Simulation of solar-powered absorption cooling system

Simulation of solar-powered absorption cooling system

Renewable Energy 28 (2003) 1277–1293 www.elsevier.com/locate/renene Simulation of solar-powered absorption cooling system Ibrahim Atmaca ∗, Abdulvaha...

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Renewable Energy 28 (2003) 1277–1293 www.elsevier.com/locate/renene

Simulation of solar-powered absorption cooling system Ibrahim Atmaca ∗, Abdulvahap Yigit Department of Mechanical Engineering, Faculty of Engineering and Architecture, Uludag University, TR–16059, Bursa, Turkey Received 14 March 2002; accepted 30 October 2002

Abstract With developing technology and the rapid increase in world population, the demand for energy is ever increasing. Conventional energy will not be enough to meet the continuously increasing need for energy in the future. In this case, renewable energy sources will become important. Solar energy is a very important energy source because of its advantages. Instead of a compressor system, which uses electricity, an absorption cooling system, using renewable energy and kinds of waste heat energy, may be used for cooling. In this study, a solar-powered, single stage, absorption cooling system, using a water–lithium bromide solution, is simulated. A modular computer program has been developed for the absorption system to simulate various cycle configurations and solar energy parameters for Antalya, Turkey. So, the effects of hot water inlet temperatures on the coefficient of performance (COP) and the surface area of the absorption cooling components are studied. In addition, reference temperatures which are the minimum allowable hot water inlet temperatures are determined and their effect on the fraction of the total load met by non–purchased energy (FNP) and the coefficient of performance are researched. Also, the effects of the collector type and storage tank mass are investigated in detail.  2002 Elsevier Science Ltd. All rights reserved. Keywords: Solar energy; Absorption; Cooling; Simulation



Corresponding author. E-mail address: [email protected] (I. Atmaca).

0960-1481/03/$ - see front matter  2002 Elsevier Science Ltd. All rights reserved. doi:10.1016/S0960-1481(02)00252-5

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Nomenclature A Ac C COP D FNP FR h hm I k m m ˙ Q Qaux QL Qload Qu r R T Tin T0 Tref Ts u U X z hc ⌬t a l

constant collector area (m2) concentration of absorbate in solution (mol/m3) coefficient of performance mass diffusivity (m2/s) the fraction of the total load met by non-purchased energy (%) flow ratio (XA / (XG⫺XA)) heat transfer coefficient (W/m2K) mass transfer coefficient (m/s) solar insolation (W/m2) heat conductivity (W/mK) storage tank mass (kg) mass flow rate (kg/s) heat transfer rate (W) auxiliary heater capacity (kW) extracted energy from the storage tank (kW) generator load (kW) useful energy (kW) cylindrical coordinate (m) radius of the tube (m) temperature (K, °C) collector inlet temperature (°C) the environment temperature (°C) reference temperature (°C) storage tank temperature (°C) flow velocity (m/s) overall heat transfer coefficient (W/m2K) weight fraction of LiBr coordinate along the tube (m) collector efficiency time period (h) heat diffusivity (m2/s) heat of absorption (kJ/m3)

Subscript A b C ch,in cool,in

absorber bulk condenser chilled water inlet cooled water inlet

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E G h,in i

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evaporator generator hot water inlet interface

