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

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In fact there is a tradeoff between direct heat and mass transfer at the liquid-gas interface and the energy flow from liquid to gas across the solid medium which creates the multiple effects of heating and humidification (MEHH) in the evaporator and similarly multi-effects of cooling and dehumidification in the condenser (MECC). 3.3.2 Macro scale level The change of PCM temperature and enthalpy is both space and time dependent as a result of the transient conduction inside the PCM beads, and thermal stratification in water and gas phases along the packing height. However, due to the stratification of the heat source temperature (i.e. the hot water), stratification exists along the packing height in the PCM layers. This in turn creates a multiple effect heating/humidification or cooling/dehumidification in the evaporator and condenser respectively. The MEHH in the evaporator is illustrated in figure (3.3). The conductive packing media proposed theoretically serve as temporary heat exchangers that virtually increase the sensible and latent heat transfer interfacial area between water and air and create a multi-effect heating and humidification and multi-effect cooling and dehumidification along humid air passages in the evaporator and condenser, respectively. Such combined effects of creating additional parallel and serial interfaces for heat and mass transfer along the packing heights would increase the air carrying capacity for water vapor and maximize evaporation and condensation rates in the closed air loop cycle. Figure 3.4: Dependency of air moisture carrying capacity on the temperature When unsaturated airflow gets in contact with hot saline water, a certain quantity of water vapor diffuses from water into the air, which results in a temperature reduction of water as well as a higher salt concentration. The maximum amount of water vapor is limited by the saturation conditions of air, which is mainly dependent on its temperature under constant pressure. Once the air is saturated, it can not carry more water vapor unless its temperature is further increased. As a well known fact, the air carrying capacity for water vapor has an exponential dependence on its temperature. In order to achieve certain humidity content, the air temperature could 59
e increased at once to the corresponding temperature (as represented by the dashed arrow on figure (3.3) or in a multiple stages of heating and humidification as depicted by the successive red and blue solid arrows on figure (3.3). In the first case, the heat source is required to be available at a high temperature, which in turn will be reflected on increasing the collector area and hence the principle and operating costs. In the latter case the heat source can have a lower quality, such as a waste heat source or simple flat plate solar collector. In order to achieve the maximum amount of water vapor in the evaporation chamber, the incoming water temperature should be at a temperature high enough so that the maximum amount of water can be evaporated to the surrounding air. The maximum temperature of the incoming water cannot exceed 85°C due to the high material costs associated with such high temperatures. Moreover, the carrying capacity of moist air approaches its highest at approximately 87 °C (see figure 3.4). Therefore it is not expected (in a water heated HDH system) that the air reaches its theoretical maximum carrying capacity at the outlet of the evaporator, since there is always a temperature difference between the hot and cold streams in a heat and mass exchanger. This leads to the fact that if the air is to be heated alternatively or simultaneously with heating water, it may be theoretically possible to reach the maximum carrying capacity for humidity content. However, solar air collectors have low thermal efficiencies, especially at higher outlet air temperatures, considerably lower than that of the water solar heaters. Thus heating air requires larger solar collector fields, which will add a considerable fraction to the specific desalinated water cost. The resulting prohibitive cost renders this option almost impossible. This particular problem inspired the concept of utilizing the a conductive packing media such as PCM capsules in the evaporator to create the MEHH locally (as illustrated by figure (3.3), without the need to extract air outside the evaporator at intermediate points, heat it in external air heaters and return it back to the same point, which would neither be technically nor economically feasible. Moreover, on the macro scale the natural draught is driven by the temperature and density gradients along and between both the evaporator and condenser. The principle behind natural convective heat transfer in the evaporator and condenser is that systems which are influenced by gravity and have materials with different densities are subject to different rates of heat transfer due to uneven heating (i.e. stratification) of the materials present. As a result, particles which are less dense such as humid air rise up and displace the colder more dense air particles, which sink down. Such phenomenon induces air currents between the evaporator and condenser and can be utilized for self-propulsion of the air into the HDH system. Existence of thermal stratification and the MEHH and MECC could have an influence on boosting the natural convection flow, which in turn increases the distillation rate. 60
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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
Page 28 and 29: Desalination development focuses on
Page 30 and 31: Introduction of the indirect type s
Page 32 and 33: 2 Literature Survey 2.1 Introductio
Page 34 and 35: There is actually a commercial prod
Page 36 and 37: Q m H c ( T T1) c ( T 2 T ) (2.2)
Page 38 and 39: (solid-liquid) have a storage densi
Page 40 and 41: (iii) a sharp melting temperature.
