for the cylindrical receiver was found to be 2040 suns and the geometrical concentration was 860 suns. With these concentrations it is possibly to think in temperatm'es as high as 1000 K or even greater. So far, the results indicates that a simplified and cheap systems can be built with those antennas to allow the conversion of solar to thermal energy for high working temperatures.

1. INTRODUCTION Some researchers in Mexico are interested in finding a new use for several parabolic antennas that have been removed by a telecommunication company in Mtxico named Telcom, due to the installation of new systems. Aluminum and other light materials make the antennas. The dimensions of each one are 3.32 m in diameter (an aperture area of 8.66 m:) with a focal length of 0.83 m, which gives an aperture angle of 90 ~ and with a total weight of 210 kg. The Japanese NEC Company made those antennas in 1964 and it is estimated that there are around 300 of those antennas in Mexico. The Autonomous Metropolitan University got free ten antennas and all of them are in very good conditions. Now, it is being investigated the adaptation of those parabolic antennas as parabolic dishes to be used as solar concentrators for thermal conversion of solar energy. The goal is to transform the antennas into low cost, parabolic dishes for solar thermal conversion with high surface reflectance and efficient receivers. In a previous work, Jimtnez et al (1997) showed that the antenna aluminum surface can be polished until get an average reflectance of 0.75 between 400 a 700 nm (visible region) and 0.92 for the infrared region until 3 000 rim. The cost of preparation and polish for each antenna's surface was estimated to be 200 USD, including materials and labor. To have a good surface reflectance is just the first step in the adaptation of the parabolic antenna as a solar concentrator. Now, it is necessary to quantify the deviation of the surface from a perfect paraboloid that will determine the form and dimensions of

the optimal receptor needed to capture the concentrated solar energy. This paper presents advances in the experimental and theoretical studies to determine the focal zone's dimensions and the optimal shape of the receiver and the thermal losses associated. 2. DISH EXPERIMENTAL CHARATERIZATION The ideal focal zone of the antenna as a solar concentrator is an ellipsoid, which has major and minor axes of 1.54 cm and 0.77 cm respectively. These values come from the solar rim angle (16' = 4.653 mrad) and the antenna's dimensions (rim angle of 90*). Due to manufacturer and polish processes the actual antenna's focal zone is greater and different to that ellipsoid. To determine experimentally the deviation from the ideal focal zone a laser scanning technique was used, see figure 1. Estrada et al (1998), showed that using a 100 cm long rod of 1 cm in diameter located at the principal axis of he paraboloid and the laser scanning technique, it was possible to determine the height of the focal zone around 6.3 cm. Additional to the rod, two more bodies were located as receivers at the focal zone of the concentrator and the laser scanning process was applied for each one. One was a cylinder with a diameter of 5 cm and a height of l0 cm. Another was a cone with a diameter of 10 cm and a height of 6 cm. Fig. 2 shows the three receivers and the laser beams hitting on them.

334


Fig. 2. Concentrator receivers with the reflected beams hitting on them. a) long rod, b) cylinder and c) cone.

Fig. 1: Parabolic dish in vertical position

For the two other bodies, it was found that 97% of the reflected beams hit the receivers. The geometrical concentration for the cylinder was found to be 860 suns, while for the cone it was 700

Concentrator imperfections such as slope errors, surface roughness, random facet misalignment, etc., have impact on the actual normal at the concentrator surface. These imperfections give rise to error distributions, which assign a probability to the chance that the actual surface normal will take a given direction. Concentrator imperfections can usually be model with 2-D elliptic normal error distributions. If it is assumed that the errors are truly random without a predisposition in any direction, it is appropriate to model them with 1-D circular-normal distributions. These distributions, special case of the elliptic-normal distributions, are used for most error types. The equation representing this circularsymmetric Gausssian distribution error is 1 E1D(r) = 2 m 2

exp[

r2 - 2-'~" I

(1)

SUNS.

It is important to note that the optimal rim angle for a paraboloid concentrator is 45* and its receiver is normally design as a cavity with a fiat aperture perpendicular to the paraboloid principal axis. However, the antenna dish has a rim angle of 90 ~ implying that an external receiver should be uses to capture the concentrated solar energy. This explains the shapes of the receivers used. 3. DISH T H E O R E T I C A L CHARATERIZATION The theoretical analysis of the parabolic concentrator was done using a software package to facilitate optical analysis of 3-D distributed solar energy concentrators developed by Romero [2] and named CIRCE2 (Convolution of Incident Radiation with Concentrator Errors). This software allows a user to efficiently build and analyze a variety of point-focus solar concentrator. CIRCE2 was set up for each of the three receivers along with the paraboloid dish parameters. For each case, the program was run and the flux distributions along the exterior receiver's surfaces were obtained. Because the axisymmetrical configuration of the receivers the flux distributions are just a fimction of the Z coordinate (see fig. 2).

where r is the local radial coordinate where the solar ray is hitting the surface and 6 is the standard deviation or dispersion parameter associated to the concentrator. Eq. 1 also called the error function, characterizes the optical errors of the solar concentrator and the parameter ~ defines how big is the deformation of the reflected sun cone due to the surface's errors. For an ideal concentrator ~ is zero and thus Em is from Eq. 1 equals to zero. To run CIRCE2 it is necessary to define the value of a. Also, Estrada et al (1998), showed that using CIRCE2, it was possible to determine a ~ value of 2 mrad for the antenna which produces a width distribution of 6.1 cm being in agreement with the 6.3 cm found experimentally. Therefore, it is established that the dispersion parameter a associated with the antenna dish is 2 mrad. Fixing G to the value of 2 mrad, CIRCE2 was set up for the cylinder and the cone receivers. Fig. 3 shows the flux distributions as a function of the axial Z-coordinate. The maximum local concentration obtained for the cylinder and the cone were 2040 and 907 suns respectively. In both cases, the distributions have a maximum at the intermediate zone of the receiver's surface loaded to the positive direction of Z. It is clear from this figure that the cylindrical receiver is the one that gave the higher concentration and the one that should be recommended as a system's receiver.


Q = mCp

AT At

335

(3)

where Q is the energy incident into the cylinder, m and Cp. are the mass and the heat capacity of the cylinder, AT is the difference between the average temperature of the solid and its initial temperature, and At is the increment in time.

Fig. 3. Cylinder and cone flux distributions as a function the axial Z-coordinate for o = 2 mrad.

4. T H E R M A L CHARACTERIZATION For the thermal characterization, an iron steel cylinder 5 cm in diameter and 7 cm long was located at the concentrator "s focal zone. Six thermocouples type K were located in the cylinder. Three of them .3 cm from the cylinder's surface, and the other three 2.5 cm from the surface, that is, at the cylinder's center. Figure 4 shows the thermocouple positions and named them. Fig.5 Temperature distributions for thermocouples T2, T3 and T4, for several times.

For the experiment shown in Fig 4. m = 1.66 kg, Cp= 473 J/kg, Tp = 98 C (Tp is the weighted average temperatures of the solid), To = 35 C and At = 10 sec., then Q = 4.946 kW. The energy incident at the concentrator was 822 W/m 2, and the concentrator area is 8.6 m 2, thus the input energy was 7.069 kW. This gives a thermal efficiency of the concentrator of 0.7 % for this conditions. This could be understood as the upper limit for the thermal efficiency of the concentrator, for a receiver working a much higher temperature, the thermal efficiency is expected to decrees. Fig. 4. Cylindrical receiver and thermocouple positions.

The concentrator was aligned with the solar rays and the concentrated solar radiation was allowed to strike the cylindrical receiver. Fig 4 shows the temperature distributions for thermocouples T2, T3 and T4, for several times. It is clear comparing Fig. 3 and Fig. 4., that the flux distribution is in correspondence with the rise in temperature at the first seconds of the transient processes. If we assume that at the first seconds of the solar exposition, all the losses by convection and radiation are neglected, then all the energy incident to the receiver will increase the internal energy of the receiver's mass, that is, the energy balance gives Q=mCp

or

dT dt

(2)

5. CONCLUSIONS The laser scanning technique as presented in this paper was adequate and it allowed finding the height of the focal zone to be around 6.3 cm. The error function of the concentrator was determined giving a value for the dispersion parameter (6) of 2 mrad. The flux distributions for a cylindrical and a conical shape receivers were analyzed and it was found that the cylindrical receiver is the one that gave the higher concentration and thus the one that should be recommended as a receiver for the dish concentrator. The pick local concentration for the cylindrical receiver was found to be 2040 sun and the geometrical concentration was 860 suns. And experiment was conducted to show that a upper efficiency of 70% can be reached by this concentrator. It is possibly to think in temperatures as high as 500 C or even greater, even though the efficiency can be low. So far, the results indicates that a simple and cheap systems can be built with those antennas to allow the conversion of solar radiation to thermal energy for medium working temperatures. The next step is define a simple and cheap structure to support the

336


antenna with a control system, as well as the heat exchanger for the receiver to be used in a useful thermal processes. A cost effective study should be made.

6. REFERENCES Jim6nez, M.R., F. Hemfindez and R. Dorantes, Construcci6n de un Concentrador Solar a Partir de una Antena Parab61ica., Proceedings of the 21 st National Week of Solar Energy, Mexican Solar Energy Society, 1997. Romero, V.J., Cice2/Dekgen2: A Software Package for Facilitated Optical Analysis of 3-D Distributed Solar Energy Concentrators. Sandia National Laboratories, SAND91-2238, 1994. Estrada, C.A., R. Dorantes, F. Hem~indez, M. Martin and E. Rinc6n, 1998. A Parabolic Dish Concentrator From a Telecommunication Antenna. Proceedings of the 1998 Annual Conference, American Solar Energy Society. Vol 1, pags.267-270. Editado por R. Campbell-Howe, T Cort6z and B. WilkinsCrowder ASES, USA.


337

EFFICIENCY IMPROVEMENT OF PARABOLIC TROUGH COLLECTORS BY MEANS OF ADDITIONAL END REFLECTORS

TH. FEND, J. LEON l, P. BINNER, R. KEMME 1, K.-J. RIFFELMANN AND R. PITZ-PAAL Solare Energietechnik, Deutsches Zentmm ftir Luft- und Raumfahrt e.V. (DLR), Linder H6he, D-51147 K61n, Germany, Phone: +49 2203 601 2101, Fax: +49 2203 66900, E-Mail: [email protected] 1plataforma Solar de Almeria, Carretera de Senes s/n, Tabemas (Almeria, Spain) An additional collector element designed for the attachment to existing north south aligned parabolic trough collectors was tested on the Spanish Test-Center Plataforma Solar de Almeria (PSA). It may be installed at the north end of a collector module and reduces end losses occurring during non-normal incident solar radiation. Thermal efficiency tests were performed by nmning a test loop including a 48 m parabolic trough collector of the LS-3 type. It could be shown by calculations and experiments, that with the use of the additional reflector collector efficiency increases significantly. From the data acquired during the experiments an annual efficiency could be calculated, which is 42.6 % compared to 39.9 % without the additional reflector. Abstract.

1. I N T R O D U C T I O N incident angle (Dudley et.al., 1994). One consequence, which has an influence on these factors, is the occurrence of end losses. They are caused by radiation penetrating through the aperture area at the north end of the collector (for the northern hemisphere), which cannot be concentrated on the absorber tube. Additionally, north-south oriented troughs with a one axis tracking system loose radiation penetrating through the north end of the collector. As has been described in more detail in previous publications (Binner, 1997; Fend et. al., 1998b), an additional end reflector at the north end of the collector may be used to eliminate both kinds of these losses (Figure 1). Calculations show that flux on the absorber tube increases by nearly a factor of two in a zone near the additional mirror.

Parabolic trough collectors concentrate direct solar radiation on a line focus, where absorber tubes convert radiation into heat, which is transported by a liquid medium, most frequently oil or water. The generated thermal energy may be used for process heat or electricity generation (Becker and Gupta, 1995). For economic reasons parabolic trough systems follow the sun by means of one axis tracking systems, which leads to non-zero angles of incidence of the solar radiation in most times of the day. When the troughs are aligned in a North-South direction, the annual average of this angle of incidence is minimized. However, the angle impacts the optical efficiency of the collector. Geometrical and optical losses are usually characterized by so called IAM-factors (Incident _Angle Modifier) as functions of the

/

,, , , / / / /

,

9

,,/,,/,,/

,

~X

/

/

receiver J r J"" i

/

/

concentrator

I end reflector

Figure 1: Reflection at the front side of the collector with end reflector


338

Measurements carried out on a parabolic trough test-loop with an aperture of 2 m2 and an 0.6 m2 end-reflector confirmed these calculations (Fend et. al., 1998). The biggest plant using parabolic trough technology is in operation in the SEGS-plants at Kramer Junction, California, since 1987 (Cohen et.al., 1998). Since this time the plant has produced an average annual electricity output of approximately 400 GWh. 2. DESCRIPTION COMPONENTS

OF

THE

TEST

LOOP

Israeli company LUZ-Intemational Limited and is the third generation of this collector technology. It consists of parabolic collector modules of 48 m length and an aperture width of 5.76 m. 112 thick glass mirror facets produced by Flachglas AG, Germany concentrate direct solar radiation on a heat collecting element (HCE) consisting of an evacuated glass envelope with an inner steel tube, which is coated with a selective black surface (see Table 1 for further technical data). The HCE is designed for Temperatures of up to 400~ Usually oil is used as a heat transfer medium. One module of this collector type was erected at the Plataforma Solar de Almeria (PSA) in 1997 for testing purposes (Figure 2).

2.1 CollectorThe so-called LS-3 collector used at the SEGS plants mentioned above was developped by the US-

Figure 2: Photograph of the 50 m LS-3 test-loop on the PSA

Table 1:LS-3 Collector Technical Data Aperture Width Total Aperture Focuslength Total Collector Length Number of Mirror Facets

[m] [m [In] [m]

Dimensions of Mirror Facets

[m]

Number of Receiver Elements Length of Receiver Element Outer Absorber Tube Diameter Absorber Tube Wall Thickness Glass Envelope Diameter Glass Envelope Wall Thickness Thermo Fluid Maximum Outlet Temperature Maximum optical efficiency Maximum Collector Efficiency @400 ~

[in] [mini [mini [mini [mm] [~ [%] [%]

5,76 272,5 1,71 48 112 1,70xl,63 1,70xl,50 12 4 70 2,25 115 3 400 78 68


2.2 Test Loop The structure of the test-loop is drawn schematically in Figure 3. An electrical heater is integrated to enable a quick warm up to working temperatures of up to 400~ As a heat transfer medium Syltherm 800 is used, which is

339

stable at temperatures of up to 400~ The pump enables maximum volume flows of 10 m3/h. Volume flow, collector position, heater and cooler can be controlled via Computer.

