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Desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas
Sustainable Energy Technologies and Assessments 1 (2013) 54–62
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Original Research Article
Desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas Ahmed Hamza H. Ali ⇑ Department of Energy Resources and Environmental Engineering, Egypt-Japan University of Science and Technology (E-JUST), P.O. Box 179, New Borg El-Arab City, Alexandria 21934, Egypt
a r t i c l e
i n f o
Article history: Received 23 July 2012 Revised 26 January 2013 Accepted 27 January 2013
Keywords: Desiccant cooling Nocturnal radiative cooling Solar cooling Air comfort Hot arid areas
a b s t r a c t In this study, the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas of Upper Egypt is carried out using the analytical techniques. The system is investigated based on a daily cycle one-way airflow direction and the measured weather data. In addition, a mathematical model for analyzing the heat and mass transfer in the system during adsorption (nighttime mode) and regeneration (daytime mode) is established. The obtained results were verified with correspondent experiments and a good agreement existed. Thereafter, the model is used to investigate the system feasibility for air comfort application. It is found that this system can be used in hot arid areas of Upper Egypt at which the ambient air relative humidity is very low in summer daytime and gradually increases during nighttime that makes utilization of this system effective. Also, the model is used to investigate the effect of the air mass flow rate on this system performance and the results show that, it is preferred to use a low air mass flow rate in order to obtain a lower outlet air temperature which is less than the ambient air by 5.5–7 °C and below 40% relative humidity. Ó 2013 Elsevier Ltd. All rights reserved.
Introduction Passive cooling of buildings within the human comfort zone is an important subject due to its goal of keeping the environment out of contamination. The upper atmosphere (the sky) is a heat sink for nocturnal emitted longwave radiation by terrestrial units to produce a net cooling. Desiccant based systems process atmospheric air by removing moisture from the air to a desired level and then cooling it using a heat exchanger, followed by some evaporative cooling. However, the desiccant based systems are generally effective in hot humid areas. Rotary desiccant wheels and fixed desiccant beds are the most common desiccant dehumidifier configurations. Pesaran et al. [1] presented a comprehensive desiccant cooling bibliography containing 1176 pieces of literature available up to the mid-1997. Among the commercially available systems, rotary desiccant wheels allow continuous operation, while a fixed desiccant bed is flexible in positioning but cannot run continuously. Usually, more than one fixed desiccant bed unit is used to compensate for the non-continuous operation drawback. One fixed desiccant bed can be in regeneration while another is in dehumidification process. A different desiccant cooling idea called desiccant enhanced nocturnal radiation cooling and dehumidifica-
⇑ Tel.: +20 3 4599839; fax: +20 12 23971265. E-mail address: [email protected] 2213-1388/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.seta.2013.01.003
tion has been proposed descriptively by Fairey et al. [2,3] and Swami et al. [4]. Swami [5] quantified this idea and clarified its benefits when applied to a residence in hot humid areas. The idea of the integrated desiccant/enhanced nocturnal radiative cooling-solar regenerated system is summarized as follows. A unit is composed of a plane structure filled with desiccant material and topped with a highly conductive and emissive radiator plate. Ambient air is withdrawn through the desiccant bed during nighttime where dehumidification with probably some cooling occurs. The desiccant regeneration cycle occurs during daytime using ambient air and solar radiation. The dehumidified air at night can be further cooled evaporatively and stored to be utilized later to cool space within building during the warm daytime or it can be used directly to cool and dehumidify the building during the night time. Lu et al. [6,7] experimentally investigated a desiccant enhanced nocturnal radiation cooling and the dehumidification system. They reported that there are several characteristics of this system: firstly, it can continuously supply the conditioned space with processed (or dehumidified) air 24 h a day; secondly, it contains many functions which are suitable for public use like air conditioning, drying, dehydration and active cooling. Thirdly, since the only energy consumption of the system is from the ventilation fans plus the enormous free solar energy input to regenerate the system, its energy efficiency is incredibly high, and finally, a large potential for dehumidification in such a humid area as Taiwan is expected.
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Nomenclature area, m2 free flow area inside the bed, m2 surface area of the desiccant bed, m2 plate base surface area and gfinAfins, m2 specific heat of humid air, J/kg K specific heat of the desiccant, J/kg K specific heat of dry air, J/kg K specific heat of water vapor, J/kg K specific heat of liquid water, J/kg K channel height, m configuration factor adsorption heat, J/(kg of adsorbed water) enthalpy of humid air, J/kg convective heat transfer coefficient, W/m2 K enthalpy of wet desiccant (energy content per unit mass of dry desiccant), J/kg hg specific enthalpy of water vapor, J/kg mass transfer coefficient, kg/m2s hm i, j and k element numbers L length of the unit, m l offset fins spacing, m _ air m mass flow rate of the air stream, kg/s _ water m mass flow rate of water from the bed to the air stream (Fig. 2), kg/s Noss number of surrounding elements q heat transfer flux, W/m2 qsolar total incident solar radiation flux on the plane of the unit, W/m2 Q heat transfer rate, W RH relative humidity of hypothetical air layer A Abed,air Abed,s Apl-iner Cpair Cdes Cpair,dry Cpwv Cw D F qads Hair hbed,air Hdes
Saito [8] and Techajunta et al. [9] investigated the regeneration in an integrated desiccant/collector system using a direct solar energy simulator as a heat source through experiments and mathematical simulations. Their desiccant bed configuration was a layer of adsorbent laying along the diagonal of the collector under glass cover and subjected directly to the radiation flux and their units have no thermal radiation cooling capability. In their model they analyzed the regeneration process only and considered the whole system is one lump in the energy balance equation. In the desiccant enhanced nocturnal radiative cooling-solar collector system air is used as a working fluid and the heat transfer medium for the heat of adsorption and regeneration to/from the absorber/radiator plate. The disadvantage of the air as a heat transfer medium is the low convective heat transfer coefficient due to its low thermophysical properties. Throughout the literature on utilization of the desiccant system it is clear that the desiccant dehumidification cooling techniques are more effective when utilized in hot humid areas. While, the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas is not available in the literature. Therefore, in this study, using the analytical techniques, an evaluation study is carried out with a detailed simulation model considering the heat transfer modes within the suggested system, and, mainly implementing the feasibility of this system for air comfort application in hot arid areas of Upper Egypt. In Upper Egypt, the ambient air relative humidity is very low in summer daytime and slightly increases by nighttime, which makes utilization of this system feasible for air cooling and dehumidification. The suggested configuration is an absorber/radiator plate without upper glazing in order to utilize the convection losses to the ambient to enhance the plate cooling
t Tair Tbed Vbed W x
time, s temperature of the air stream, °C temperature of desiccant bed, °C volume of desiccant bed element, m3 channel width, m length scale m
Greek symbols a absorptance, – d offset fin height, m e emittance, – r Stefan–Boltzmann constant, (5.67051 108), W/m2 K4 gfin fin efficiency xair humidity ratio of air stream xbed desiccant bed water content on dry basis xs,bed humidity ratio of hypothetical air layer qbed bulk density of the desiccant bed, kg/m3 Subscripts air dry air stream amb ambient bed desiccant bp back plate conv convection pl absorber/radiator plate rad thermal radiation ss surrounding surfaces s surface of the desiccant bed sky sky wv water vapor
and also to enhance nocturnal radiative cooling. This absorber/ radiator plate is designed to have a set of offset fins on its backside parallel to the flow direction in order to improve the convected heat transfer process from the plate to the air stream. It is also aimed to examine the effect of air mass flow rate on the system performance. System concepts and description of the operating modes The nocturnal radiative cooling of a surface exposed to the sky can be used to lower the temperature of a fluid beneath the radiator plate below that of surrounding ambient air temperature. Detailed explanation of this phenomenon was presented by Ali et al. [10]. As the radiator plate temperature rises the radiative heat rejection rate to the sky increases. Desiccation is a process by which highly hygroscopic materials adsorb moisture from the air. This process is combined by a release of thermal energy roughly equal to the normal heat of water vapor condensation and heat of wetting. A desiccant system can fall into one of two generic operating concepts adiabatic or isothermal. Since the desiccation process in isothermal systems operates at a lower temperature than that in adiabatic systems as reported by Lavan et al. [11], the moisture absorption capacity of isothermal systems is somewhat greater than that of adiabatic systems. However, the concept of the desiccant enhanced nocturnal radiative cooling-solar collector system attempts to take advantage of approaching isothermal desiccation case through nocturnal radiative cooling. The radiator plate is coupled thermally to the desiccant bed during the adsorption process leading to a radiator temperature which tends to remain equal to or less than the ambient by about 3–5° due to heat of adsorption. Thus, a combination of the moisture removal
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Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
qrad,pl,sky
A
Tair,in
D
air,in
qsolar
element i
qconv,pl,amb
absorber/radiator plate
Tair,out
C
air,out
Daytime mode
x
back plate
Thermal Insulation
B
Nighttime mode
(a)
l fin desiccant bed
E
L (b)
Fig. 1. Schematic of the proposed desiccant nocturnal radiative cooling-solar collector system (a) cross section and (b) plane view.
capabilities of desiccant materials and, heat rejection capabilities of nocturnal sky radiation is achieved in such a way that each process complements and enhances the other. The proposed desiccant enhanced nocturnal radiative coolingsolar collector system for the present analysis is shown in Fig. 1. It consists of an absorber/radiator plate made of an aluminum sheet coated with black paint at both sides and has a set of offset fins on its backside parallel to the flow direction. A back plate forms with the absorber/radiator plate a channel having six sections four of them contain desiccant bed laying along the diagonal (best configuration due to results of Swami [5]) across the air flow path. The absorber/radiator plate is exposed to the ambient while the back plate of the channel is thermally well insulated. Ambient air flows through the system from the left or the right side (location A or B) based on the operation mode. At nighttime ambient air is flown from location B to location C for adsorption mode and is dehumidified by the desiccant bed. The bed releases heat, which warms the air and consequently the absorber/radiator plate. Thus, the warmer absorber/radiator plate has a greater cooling potential with respect to the ambient and from location C to location A the rest of absorber/radiator plate (two sections) serves to cool the air stream without dehumidification. The out-coming air at location A is more dry and cool than at the inlet location B. During the daytime, regeneration mode, the ambient air is flown through the system from location A to location C, where it is heated in the first two sections of the absorber/radiator plate by the absorbed solar radiation. Thus, at the beginning of the section having desiccant bed at location C, this air has a relatively high temperature and low relative humidity. This warm air combined with the heat exchange between the absorber/radiator plate and bottom plate with the desiccant bed serves to regenerate the bed from location C to location B. The air stream is then exhausted to the ambient at location B.
