Performance of solar-assisted hybrid air-conditioning liquid desiccant system in Beirut N. Ghaddarl, K. Ghali2 & A. ~ajm' I American University of Beirut, Faculty of Engineering and Architecture, Beirut 11072020, Lebanon 2 Beirut Arab University, Faculty of Engineering, Beirut, Lebanon Abstract In this work a model of a solar-operated liquid desiccant system (using calcium Chloride) for air dehumidification is developed. The desiccant system model is integrated with a solar heat source for performance evaluation at a wide range of recorded ambient conditions for Beirut city. The desiccant system of the current study is part of a hybrid desiccant-vapor compression system for a high latent load application, namely a small restaurant with an estimated cooling load of 11.39 TR (40 kw), including reheat. The relevant parameters of the desiccant system are optimized at peak load, and it is found out that there is an important energy saving if the ratio of the air flow rate in the regenerator to that in the dehumidifier is about 0.3 to 0.4. The COP of the desiccant unit is 0.45 for the restaurant. The size of the vapor compression unit of the restaurant is reduced to eight TR when supplemented by a desiccant system. The transient simulation of the solar desiccant system is performed for the entire cooling season. The solar fraction for the restaurant is 0.19, 0.38, and 0.54, for a collector area of 28.72, 57.44, and 86.16 m', respectively. The economic benefit of the desiccant system is positive. For a gas price of 0.5638 $/kg, the payback period for the restaurant turned out to be immediate if the energy is supplied solely by natural gas, and 11 years if an 86.16 m2 solar collector is implemented to reduce the fuel consumption.
242 Energv and the Environment 1 Introduction The residential and commercial sectors in Lebanon consumed 30xl0~~i~ajoules in 1994, representing 30% of the total energy consumption in the country [l]. Eighty percent of the total electricity consumption in the residential and commercial sector is distributed as follows: electric heaters for space heating 31%, electric domestic hot water systems 22%, air conditioning A/C 13%, lighting 8.5%, refrigerator 6% [l]. The continuing rise in energy demand, costs and the associated environmental problems, notably climate change, is causing increased emphasis on the design of energy efficient air-conditioning systems for both industrial and comfort applications. In humid climates such as the Lebanese coast, the humidity issues are a major contributor to energy inefficiency in HVAC devices. The high humidity of the outside air combined with ventilation requirement increases the latent load. Most conventional air-conditioning systems are not designed to independently control temperature and humidity. The use of desiccant pre-conditioning of supply air can improve the humidity control capabilities of overall refrigeration and air conditioning systems and reduce energy costs. Furthermore, incorporation of desiccant preconditioning into such systems allows higher percentage of fresh air in the supply stream [2]. With recent advances in desiccant dehumidification, liquid sorbent equipment is becoming even more attractive for air-conditioning applications [3-71. Another attractive feature of desiccant dehumidification systems is their suitability for solar or other low-grade thermal energy applications 171. Adebiyi and Russell [8] conducted a comparative second law analysis for various assisted air-conditioning systems including vapor compression, absorption, and desiccant systems. Jurinak et a1 [9] presented a performance comparison between desiccant and vapor compression systems for residential applications and the cost and environmental impact of using solar energy for regeneration. Ghaddar et al. [l01 performed a study on solar-operated Lithium-Bromide absorption cycle performance under climatic conditions of Beirut. A desiccant-based airconditioning system was proposed and evaluated by Kinsara et al. [ll]. The proposed configuration utilized a vapor compression refrigeration system for sensible cooling and a liquid desiccant system for dehumidification. In this work, the transient performance of a calcium-chloride desiccant system is studied. Ambient summer conditions of Beirut are considered. The desiccant system use is studied for a high latent load application where the desiccant system takes charge of the major part of the latent load and a portion of the sensible load for restaurant that has a cooling load requirement of 11.36 T.R. (= 40 KW). On a daily basis, the cooling system for the restaurant operates for 18 hours (from 6a.m. to 12 p.m.). Solar energy supplies part of the regeneration heat. The auxiliary source of energy is natural gas. An economic feasibility analysis will be performed of the solar-desiccant system as compared to the traditional vapor compression system.