Figure 3: Schematic drawing of the PSA's LS-3 test-loop

2.3 Measurement and Data Acquisition Temperature of the thermofluid is measured at the collector inlet and outlet as well as in the middle of the collector by thermocouples. A Volume flow sensor is integrated behind the pump. System pressure is measured before the pump. Signals from these sensors as well as Direct Solar Radiation (DNI) and some more weather data are monitored by a Data Acquisition System (DAS), which is connected to a PC. Data from the DAS is acquired every second, 5 minute average values are generated and stored.

2.4 End-Reflector For the LS-3 test-collector described above a 6 m2 end reflector was designed and installed in April 1998 (Figure

4). Special attention was paid on costs, lightweight and simplicity of installation. The collector consists of 8 facets of high specular aluminium coil material with average solar reflectance values of 88-90 %. Durability and optical performance properties have been investigated in previous investigations (Fend et.al., 1998a). The facets were glued on aluminum frames, which were connected to steel supports by screws allowing a precise adjustment with respect to the geometry of the concentrator and the absorber. Each facet has been adjusted separately with a laser beam simulating the sun's radiation from different incident angles. The steel supports were easily connected to the LS-3 by screws without dismounting any part of the framework of the existing collector (Figure 3).

Figure 4: Front and backside view of the end reflector for the LS-3 test loop at the PSA


340

4. RESULTS AND DISCUSSION Efficiency tests have been performed in February, April and July 1998. Due to accuracy problems of the tracking unit and some misalignment of the LS-3 mirror facets, the optimum possible solar to thermal efficiency of more than 60% was not reached. Efficiency rl of the collector module was calculated from volume-flow 17, density p , specific heat c t, and temperature data of the heat transfer fluid as well as from Direct Normal Incident Solar Radiation data (I) using

q

~

#. p.c

A denotes the aperture area of the collector. Five days of reference testing were followed by 3 test days with endreflector. At each day incident angles from about 40 ~ in the morning to 0 ~ at solar noon were observed. As an example, measured data of one day reference testing are shown in Figure 5. After a phase of additional electrical pre-heating in the morning, an inlet temperature of 250 ~ C was adjusted. Maximum efficiency of this day (44%) was reached at 2.00 p.rm, 30 minutes aider solar noon. Due to geometrical losses efficiency in the morning hours is markedly lower.

"(L., -T,n) I.A

Figure 5: Efficiency, Volume Flow and Temperature data measured at the LS-3 test-loop at the PSA on February 19, 1998


-

1.2-

~

~

~

~

.

.

.

.

.

! end reflector

,

..u

~. e~

.

--

~i

t

.

...............

*' 0.4 Q >

341

,m

t~ 0

m,,

~

'

no lend reflector

Im

0 0

10

20

30

40

Incident angle

Figure 6: Relative Efficiency as a function of incident angle of the LS-3 Collector at the DISS-Reference loop of the PSA comparing the performance of an additional end-reflector with the original configuration To characterize the optical properties of a parabolic trough collector, solar to thermal efficiency rl is often replaced by relative efficiency rl/rl0 with 110 denoting collector efficiency for 0 = 0 ~ incident angle of the solar radiation. Relative efficiency values rl/rl0 are shown in Figure 6 as a function of the incident angle. Each data point was taken from the above mentioned 8 testing periods. In order to generate the data points in Figure 6, 20 min average values were taken from the data of each day, which was monitored in 5 min intervals. Finally the data points were fitted with second degree polynomial functions. The positive effect of the end-reflector is significant. An efficiency increase of up to 20% can be observed. 5. CONCLUSIONS A prototype of an end reflector could be tested successfully on the LS-3 test loop at the PSA. Thermal tests have shown, that collector efficiency can be increased significantly. Furthermore it could be shown, that high specular aluminium coil material can be successfully applied in parabolic trough technology. To calculate a possible annual efficiency of a collector of the LS-3 type, the data shown in Figure 6 were applied on a north south collector located 35 ~ northern latitude. An annual efficiency of 42.6 % was estimated compared to 39.9 % without an additional end-reflector. However, tests were performed on PSA's LS-3 test loop with an overall length of 50 m. Commercial plants usually consist of longer troughs. Assuming an annual DNI of 1800 KWh/m 2 the end-reflector would produce an annual yield of 13400 KWh. Costs of the end-reflector have been estimated to approximately 1500 Euro.

References Becket, M., Gupta, B. (1995) Solar Energy Concentrating Systems, p25, p40, C.F. Miiller, Heidelberg Binner, P. (1997) Wirkungsgradsteigerung von Parabolrinnenkonzentratoren durch Zusatzreflektoren, DLR, IB 375-97/01 Dudley, V.E., Kolb, G.J., Sloan, M., Keamey, D. (1994) SEGS LS-2 Solar Collector, Sand94-1884 Duffle, J.A. Beckmann, W.A. (1991) Solar Engineering of Thermal Processes, John Wiley & Sons, Inc., New York, p 22 Fend, Th., Jorgensen, G., Bthmer, M., Kr/imer, T., Rietbrock, P.M. (1998a) First Surface Aluminium Mirrors: an Assessment for Solar Outdoor Applications, Proc. Eurosun '98, Sept. 14-17, Portoroz, Slowenia Fend, Th., Le6n, J., Bthmer, M., Binner, P., Deidewig, F., Kemme, R. (1998b) Development and Test of an End Reflector for Parabolic Trough Collectors, 9th Int. Symp. - Solar PACES - Solar Thermal Concentrating Technologies, 22-26 June 1998, Odeillo - Font-Romeu Geyer, M., Hollfinder, A., Aringhoff, R., Nava, P. (1998) H/ilfte des weltweit produzierten Solarstroms. Sonnenenergie 3, 33-37


342

EXPERIMENTAL PERFORMANCE OF A PV V-TROUGH SYSTEM N. Fraidenraich and E. M. de S. Barbosa

Research Group on Alternative Sources of Energy- FAE, Universidade Federal de Pernambuco- UFPE. Av. Prof. Luiz Freire 1000 - Cidade Universit~a, 50 740 540 Recife-PE/Brasil Phone / Fax: +55 081 271 (8252) / (8250) e-mail: [email protected]

Abstract - We describe the design procedure and the experimental facility built to asses the technical viability of a Vtrough tracking PV system. Experimental results for radiation collected at the cavities' aperture and absorber regions are presented. The optical properties of the cavity are described theoretically and compared with experimental results as well as the output electric power when the PV system is connected to a bank of batteries. Both, optical and electrical estimations have a good agreement with experiment. Thus, the simple theoretical tools used to analyze the system's performance can be reliably used for design and evaluation purposes. Experimental results, represented as a function of diffuse to hemispherical radiation ratio, yield a set of empirical correlations which have been used to simulate the system behavior along the year. Results of simulation show that the annual benefit in electric energy due to tracking is equal to 26 % and the increase of electric energy output for the V-trough system is 72%, both with respect to an horizontal fixed collector.

1. INTRODUCTION

to determine the cavity parameters (Fraidenraich, 1998).

Photovoltaic systems combined with V-trough cavities and N-S solar tracking c a n have their energy output significantly increased when compared to fixed fiat plate systems. If manufacturing costs of V-trough concentrators and tracking mechanisms are kept low, the cost-benefit relation can be rather favorable. Previous results, obtained with a small prototype of photovoltaic V-trough tracking generator, working in the local sunny climate, shows that the energy output almost doubles as compared to fixed fiat plate PV systems (we verified an average increase of 52%, due to concentration, and 30 % increase due to solar tracking) (Fraidenraieh and Krenzinger, 1997). Several works have been published analyzing the benefits of PV systems coupled to Vtrough cavities. The performance of the largest photovoltaic system with tracking and mirror boosters (Carrisa Plains, CA) is described by Berman and Mitchell, (1989). The potential of this technology for different climates has been analyzed by Nann, (1990) and the economic potential for European climate, for systems with passive tracking, has been studied by Klotz (1995). The objective of this work is to evaluate the performance of a PV system with tracking and concentration. Experimental results, represented as a function of diffuse to hemispherical radiation ratio, yield a set of empirical correlations used to simulate the system behavior along the year. We describe, at first, the design procedure of the V-trough cavity. Immediately, we present results for the collection of solar radiation at the cavity absorber and aperture as compared with hemispherical solar radiation. We analyze then the optical efficiency of the cavity, the module's temperature when working with and without concentration, the electric output of the PV generator and finally we present a simulation of the monthly and annual performance of the system.

Design criteria, satisfied by the concentrator cavities are: (a) Light distribution on module's surface is uniform Co) Heat at the absorber region is dissipated by passive means (c) Small deviations of tracker-sun alignment are allowed, still satisfying uniformity of illumination on absorber region. The concentration ratio (C) and vertex angle ( 9 which minimize the cost of energy are chosen as the best option in terms of cavity design. For the local climate (Recife, latitude 8,05 South, Kh=0.55) the geometric parameters of the cavity, optimized with that procedure, are C = 2.2 and 9 = 30 ~ The optical behavior of the cavity is properly described by the angular acceptance function, F(0i), which gives the fraction of light rays, incident on the aperture at angle 0 i , able to reach the absorber. The angle 0i denotes the projection of the incidence angle on the eavity's cross section. The function F(0 i ) , for a cavity with uniform illumination at the absorber, is illustrated in Fig. 1. At incident angles 10il 2 and an additional non-imaging concentrator. As the primary concentrator we have designed and constructed a novel tower concentrator. A 3.4m diameter primary mirror was mounted on a commercial two-axis positioner. Unlike the common zenith mounting, the positioner fixed axis is directed southwards, 32~ above the horizon. With this novel mounting, the concentrator is the first implementation of the Astigmatic Corrected Target Aligned (ACTA) design, which flattens the irradiation density variations during the day. 61 primary mirror segments are each mounted on a separate two-axis mount, and aligned to compensate for astigmatism. The segments are spherically curved with 17m radius of curvature, while their vertexes are placed on an 8.5 m radius spherical cap. A four-segment plane mirror reflects the light towards a horizontal focal plane. We have measured the focal spot power distribution during a full day and found good agreement with optical design calculations. Also, absorbed solar power into a rectangular 8.9x9.1 cm2 aperture were within :k30% of the average during four hour period near solar noon. Peak solar concentration in the focal plane exceeded 400 suns.

1. INTRODUCTION Solar pumped lasers may find applications at a variety of space technologies as well as on earth (Brauch et al 1991, Duchet 1992, Hall 1992). Being virtually the only energy source in space, solar energy may be used to pump solid state lasers either directly or indirectly. In indirect pumping, solar cells are used to generate electric energy, and diode lasers would convert the electric power into a pumping light. Direct solid state laser pumping with solar light is inherently more efficient, much simpler, and more reliable. Existing solar energy facilities are too large for efficiency demonstration of solar pumped lasers, and their operation is too expensive for durability studies. Therefore, for the purpose of solar pumped laser development we have designed and constructed a special solar concentrator. To pump solid state lasers, high solar concentrations are needed (Thompson et al, 1993). The desired concentration may be obtained by combining primary concentrator with f/D>2 with a non-imaging optics concentrator (Gleckman 1988). For the primary concentrator, we chose a tower configuration, with a folding mirror, which directs the solar light towards a fixed horizontal plane. Such a tower configuration is more convenient for conducting experiments than a dish or a cassagrainian configuration, in which the focal plane moves with the solar orbit. In a tower configuration, the deflection of the solar light from its incoming direction causes astignmtic aberration and focal spot increase. With a primary mirror segmentation, the focal spot may decrease to a minimum at certain hours and for a certain season by segment alignment (canting, hereafter). Usually, this is done for the solar noon, and the focal spot area

increases in the afternoon and in the morning. Several realignments are required over the year to compensate for sun seasonal inclination variation. The focal spot increase is reduced by a novel design, Astigmatic Corrected Target Aligned (ACTA) heliostat (Ries and Schubnell 1990, Zaibel et al 1995). We adopted the ACTA concept for our concentrator, and demonstrated its first implementation. The next section compares conventional mounting with an ACTA configuration, and presents the predicted solar concentration for both configurations. The concentrator construction and the performance are depicted in the third section. 2. THEORY We start with a qualitative explanation for the difference between conventional and ACTA mounting. Then, we give the design parameters used in the optical analysis, and present its results. In a conventional tower configuration, the fixed azimuth axis is vertical, and the elevation axis is horizontal, as shown in Fig. l a. Fig. lb shows an ACTA tower configuration, in which the fixed axis is tilted southwards to the target folding mirror and the second "elevation" axis is normal to the fixed axis. To discuss both configurations, we point out that in tower configuration incidence plane is defined by the centers of the primary mirror, the sun and the target. Incidence angle is the angular deviation of the target direction from the sun direction. For both configurations, the incidence plane is normal to the ground at canting time of solar noon. To simplify the discussion we designate the central highest, lowest, leflmost and rightmost segment mirrors at canting time as Up, Down, Left


and Right segments. The canting process is symmetrical relative to a symmetry line which connects the centers of the Up and Down segments. Consequently, the Left and Right segments, for example, are aligned towards the central segment with the same angles, but from opposite sides.

355

canting continues to be effective, and astigmatic correction deteriorates only due to incidence angle slow variation. Before presenting the optical analysis results, we describe the primary mirror in more detail. The 3.4 m diameter primary mirror is mounted on a two-axis positioner. The positioner fixed axis is directed southwards, 32~ above the horizon. The primary mirror is composed of 61 hexagonal segments, 360 mm between sides, each mounted on a separate two-axis base. The segments are spherically curved at 17 m radius of curvature, while their vertexes are placed on an 8.5 m radius spherical cap, in similarity to former solar furnaces (Diver et al 1983, Thompson et al 1992). The optical design calculation were done for a semi-annual canting regime, with a winter canting and a summer one. The winter canting is for 22~ incidence angle, compatible with solar noon on November 7 and on February 7. The summer canting is for 42~ incidence angle, optimal for the solar noon on May 7 and on August 7. The focal point distribution for several hours during the day, and for various days during the year was calculated by OMEK OPTICS using in house ray tracing code. Fig. 2 compares calculated focal distribution with experimental ones, which will be discussed in the third section. The summer canting calculations are for 2 hours before solar noon (a), solar noon (b) and 3 hours after solar noon (c), all during June 23.