Analytical approach Simulation of this suggested system is carried out based on a daily cycle one-way airflow direction and the measured weather data for Assiut, Egypt as the model inputs. The system attempts to combine the moisture removal capabilities of desiccant materials, and heat rejection capabilities of nocturnal sky radiation in such a way that each process complements and enhances the other, and at the same time provides isothermal desiccation which, allows for lower regeneration temperature that provides an avenue for using a solar driven desiccant system. A mathematical model for analyzing the heat and mass transfer, for the proposed system shown in Fig. 1, during adsorption (nighttime mode) and regeneration (daytime mode) is obtained by applying the energy and mass balances on its parts. This is done in three parts, the first part is for
the plates, the second is for the section not having desiccant bed and the third part is for the section having desiccant beds at which the mass balance is applied. The model is established based on the following assumptions. One-dimensional temperature and moisture gradient in both air stream and desiccant bed are considered along the flow direction and the flow channel is divided into n equal elements. Accumulations of energy and moisture in the air elements are neglected. The axial heat conduction in the plates and heat and mass transport in the air stream along the x-axis are neglected. The convection heat and mass transfer coefficients between the air stream and the desiccant bed are constant over the sections between locations B and C of Fig. 1. The desiccant bed is in equilibrium with a thin air layer on the surface of desiccant particles. The mass transfer driving force is the water vapor pressure difference between this air layer and the processing moist air stream. The pressure drop along the desiccant beds does not affect the air stream velocity. For the offset fins, the fins are considered to have the base temperature of the absorber/radiator plate with surface area equal to gfinAfin. Conduction between plates and desiccant bed ends is ignored. Each desiccant bed element is considered as a lumped heat capacity system. Surfaces inside the duct are black for longwave radiation. Energy storage in plates and fins is neglected. Energy balance For plates The energy balance equations are applied on the plate elements as follows: For the absorber/radiator plate element (i).
apl;i Apl;i qsolar hpl;amb Apl;i ðT pl;i T amb Þ repl;i Apl;i F pl;sky ðT 4pl;i T 4sky Þ Noss X hpl;air Apl-iner;i ðT pl T air Þi r Apl-iner;i F pl;ij ðT 4pl;i T 4j Þ ¼ 0
ð1Þ
j¼1
The terms in Eq. (1) are as follows; the first one is the solar radiation absorbed on the plane of the outer surface of the absorber/radiator plate; the second is representing the convection to the ambient air, and, the value of hpl,amb is determined using the formula presented in Duffie and Beckman [12]; the third is the net thermal radiant energy exchange with the sky; details of calculation sky temperature are in Ali et al. [10]. While, the fourth
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Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
term is the convection heat transferred to the air stream in the channel, and, the value of hpl,air is calculated using the formula presented in Diab [13]. The fifth one is the thermal radiant exchange between the absorber/radiator plate element i and all other solid surfaces (Noss) including any desiccant surface it sees. For the back plate element (i).
hbp;air Abp;i ðT bp T air Þi þ r
Noss X Abp;i F bp;ij ðT 4bp;i T 4j Þ ¼ 0
ð2Þ
where hbed,air is the convective heat transfer between the desiccant bed and the flowing air and is given in [4]; the third term in Eq. (4) and the fourth term in Eq. (5) are the energy transferred to the air stream due to the heat of adsorption that is a summation of the normal heat of condensation plus the heat of wetting; the first term in Eq. (5) is the rate of energy storage in the desiccant bed; the second term in Eq. (5) is the net thermal radiant exchange from the desiccant bed element i to all other solid surfaces (Noss) including any desiccant surface it sees.
j¼1
The two terms in Eq. (2) are defined as follows: the first one is the heat convected to the air stream from the back plate in the duct, and, the value of hbp,air is calculated using the correlation equation for the data of Heaton et al. [14], while the second one is the net thermal radiant exchange between the back plate element i and all other solid surfaces (Noss) including any desiccant surface it sees. For sections not having desiccant beds The energy balance equation is applied on the air stream element as follows: For air stream element (i).
hpl;air Apl-iner;i ðT pl T air Þi þ hbp;air Abp;i ðT bp T air Þi _ air Cpair ðT air;in T air;out Þi ¼ 0 þm
ð3Þ
The terms in Eq. (3) are defined as follows; the first is the convected heat to the air stream in the duct from the absorber/radiator plate; the second is the convected heat to the air stream in the duct from the back plate; the third is the net energy transfer due to inflow and outflow. For sections having desiccant bed. The equations governing the energy balance for the desiccant bed and air stream elements i, shown in Fig. 2, are given as follows: (i) For air stream node (i)
_ air ðHair;in Hair;out Þi þ hbed;air Abed;airÞi ðT bed T air Þi m qads hm Abed;airÞi ðxs;bed xair Þi þ hpl;airÞi Apl;i ðT pl T air Þi þ hbp;airÞi Abp;i ðT bp T air Þi ¼ 0
ð4Þ
(ii) For desiccant bed node (i)
qbed V bed
Noss X @Hbed þ r Abed;sÞi F bed;ik ðT 4bed;i T 4k Þ @t k¼1
þ hbed;air Abed;air ðT bed T air Þi qads hm Abed;airÞi ðxs;bed xair Þi ¼ 0
ð5Þ
The first term in Eq. (4) is the net energy transfer due to inflow and outflow; the second term in Eq. (4) and the third term in Eq. (5) are the net convected heat transfer from the bed to the air stream,
Mass balance The equations governing the mass transfer for the desiccant bed and air stream elements i, shown in Fig. 2, are given as follows: For sections having desiccant bed (i) For air stream node (i)
_ air ðxair;in xair;out Þi hm Abed;air;i ðxs;bed xair Þi ¼ 0 m
ð6Þ
(ii) For desiccant bed node (i)
qbed V bed
@ xbed @t
hm Abed;air;i ðxs;bed xair Þi ¼ 0
ð7Þ
The first term in Eq. (6) is the water mass transfer rate due to the inflow and outflow of the control volume. The second term in Eqs. (6) and (7) is the convective mass transfer from the bed to the air stream, where hm is the convective mass transfer between the desiccant bed and the flowing air and is given in [4]. The first term in Eq. (7) is the rate of water storage for the desiccant bed. The value of Abed,air,i is the product of the transfer area of unit volume of the desiccant bed multiplied by the desiccant bed element volume; it is a desiccant property. In Eqs. (4)–(7), the positive signs are for the desiccant bed regeneration and negative signs for air dehumidification modes. Desiccant equilibrium properties The set of equations governing the dynamics of the sorption process has to be solved along with equilibrium sorption isotherms of the desiccant bed. The sorption isotherms for silica gel bed relating the air stream relative humidity with the desiccant bed humidity ratio and the heat of adsorption of water vapor in silica gel reported by Pesaran and Mills [15,16] and used in this study are given by:
RH ¼ 0:0078 0:05759xbed þ 24:1655x2bed 124:478x3bed þ 204:226x4bed hads ¼ ð12; 400xbed þ 3500Þ hads ¼ ð1400xbed þ 2950Þ
ð8Þ
xbed < 0:05 xbed > 0:05
ð9aÞ ð9bÞ
The ASHRAE [17] psychometrics for moist air is used to calculate xair in Eqs. (6) and (7). The enthalpy values of both the moisture air Hair and wet desiccant bed Hdes are determined from the following equations:
Hair ¼ Cpair;dry T air þ xair ðhg þ Cpwv T air Þ
ð10Þ
Hdes ¼ ðC des þ xdes C w ÞT des
ð11Þ
Numerical method of solution
Fig. 2. Energy and mass transfer for desiccant bed control volume.
