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Thermodynamic investigation of LiCl-H2O working pair in a double effect absorption chiller driven by parabolic trough collectors
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Evangelos Bellosa, Christos Tzivanidisa, Sasa Pavlovicb, Velimir Stefanovicb
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a
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Thermal Department, School of Mechanical Engineering, National Technical University of Athens, Zografou, Heroon Polytechniou 9, 15780 Athens, Greece. b Department of Energetics and Process Technique, University of Niš, Serbia
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Tel: +30 210 772 2340 , Fax: +30 210 772 1260
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Abstract In this study, an alternative working pair, the LiCl-H2O is investigated in a doubleeffect absorption chiller which is powered by parabolic trough collectors. Moreover, the conventional working fluid LiBr-H2O is examined under the same operating conditions in order to perform a suitable comparison. Three different condensation temperature levels are examined (30oC-35oC-40oC) and four evaporating temperature levels (5.0oC-7.5oC-10.0oC-12.5oC), while the generator temperature varies in the allowed range in every case. The analysis is performed in steady state conditions with EES (Engineering Equator Solver). The final results indicate that the solar cooling performance is 8% higher with the LiCl-H2O compared to the operation with LiBrH2O. This significant performance enhancement is able to make the solar cooling a more sustainable solution using LiCl-H2O. Furthermore, the demanded specific collecting area is calculated close to 1 m2∙kW-1 which is a relatively low value. This result is based on the high thermal efficiency of the parabolic trough collectors and on the increased performance of the operation with LiCl-H2O.
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Keywords LiCl-H2O, Double-effect absorption chiller, Exergetic analysis, PTC, EES
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1. Introduction The worldwide energy consumption has an increasing rate due to the development of many countries as China and the new lifestyle trends which demand greater energy consumption in order to fulfill the higher thermal comfort standards [1-3]. The building sector is responsible for approximately the 35% of the total energy consumption; a high percentage which is greater in developed countries [4-5]. Thus, the utilization of alternative and renewable energy sources seems to be the only sustainable way for facing the existing situation [6]. Solar cooling is a promising way for utilizing the available solar irradiation during the summer. This technology is ideal for areas with high solar potential because there is high compatibility between source supply and load demand [7]. Additionally, solar cooling aids the decrease of high peaks in electricity consumption during the summer which creates severe problems in the grid energy distribution [8].
Corresponding author: Evangelos Bellos (
[email protected])
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Solar cooling technologies are based on sorption machines, with absorption chillers to be the most mature and widespread technology. The most usual working pairs in the absorption chillers are lithium bromide-water (LiBr-H2O) for temperature levels greater than 4oC and water-ammonia (H2O-NH3) for producing cooling in extremely low-temperature levels [9]. Moreover, the last years a lot of research has been focused on the examination of alternative working pairs [10-12] as the lithium chloride-water (LiCl-H2O), lithium nitrate-ammonia (LiNO3-NH3), ammonia-calcium chloride (NH3CaCl2), ammonia/sodium thiocyanate (NH3-NaSCN) and many others. Among them, the lithium chloride-water (LiCl-H2O) seems to be an efficient alternative working fluid [11-14], except the conventional lithium bromide-water (LiBr-H2O). Furthermore, LiCl-H2O presents greater long-term stability [15] and higher internal energy storage capacity [11] than LiBr-H2O; two important factors for selecting LiClH2O as a promising working pair in absorption chillers.
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The last years, a lot of research has been focused on the evaluation of absorption chillers, as well as in other sorption machines, operating with LiCl-H2O. Borge et al. [16] examined a solar absorption cooling system operating with LiCl-H2O which rejects heat to an outdoor swimming pool. Fong et al. [11] examined two solar assisted cooling systems for Hong Kong subtropical climate. More specifically, an absorption and an adsorption chiller operating with LiCl were examined and finally, the absorption technology was proved to be the best solution leading to 12.2% yearly savings compared to a conventional system while the adsorption system led to 7.1% energy savings. Gunhan et al. [12] examined a novel solar absorption chiller with the LiCl-H2O working pair which does not face crystallization problems. This system is examined exergetically and finally, the maximum exergetic destruction was found on the solar collectors. She et al. [17] suggested a double effect absorption chiller which uses both LiCl-H2O and LiBr-H2O and it is able to operate with low heat source temperature levels. The final results proved that this system has a higher coefficient of performance compared to the traditional double effect absorption chiller operating with LiBr-H2O in the respective conditions. In the same direction, many comparisons between LiCl-H2O and LiBr-H2O have been performed. Gogoi and Konwar [18] compared these working pairs in a single effect absorption chiller and according to their results, LiCl-H2O performs better than LiCl-H2O with a difference close to 8%. Moreover, Parham et al. [15] proved that LiCl-H2 O demands lower temperature levels in generator than the LiCl-H2O in order to the same COP to be achieved. This result means that the chiller exergetic performance is higher for operation with the LiClH2O working pair.
