An Investigation of a Falling Film Desiccant Dehumidification/ Regeneration Cooling System

Transcription

1 Heat Transfer Engineering, 28(2): , 2007 Copyright C Taylor and Francis Group, LLC ISSN: print / online DOI: / An Investigation of a Falling Film Desiccant Dehumidification/ Regeneration Cooling System SCOTT FEYKA Remec Defense and Space, San Diego, California, USA KAMBIZ VAFAI Department of Mechanical Engineering, University of California at Riverside, Riverside, California, USA The present work outlines a liquid desiccant refrigeration cycle with a significantly higher efficiency, as compared to a traditional chiller cycle operating between the same two thermal reservoirs. The present investigation includes a general cooling design under nominal values. A parametric study enables the determination of the overall cooling and humidity levels as compared to the ambient conditions. Typical cooling achieved by this analysis is approximately 17 C below the ambient temperature, and the relative humidity of the inlet air can be reduced up to 22 percent. The proposed cycle yields a percent higher coefficient of performance as compared to a chiller cycle operating between the same temperatures. The proposed dehumidifier displays close agreement when compared to a standard cross-flow heat exchanger and the general trends and performance of this cycle compare well with previous studies on falling film dehumidifiers. INTRODUCTION Desiccant dehumidification has been studied extensively during the past several years as a way to dramatically reduce energy consumption of a vapor compression cooling cycle. In this type of system, the water vapor is removed from the air as it is routed to the evaporator in an effort to reduce the cooling load on the air conditioning system. This modified air conditioning cycle, known as a hybrid cooling system, has been shown to help increase the coefficient of performance (COP) of the cycle while reducing evaporator and condenser sizes. Analysis of a hybrid system [1] has revealed that the compressor power consumption can be reduced by 25% with an evaporation and condensation area reduction of 34% as compared to a standard vapor compression cycle. Two types of liquid desiccant structures have been studied. The first involves a packed bed design first proposed by Lof [2] in 1955 using a triethylene glycol solution as the desiccant. This design incorporates a porous packing material, sprinklers, and a solution pump. The concentrated liquid desiccant is sprayed over the packing material, which forms a thin film along the bed. Address correspondence to Prof. Kambiz Vafai, Department of Mechanical Engineering, University of California at Riverside, A363 Bourns Hall, Riverside, California, USA. vafai@engr.ucr.edu Similar to a cooling tower design, the inlet air is forced through the packed bed structure in a counter-flow direction, where the weight of the desiccant enables collection at the bottom of the bed. The packed bed structure has been extensively studied by Factor and Grossman [3] with exit condition predictions using a one-dimensional model. Their analysis included a set of differential equations derived from energy and mass balances along a differential slice of the packed column. A finite difference scheme was used to solve these sets of differential equations, revealing the temperature and humidity profiles along the bed. Along the same lines, Radhwan et al. [4] solved a set of four first-order differential equations based on energy and mass balances along a differential slice. They observed that performance enhancements occur by an increase in height and transfer area of the packed bed. Falling film dehumidification was developed in response to the relatively large pressure drops and occasional inconsistent mixing of a packed bed structure. In addition, the falling desiccant film helps to purify and filter the conditioned air from large organic molecules and prevents the growth of mold, mildew, bacteria, and other airborne microorganisms [5]. The present work is based on a numerical and experimental study of a cross-flow fin tube falling film desiccant dehumidifier designed and analyzed by Park et al. [6]. The results of this two-dimensional model under isothermal channel flow 163