Simulation of this system was carried out based on a daily continuous operation, one direction of airflow from ambient and inlet to the system is in direction from location B to location A during
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
Results and discussions Assessment of the theoretical results Before reporting and discussing the feasibility of using this system as well as the effect of varying the air mass flow rate on its performance, the results of the model are compared with the corresponding measured data available in the nearest similar system of Lu and Yan [6]. The unit dimensions of [6] were input to the present model while assuming any missed data. The dimensions of the experimental setup of Lu and Yan [6] which were used in model verification are as follows: total channel flow length L = 3.6 m, flow lengths without desiccant bed E = 2.0 m, channel width W = 1.4 m and channel height D = 20 mm; and, the weather data as present in Lu and Yan [6]. Fig. 3 shows the input weather data to the model and a comparison between the calculated values of the outlet air dry bulb temperature and absolute humidity ratio with corresponding experimentally measured values of [7]. It is clear from the figure that, the predicted humidity ratio values are in good agreement with the measured data. Also, it is seen from the figure that the predicted air dry bulb during daytime is slightly higher by 1 to 1.5 C than the measured one, while its values during nighttime are in reasonable agreement with the measured data. This slight deviation during daytime can be attributed to using the offset fins on the backside of the absorber leading to a higher rate of heat transfer to the air and in the same time decreasing the convection losses from the absorber plate to the surrounding. However, these agreements of the calculated values with the measured data give confidence in the analysis techniques to investigate the feasibility of using this system in Upper Egypt and the effect of varying the air mass flow rate on its performance. Performance characteristics and system feasibility The outlet absolute humidity ratio and air dry bulb temperature are the characteristics of the performance for the suggested system, desiccant enhanced nocturnal radiative cooling-solar collector was intended to be used for air comfort application in hot summer month in hot arid areas of Upper Egypt. An overall look at the year
60
air,out air,out
50 40
0.03 30
0.02 20
0.01
10 3
air mass flow rate=0.0112 kg/s (0.56 m /min) 0
2
4
6
900
Tout [Lu et al., 1995] Tout present 0.04
air,in
0
1000
[Lu et al., 1995] present
8
10
12
14
16
18
20
22
800 700 600 500 400 300 200 100
24
Time, hr Fig. 3. Verification of the calculated parameters and corresponding data of [6].
2
qsolar Tin
o
Dry Bulb Temperature, C
70
study, are as follows: total channel flow length L = 1.2 m, flow length without desiccant bed E = 0.6 m, channel width W = 0.75 m, channel height D = 50 mm, fins pitch l = 148 mm, fin height d = 40 mm and fin length lfin = 90 mm.
Total Solar Insolation (W/m )
nighttime, while, it is in the opposite direction from location A to location B during daytime. In practice, for this system, the air is in cyclic mode to provide a continuous operation. With the system continuing operation, the last condition that forms the dehumidification process at end of nighttime mode is the initial conditions for the regeneration process during the next daytime mode. The temperatures of the solid boundary elements were obtained by solving the energy balance Eqs. (1)–(5) for n-module model (in this study n = 20,000) which generates (4n) finite central difference nonlinear equations in (4n) unknowns. These (4n) unknowns are the temperatures of the absorber/radiator plate, air stream bulk temperature, desiccant bed and back plate elements respectively. For solving this set of nonlinear heat balance equations, the multidimensional secant method – Broyden’s method – which is presented in Press et al. [18] was used. The calculation is carried out in quasi steady at one minute time step. While, this set of equations is a function of the air and desiccant bed moisture content, the moisture content of the air and desiccant bed elements are calculated first for a certain time, keeping the temperature constant, from the mass balance Eqs. (6) and (7). These computation procedures are repeated until the change in the temperatures of the solid boundary elements is within the allowable error (below 105). The desiccant bed material considered in this study is silica gel grade 01, the properties of this grade and the values of both the convective heat and mass transfer coefficients, which were defined, based on Pseudo-Gas-side Controlled model (PGC) were reported by Helen-Xing [19]. The values of hpl,air and Tsky in Eq. (1) were determined using the formulae presented by Duffie and Beckman [12]. The calculation starts with all system components having the same initial condition equal to that of the ambient and the bed water content on dry mass basis xbed(s = 0) = 0.25. Simulation should be based on actual hourly values of weather data in order to decide the feasibility of this system. Therefore, in this study, the hourly average values of 30 years measurements of the weather parameters for Assiut, at Upper Egypt (Lat. 27°120 and long. 31°100 ). The method to calculate these hourly average values that is reported in Degelman [20,21] is used in this study. The radiative (optical) properties used in the model are considered as follows: the absorptance (emittance) of the metallic surfaces, which are coated with a commercial black paint is 0.94 and the emittance of the silica gel is 0.9 as given by Fairey et al. [3]. The flow (air) mass flow rate varied from 0.0094 to 0.022 kg/ s corresponding to air volume flow rate ranging from 0.479 to 1.12 m3/min. The dimensions for Fig. 1, which are used in this
Humidity Ratio
58
59
Solar Radiation Wind speed
1200 1100 1000 900 800 700 600 500 400 300 200 100 0
60
50
40
30
20
10
0
6
12
June
18
24
2
Air dry bulb temprature Air RH %
o
T ( C), RH% and Vw (m/s)
70
Total Solar Insolation (W/w )
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
30
Fig. 4. Weather data for June.
Tambient 80
ambient
Tout
Tlocation
C
out
60 50 40 0.01
30 20
0.008
10 0
0.006
Humidity Ratio (kg/kg)
o
Dry Bulb Temperature, C
70
0.004
6
12
June
18
24
30
Fig. 5. Predicted absolute humidity ratio and air dry bulb temperature at outlet and at location C for air mass flow rate of 0.01251 kg/s (0.64 m3/min).
weather data for Assiut, Egypt indicates that the maximum air dry bulb temperature was recorded through the month of June. The weather data of June are shown in Fig. 4 and are the input to the model. The predicted outlet absolute humidity ratio and air dry bulb temperature from the system and the statues of the dry bulb temperature at location C of Fig. 1 for the air mass flow rate of 0.01251 kg/s (0.64 m3/min) are shown in Fig. 5. As seen in Fig. 4, the ambient air dry bulb temperature during nighttime varied between 30 and 16 °C and the relative humidity reaches up to 47%. However, these weather conditions are applicable to desiccant systems using silica gel as a desiccant and work well. As seen from Fig. 5, during nighttime when the ambient air passes through the sections having desiccant bed (from B to C of Fig. 1) its dry bulb temperature and absolute humidity ratio are lowered with values ranging from 3 to 4 C and 0.13 to 0.75 g/kg respectively, further decrease in its dry bulb temperature occurs when it passes through the sections not having desiccant bed (from C to A of Fig. 1) with values ranging from 2.5 to 3 C. As a result, the total air dry bulb temperature drop ranged from 5.5 to 7 C. These values indicate that as the ambient air is withdrawn through the suggested system during nighttime it is dehumidified and cooled. However, this cool and dehumidified air at nighttime can be further cooled evaporatively and stored to be utilized later to cool space during the warm daytime or it can be used directly to cool and dehumidify
the building during the nighttime. Therefore, this system is feasible in use for air comfort applications in hot arid areas of Upper Egypt, at which, the ambient air relative humidity is very low in summer daytime and increases by nighttimes, which makes utilization of this system effective. In order to clarify and magnify the characteristic details of this system performance, weather data for three consecutive days (72 h) were input to the model. The corresponding predicted results of outlet air absolute humidity and air dry bulb temperature in the case of air mass flow rate of 0.01251 kg/s (0.64 m3/min) are shown in Fig. 6. As seen from the figure, at nighttime when the ambient air passes through the sections having desiccant bed its temperature is decreased. This is because it has a low absolute humidity value, and consequently, a small amount of water vapor is adsorbed to the desiccant bed leading to a released adsorption heat less than the net thermal radiation loss to the sky which causes a net cooling effect. As the air passes through the section dose not having desiccant bed (whose surface area of this section is equal to one third of the total system surface area as shown in Fig. 1) it is further cooled by 3–4 C by the effect of the net thermal radiation losses to the sky. Determination of the night sky radiation cooling rate is carried out ranging from 75 to 85 W/m2. These values are expected in hot arid areas of Upper Egypt due to a small amount of downward thermal radiation because of the low ambi-
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
3
1200 q solar T ambient 1100 1000 ambient 900 T out T at C 800 out 700 600 0.006 500 0.005 400 0.004 300 0.003 200 0.002 100
/min)
70 60 50 40
Humidity Ratio
o
Dry Bulb Temperature, C
air mass flow rate=0.01251 kg/s (0.64 m
30 20 10 0
0
6
12
18
24
30
36
42
48
54
60
66
2
80
Total Solar Insolation (W/m )
60
72
Time, hr Fig. 6. The predicted absolute humidity and air dry bulb temperature at outlet for the presented values of weather data and air mass flow rate.
ent water vapor content, at the daytime, at which the desiccant bed regeneration process takes place, the ambient air is drawn and passed through the sections not having desiccant bed. The air at outlet from these sections gains temperature differences higher than when it passes through the section having desiccant bed (see Fig. 6) although the surface area subject to solar insolation is twice as that of the sections not having desiccant bed. This is due to the major part of the heat of regeneration (desorption heat) which is transferred from the air in the section having the desiccant bed to the desiccant bed. Also, it can be noticed from Figs. 5 and 6 that the nighttime dehumidification performance has a close relation with the regeneration process during daytime. The regeneration process depends on the solar radiation flux, during the daytime which has a direct consequent on the air dehumidification at nighttime. The psychometrics of the temperature humidity paths at two different times for the suggested system during the nighttime is
shown Fig. 7. It can be seen from the figure that, at location B (at 22:00), (see Fig. 1 for the locations B, C and A), the ambient air at entrance to the unit is at a dry bulb temperature of 30 °C and RH = 17%. It passes through the sections having desiccant bed and leaves these sections at location C (at 22:00) at which its absolute humidity is lowered and it is cooled by thermal radiation losses to the sky, to a dry bulb temperature of 25.9 °C and RH = 20.8%. Then, the air passes through the sections not having desiccant bed and it is further cooled with no changes in its absolute humidity value to a dry bulb temperature of 23.9 °C and RH = 23.5%. While, at point B (at 4:00) the ambient air is at a dry bulb temperature of 23 °C and RH = 24.8%. It enters the unit and passes through the sections having desiccant bed and leaves these sections at point C (at 4:00) at which its absolute humidity is lowered and it is cooled by thermal radiation losses to the sky, to a dry bulb temperature of 18 °C and RH = 32.9%. Then, the air passes through the sections not having desiccant and it is further cooled
Fig. 7. Psychometrics of the temperature humidity paths at two different times for the suggested system during the nighttime.