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The most usual absorption chiller is the single effect absorption chiller which performs with a coefficient of performance (COP) close to 0.7 [19-20]; a low value compared with the compression systems where the COP is close to 3 [21]. The performance of the absorption chillers can be improved by using a multi-stage machine (double-effect or triple-effect). However, these machines need higher heat source temperature levels in order to operate efficiently [22]. Following this demand, 2
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the use of concentrating collectors is the most reliable and efficient solution in order to design solar cooling systems. Parabolic trough collectors (PTC) are the most mature solar concentrating technology and there are numerous applications which utilize PTC for producing heat at medium and high temperatures [23]. Moreover, evacuated tube collectors with smaller concentrator are another reliable solution for medium temperature levels. The advantage of PTC is the higher thermal performance and the possibility of operating at high temperature levels. On the other hand, ETC is able to operate without tracking system and they can utilize the solar diffuse irradiation partially [24]. It is obvious that these two solar collectors have different advantages and they are suitable solutions for solar cooling systems which demand high-temperature levels. It is important to add that the single-effect systems can operate also with conventional flat plate collectors because of the lower heat source temperature level demand.
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The last years, a lot of research has been focused on solar cooling systems with multieffect absorption chiller because of their high performance, compared to the single effect systems. As it is stated below, these systems cannot usually operate with flat plate collectors and so concentrating collectors as parabolic trough collectors, linear Fresnel reflectors and evacuated tube collector with small concentrators can be used. Li et al. [25] examined a double effect absorption chiller operating with LiBr-H2O coupled with ETC and they concluded that the heat source temperature should be between 110 to 130oC. Bermejo et al. [26] examined a similar system powered by linear Fresnel reflectors and they finally proved that the daily chiller COP was up to 1.25. In a recent study for the Greek climate, Drosou et al. [27] examined the use of PTC with a double absorption chiller for space cooling of office buildings. They proved that with a collecting area of 1716m2 the solar energy can cover the 50% of the cooling load which is equal to 1147 kW. Shirazi et al. [22] performed an energetic and financial analysis of the various solar cooling systems and they examined in the system performance many parameters as the collecting area, the storage tank volume and the solar potential. They proved that ETC is the best solution for single-effect absorption chillers and PTC has to be coupled with double or triple machines.
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As it is obvious from the literature, the combination of PTC and double effect absorption chillers is a promising solution in solar cooling systems. However, only LiBr-H2O and H2O-NH3 have been analyzed in these systems and there is not any study testing the use of LiCl-H2O in a double-effect solar cooling system powered by concentrating collectors. This study aims to compare the LiCl-H2O and LiBr-H2O working pairs in a double–effect absorption chiller powered by commercial PTCs (Eurotrough ET-150). These working pairs are tested under the same operating conditions and the system performance is examined in energetic and exergetic terms. The optimum heat source temperature which maximizes the exergetic performance and simultaneously minimizes the demanded collecting area is determined in any case. The optimum operating conditions for the examined working pairs are compared under various evaporating and condensing temperature levels. The final results of this 3
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study can be used for designing double effect absorption chillers with LiCl-H2O and for selecting the optimum range of the heat source temperature levels in every case. Furthermore, it is essential to state that the emphasis is given in the thermodynamic comparison of these working pairs and thus the analysis is performed under steady state conditions.
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2. The examined model In this study, a double effect absorption chiller powered by parabolic trough collectors (PTCs) is the examined system. The PTCs are able to produce heat in medium-high temperature levels with sufficient performance. The working fluid is selected to be Therminol VP-1 [28] because the operation temperatures are greater than 100 oC. The examined PTC is the Eurotrough ET-150 which is a commercial collector [29]. The storage tank includes the stored thermal oil and it is modeled with the mixing zones model. This is an insulated storage tank in order the thermal losses (Qloss) to be reduced. The absorption chiller is a double effect parallel flow absorption chiller which demands high heat source temperature levels. Figure 1 illustrates the examined system with many details. This system produces cooling in relatively hightemperature levels, from 5oC to 12.5oC in the present study, and it is ideal for cooling applications.
