Performance analysis of a liquid desiccant evaporative cooling airconditioning system powered by solar energy

Transcription

1 Performance analysis of a liquid desiccant evaporative cooling airconditioning system powered by solar energy F. Zhang 1, Y. Yin 1,2* and X. Zhang 1,2 1 School of Energy and Environment Southeast University, 2 Sipailou Road, Nanjing 2096, China 2 MOE Key Laboratory of Energy Thermal Conversion and Control, Southeast University, 2 Sipailou Road, Nanjing 2096, China * Corresponding author: y.yin@seu.edu.cn SUMMARY Solar cooling is one of the promising solutions to the worsening energy and climate issues. A solar-powered liquid desiccant evaporative cooling air-conditioning system with solution recirculation was proposed. Considering the coupled characteristics of the solar collecting subsystem and the liquid desiccant cycle, a detailed mathematical model of the whole system was established. Based on this model, hourly ambient parameters from three days ( 22 nd, 13 th and 28 th ) in the typical meteorological year of Nanjing were chosen to investigate the thermal performance of the proposed system under hot and humid climate conditions. The results show that the supply air produced by this system can be directly used to handle the cooling load of the airconditioned space in most time. The average thermal coefficients of performance (COPs) in the three selected typical days are 0.59, 0.67 and 0.64, respectively. The average solar COPs are 0., 0.33 and 0.31, respectively. INTRODUCTION Heating, ventilation and air conditioning systems play a significant role in ensuring human thermal comfort and are among the largest energy consumers of the building sector (Vakiloroaya et al. 14) which consumed -40% of total energy use in developed countries (Pérez et al. 07). Increased emphasis is put on the design of energy-efficient air-conditioning systems due to the continuing increase in energy demand, costs and the associated environmental problems (Ghaddar et al. 03). Thermal-driven cooling and air-conditioning system can be an alternative to the conventional vapor-compression system driven by electricity. Among all kinds of the thermal-driven cooling systems, the solar-powered cooling system shows prominent advantages. It is environmentally friendly due to no use of refrigerant with high ODP (Ozone Depression Potential) and GWP (Global Warming Potential). Besides, the solar energy supply and cooling demand match very well. Therefore, solar cooling/air conditioning is one of the most promising solutions to the deteriorating energy and climate issues. Solar-powered cooling systems mainly include absorption cooling system, solid desiccant cooling system and liquid desiccant cooling system. Extensive researches on solarpowered absorption cooling system have been conducted in the past few years (Xu et al. 11, Gomri 13, Fong and Lee 14, Zhang et al. 15, Bellos et al. 16, Shirazi et al. 16). Zhang et al. (15) proposed a heat pump solar absorption refrigeration system with a multi-stage heat storage system. The grade of the solar energy is elevated by the heat pump system to drive the generator, because the required generation temperature for the absorption cycle is relatively high. Bellos et al. (16) tested four different types of solar collectors in a single effect absorption chiller and concluded that the evacuated tube collector is the most beneficial choice. For seeking the better operation mode of the solar-powered absorption chiller, three control strategies were proposed and compared in the solar loop of the plant (Shirazi et al. 16). Compared with the closed absorption chiller, the desiccant cooling system can be driven by lower-temperature heat source, and therefore has the higher potential to make full use of solar energy (Ameel et al. 1995, Grossman 02). Ge et al. (12) proposed a novel solar driven desiccant coated heat exchanger cooling system and conducted a performance analysis of the proposed system under Shanghai summer condition. The simulation results validated its feasibility when applied under hot and humid climate condition. Since the liquid desiccant can be