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Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
3
Tout,(0.48 m /min) 3 T C,(0.48 m /min) 3 Tout,(0.798 m /min) 3 TC (0.798 m /min)
o
Dry Bulb Temperature, C
90 80 70 60 50 40 30 20
3
10 0
Tout,(1.12 m /min) 3 TC (1.12 m /min)
Tambient 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 8. Effect of air volume flow rate on the predicted outlet air dry bulb temperature.
0.01
3
ω out,(0.48 m /min) 3 ω out,(0.798 m /min) 3 ω out,(1.12 m /min)
Humidity Ratio, kg/kg
0.009 0.008 0.007 0.006 0.005 0.004 0.003 0.002 0.001 0
ω ambient 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 9. Effect of air volume flow rate on the predicted outlet air absolute humidity ratio.
with no changes in its absolute humidity value to a dry bulb temperature of 16.7 °C and RH = 35.7%. However, at both times, the outlet air from the system if it is supplied directly to the residential space is still within the range of air supply conditions required for air conditioning. Effect of air mass flow rate on the performance In order to investigate the effect of varying the air mass flow rate on the predicted performance parameters (outlet air dry bulb temperature and absolute humidity) the weather data presented in Fig. 6 are input to the model. The air mass flow rate is varied with values of 0.0093, 0.0154 and 0.022 kg/s corresponding to air volume flow rate values of 0.48, 0.798 and 1.12 m3/min, and the output results are presented in Figs. 8 and 9. Also, the effect of the air mass flow rate on the desiccant bed water content is shown in Fig. 10, knowing that the initial bed water content on dry basis is 0.25 and its temperature is equal to that of the ambient at that time. It can be seen from Figs. 8 and 9 that, lower air volume flow rate produced minimum outlet temperature values during nighttime and the highest outlet temperature values during the day time as it is expected. At the same time, it has the minimum values of water vapor adsorption (nighttime) and desorption (daytime),
i.e. highest values of the outlet air absolute humidity ratio. This can be attributed to the lower values of both the convective heat and mass transfer coefficient between the air and the bed particles as it is dependent on the flow Reynolds number based on the desiccant particle diameter. However, as the air volume flow rate increases, the outlet air dry bulb temperature values during nighttime increase and during the daytime decrease followed by higher values of water vapor adsorption (nighttime) and desorption (daytime), i.e. lower values of the outlet air absolute humidity at nighttime. This is due to the increases in the values of both the convective heat and mass transfer coefficient between the air and the desiccant bed particles. It can be noticed from Fig. 9 that, the integration over time for the air absolute humidity ratio during regeneration period (daytime) is higher in comparison with the integration over time for the air absolute humidity ratio during adsorption period (nighttime). This is because the ambient air during the day time desorbed the water vapor which had been adsorbed during the previous nighttime plus a few of the initial water content of desiccant bed. The results seen in Fig. 10 are explained based on the low ambient air absolute humidity in hot arid areas, and as the air volume flow rate increases the initial bed water content during desorption decreases. However, the bed after a couple of days reaches to the equilibrium condition, i.e. the
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Bed Water Content, kg/kg
0.3 0.25 0.2 0.15 0.1 3
ω bed,(0.48 m /min) 3 ω bed,(0.798 m /min) 3 ω bed,(1.12 m /min)
0.05 0 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 10. Effect of air volume flow rate on the predicted desiccant bed water contents.
adsorbed water during nighttime will be equal to the desorbed water during daytime. Therefore, based on the above results of the effect of the air mass flow rate, in case of using the desiccant nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas, it is preferred to use a low air mass flow rate in order to obtain a lower outlet air temperature which is less than the ambient air by 5.5–7 °C and below 40% relative humidity. Conclusions This study examines the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas of Upper Egypt. A mathematical model analyzing the heat and mass transfer, for the proposed system during adsorption (nighttime mode) and regeneration (daytime mode) is established. The actual hourly values of weather data, of the average of 30 year measurements were used as the input to model. The model output results were verified with correspondent experimental ones and a good agreement existed. The model was used to investigate this system feasibility for air comfort application in hot arid areas of Upper Egypt and it was found that, this system is feasible to be used in hot arid areas of Upper Egypt. This is because the ambient air relative humidity is very low in summer daytime and increase by nighttime, which makes utilization of this system effective. The model is used to investigate the effect of the air mass flow rate on this system performance and the results indicate that, it is preferable to use a low air mass flow rate in order to obtain outlet air having a dry bulb temperature less than the ambient air with values ranging from 5.5 to 7 C at the same time it has relative humidity not higher than 40%.
References [1] Pesaran A, Terry RP, Czanderna AlW. Desiccant cooling: state-of-the-art assessment. DOE-NREL’s Report No. TP-254-4147, 1998. [2] Fairey P, Kerestecioglu A, Vieira R. Desiccant enhanced nocturnal radiation – a new passive cooling concept. In: Proceedings of the 10th national passive solar conference, October 15–20, Raleigh, NC, USA, 1985. p. 271–6.
[3] Fairey P, Kerestecioglu A, Vieira R. Analytical investigation of the desiccant enhanced nocturnal radiation cooling concept. FSEC-CR-152-86. Florida Solar Energy Center, Cape Canaveral, FL. Final Report, DOE Contract No. DE-AC0384SF12216, 1986. [4] Swami M, Rudd A, Fairey P, Patil S, Kerestecioglu A, Chandra S. An assessment of the DESiccant enhanced RADiative (DESRAD) cooling concept and description of the diurnal test facility. FSEC-CR-237-88. Florida Solar Energy Center, Cape Canaveral, FL. DOE Contract No. DE-FC03-86SF16305, 1989. [5] Swami M. Effect of desiccant geometry on the performance of the desiccant enhanced nocturnal radiative cooling system. In: Proceedings of the 1990, ASME international computers in engineering conference and exhibition, August 5–9, 1990. p. 425–31. [6] Lu S-M, Yan W-J. Development and experimental validation of a full-scale solar desiccant enhanced radiative cooling system. Renewable Energy 1995;6:821–7. [7] Lu S-M, Shyu R-J, Yan W-J, Chung TW. Development and experimental validation of two novel solar desiccant-dehumidification–regeneration systems. Energy 1995;20:751–7. [8] Saito Y. Regeneration characteristics of adsorbent in integrated desiccant/ collector. J Sol Energy Eng 1993;115:169–75. [9] Techajunta S, Chirarattananon S, Exell RHB. Experiments in a solar simulator on solid desiccant regeneration and air dehumidification for air conditioning in a tropical humid climate. Renewable Energy 1999;17:549–68. [10] Ali A Hamza H, Taha IMS, Ismail IM. Cooling of water flowing through a night sky radiator. Sol Energy 1995;55:235–53. [11] Lavan A, Monnier JB, Worek WM. Second law analysis of desiccant cooling systems. J Sol Eng 1982;104:229–36. [12] Duffie JA, Beckman WA. Solar engineering of thermal processes. second ed. John Wiley & Sons, Inc.; 1991. [13] Diab MR. Experimental and analytical study of the heat transfer characteristics of a solar air heater incorporated a finned absorber. Ph.D. thesis, Purdue University, 1981. [14] Heaton HS, Reynolds NC, Kays WM. Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int J Heat Mass Transfer 1964;7:763–81. [15] Pesaran AA, Mills AF. Moisture transport in silica gel packed beds: I. Theoretical study. Int J Heat Mass Transfer 1987;30:1037–49. [16] Pesaran AA, Mills AF. Moisture transport in silica gel packed beds: II. Experimental study. Int J Heat Mass Transfer 1987;30:1051–60. [17] ASHRAE. 2005 ASHRAE handbook – fundamentals. Atlanta, GA: American Society of Heating Refrigerating and Air-Conditioning Engineers; 2005. [18] Press WH, Teukolsky SA, Vetterling WT, Flamery BP. Numerical recipes in FORTRAN – the art of scientific computing. 2nd ed. Cambridge University Press; 1992. [19] Helen-Xing H-Y. Desiccant dehumidification analysis. M.Sc. thesis, Massachusetts Institute of Technology, 2000. [20] Degelman LO. Enercalc: a weather and building energy simulation model using fast hour-by-hour algorithms. In: Proceedings of the fourth national conference on microcomputer applications in energy, April 25–27, Tuscon, Arizona, 1990. [21] Degelman LO. A statistically-based hourly weather data generator for driving energy simulation and equipment design software for buildings. In: Proceedings of the building simulation ‘91, international building performance simulation association (IBPSA), August 20–22, Nice, SophiaAntipolis, France, 1991.