Energy and the Environmmt 243 2 Problem formulation The liquid desiccant in our work is Calcium chloride solution. Figure 1 shows a schematic of a solar liquid desiccant air dehumidification system that consists of two loops: the air dehumidification loop and the liquid desiccant regeneration loop. The regeneration heat is supplied from a solar collector system and a backup auxiliary heater. The air flows from the bottom of the packed bed in a counter direction to the liquid desiccant [12]. The packing material is used to increase the transfer area between the air and the liquid desiccant. The solarassisted desiccant system design and performance will be analyzed for a high latent load application to evaluate the system feasibility in Beirut. The case is for a small restaurant of area of 64 m'. The cooling load calculations are performed at the outdoor climatic conditions of Beirut as Tdb=34"C and TWb=26"C, and the recommended indoor standard design conditions at Tdb=25"C and 55% relative humidity [12]. The design parameters of the restaurant hybrid-desiccant cooling system are determined based on steady state analysis to meet their peak cooling load requirements. Then the feasibility of solar-assisted desiccant system is studied through a transient analysis of the system performance over the operational period from May till October. 2.1 Peak cooling load requirement of the restaurant The case of high latent load is for a 64 m2 restaurant at ground floor of a building. It has a length and a width of 8 m, and a height of 3 m. The top and west neighboring sides are air-conditioned zones. The bottom side is a non airconditioned basement. The kitchen is an un-conditioned zone. The restaurant heat conductance values (U) are: Roof (U=0.511 w/m2.k), Wall (U=2.42 w/m2.k), Partition (U=2.76 w/m2.~), Floor (U= 1.53 w/m2.k), and Windows: (U=5.95 w/m2 K). The Effective Sensible Heat Factor (ESHF) is 0.639 for the restaurant, requiring reheat of the supply air when traditional vapor compression is used alone to meet the cooling load. For economic analysis and performance comparison between hybrid system and vapor compression system, the peak load vapor compression cooling unit size is calculated according to ASHRAE standards [12]. The restaurant fresh air requirement equals to 0.645 m3/sec and is equal to the supply air of the system. The peak load occurs at 3 p.m. in July and August and is equal to 33.77 kw consisting of a sensible load of 15.25 kw and a latent load of 18.52 kw. Adding 10% to the load to account for leakage, duct losses and fan heat gain, a 37.15 kw total cooling load is obtained. Since ESHF is low, the equipment capable of handling the load would have a low apparatus dew point temperature. Then reheat is necessary and estimated at 2.02 kw. The vapor compression unit size for the restaurant is 40 KW. 2.2 Desiccant system design parameters To determine the design parameters of the desiccant cooling system, the steady state thermodynamic cycle model is simulated to meet the peak load
244 Energv and the Environment Figure 1: The liquid desiccant system and packed beds. independent of the solar heat source. The model is developed performing standard mass and energy balances over each component of the desiccant cooling system. The process of air dehumidification or liquid desiccant regeneration in the packed beds of the dehumidifier and the regenerator is simulated using mathematical model of Radhwan et a1 [13,14]. The one-dimensional flow model uses a fourth order Runge-Kutta scheme to integrate the moisture and energy balances of air and liquid desiccant in the packed beds during the air dehumidification mode and the solution generation mode. The numerical model calculates the air and liquid desiccant temperatures, the air humidity ratio, and the liquid desiccant moisture content in the packed beds at various bed heights as well as inlet conditions of the liquid desiccant and air and solution flow rates. The Radhwan et al. model [l41 is integrated for the full desiccant cycle for design parameters evaluation. The main design