regenerated at a lower temperature in comparison with the solid desiccant, many researchers focus on the investigation of the solarpowered liquid desiccant cooling system (Gommed and Grossman 12, Qi et al. 12, Qi and Lu 14). Qi and Lu (14) simulated and optimized the operation performance of liquid desiccant air-conditioning system driven by solar energy in Hong Kong. However, the hourly performance of the system is beyond the scope of their study. In this paper, a solar-powered liquid desiccant evaporative cooling air-conditioning (LDECAC) system with solution recirculation was proposed. Considering the coupled characteristics of the solar collecting subsystem and the liquid desiccant cycle, the detailed mathematical model of the whole system was established. Hourly ambient parameters in three days ( 22 nd, 13 th and 28 th ) from the typical meteorological year of Nanjing were chosen to investigate thermal performance of the whole system. The supply air state, system cooling capacity, thermal coefficient of performance of the liquid desiccant subsystem, and solar coefficient of performance of the whole system were used to give a comprehensive evaluation. METHODOLOGY System description Schematic diagram of the proposed solar-powered LDECAC system is shown in Fig. 1. This hybrid system consists of four subsystems: 1) liquid desiccant subsystem which handles the total latent load; 2) evaporative cooling subsystem which handles the sensible heat load; 3) solar collecting subsystem which provides the hot water for the liquid desiccant subsystem; and 4) cooling water subsystem (omitted in Fig. 1 for simplicity). ISBN: COBEE18-Paper6 page 904

2 The liquid desiccant used in this system is LiCl aqueous solution. It possesses a relatively low water vapor partial pressure even at ambient temperature, so it can extract water vapor from the humid air in an efficient way when cooled by natural cooling source. Besides, its water vapor partial pressure would be higher than that of ambient air when heated to above 55 ºC, thus facilitating the use of flatplate solar collector for its regeneration. The selected mass concentration of the liquid desiccant is dependent both on the temperature of the cooling source and heating source. Normally, the adopted mass concentration of LiCl aqueous solution in this system is about 40% under hot and humid climate conditions. W3 W2 W4 Heat source Water tank 9 W1 S5 1 2 S1 A7 A1 A1 A2 A3 A4 Reg Deh EA 3 OA OA EA A5 S4 S8 A/W HE 8 S3 S7 DECI DECII 6 S6 S/S HE S2 7 Air 4 Liquid desiccant Water 5 OA-Outdoor air; SA-Supply air; EA-Exhaust air; Deh-Dehumidifier; Reg-Regenerator; S/S HE-Solution/Solution heat exchanger; A/W HE-Air/Water heat exchanger; DECI-Direct evaporative cooler I; DECII-Direct evaporative cooler II; 1-Solution heater; 2-Solution cooler; 3-Fan; 4-Valve; 5-Solution pump; 6-Concentrated solution tank; 7-Diluted solution tank; 8-Water pump 9-Solar collector; -Auxiliary heater Figure 1. Schematic diagram of solar-powered liquid desiccant evaporative cooling air-conditioning system. The liquid desiccant loop is as follows: The coldconcentrated solution (state S1) enters the dehumidifier from the top and comes into direct contact with the process air in a cross-flow direction, becoming warm and diluted after absorbing the moisture from the process air. The warmdiluted solution leaving the dehumidifier at state S2 is then divided into two parts. Only a relatively small fraction is preheated to state S3 in the solution-solution heat exchanger by an equally small fraction of the hot-concentrated solution (state S6) leaving the regenerator. At the same time, the remaining part is mixed with the pre-cooled concentrated solution at state S7 and becomes mixed solution at state S8. This mixed solution is further cooled to state S1 again, thus completing a self-cycle at the dehumidification side (S1 S2 S8 S1). At the regeneration side, the diluted solution at state S3 is mixed with a relatively large fraction of concentrated solution at state S6 and then becomes mixed solution at state S4. This mixed solution is further heated to a specified set point temperature in the