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Figure 1. The examined solar cooling system
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Cold thermal oil with temperature (Tcol,in) leaves the bottom of the storage tank and enters in the collector field. The captured solar irradiation from the PTCs warms the flowing thermal oil and its temperature levels reach the value (Tcol,out). It is essential to state that the parabolic trough collectors are able to utilize only the direct beam irradiation (Gb). In the storage tank, there are mixing zones and by this modeling, the temperature levels inside the storage tank are calculated (Tst 1, Tst2 and Tst3). Thermal oil with temperature (Ts,in) leaves the over part of the storage tank and it goes to the absorption chiller. More specifically, this fluid enters to the high generator of the system and it gives the appropriate heat to the chiller. After the high generator, the thermal oil has lower temperature level (Ts,out) and it returns to the storage tank.
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The examined absorption chiller is a complex system which includes an absorber, two generators, two condensers, two solution heat exchangers, as well as circulators and throttling valves. Saturated weak (Xw) solution (State Point 1 or S.P. 1) leaves the absorber and its pressure increases with a circulation pump (S.P. 2). This solution continues its flow in the low heat exchanger where its temperature increases (S.P. 3). On the other side, strong (Xs) solution (S.P. 4) leaves the low generator and enters in the low heat exchanger. In this device, heat is transferred from the strong and hot solution to the cold and weak solution. In the outlet of this device, the strong solution has a lower temperature than in its inlet (S.P. 5). The next step is the expansion of the throttling valve (S.P. 6) and finally, it returns to the absorber.
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The low generator is an important device of this system where many streams are engaged. The demanded heat input (Qgl) in this device is given by the high condenser (Qch) which operates in higher temperature level (Tch > Tgl), and thus Qgl is equal to Qch. Weak solution (S.P. 11) leaves the low generator and continues its flow to the circulation pump (S.P 12) and to a high heat exchanger (S.P. 13). On the other side, strong solution leaves the high generator (S.P. 14) and it enters to the high heat exchanger where its temperature is getting lower (S.P. 15) and after the throttling valve (S.P. 16) it enters to the low generator.
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In the high generator, the demanded heat input (Qgh) is given by the heat source (solar energy) and the suitable quantity of water is evaporated (S.P. 17). This superheated steam goes to the high condenser where heat is rejected to the low generator (Qch) and finally saturated water (S.P. 18) leaves the high condenser. This stream continues its flow in the throating valve (S.P. 19) and after this device, it enters in the low condenser. Moreover, superheated steam from the low generator (S.P. 7) enters in the low condenser in order heat to be rejected in the ambient (Qcl). The saturated water (S.P. 8) enters in the throttling valve and water/steam mixture is created (S.P. 9). This mixture absorbs the cooling load in the evaporator (Qe) and it is transformed to saturated steam (S.P. 10). This quantity returns to the absorber where heat is rejected in the environment (Qa) and the absorption cycle closes.
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3. Mathematical background In this section, the mathematical modeling of the presented system (Figure 1) is given with many details. All the used equations are presented and explained in order to be clear the way that the calculations have been made.
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3.1 Solar system modeling The useful heat from the solar collectors can be calculated with the energy balance in the fluid volume, as equation 1 shows:
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Qu mcol c p Tcol, out Tcol,in ,
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The available solar irradiation is the product of the collecting area (Aa) and the solar beam irradiation (Gb), as equation 2 shows:
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Qsol Aa Gb ,
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The thermal efficiency (ηth) of the solar collector is the ratio of the useful heat to the available solar irradiation:
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th
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The mass flow rate in the PTC circuit (mcol) is determined by the usual formula which is given below [13]:
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mcol kg s 1 0.02 kg s 1 m2 Aa m2 ,
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In the present study, the thermal efficiency curve equation is used. More specifically, this equation correlates the thermal efficiency with the thermal oil inlet temperature (Tcol,in), the ambient temperature (Tam) and the available solar beam irradiation (Gb). Equation 5 shows the formula for the thermal efficiency (ηth) calculation [30]:
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Tcol Tam T T , (5) th 0.75 K 0.000045 Tcol,in Tam 0.039 col,in am 0.3 ,in Gb Gb
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The exergetic output of the solar collector can be calculated according to equation 6 [20]:
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T Eu Qu mcol c p Tam ln col, out , Tcol,in
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The exergy flow of the solar energy can be estimated by the Petela formula [31]. This equation seems to be different from the usual exergetic definition because the sun is a radiative reservoir and not a heat reservoir.
(1)
(2)
Qu , Qsol
(3)
(4)
2
(6)
6
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4 T 1 T 4 Esol Qsol 1 am am , 3 Tsun 3 Tsun
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It is essential to state that the temperature levels in equations 6 and 7 have to be in Kelvin units and the sun temperature (T sun) is the mean temperature in its outer surface which is taken as 5770 K [31].