2 164 S. FEYKA AND K. VAFAI Figure 1 Proposed desiccant cooling cycle under nominal operating conditions. indicate enhanced performance under lower airflow using triethylene glycol as the desiccant. This effect was primarily due to increased contact time between the air and desiccant. The model was validated from experimental data for a cross-flow dehumidifier using both triethylene glycol and lithium bromide as the desiccant. A comprehensive and detailed simulation of a cross-flow dehumidifier by Ali et al. [7] confirmed Park s results with additional performance recommendations. Ali et al. s [7] study demonstrated that further dehumidification and cooling could be achieved by increasing the height and length of the channel while decreasing the channel width. In addition, they also demonstrated marginal heat and mass transfer enhancements by the addition of conductive ultra fine particles within the liquid desiccant [8]. Ali and Vafai [9] established the optimal airflow direction and plate inclination effects on a falling film dehumidifier, and found that optimal cooling and absorption occur under parallel flow configurations. In addition, they demonstrated the occurrence of heat and mass transfer enhancements under an inclined configuration due to the heightened flow velocities through the channel. The falling film model was modified and studied by Rahamah et al. [10] to incorporate fin temperature variations using an annular fin approximation under a parallel flow configuration. This parametric study provided correlations along with general device performance and limitations using a calcium chloride desiccant solution. ANALYSIS The proposed cycle is shown in Figure 1 under nominal operating conditions. The cycle contains two main loops: a chilled water loop and a desiccant loop consisting of an aqueous solution of calcium chloride. As a result of the relatively high specific heat of the desiccant solution, optimal inlet desiccant temperatures were suggested as ranging from C [3, 4]. A nominal inlet desiccant temperature is taken as an average within this suggested range as 25 C. The circulating chilled water helps to maintain low base and desiccant temperatures to promote heat and mass transfer through the dehumidifier. The heated exit water temperature is then cooled back to the initial temperature by a chiller. As the desiccant absorbs moisture from the air, significant dilution occurs within the desiccant (increased water concentration). This weak solution is then regenerated to initialize

3 S. FEYKA AND K. VAFAI 165 Figure 2 (a) Illustration of multiple fin parallel flow dehumidifier with exploded views of accompanying control volumes. (b) Cross-sectional view of the annular fin approximation with accompanying geometry. desiccant concentration. It was found that regeneration is highly dependent on elevated base temperatures; however, the inlet desiccant temperature has a minimal effect. In order to maintain and adjust regeneration base temperatures, heated water is circulated through a fin-tube regenerator. Dehumidifier/Regenerator Analysis As seen in Figure 2a, the heat removed from the water flow (CV 1) is convected through the fin. The exit bulk water temperature emerging from CV 1 is determined from an energy balance using an annular fin approximation, as shown in Figure 2b. Utilizing an analytical expression for the heat flux, given by Brown [11], and assuming an insulated fin tip results in the following: ṁ w c p,w (T f,o T w,i ) = πd p k f Bθ b m K 1(mR p )I 1 (mr 2 ) I 1 (mr p )K 1 (mr 2 ) (1) K 0 (mr p )I 1 (mr 2 ) + I 0 (mr 2 )K 1 (mr 2 ) where I o and I 1 are modified first-order Bessel functions of the zero and first kind and K o and K 1 are modified second-order Bessel functions of the zero and first kind, respectively. The excess base temperature, θ b, is defined as: θ b = T b T (2) where the base temperature, T b, is approximated as an average of the inlet and exit bulk water temperatures from CV 1. The constant, m, is defined as: 2 h a,dh m = (3) k f B where B is the fin thickness, k f is the thermal conductivity of the fin, and h a,dh is the average convection coefficient from the desiccant film. Following an energy balance around the pipe (CV 2 of Figure 2a), the bulk temperature of the chilled water leaving CV 2 is defined as: T ave T w,n T ave T f,o ( ) πdp f s = exp Ū ṁ w c p,w where T ave is taken as the average inlet and exit air temperatures, neglecting the airside pressure drop due to the pipe; Ū is the overall heat transfer coefficient; f s is the fin spacing; D p is the pipe diameter; and the subscript n is the fin number. The average surface temperature of the pipe was taken as [12]: T s = T w,n + T f,o (5) 2 where the exit bulk water temperature of the pipe, T w,n,isobtained from Eq. (4). Assuming steady, incompressible fully developed internal flow, the Reynolds number of the circulating chilled water can be defined as: Re w = 4ṁ w (6) μ w πd p where ṁ w is the mass flow rate and μ w is the dynamic viscosity of the chilled water. The heat removed from the air is a combination of the fin and pipe convective effects as seen from an energy balance around CV 3 (See Figure 2a): ṁ a,dh c p,a (T a,i T a,o ) = 2 h a,dh HW(T a,i T b ) + πd p f s h bp ( T ave T s ) (7) (4)