solution heater by the hot water produced in the solar collecting subsystem and enters the regenerator at state S5. This hot-diluted solution desorbs its moisture to the regeneration air, becomes reconcentrated and reaches state S6, thus completing a selfcycle at the regeneration side (S6 S4 S5 S6). The hot water loop is as follows: The relatively hot water (state W1) enters the solar collector and absorbs the thermal energy converted from solar radiation, leaving the solar collector with higher temperature (state W2). When the solar radiation is unavailable or low, the hot water will be further heated to state W3 through the auxiliary heater. Hot water at state W3 enters the solution heater and releases its heat to the solution, and then returns to the water tank at state W4, thus completing the hot water loop. A6 SA The air handling process is as follows: Outdoor air at state A1 passes through the dehumidifier, where its humidity ratio is reduced. The dry air leaving the dehumidifier at state A2 is sensibly cooled to state A3 in the air-water heat exchanger and then is divided into two streams. One stream enters the DECI to produce cooling water used in the air-water heat exchanger and is rejected to the environment at state A4 after humidified and heated. The other passes through the DECII to be further cooled to state A5. To obtain a supply air with desired temperature and humidity ratio, a fraction of dry air at state A3, without entering the DECII, is mixed with the air leaving the DECII at state A5. This mixed air at state A6 is then delivered into the air-conditioned room. Besides, another outdoor air stream (state A1) passes through the regenerator to re-concentrate the diluted solution and then it is rejected to the environment at state A7. It should be pointed out that the temperature of the exhaust air at state A7 is still high and had better be recovered when the regeneration heat is inadequate. This can be solved by setting an air-air heat exchanger to preheat the regenerator inlet air with the air leaving the regenerator. Mathematical model In this section, mathematical models of major components in solar-powered LDECAC system, namely, dehumidifier, regenerator, direct evaporative coolers, heat exchangers, solar collector and water tank, are established in order to simulate the whole system numerically. Adiabatic packed-bed dehumidifier and regenerator with cross-flow configuration were used in this study. The NTU- Le f model established in (Yin and Zhang 08) was adopted to describe the coupled heat and mass transfer processes between air and liquid desiccant. The numerical method used to solve the NTU-Le f model to obtain outlet air and solution parameters of dehumidifier/regenerator has been described in detail by Liu et al. (07). Besides, the thermophysical properties of LiCl solution can be obtained from Conde (04). The mixing processes at the dehumidification and regeneration sides follow the energy conservation equation and the mass conservation equations of solute and solution. The solution recirculation ratio at the dehumidification/regeneration side (R s,deh /R s,reg ) can be expressed by Eqs. (1) and (2), respectively. R m m S8 S7 s,deh (1) ms2 R m m S4 S3 s,reg (2) ms6 where m s is the mass flow rate of solution, subscripts S2-S8 refer to the corresponding liquid desiccant states, as shown in Fig. 1. As there are similarities of coupled heat and mass transfer characteristics between air-water and air-liquid desiccant direct-contact devices (Liu et al. 09), the NTU-Le f model can also be used to describe the direct evaporative coolers. For simplicity, energy efficiency model was selected to describe the heat transfer process in all kinds of heat exchangers. Considering the practical situation, the efficiency of the solution-solution heat exchanger, air-water heat exchanger, solution-hot water heat exchanger and airair heat exchanger at the regeneration side was set as 0.75, 0.65, 0.75 and 0.55, respectively. ISBN: COBEE18-Paper6 page 905