Fig. 1 Tower configurations: Conventional versus Astigmatic Corrected Target Aligned mounting

When the sun moves from its southward direction, the canted primary mirror revolves around the two axes to follow solar orbit. However, the exact mode of revolution is different for each configuration. In the conventional configuration, the Up, Down, Left and Right segments stay always in their respective highest, lowest, leftmost and rightmost positions, while the incidence plane stops to be normal to the ground. Consequently, the incidence plane is not anymore situated symmetrical to the primary mirror, the astigmatic correction deteriorates, and the focal spot size increases. In ACTA configuration, the incidence plane contains the target oriented fixed axis. To follow solar orbit, the primary mirror revolves around the fixed axis, keeping the symmetry line on the incidence plane, while it rotates around the "elevation" axis to compensate for the slowly varying incidence angle. Thus, the

Fig. 2 Calculated versus measured light distribution at focal plane at June 23. a. 2 hours before noon. b. solar noon. c. 3 hours after solar noon. Calculation square frame sides are 200, 200, 300mrn, respectively. Fig. 3 shows the hourly variation of the focal spot power density for May end, demonstrating that power density is within :k25% of the average power density during four hours around solar noon. For comparison, we also present calculated focal spot power density for conventional tower configuration of an otherwise identical design, clearly manifesting the advantage of ACTA over the conventional mounting.

356


Fig. 3 Peak solar density at focal spot during the day for ACTA (solid line) and conventional (dashed line) mountings

3. CONSTRUCTION AND PERFORMANCE The primary mirror was mounted on a commercial two-axis ORBIT Advanced Technologies AI~035-1SL antenna positioner. Serving as a basis for the positioner, a steal construction beam was fixed to a concrete base, which was forged to a concrete roof of a three stage building. For southward orientation, we looked at Polaris with a theodolite, and used time dependent astronomic calculation to correct for Polaris north deviation at Rotem Industrial Park.

Fig, 4 The 61 segments primary mirror with individual bases on hexagonal steal frame and 9 parallel beams The primary segments are individually mounted on a two-axis base. The base mounts are connected to a hexagonal steal frame through 9 parallel beams as shown in Fig. 4. A large steal

construction contains the experiment booth and serves as a rigid base for the folding mirror and for a 150X112 cm2 optical table located inside the booth. The construction is rigidly fixed to four concrete cubes, which are forged to the concrete roof as well. The folding mirror frame can rotate around an east-west and around south-north axis. Four glass-on-aluminum plane mirrors serve as reflecting surfaces. Mechanical attachment enables mutual alignment of the four segments to constitute a single plane. The solar light enters the experiment booth through a special hole in its roof, and the focal spot plane is the entrance plane for the collecting device, which is a Compound Parabolic Concentrator (CPC) for solid state laser pumping. Before installing the folding mirror, we have mounted a Doraboard 1200 target on the direct focal plane, and used removable plastic covers on the segments to expose the target to a separate segment at each time. This was used initially for canting, and later to evaluate the hourly focal spot variation. To that end, we have exposed five segment mirrors; the central mirror and the aforementioned extreme segments. The resulted focal spot envelope on the Doraboard target was video recorded, and a SPIRICON beam analyzer was used to evaluate the intensity distributions. Typical results compare well with the calculation results of the former section, as shown in Fig. 2. To follow the solar orbit, we used the ORBIT option of the positioner. The solar orbit in the zenith-north-east coordinate system is calculated using an astronomical software, US Naval Observatory MICA1.5 (1998). An OMEK OPTICS code (ZAVIT1-2.0) bisects the incidence angle to account for reflection, and transforms the bisector coordinates to a southwards tilted coordinate system. The calculating PC computer sends a weekly table of coordinates to the positioner controller for off-line tracking. A 2 kVA Gamatronic Uninterrupted Power Supply (UPS) drives the positioner to ensure fast removal of the high intensity focal spot under electrical grid failure. To monitor the ACTA concentrator operation, we have installed three auxiliary systems. The first one is a small meteorological station equipped with a Young 06201 wind tracker, Exetech temperature and humidity digital meter, and EPLAB NIP pyrheliometer on ST-1 equatorial mount. The second system is a vision system with which the worker inside the booth keeps control of the positioner motion. One of its two video monitors shows the scene of Fig. lb, while the other one views the focal spot plane. The third auxiliary system is a cooling system, which removes excess heat from various part of the CPC and laser head, and monitors temperatures and cooling liquid flow rates. National Instruments DAQ AT-MIO-64E card and LabVIEW 5.01 software are used for data acquisition, data presentation and calorimetric calculations. Detailed description of the cooling system is due to appear in a forthcoming report on the Rotem solar pumped laser. With the cooling system, we could measure the absorbed solar power into the 8.9X9.1 cm2 CPC aperture. In Fig. 5 we show the measured absorbed solar power on May 19 as a function of the solar hour. The power is shown to be within :~.30% of the average from 10:00 to 14:00, in good agreement with the calculation of Fig. 3. It should be noted that solar insolation


variation was not taken into account in the calculation, while the insolation manifested a +4% variation between 10:00 and 14:00 (solar time), which affected the absorbed power to the same extent.

357

Gleckman P., Achievement of ultrahigh solar concentration with potential for efficient laser pumping, Applied Optics,

27,4385-4391, 1988. Hall R. B., Lasers in industrial chemical synthesis,

Laser

Focus, pp. 57-62, September 1992. US Naval Observatory Multiyear Interactive Computer Almanac 1990-2005, Willman-Bell, Richmond 1998. Thompson G. A., Krupkin V., Oron M., Yogev A., High power solar pumped solid state lasers, in CLEO, (OSA), 11,

Q 500 a..~ 400 300

590-592, 1993. Thompson G. A., Krupkin V., Yogev A., Oron M., Solar Pumped Nd:Cr:GSGG parallel array laser , Optical

200

0 ~ .o
that is being built in Auroville

1. INTRODUCTION This paper describes a solar bowl that is being built in Auroville, near Pondichery in India. This bowl project has been initiated and realised by the CSR (Center for Scientific Resarch), an Indian research institute based in Auroville (Tamil Nadu) working essentially on renewable energies and appropriate technologies. This project has been financed by the

Government of India under the authority of the Ministry of Non conventional Energy Sources.

2. CHOICE OF THE SOLAR BOWL TECHNOLOGY The solar bowl is one of the wellknown thermal solar concentrating technology, but it has not been developed as much

406


as the solar troughs, solar dishes and solar tower. A solar bowl is a fixed spherical cap covered with mirrors, it concentrates the sun rays on a radial line moving with the hour and date of the year. A boiler is hanged inside the bowl and tracks the sun in order to always be situated on this focus line.

A few bowls have been constructed in the 80's: - a small bowl of 3.5 m aperture diameter in Auroville, - a 20 m diameter bowl in Texas (Reicher, 1982), - a 10 m diameter bowl, called Pericles (Authier, 1982), built by a team from the CNRS (Centre National de la Recherche Scientifique) in Marseilles (France). It was later dismantled and rebuilt in Recife (Brasil) where it worked satisfactorily for 2 years. These two last bowls were extensively studied and tested. The conclusion was that the bowl technology could be advantageous in developing countries, especially because the low cost of civil work would allow the construction of a cheap spherical cap (Lodhi and O'Hair, 1982). But the proof was still to be made, the CSR wanted to try it out.

3. DESIGN AND SIZING OF THE SOLAR B O W L 3.1. Choice of the application At that time, the CSR was involved in the construction of a collective kitchen using steam for cooking. We decided to merge the two projects and to install a solar bowl on top of the kitchen building to provide the steam. Therefore, we had the opportunity of an application using directly the heat produced and functioning in actual conditions. This bowl had to be cheap and realised only with Indian equipment in order to be replicable afterwards, but it also had to be an operational tool, easily controlled and reliable with a reduced and simple maintenance. This implied solutions both simple and largely automated ... which was a real challenge for the equipment and for the staff as well. 3.2. Designing and sizing the solar bowl Other existing collective kitchen using steam cooking could give us the amount of steam used to cook one I ndian meal. Our collective kitchen is supposed to cook 1000 meals twice a day. We wanted to be able to provide the totality of the required steam with the solar system on clear days, but we also decided to complete the production of steam with a conventional boiler in case of a bad weather. People should be able to eat everyday, even during the monsoon! We designed a bowl of 15 meter diameter, 120 degrees aperture angle. It should supply the cooking pots with 600 kg of steam at 110~ 1.2 bars. This evaluation is based on the results of the few bowls previously built around the world, in correlation with

their size and the solar radiation available on the site. The bowl was nicely integrated in the roof of the kitchen, it is tilted of 12~ (latitude of Auroville).

Geometrical characteristics of the bowl concentrating surface - Aperture diameter: 15 m - Aperture angle: 120 ~ - Aperture area: 176.6 m 2 - Curve radius: 8.66 m - Tilt towards the South: 12 ~

4. REALISATION OF THE REFLECTIVE SURFACE In order to have an operational tool, we used as much as possible proven solutions; but we had to innovate to realize the reflective surface. In the other bowls the reflective surface was made of curved mirrors fixed on a metallic frame. In our case, we chose a different solution in order to get a minimum price in a rural Indian environment. The reflective surface is made of small flat mirrors fixed on a spherical cap built with civil work. This solution had not been applied anywhere before, so this is the more innovative part of our project. We had to define and experiment original solutions for the equipment and procedures to be used on the site to realise a reflective surface of a high quality. The main research work was to transform theoretical solutions into concrete realisations. Therefore we made a lot of tests beforehand but our work benefited essentially of the CSR experience in realising appropriate technology projects in rural India.

4.1. Making of the shell A structure made of compressed earth bricks stabilised with cement supports a spherical shell cap made of prefabricated ferrocement elements. Then this shell is covered with a cement plaster and small flat mirrors are glued on it. 4.2. Mirror size We tested different size of mirrors, and finally chose a base of 15 cm x 15 era. This leads to a reasonable number of 10 500 mirrors and an acceptable angular error due to the fact that the mirrors are flat instead of being curved. In fact, they are not squares but trapezoids of decreasing surface from near-squares of 225 cm 2 at the edge of the bowl down to triangles of 140 cm 2 at the bottom. 4.3. Mirror protection We had to undergo a lot of experiments to find out how to avoid the damaging of the silvering by the glue or the ions


released by the cement. We finally decided to protect each mirror back by gluing a glass sheet of the same size.

Mirrors 10 500 mirrors Trapezoids of 140 to 225 cm 2 (from squares of 15 cm side at the edge to triangles of 10 cm side at the bottom) Float glass mirrors 3 mm thick glued to a 2 mm thick glass sheet Mirror protection

So, two stages of preparatory work are necessary before setting up the mirrors on the cement shell: 1) cutting the glass and mirror sheets in precisely defined trapezoids 2) gluing together with araldite a mirror and its glass protection 4.4 Setting up the mirrors Then these couples glass+mirrors are set up one by one on the cement shell. We had to define a procedure in order to place them precisely in their right position and orientation.

The correct position of the mirror on the shell is defined with the help of cardboard lines glued provisorily on the cement The orientation is correct when the center of the mirror plane faces exactly the virtual center of the bowl spherical surface. This is obtained with the help of a device fixed on a pivot situated at the virtual center of the sphere. This device is composed of a small laser pointer and a circular white target on either side of the sphere center. The target center is symmetrical to the laser emitted dot. The laser beam is pointed towards the center of a mirror. If the mirror is properly orientated, the reflected laser beam hits the center of the target. The following five steps have to be applied to set up each mirror: 1) positioning the laser beam 2) putting 6 spots of silicone glue on the back of the couple mirror+glass 3) gluing the couple mirror+glass in its exact place on the cement surface 4) checking the reflected laser spot on the target 5) pressing lightly on the mirror in order to shit~ this spot to the center of the target. We applied this procedure 10 500 times, we got an average value of 3 angular minutes for the precision of the mirror orientation.

407

A team of 4 persons was necessary to set up 20 mirrors an hour. Since it is too hot to work inside the bowl during daytime, a trained team could set 100 mirrors everyday from 5 pm to 10 pm. For 10500 mirrors, we needed 105 working days.., it actually took almost six months. It is obvious that such a solution would be a nonsense in a developed country, but it is indeed quite adapted technically, economically and socially to rural India.

5. THE HEATING SYSTEM The heating system is composed of two loops : - a primary loop containing a thermal fluid called Therminol 66, including the solar boiler and a heat storage tank, able to store one hour of peak steam production, a secondary loop containing water. Steam is produced in a heat exchanger situated inside the heat storage tank and then transported to the utilities. This loop includes also a diesel-fired boiler to complement or replace the solar boiler steam production in case of clouds, especially during the monsoon period. -

The solar boiler is supposed to heat a variable flow of Therminol 66 from 150~ to 250~ Its design derives closely from the boiler used in the Pericles bowl. It is composed of a cylindrical low concentration part topped by a high concentration conical part. A strap of 3 small parallel pipes is wrapped around a supportive sleeve. The shape and dimensions of the boiler have been optimized with the aid of a statistical method applied to the reflected sun rays. The diameter and number of small pipes are determined by the fluid temperatures and flow rates inside the boiler. These calculations are based on the studies made by the Pericles team: an optical simulation of the collector (Authier, 1979) and a modelisation of the thermal fluid flow (Pouliquen and Authier, 1979). The boiler (and its fluid feeding pipes) is attached on an arm hold by a pivot fixed at the virtual center of the spherical surface of the bowl. The whole is supported by a strong monopod standing inside the bowl, away from the boiler trajectory. The pivot allows the tracking of the sun. The arm is pivoting in the North-South and East-West directions to follow the sun movement during the day throughout the year. The tracking system will be driven by a step by step motor and controlled by a computerised system. 6. SOME HINTS ABOUT STEAM C O O K I N G Cooking with steam is fast, clean, safe, healthy and more energy efficient than any burner heating only the bottom of the pot. Two types of containers are used:


408

- double-jacketed vats where steam condenses between the two jackets and is then sent back to the exchanger to be recycled. They are used to cook vegetables, lentils or noodles. - steam ovens where the steam is set free inside the oven and not recycled after use. These are preferably used to cook idlis (a popular Tamil breakfast) or rice in large cylindrical vessels containing the rice and its cooking water.

7. P R E S E N T

STATE OF THE WORK

The collective kitchen is functioning since December 1997 with steam entirely provided by the diesel-fired boiler. The reflective surface of the bowl is completed. The monopod and the solar boiler have been manufactured, they are now ready to be assembled. The tracking system is in its progrzmming and testing phase. The bowl should be operative in October 1999. REFERENCES

Reicher J.D. (1982) The Crosbyton Solar Power Project : Fixed Spherical Mirror / Tracking Receiver, Department of Electrical Engineering, Texas Tech University, Lubbock, Texas. Authier B. (1982) Le collecteur sph6rique fixe Mini-Pericles.

Revue Internationale d'H~liotechnique, 1er semestre 1982. Lodhi M.A.K. and O'Hair E.A. (1982) The solar bowl technology transfer to developing nations: a case study of Pakistan. Texas Tech University, Lubbock, Texas. Authier B. ( 1 9 7 9 ) Optical simulation for a fixed spherical solar collector. Applied Optics, Vol 18. N ~18. Pouliquen D. and Authier B. (1979) Mod61e d'6coulement fluidique dam les chaudi~res ~ veine h61ico'idale, Revue de Physique Appliqu~e, Janvier 1979, Tome 14, pp 91-95.