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The exergetic efficiency of the collectors is the ratio of the useful exergy flow to solar exergy flow, as equation 8 presents:
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ex, col
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The thermal and the exergetic performance of the collectors are presented in figure 2. It is obvious that higher operating temperature levels lead to lower thermal performance but to higher exergetic performance. These results show that the performance of the collector is fully depended on the operating temperature level (Tcol,in).
Eu , Es
(7)
(8)
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Figure 2. Thermal and exergetic performance of the examined parabolic trough collector for various operating conditions (θ=0o)
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3.2 Storage tank modeling In this study, the storage tank simulation is based on the mixing zones modeling. The storage tank is separated into zones where a uniform temperature level is assumed. The energy balance in every mixing zone is made and all the energy balance equation creates a system of differential equations. Equations 9 to 11 present the mixing zone modeling of this study. It is essential to state that three zones have to be selected, something that had been done in other literature studies [13, 20].
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M T col cp Tcol,out Tst1 m s cp Tst 2 Tst1 U L Ast1 Tst1 Tam , c p st1 m 3 t
(9)
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T M col cp Tst1 Tst 2 m s cp Tst3 Tst 2 U L Ast 2 Tst 2 Tam , c p st 2 m 3 t
(10)
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T M col cp Tst 2 Tst3 m s cp TS ,out Tst3 U L Ast3 Tst3 Tam , c p st3 m 3 t
(11)
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The analysis is performed in steady state conditions something that means that the terms with time variation are equal to zero. The present modeling is coupled with the following assumptions. The heat source temperature (TS,in) is equal to the temperature level in the first mixing zone (Tst1) and the temperature at the inlet of the collector field (Tcol,in) is equal to the temperature of the third zone (Tst3). The thermal losses (Qloss) of the storage tank are the summary of the thermal losses of all the zones, as equation 12 shows:
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Qloss U L Ast 2 Tst 2 Tam U L Ast 2 Tst 2 Tam U L Ast3 Tst3 Tam ,
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The thermal loss coefficient (UL) includes the convection, conduction and radiation losses and it is a simple and accurate way for calculating the thermal losses in every zone.
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The storage tank volume (V) is calculated according to equation 13, which has also been used in other studies [13, 20, 32].
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V m3
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The storage tank has been assumed to be a cylinder with length L and diameter D. In this case, the total volume of the tank is calculated as:
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V
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The length of the tank is assumed to be equal to the diameter, an assumption that has negligible influence on the final results. The total mass (M) that is stored in the tank can be calculated according to equation 15, using the fluid density (ρ):
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M V ,
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The outer surfaces of the zones (Ast) are calculated by the equations 16 to 18:
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Ast1
Aa m 2 , 30
D2 4
4
(13)
L ,
D2
(12)
(14)
(15)
DL 3
,
(16)
8
DL
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Ast 2
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Ast3
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3.3 Absorption chiller modeling In this section, the equation which describes the modeling of the double effect absorption chiller are presented with many details. Energy rate and mass flow rate balances have to be made in all the devices in order finally all state points of figure 1 to be determined. Before describing the mathematical equations, it is important to give the main assumptions of this study [13, 20, 33]:
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- The points 1, 11, 4, 14 are assumed as a saturated solution.
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- The points 8 and 18 are assumed to be saturated water, while point 10 is saturated steam.
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- The expansions in throttling valves are assumed to be done without energy losses, which mean that there is no change in the fluid enthalpy.
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- The pumping work is assumed to be negligible.
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- The demanded input energy in the low generator is given by the high condenser.
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- The heat exchange in the solution heat exchangers is adiabatic without thermal losses to the ambient.
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- The absorber and the low condenser operate at the same temperature level which is called heat rejection temperature level.
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- The cooling load is assumed to be 100 kW in all the cases; an accepted value for double-effect absorption chillers.
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It is essential to state that the modeling of this cycle can be performed with seven different mass flow rates: m1 , m4 , m7 , m10 , m11 , m14 and m17.
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3.3.1 Absorber The energy balance in the absorber is given by the next equation. In the absorption process, heat is rejected to the ambient.
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Qa m10 h10 m4 h6 m1 h1 ,
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The total mass flow rate balance is given by equation 20 and the mass flow rate balance for LiCl or LiBr mass is given in equation 21:
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m1 m4 m10 ,
3
D2 4
,
(17)
DL 3
,
(18)
(19)
(20)
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m1 X w m4 X s ,
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3.3.2 Generators In the present chiller there are two generators, the low and the high. The energy balances in the low and the high generators are given by equations 22 and 23 respectively.