3 Vacuum tube solar collector is adopted in this proposed system. Efficiency equation of this kind of collector is given by Ge et al. (12): w,sc,in amb T.. T sc (3) I where η sc is the efficiency of solar collector, I s means solar radiation density (W/m 2 ), T w,sc,in is the temperature of water at the inlet of solar collector, and T amb is ambient temperature. The temperature of hot water at the outlet of solar collector T w,sc,out can be calculated by Eq. (4): T s AI sc s sc w,sc,out Tw,sc,in (4) mw,sc cp,w where A sc is the area of the solar collector, c p,w is the specific heat capacity of water, and m w,sc is the mass flow rate of water in the solar collector. For the water tank, Eq. (5) is adopted to describe its dynamic performance: dtw,wt wvwt cp,w mw,sc cp,w( Tw,wt,in Tw,wt,out ) (5) dt where ρ w is water density, V wt is the volume of water tank, T w,wt,in and T w,wt,out are the temperature of water at the inlet and outlet of the water tank, respectively. It should be mentioned that the water tank is considered to be a fully mixed model in this paper. Performance index Three performance indexes, including the system cooling capacity (Q 0 ), the thermal coefficient of performance (TCOP) of the liquid desiccant subsystem, and the solar coefficient of performance (COP s ) of the whole system, were used to give a comprehensive evaluation of the thermal performance of this solar-powered LDECAC system. They can be calculated by the following Eqs. (6) to (8): Q m (1 R )( h h ) (6) 0 a,deh a amb sup TCOP Q Q 0 (7) s reg Q0 COPs I A where m a,deh is the mass flow rate of air entering the dehumidifier, h amb and h sup refer to the specific enthalpy of the ambient air and the supply air, respectively. R a refers to the working to intake air flow ratio, defined as the mass flow rate ratio of the air entering DECI to the air entering air-water heat exchanger. Q reg refers to the thermal energy consumed in the regeneration process and can be calculated by Eq. (9): sc reg s,reg,in S5 S4 (8) Q m ( h h ) (9) where m s,reg,in refers to the mass flow rate of regenerator inlet solution, subscripts S4 and S5 refer to the solution entering the solution heater and that entering the regenerator, respectively, as shown in Fig. 1. Modal Validation and simulation flowchart Mathematical model of the cross-flow dehumidifier and regenerator has been validated by Zhang et al. (17). Considering that the routine models of other components are mature, the validations of them are omitted. Consequently, models of the various components are reliable when used in the performance analysis of the whole system. The flow chart for calculating the thermal performance of the liquid desiccant and evaporative cooling subsystems is given in (Zhang et al. 17). The regeneration temperature in (Zhang et al. 17) is specified in advance. However, in this paper, the regeneration temperature is determined by the coupled characteristics of the liquid desiccant subsystem and solar collecting subsystem. Therefore, another iterative calculation is needed to obtain the regeneration temperature. RESULTS AND DISCUSSION In this section, the thermal performance of the proposed system under Nanjing summer condition was investigated using MATLAB software. Hourly ambient parameters in three days ( 22 nd, 13 th and 28 th ) from the typical meteorological year of Nanjing were chosen for the performance analysis. Fig. 2 shows the hourly ambient conditions from 4:00 to 17:00. It can be seen from Fig. 2(a) and Fig. 2(b) that both the ambient air dry-bulb temperature (T amb ) and humidity ratio (ω amb ) increase from to, then decrease slightly from to. The highest values of T amb in these three selected days are 27.4 ºC, 35.1 ºC and 34.0 ºC, respectively, and these of ω amb are 17.0 g/kg, 23.1 g/kg and 21.0 g/kg, respectively. As shown in Fig. 2(c), the values of solar radiation density (I s ) are similar (the highest value is about 800 W/m 2 ) in and, which are higher than I s in (the highest value is about 670 W/m 2 ). As the proposed system is suitable for the hot and humid climate conditions, the typical days in and are selected to investigate its performance. Although the T amb in is lower than the temperature of the indoor air design conditions (26 ºC, 55%), the ω