1. Introduction The possible use of solar energy as the main heat input for a cooling system has led to several studies of available cooling technologies that use solar energy. Various types of solar-powered systems are available for cooling applications. One widespread application of a solar-powered system is for absorption cooling which is an alternative approach to cooling that is largely thermally driven and requires little external work. Recently, solar energy has received interest as an alternative energy source for cooling systems, especially in places where electricity is expensive or in short supply. With the use of solar energy, usage of conventional energy sources and its peak demand will be reduced. Several computer models for describing the performance characteristics of absorption chillers were developed by Egrican et al. [1], Wordono et al. [2] and Mostofavi et al. [3]. In addition, the use of solar energy in an absorption heat pump system has been investigated by Ileri [4, 5] and Ergul [6]. Also, Ghaddar et al. [7] have carried out research into solar absorption system performance in Beirut. In order to improve the present understanding of the system dimensions and productive usage of solar energy, much more investigation still has to be done. In this study, a solarpowered, single stage, absorption cooling system, using a water–lithium bromide solution is simulated. The main focus of this study is concentrated on the simulation of the absorption cooling system with a detailed solar energy process. The purpose of this paper is to investigate the effect of hot water, which is supplied from the solar energy, inlet temperatures on the coefficient of performance (COP) and the effect of the heat transfer surface area of the absorption cooling components based on a 10.5 kW constant cooling load. In addition, using a constant load of 10.5 kW cooling capacity, reference temperatures, which are the minimum allowable hot water inlet temperatures, are determined and their effect on the fraction of the total load met by nonpurchased energy (FNP) and the coefficient of performance are investigated. Also, the effect of the collector type and storage tank mass are researched in detail. Consequently, the surface areas of the absorption cooling system components, which can be used to optimise the commercial solar-powered absorption cooling systems, are presented and the possibility of using solar energy in this system is shown. 2. Absorption cooling system modeling A schematic diagram of a single-stage absorption cooling cycle, using lithium bromide and water as the working fluid, is shown Fig. 1. The lithium bromide–water

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Fig. 1.

Schematic diagram of the absorption cooling system.

solution works as an absorbent while water is a refrigerant fluid. As is known, the system includes heat exchangers, a pump, valves and piping. In addition, the solarpowered system includes a collector, a storage tank and an auxiliary heater. The generator of the absorption cooling system uses solar energy, which is collected in solar system collectors, for driving vapor from the liquid solution. Usage of solar energy in the system is shown in Fig. 2 in detail. The vapor flows into the condenser while the heated liquid passes back to absorber through an absorbent-throttling valve, thus completing the absorber loop. The throttling valve maintains the pressure difference between the absorber and the generator. Conveying hot absorbent from the generator into the absorber wastes a considerable amount of thermal energy. A liquidto-liquid heat exchanger, which transfers energy from this stream to the weak concentration solution pumped back to the generator, saves a major portion of the energy. In the absorber, the steam is absorbed by the high-concentration solution. The solution is weakened by the absorption of water vapor coming from the evaporator. The vapor is condensed in the condenser and then flows to the evaporator. An expan-

Fig. 2.

Usage of solar energy in the system in detail.

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sion valve, between the condenser and the evaporator, is used to decrease the pressure of the refrigerant (water vapor), which also decreases its temperatures. The evaporation of the water in the evaporator produces the desired cooling effect. Simplified mathematical models are used for the system components. Each of these components is considered to be a heat exchanger. First, the overall heat transfer coefficient is predicted and according to this prediction, approximate dimensions are determined by using the LMTD method for each component. By means of these dimensions, the real overall heat transfer coefficient is calculated for each component. Calculations have been continued until the real overall heat transfer coefficient has converged with the predicted value and as convergence is supplied, the dimensions of the system are determined. Coefficient of performance (COP) is defined according to Fig. 1 as follows: COP ⫽

QE QG

(1)

where QE is the refrigeration capacity and QG is the rate of heat transferred to the generator from an energy source. For two-phase flows, which include evaporation and condensation in horizontal tubes, the inside heat transfer coefficient is calculated as suggested by Kakac [8] and the Heat Exchanger Design Handbook [9]. For each component, a shell-side heat transfer coefficient equation is obtained from Kern et al. [10]. For other components except the absorber, equations inside the heat transfer coefficient calculation are taken from Incropera et al. [11]. The condenser is investigated in two zones as shown in Fig. 3. The refrigerant state entering the condenser is superheated. A short distance after entering the condenser the refrigerant is cooled to the saturation point. Condensation then occurs over most of the heat exchanger length as the refrigerant goes from 100% to 0% quality.

Fig. 3.

Typical temperature profiles for a condenser.