409

THE DEVELOPMENT AND TESTING OF SMALL CONCENTRATING PV SYSTEMS George R. Whitfield, Roger W. Bentley, Clive K. Weatherby and Alison Hunt Dept. of Cybernetics, Univ. of Reading, Whiteknights, PO Box 225, Reading, RG6 6AY, U.K. Tel + 44 118 931 8223, Fax + 44 118 931 8220, e-mail [email protected]

Hans-Dieter Mohring, Fritz H. Klotz and Peter Keuber ZSW, Hessbruehlstr. 2 l c, D-70565 Stuttgart, Germany. Tel + 49 711 7870 272, Fax + 49 711 7870 230, e-mail [email protected]

Juan Carlos Mifiano and Elisa Alarte-Garvi I.E.S., Universidad Politecnica de Madrid, E.T.S.I. Telecomunicacion, Ciudad Universitaria, E-28040, Madrid, Spain. Tel + 3491 336 7222, Fax + 3491 544 6341, e-mail [email protected] Abstract - Spreadsheets have been used to compare some 90 possible small PV concentrator designs that might be

suitable for use at remote sites. They have apertures of about 2m 2, use BP Solar LBG cells, and employ small aperture modules to reduce heat sinking and construction costs. Designs include fixed V-troughs and CPC's, single axis tracked cylindrical lens and mirror systems, and 2-axis tracked spherical-symmetry systems. Performance and volume production costs were estimated. Four promising systems were constructed as prototypes. A- Point-focus Fresnel, 2-axis tracking; Cg = 32x; and 69x with secondaries. B - Line-focus mirror parabolic trough, 1-axis tracking, Cg = 20x. C - SMTS ('Single mirror two-stage'), 1-axis tracking, Cg = 30x. F - Multiple line-focus mirror parabolic trough, E-W 1/day manual tracking, Cg = 6x. The prototypes were tested at Reading, and three for up to a year's field trial at ZSW's test site, Widderstall, in Germany. The best module efficiencies, normalised to 25~ and excluding the end losses of linear systems, were: 12.5%, 13.2%, 13.6% and 14.3% for collectors A, B, C, and F, respectively. The tests have shown the practicality and robustness of the designs, and the performances of collectors B, C and F are only 10% below the estimates in the spreadsheet calculations. The best of the collectors have costs in the region of 1.5 to 1.8 $/Wp, yielding energy costs at a good site (excluding BOS and overheads) of between 5 and 7 cents/kWh. A conventional PV array costs 4.3 $/Wp, and 18 cents/kWh.

1. INTRODUCTION Photovoltaic systems have advantages as sources of small amounts of electrical power in remote areas, but conventional solar panels are expensive. Since lenses and mirrors in volume cost only about 1/20 as much as solar cells, it should be possible to reduce the cost of PV electricity by using them to concentrate the sunlight from a large area onto a small area of solar cells. Many concentrators have been developed in the past, but they have usually been not much cheaper than conventional solar panels, because concentrator solar cells have cost much more than one-sun cells and the optical and tracking systems have been expensive. Recent developments, such as BP Solar's Laser Buried Grid cells (Mason, Bruton and Heasman, 1995), have made it possible to manufacture solar cells little different in design and cost from one-sun cells that can be used at concentration ratios up to 40x. Using these cells with optical and tracking systems that are no better than required there is considerable scope for cost reduction. This JOULE III project built on the progress made in a previous JOULE II project, EUCLIDES (Sala et al., 1997), led by BP Solar and joint with UPM, ZSW and Reading University, which developed a PV concentrator for large grid-connected systems. A 480 kWp demonstration plant, based on the EUCLIDES work, has been built in Tenerife (Sala et al., 1998). In the present project the objective instead has been to reduce the cost of PV electricity by developing small concentrating systems of about 2 ~ aperture designed for use in remote areas. Specifically, the aim has been to produce a small number of prototypes of such systems sufficiently near practical

production to be of interest to industrial companies. We have examined a wide range of possible concentrators using BP Solar LBG cells. For each system, performance and cost have been estimated on common basis, assuming large-scale production. No such comparison of a wide range of systems been made for many years. 2. THE COMPARATIVE ANALYSIS OF A WIDE RANGE OF POSSIBLE CONCENTRATING SYSTEMS Insolation data were assembled for three representative sites, and a database of materials costs was generated. Using these data, some 90 possible small PV concentrator designs suitable for use at remote sites were then compared (Whitfield et al., 1997, 1998). The systems envisaged have apertures of 2 r~, use BP Solar LBG cells, and are composed of smaller aperture modules to reduce heat sinking and construction difficulties. Designs considered included: - fixed V-trough and CPC systems, - cylindrical single-axis tracked systems up to 30x, - spherical-symmetry 2-axis tracked systems up to 69x, - secondary optics, and a variety of manual and automatic tracking strategies. The performance of each system was calculated on a spreadsheet, taking due account of its dimensions, the properties of the optical components used, the efficiency of the solar cells, and the energy capture of the tracking strategy chosen. Each system was then given outline mechanical design, bearing in mind ease of mass production and long service fife. Using this mechanical design, each system's material cost was calculated, and with advice from a consulting engineer, the manufacturing


410

cost, in mass production, was estimated. The total cost was combined with the estimated performance to give the resulting cost per peak watt, and per kWh. Typical results are given in Table 1, which shows the cost per peak watt and per kWh for the best of the collectors considered and a few others. The main conclusions from the analysis are:1. Concentrating collectors earl be much cheaper than conventional planar collectors, by a factor between 2 and 3. 2. To obtain this improvement, they must be made in large numbers. 3. There is a wide range of good designs. The best have relatively high concentration ratios and imaging primary optical systems. 4. Mirrors are usually more cost-effective than lenses.

5. Secondary optical elements often improve the performance. 6. Automatic tracking systems are better than manual tracking, in spite of their extra cost. 3. SELECTION OF COLLECTORS FOR PROTOTYPES

From the information in section 2, the project group then unanimously (!) selected six collectors, A to F in Table 1, for more detailed analysis; of these six, four collectors were built as prototypes. Selection criteria included not only low cost, but also innovative optics (SMTS) (Alarte, Benitez and Mifiano, 1998), and innovative tracking (once per day movement). Calculations were made to optimise the parameters of each

Table 1. Results for the best collectors analysed, and a few others. The last line is a conventional planar array Letter for Prototype A A B C B

D

D E

C G F F

Secondary Optics

Mounting

Cone Ratio

Cost $AVp

Point focus Fresnel lens Point focus Fresnel lens Weatherby's Cylindrical Paraboloid SMTS Collector, Plastic, 0.6m Cylindrical Paraboloid: Multiple offset Cylindrical Paraboloid Cylindrical Paraboloid Cylindrical Paraboloid: Multiple offset Cylindrical Paraboloid Linear Fresnel lens Cylindrical Paraboloid Curved TIR Lens Cylindrical Paraboloid

Pt-focus solid CPC No No

Gimbals Gimbals Polar

69 36 19

1.46 1.48 1.62

12.1 12.2 14.0

6.2 6.3 7.2

5.4 5.4 6.2

Yes No

2-axis Polar

30 20

1.78 1.78

14.7 15.4

7.6 7.9

6.6 6.8

No Point-Focus CPC Mirror CPC

2-axis Polar Polar

20 65 27

1.95 1.78 1.88

16.1 16.2 16.3

8.3 8.3 8.4

7.2 7.2 7.2

Solid CPC Solid CPC No No Mirror CPC

Polar Gimbals Polar Polar 2-axis

37 37 20 28 25

1.90 2.02 1.95 1.97 2.06

16.4 16.7 16.8 17.0 17.0

8.4 8.6 8.6 8.7 8.8

7.3 7.4 7.5 7.6 7.6

SMTS Collector, Alum, 0.3 m SMTS Collector, Alum., 0.3m: Alfilm Cylindrical Paraboloid

Yes Yes

Polar Polar

30 30

2.00 2.01

17.3 17.3

8.9 8.9

7.7 7.7

Oil filled CPC

Polar

37

2.04

17.6

9.1

7.8

Point-Focus CPC No No No

Chinese Polar Polar Polar

50 15 30 2

1.72 2.18 2.39 4.31

19.5 18.8 21.6 23.7

9.6 9.7 10.6 13.9

8.3 8.4 9.2 12.7

Mirror CPC

E-W axis 1/day E-W axis 1/day Fixed at latitude

8

2.52

31.4

15.8

13.6

6

2.64

32.8

16.6

14.2

1

4.31

30.5

19.6

18.1

Primary Optics

Cylindrical Paraboloid Curved Fresnel lens SMTS Collector, Plastic, 0.6m V-trough, Screen Printed Single Crystal Cells Cylindrical paraboloid: multiple offset Cylindrical paraboloid: multiple offset Flat, Screen Printed Single Crystal Cells

No No

Cost Cents/kWh Wid. Man. Aim.

Note. The costs given in the table are for cells, optical systems, mountings and trackers only; balance of system costs are omitted as they are similar for all types of collector. The cost in $/Wp is based on direct beam radiation at 850 W/m2. The cost in cents/kWh is site-specific; the three columns are for Widderstall, near Stuttgart, a relatively cloudy site, Manfredonia in Italy, and Almeria in Southern Spain, a particularly good site.


collector, particularly the concentration ratio and aperture, to maximise the performance/cost ratio. In the detailed design, attention was paid to ease of large scale manufacture; for example cylindrical parabolic mirrors could be pressed from aluminium sheet or moulded in plastic, and lenses could be made by a rolling process. But the press tools and moulds were too expensive for these techniques to be used for our prototypes. So the prototypes were built with supporting ribs or glass fibre members to maintain the optical shapes. Care was also taken to use techniques and materials that would ensure a long working life and freedom from corrosion.

411

the cost of production systems. So commercial systems were purchased from the USA. Collector A uses a 2-axis active tracking system from Wattsun, with a sensing head on the inner gimbal driving right ascension and declination servos. Collectors B and C use an open loop microprocessor-controlled system, developed initially by Maish of Sandia (Maish, 1991), and marketed by Enhancement Electronics; this drives a Wattsun servo, rotating the collector about a polar axis. The systems were reliable and weatherproof, but were stronger, heavier and more expensive than was required for small systems. The microprocessor-controlled systems were difficult to set up, but ran well once this had been done.

3.1 The choice of mirror surface For the reflector surface of the three mirror collectors (13, C and 10 we essentially had the choice between four materials; where the costs shown are our assumed very large order prices. 3M's ECP305+ silver film, exterior grade; p ~=- 0.94; $ 25/m z 3M's SA-85P aluminium, exterior grade; p ~-- 0.85; $ 7/m2 3M's SS-95P silver film, interior grade; p ~=- 0.94; $ 9/m 2 Anocoil anodised aluminium; p ~=- 0.84; $ 20/m 2 Note: The first three minors are reflective film, so the cost of substrate and laying down must be added; this will be typically $12/m 2. In collectors B and F the parabolic troughs are covered with glass or acrylic sheet, so the reflective surface need not be especially weatherable. 3M's ECP305+ was difficult to obtain, as 3M have ceased manufacture, but UPM, via Mifiano, kindly supplied a roll for use on collector C, where the reflector is in the open air. A comparison of ECP305+, SA-85P and SS-95P was carried out, using the cost/performance spreadsheets for collector B; for this comparison, the reflectivities were reduced to 0.90, 0.81 and 0.88 respectively, to allow for mirror profile errors. The annual energy cost, in c/kWh, was highest for ECP305+, marginally less for SS-85P, and about 7% less for SS-95P. Anocoil anodised aluminium was not include in this analysis, as it was more expensive than SA-85P, which has a higher performance. Therefore for collectors B and F we used the SS-95P interior grade film. This is only supplied pre-glued to 0.5 mm thick aluminium sheet. This had the advantage that we did not have to do the laminating ourselves, but 0.5 mm is thinner than we would have liked for heat-sinking. 3.2 Heat sinking Heat sinks for the cell strings were designed to hold the cells at temperatures similar to those of conventional flat arrays, typically 30-40 ~ above ambient. This requires an effective cooling area approximately twice the projected area of the array; for a flat array this is just the front and rear surfaces; for collectors A, B and F, there were suitable aluminium surfaces forming the structure or mirror surface of the collector. Studies carried out in the previous (EUCLIDES) project (Whitfield et al., 1997) and thermal tests on small mock-up collectors showed that adequate cooling would be obtained with the 0.5 mm aluminium sheet if the aperture was not above about 15 cm. So this aperture was chosen for collectors B and F. 3.3 Trackers Time did not permit the development of custom-built trackers for each collector, although they will be necessary to reduce

3.4 Solar cells The solar cells used for the project were BP Solar 'Saturn' laser buried grid cells (Mason, Bruton and Heasman, 1995). These cells, developed from work by Prof. Green of the University of New South Wales, Australia, can be made for little more than the cost of conventional screen-printed one-sun cells, but have inherently higher efficiency and lower series resistance; this makes them particularly suitable for moderate concentration ratio concentrators such as ours. The cells were cut to length 50 mm and widths appropriate for each collector, retaining a full length bus bar down one side. They were connected in strings with conventional tabbing strips, using thicker than usual material to keep the resistance down. The strings were mounted on machined aluminium carriers, using a thermally conducting electrically insulating tape, and encapsulated in transparent silicone, with glass or Tefzel covers. This technique was satisfactory, but was rather slow and messy. 3.5 End losses of cylindrical collectors. Collectors with cylindrical optics, such as linear Fresnel lenses or cylindrical paraboloid mirrors, concentrate the light into a line focus; the solar cell string is mounted along this line, and the collector is rotated about an axis parallel to this line to keep the focal line on the cell string. Thus a single set of bearings and a single tracker suffice.