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Qgl m4 h4 m11 h11 m7 h7 m1 h3 m14 h16 ,
(22)
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Qgh m14 h14 m17 h17 m11 h13 ,
(23)
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The total mass flow rate balance in high generator is given by equation 24 and the mass flow rate balance for LiCl or LiBr mass is given in equation 25:
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m11 m14 m17 ,
(24)
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m11 X w m14 X s ,
(25)
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3.3.3 Condensers There are two condensers in this system, the low and the high condenser. It is essential to state that the high condenser rejects heat to the low generator and the low condenser rejects heat to the ambient. Equations 26 and 27 show the low and the high condenser heat output calculations
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Qcl m7 h7 m17 h19 m10 h8 ,
(26)
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Qch m17 h17 h18 ,
(27)
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The mass flow rate balance in the low condenser is also needed in the present modelling:
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m10 m7 m17 ,
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3.3.4 Heat exchangers The heat exchangers are important for increasing the performance of the absorption chiller. In every heat exchanger, the energy balance between the two streams and the heat exchanger efficiency are the needed equations for presenting a completed modelling. These equations for the low heat exchanger are presented below:
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m1 h3 h2 m4 h4 h5 ,
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hex,l
320
The respective equations for the high heat exchanger are given below:
h4 h5 , h4 h2
(21)
(28)
(29) (30)
10
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m11 h13 h12 m14 h14 h15 ,
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hex, h
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3.3.5 Other equations The energy balance in the evaporator connects the refrigerant mass flow rate (m10) and the cooling load (Qe), as equation 33 shows:
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Qe m10 h10 h9 ,
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The throttling valves do not change the enthalpy of the fluid. Equations 34 to 37 present this modelling:
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h5 h6 ,
(34)
330
h8 h9 ,
(35)
331
h15 h16 ,
(36)
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h18 h19 ,
(37)
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The work in the pumps is assumed to be negligible and so equations 38 and 39 can be written as:
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h2 h1 Wl h1 ,
(38)
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h12 h11 Wh h11 ,
(39)
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Moreover, some important assumptions that have been explained above are useful also to be given with the following equations. More specifically, equation 40 shows that the high condenser rejects heat to the low generator:
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Qch Qgl ,
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The temperature level in the absorber is the same as in the low condenser:
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Ta Tcl ,
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Another important assumption that is made is that the temperature level between the high condenser (Tch) and the low generator (Tgl) is about 10 K, as equation 42 shows.
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Tgl Tch 10 ,
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3.4 Modeling of the heat exchange in high generator In this section the equations for the modelling of the heat exchange in the high condenser are presented [13, 20, 34]. The temperature difference between the
h14 h15 , h14 h12
(31) (32)
(33)
(40)
(41)
(42)
11
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generator temperature and the high heat source temperate (T S,in) is assumed to be 10 K as equation 43 shows:
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Tgh TS ,in 10 ,
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The difference between the inlet and the outlet of the heat source in the high generator is 7K.
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TS ,out TS ,in 7 ,
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The mass flow rate of the heat source is calculated by equation 45, if the high generator heat demand is known.
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Qgh mS c P TS ,in TS ,out ,
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3.5 Energetic and exergetic indexes The absorption chiller performance can be determined by the coefficient of performance (COP) and its exergetic efficiency. Equation 46 shows the calculation of the COP for the chiller. It is important to state that the work of the pump is neglected:
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COP
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The useful exergetic output of the chiller (E e) is calculated using the cooling load in the evaporator, as equation 47 shows [13, 20]:
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T Ee Qe 1 am , Te
366
The exergetic input in the high generator (E gh) is given as:
367
T E gh Qgh 1 am , T gh
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It is important to state that the temperature levels in equations 47 and 48 have to be in Kelvin units. The exergetic performance of the chiller (ηex,ch) is the ratio of the exergetic output in the evaporator (E e) to the exergetic input in the high generator (Egh).
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ex,ch
373
Using equations 46 to 48, the above equation can be written as:
Qe , Qgh
Ee , E gh
(43)
(44)
(45)
(46)
(47)
(48)
(49)
12
T 1 am Te COP , Tam 1 T gh
374
ex,ch
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In the total system (solar cooling system), similar indexes can be determined. More specifically, the solar coefficient of performance (SCOP) is given as:
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SCOP
378
The exergetic efficiency of the solar cooling system is given by equation 52:
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ex , sol
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4. Methods 4.1 Methodology explanation The followed modelling, which is based on the equations of section 3, is an accepted modeling according to the literature [13, 20]. In the present analysis, firstly a parametric analysis of the absorption chiller is performed and after a parametric analysis of the solar cooling system is given.