4 166 S. FEYKA AND K. VAFAI where W is the channel width and h bp is the airside convection coefficient of the bare pipe. The convection coefficient, h c, has been correlated under low free stream turbulence for flow around a cylinder provided by Churchill and Bernstein [13], assuming a constant surface temperature, negligible flow blockage and end-effects, small temperature differences, and negligible free convection. The convection coefficient for a fin-tube falling film dehumidifier was correlated by Rahamah et al. [10] as: h a,dh = ( kf D h ) Pr 1.84 a Re 0.9 a,dh ( ) H 0.69 ( fs ) where D h is the hydraulic diameter of the channel and H is the channel height. Assuming steady, incompressible, fully developed flow, the Reynolds number describing the air and desiccant flow through the channel is defined as: Re a = 4ṁ a (9) μ a W Re d = 4ṁ d (10) μ d W where μ a and μ d are the dynamic viscosity of air and desiccant, respectively. Applying a moisture balance around the channel (see Figure 2a), the exit humidity ratio can be determined as: ω = p z (12) p t p z where p t is the total vapor pressure (Pa). The partial vapor pressure of the desiccant, p z, is equal to [10]: ( p z = p ws Z 2 + Z T ) b 40 (13) 350 where p ws is the saturated vapor pressure of water and Z is the salt concentration of the desiccant. From a Sherwood number correlation provided by Rahamah et al. [10], the mass convective term is defined as: ) ( ) Ca H 0.69 ( ) h ma = ( Sc 1.84 Re 0.9 fs D h (8) term, as both sides of the channel are considered: ṁ d (h d,o h d,i ) = 2 h a,dh HW(T a,i T b ) + 2ṁ a,dh (ω o ω i )h fg 2 h d HW(T d,i T b ) (15) where h fg is the heat of vaporization of water and T d,i and T d,o are the inlet and exit desiccant temperatures, respectively. A general convection coefficient from the desiccant to the fin is given by Rahamah et al. [10] for both the dehumidification and regeneration cases as: h d = k d (Re d Pr d ) 0.77 (16) D h where k d is the thermal conductivity of the desiccant and Re d and Pr d are the Reynolds and Prandtl numbers of the desiccant, respectively. In addition, exit desiccant concentration predictions were obtained from a mass balance around the channel (see Figure 2a): ṁ a,dh (ω i ω o ) = 2ṁ d (ξ o ξ i ) (17) where ξ is the water concentration of the desiccant and ṁ d is the mass flow rate of the desiccant. It should be noted that the inlet desiccant temperature for the dehumidifier will tend to fall below the suggested optimal range as it rejects heat to the regeneration desiccant stream. As mentioned from previously published data [10], this may lead to desiccant crystallization depending on the salt concentration and temperature of the solution. In order to stabilize localized desiccant areas from crystallization, the proposed design includes a ṁ a,dh (ω i ω o ) = 2 h ma HW(ω i ω ) (11) pre-heater, which uses the ambient air to heat the desiccant back where ω is the humidity ratio of the air stream. ω is the interfacial humidity ratio [14]: the dehumidifier is also reduced, leading to slightly higher cycle to quiescent conditions. As a consequence, the sensible load of efficiencies. From an energy balance around the pre-heater (see Figure 1), the inlet air temperature to the dehumidifier is defined as: 0 = ṁ d (h d,pi h d,po ) + ṁ a (h a, h a,po ) (18) Assuming a constant specific heat of air, the above energy balance reduces to: ṁ a c p,a (T a,po T ) = ṁ d (h d,i h d,o ) (19) where T is the ambient air temperature, T a,po is the pre-heater exit air temperature, and h d,pi and h d,po are the inlet and exit desiccant enthalpies of the pre-heater, respectively. Based on Eq. (11), desiccant regeneration occurs as the interfacial humidity ratio exceeds the humidity levels of the air (14) stream. The exit regeneration air temperature was determined from an energy balance around CV 3 (see Figure 2a): where Sc is the Schmidt number and C a is the diffusion coefficient of air. ṁ a,r c p,a (T T a,o ) The exit desiccant temperature was determined from an energy balance around half of the channel considering both evap- a,r HW(T T b ) + πd p f s h bp ( T ave,r T s,r ) (20) = 2 h orative and convective effects at the desiccant/fin and the desiccant/air interfaces. Using the channel centerline for symmetry, regeneration air temperatures, and the average pipe surface tem- where T ave,r is taken as an average between the inlet and exit the exiting desiccant enthalpy is simply double each convective perature, T s,r, is taken from Eq. (5) for the regeneration case.