Figure 1. Shading loss: The cell string has to be shorter than the collector If, as is common, the chosen axis points at the pole, then tracking is perfect at the equinox, when the sun's declination is zero; the sun's rays are perpendicular to the axis of the collector, and the cell string can be as long as the minor. But at any other date, when the declination is not zero, the focal line is displaced towards or away from the pole. The worst case is at the solstice, when the declination ~5reaches its maximum value of 23.5 ~ Then in a typical case, Figure 1, a length 2h tan8 of the cell string is not illuminated. At the other equinox, an equal length is shaded at the other end of the cell string. Since one shaded cell blocks the whole output current of the string, it is usual to reduce the cell string length to d - 2 x 2h tan~, so that the whole of the cell string is always illuminated. The output of the collector is reduced proportionately. This end loss has been

412


computed for each collector and included in the spreadsheet analyses. With long narrow collectors the loss is small, but with wider apertures it becomes significant. For collector B, a narrow offset paraboloid, the correction for end losses is a factor of 89%, but for collector C, the wide SMTS collector, it is 65%. This end loss can be eliminated by using a gimbal mounting and two-axis tracking, but this adds cost and mechanical complication. But a partial solution is to use a full length cell string and bypass diodes, Figure 2. Suppose that there are n cells in the centre part of the string that is never shaded, and m cells at each end that are shaded at the solstice. Then a full string is n + 2m cells, and the correction for end losses if the shaded cells are omitted is n/(n+2m). ~

~

t

The collector of this project differs from previous similar systems primarily in the use of only moderate concentration levels, and hence the ability to use the commercially-priced essentially one-sun technology LBG cells. It has the potential to incorporate a secondary optical element, to raise the concentration ratio, or to increase tolerance to tracking errors. Half the collector was fitted with secondary concentrators, and the other half was not. This collector is an easy device to manufacture in volume. The acrylic lenses are made by 3M's calendaring process, already used in volume for street signs, while the aluminium sheet housings can be pressed to shape. The cells are relatively large, so either manual or automated tabbing, encapsulation and mounting are straightforward. 4.2 Collector B, Multiple offset cylindrical paraboloid, one

axis tracking

Figure 2. The longer cell string with bypass diodes If the shaded cells are fitted in the collector, then at the equinox, all the cells are illuminated and there is no end loss. In the worst case, at a solstice, the m cells at one end only are shaded, and the output is that from n + m cells. One bypass diode will be turned on, and the voltage drop across it will be about the same as the output voltage of one cell. So the output at a solstice will be that from n + m - 1 cells. Away from the solstice, fewer than m cells will be shaded and the output will increase. So the end loss will be less than half than that when the shaded cells are omitted; for the SMTS collector the correction factor is increased from 65% to about 83%. (A full calculation of the loss factor depends on the distribution of sunlight through the year, and has not been attempted). The cost of the extra 2m solar cells is small, and the cost of the diodes is negligible. The cost per peak watt of collector C, the SMTS collector, is reduced from 2.39 $/Wp to about 2.06 $/Wp. Of course the voltage of the peak power point changes as cells are shaded, and so a true peak power tracker and appropriate array to load converter are needed in the system; but they should be provided anyway for an optimal design. 4. DISCUSSION OF SPECIFIC COLLECTORS

4.1 Collector A, Point focus Fresnel lens, with two axis tracking This collector (Hunt, 1998) is a two-axis tracked, point focus system using flat acrylic Fresnel lenses. As such, it is similar to many previous systems, including: the original Martin Marietta Soleras collectors (Salim and Eugenio, 1990), various Sandia designs (Chiang and Quintana, 1990), and collectors from Alpha Solarco (Carroll, Schmidt and Bailor, 1990) and Midway. The advantages of this design include:- maximum beam insolation collection due to two-axis tracking potential for simple mass-produced optics, the use of the housing as heat sink; while the main disadvantages include:- the increased cost of the second axis tracking, the fact that flatFresnel lenses offer reduced efficiency at practical f-numbers.

This collector (Weatherby and Bentley, 1998) was designed for polar axis tracking. The designed concentration ratio was 20x. A single module consisted of two cylindrical parabolic mirrors, each with an aperture of 150mm x 1.5m, formed from thin (0.5mm)sheet aluminium alloy laminated with SS-95P reflective material. The mirrors were arranged back to back so that the two 7.5ram wide strings of cells, located near to the focal lines, were on the outer walls of the module, to facilitate the cooling of the cells. The cooling was further enhanced with a fin extending normal to the sun. A simple glass or acrylic cover was used to protect the optics and add strength to the module. The cell strings of two modules were connected in series forming one integrated double module of about 28Vor Advantages of this collector are the simple modular construction and the ability to mass-produce the housings by an inexpensive stamping method. The cooling arrangement removes the need for bulky expensive heat sinks whilst keeping the cells at temperatures no higher than those of conventional flat panels. To avoid the cost of the large press tools necessary for mass production, a series of fibre-glass ribs were moulded and used to retain the mirrors in their parabolic form. One disadvantage of this technique is that the ribs obstruct the convective flow of cooling air. It is expected that mass-produced modules, with no fibs, will give better cooling and even better performance than the prototype. 4.3 Collector E Multiple offset cylindrical paraboloid, moved manually once per day Of the four prototype collectors, this one (also Weatherby and Bentley, 1998) is designed to be the simplest and most robust, as no automatic tracker is required. It consists of parabolic troughs that run E-W, so that as the sun moves approximately along the axial plane of the parabola the focal line moves mainly E-W along the line of cells. The collector tilt is designed to be manually re-aligned every day, or every few days, to accommodate the changing declination of the sun. Even so, except at the equinox, the sun moves along a small circle, and so moves a little in a N-S direction off the great circle defined by the axial plane of the parabola. The width of the cells therefore has to be greater than the width of the focal line, to allow for this N-S movement. The chosen geometrical concentration ratio is 6, giving an acceptance angle of about 7~, which allows the sunlight to fall on the cell string for about 8 hours on most days.

413


The design of the collector envisages a very simple process when in mass-production: - the full 2m 2 reflective sheet is stamped to profile, - the cell strings are attached, - the glass cover is bonded in place. For the prototype, we could not afford the tool for a stamping, and a rather elaborate fabrication of the parabolas, supported by a lattice of ribs and cross-members, was used instead.

can be allowed for in subsequent analysis. Collector A (section 4.1 above), provided experience with a two-axis tracker. 4.5 Collector D, Cylindrical paraboloid, with secondary

concentrators and single axis tracking This was another promising design, slightly more costly than A, B and C, but cheaper than F and G. It was not selected for manufacture in this project because we had already chosen two better mirror systems, and we wanted to include a system with manual movement instead of automatic tracking. 4.6 Collector F~ Linear Fresnel lens, with solid CPC secondaries and two axis tracking This collector, although more costly than some of the others, has the advantage of being simple and totally enclosed. Some work was done on the manufacture of prototype linear lenses, by machining them from acrylic sheet, but it was not possible to polish the milled surfaces to a sufficiently good finish in a reasonable time. Some lenses are commercially available, but they are designed for other purposes and are not of the correct dimensions for our collector. Special lenses could be made to our design by 3M, but the cost of a prototype was prohibitive. So this design was abandoned.

Figure 3. Cross section and workings of the SMTS collector 4.4 Collector C, Single mirror two stage (SMTS) collector, single axis tracking The SMTS collector, Figure 3, (Alarte, Benitez and Mifiano, 1998) presents some characteristics that make it a good candidate for achieving the required cost reduction. In particular: a) Since the cell strings and the heat sinks are at the mirror edges, the heat sinks can form part of the mechanical structure of the collector, reducing the mass and cost of other structural material. b) There is only one mirror per concentrator, in such an arrangement that two concentrators share their mirror so one mirror works as a first stage for one concentrator and as a second stage for the other. Therefore it gets a high acceptance angle (around 90% of the theoretical maximum corresponding to its concentration) yet maintaining the structural simplicity of a parabolic trough collector (whose acceptance angle reaches no more than 50% of the theoretical maximum). This high concentration-acceptance angle product makes the collector less sensitive to tracking or manufacturing errors, which reduces again the manufacturing cost. The spreadsheet analysis showed that it would be most costeffective to use a two-axis tracking system; however for mechanical convenience a single-axis polar mounting was chosen for the prototype. The loss, due mainly to end losses,

4.7 Collector G, V-trough, with single axis tracking This collector was already being investigated by ZSW, as a cheap alternative to fiat plate systems. Although not so cheap as the best collectors, it is much nearer to practical production, as the solar cell arrays can be conventional planar arrays and the optical design and construction of the mirrors is not critical. The tracker was to be thermo/hydraulic, which does not require an external power supply, and promises to be cheaper in production than an electronic system. But the prototype, being built under another programme, was not ready in time for this project. 5. THE PROTOTYPES BUILT The characteristics of the prototypes are summarised in Table 2, and photographs of them are shown in Figures 4 to 7. 6. TESTING All the prototype collectors were built at Reading. During construction, the solar cells were all individually tested on a computer controlled laboratory tester, to reject faulty ones (in fact, none) and to select matched sets for each string. The variation between cells was small, and perhaps due to measuring errors; it was probably unnecessary to select at all. The peak efficiency of the cells used was about 18 %, corrected to 25~ at an insolation of 15 kW/m 2, failing to about 17 % at

Table 2. The prototype collectors

Code A2 A1 B C F

Primary Optics Point-focus Fresnel Point-focus Fresnel Parabolic mirror SMTS Parabolic mirror

2ndry

Conc.

Optics No Yes No Yes No

Ratio 32x 69x 20x 30x 6x

Module Width (In) 0.225 0.225 0.150 0.300 0.150

Aperture

(mz) 0.81 0.81 1.80 2.40 1.80

Axes .Tracking 2-axis: cont. 2-axis:cont. Polar: cont. Polar: cont. E-W: 1/day

414


Figure 4. Collector B. Multiple cylindrical semi-parabolic mirrors, tracking about a polar axis.

Figure 5. Collector C. The single-mirror two-stage collector, tracking about a polar axis.

Figure 6. Collector F. Multiple cylindrical semi-parabolic mirrors, moved manually every day about an E-W axis

4 and 3 7 kW/m2. The cell strings were tested at 1-sun, to check for faulty connections, before being fitted to the collectors. As soon as they were completed, the collectors were tested for a short period, measuring spot performance and I-V curves. The results are summarised in Table 3, 'Measured Efficiency', 'Best Module' and 'Collector'. Three of the collectors, A, B, and F were then sent to ZSW's test site at Widderstall for long term testing. A computer-based data logger was used to record, at one minute intervals, output current, voltage and power at the peak power point, cell temperature, tracking accuracy, and a range of meteorological data, notably direct and global insolation, ambient temperature, and wind strength and direction. From time to time I-V curves were taken. Typical results for insolation and output power for a clear warm summer day (8th August 1998), are shown in Figures 8 and 9. Beam radiation was above 800 W/m2 for more than 8 hours. The daily insolation sums were: diffuse horizontal: 859 Wh/m2, direct normal: 10180 Wh/m2, and total normal: 11393 Wh/m"~. Ambient temperature was between 20~ and 29~ and wind velocity between 1 m/s and 2.5 m/s while the beam radiation exceeded 800 W/m2. The 2-axis tracked collector, A2 was in operation for about 12 hours; the end switch stopped further tracking in the evening. The fixed collector F was active for approximately 8 hours. The polar axis tracked collector B generated power for more than 12 hours. The total generated energies for this day were 725 Wh for Collector F, 505 Wh for Collector A2 and 1239 Wh for Collector B. The output power of collectors A2 and F were lower than they were on a winter day, 25 th March, mainly due to the high summer temperature. Collectors B and F, tested at Widderstall proved to be reliable practical units. They withstood a year of weather with no significant deterioration. Heat sinking has proved satisfactory, with cell temperatures 30~ to 45~ above ambient, similar to conventional flat panels. The test results are summarised in Table 3. The best module efficiencies, normalised to 25~ andexcluding the end losses of linear systems, were: 12.5%, 13.2%, 13.6% and 14.3% for collectors A2, B, C, and F, respectively. Full systems performed somewhat worse than their best modules, due to manufacturing variations. There were two serious problems revealed by the tests:1. Collector A, using point focus Fresnel lenses had a concentration ratio of 32x without secondaries, and 69x with secondaries, too high for the cells used, which were optimised for 15 suns, causing a loss of efficiency at the higher concentration. The cure is to use cells optimised for a higher concentration ratio, or to redesign the optical system for a lower concentration ratio. 2. Collector C, the single mirror two stage concentrator was made of fibreglass (GRP) mouldings, coated with a silvered plastic reflecting film. The intensity of sunlight on the secondary part of the mirror was sufficiently great to burn the GRP wherever there was a flaw in the mirror. So, although the optical performance was very good, the collector would not be reliable in use. The cure is to build the mirror from aluminium, to spread any local heating over a reasonable surface area.

415


Figure 8. Insolation on a clear summer day, 8th August 1998.

Figure 7. Collector A. Multiple circular Fresnel lenses, mounted in gimbals, with closed-loop active tracking about both axes. 7. C O N C L U S I O N S Several designs of small concentrator systems can be significantly cheaper than conventional planar arrays, reducing cost/watt and cost/kWh by a factor of 2 or 3. To achieve such reduced costs, the concentrators should be designed to use minimum amounts of material, and be manufactured in such a way, and in sufficient quantity, as to keep down the manufacturing cost.

Figure 9. The output of collectors A2, B and F on 8th August 1998.

Table 3. S u m m a r y o f costs and performances o f the prototype collectors

Code A2 AI B C F

Optics Pt-focus Fresnel Pt-focus Fresnel I Parabolic mirror I SMTS I Parabolic mirror

Conc~ Ratio 32x 69x 20x 30x 6x

From Spreadsheets Effcy. (%)* $/Wp I c/kWh* 16.0 1.48 [ 5.4 15.5 1.46 [ 5.4 15.1 1.62 I 6.2 15.1 2.39 I 9.2 15.4 2.64 I 14.2

Measured Efficiencies (%)* Best Mod. 12.5 9.8 1'3.2 1'3.6 14.3

I

Coll. I ZSW spot 12.3 [ 8.0-11.0 9.4 [ 2.5 -5.0 [10.619.4-11.0 . 12.2 I 13.4

. I 9-13

ZSW lon~* 7.0 - 8.5 2.2- 3.2 7.7 - 8.4 6.9 - 9.2

* Notes: - The spreadsheet c/kWh is for a good site (Almeria), and is a comparative number: it includes cells, optics, housing, structure, tracking, and manufacture, but does not include, e.g., power conditioning, land or overheads. The comparable figure for a flat array is 18.1 c/kWhr. - The spreadsheet efficiency is a spot (i.e. instantaneous) value. The annual efficiency assumed is sometimes lower due to tracking and other losses. The figures are normalised for cells at 25~ and do not include the 'end losses' of the linear systems (as these reflect each system's physical shape, rather than its cell and optical performance). - The measured efficiencies are: - Best Mod.: the better, or best, of the prototype modules, - Coll.: the complete prototype collector, - ZSW spot: typical spot efficiencies from the ZSW field test results. These three efficiencies are also for cells at 25~ and exclude 'end losses', and thus can be compared with the original 'spreadsheet' estimate. - The final efficiency ('ZSW long') is calculated from the total power output from each collector during its period of testing at the ZSW test site, divided by the total of the beam radiation normal to that collector's aperture over the same period. It has not been adjusted back to 25~ nor are end losses subtracted, so is not comparable with the figures given previously. However, this figure can be combined with the annual energy incident on each of the collector types to indicate that system's typical long-term output under field conditions.