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In the absorption chiller investigation, numerous combinations of operation conditions have been investigated. Three low condensation temperature levels (30 oC-35oC40oC), four evaporator temperature levels (5.0 oC-7.5oC-10.0oC-12.5oC) and two working pairs (LiCl-H2O and LiBr-H2O) create 24 different study cases. For every studied case, the generator temperature is examined parametrically (with step 5K) in the possible operational range. In every studied case, the generator temperature which maximizes the exergetic performance of the chiller is assumed to be the optimum one.
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In the solar cooling system investigation, the evaporator temperature level was selected to be equal to 10oC [13, 20], as the most representative value, while the low condensation temperature was examined parametrically (30oC-35oC-40oC) and both LiCl-H2O and LiBr-H2O are investigated. So, six different cases are examined and in every case, the heat source temperature (the temperature level of the thermal oil in the high generator inlet) is examined parametrically in the respective range. The heat source temperature levels which lead to a maximum exergetic efficiency of the solar system is selected as the optimum. In the final evaluation of the examined cases, the demanded collecting are is an important criterion which is taken also into account.
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In the analysis that is presented, some parameters have selected to be constant for all the examined cases in order to perform a suitable comparison between the examined working pairs. Table 1 includes these values [13, 20].
Qe , Qsol
Ee , Esol
(50)
(51)
(52)
13
405
Table 1. Constant parameters for all the examined cases Parameter Values ηhex,l 65 % ηhex,h 65 % Tam 30 oC Gb 1 kW∙m-2 UL 0.7∙10-3 kW∙m-2∙K-1 Qe 100 kW Tsun 5770 K ο K(θ=0 ) 1
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The developed model includes many equations which are solved in the EES (Engineering Equation Solver) [35]. This is a powerful tool for solving a great system of non-linear equations and also this program includes subroutines for fluid properties. The properties for Therminol VP-1, LiCl-H2O and LiBr-H2O are taken from the references [28, 36-37] respectively. More specifically, it is essential to explain that the used LiCl-H2O properties in this study are taken from Patek and Klomfar study [36] and these properties are in accordance with the recent study of Li et al. [38]. Figure 3 shows the flow chart of the developed model and the basic equations which are needed for the calculations are included in this figure.
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Figure 3. The flow chart of the developed model in EES
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4.2 Absorption chiller modeling validation The developed model for the absorption chiller is compared with a respective system in the literature. More specifically, the same double effect absorption chiller has been analyzed in the Ref. [39] with many details. In order to conduct a suitable test of the developed model, the same operating conditions were inserted in the model and the results are compared with the respective from the literature [39]. Table 2 includes the results of this comparison. More specifically, the inputs and the outputs, as well as the deviation between the literature and the present study results are given. This test has been performed for operation with LiBr-H2O pair. The comparison proves that the 15
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developed model gives reliable results with the maximum deviation to be up to 2%. More specifically, for the COP, which is the most important parameter, the deviation is about 1.06%, a very small and accepted value.
OUTPUTS
INPUTS
430
Table 2. Validation of the developed model Parameters Literature [39] Present study Te (oC) 5.0 5.0 o Tcl ( C) 30.0 30.0 Ta (oC) 30.0 30.0 o Tgh ( C) 145.0 145.0 ηhex,l (%) 50.0 50.0 ηhex,h (%) 50.0 50.0 COP 1.325 1.339 Xw (%) 52.77 52.57 Xs (%) 61.97 61.82 Pmin (kPa) 0.88 0.87 Pm (kPa) 4.17 4.25 Pmax (kPa) 64.23 63.79 Qe (kW) 354.4 355.0 Qg (kW) 267.5 265.2 Qa (kW) 436.2 438.1 Qcl (kW) 185.7 182.1
Deviation (%) 1.06 0.38 0.24 1.14 1.92 0.69 0.17 0.86 0.44 1.94
431 432 433 434 435 436
5. Results In this section, the results of the simulations are presented. In section (5.1) the performance of the chiller is presented and in section (5.2) the solar cooling system performance is given. The last section of this section (5.3) is devoted to discussing the results in order the important points of this study to be noticed.
437 438 439 440
5.1 Absorption chiller performance In this section, the performance of the double effect absorption chiller operating in various conditions is presented. Figures 4 to 6 show the coefficient of performance of the chiller and figures 7 to 9 depict the exergetic performance of the chiller.
441 442 443 444 445 446 447 448
In figures 4 to 6, the COP for various evaporating temperature levels is given for both LiCl-H2O and LiBr-H2O working pairs. The low condensation temperature is different from these figures and it takes the values 30oC, 35oC and 40oC in figures 4, 5 and 6 respectively. Practically, this temperature level is the heat rejection temperature and it influences on the results, as it is obvious from figures 4 to 6. Lower heat rejection temperature level leads to the higher coefficient of performance, something that is logical and accepted. Moreover, higher condensation temperature needs greater generator levels, as it is shown in figures 4 to 6.