5 S. FEYKA AND K. VAFAI 167 The convection coefficient [10] for the regenerator is given as: h a,r = k f Pr 1.51 a Re 0.9 a,r (21) D h The same basic conservation laws are applied for regeneration performance predictions in the form of energy and mass balances, starting with a moisture balance given by: ṁ a,r (ω o ω i ) = 2 h ma HW(ω i ω ) (22) The exit desiccant temperature was determined from an energy balance along the film thickness as given below: ṁ d c p,d (T d,o T d,i ) = 2 h d HW(T b T d,i ) 2 h a,r HW(T b T ) 2ṁ a,r h fg (ω o ω i ) (23) From a mass balance along the desiccant film: ṁ a,r (ω o ω i ) = 2ṁ d (ξ o ξ i ) (24) In order to maintain cycle reliability and performance, the regenerator design is set so that the exit desiccant concentration from the regenerator will regain the inlet dehumidifier condition. Based on this condition, Eqs. (22) and (24) were solved simultaneously and equated to the definition of the interfacial humidity ratio, as defined by Eq. (12). When combined with a partial pressure desiccant correlation based from tabulated data [15], this procedure yields the required base temperature necessary to satisfy the exit concentration constraint. flow heat exchanger. The required power from the compressor, assuming isentropic and adiabatic compression, is: Ẇ c = ṁ R (h 2s h 1 ) (25) where h 2s is the isentropic exit enthalpy of the compressor (State 2s of Figure 3) and ṁ R is the refrigerant mass flow rate. Equation (25) can be expressed in terms of compressor efficiency as: Ẇ c = η c ṁ R (h 2 h 1 ) (26) where η c is defined as the compressor efficiency. For this study, a nominal compressor efficiency of 80 percent is considered. The refrigerant mass flow rate is obtained from an energy balance around the evaporator (CV 4 of Figure 3). As the refrigerant s cooling effect is primarily due to its phase change, complete phase change from liquid to saturated vapor is assumed under constant temperature conditions, leading to: Electric Heating ṁ w c p,w (T w,o T w,i ) = ṁ R h fg (27) The required power to the heater is determined as: Q EH = ṁ w c p,w (T b T w,ro ) (28) where T w,ro is the exit water temperature from the regenerator. Similarly, the heat removed from the water is transferred to the air as given from an energy balance around the regenerator, assuming that the specific heat of air remains constant: ṁ w c p,w (T b T w,ro ) = ṁ a,r (h a,ro h a,ri ) (29) Chiller Modeling and Analysis Chiller modeling (see Figure 3) and power consumption predictions are based on an analysis of a vapor compression cycle operating between the inlet and exit dehumidifier water temperatures. As seen in Figure 3, the evaporator performs as a counter- leading to: Q EH = ṁ a,r (h a,ro h a,ri ) (30) where ṁ a,r is the mass flow rate of air through the regenerator, and h a,ri and h a,ro are the inlet and exit enthalpies of air through the regenerator, respectively. The enthalpy of the regenerator air is defined as: h a,r = h dry + ω a h v (31) where the enthalpy of dry air, h dry, has been correlated based on tabulated data [16]. The enthalpy of the moisture, h v, has been approximated as the enthalpy of saturated water vapor and correlated from tabulated data [17] for liquid water. COP The coefficient of performance (COP) of the proposed cycle has been defined for a combined cycle with both work and heat added. As the steady-flow work is proportional to the specific volume of the desiccant, the work input from the liquid pumps is very small and neglected in the cycle analysis. Therefore, the COP can be approximated as: Figure 3 Dehumidifier section with vapor compression chiller cycle. COP = Q L Q EH + Ẇ c (32)