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In the absorber of the cooling system, heat and mass transfer take place together. A falling-film, vertical-tube absorber is modelled. A liquid solution film flows down around the vertical tube and the absorption of the vapor takes place at the liquid– vapor interface. Cooling water flows counter-current in the tube. The heat of absorption is released at the interface and transferred through the liquid film. The absorption heat is rewaved by the cooling water. In this study, the steady state operation of the absorber system is considered. It is assumed that the flow of the liquid film is laminar and fully developed. No shear forces are exerted on the liquid by vapor at the liquid–vapor interface. All properties of the lithium bromide–water solution and water are calculated from equations as suggested Egrican et al. [1]. If the liquid is Newtonion and has constant physical properties, the energy and diffusion equations are, respectively;

冉 冉

冊 冊

u

∂T ∂2T ∂T ⫽a 2⫹ ∂z ∂r r∂r

u

∂C ∂2C ∂C ⫹ ⫽D ∂z ∂r2 r∂r

(2) (3)

It can be assumed that vapor pressure equilibrium exists at the vapor–solution interface, and a linear relationship between the interfacial temperature and concentration can be taken as follows: Ci ⫽ A1Ti ⫹ A2 ; l ⫽ constant

(4)

where l is the heat of absorption. Interfacial temperature and concentration are related to each other by the following equations: k

冉冊 冉 冊

∂T ∂C ⫽D l ∂r ∂r

(5)

The Crank–Nicolson finite difference method is used to solve these partial differential equations with appropriate boundary conditions. These differential equations were expressed in finite difference form in cylindrical coordinates. Solutions for temperature and concentration distributions were obtained by solving discretized equations in the flow field. These solutions were used in determining heat transfer coefficients from the bulk of the fluids to the wall and the mass transfer coefficient from the interface to the bulk of the liquid. Heat and mass transfer coefficients can be calculated by using the previously obtained temperature and concentration distributions. ⌬C ⫺D |r=interface ⌬r hm ⫽ (Ci⫺Cb) and

(6)

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⌬T | ⌬r r=surface (Tb⫺Ti)

⫺k h⫽

(7)

where Tb and Cb are the bulk temperature and concentration, respectively [12]. The heat transfer coefficient is changed from 1500 W/m2K to 2000 W/m2K as between 80 and 95°C hot water inlet temperatures.

3. Solar energy modeling Usage of solar energy in the system is shown in Fig. 2 as a schematic diagram. The useful energy collected is transferred to the hot liquid storage tank from which the generator of the absorption system is supplied with input thermal energy. This investigation covers the environmental and atmospheric circumstances of Antalya, Turkey. Necessary equations for calculating the solar insolation, I, are taken from Duffie and Beckman [13]. Solar insolation, I, for Antalya is presented in Table 1. In this table, the angle of incidence of the collector is taken as 35° which is the optimum incidence angle for Turkey from May to September which includes the summer and spring seasons. Solar collectors are modelled with the formulation as suggested by Duffie and Beckman [13]: Qu ⫽ Ac·I·hc

(8)

where Qu is the useful energy collected in system collectors, Ac is the collector area and the collector efficiency, hc, is;



hc ⫽ FR· (ta)⫺UL·



Tin⫺T0 I

(9)

Table 1 Solar insolation, I (W/m2), for Antalya, Turkey Hour

May

June

July

August

September

7:00 8:00 9:00 10:00 11:00 12:00 13:00 14:00 15:00 16:00 17:00

198 371 526 636 701 728 701 636 526 371 198

218 396 551 663 729 756 729 663 551 396 218

216 401 566 684 753 780 753 684 566 401 216

195 393 574 707 784 814 784 707 574 393 195

143 341 533 675 757 787 757 675 533 341 143

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In this equation, Tin is the collector inlet temperature and T0 is the environment temperature. Three types of collectors, as suggested by Ileri [5], are considered in the simulation. Types of collectors with their characteristics are presented in Table 2. A water storage tank is placed after the solar collectors as shown in Fig. 2. Perfect mixing within the tank is assumed. If the rate of heat addition and removal for a reasonable time period of ⌬t are assumed to be constant, equations can be written for each time interval as suggested by Duffie and Beckman [13]: Ts,new ⫽ Ts,old ⫹

⌬t ·[Q ⫺QL⫺(U·A)s·(Ts⫺T0)] (m·cvw)s u

(10)