parameters of the packed beds are: (a) The flow rate ratio "r" of the mass flow rate of CaC12 liquid ''G? to the mass flow rate of air "G,' in the dehumidifier; (b) the ratio "F of outdoor air to total air (outdoor + return) in the dehumidifier; and (c) the ratio "q" of total air flow rate in the regenerator to that in the dehumidifier. Inlet conditions of both air and liquid desiccant to the dehumidifier are governed by ambient dry bulb and wet bulb temperatures. The temperature of air at inlet to the dehumidifier Taid and its specific humidity w,id depend on whether the air stream consists of outdoor air, return air from the conditioned space or a combination of both. The moisture content of the desiccant solution at inlet to the dehumidifier Ylid is set to 1.5 to avoid the problem of crystallization of CaC12 [13,14]. The length of the packed dehumidifierlgenerator bed is 1.5 m and the diameter is 1 m. The desiccant system may need the use of two packed beds in the dehumidifier andlor regenerator to meet the cooling load requirement. The rate of regeneration heat needed to attain a regeneration temperature of TXg is: PR = 4 (cp/ + ~ ~ k, d- k. c ) ~ ~ (1)
Energy and the Environmmt 245 where c,] and c, are the constant-pressure specific heats of the pure liquid desiccant and water, respectively, Ylod is the moisture content at outlet of dehumidifier, Tlohe is the liquid temperature at outlet of heat exchanger and m, is the liquid flow rate of CaC12. The coefficient of performance of the system and its operating cost per hour is defined as follows: COP =- Qc (2) Q R + P,,, Cost = 3600Q~C, + H, 1000 where P,,, is the total electric power consumption of the desiccant system in Watts, QC is the total saving in cooling load by the desiccant system in Watts, QR is the rate of the regeneration heat added in Watts, H, the heating value of 1 kg of Butane (C4 Hlo, Hg = 49.593 X 106 Jkg) [4], C, is the cost of 1 kg of Butane (=$0.5638), C, is the cost of 1 kw-hr of electricity in Beirut (= $0.1327). 2.3 Integrated solar-assisted desiccant system model Once the design parameters of the desiccant cooling system are set based on the steady state peak load analysis, then the desiccant system transient performance is studied. A feasibility study of the solar-assisted desiccant system is performed at three different solar collector areas and costs for the months when cooling is needed. The solar collector system is composed of several units of widely spread non-selective single glazing type that collects and stores solar energy in a water storage tank. Each unit has an effective surface area of 2.872 m2. The absorber plate is a black painted steel plate with an emissivity of 0.91 and absorptivity of 0.91 and it has 12 welded longitudinal copper tubes of 15 mm diameter. The collectors are used in forced circulation mode and are set at @-15O tilt angle facing south where 0 is the latitude and is equal 33.9O for Beirut. The transient performance of the collector-tank system is simulated numerically using the theory of Hottel and Whillier presented by Duffie and Beckman [15]. The useful heat gain of water in the collector is obtained by applying an energy balance that indicates the distribution of incident solar energy into useful heat gain, thermal losses and optical losses with a time step between 0.0833 up to 0.15 hour depending on system flow rate. The details of the model can be found in References [l51 and [16]. Two thermostats measure the storage tank temperature and the liquid temperature at inlet to the solar heater. If the difference between the temperatures is greater than 3"C, the liquid is allowed into the solar heater, otherwise it bypasses it and the liquid goes directly into the auxiliary heater as shown in Fig.1. The coefficient of performance of the desiccant system based on the daily and monthly purchased auxiliary energy COP, is calculated similar to equation (2), while replacing QR by Q,. The yearly solar fraction is then defined as the ratio of the cooling provided by solardriven desiccant system to the total cooling load.