416


Prototypes of four designs of concentrator were built by hand. Design and construction, though taking longer than anticipated, presented few problems. Three were tested for some months at ZSW's test site at Widderstall. Overall, the prototypes have behaved roughly as we expected; they have proved to be robust and reliable, and capable of operating for long periods in the field. The performance figures from the tests generally support the estimates entered in the spreadsheet calculations. For collectors B, C and F, best prototype module outputs ranged from 7% to 13% below the expected performance figures. Collector A, where the cells were run well above their design concentration, gave a performance (without and with secondaries) of 22% and 37% below expectation. On the basis of the spreadsheet data, the best of the collectors have costs in the region of 1.5 to 1.8 $/Wp, yielding energy costs at a good site (excluding BOS and overheads) of between 5 and 7 cents/kWh. The corresponding figures for a fixed planar PV array are 4.3 $/Wp, and 18 cents/kWh.

REFERENCES Matte E., Benitez B. and Mifiano J. C. (1998). Design, Construction and Measurement of a Single-Mirror Two-Stage (SMTS) Photovoltaie Concentrator. In Proc. 2nd World Confr. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2245-2247. Carroll D, Sehmidt E. and Bailor B. (1990). Production of the Alpha Solarco proof-of-concept array. In Proceedings of 21st IEEE PVSC Vol 2, May, Kissimimee, Florida, USA, pp 1136 -1141. Chiang C. and Quintana M. (1990). Sandia's concept-90 photovoltaic concentrator module. In Proceedings of 21a 1EEE PVSC Vol 2, May, Kissimimee, Florida, USA, pp 887891. Hunt A. C. (1998). Design and Manufacture of a Point-Focus Fresnel Lens Concentrator for a Stand-Alone PV System. In Proc. 2nd World Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2185-2188. Maish A.B. (1991). The Solartrak solar array tracking controller. Sandia Report, SANDgO--1471. UC-275. Mason N. B., Bruton T. M. and Heasman K. C. (1995). Optimisation of low-cost concentrator solar cells. In Proc. 13th European PV Solar Energy Confr., Nice, H.S. Stephens, Bedford, pp 2110-2112. Sala G. et al. (1997) Description and performance of the EUCLIDES concentrator. In Proc. 14th European PV Solar Energy Confr, Barcelona, H.S. Stephens, Bedford, pp 352 -355. Sala G. et al. (1998). 480 kW peak EUCLIDES concentrator power plant using parabolic troughs, In Proc. 2ad Worm Conf. on PVSolar Energy Conversion, Vienna, EC, Luxembourg, p p 1963-1968. Salim A. and Eugenio N. (1990). A comprehensive report on the performance of the longest operating 350kW concentrator photovoltaie power system. Solar Cells 29, pp 1-24. Weatherby C. K. and Bentley R. W. (1998). Further development and field test results of two low-material-cost parabolic-trough concentrators. In Proc. 2nd Worm Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2189-2192.

Whitfield G. R. et al. (1997). The development of optical concentrators for small PV systems. In Proc. 14th European PVSolar Energy Confr., Barcelona, H.S. Stephens, Bedford, pp 336-339. Whitfield G. R. et aL (1998). Development and testing of optical concentrators for small PV systems. In Proc. 2nd World Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2181-2184.

Acknowledgement The support of the EC for this work, under contract JOR3CT96-0101, is gratefully acknowledged.


XXI. Active Cooling, Refrigeration and Dehumification

417

418




419

THERMODYNAMIC DESIGN OF A SOLAR REFRIGERATOR TO PRESERVE SEA PRODUCTS

Hector D. Arias-Varela, Wilfredo Soto-Gomez and Oscar Castillo-Lopez Metal-mechanical Department, Institute of Technology of Tijuana, Calzada del Tecnologico s/n, Tijuana, Baja California., Mexico 22370, Phone/Fax (66) 833519, E-mail: [email protected]

Roberto Best-Brown Centro de Investigacion en Energia, Universidad Nacional Autonoma de Mexico, Apartado Postal No. 34, Temixco, Morelos, Mexico 62580, E-mail: [email protected] Abstract - This project was developed to determine a means of providing refrigeration to communities lacking conventional energy sources. The design of an absorption refrigeration system operating with solar energy was carried out. The refrigerator is used to conserve sea food. The system is adapted to an industrial size cold-storage room. A maximum of 200 kg of fish in ice may be introduced to this room daily, up to a total capacity of 2 tons. The lowest temperature the evaporator reaches is -10 C, the high and low system pressures are 13.4 atm and 2.87 arm respectively. The refrigerant-absorbent mixture is ammonia and water, where the refrigerant is ammonia. The design of this system requires six effective solar hours to generate the refrigerant needed by the refrigerator to work eighteen hours daily. Evacuated tube solar collectors are used. Only solar energy is used to operate the system. To compare the cost effectiveness of this solar refrigerator with a vapor compressor refrigerator of the same capacity, the following was considered: the vapor compression refrigerator requires electricity generated by internal combustion plant. The period of comparison is twenty five years with a MARR of 4.5%. Initially, solar energy refrigeration is more monetarily expensive, but less expensive ecologically than conventional refrigeration. However, at twenty three years of operation they become the same monetarily. Beyond twenty three years, conventional refrigeration is more expensive. I. INTRODUCTION Santa Clam Ca. Recreation Center, cooling space 2508 m 3, flat plate collectors copper/copper, collection area 650 m 2, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 25 tons (USA). -

Most conventional refrigeration systems operate with electricity, however, there are regions where it is difficult or not cost efficient to provide electric service. In addition, the cost of generating electricity is high, both economically and ecologically. Therefore, a project was developed to determine a means of providing refrigeration to communities lacking conventional energy sources. The design of an absorption refrigeration system operating with solar energy was carried out. The refrigerator is used to conserve sea food. Solar refrigeration by absorption is considered to be an alternative to substitute conventional refrigeration equipment and a way to save electricity or make refrigeration possible in areas without electricity. While research on solar cooling has been carried, a design that operates efficiently and is economically within reach of its users has not been mass produced. The majority of research on solar cooling has focused on lithium bromide and water where water is the refrigerant. This limits the operation of the system to temperatures at or above 5 C. Moreover, even the systems designed for high temperature have low efficiency and little competitiveness in the market (Hacuz, 1982). A list of different experimental prototypes related with solar refrigeration around the world are listed as follows: - University of Ohita Building, cooling space 1860 m 3, flat plate collectors copper/copper, collection area 513 m 2, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 30 tons (Japan). - Solar House Hirakata, cooling space 118.5 m 3, evacuated tube collectors copper-aluminum, area of collection 46.6 m z, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 2 tons (Japan).

- Brisbane Solar House, cooling space 123 m 3, flat plate collectors copper/copper, collection area 6 m 2, and a refrigeration unit working by absorption with lithium-bromide-water with a capacity of 2 tons (Australia). - Carrier Corporation Phoenix Arizona, cooling space 3700 m 3, linear lenses, collection area 133.8 m2, and a refrigeration unit working by absorption with lithium-bromide-water with a capacity of 13.5 tons (USA) (Huacuz 1992), (Bogart, 1992) (Mongomery, 1991) and (instituto de Refrigeracion y Aire Acondicionado, 1989). This paper concentrates on the thermodynamic design and analysis of a solar refrigeration system. The project is unique since the system is adapted to an industrial size cold-storage room and developed to operate under the following conditions: temperatures less than 0 C in the evaporator, 18 hours of operation per day, storage of refrigerant, approximately 6 effective solar hours of solar collection with evacuated tubes and the use of solar energy as the only source of power to operate the system. The components of the solar refrigerator are shown in Figure 1. The elements of the system located in the high pressure region are the following: - Evacuated Tube Solar Collectors. The function of this component is to heat the thermal oil used by the generator.


420

- Separator. This is a more sophisticated heat exchanger that ensures approximately one hundred percent of ammonia separation from water.

- Absorber. This element contains enough water to absorb the mixture of water-ammonia or ammonia that comes from the mixer. After the absorption takes place, the solution with a high concentration of ammonia is send to the tank of strong solution. - Tank of Strong Solution. This tank keeps the ammonia-water solution. This solution is released as required and pumped to the high pressure region.

- Condenser. This heat exchanger uses water in order to condense the ammonia that will be used as the refrigerant. After the refrigerant is condensed it is sent to the refrigerant tank.

- Pump. This component pumps the solution to the generator w (in the high pressure region). The pump is powered by electricity which is generated with photo-voltaic panels.

- Generator. This element is a heat exchanger whose main pm]x)se is to heat the ammonia-water mixture in order to separate the water from the ammonia.

2. THERMODYNAMIC ANALYSIS OF THE SYSTEM 1. S0I~ COI,LXCT01~S

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First, the balance of energy was calculated. The dimensions and characteristics of the refrigerator are shown in Figure 2. The walls are constructed of brick and concrete and the interior lined styrofoam sheets for insulation. The thickness of the insulation is l0 em. The roof consists of a concrete with gravel with a thickness of l0 em and styrofoam insulation of the same thickness. The construction of the floor is similar to the roof. The difference is that the insulation is placed between the concrete floor and the ground. The height of the refrigerator is 2.3 m. An interior temperattae of 1 C must be maintained. The exterior atmospheric temperature is considered to be 40 C, which is the average high temperature in the region. The total heat flow for the building was calculated to be 1.5 kw. 4.1m

-!-

_,_1.9

m_,

Figure 1. Components of the solar refrigerator.

- Refrigerant Tank. This tank keeps the refrigerant that is released as required. ARer this tank, there is a pre-cooler that causes the temperature of the ammonia (at liquid state) increase.

5. Om

4.5m

- Pre-cooler. This heat exchanger uses the ammonia that returns from the evaporator at high temperature in order to elevate the temperature of the refrigerant that comes from the refrigerant tank and passes to the expansion valve. - Expansion Valve. This valve regulates the refrigerant flow and lowers the pressure of the system before the ammonia enters the evaporator.

To

= 40 C

Figure 2. Dimensions of storage room. - Evaporator. Through this element, the refrigerant that absorbs the surrounding heat circulates so that the space around the evaporator becomes colder. The components of the system located in the low pressure region are the following: - Mixer. This component mixes the ammonia that comes from the Pre-Cooler with the water that has been separated from the ammonia in the high pressure region. Then, the mixture is absorbed by the Absorber.

Next, the refrigeration load was calculated. The system=s total refrigeration load originates from the following sources of heat: heat transmission between the difference in interior and exterior temperatures, load of the products which are intrOduced at room temperature and that must be cooled to the temperature of the refrigerator, air infiltration, supplementary loads caused by electric lights, motors, tools, and even people. The energy needed to reduce the temperature of the fish from 35 C to 1 C was calculated based on the conservation of 200kg of fish per day. The product load was found to be equal to 6.9 kW-h. The load caused by air infiltration was considered to be within the range of

421


design safety factor along with the supplementary load since the volume of the refrigerator is relatively small and the number of times the doors will be opened is minimal. In addition, the refrigerator does not have much supplementary equipment. The total load of the refrigerator was found to be 2.0kW.

calculations necessary for the fabrication of these parts were carried out. Table 1. Energy Balance Results. 1

:

:

Enrgy Op. Time Energy In Energy Out Type KVV h KW-h KW-h QGE 14.02 6 84.12 QEV 2.00 18 36.00 v~ol o.os s o.3~ 3222 -6~-6- ~ s.3z 6 --29.52 ....QRE 4.92 6 --46.62 QAB 2.59 18 --1 2.00 QSD 2.00 6 --1 20.36 Total i 1 20.43

The following characteristics were also calculated:

. . .

. . .

1) Total mass of ammonia, 100kg 2) Concentration of ammonia and water, 0.42 3) Condensation pressure at 34 C, 12.94 atm 4) Total mass of the solution, 961.29 kg 5) Energy required by the generator, 14.02 kW 6) Rectification energy, 4.927 kW 7) Condensation energy, 5.37 kW 8) Absorption energy, 2.5867 kW 9) Energy transferred by the weak solution, 2.00 kW 10) Work by the pump, 0.0523 kW

_ . _

The thermodynamic properties of the system are indicated and explained in Figure 3. Based on the energy balance, the system is considered to respect the First Law of Thermodynamics. The energy types entering the system, inputs, are the heat of generation (Q~3E), the heat of evaporation (QEv) and the work of the pump (WBO). The energy types going out the system, outputs, are the heat of condensation (Qco), the heat of rectification (QRE), the heat of absorption (Q~), and the heat transfered by the weak solution while in the weak solution tank (Qws). The energy balance results are shown in Table 1 where it is observed that input of energy, 120.43 KW-H, is almost the same as the output, 120.36 KW-H. "

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The Solar Collectors. After analyzing the efficiency and cost of evacuated tube solar collectors manufactured by various companies the decision was made to use Suntube which is designed to provide optimal efficiency for the collection of solar energy at a minimum cost (Nipon Electric Glass, 1990). It allows a 30% reduction in the required collection area. In addition it has a long service life due to its closed construction which effectively prevents problems such as corrosion and leaks. The solar collector tube may be installed horizontally which not only permits the use of virtually any part of the surface, but also assures that the collector can be integrated into the existing building. Due to its e ~ i a l layer it has a high rate of absorption (0.91 minimum) and a low rate of emissions (0.12 maximum). Figure 4 show a diagram with which the efficiency can be determined as a function of irradiation, air temperature, and the average operating temperature of the fluid. The area of the solar collectors required by the system was determined based on the quantity of energy demanded and the mass flow of required heat exchange fluid. The energy demand of the generator is 16 KW ( considering a safety factor of 15%). The quantity of heat exchange fluid required can be calculated based on the proposed entrance and exit temperatures of the oil in the generator. The oil used is Marlotherml. The average specific heat is calculated to be 0.825 kcal/kg C. For 16 KW, a result of 1667.87 kg/h is obtained. Since the average density of Marlotherml (110-120 C) is 912 kg/m2, the volume required is calculated to be 1819.5 l/h.

18 18 18 18 18 18 s

18 18 _[ ......................