In the above equations; QL is the extracted energy from the storage tank, Ts is the main storage temperature for the period and m is the storage tank mass. (U.A)s is taken as 11.1 W/K as suggested by Duffie and Beckman [13]. Here, the heat transfer coefficient is assumed to be 0.72 W/m2K for natural convection. In spite of the fact that (U.A)s depends on the surface area of the storage tank and hence the volume since the surface area of the storage tank shows little change for selecting storage tank mass, an average storage tank surface area and an average (U.A)s are assumed. An auxiliary heater at the exit of the storage tank boosts the temperature of the hot water from the storage tank temperature to the allowable reference temperature when the storage tank temperature drops below the allowable reference temperature. For this reason, the reference temperature can be considered to be exactly the minimum allowable hot water inlet temperature. The auxiliary heater capacity is calculated as follows: Qaux ⫽ m ˙ ·cpw·(Tref⫺Ts)

(11)

where m ˙ is the mass flow rate used by the generator. The fraction of the total load met by non-purchased energy FNP is calculated as; Qaux FNP ⫽ 1⫺ Qload

(12)

where Qload is exactly the same as QG which is the generator load. Simulation includes inputs and outputs. The inputs of the system include the evaporator temperature (Te), refrigeration capacity (QE), cooling water inlet temperature (Tcool,in) and flow rates. In the same time, storage tank temperature, which is connected to storage tank mass, collector type and area, is calculated from eq. (10) for hourly input from 10:00 to 18:00 for about 8 hours a day. By these inputs, the chilled Table 2 Solar collector characteristics Collector type

Collector description

FR(ta)

FRUL

A B C

Evacuated, selective surface Double glazed Single glazed

0.70 0.75 0.90

3.3 6.5 10.0

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water inlet temperature (Tch,in), condenser (absorber) temperature (Tc) and generator temperature (Tg) are calculated. Chilled water inlet temperature is taken as 10 K higher than evaporator temperature. Condenser (absorber) temperature is assumed as 15 K higher than cooling water inlet temperature. Generator temperature is 10 K less than hot water inlet temperature (Th,in). The system has been designed for 10.5 kW constant cooling load. Cooling water, chilled water and hot water flow rates are taken as 0.3, 0.4 and 0.4 kg/s, respectively. The outputs include the surface area of the system components, coefficient of performance (COP) and the fraction of the total met by non-purchased energy (FNP).

4. Results and discussion Simulation results are discussed in this section under design considerations which are presented above. Hot water inlet temperature effects on COP, flow ratio FR and surface area of the system components are shown in Figs 4 and 5, respectively. It can be seen that an increase in this temperature increases the mass fraction of concentration solution (XG), so the flow ratio (FR) decreases. COP increases now that the decreasing flow ratio causes the generator load to show a decreasing trend. As shown in Fig. 5, the hot water inlet temperature is found to affect some of the surface area of the system components. An increase in this temperature decreases the absorber and solution heat exchanger surface area while the dimensions of the other components remain approximately unchanged. This is due to the effect of the flow ratio. By increasing the hot water inlet temperature, thermal energy extracted from the absorber, depending on the flow ratio, decreases. Further, an increase in this temperature increases the absorber temperature and in this way, the average logarithmic temperature difference (⌬Tml) boosts in absorber and solution heat exchanger. By decreasing

Fig. 4. The effect of the hot water inlet temperatures on the system COP and FR (Te=280 K, QE=10.5 kW, Tc=306 K).

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Fig. 5. The effect of the hot water inlet temperatures on the surface area of the system components (Te=280 K, QE=10.5 kW, Tcool,in=291 K).

the heat capacity and increasing ⌬Tml, heat transfer surface area normally decreases in these components. The investigation was carried out between 353 and 368K (approximately 80–95°C) because increasing the system COP slows down considerably after a certain temperature. In addition, boiling occurs above 95°C and this is not desired. To investigate the effect of storage tank reference temperature on the FNP and COP, three reference temperatures of the storage tank, 80, 85 and 90°C, were taken into consideration. In this research, it is assumed that the collector area is 50 m2 and the storage tank mass is 75 kg per square meter of collector area. In addition, collector A is preferred as a collector type which has high efficiency. The results are given in Fig. 6 for FNP. As seen from Fig. 6, the 80°C reference temperature seems to be the best. It gives better results than that of the higher reference temperatures. In this study, a reference temperature below 80°C wasn’t investigated, because it could cause a decrease in

Fig. 6. The effect of the reference temperatures on FNP (collector A, 3750 kg storage tank mass and 50 m2 collector area).