246 Energv and the Environment The temporal calculations of the solar collector-tank system proceeds using first order Euler-Forward integration scheme at an initial temperature of the storage tank and inlet water temperature to the collectors equal to the ambient temperature for the given day. The hourly direct and diffuse solar radiation incident on the collectors and the values of ambient temperatures, wind speed and direction are derived directly from actual hourly measured weather data files of Beirut. In the transient analysis, the ambient conditions are not constant but are equal to the hourly-measured values by the A.U.B. weather and solar station for the typical day of the month. The liquid temperature at outlet of the heater is assumed constant over the whole period of operation, and is equal to the regeneration temperature obtained in the steady-state analysis. To close the calculations iteration loop, the moisture content of the desiccant exiting the regenerator should match the moisture content of the desiccant at inlet of the dehumidifier. The system is assumed to be quasi-steady-state; i.e. the variables, while varying from hour to hour, are considered constant during every hour of analysis. A sensitivity analysis is performed on the system with regard to the time step, where time steps of 15, 30, and 60 minutes are used. The results showed less than 1% difference in the calculated values of heat rates and COP. The collector-tank simulation has been previously verified with experimental data on solar water heaters without load extraction [16,17]. 3 Results and discussion The steady state results are presented to show the bases upon which the desiccant system design parameters are selected for the restaurant at the outdoor climatic conditions of Beirut. Transient simulations' results of the solar-assisted desiccant system will follow with emphasis on economical analysis. 3.2 Restaurant hybrid desiccant cooling system For the restaurant, F is fixed at the value of unity because of the required high ventilation. The parameter r is kept constant at 0.8, and the flow ratio parameter q is varied from 0.3 to 0.7, with two beds in the dehumidifier and one bed in the regenerator. Figure 2(a) shows the variation of (a) COP with q and Fig. 2(b) shows the variation of the regeneration temperature and the liquid temperature at inlet to the heater with q. The COP peaks at a value of q equal to 0.3, which is taken as the design value for the desiccant system in this work. 3.2 Transient analysis and solar heating system from May to October The design parameters of the desiccant system are set to r=0.8, Treg= 46.4OC and q=0.3 while using two beds in the dehumidifier and one bed in the regenerator. The transient response of the selected system is studied for a typical day of every month. The hourly-measured dry bulb and dew point temperatures for a typical day must be close to the monthly-averaged values. The chosen typical days are: May 16'~, June 15", July 16", August 16'~, September 16", and October 16". The weather data of Beirut corresponds to the year 1998.
Energy and the Environmmt 247 Figure 2: The variation of (a) COP vs. the ratio q of total air flow rate in the regenerator to that in the dehumidifier, and (b) the regenerator temperature and the liquid desiccant temperature at heater inlet versus 4- Figure 3: COP for a representative Figure 4: Q,,, and Q, for a day in July day of each month. (30 units). For the solar system, three areas of solar collectors are considered using 10, 20 and 30 units where the unit area is taken as 2.872 m2. Figure 3 shows the hourly COP of the desiccant system for the 6 representative days, from May to October for the restaurant. The hourly quantity of auxiliary heat required, and the hourly energy supplied by the solar storage tank in the month of July are then shown in figure 4 for collector area 86.16 m'. With 20 units of solar collectors, the storage tank supplied totally the regeneration heat for 6 hours in July, while with 30 units, the period jumped to 10 hours. Figures Sdepict the variation of the monthly and yearly solar fraction with the area of collectors. 3.3 Economic analysis It is important to perform an economic feasibility analysis of the solar-desiccant system as compared to the traditional vapor compression system. The solardesiccant system is cost effective when its additional cost is less than the present value of the energy savings.
248 Energv and the Environment Figure 5: A plot of MSF for (a) May, June, July, and YSF, and (h) August, September, and October. The cost of the vapor compression system for the restaurant, having a size of 11.37 T.R. (40 kw) chiller, is taken equal to $10,000 (The cost is slightly increased to take into account the reheat coil). When used as a hybrid system, the desiccant dehumidifier reduces the size of the vapor compression system of the restaurant to 8.14 TR. The cooling system size is taken equal to 8 T.R., having a cost of 8 X $833.33 = $6667 for the restaurant. The temperature is allowed to swing slightly at peak load. The desiccant system consists of two dehumidifier beds, one regenerator bed, a heat exchanger, an auxiliary heater, and a cooling tower. The estimated cost of a 3 TR system is $3000 [IS]. The cost of the solar system is composed of the cost of the solar collectors (= $200/m2) and that of the insulated galvanized steel storage tank (=$500/m3). Installation cost is assumed to be included in all these items. The operating energy cost of the combined desiccant-solar system, over the whole season, is given by: where E,,, is the total electric energy consumption of the season. The operating cost of the traditional vapor compression system is given by: Cost, Qcce (5) = 3. 6 ~ 1 0 ~ ~ 0 ~ ~ ~ where COP, is the COP of the vapor compression system (= 2.5). The total cooling load of the restaurant is 422.35 GJJseason. Therefore the seasonal operating cost of the vapor compression system of the restaurant is equal to 7683.0 $, including a reheat cost of $1455.7 (10969.9 kw-hr). The performance of the desiccant system is studied at three gas costs (0.1410, 0.2819, 0.4229, and 0.5638 US $/kg). If the desiccant system is used to reduce the restaurant cooling load, there is a saving in the operating cost of the vapor compression system equal to $3819. llyear.