6 168 S. FEYKA AND K. VAFAI From an energy balance around the dehumidifier, the heat load of the dehumidifier, Q L, is the total amount of heat removed from the chilled water: Therefore, Q L = ṁ w c p,w (T w,o T w,i ) (33) COP = ṁ w c p,w h fg (T w,o T w,i ) Q EH h fg + η c ṁ w c p,w (T w,o T w,i )(h 2 h 1 ) (34) RESULTS AND DISCUSSION A parametric study is employed to investigate the overall cycle performance. Although the main focus of this investigation is to show improvements of the overall cycle efficiency (COP) as compared to a refrigeration cycle, the study also describes nominal cycle operating parameters and general cooling and dehumidification trends. Effects of Airflow on Conditioned Air Humidity and Temperature As expected [6 10], the cooling and dehumidification of the inlet air is strongly dependent on the airflow rate. On the other hand, the chilled water flow rate has a modest effect on the conditioned air temperature and humidity, as shown in Figure 4. The inverse dependence of airflow on the rate of heat and mass transfer contribute significantly to its convex structure, as seen in Figure 5. From previous studies of airflow analysis within a falling film dehumidifier [6 10], the effects of tube geometry are often neglected by assuming small pipe diameters. In addition, the regeneration airflow rate plays a significant role on the conditioned air temperature and humidity (see Figure 6) Figure 5 Effects of airflow rate and chilled water temperature on conditioned temperature and humidity levels. as a result of its control over the inlet desiccant stream to the dehumidifier. Effects of Chilled Water Temperature on Conditioned Air The water temperature tends to play a significant role in the overall cycle cooling and dehumidification. Because the mass diffusion and heat transfer both depend on the interfacial temperature and humidity differences, cooler air temperatures at lower humidity can be achieved as the inlet water temperature is lowered (see Figure 5). Although the cooling and dehumidification process is enhanced under lower circulating water temperatures, additional energy is required from the chiller to remove the excess heat. This effect helps to explain the gradual decline in efficiency (see Figure 7) under lower airflow where the chiller load is clearly noticeable. Figure 4 Effects of air and chilled water flow rates on conditioned temperature and humidity levels. Figure 6 Effects of regeneration and dehumidifier flow rates on conditioned air temperature and humidity.

7 S. FEYKA AND K. VAFAI 169 Figure 7 Effects of airflow rate and chilled water temperature on the proposed cycle COP in comparison to a standard chiller cycle COP. Effects of Dehumidifier Sizing on Conditioned Air Humidity and Temperature The width of the dehumidifier also plays a significant role [7] in cooling and cycle performance. As seen in Figures 8 and 9, modest increases in the width lead to substantial cycle performance and efficiency enhancements due to the heightened surface area of the fin. In addition, the relative length, or number of fins, of the dehumidifier seems to play an insignificant role in both cycle performance and efficiency. This suggests that the amount of conditioned air that the cycle can provide (volumetric flow rate) is dependent on width but is largely independent of length. Effects of Airflow on Regeneration Base Temperature Regeneration base temperature variations are required to satisfy the exit desiccant concentration constraints. Indirectly, the Figure 9 Effects of dehumidifier width and airflow rate on the proposed cycle COP in comparison to a standard chiller COP. airflow rates through the dehumidifier strongly influence the required base temperature as a result of enhanced rates of absorption under lower airflows (see Figure 10). By lowering the conditioned flow rate, the accompanying diffusion enhancement yields a higher water concentration in the desiccant. As a result, higher regeneration base temperatures are required to expel the elevated water concentration. Conversely, lower regeneration base temperatures are required under higher conditioned flow rates as the desiccant absorbs less water. As depicted in Figure 10, regenerator performance and base temperatures are also largely dependent on regeneration airflow rates. Effects of Airflow Rates on Cycle COP Traditionally, additional cycle efficiencies may result from either a reduction in energy delivered to the cycle or by an increase Figure 8 Effects of airflow rate and width on conditioned air temperature and humidity levels. Figure 10 Effects of regeneration and dehumidifier airflow rates on regeneration base temperatures and heater power consumption.