Page 42 and 43: melt and freeze without segregation
Page 44 and 45: However, the stored heat is less va
Page 46 and 47: parameters on the heat transfer and
Page 48 and 49: categories. The first category is b
Page 50 and 51: to 550 kWh/m 3 of distilled water a
Page 52 and 53: The enthalpy and humidity of the sa
Page 54 and 55: Air mass flow rate [kg/h] GOR [-] 1
Page 56 and 57: On the other hand, the efficiency o
Page 58 and 59: 2.8.4.2 Effect of air to water mass
Page 60 and 61: distillate rate. The system was not
Page 62 and 63: were; EnviPac Gr.1, 32 mm, and VFF
Page 64 and 65: Figure 2.15: Height versus air temp
Page 66 and 67: The performance of the condenser is
Page 68 and 69: Klausner and Mei [136] described a
Page 70 and 71: The impact of the collector cost ca
Page 72 and 73: exchanger and energy storage in one
Page 74 and 75: 3 Scope of the Study This chapter p
Page 76 and 77: packing where it gets in direct con
Page 80 and 81: The energy flow between all and eac
Page 82 and 83: 4 Theoretical Analysis and Modeling
Page 84 and 85: The two-energy equation models whic
Page 86 and 87: solid-liquid phase change inside th
Page 88 and 89: assumed by conduction in both axial
Page 90 and 91: the air pressure drop ∆P per unit
Page 92 and 93: evaporation rate per unit volume of
Page 94 and 95: where β t , β c , ρ ref , c ref
Page 96 and 97: the viscous transport and introduce
Page 98 and 99: liquid and solid phases, respective
Page 100 and 101: 4.2.3.3 Liquid and gas hold-ups Liq
Page 102 and 103: Several models and approaches for m
Page 104 and 105: where G=ρU d /ε g is the pore mas
Page 106 and 107: k k eff s 1 3 2 (4.55) This cor
Page 108 and 109: uncertainties associated with the f
Page 110 and 111: 1 2 (4.62) d1 capp c2 c1 c2 1
Page 112 and 113: [58] evaluated the cooling tower ai
Page 114 and 115: 1 exp NTU 1 C r, cond cond (4
Page 116 and 117: A HE QHE U T HE LM (4.81) Where Q
Page 118 and 119: c M f b (4.86) M b S S f where r
Page 120 and 121: that form the study logical foundat
Page 122 and 123: 5 Experimental Analysis This chapte
Page 124 and 125: 5.2 Measuring Techniques K-type the
Page 126 and 127: 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
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Next time step validation were the
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The evaporator and condenser both w
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Figure 6.2: Evolution of inlet and
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Figure 6.4: Evolution of inlet and
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temperature is at its maximum level
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Bottom Top Figure 6.7: Evolution of
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7 Parameters Analysis of the Evapor
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Figure 7.1: Effect of packing size
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e 0.45 kPa under natural convection
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aspect ratio under constant bed vol
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Figure 7.7: Effect of inlet hot wat
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well as the time required to reach
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Figure 7.10: Effect of packing medi
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40°C respectively. This gives flex
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the condensate rate, productivity f
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esult, rather than PCM media which
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Figure 7.18: Effect of PCM thermal
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8 Optimization of HDH plant In this
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The MATLAB model describes how the
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less steep than that of the distill
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8.3.4 Effect of hot water flow rate
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storage materials and thus it can h
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for latent heat storage would be ob
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When T m becomes sufficiently highe
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9 Summary and conclusions 9.1 Summa
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productivities at any cooling water
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[13] Vafai K., Sozen M. An investig
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[48] Crone S., Bergins C., and Stra
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[79] Tiwari G.N. and Yadav Y.P. (78
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[109] Mehling H., Hiebler S. and Zi
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[139] Belessiotis V, Delyannis E (2
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Appendices Appendix A A1. PCM Therm
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A3. Constants and Thermo-physical p
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From the above correlation, the con
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The Ergun analogy factor J h repres
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Equation (C.7) represents an overal
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Appendix D D1. Cooling tower theory
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A m w = the plan or cross sectiona
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