Enrrgy and the Environmmt 249 Figure 6: A plot of LCS versus Figure 7: Effect of the energy collector area for the inflation rate on the restaurant. LCS. Figure 8: (a) Effect of market discount rate on LCS and (b): Effect of the life cycle on LCS. In this study, an initial extra expenditure "C' is invested at a market discount rate d = 0.08, in order to make a saving "S" in the yearly energy bill that is inflating at a rate e equal to 0.05 over a period N = 20 years. The system is assumed to have no resale value. The life cycle savings is then given by [15]: where LCS =-C + PWF(N,e,d)S (64 PWF(N,e,d) = if efd N if e=d (6b) l+e The LCS of the restaurant is shown in Fig. 6 as a function of the solar collector system area. The desiccant system alternative is cost effective for the restaurant as a hybrid system. Figure 7 shows the effect of the variation of the energy inflation rate on the LCS of the restaurant system, at C, equal to 0.5638 $/kg and a market discount rate of 8%. Figures 8(a) and 8(b) present the effect of variation of the market discount rate and the life cycle on the LCS, at a gas cost
250 Energv and the Environment of 0.5638 $/kg for the restaurant case. The return of investment R01 is immediate, since savings exist also in the capital investment as well as the operating cost for the no solar alternative. The (ROI) becomes 21.6 %, 18.0 %, and 15.6 %, when the solar collector area varies from 28.72 m2 to 57.44 m', and 86.16 m', respectively. The payback period is zero years for the no solar alternative, whereas it assumes the values of about 6.14, 9.0, and 11.0 years for 10, 20, and 30 units of solar collectors, respectively. The liquid desiccant cooling, when used as a part of a hybrid system, is very attractive because it eliminates the need for reheat in the vapor compression system. 4 Conclusions A model of a solar-operated liquid desiccant system (using calcium Chloride) for air dehumidification is developed. The use of liquid desiccant dehumidification systems of supply air is a viable alternative to reduce the latent heat load on the HVAC system, improve efficiency and reduce operating cost. The regeneration temperature of liquid desiccants is in the range of 45-50 C. Waste heat is readily available in that range giving liquid desiccant systems great potential for energy conservation. Acknowledgment The authors would like to acknowledge the financial support of the Lebanese National Council for Scientific Research under grant No. 323041. References "First National Communication", MOE, Lebanon - UNDP & IPCC, pp. 1999. Meckler, M. "Desiccant Outdoor Air Pre-conditioners Maximize Heat Recovery Ventilation Potentials". ASHRAE Transaction Symposia: SD-95-9- 4, 1995. Collier R.K., Jr., D. Novosol, and W.M. Worek. "Simulation of Open-cycle Desiccant Cooling System performance." ASHRAE Transactions, Vol. 96, Par 1, pp.1262-1268, 1990. Albers W.F., J.R. Beckman, R.W. Farmer, K.G. Gee. "Ambient Pressure, Liquid Desiccant Air Conditioner" ASHRAE Transactions, Vol. 99, Par 1, pp. 603-608, 1991. Griffiths, W.C. 'Desiccant Dehumidification Reduces Refrigeration Loads" Energy Engineering, Vol. 86, No. 4, pp. 39-49, 1989. Peng C.S. P. and J.R. Howell. "The Performance of Various Types of Regenerators for Liquid Desiccants" ASME J Solar Engineering, Vol. 106, pp.133-144, 1984. Buzweiler, U. "Air conditioning with a Combination of Radiant Cooling, Displacement, ventilation, and Desiccant Cooling." ASHRAE Transactions, No. 4, 1999.
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