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Sustainable Energy Technologies and Assessments journal homepage: www.elsevier.com/locate/seta
Original Research Article
Desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas Ahmed Hamza H. Ali ⇑ Department of Energy Resources and Environmental Engineering, Egypt-Japan University of Science and Technology (E-JUST), P.O. Box 179, New Borg El-Arab City, Alexandria 21934, Egypt
a r t i c l e
i n f o
Article history: Received 23 July 2012 Revised 26 January 2013 Accepted 27 January 2013
Keywords: Desiccant cooling Nocturnal radiative cooling Solar cooling Air comfort Hot arid areas
a b s t r a c t In this study, the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas of Upper Egypt is carried out using the analytical techniques. The system is investigated based on a daily cycle one-way airflow direction and the measured weather data. In addition, a mathematical model for analyzing the heat and mass transfer in the system during adsorption (nighttime mode) and regeneration (daytime mode) is established. The obtained results were verified with correspondent experiments and a good agreement existed. Thereafter, the model is used to investigate the system feasibility for air comfort application. It is found that this system can be used in hot arid areas of Upper Egypt at which the ambient air relative humidity is very low in summer daytime and gradually increases during nighttime that makes utilization of this system effective. Also, the model is used to investigate the effect of the air mass flow rate on this system performance and the results show that, it is preferred to use a low air mass flow rate in order to obtain a lower outlet air temperature which is less than the ambient air by 5.5–7 °C and below 40% relative humidity. Ó 2013 Elsevier Ltd. All rights reserved.
Introduction Passive cooling of buildings within the human comfort zone is an important subject due to its goal of keeping the environment out of contamination. The upper atmosphere (the sky) is a heat sink for nocturnal emitted longwave radiation by terrestrial units to produce a net cooling. Desiccant based systems process atmospheric air by removing moisture from the air to a desired level and then cooling it using a heat exchanger, followed by some evaporative cooling. However, the desiccant based systems are generally effective in hot humid areas. Rotary desiccant wheels and fixed desiccant beds are the most common desiccant dehumidifier configurations. Pesaran et al. [1] presented a comprehensive desiccant cooling bibliography containing 1176 pieces of literature available up to the mid-1997. Among the commercially available systems, rotary desiccant wheels allow continuous operation, while a fixed desiccant bed is flexible in positioning but cannot run continuously. Usually, more than one fixed desiccant bed unit is used to compensate for the non-continuous operation drawback. One fixed desiccant bed can be in regeneration while another is in dehumidification process. A different desiccant cooling idea called desiccant enhanced nocturnal radiation cooling and dehumidifica-
⇑ Tel.: +20 3 4599839; fax: +20 12 23971265. E-mail address: [email protected] 2213-1388/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.seta.2013.01.003
tion has been proposed descriptively by Fairey et al. [2,3] and Swami et al. [4]. Swami [5] quantified this idea and clarified its benefits when applied to a residence in hot humid areas. The idea of the integrated desiccant/enhanced nocturnal radiative cooling-solar regenerated system is summarized as follows. A unit is composed of a plane structure filled with desiccant material and topped with a highly conductive and emissive radiator plate. Ambient air is withdrawn through the desiccant bed during nighttime where dehumidification with probably some cooling occurs. The desiccant regeneration cycle occurs during daytime using ambient air and solar radiation. The dehumidified air at night can be further cooled evaporatively and stored to be utilized later to cool space within building during the warm daytime or it can be used directly to cool and dehumidify the building during the night time. Lu et al. [6,7] experimentally investigated a desiccant enhanced nocturnal radiation cooling and the dehumidification system. They reported that there are several characteristics of this system: firstly, it can continuously supply the conditioned space with processed (or dehumidified) air 24 h a day; secondly, it contains many functions which are suitable for public use like air conditioning, drying, dehydration and active cooling. Thirdly, since the only energy consumption of the system is from the ventilation fans plus the enormous free solar energy input to regenerate the system, its energy efficiency is incredibly high, and finally, a large potential for dehumidification in such a humid area as Taiwan is expected.
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Nomenclature area, m2 free flow area inside the bed, m2 surface area of the desiccant bed, m2 plate base surface area and gfinAfins, m2 specific heat of humid air, J/kg K specific heat of the desiccant, J/kg K specific heat of dry air, J/kg K specific heat of water vapor, J/kg K specific heat of liquid water, J/kg K channel height, m configuration factor adsorption heat, J/(kg of adsorbed water) enthalpy of humid air, J/kg convective heat transfer coefficient, W/m2 K enthalpy of wet desiccant (energy content per unit mass of dry desiccant), J/kg hg specific enthalpy of water vapor, J/kg mass transfer coefficient, kg/m2s hm i, j and k element numbers L length of the unit, m l offset fins spacing, m _ air m mass flow rate of the air stream, kg/s _ water m mass flow rate of water from the bed to the air stream (Fig. 2), kg/s Noss number of surrounding elements q heat transfer flux, W/m2 qsolar total incident solar radiation flux on the plane of the unit, W/m2 Q heat transfer rate, W RH relative humidity of hypothetical air layer A Abed,air Abed,s Apl-iner Cpair Cdes Cpair,dry Cpwv Cw D F qads Hair hbed,air Hdes
Saito [8] and Techajunta et al. [9] investigated the regeneration in an integrated desiccant/collector system using a direct solar energy simulator as a heat source through experiments and mathematical simulations. Their desiccant bed configuration was a layer of adsorbent laying along the diagonal of the collector under glass cover and subjected directly to the radiation flux and their units have no thermal radiation cooling capability. In their model they analyzed the regeneration process only and considered the whole system is one lump in the energy balance equation. In the desiccant enhanced nocturnal radiative cooling-solar collector system air is used as a working fluid and the heat transfer medium for the heat of adsorption and regeneration to/from the absorber/radiator plate. The disadvantage of the air as a heat transfer medium is the low convective heat transfer coefficient due to its low thermophysical properties. Throughout the literature on utilization of the desiccant system it is clear that the desiccant dehumidification cooling techniques are more effective when utilized in hot humid areas. While, the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas is not available in the literature. Therefore, in this study, using the analytical techniques, an evaluation study is carried out with a detailed simulation model considering the heat transfer modes within the suggested system, and, mainly implementing the feasibility of this system for air comfort application in hot arid areas of Upper Egypt. In Upper Egypt, the ambient air relative humidity is very low in summer daytime and slightly increases by nighttime, which makes utilization of this system feasible for air cooling and dehumidification. The suggested configuration is an absorber/radiator plate without upper glazing in order to utilize the convection losses to the ambient to enhance the plate cooling
t Tair Tbed Vbed W x
time, s temperature of the air stream, °C temperature of desiccant bed, °C volume of desiccant bed element, m3 channel width, m length scale m
Greek symbols a absorptance, – d offset fin height, m e emittance, – r Stefan–Boltzmann constant, (5.67051 108), W/m2 K4 gfin fin efficiency xair humidity ratio of air stream xbed desiccant bed water content on dry basis xs,bed humidity ratio of hypothetical air layer qbed bulk density of the desiccant bed, kg/m3 Subscripts air dry air stream amb ambient bed desiccant bp back plate conv convection pl absorber/radiator plate rad thermal radiation ss surrounding surfaces s surface of the desiccant bed sky sky wv water vapor
and also to enhance nocturnal radiative cooling. This absorber/ radiator plate is designed to have a set of offset fins on its backside parallel to the flow direction in order to improve the convected heat transfer process from the plate to the air stream. It is also aimed to examine the effect of air mass flow rate on the system performance. System concepts and description of the operating modes The nocturnal radiative cooling of a surface exposed to the sky can be used to lower the temperature of a fluid beneath the radiator plate below that of surrounding ambient air temperature. Detailed explanation of this phenomenon was presented by Ali et al. [10]. As the radiator plate temperature rises the radiative heat rejection rate to the sky increases. Desiccation is a process by which highly hygroscopic materials adsorb moisture from the air. This process is combined by a release of thermal energy roughly equal to the normal heat of water vapor condensation and heat of wetting. A desiccant system can fall into one of two generic operating concepts adiabatic or isothermal. Since the desiccation process in isothermal systems operates at a lower temperature than that in adiabatic systems as reported by Lavan et al. [11], the moisture absorption capacity of isothermal systems is somewhat greater than that of adiabatic systems. However, the concept of the desiccant enhanced nocturnal radiative cooling-solar collector system attempts to take advantage of approaching isothermal desiccation case through nocturnal radiative cooling. The radiator plate is coupled thermally to the desiccant bed during the adsorption process leading to a radiator temperature which tends to remain equal to or less than the ambient by about 3–5° due to heat of adsorption. Thus, a combination of the moisture removal
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Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
qrad,pl,sky
A
Tair,in
D
air,in
qsolar
element i
qconv,pl,amb
absorber/radiator plate
Tair,out
C
air,out
Daytime mode
x
back plate
Thermal Insulation
B
Nighttime mode
(a)
l fin desiccant bed
E
L (b)
Fig. 1. Schematic of the proposed desiccant nocturnal radiative cooling-solar collector system (a) cross section and (b) plane view.
capabilities of desiccant materials and, heat rejection capabilities of nocturnal sky radiation is achieved in such a way that each process complements and enhances the other. The proposed desiccant enhanced nocturnal radiative coolingsolar collector system for the present analysis is shown in Fig. 1. It consists of an absorber/radiator plate made of an aluminum sheet coated with black paint at both sides and has a set of offset fins on its backside parallel to the flow direction. A back plate forms with the absorber/radiator plate a channel having six sections four of them contain desiccant bed laying along the diagonal (best configuration due to results of Swami [5]) across the air flow path. The absorber/radiator plate is exposed to the ambient while the back plate of the channel is thermally well insulated. Ambient air flows through the system from the left or the right side (location A or B) based on the operation mode. At nighttime ambient air is flown from location B to location C for adsorption mode and is dehumidified by the desiccant bed. The bed releases heat, which warms the air and consequently the absorber/radiator plate. Thus, the warmer absorber/radiator plate has a greater cooling potential with respect to the ambient and from location C to location A the rest of absorber/radiator plate (two sections) serves to cool the air stream without dehumidification. The out-coming air at location A is more dry and cool than at the inlet location B. During the daytime, regeneration mode, the ambient air is flown through the system from location A to location C, where it is heated in the first two sections of the absorber/radiator plate by the absorbed solar radiation. Thus, at the beginning of the section having desiccant bed at location C, this air has a relatively high temperature and low relative humidity. This warm air combined with the heat exchange between the absorber/radiator plate and bottom plate with the desiccant bed serves to regenerate the bed from location C to location B. The air stream is then exhausted to the ambient at location B.