For the thennic and mechanical analysis, the main components of the system are the following: Solar Collectors, Generator, Separator-Rectifier, Condenser, Condensation Tank, Cooler, Evaporator, Absorber, Strong solution tank, Mixer, and Weak Solution Tank. The other components of the system are considered to be complementary equipment.

i___i

Figure 3. Results of thermodynamic analysis. 3. THERMIC AND MECHANICAL ANALYSIS This section examines the thermic and mechanical characteristics of the elements of the system. For the selection of these elements, existing commercial equipments is considered first. For elements that are not available commercially, design


422

Length without fins (cm): Total length (m): Fins per inch (2.54 cm): Height of fins (cm): Type of fins: Tube material: Number of Tubes:

.11.oo ~

K

60

zo ~

" (c)

~

4t0

..$

'" -

I

\

"x, "%

Figure 4. Solar collectors efficiency. To determine the efficiency of the solar collectors using Figure 4, it is necessary to know the temperature difference and the solar radiation. The temperature difference was calculated to be 83 C. The average solar radiation in W/m2 for the region studied was obtained l~om a manual of the Institute of Engineering of the UNAM ACalculo de la Radiation Solar Instantanea en la Republica Mexicana`@ (Fernandez, 1990). For an average of six hours of effective solar radiation, the value of 597.7 kcal/m2h. In this case the efficiency of the solar collectors is E = 0.4. The energy absorbed is 278 W/m2. This result in a total absorption area of 57.55 m2 to satisfy the demand of 16 KW, or 32 collectors arranged in four rows of eight. The use of oil in the solar collectors causes a drop in pressure which was calculated to be 67.87 lb/pig 2 This drop in pressure requires a 3/4hp motor. The Generator. A high pressure generator with carbon harden steel tubes and fins was selected because of the use of a solution containing ammonia. The Separator-Rectifier. Due to the difficulty of finding a separator-rectifier that is able to cool the vapor to allow all of the water to condense and produce a vapor that is pure ammonia, the decision was made to design it and order its construction by an industrial shop. The separator consists of a distillation column of carbon hardened steel. The nucleus of the column has a series of washers the facilitate the separation of the vapor. The Condenser. A heat exchanger of armor plating and tubes was chosen since a heat exchanger of double tubes required a length of nearly 30 meters. The Refrigerant Tank. The condensation storage tank must be able to store 170 1 of ammonia along with a addition 15% of the gases that are not able to be condensed or 25.51. The Precooler. The precooler of the liquid refrigerant uses 1.53 m long coiled tubing with a diameter of 5.5 mm in a carbon hardened steel tube 0.5 meters long with a diameter of 6.5cm. The Evaporator. An characteristics was chosen:

evaporator with the following

Finned tubing: Fins welded with high l~equency resistance in a helicoidal form Exterior diameter (cm): 3.34 Wall thickness (cm): 0.33 Length with fins (m): 4.89

10.16 4.99 5 1.90 Solid I-IF Carbon hardened steel 3

The absorber. The strong solution storage tank must guarantee the storage of 1,137.5 1 of a solution of ammonia and water at a pressure of 3 arm and a temperature of 34 C in addition to 10 % of the non condensable gases (113.75 1) for a total of 1,252.25 1. It is a carbon hardened horizontal cylinder 1.22 rots long. The mixer. A maximum of 200 kg of fish in ice may be introduced to this room daily, up to a total capacity of 2 tons. The lowest temperature the evaporator reaches is -10 C, the high and low system pressures are 13.4 arm and 2.87 arm respectively. The refrigerant-absorbent mixture is ammonia and water, where the refrigerant is ammonia. The design of this system requires six effective solar hours to generate the refrigerant needed by the refrigerator to work eighteen hours daily. Evacuated tube solar collectors are used. A photovoltaic system is also required to power the principle electrical components, the motors of the hydraulic system, each with a different capacity. The equipment is as followed: a pump for oil, 3/4 h.p., 750 W, operating 6 hours per day; a high capacity pump for water, 500 W, operating 6 hours pers day; a pump for NH3-H20, 1/4 h.p., 250 W, operating 6 hours per day; and a low capacity pump for water, 1/8 h.p., 125 W, operating 12 hours per day. The total amount of energy required is the sum of the potentials multiplied by the operating hours, 10,500 W. To detemdne the energy total that must be provided, it is also necessary to consider the energy that the components of the photovoltaic system consume. This self-consumption energy can be calculated with the energy efficiency of the elements of the system. The values are as follows: cmrent invertor 95%, charge controller 95%, lead acid batteries 90%. The total energy that must be furnished by the photovoltaic system is 12, 927.054 W. The peak potential, in KWp, is proportional to the average total energy required per day, over the average solar radiation per day, expressing the solar radiation in KWh/day. Therefore, the peak potential is 2,5854 KWp. The selected electric equipment works within the range of 115-125 volts. For calculations a nominal voltage of 125 volts is used. Therefore, the current demanded by the system is 20.68 amp. The proposed batteries are lead acid 12.5 volts DC, the recommended charging voltage is 14.5 volts. In addition, if the nominal voltage required is 125 volts, it is necessary to connect in a series 10 batteries of 12.5 volts. moreover, to charge 10 batteries in a series 145 volts are require& If the batteries are 12.5 volts/200 amp-h and if a battery bank with 3- day autonomy is desired, the bank has to be three parallel lines with a series of 10 batteries in each line. The proposed battery is made by SIEMEN with a potential of 53 W, 17 volts and 3 amps. The photovoltaic arrangement needed to provide 145 volts and 20.68 amps consists of seven parallel line with nine modules in a series in each one of them. Overall, 63 modules are necessary to supply the required energy. 4. COST ANALYSIS In this section a cost analysis which compares the solar absorption refrigerator designed for this project for which the


only source of energy is solar with a conventional vapor compression refrigerator which uses a gasoline-powered electric generator was carried out. To compare the cost effectiveness of this solar refrigerator with a vapor compressor refrigerator of the same capacity, the following was considered: the vapor compression refrigerator requires electricity generated by internal combustion plant. This uses fuel whose price increases at an annual inflation index of 9.1% (average in Mexico for the last 10 years). The period of comparison is twenty five years with a minimal attractive rate of return (MARR) of 4.5% per year. All monetary values are stated in U.S. dollars since it is a more stable currency. The approximate cost of the solar refrigeration system was obtained using the prices of the components of the system: 32 solar collection modules, $17,280.00; oil pump, $540.00; generator $3,270.00; separator-rectifier, $1,166.67; condenser, $1,855.26; condensation storage tank, $933.63; pre-cooler, $318.34; evaporator $2,112.18; absorber, $1,985.53; strong solution storage, $2,235.60; mixer $2,485.00; weak solution storage $2,005.60; expansion valve (refrigerant), $135; expansion valve (weak solution), $65.00; pump (solution) $375; high capacity pump (water), $440.00, low capacity pump (water), $214.00; cooling tower, $2,471.73; photovoltaic system, $55,113.34. The total cost is $95,001.88 U.S. Dollars. The cost of the equipment necessary for the connections is not included, neither is the cost of the fluids. The price of the equipment does not include installation. The cost of the solar energy is $0.00 dollars. The cost of a conventional system with the same characteristics as the solar refrigeration system is $7,991.44 U.S. dollars. However, the replacement of equipment has to be considered. The gasoline powered generator which its cost is $960.00 has to be replaced every five years. The compressor unit and other components which their costs add up to $4,000.00 have to be replaced every 10 years. The cost of the gasoline required to operate the generator must also be calculated. It is estimated that it will consume 1.5 liters per hour of operation and run 9 hours a day for 25 years. Initially, solar energy refrigeration is more monetarily expensive, but less expensive ecologically than conventional refrigeration. However, at 23 years of operation they become the same monetarily. Beyond 23 years, conventional refrigeration is more expensive as shown in Figure 5.

423

Cost Break-Even Point

1l l . l l m Im tl am ,m 0

q.,n

II.II r ql W

u.,

O J~ ~1~

I1.11

J t

)

5

I'

*J

tl

1)

15

t?

I d) 21

~:) ~:S

years -------

Conventional

Solar

[

Figure 5. Costs of the conventional and solar refrigerators.

5. CONCLUSIONS The construction of a solar refrigerator to conserve sea products is technologically feasible and it can be an alternative to provide refrigeration to rural communities which do not have electric service. Although the initial cost of the solar refrigerator is relatively high and it makes this alternative less attractive economically, it could be an acceptable solution since the cost of the conventional and solar refiigerators become the same at approximately 23 years of operation with a MARR of 4.5%. In addition, the costs of the solar refrigerator can be reduced significantly if this equipment is mass produced. Many of its components can be improved and standardized in the future which will lower the costs of the solar refrigerator. REFERENCES

Bogart Marcel (1992) Ammonia Absorption Refn'geration in Industn'alprocesses, 1st edn, pp. 5-15. GULF, New York. Fernandez Zayas Jose (1990) ACalculo de la Radiacion Solar Instantanea en la Republica Mexicana@ Instituto de Ingenieria UNAM Serie No. 472 Huacuz Jorge (1992) AEstudios de Refrigeration Solar@ Boletin del Instituto de Investigaciones Electricas, Vol. 6 Issue No. 2. lnstituto de Refi'igeracion de Aire Acondicionado (1989) Manual de Refrigeracion y aire Acondicionado, 2nd edn, pp. 213-218. Prentice Hall, Mexico D.F. Montgomery Richard (1991) Energia Solar. Seleccion del Equipo, Instalacion y Aprobechamiento, 1st edn, pp.86-90. LIMUSA, Mexico D.F. Nipon Electric Glass (1990) ATechnical Data for NEG Evacuated Tube Solar Collector and Solar Collector Module LD-2800/DP62800@ Nipon Electric Glass Co., Ltd.

424


DEMONSTRATION OF A NEW TYPE OF ICPC IN A DOUBLE-EFFECT ABSORPTION COOLING SYSTEM

Roland Winston and Joseph J. O'Gallagher Enrico Fermi Institute, University of Chicago 5640 South Ellis Avenue, Chicago IL 60637 USA Telephone: 1-773-702-7756, FAX: 1-773-702-7756 E-mail: [email protected]

William S. Duff Dept. of Mechanical Engineering, Colorado State University Fort Collins CO 80523 USA Telephone: 1-970-493-1321, FAX: 1-970-495-0657 E-mail: [email protected], edu

Tom Henkel and Julius Muschaweck Solar Enterprises International, c/o Richardson Electronics, 40W267 Kesshnger Road, LaFox IL 60147, USA Telephone: 1-630-208-2577, E-mail: [email protected]

Rich Christiansen Ohio State University

Jim Bergquam Sacramento State University

Abstract- Recently two new technologies, a new ICPC solar collector design and the solar operation of a double effect chiller, were commercially demonstrated for the first time. A new family of ICPC designs was developed which allows a simple manufacturing approach to be used and solves many of the operational problems of previous ICPC designs. A low concentration member of this family of designs that requires no tracking was used in the demonstration. For the demonstration, an off-the-shelf 20 ton double effect direct fired absorption chiller was modified to operate with hot water. The new ICPC design and double effect chiller was able to produce cooling energy for the building using a collector field that was about half the size of that required for a more conventional collector and chiller.

1. INTRODUCTION Two new technologies, a new ICPC solar collector and the solar operation of a double effect (2E) chiller, were commercially demonstrated for the first time. In early 1998 the collector and a 2E chiller were installed in an office building in Sacramento, California. This paper describes the demonstration project and reports on component and system performance. Research on Integrated Compound Parabolic Concentrator (ICPC) evacuated solar collectors has been going on for more than twenty years. See (Garrison, 1979) and (Snail et al, 1984). With a good selective surface, a non-tracking ICPC evacuated solar collector can provide highly efficient daily collection of solar energy at temperatures up to 250C. A new family of ICPC designs was recently developed by researchers at the University of Chicago and Colorado State University for Solar Enterprises International. These designs allow a relatively simple manufacturing approach to be used and solve many of the operational problems of previous ICPC designs. A low concentration member of this family of designs that requires no tracking was used in the demonstration. The 336 tube ICPC collector away was fabricated for the demonstration by Solar Enterprises International. This design and the fabrication approaches are described in (Duff et al, 1997) and (Winston et al, 1997). Double effect absorption chillers have been commercially available for almost ten years. These 2E absorption chillers

require substantially higher operating temperatures (around 155C) than single effect (1E) chillers (around 80C). However, 2E chillers produce nearly twice as much cooling for the same energy input. Moreover, ICPC collectors operate as efficiently at these higher temperatures as do more conventional collectors at lower temperatures. Thus, with the advent of the new ICPC design, it became possible to produce cooling with a 2E chiller using a collector field that was about half the size of that required for a 1E chiller with the same cooling output. Various commercially available 2E chillers are either steam, hot water or direct fired. Steam and direct fired 2E absorption chillers are available in sizes ranging from 20 tons to 600 tom and above. However, at the time the project was initiated hot water fired 2E absorption chillers were only available in 280 ton and larger sizes. Thus, since the solar collector array produced heated water, an off-the-shelf 20 ton 2E direct fired absorption chiller had to be modified to operate with hot water. 2. ICPC SOLAR COLLECTORS The demonstration solar collector array consists of three banks of eight modules. Each module has 14 ICPC evacuated collector tube solar collectors. The evacuated tubes in the 106 m 2 (1150 It 2) aperture area array were oriented lengthwise north to south and sloped at 15 degrees to optimize solar energy collection during the air-conditioning season. Each of the 336 125 mm (5 inch) diameter 2.78 meter (9 foot, 1.5 inch) long soda-lime glass evacuated solar collector tubes


contains a thin wedge shaped fin absorber and matched CPC reflector running the length of the tube. The inside bottom half of the glass tube is silvered to form the CPC reflector. A 12 mm tube that is plugged at one end is bonded to the absorber. A small feeder tube is placed inside the 12 mm tube to allow fluid to flow into and out of the evacuated tube. The module manifolds are a concentric tube-inside-tube design as well. Two absorber orientations were produced, one with a vertical absorber fin and one with a horizontal fin. The north bank of the collector array consists of all horizontal fin tubes and the middle bank consists of all vertical fin evacuated tubes. The south bank contains a mixture of fin orientations. A crosssection of the two collector tubes illustrating the two orientations is shown in figure 1.

Prior to the demonstration, fourteen tube modules of each absorber orientation were tested on the Sandia National Laboratories two-axis tracking (AZTRAK) platform. This platform allows a non-tracking solar collector to be positioned at any angle between zero and 90 degrees in any orientation. See (Winston et al, 1997).

3. COOLING SYSTEM A McQuay 20 ton LiBr and water 2E chiller was modified for firing by solar hot water in the 145 to 165C range. See (CEC Final Report, 1998). Auxiliary energy was provided to the chiller generator or to storage from a 70 KW (240,000BTU/hour) hot water boiler. Depending on the state of the system at the time, auxiliary energy was used to boost the temperature of the hot water delivered to the generator or to increase the storage temperature. The collector loop and chiller generator loop were interfaced by a pressurized and insulated 3785 liter (1000 gallon) horizontal cylindrical storage tank. 4.

Fig. 1. New ICPC Design Used in the Sacramento Demonstration Showing both Vertical and Horizontal Fin Orientations

Fig. 2. Sacramento Demonstration Collection Efficiencies and Array Temperature for August 14, 1998.

425

DATA ACQUISITION SYSTEM

The primary data acquisition system (DAS) was provided by the National Renewable Energy Laboratory (NREL). The DAS consisted of thermocouples, thermopiles, flowmeters and a data recorder which could be remotely accessed and downloaded via modem. The DAS was put into full operation in August 1998.


426

5.

