Experimental and Numerical Analysis of a PCM-Supported ...

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productivities at any cooling water mass flow rate is coincidently the same as the rank of their thermal conductivities. The thermal conductivity of the packing media should not be less than 2W/m/K for better performance of the condenser. The role of MEHH is nearly absent or even negatively influencing the distillate rate at very low hot water mass flow rates and low temperature due to cooling effects on the air in the lower part of the evaporator. Both experimental and theoretical analysis indicated that increasing the specific surface area of the packing can compensate and outpace the role of MEHH, as the conductive packing has rather limited influence under such conditions. Therefore, when operating HDH plants under forced convection, industrial packing media which have high specific surface area and lower cost would be ideal candidates for the evaporator and condenser. However, conductive packing media may have a potential that it could be utilized to improve the natural flow characteristics in other important applications such as cooling towers. Moreover, the study revealed that the inlet hot water temperature and air to water mass flow ratio have rather stronger influence on the system performance than the thermal conductivity of the packing. The numerical analysis indicated that the optimum value of water to air mass flow ratio is lies between 1.3 and 1.5 for the evaporator and between 1.5 and 2 for the condenser. The results obtained from the parametric study on the complete solar HDH system indicates that the crucial parameters have been identified for practical purposes under real weather conditions. The ideal HDH performance is approached for: brine concentration factor (r c )=2, air to hot water mass flow ratio=1.5, cooling water to air mass flow ratio=2, thermal storage volume 3-5 m 3 per m 3 of daily fresh water production capacity depending on the local climatic conditions. The once-through flow arrangement in the external PCM thermal buffer with a flat plat collector has a limited efficiency when combined with the HDH plant in comparison with water storages with the same volumetric capacity. For optimal operation of the external PCM thermal buffer, the Stefan number must be small enough, as this means that there is a great amount of latent heat that can be stored (relative to sensible heat). Utilization of PCM in the thermal storage is recommended only for high energy density, and in applications where there is a narrow operational temperature change around the melting range. It is therefore of interest to look in future studies at the techno-economical aspects for different simulation scenarios and coupling setups of the external thermal buffer and concentrating solar power collectors with the HDH system. 179
References [1] Hudges C.N., Tompson T.I., Groh J.E., and Frieling D.H., Solar distillation utilizing multiple-effect humidification., U.S. Dept. of Interior Research and Development, Report No. 194, 1966. [2] Al-Hallaj S. and Selman J.R. (2002). A comprehensive study of solar desalination with humidification-dehumidification cycle. MEDRC Series of R&D Reports, Project: 98-BS-032b. [3] Al-Hallaj S., Parekh S., Farid M.M. and Selman J.R. (2006). Solar desalination with humidification-dehumidification cycle: Review of economics. Desalination 195: 169-186. [4] Müller-Holst H., Engelhardt M. and Scholkopf W. (1999). Small-scale thermal seawater desalination simulation and optimization of system design. Desalination 122: 255-262. [5] Müller-Holst H., Engelhardt M., Herve M. and Scholkopf W. (1998). Solar thermal seawater desalination systems for decentralized use. Renewable Energy 14(1-4): 311-318. [6] Gajbert H. and Fiedler F. (2003). Solar combisystems, State of the Art Report. In: Solar Energy State of the Art. Denmark Tekniske Universitet, Internal Report BYG·DTU SR-03-14 2003, Editors: Furbo S., Shah L.V. and Jordan U., ISSN 1601 – 8605. [7] Knudsen S. and Dennis M. (2003). Thermal storage for small solar heating systems, state of the art report, In: Solar Energy State of the Art. Danmark Tekniske Universitet, Internal Report BYG·DTU SR-03-14 2003, Editors: Furbo S., Shah L.V. and Jordan U., ISSN 1601 – 8605. [8] Nagano K. (2003). Development of a high performance solar hot water supply system using evacuated solar tubes and latent thermal energy storage. Proceedings of International Solar Energy Society, Gotegorg, 2003. [9] Hassabou A., Spinnler M., Hanafi A., Polifke W. Experimental analysis of PCMsupported humidification-dehumidification desalination system. IDA Word Congress–Atlantis, The Palm-Dubai, UAE November 7-12, 2009. REF:IDAWC/DB09-289. [10] Schlunder E.U., Equivalence of one- and two-phase models for heat transfer processes in packed beds: one-dimensional theory, Chem. Eng. Sci. 30 (1975) 449-452. [11] Kaviany M., Principles of Heat Transfer in Porous Media, second ed., Springer, New York, 1995, pp. 391:424. [12] Arkar C., Medved S. Influence of accuracy of thermal property data of a phase change material on the result of a numerical model of a packed bed latent heat storage with spheres. Thermochimica Acta 438 (2005) 192-201. 180
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Technische Universität München In
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It is my pleasure to thank Prof. Th
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Contents List of figures ..........
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5.3 Functional description ........