amb in is relatively high, and therefore the outdoor air should be dehumidified before supplied to the air-conditioned room. Considering that the proposed system may perform badly in due to the relatively low solar radiation, a typical day from is also chosen to investigate the applicability of this system. Geometrical size of the several key components in the liquid desiccant and evaporative cooling subsystems as well as the corresponding operating parameters can be founded in (Zhang et al. 17). The solution recirculation ratio (R s,deh /R s,reg ) was set at 0.5. For solar collecting subsystem, the area of solar collector was set at 0 m 2 and the volume of water tank was 7.5 m 3. The mass flow rate of water in both solar collector and solution heater was set at 1.33 kg/s. It should be pointed out that the auxiliary heater was not activated during the simulation for the purpose of exploring the solar driving performance of the system, which was also adopted by Ge et al. (12) and Xu and Wang (17). During the simulation, the system operation time was set from 9:00 to 17:00 for 8 hours. The initial temperature of the water storage tank (T w,wt ) was not set at the ambient temperature, but close to T w,wt at the end of the operation time. This is equivalent to assume that the situation in the investigated day is similar to that of the day before and the water storage tank is in good insulation. Even if the above assumptions are not satisfied, the solar collecting subsystem can be operated in advance to preheat the water to the set temperature. The initial T w,wt for 22 nd, 13 th and 28 th are 35 ºC, 50 ºC and 48 ºC, respectively. ISBN: COBEE18-Paper6 page 906

4 th 4 International Conference On Building Energy, Environment (a) 36 Figure 3. Calculated hourly supply air state from to on psychrometric chart 34 Tamb (oc) 32 As shown in Fig. 3, almost all of the supply air states fall into the area below the isenthalpic line, indicating that the supply air can be directly supplied to the air-conditioned room to handle the cooling load. Only a few points are very close to the isenthalpic line, which occur at time 9:00 and 17:00. This is because at these time, the regeneration temperature (Ts,reg,in) is relatively low so that the system cannot operate in an effective way. As presented in Fig. 4, the values of Ts,reg,in at 9:00 in the three months are only 36.8 ºC, 55.9 ºC and 52.5 ºC, respectively. Besides, it can also be seen from Fig. 3 that although the enthalpy of ambient air (hamb) in is significantly lower than that in and, the enthalpy of supply air (hsup) is similar in these three months. As the solar radiation in and is obviously higher than that in (as shown in Fig. 2), the values of Ts,reg,in in and are significantly higher than that in (as shown in Fig. 4). Due to the combined effects of the hamb and Ts,reg,in, the values of hsup in these three months are very similar :00 06:00 08:00 :00 12:00 14:00 16:00 18:00 :00 (b) 24 ωamb (g/kg) :00 06:00 08:00 :00 12:00 14:00 16:00 18:00 :00 65 Ts,reg,in (oc) 60 (c) 800 Is (W/m2) :00 :00 11:00 12:00 13:00 14:00 15:00 16:00 17: Figure 4. Variations of regeneration temperature (Ts,reg,in) from 9:00 to 17:00 04:00 06:00 08:00 :00 12:00 14:00 16:00 18:00 :00 Figure 2. Hourly parameters of (a) ambient air dry-bulb temperature (Tamb), (b) ambient air humidity ratio (ωamb), and (c) solar radiation density (Is) Supply air state Fig. 3 shows the calculated supply air temperature (TSA) and humidity ratio (ωsa) on the psychrometric chart. The indoor air design conditions are set as 26 ºC and 55% RH, and the corresponding isenthalpic line is also plotted in Fig. 3. It should be pointed out that the supply air state mentioned here is state A3 shown in Fig. 1. When the by-pass flow ratio of the DEC at the last stage is changed from 0 to 1, the possible states that the supply air can reach is the approximate isenthalpic line from state A3 to the saturated point. Only state A3 is plotted here for the convenience of reading figures. 