449 450 451 452
Another important conclusion from figures 4 to 6 is about the COP curves shape. In every curve, after a critical generator temperature level, the COP presents small increase and the curve tends towards to horizontal. This observation shows that in every case there is a minimum generator temperature level which leads to a sufficient 16
453 454 455 456
COP for the chiller. This phenomenon is more intense in low evaporator temperature levels. At this point, it is also useful to state that higher evaporating temperature leads to higher COP because the cooling production is easier to be achieved at higher temperatures.
457 458 459 460 461
The working pair LiCl-H2O is proved to be the best working fluid in all the examined operating conditions. On the other hand, according to figures 4 to 6, this working pair is able to operate in a smaller range of generator temperature level because of crystallization problems. This is an important restriction in its application and it has to be taken into account in the design of these systems.
462 463 464
Figure 4. Coefficient of performance of the chiller for low condenser temperature equal to 30oC
17
465 466 467
Figure 5. Coefficient of performance of the chiller for low condenser temperature equal to 35oC
468
469 470 471
Figure 6. Coefficient of performance of the chiller for low condenser temperature equal to 40oC
472 473 474 475
Figures 7 to 9 illustrate the exergetic performance of the chiller for all the examined cases. These figures are very interesting because the curves are maximized for a specific generator temperature level in every case. LiCl-H2O is the working pair with the highest exergetic performance for all the examined operating conditions, a result 18
476 477 478 479 480
which is in accordance with the energetic investigations based on the COP index. The optimum operating temperature for the high generator is lower for the LiCl-H2O compared to the respective for LiBr-H2O. This result indicates that the LiCl-H2O needs lower exergetic input in order to operate optimally, something important which leads to higher exergetic performance.
481 482 483
The exergetic performance is influenced by the heat rejection temperature level (T cl). Greater values of this parameter lead to lower exergetic results due to the lower COP, as equation 50 indicates.
484 485 486 487 488 489 490 491 492 493 494 495 496
Comparing the curves of different evaporating temperature levels, the analysis is more complex. By evaluating the maximum points of the curves, lower evaporating temperature level leads to lower exergetic performance because greater generator temperature level (Tgh) is demanded. On the other hand, in high generator temperatures, the lower evaporating temperature levels are connected with higher exergetic performance. Thus, in every studied case, a specific analysis is needed in order to determine the optimum generator temperature level because this parameter is fully connected with the temperature level in the evaporator, as well as the heat rejection temperature level. Table 3 includes all the results for the exergetic comparison. The optimum operating conditions (optimum points from figures 7 to 9) are given in this table. For all of these cases, the COP deviation between the two examined working pairs is presented. Finally, the working pair LiCl-H2O is proved to be more efficient with a mean enhancement of 7.52%, compared to LiBr-H2O.
497 498 499
Figure 7. Exergetic performance of the chiller for low condenser temperature equal to 30oC
19
500 501 502
Figure 8. Exergetic performance of the chiller for low condenser temperature equal to 35oC
503
504 505 506
Figure 9. Exergetic performance of the chiller for low condenser temperature equal to 40oC
507 508
20
509
Table 3. Results for optimum exergetic operation of the chiller Cases LiCl-H2O LiBr-H2O Deviation Tcl Te Tgh Tgh COP ηex,ch COP ηex,ch COP (oC) (oC) (oC) (oC) (%) 5.0 110 0.3283 1.447 115 0.2836 1.346 7.50 7.5 100 0.3378 1.397 105 0.2970 1.334 4.72 30 10.0 95 0.3527 1.491 100 0.3120 1.439 3.61 12.5 90 0.3602 1.570 90 0.3296 1.437 9.25 5.0 130 0.2385 1.383 135 0.2057 1.268 8.91 7.5 120 0.2435 1.365 130 0.2100 1.341 1.79 35 10.0 115 0.2482 1.446 120 0.2149 1.343 7.67 12.5 110 0.2464 1.511 110 0.2175 1.334 13.27 5.0 150 0.1842 1.345 160 0.1561 1.264 6.41 7.5 150 0.1818 1.462 150 0.1576 1.267 15.39 40 10.0 135 0.1828 1.384 140 0.1575 1.264 9.49 12.5 125 0.1805 1.366 135 0.1559 1.337 2.17
510 511 512 513 514 515 516 517
5.2 Solar cooling system performance In this section, the investigation of the solar cooling system is presented. The working pairs LiCl-H2O and LiBr-H2O are compared under the same operating conditions (evaporator temperature level and heat rejection temperature level) and they are analyzed parametrically with the heat source temperature level (Ts,in). The curves in figures 10 to 14 show results for three heat rejection temperatures levels (30oC - 35oC - 40oC) with the evaporating temperature level to be equal to 10oC.