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List of Figures Figure 1.1: Cost br
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78cm; Top: Evaporator, Bottom: Cond
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xii
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m Mass flow rate; [kg.s -1 ] m eva
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ed c coll cond cw d d e or eff evap
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1 Introduction The first part of th
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summarized in figure (1.2). Thermal
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1.3.2 Selection criteria of desalin
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of energy storage or hybridization
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Desalination development focuses on
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Introduction of the indirect type s
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2 Literature Survey 2.1 Introductio
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There is actually a commercial prod
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Q m H c ( T T1) c ( T 2 T ) (2.2)
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(solid-liquid) have a storage densi
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(iii) a sharp melting temperature.
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melt and freeze without segregation
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However, the stored heat is less va
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parameters on the heat transfer and
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categories. The first category is b
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to 550 kWh/m 3 of distilled water a
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The enthalpy and humidity of the sa
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Air mass flow rate [kg/h] GOR [-] 1
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On the other hand, the efficiency o
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2.8.4.2 Effect of air to water mass
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distillate rate. The system was not
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were; EnviPac Gr.1, 32 mm, and VFF
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Figure 2.15: Height versus air temp
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The performance of the condenser is
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Klausner and Mei [136] described a
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The impact of the collector cost ca
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exchanger and energy storage in one
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3 Scope of the Study This chapter p
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packing where it gets in direct con
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In fact there is a tradeoff between
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The energy flow between all and eac
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4 Theoretical Analysis and Modeling
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The two-energy equation models whic
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solid-liquid phase change inside th
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assumed by conduction in both axial
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the air pressure drop ∆P per unit
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evaporation rate per unit volume of
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where β t , β c , ρ ref , c ref
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the viscous transport and introduce
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liquid and solid phases, respective
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4.2.3.3 Liquid and gas hold-ups Liq
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Several models and approaches for m
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where G=ρU d /ε g is the pore mas
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k k eff s 1 3 2 (4.55) This cor
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uncertainties associated with the f
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1 2 (4.62) d1 capp c2 c1 c2 1
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[58] evaluated the cooling tower ai
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1 exp NTU 1 C r, cond cond (4
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A HE QHE U T HE LM (4.81) Where Q
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c M f b (4.86) M b S S f where r
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that form the study logical foundat
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5 Experimental Analysis This chapte
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5.2 Measuring Techniques K-type the
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The product condensed water was mea
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operating limits. In light of the l
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In the last two cases, the inlet ho
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performance and its threshold value
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It could be interpreted that this p
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Evaporator Condenser Figure 5.8: Av
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within the inaccuracy margins. The
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The last two cases of boundary cond
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natural air flow in the system. In
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6 Model Implementation and Validati
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x t u x u 0 0 , (6.5) and for a s
Page 148 and 149: Next time step validation were the
Page 150 and 151: The evaporator and condenser both w
Page 152 and 153: Figure 6.2: Evolution of inlet and
Page 154 and 155: Figure 6.4: Evolution of inlet and
Page 156 and 157: temperature is at its maximum level
Page 158 and 159: Bottom Top Figure 6.7: Evolution of
Page 160 and 161: 7 Parameters Analysis of the Evapor
Page 162 and 163: Figure 7.1: Effect of packing size
Page 164 and 165: e 0.45 kPa under natural convection
Page 166 and 167: aspect ratio under constant bed vol
Page 168 and 169: Figure 7.7: Effect of inlet hot wat
Page 170 and 171: well as the time required to reach
Page 172 and 173: Figure 7.10: Effect of packing medi
Page 174 and 175: 40°C respectively. This gives flex
Page 176 and 177: the condensate rate, productivity f
Page 178 and 179: esult, rather than PCM media which
Page 180 and 181: Figure 7.18: Effect of PCM thermal
Page 182 and 183: 8 Optimization of HDH plant In this
Page 184 and 185: The MATLAB model describes how the
Page 186 and 187: less steep than that of the distill
Page 188 and 189: 8.3.4 Effect of hot water flow rate
Page 190 and 191: storage materials and thus it can h
Page 192 and 193: for latent heat storage would be ob
Page 194 and 195: When T m becomes sufficiently highe
Page 196 and 197: 9 Summary and conclusions 9.1 Summa
Page 200 and 201: [13] Vafai K., Sozen M. An investig
Page 202 and 203: [48] Crone S., Bergins C., and Stra
Page 204 and 205: [79] Tiwari G.N. and Yadav Y.P. (78
Page 206 and 207: [109] Mehling H., Hiebler S. and Zi
Page 208 and 209: [139] Belessiotis V, Delyannis E (2
Page 210 and 211: Appendices Appendix A A1. PCM Therm
Page 212 and 213: A3. Constants and Thermo-physical p
Page 214 and 215: From the above correlation, the con
Page 216 and 217: The Ergun analogy factor J h repres
Page 218 and 219: Equation (C.7) represents an overal
Page 220 and 221: Appendix D D1. Cooling tower theory
Page 222 and 223: A m w = the plan or cross sectiona
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