8 170 S. FEYKA AND K. VAFAI COMPARISON Figure 11 Effects of regeneration and dehumidifier airflow rates on the proposed cycle COP in comparison to a standard chiller cycle COP. in the cycle s cooling capacity. As described earlier, the diffusion and evaporation rates are highly dependent on the regeneration airflow rate. The total cycle COP steadily increases as regeneration airflow increases, as shown in Figure 11. This effect is largely due to the heightened desiccant temperatures from the regenerator under lower airflow. The study concludes that a liquid desiccant cycle may improve cycle performance by percent, with a 22 percent humidity reduction under nominal conditions. In Figure 12, the proposed cycle and a traditional chiller cycle are compared to the maximum possible efficiency under varied airflow. As can be seen, the liquid desiccant cycle offers a substantial enhancement over the traditional chiller cycle. Figure 12 Effects of airflow rate on efficiency gains over a standard chiller cycle and relative comparison to a reversed Carnot efficiency. The model used in this study was compared to a chiller cycle using a mixture of monoethylene glycol [18] and water as the working fluid. The ε NTU method is employed to determine the amount of heat transferred to the chilled water. This technique defines the rate of heat transfer as the maximum possible temperature difference between the corresponding fluids adjusted by an effectiveness constant. Because the effectiveness is defined as a ratio between the actual to the maximum possible heat transfer, this term acts as a scaling factor varying from zero to unity. Q EH = εṁ a,eh c p,a (T a,i T w,i ) (35) where the effectiveness constant, ε, was taken from a correlation [19] for a cross-flow heat exchanger assuming a mixed air stream and unmixed water stream. The NTU method is based on an energy balance between the two fluid streams, assuming uniform flow distribution of the cold stream and uniform distribution of the heat transfer area while neglecting mass transfer effects [19]. As the NTU is a function of the airflow rate, specific heat, and overall heat transfer coefficient (UA), the UA is defined as the total thermal resistance from the ambient to the chilled water bulk temperature as given below: 1 UA = ( π h bpd p f s kf + h f ln ( R 2 )) R p k f + h f ln ( R 2 ) + R p + 2(R2 R p )π 2 BD p h bp f s h f k f 1 πd p h w f s (36) where the convection coefficient of water, h w, is based on Dittus and Boelter s [20] correlation and h f is obtained from a Nusselt number correlation [21] for channel flow between two isothermal plates. The exit air temperature was determined from an energy balance over the channel as: ṁ a c p,a (T a,o T a,i ) = εṁ a c p,a (T a,i T w,i ) (37) Along the same lines as Eq. (32), the chiller cycle efficiency is defined as a ratio of the cooling load to the power consumption of the chiller: COP = ṁac p,a (T a,i T a,o ) (38) η c ṁ R (h 2 h 1 ) From an energy balance between the refrigerant and chilled water, the refrigerant flow rate is defined as: ṁ R = ṁwc p,w (T w,o T w,i ) (39) h fg As such, the chiller cycle efficiency can be presented as: h fg COP = (40) η c (h 2 h 1 ) A general comparison between the two contrasting methods under identical operating conditions is shown in Figure 13. The enhanced performance of the desiccant fin tube model is apparent as both heat and mass transfer effects are considered. As

9 S. FEYKA AND K. VAFAI 171 Figure 13 Comparison between a standard heat exchanger model and the proposed fin tube dehumidifier model. can be seen, the general cooling trends between both methods compare particularly well. CONCLUSIONS H fin height, m h enthalpy, kj/kg h average convection coefficient, W/m 2 K h fg heat of vaporization, kj/kg k thermal conductivity, W/m K m fin constant (m 1 ) defined by Eq. (3) or mass, kg ṁ mass flow rate per unit width, kg/s m N fin number NTU number of units transferred p pressure, kpa Pr Prandtl number Q rate of heat transfer, kj/s q heat flux, W/m 2 R radius, m Re Reynolds number Sc Schmidt number T temperature, C T average bulk temperature, C UA overall heat transfer coefficient, W/ C W channel width, m Ẇ power, kj Z salt concentration of the liquid desiccant, m salt /m solution A liquid desiccant cooling cycle has been investigated and compared to a standard chiller cycle for overall performance and efficiency. The results under nominal values indicate that the proposed cycle has a maximum cooling capacity of 17 C and a humidity reduction of 22 percent while boosting