Analytical approach Simulation of this suggested system is carried out based on a daily cycle one-way airflow direction and the measured weather data for Assiut, Egypt as the model inputs. The system attempts to combine the moisture removal capabilities of desiccant materials, and heat rejection capabilities of nocturnal sky radiation in such a way that each process complements and enhances the other, and at the same time provides isothermal desiccation which, allows for lower regeneration temperature that provides an avenue for using a solar driven desiccant system. A mathematical model for analyzing the heat and mass transfer, for the proposed system shown in Fig. 1, during adsorption (nighttime mode) and regeneration (daytime mode) is obtained by applying the energy and mass balances on its parts. This is done in three parts, the first part is for
the plates, the second is for the section not having desiccant bed and the third part is for the section having desiccant beds at which the mass balance is applied. The model is established based on the following assumptions. One-dimensional temperature and moisture gradient in both air stream and desiccant bed are considered along the flow direction and the flow channel is divided into n equal elements. Accumulations of energy and moisture in the air elements are neglected. The axial heat conduction in the plates and heat and mass transport in the air stream along the x-axis are neglected. The convection heat and mass transfer coefficients between the air stream and the desiccant bed are constant over the sections between locations B and C of Fig. 1. The desiccant bed is in equilibrium with a thin air layer on the surface of desiccant particles. The mass transfer driving force is the water vapor pressure difference between this air layer and the processing moist air stream. The pressure drop along the desiccant beds does not affect the air stream velocity. For the offset fins, the fins are considered to have the base temperature of the absorber/radiator plate with surface area equal to gfinAfin. Conduction between plates and desiccant bed ends is ignored. Each desiccant bed element is considered as a lumped heat capacity system. Surfaces inside the duct are black for longwave radiation. Energy storage in plates and fins is neglected. Energy balance For plates The energy balance equations are applied on the plate elements as follows: For the absorber/radiator plate element (i).
apl;i Apl;i qsolar hpl;amb Apl;i ðT pl;i T amb Þ repl;i Apl;i F pl;sky ðT 4pl;i T 4sky Þ Noss X hpl;air Apl-iner;i ðT pl T air Þi r Apl-iner;i F pl;ij ðT 4pl;i T 4j Þ ¼ 0
ð1Þ
j¼1
The terms in Eq. (1) are as follows; the first one is the solar radiation absorbed on the plane of the outer surface of the absorber/radiator plate; the second is representing the convection to the ambient air, and, the value of hpl,amb is determined using the formula presented in Duffie and Beckman [12]; the third is the net thermal radiant energy exchange with the sky; details of calculation sky temperature are in Ali et al. [10]. While, the fourth
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Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
term is the convection heat transferred to the air stream in the channel, and, the value of hpl,air is calculated using the formula presented in Diab [13]. The fifth one is the thermal radiant exchange between the absorber/radiator plate element i and all other solid surfaces (Noss) including any desiccant surface it sees. For the back plate element (i).
hbp;air Abp;i ðT bp T air Þi þ r
Noss X Abp;i F bp;ij ðT 4bp;i T 4j Þ ¼ 0
ð2Þ
where hbed,air is the convective heat transfer between the desiccant bed and the flowing air and is given in [4]; the third term in Eq. (4) and the fourth term in Eq. (5) are the energy transferred to the air stream due to the heat of adsorption that is a summation of the normal heat of condensation plus the heat of wetting; the first term in Eq. (5) is the rate of energy storage in the desiccant bed; the second term in Eq. (5) is the net thermal radiant exchange from the desiccant bed element i to all other solid surfaces (Noss) including any desiccant surface it sees.
j¼1
The two terms in Eq. (2) are defined as follows: the first one is the heat convected to the air stream from the back plate in the duct, and, the value of hbp,air is calculated using the correlation equation for the data of Heaton et al. [14], while the second one is the net thermal radiant exchange between the back plate element i and all other solid surfaces (Noss) including any desiccant surface it sees. For sections not having desiccant beds The energy balance equation is applied on the air stream element as follows: For air stream element (i).
hpl;air Apl-iner;i ðT pl T air Þi þ hbp;air Abp;i ðT bp T air Þi _ air Cpair ðT air;in T air;out Þi ¼ 0 þm
ð3Þ
The terms in Eq. (3) are defined as follows; the first is the convected heat to the air stream in the duct from the absorber/radiator plate; the second is the convected heat to the air stream in the duct from the back plate; the third is the net energy transfer due to inflow and outflow. For sections having desiccant bed. The equations governing the energy balance for the desiccant bed and air stream elements i, shown in Fig. 2, are given as follows: (i) For air stream node (i)
_ air ðHair;in Hair;out Þi þ hbed;air Abed;airÞi ðT bed T air Þi m qads hm Abed;airÞi ðxs;bed xair Þi þ hpl;airÞi Apl;i ðT pl T air Þi þ hbp;airÞi Abp;i ðT bp T air Þi ¼ 0
ð4Þ
(ii) For desiccant bed node (i)
qbed V bed
Noss X @Hbed þ r Abed;sÞi F bed;ik ðT 4bed;i T 4k Þ @t k¼1
þ hbed;air Abed;air ðT bed T air Þi qads hm Abed;airÞi ðxs;bed xair Þi ¼ 0
ð5Þ
The first term in Eq. (4) is the net energy transfer due to inflow and outflow; the second term in Eq. (4) and the third term in Eq. (5) are the net convected heat transfer from the bed to the air stream,
Mass balance The equations governing the mass transfer for the desiccant bed and air stream elements i, shown in Fig. 2, are given as follows: For sections having desiccant bed (i) For air stream node (i)
_ air ðxair;in xair;out Þi hm Abed;air;i ðxs;bed xair Þi ¼ 0 m
ð6Þ
(ii) For desiccant bed node (i)
qbed V bed
@ xbed @t
hm Abed;air;i ðxs;bed xair Þi ¼ 0
ð7Þ
The first term in Eq. (6) is the water mass transfer rate due to the inflow and outflow of the control volume. The second term in Eqs. (6) and (7) is the convective mass transfer from the bed to the air stream, where hm is the convective mass transfer between the desiccant bed and the flowing air and is given in [4]. The first term in Eq. (7) is the rate of water storage for the desiccant bed. The value of Abed,air,i is the product of the transfer area of unit volume of the desiccant bed multiplied by the desiccant bed element volume; it is a desiccant property. In Eqs. (4)–(7), the positive signs are for the desiccant bed regeneration and negative signs for air dehumidification modes. Desiccant equilibrium properties The set of equations governing the dynamics of the sorption process has to be solved along with equilibrium sorption isotherms of the desiccant bed. The sorption isotherms for silica gel bed relating the air stream relative humidity with the desiccant bed humidity ratio and the heat of adsorption of water vapor in silica gel reported by Pesaran and Mills [15,16] and used in this study are given by:
RH ¼ 0:0078 0:05759xbed þ 24:1655x2bed 124:478x3bed þ 204:226x4bed hads ¼ ð12; 400xbed þ 3500Þ hads ¼ ð1400xbed þ 2950Þ
ð8Þ
xbed < 0:05 xbed > 0:05
ð9aÞ ð9bÞ
The ASHRAE [17] psychometrics for moist air is used to calculate xair in Eqs. (6) and (7). The enthalpy values of both the moisture air Hair and wet desiccant bed Hdes are determined from the following equations:
Hair ¼ Cpair;dry T air þ xair ðhg þ Cpwv T air Þ
ð10Þ
Hdes ¼ ðC des þ xdes C w ÞT des
ð11Þ
Numerical method of solution
Fig. 2. Energy and mass transfer for desiccant bed control volume.