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Fig. 7.

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The effect of reference temperatures on the storage tank temperatures for a day in June.

the system COP. However, below the 80°C reference temperature, sufficient refrigerant vapor isn’t driven from the liquid solution. Storage tank temperature changes according to reference temperature for a day in June are shown in Fig. 7. It can be seen that the storage tank temperature increases for high reference temperatures since generator load decreases and further, thermal energy extracted from the storage tank decreases at high reference temperatures because of greater usage of the auxiliary heater (Fig. 8). But, for the reference temperature of 90°C, in spite of the high storage tank temperature, it doesn’t reach 90°C. For this reason, the amount of auxiliary heater capacity increases as shown in Fig. 8, and so FNP decreases. In addition, COP is investigated for a day in August as

Fig. 8.

The effect of reference temperatures on the auxiliary heater capacity for a day in June.

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Fig. 9.

The effect of the reference temperatures on the system COP for a day in August.

shown in Fig. 9. As shown, unlike the effects caused by reference temperature on FNP, increasing the reference temperatures boosts the system COP. Also, COP remains unchanged at this high reference temperature (90°C). To investigate the storage tank mass effects on the FNP, it is assumed that the collector area is 50 m2, the reference temperature is 85°C, and the storage tank mass is taken as 75, 100 and 125 kg per square meter of collector area. Collector A is chosen as a collector type. The effect of the storage tank mass on the FNP is shown in Fig. 10. As seen from Fig. 10, an increase in the storage tank mass causes a decreasing trend in FNP. The reason for this can be seen by the inspection of the

Fig. 10. The effect of the storage tank mass on FNP (85°C reference temperature, 50 m2 collector area and collector A).

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daily variations of tank temperatures. Storage tank temperature changes according to storage tank mass for a day in June are shown in Fig. 11. There is a reverse order between the storage tank mass and the change of the storage tank temperature. This means that increasing the storage tank mass results in less change in the storage tank temperature. As a result of this, the smallest tank (3750 kg) reaches maximum and minimum temperatures in comparison with the others during the day time. It can be seen that the highest storage tank temperature at the end of the working hour is supplied at the highest storage tank mass. In this study, a storage tank mass below 3750 kg isn’t investigated, because it could cause an excessive decrease in the storage tank temperature at the end of the working hours. As seen from Fig. 11, since storage tank temperature exceeds the reference temperature (85°C for this research) for certain hours at 3750 kg storage tank mass, the auxiliary heater load decreases and so, FNP increases. To investigate the effects of the collector type on the FNP, three types of collectors which are presented in Table 2 are investigated. In this investigation, the 80°C reference temperature and 50 m2 collector area are selected. It is assumed that storage tank mass is 75 kg per square meter of collector area. Results are given for FNP in Fig. 12. It can be seen from Fig. 12 that collector type is very important for solarpowered absorption cooling systems. Changes in the collector efficiency for three types of collectors for a day in August are presented in Fig. 13. As shown, when collector A which has high efficiency is used, it can be seen from Figs 14 and 15, respectively, that useful energy collected increases and as a result of this increase storage tank temperature reaches a higher value than the values of collectors B and C. Therefore, FNP increases. It can be seen from Fig. 15 that when collector A is used, storage tank temperatures exceed the reference temperature (80°C for this investigation) except the value at 17:00. Only at this hour is the auxiliary heater needed.

Fig. 11. The effect of storage tank mass on the storage tank temperatures for a day in June.

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Fig. 12. The effect of the collector type on FNP (80°C reference temperatures, 50 m2 collector area and 3750 kg storage tank mass).

Fig. 13. The effect of the collector type on the collector efficiency for a day in August.