15 alpic Dry Bulb Temperature (oc) % line Q0 (kw) Humidity Ratio (g/kg) ISBN: % 0% % 40% Isent h Indoor Air Fig. 5 shows the hourly cooling capacity (Q0) of the proposed system. It can be seen that the values of Q0 in both and are the highest at 13:00, when the Ts,reg,in reaches its highest value. For, the highest value of Q0 occurs at 11:00 and 15:00, when Ts,reg,in is the highest. In fact, both Q0 and Ts,reg,in have very similar trend during different months. In addition, the highest values of Q0 in these three months are 43.5 kw, 64.6 kw and 66.1 kw, respectively. With the increase of Tamb and ωamb, the Q0 increases and therefore, the proposed system can make full use of its own cooling potential when applied in hot and humid climate conditions. OA in SA in OA in SA in OA in SA in Overall system performance 09:00 :00 11:00 12:00 13:00 14:00 15:00 16:00 17: Figure 5. Variations of system cooling capacity (Q0) from 9:00 to 17:00 COBEE18-Paper6 page 907

5 CONCLUSIONS 4 th International Conference On Building Energy, Environment Figure 6. Variations of thermal coefficient of performance (TCOP) and solar coefficient of performance (COP s ) from 9:00 to 17:00 Fig. 6 shows the variation trends of the thermal coefficient of performance (TCOP) and the solar coefficient of performance (COP s ). In general, the TCOP gradually increases from 9:00 to 17:00; while for COP s, it decreases slightly from 9:00 to 13:00 and then shows a relatively sharp increase. The reasons for above variation trends of TCOP and COP s are as follows. As mentioned above, Q 0 and T s,reg,in shows very similar trend. That is also true for Q 0 and the thermal energy consumed in the regeneration process (Q reg ). The values of Q 0 depend on both Q reg and ambient conditions. Before 13:00 (in and ), both Q 0 and Q reg increases. As the rise of h amb would also increase Q 0, Q 0 increases more than Q reg, leading to the rise of TCOP. After 13:00, the Q reg decreases more than Q 0, which also causes the rise of TCOP. As for COP s, its value depends on both Q 0 and I s when the area of solar collector is constant. As the value of I s changes faster than that of Q 0, I s increases more than Q 0 before 13:00, and decreases more significantly after 13:00. Therefore, the COP s decreases first and then increases obviously. η sc :00 :00 11:00 12:00 13:00 14:00 15:00 16:00 17:00 Figure 7. Variations of solar collector efficiency (η sc ) from 9:00 to 17:00 It can also be seen from Fig. 6 that the values of TCOP in and are always higher than that in. The average values of TCOP for the three selected typical days ( 22 nd, 13 th and 28 th ) are 0.59, 0.67 and 0.64, respectively. In general, the values of COP s in and are also higher than that in, but the difference among them is slight. The average values of COP s for the three days are 0., 0.33 and 0.31, respectively. Besides, as shown in Fig. 7, the values of solar collector efficiency (η sc ) in the three months are also similar, and the average value is about 0.5, indicating the solar collector subsystem can perform well when coupled with liquid desiccant dehumidification subsystem in hot and humid region like Nanjing. A solar-powered liquid desiccant evaporative cooling air conditioning system with solution recirculation was proposed. The detailed mathematical model of the whole system was established and utilized to investigate its thermal performance under Nanjing summer condition. Hourly ambient parameters in three days ( 22 nd, 13 th and 28 th ) from the typical meteorological year of Nanjing were chosen for the performance analysis. The supply air state, system cooling capacity (Q 0 ), thermal coefficient of performance (TCOP) of the liquid desiccant subsystem, and solar coefficient of performance (COP s ) of the whole system were used to give a comprehensive evaluation. The following conclusions can be drawn: (1) In most time, the supply air produced by the proposed system can be directly used to handle the cooling load of the air-conditioned space. Although the enthalpy of ambient air in is significantly lower than that in and, the enthalpy of supply air is similar in these three months due to their different solar radiation. (2) System cooling capacity and regeneration temperature show very similar trend