Simulation of this system was carried out based on a daily continuous operation, one direction of airflow from ambient and inlet to the system is in direction from location B to location A during
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
Results and discussions Assessment of the theoretical results Before reporting and discussing the feasibility of using this system as well as the effect of varying the air mass flow rate on its performance, the results of the model are compared with the corresponding measured data available in the nearest similar system of Lu and Yan [6]. The unit dimensions of [6] were input to the present model while assuming any missed data. The dimensions of the experimental setup of Lu and Yan [6] which were used in model verification are as follows: total channel flow length L = 3.6 m, flow lengths without desiccant bed E = 2.0 m, channel width W = 1.4 m and channel height D = 20 mm; and, the weather data as present in Lu and Yan [6]. Fig. 3 shows the input weather data to the model and a comparison between the calculated values of the outlet air dry bulb temperature and absolute humidity ratio with corresponding experimentally measured values of [7]. It is clear from the figure that, the predicted humidity ratio values are in good agreement with the measured data. Also, it is seen from the figure that the predicted air dry bulb during daytime is slightly higher by 1 to 1.5 C than the measured one, while its values during nighttime are in reasonable agreement with the measured data. This slight deviation during daytime can be attributed to using the offset fins on the backside of the absorber leading to a higher rate of heat transfer to the air and in the same time decreasing the convection losses from the absorber plate to the surrounding. However, these agreements of the calculated values with the measured data give confidence in the analysis techniques to investigate the feasibility of using this system in Upper Egypt and the effect of varying the air mass flow rate on its performance. Performance characteristics and system feasibility The outlet absolute humidity ratio and air dry bulb temperature are the characteristics of the performance for the suggested system, desiccant enhanced nocturnal radiative cooling-solar collector was intended to be used for air comfort application in hot summer month in hot arid areas of Upper Egypt. An overall look at the year
60
air,out air,out
50 40
0.03 30
0.02 20
0.01
10 3
air mass flow rate=0.0112 kg/s (0.56 m /min) 0
2
4
6
900
Tout [Lu et al., 1995] Tout present 0.04
air,in
0
1000
[Lu et al., 1995] present
8
10
12
14
16
18
20
22
800 700 600 500 400 300 200 100
24
Time, hr Fig. 3. Verification of the calculated parameters and corresponding data of [6].
2
qsolar Tin
o
Dry Bulb Temperature, C
70
study, are as follows: total channel flow length L = 1.2 m, flow length without desiccant bed E = 0.6 m, channel width W = 0.75 m, channel height D = 50 mm, fins pitch l = 148 mm, fin height d = 40 mm and fin length lfin = 90 mm.
Total Solar Insolation (W/m )
nighttime, while, it is in the opposite direction from location A to location B during daytime. In practice, for this system, the air is in cyclic mode to provide a continuous operation. With the system continuing operation, the last condition that forms the dehumidification process at end of nighttime mode is the initial conditions for the regeneration process during the next daytime mode. The temperatures of the solid boundary elements were obtained by solving the energy balance Eqs. (1)–(5) for n-module model (in this study n = 20,000) which generates (4n) finite central difference nonlinear equations in (4n) unknowns. These (4n) unknowns are the temperatures of the absorber/radiator plate, air stream bulk temperature, desiccant bed and back plate elements respectively. For solving this set of nonlinear heat balance equations, the multidimensional secant method – Broyden’s method – which is presented in Press et al. [18] was used. The calculation is carried out in quasi steady at one minute time step. While, this set of equations is a function of the air and desiccant bed moisture content, the moisture content of the air and desiccant bed elements are calculated first for a certain time, keeping the temperature constant, from the mass balance Eqs. (6) and (7). These computation procedures are repeated until the change in the temperatures of the solid boundary elements is within the allowable error (below 105). The desiccant bed material considered in this study is silica gel grade 01, the properties of this grade and the values of both the convective heat and mass transfer coefficients, which were defined, based on Pseudo-Gas-side Controlled model (PGC) were reported by Helen-Xing [19]. The values of hpl,air and Tsky in Eq. (1) were determined using the formulae presented by Duffie and Beckman [12]. The calculation starts with all system components having the same initial condition equal to that of the ambient and the bed water content on dry mass basis xbed(s = 0) = 0.25. Simulation should be based on actual hourly values of weather data in order to decide the feasibility of this system. Therefore, in this study, the hourly average values of 30 years measurements of the weather parameters for Assiut, at Upper Egypt (Lat. 27°120 and long. 31°100 ). The method to calculate these hourly average values that is reported in Degelman [20,21] is used in this study. The radiative (optical) properties used in the model are considered as follows: the absorptance (emittance) of the metallic surfaces, which are coated with a commercial black paint is 0.94 and the emittance of the silica gel is 0.9 as given by Fairey et al. [3]. The flow (air) mass flow rate varied from 0.0094 to 0.022 kg/ s corresponding to air volume flow rate ranging from 0.479 to 1.12 m3/min. The dimensions for Fig. 1, which are used in this
Humidity Ratio
58
59
Solar Radiation Wind speed
1200 1100 1000 900 800 700 600 500 400 300 200 100 0
60
50
40
30
20
10
0
6
12
June
18
24
2
Air dry bulb temprature Air RH %
o
T ( C), RH% and Vw (m/s)
70
Total Solar Insolation (W/w )
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
30
Fig. 4. Weather data for June.
Tambient 80
ambient
Tout
Tlocation
C
out
60 50 40 0.01
30 20
0.008
10 0
0.006
Humidity Ratio (kg/kg)
o
Dry Bulb Temperature, C
70
0.004
6
12
June
18
24
30
Fig. 5. Predicted absolute humidity ratio and air dry bulb temperature at outlet and at location C for air mass flow rate of 0.01251 kg/s (0.64 m3/min).
weather data for Assiut, Egypt indicates that the maximum air dry bulb temperature was recorded through the month of June. The weather data of June are shown in Fig. 4 and are the input to the model. The predicted outlet absolute humidity ratio and air dry bulb temperature from the system and the statues of the dry bulb temperature at location C of Fig. 1 for the air mass flow rate of 0.01251 kg/s (0.64 m3/min) are shown in Fig. 5. As seen in Fig. 4, the ambient air dry bulb temperature during nighttime varied between 30 and 16 °C and the relative humidity reaches up to 47%. However, these weather conditions are applicable to desiccant systems using silica gel as a desiccant and work well. As seen from Fig. 5, during nighttime when the ambient air passes through the sections having desiccant bed (from B to C of Fig. 1) its dry bulb temperature and absolute humidity ratio are lowered with values ranging from 3 to 4 C and 0.13 to 0.75 g/kg respectively, further decrease in its dry bulb temperature occurs when it passes through the sections not having desiccant bed (from C to A of Fig. 1) with values ranging from 2.5 to 3 C. As a result, the total air dry bulb temperature drop ranged from 5.5 to 7 C. These values indicate that as the ambient air is withdrawn through the suggested system during nighttime it is dehumidified and cooled. However, this cool and dehumidified air at nighttime can be further cooled evaporatively and stored to be utilized later to cool space during the warm daytime or it can be used directly to cool and dehumidify
the building during the nighttime. Therefore, this system is feasible in use for air comfort applications in hot arid areas of Upper Egypt, at which, the ambient air relative humidity is very low in summer daytime and increases by nighttimes, which makes utilization of this system effective. In order to clarify and magnify the characteristic details of this system performance, weather data for three consecutive days (72 h) were input to the model. The corresponding predicted results of outlet air absolute humidity and air dry bulb temperature in the case of air mass flow rate of 0.01251 kg/s (0.64 m3/min) are shown in Fig. 6. As seen from the figure, at nighttime when the ambient air passes through the sections having desiccant bed its temperature is decreased. This is because it has a low absolute humidity value, and consequently, a small amount of water vapor is adsorbed to the desiccant bed leading to a released adsorption heat less than the net thermal radiation loss to the sky which causes a net cooling effect. As the air passes through the section dose not having desiccant bed (whose surface area of this section is equal to one third of the total system surface area as shown in Fig. 1) it is further cooled by 3–4 C by the effect of the net thermal radiation losses to the sky. Determination of the night sky radiation cooling rate is carried out ranging from 75 to 85 W/m2. These values are expected in hot arid areas of Upper Egypt due to a small amount of downward thermal radiation because of the low ambi-
Ahmed Hamza H. Ali / Sustainable Energy Technologies and Assessments 1 (2013) 54–62
3
1200 q solar T ambient 1100 1000 ambient 900 T out T at C 800 out 700 600 0.006 500 0.005 400 0.004 300 0.003 200 0.002 100
/min)
70 60 50 40
Humidity Ratio
o
Dry Bulb Temperature, C
air mass flow rate=0.01251 kg/s (0.64 m
30 20 10 0
0
6
12
18
24
30
36
42
48
54
60
66
2
80
Total Solar Insolation (W/m )
60
72
Time, hr Fig. 6. The predicted absolute humidity and air dry bulb temperature at outlet for the presented values of weather data and air mass flow rate.
ent water vapor content, at the daytime, at which the desiccant bed regeneration process takes place, the ambient air is drawn and passed through the sections not having desiccant bed. The air at outlet from these sections gains temperature differences higher than when it passes through the section having desiccant bed (see Fig. 6) although the surface area subject to solar insolation is twice as that of the sections not having desiccant bed. This is due to the major part of the heat of regeneration (desorption heat) which is transferred from the air in the section having the desiccant bed to the desiccant bed. Also, it can be noticed from Figs. 5 and 6 that the nighttime dehumidification performance has a close relation with the regeneration process during daytime. The regeneration process depends on the solar radiation flux, during the daytime which has a direct consequent on the air dehumidification at nighttime. The psychometrics of the temperature humidity paths at two different times for the suggested system during the nighttime is
shown Fig. 7. It can be seen from the figure that, at location B (at 22:00), (see Fig. 1 for the locations B, C and A), the ambient air at entrance to the unit is at a dry bulb temperature of 30 °C and RH = 17%. It passes through the sections having desiccant bed and leaves these sections at location C (at 22:00) at which its absolute humidity is lowered and it is cooled by thermal radiation losses to the sky, to a dry bulb temperature of 25.9 °C and RH = 20.8%. Then, the air passes through the sections not having desiccant bed and it is further cooled with no changes in its absolute humidity value to a dry bulb temperature of 23.9 °C and RH = 23.5%. While, at point B (at 4:00) the ambient air is at a dry bulb temperature of 23 °C and RH = 24.8%. It enters the unit and passes through the sections having desiccant bed and leaves these sections at point C (at 4:00) at which its absolute humidity is lowered and it is cooled by thermal radiation losses to the sky, to a dry bulb temperature of 18 °C and RH = 32.9%. Then, the air passes through the sections not having desiccant and it is further cooled
Fig. 7. Psychometrics of the temperature humidity paths at two different times for the suggested system during the nighttime.