Obtained results have been compared with similar studies, like those by Ileri [5] and Ghaddar et al. [7], by taking into consideration COP, the effect of the reference temperature and storage tank mass on the FNP. It can be seen from Table 3that, as compared with Ileri’s [5] study, our results for reference temperature effects show the same general behaviour. Similarly, Ileri’s study shows that an 80°C reference temperature is the best solution. The storage tank temperature as a function of time for the masses of 1000, 1300, 1500 and 1800 kg in August is determined by Ghaddar et al. [7]. Our results show similar trends in comparison with Ghaddar’s study that has the same temperature

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Fig. 14. The variations of the useful heat gain as a function of time for a day in August for different collector type.

Fig. 15. The variations of the storage tank temperatures as a function of time for a day in August for different collector type.

Table 3 Simulation result comparisons for reference temperature effects on the FNP for August Tref (°C)

Ileri’s study [8] FNP (%)

Present study FNP (%)

80 85 90 95

94 – 58 56

100 89 69 –

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interval. In Ghaddar’s [7] study, yearly solar fraction (YSF) was determined. YSF decreased from 0.22 to 0 for between 50 and 2500 kg storage tank mass and 50 m2 collector area. Our results for storage tank mass effects on FNP in August are shown in Fig. 10. Since our study and Ghaddar’s study were performed at different geographical regions for different storage tank masses, different results are found. Nevertheless, this study and Ghaddar’s [7] study show the same behaviour. In addition, in Ghaddar’s [7] study, COP changes approximately between 0.7 and 0.78, similar to our results. In this case, it can be said that there is a good agreement between the two studies. 5. Conclusion From the above study the following results can be drawn. 1. The hot water inlet temperature is found to affect some of the surface area of the system components. Increasing this temperature decreases the absorber and solution heat exchanger surface area, while the dimensions of the other components remain approximately unchanged. 2. Although high reference temperature increases the system COP and decreases the surface area of system components, lower reference temperature gives better results for FNP than high reference temperatures do. For this study, an 80°C reference temperature is the best choice. 3. An increase in the storage tank mass causes the FNP to show a decreasing trend. It can be said that the storage tank mass should be kept at a minimum on the condition that it shouldn’t cause the storage tank temperature to decrease excessively at the end of the working hours. In this study a 3750 kg storage tank mass seems to be the best choice. 4. Supplied results show that a solar-powered absorption cooling system requires high performance collectors. Despite their high cost, the high efficiency collectors must be selected for effective operation of the solar-powered absorption cooling system. In this study, collector A is the best choice. References [1] Egrican N, Yigit A. Simulation of an absorption cooling system. Energy 1992;17(6):513–600. [2] Wardono B, Nelson RM. Simulation of a 20-ton LiBr/H2O absorption cooling system. ASHRAE Trans 1995:96–103. [3] Mostafavi M, Agnew B. The effects of ambient temperature on the surface area of components of an air-cooled lithium bromide/water absorption unit. Appl Thermal Engng 1996;16(4):313–9. [4] Ileri A. A discussion on performance parameters for solar-aided absorption cooling systems. Renew Energy 1997;10(4):617–24. [5] Ileri A. Yearly simulation of a solar-aided R22-Degdme absorption heat pump system. Solar Energy 1995;55(4):255–65. [6] Ergul E. Simulation of a solar-aided R22-Degdme absorption heat pump system. M.S. Thesis, Middle East Technical University, 1991.

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[7] Ghaddar NK, Shihab M, Bdeir F. Modeling and simulation of solar absorption system performance in Beriut. Renew Energy 1997;10(4):539–58. [8] Kakac S. Boilers, evaporators & condensers. Florida: John Wiley and Sons, 1991. [9] Heat exchangers design handbook. Hemisphere Publishing Corporation, 1983. [10] Kern DQ, Kraus AD. Extended surface heat transfer. New York: McGraw-Hill Company, 1972. [11] Incropera FP, DeWitt DP. Fundamentals of heat and mass transfer, 3rd ed. Singapore: John Wiley & Sons, 1990. [12] Bird RB, Stewart WE, Lightfoot EN. Transport phenomena. John Wiley & Sons, Inc, 1960. [13] Duffie JA, Beckman WA. Solar engineering of thermal process. University of Wisconsin, Madison: A Wiley, Interscience Publications, 1980.