during different months. The highest values of Q 0 in these three months are 43.5 kw, 64.6 kw and 66.1 kw, respectively. (3) In general, the TCOP gradually increases from 9:00 to 17:00; while for COP s, it decreases slightly from 9:00 to 13:00 and then shows a relatively sharp increase. The average values of TCOP for the three selected typical days are 0.59, 0.67 and 0.64, respectively. The average values of COP s for the three days are 0., 0.33 and 0.31, respectively. REFERENCES Ameel, T. A., Gee, K. G., & Wood, B. D. (1995). Performance predictions of alternative, low cost absorbents for open-cycle absorption solar cooling. Solar Energy, 54(2), Bellos, E., Tzivanidis, C., & Antonopoulos, K. A. (16). Exergetic, energetic and financial evaluation of a solar driven absorption cooling system with various collector types. Applied Thermal Engineering, 2, Conde, M. R. (04). Properties of aqueous solutions of lithium and calcium chlorides: formulations for use in air conditioning equipment design. International Journal of Thermal Sciences, 43(4), Fong, K. F., & Lee, C. K. (14). Performance advancement of solar air-conditioning through integrated system design for building. Energy, 73(9), Ge, T. S., Dai, Y. J., Li, Y., & Wang, R. Z. (12). Simulation investigation on solar powered desiccant coated heat exchanger cooling system. Applied Energy, 93(5), Ghaddar, N., Ghali, K., & Najm, A. (03). Use of desiccant dehumidification to improve energy utilization in airconditioning systems in Beirut. International Journal of Energy Research, 27(15), Gommed, K., & Grossman, G. (12). Investigation of an improved solar-powered open absorption system for cooling, dehumidification and air conditioning. International Journal of Refrigeration, 35(3), ISBN: COBEE18-Paper6 page 908

6 Gomri, R. (13). Simulation study on the performance of solar/natural gas absorption cooling chillers. Energy Conversion & Management, 65(1), Grossman, G. (02). Solar-powered systems for cooling, dehumidification and air-conditioning. Solar Energy, 72(1), Liu, X. H., Jiang, Y., & Qu, K. Y. (07). Heat and mass transfer model of cross flow liquid desiccant air dehumidifier/regenerator. Energy Conversion & Management, 48(2), Liu, X. H., Li, Z., & Jiang, Y. (09). Similarity of coupled heat and mass transfer between air-water and air-liquid desiccant direct-contact systems. Building & Environment, 44(12), Pérez-Lombard, L., Ortiz, J., & Pout, C. (08). A review on buildings energy consumption information. Energy & Buildings, 40(3), Qi, R., Lu, L., & Yang, H. (12). Investigation on airconditioning load profile and energy consumption of desiccant cooling system for commercial buildings in Hong Kong. Energy & Buildings, 49(2), Qi, R., & Lu, L. (14). Energy consumption and optimization of internally cooled/heated liquid desiccant airconditioning system: a case study in Hong Kong. Energy, 73(7), Shirazi, A., Pintaldi, S., White, S. D., Morrison, G. L., Rosengarten, G., & Taylor, R. A. (16). Solar-assisted absorption air-conditioning systems in buildings: control strategies and operational modes. Applied Thermal Engineering, 92, Vakiloroaya, V., Samali, B., Fakhar, A., & Pishghadam, K. (14). A review of different strategies for hvac energy saving. Energy Conversion & Management, 77(1), Xu, S. M., Huang, X. D., & Du, R. (11). An investigation of the solar powered absorption refrigeration system with advanced energy storage technology. Solar Energy, 85(9), Xu, Z. Y., & Wang, R. Z. (17). Simulation of solar cooling system based on variable effect LiBr-water absorption chiller. Renewable Energy, 113, Yin, Y., & Zhang, X. (08). A new method for determining coupled heat and mass transfer coefficients between air and liquid desiccant. International Journal of Heat & Mass Transfer, 51(13), Zhang, F., Yin, Y., & Zhang, X. (17). Performance analysis of a novel liquid desiccant evaporative cooling fresh air conditioning system with solution recirculation. Building & Environment, 117, Zhang, X., Li, H., & Yang, C. (15). A novel solar absorption refrigeration system using the multi-stage heat storage method. Energy & Buildings, 2, ISBN: COBEE18-Paper6 page 909