RESULTS

5.1 Collection Efficiencies Performance of the ICPC array has fulfilled the expectations of the Sandia testing. Conversion efficiencies in excess of 60 percent at insolation levels of 750 to 850 W/m2 and collector temperatures of 120 to 130C were attained. As can be seen in figure 2, results for a temperature range of 115 to 130C show close agreement with the Sandia tests (Winston et al 1997). Operating results are not yet sufficient to statistically substantiate superofity of one or the other absorber fin orientations. 5.2 Daily Performance The system was designed to allow the chiller to operate on solar energy most of the time. Several circumstances led to more input energy being required by the chiller and less solar energy being delivered to the chiller than had been originally planned: 1. The desired operation of the modified chiller was not attained until late in the cooling season. Consequently, before acceptable chiller operation was achieved, the chiller operated at a lower than design COP. Thus, more input energy was often required for a given cooling load. 2. Because perfecting the performance of the chiller took longer than expected, the automated valves and heat rejection system intended for summer 1998 operation were not installed. This resulted in manual operation of solar energy system, including daily coveting and uncovering the collector array, which often reduced the amount of solar energy collected. 3. Collection efficiency varied with flow rate to the collector array. At the lower flow rates that were used much of the time, collection efficiency was reduced. In figure 2 the flow to the collectors was reduced after 11:00 where a reduction in performance, as compared to the Sandia predictions, can be observed. 4. A leak developed in the storage tank manhole cover gasket. This leak was not fixed until after cooling operation had ceased and the tank could be drained and inspected. The leak caused a loss of pressure in the tank overnight. The subsequent vaporizing of the stored fluid produced greater storage energy losses and lower system start-up temperatures the next morning. In spite of these limitations, exceptional system performance was often achieved. Even on some days when these limitations were severe, the daily performance of the system was exceptional: 1. Daily energy collected when operating at around 150C often reached nearly 50 percent of the total solar energy falling on the collectors. 2. Daily energy delivered to the storage tank when operating at and above 150C often reached nearly 40 percent of the total solar energy falling on the collectors. During the peak insolation months next summer and

under automatic control, a high value of this figure of merit is expected to be achieved routinely. 3. The chiller was brought up to the design target late in the air-conditioning season. Consequently, it had an opportunity to operate efficiently only well below maximum design load. Under these circumstances, it has produced daily COPs near and above 1.2. 4. Despite a generator energy draw that was greater than design, the chiller often satisfied the building cooling load while operating entirely or almost entirely on solar energy. Automated operation of the system has already been partially implemented for 1999 and the collectors no longer need to be covered and uncovered. Longer-term fully automated operation of the system is planned for the summer of 1999. As compared to other solar air-conditioning options, current performance of the system has been exceptional. The building air-conditioning load has been consistently met with a solar collector field that is about half the size of that required by more conventional collectors using 1E chillers. Even better performance is expected in 1999.

5.

ACKNOWLEDGEMENTS.

Mary Jane Hale, Pramod Kulkami, Prab Sethi, Don Osbom, Cliff Murley and Ranji George have all been especially supportive in this project. The California Energy Commission provided the primary funding for the research. Matching funding were provided by the National Renewable Energy Laboratory, Sacramento Municipal Utility District, South Coast Air Quality District, Sun Utility Network, Bergquam Energy and Thermal Energy Systems Specialists. REFERENCES Garrison, J. D. (1979) Optimization of Fixed Solar Thermal Collectors, Solar Energy, v23. Snail, J. J., O'Gallagher and R. Winston (1984) A Stationary Evacuated Collector with Integrated Concentrator, Solar Energy, v33. Duff, William S, R. Duquette, Roland Winston and Joseph O'Gallagher (April, 1997) Development, Fabrication and Testing of a New Design for the Integrated Parabolic Evacuated Collector, Proceedings of the ASES/ASME Solar Energy Forum, Washington D. C. Winston, R, J. J. O'Gallagher, William S. Duff and Alberto Cavallaro (April, 1997) The Integrated Compound Parabolic Concentrator: From Development to Demonstration,


Proceedings of the ASES/ASME Solar Energy Forum, Washington D. C. "Demonstration of ICPC Solar Energy Collectors for Space Heating and Cooling Using a 2E Absorption Water Chiller" (December 1998) California Energy Commission Project 50095-018 Final Report by Bergquam Energy.

427


428

INDIRECT EVAPORATIVE COOLING THROUGH A CONCRETE CEILING Baruch Givoni, Sukanya Nutalaya UCLA, Dept. Of Architecture and Urban Design, 405 Hilgard Ave. Los Angeles, USA Tel: +1-310-8252067, Fax: +1-310-8258959, Email: [email protected]

Keywords: Hot Humid Climate, Passive Cooling, Indirect Evaporative Cooling, Cooling Towers, Ceiling Heat Exchanger. ABSTRACT - Two experimental cells were used in this study. One cell remained without any cooling (control). The other cell (test cell) was cooled by the system reported in this paper. The cooling system consisted of a pond of cooled by water circulating within a concrete ceiling, acting as a heat exchanger. The water was cooled by a shower at the top of an open shaft, re-circulated by a pump. The use of the ceiling as a heat exchanger facilitates convective down flow of the cooled air and maximizes the radiant exposure of the inhabitants to the cooled ceiling surface. The performance of the cooling system was tested under two levels of heat capacity: first in the original configuration (described below) and later with added internal mass (concrete blocks). In hot climates the indoor average temperature may serve as the best indicator of the thermal performance of a passive building. On the basis of the experimental data a formula has also been developed, expressing the indoor maximum temperature of the cooled cell as a function of the outdoor average air temperature. INTRODUCTION One of the passive cooling options available in a hot humid climate is indirect evaporative cooling, which lowers the temperature without raising the indoor absolute humidity. The cooling system described below provides a design solution of this option. When fine drops of water, having a very large surface area, are sprayed vertically downward like a shower from the top of an open shaft, the falling water entrain a large volume of air, creating an inertial air flow down the shaft. The water is collected in a small pond at the bottom of the shaft and is pumped back to the shower head. The evaporation from the free drops cools the water, as well as the air in the shaft, to a level close to the ambient Wet Bulb Temperature (WBT). As any evaporation from the droplets takes place in the free air stream, any type of water, brackish or even sea water if available, could be used for cooling in this system. In fact, no difference was found in the cooling performance of the system when tested with either fresh or sea water. This system enables both direct and indirect evaporative cooling of buildings. Direct evaporative cooling utilizes the humid cool stream created by the falling drops, which is directed into a building. Direct evaporative cooling is applicable mainly in hot dry regions (Givoni & A1 Hemiddi 1995). The cooled water in the pond can be used for indirect evaporative cooling, by circulating the water, in a closed loop, through a water-to-air heat exchanger inside the cooled space. The best position of the heat exchanger is above the cooled space, in order to facilitate convective down flow of the cooled air and to maximize the radiant exposure of the inhabitants to the cool surfaces of the heat exchanger. Indirect evaporative cooling is applicable mainly in hot humid regions.

Such cooling system with specialized heat exchanger (a radiator) was tested and reported previously (Givoni et al. 1995). However, a ceiling cooled by water flow in embedded tubes can also serve as an architecturally integrated cooling element. The performance of such a system is reported in this paper. The fact that the shower system provides the same cooling performance with sea water as with fresh water has profound implication for its applicability in desert coastal regions, where fresh water is in short supply while sea water availability is unlimited. In this respect this system can be applied in places where other evaporative cooling systems, which need high quality water, are not applicable. In hot climates the indoor average temperature may serve as the best indicator of the thermal performance of a passive building. Therefore, in the analysis of the experimental data, the main emphasis is on the indoor maxima which can be obtained with this cooling system. On the basis of the experimental data a formula has been developed, expressing the indoor maximum temperature of the cooled cell as a function of the outdoor average air temperature.

The Experimental Setup The experimental setup, at UCLA, consisted of two similar test cells (available from a previous study), each 1.2xl.2x3.6 meters. One cell was used as a control (without any cooling) and the second cell (Test-cell) was cooled. The walls and floors of the two cells were identical: insulated sandwiches of plywood. The roof of the control cell was (the original) insulated sandwich, while the roof of the cooled cell was modified. It was built of concrete, insulated to provide the same U value as the roof of the control cell. Budget limitations did


not enable to construct a concrete roof also for the control cell, so the thermal masses of the two cells were different. However, the Heat Loss Coefficients (UA) of the two cells, which are not affected by the difference between the heat capacities of the cells, were measured and were found to be very similar, 11.3 and 11.0 W/C, for the Control and the Test cells, respectively. Consequently, it was possible to calculate, from the differences of the indoor average temperatures of the two cells, the average diurnal cooling performance of the system. In the second stage of the study the same amount of additional mass, in the form of 169 small concrete bricks (5xlOx20 cm), was added to each cell. In the third stage, the two cells were ventilated during the night hours while the cooling system operated continuously. Thus, the cooling performance of the

Figure 1: Embedded plastic tube circulating cooled water Figure 2 shows the data acquisition system and the measuring points.

Figure 2: Data Acquisition System and Measuring Points

429

system was measured under three configurations of the cells: First in the original construction, then with the added mass, and later with night ventilation. The shower tower was 3.6 meter-high. The tower was made of 5 cm. (2") rigid (polystyrene) insulation and contained 4 heads of nozzles on the top of shower tower. The pond (123x123x76 crn) was built of polystyrene sheets. The inside of the pond was a water-proved plywood, with two layers of plastic sheets inside forming the actual pond. The construction of the concrete roof with the embedded plastic tubes is illustrated in figure 1. Concrete bricks, 5 cm thick, were laid over a thin steel plate. Then, plastic tube circulating the cooled water was installed, and a second (5 cm) layer of concrete was poured, to create a 10 cm thick roof.

430


Overall Performance of the System with Floating Condition Figure 3 (A to C) shows samples of the diurnal temperatures during the cooling in the original configuration (3a), with the added mass (3b) and with the night ventilation (3c).

Figure 3: Diurnal Temperature Patterns


431

Table 1 shows the averages of the maximum, minimum and average temperatures, measured in the different configurations of the study. In this paper only the average temperatures will be discussed in details. Table 1: Summery of Grand Average of Maximum, Minimum and Average Temperature Conditions A)

Control- No Added Mass

AVG.

DBT

Control

Cooled

16.7

19.0

19.2

MAX. 21.8 29.0 22.6 MIN. 10.1 10.1 14.8 23.5 23.7 20.0 B) C o n t r o l - With Added AVG. Mass MAX. 26.1 29.2 25.4 MIN. 15.2 18.6 21.7 AVG. 19.4 21.8 19.4 C) Cooling- No Added Mass MAX. 24.2 31.5 22.7 MIN. 15.0 15.1 17.0 25.8 22.7 22.6 D) C o o l i n g - With Added AVG. Mass MAX. 28.8 31.2 24.5 MIN. 17.7 21.5 20.7 27.4 24.6 AVG. 27.0 E) Cooling With Mass and Night Ventilation 34.0 27.0 MAX. 34.6 22.4 22.3 MIN. 20.9 Note - The cooling effect is calculated by (Control - Cooled) temperature drop

Average Temperatures Under floating conditions the average of the "test" cell was higher by 0.2 ~ then the control cell. With cooling in the original configuration it was lower by 2.4 o C. With the added mass the cooled cell's average was lower by 3.1 C and with night ventilation the temperature reduction was 2.8 C.

Maximum Temperatures As the two cells had different masses, their maxima, as well as their minima, can not be compared directly. However, the differences between the DBT and the cooled cell maxima in the different configurations can be compared. In the floating conditions without the added mass the maximum of the "test" cell was higher by 0.6 ~ then the outdoors' maximum. With added mass, but without cooling, the indoor maximum was lower than the outdoors by 0.7 C, reflecting the suppression of the indoor swing by the mass. With cooling in the original configuration the maximum of the cooled cell was lower by 1.5 ~ than the control. With the added mass the difference was 4.3 ~ and with night ventilation it was 7.6 ~

Minimum Temperatures Under floating conditions without the added mass the minimum of the "test" cell was higher by 4.7 ~ then the outdoors' minimum. With added mass, but without cooling, the indoor maximum elevation above the DBT was 6.5 ~ again reflecting the suppression of the indoor swing by the mass. With cooling

Effect

Pond

Pond Dep.

18.0 20.0 16.3 20.3 22.5 18.1 21.8 23.8 19.6

-1.4 -4.2 +1.3 -2.3 -6.3 +0.4 -5.2 -10.8 +1.3

-2.4 -8.8 +1.9 -3.1 -6.7 +0.8 -2.8 -7.0 +0.1

in the original configuration the minimum of the cooled cell was 2 ~ higher than the DBT, and with the added mass the elevation was 3 ~ With night ventilation the indoor minimmn elevation was 1.4 ~

The Cooling Effectiveness of the System Without cooling (floating conditions) the indoor averages of the two cells were almost identical and about 2.5 ~ above the outdoors' average, as a result of solar energy absorption in the envelope of the cells. With cooling, the indoor temperature of the cooled-cell was lower than that of the control. The cooled pond is the source of the cooling and thus, the lowest attainable average indoor average temperature would be that of the cooled water (Tpond) average. Consequently, the overall thermal effectiveness of the cooling system was calculated by the formula:

Effectiveness = (Tcontrol - Tcooled) / (Tcontrol -

Tpond) The experimental thermal effectiveness of the system, in the configurations without night ventilation, it was about 0.6. The average cooling energy provided by the system can be calculated from the differences between the average indoor temperature of the cells and the average DBT, and the cell's BLC:

Cooling- (Tcontrol- Tcooled) * BLC; (Whr/Day)


432

Formula Predicting the Maximum Temperature of the Cooled Cell

The indoor maximum in a given configuration is almost parallel to the pattern of the DBT average. Its relationship to the outdoors' maximum is weaker.

Figure 4 shows the daily indoor maximum and the DBT maximum, average and minimum temperatures, during the various periods with the different configurations. It illustrated the following relationships between the patterns of the indoor maximum and the DBT:

b) The day to day changes in the indoor maximum, especially in the configuration with the added mass, are smaller than the corresponding changes of the outdoors' average.

40

35

30

25 U

15

1

PL "q

The initial and boundary conditions are following: 9 0

1999 ISES Solar World Congress

Solar

Solar

Solar

Solar

Solar

Solar

Solar

SOLAR

Solar

Solar

Solar

Solar

Solar

1999

№1999

Solar

Third World Congress on Oxidation Catalysis

Third World Congress on Oxidation Catalysis

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

1999 ISES Solar World Congress

Solar

Solar

Solar

Solar

Solar

Solar

Solar

SOLAR

Solar

Solar

Solar

Solar

Solar

1999

№1999

Solar

Third World Congress on Oxidation Catalysis

Third World Congress on Oxidation Catalysis

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

Solar

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