the overall cycle efficiency by percent. The proposed cooling performance of the dehumidifier varies substantially with airflow and geometry, depending on its application. Furthermore, the fin tube desiccant model compares well with similar methods using a standard heat exchanger analysis. Although liquid desiccant cooling is relatively ineffective for required temperature reductions above 20 C, it is well suited for tropical climates with small diurnal temperature swings and high relative humidity. Furthermore, cycle efficiency enhancements may be achieved by using ammonia as the chiller refrigerant due to its high evaporative energy [22], resulting in both lower conditioned air temperatures and overall reduction in compressor power. Additional efficiencies may be realized, as higher performance desiccants are selected with better absorption properties while maintaining low regeneration temperatures. With these additional improvements, liquid desiccant refrigeration and freezing may be possible while maintaining a COP above a standard vapor compression or chiller cycle. NOMENCLATURE B fin thickness, cm C diffusion coefficient, m 2 /s c p specific heat, kj/kg K CV control volume D diameter, m fin spacing, m f s Greek Symbols δ thickness, m ε effectiveness constant η thermal efficiency μ dynamic viscosity, kg/s m θ b excess base temperature, C ω humidity ratio, m w /m da ω interfacial humidity ratio, m w /m da ξ water concentration of the liquid desiccant, m w /m solution ρ density, kg/m 3 Subscripts a air b base bp bare pipe c compressor d desiccant da dry air dh dehumidifier EH electric heater f fin h hydraulic i in L load ma mass transfer n fin number o out p pipe, preheater R refrigerant r regenerator

10 172 S. FEYKA AND K. VAFAI s surface t total v vapor w water ws wet saturation z salt concentration of the liquid desiccant 1 saturated refrigerant vapor 2 annular fin 2s isentropic vapor compression 3 saturated liquid refrigerant 4 saturated liquid refrigerant ambient REFERENCES [1] Howell, J. R., and Peterson, J. L., Preliminary Performance Evaluation of a Hybrid Vapor Compression/Liquid Desiccant Air Conditioning System, ASME Paper 86-WA/Sol. 9, Anaheim, California, [2] Lof, G. O. G, Cooling with Solar Energy, Proceedings of the World Symposium on Applied Solar Energy, Phoenix, Arizona, November 1 5, pp , [3] Factor, H. M., and Grossman, G., A Packed Bed Dehumidifier/Regenerator for Solar Air Conditioning with Liquid Desiccant, International Journal of Solar Energy, vol. 24, pp , [4] Radhwan, A. M., Gari, H. N., and Elsayed, M. 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F., Transfer Processes: An Introduction to Diffusion, Convection, and Radiation, 2nd ed., Hemisphere Publishing Corp., Washington, pp , [22] Haar, L., and Gallagher, J. S., Thermodynamic Properties of Ammonia, Journal of Physical Chemistry Reference Data, vol. 7, pp , Scott Feyka is a mechanical engineer for an aerospace subcontractor located in San Diego, California. He received his Masters degree in mechanical engineering from the University of California, Riverside, in 2005, and his bachelor s degree in Electrical Engineering from the University of San Diego in Currently, he is a lead mechanical engineer for the development of Integrated Microwave Assemblies, where his specialization in heat transfer and thermodynamics plays a critical role in product modeling and analysis. Kambiz Vafai started his university career at The Ohio State University, where he received the outstanding research award as an assistant, associate, and full professor. He has authored more than two hundred journal publications, book chapters, books, and edited symposium volumes, and has been an invited visiting professor at the Technical University of Munich in Germany, University of Bordeaux, and Paul Sabatier University (multiple times) in France and Technical University of Naples in Italy. He is a Fellow of ASME (1992), Fellow of AAAS (2002), Associate Fellow of AIAA (1998), and Fellow of WIF (2003). He has been selected by the ISI among the very limited number of highly cited scientists. He has received the ASME Heat Transfer Classic Paper Award in 1999 and the Heat Transfer Memorial Award (2006). He is the editor in chief of the Journal of Porous Media, editor of both editions of the Handbook of Porous Media, and on the editorial board of several technical journals. He has been an associate editor of the ASME Journal of Heat Transfer. He is currently Professor of Mechanical Engineering at the University of California, Riverside.