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3
Tout,(0.48 m /min) 3 T C,(0.48 m /min) 3 Tout,(0.798 m /min) 3 TC (0.798 m /min)
o
Dry Bulb Temperature, C
90 80 70 60 50 40 30 20
3
10 0
Tout,(1.12 m /min) 3 TC (1.12 m /min)
Tambient 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 8. Effect of air volume flow rate on the predicted outlet air dry bulb temperature.
0.01
3
ω out,(0.48 m /min) 3 ω out,(0.798 m /min) 3 ω out,(1.12 m /min)
Humidity Ratio, kg/kg
0.009 0.008 0.007 0.006 0.005 0.004 0.003 0.002 0.001 0
ω ambient 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 9. Effect of air volume flow rate on the predicted outlet air absolute humidity ratio.
with no changes in its absolute humidity value to a dry bulb temperature of 16.7 °C and RH = 35.7%. However, at both times, the outlet air from the system if it is supplied directly to the residential space is still within the range of air supply conditions required for air conditioning. Effect of air mass flow rate on the performance In order to investigate the effect of varying the air mass flow rate on the predicted performance parameters (outlet air dry bulb temperature and absolute humidity) the weather data presented in Fig. 6 are input to the model. The air mass flow rate is varied with values of 0.0093, 0.0154 and 0.022 kg/s corresponding to air volume flow rate values of 0.48, 0.798 and 1.12 m3/min, and the output results are presented in Figs. 8 and 9. Also, the effect of the air mass flow rate on the desiccant bed water content is shown in Fig. 10, knowing that the initial bed water content on dry basis is 0.25 and its temperature is equal to that of the ambient at that time. It can be seen from Figs. 8 and 9 that, lower air volume flow rate produced minimum outlet temperature values during nighttime and the highest outlet temperature values during the day time as it is expected. At the same time, it has the minimum values of water vapor adsorption (nighttime) and desorption (daytime),
i.e. highest values of the outlet air absolute humidity ratio. This can be attributed to the lower values of both the convective heat and mass transfer coefficient between the air and the bed particles as it is dependent on the flow Reynolds number based on the desiccant particle diameter. However, as the air volume flow rate increases, the outlet air dry bulb temperature values during nighttime increase and during the daytime decrease followed by higher values of water vapor adsorption (nighttime) and desorption (daytime), i.e. lower values of the outlet air absolute humidity at nighttime. This is due to the increases in the values of both the convective heat and mass transfer coefficient between the air and the desiccant bed particles. It can be noticed from Fig. 9 that, the integration over time for the air absolute humidity ratio during regeneration period (daytime) is higher in comparison with the integration over time for the air absolute humidity ratio during adsorption period (nighttime). This is because the ambient air during the day time desorbed the water vapor which had been adsorbed during the previous nighttime plus a few of the initial water content of desiccant bed. The results seen in Fig. 10 are explained based on the low ambient air absolute humidity in hot arid areas, and as the air volume flow rate increases the initial bed water content during desorption decreases. However, the bed after a couple of days reaches to the equilibrium condition, i.e. the
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Bed Water Content, kg/kg
0.3 0.25 0.2 0.15 0.1 3
ω bed,(0.48 m /min) 3 ω bed,(0.798 m /min) 3 ω bed,(1.12 m /min)
0.05 0 0
6
12
18
24
30
36
42
48
54
60
66
72
Time, hr Fig. 10. Effect of air volume flow rate on the predicted desiccant bed water contents.
adsorbed water during nighttime will be equal to the desorbed water during daytime. Therefore, based on the above results of the effect of the air mass flow rate, in case of using the desiccant nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas, it is preferred to use a low air mass flow rate in order to obtain a lower outlet air temperature which is less than the ambient air by 5.5–7 °C and below 40% relative humidity. Conclusions This study examines the feasibility of implementing desiccant enhanced nocturnal radiative cooling-solar collector system for air comfort application in hot arid areas of Upper Egypt. A mathematical model analyzing the heat and mass transfer, for the proposed system during adsorption (nighttime mode) and regeneration (daytime mode) is established. The actual hourly values of weather data, of the average of 30 year measurements were used as the input to model. The model output results were verified with correspondent experimental ones and a good agreement existed. The model was used to investigate this system feasibility for air comfort application in hot arid areas of Upper Egypt and it was found that, this system is feasible to be used in hot arid areas of Upper Egypt. This is because the ambient air relative humidity is very low in summer daytime and increase by nighttime, which makes utilization of this system effective. The model is used to investigate the effect of the air mass flow rate on this system performance and the results indicate that, it is preferable to use a low air mass flow rate in order to obtain outlet air having a dry bulb temperature less than the ambient air with values ranging from 5.5 to 7 C at the same time it has relative humidity not higher than 40%.
References [1] Pesaran A, Terry RP, Czanderna AlW. Desiccant cooling: state-of-the-art assessment. DOE-NREL’s Report No. TP-254-4147, 1998. [2] Fairey P, Kerestecioglu A, Vieira R. Desiccant enhanced nocturnal radiation – a new passive cooling concept. In: Proceedings of the 10th national passive solar conference, October 15–20, Raleigh, NC, USA, 1985. p. 271–6.
[3] Fairey P, Kerestecioglu A, Vieira R. Analytical investigation of the desiccant enhanced nocturnal radiation cooling concept. FSEC-CR-152-86. Florida Solar Energy Center, Cape Canaveral, FL. Final Report, DOE Contract No. DE-AC0384SF12216, 1986. [4] Swami M, Rudd A, Fairey P, Patil S, Kerestecioglu A, Chandra S. An assessment of the DESiccant enhanced RADiative (DESRAD) cooling concept and description of the diurnal test facility. FSEC-CR-237-88. Florida Solar Energy Center, Cape Canaveral, FL. DOE Contract No. DE-FC03-86SF16305, 1989. [5] Swami M. Effect of desiccant geometry on the performance of the desiccant enhanced nocturnal radiative cooling system. In: Proceedings of the 1990, ASME international computers in engineering conference and exhibition, August 5–9, 1990. p. 425–31. [6] Lu S-M, Yan W-J. Development and experimental validation of a full-scale solar desiccant enhanced radiative cooling system. Renewable Energy 1995;6:821–7. [7] Lu S-M, Shyu R-J, Yan W-J, Chung TW. Development and experimental validation of two novel solar desiccant-dehumidification–regeneration systems. Energy 1995;20:751–7. [8] Saito Y. Regeneration characteristics of adsorbent in integrated desiccant/ collector. J Sol Energy Eng 1993;115:169–75. [9] Techajunta S, Chirarattananon S, Exell RHB. Experiments in a solar simulator on solid desiccant regeneration and air dehumidification for air conditioning in a tropical humid climate. Renewable Energy 1999;17:549–68. [10] Ali A Hamza H, Taha IMS, Ismail IM. Cooling of water flowing through a night sky radiator. Sol Energy 1995;55:235–53. [11] Lavan A, Monnier JB, Worek WM. Second law analysis of desiccant cooling systems. J Sol Eng 1982;104:229–36. [12] Duffie JA, Beckman WA. Solar engineering of thermal processes. second ed. John Wiley & Sons, Inc.; 1991. [13] Diab MR. Experimental and analytical study of the heat transfer characteristics of a solar air heater incorporated a finned absorber. Ph.D. thesis, Purdue University, 1981. [14] Heaton HS, Reynolds NC, Kays WM. Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int J Heat Mass Transfer 1964;7:763–81. [15] Pesaran AA, Mills AF. Moisture transport in silica gel packed beds: I. Theoretical study. Int J Heat Mass Transfer 1987;30:1037–49. [16] Pesaran AA, Mills AF. Moisture transport in silica gel packed beds: II. Experimental study. Int J Heat Mass Transfer 1987;30:1051–60. [17] ASHRAE. 2005 ASHRAE handbook – fundamentals. Atlanta, GA: American Society of Heating Refrigerating and Air-Conditioning Engineers; 2005. [18] Press WH, Teukolsky SA, Vetterling WT, Flamery BP. Numerical recipes in FORTRAN – the art of scientific computing. 2nd ed. Cambridge University Press; 1992. [19] Helen-Xing H-Y. Desiccant dehumidification analysis. M.Sc. thesis, Massachusetts Institute of Technology, 2000. [20] Degelman LO. Enercalc: a weather and building energy simulation model using fast hour-by-hour algorithms. In: Proceedings of the fourth national conference on microcomputer applications in energy, April 25–27, Tuscon, Arizona, 1990. [21] Degelman LO. A statistically-based hourly weather data generator for driving energy simulation and equipment design software for buildings. In: Proceedings of the building simulation ‘91, international building performance simulation association (IBPSA), August 20–22, Nice, SophiaAntipolis, France, 1991.