518 519 520 521 522 523 524 525 526
The energetic and exergetic performance of the system, which is determined by COP (Figure 10) and exergetic efficiency (Figure 11), prove that the LiCl-H2O working pair is better than the LiBr-H2O in all the examined cases. According to figures 10 and 11, higher heat rejection temperature level (T cl) leads to lower COP and to lower exergetic performance respectively. Moreover, these figures prove that higher heat rejection temperature needs higher heat source temperature level for the system operation. The curves in figures 10 and 11 have an increasing rate with the increase of the heat source temperature up to a point and after this point, they tend towards to horizontal.
527 528 529 530 531 532
Similar conclusions can be extracted from the results presented in figure 12. The solar coefficient of performance (SCOP) is given in this figure and the curves for LiCLH2O prove the higher performance of this working pair. Moreover, the curves of figure 12 are similar to the curves of figure 10, but they correspond to lower value. The SCOP of the system is lower than the COP of the chiller because SCOP includes the thermal losses of the solar collector and of the storage tank.
533 534 535 536
Figure 13 shows the results for the exergetic performance of the chiller in the solar cooling system for various heat source temperature levels. In every case, there is a specific temperature level which maximizes the performance of the chiller but this is not the optimum heat temperature level of the system. The optimum operating 21
537 538 539 540 541 542
temperature level (Ts,in) is the one which maximizes the exergetic performance of the total cooling system, including the solar collectors, the storage tank, and the absorption chiller. These optimum temperature levels, which are different for different operating conditions, are determined by the results in figure 11. It is important to state that the maximization of the solar exergetic efficiency (ηex,sol) is fully connected with the maximization of the SCOP, as the final results prove.
543 544 545 546 547 548 549 550
Figure 14 depicts the demanded collecting area of PTC for producing the needed cooling load. All the curves in figure 14 are minimized for a specific heat source temperature level (different for every curve) which is the same temperature level for the maximization of the system exergetic performance and of the SCOP. Thus, in every case, the system has to be designed to operate close to the optimum heat source temperature level in order to minimize the collecting area. This design leads to minimization of the investment cost and to smaller installation, while the cooling production is kept constant in all the examined cases.
551 552 553 554 555 556 557 558 559 560 561 562 563 564 565
Table 4 shows the results for the optimum operation of the double effect absorption chiller coupled with parabolic trough collectors and for operating with the two examined working pairs. The results show that different heat source temperature levels are needed when there are different operating conditions. The working pair LiCl-H2O demands lower heat source temperature levels than the LiBr-H2O. This result shows that the LiCl-H2O working pair needs lower exergetic input in the chiller in order to produce the same output and so this is proved to be the more efficient solution. Moreover, it is obvious that higher condensation temperature (T cl) or heat rejection level leads to higher optimum heat source temperature level. Furthermore, the demanded collecting area is lower for the LiCl-H2O case with a mean difference of 7.59%. This difference is smaller for lower heat rejection levels. More specifically, for heat rejection at 30oC, the difference in the collecting area is 5.04%, while for 35oC and 40oC it is about 6.92% and 10.82% respectively. In the comparison of the other indexes, it is remarkable to state that the mean enhancement of the SCOP for operation with LiCl-H2O is about 7.55%.
22
566 567 568
Figure 10. Coefficient of performance of the chiller in the solar system for various low condenser temperature levels
569
570 571 572
Figure 11. Exergetic performance of the solar system for various low condenser temperature levels
23
573 574 575
Figure 12. Solar coefficient of performance of the solar system for various low condenser temperature levels
576
577 578 579
Figure 13. Exergetic performance of the chiller in the solar system for various low condenser temperature levels
580
24
581 582 583
Figure 14. Demanded collecting area of the solar cooling system for various low condenser temperature levels
584 585
Tcl (oC) 30 35 40
Table 4. Comparison of the cases with minimum collecting area Tgh Aa Working Deviation of Aa COP SCOP ηex,sol pair (%) (oC) (m2) LiCl-H2O 125 1.642 1.140 0.0764 87.69 5.04 LiBr-H2O 135 1.577 1.086 0.0727 92.11 LiCl-H2O 150 1.594 1.082 0.7250 92.43 6.92 LiBr-H2O 160 1.305 1.012 0.6780 98.83 LiCl-H2O 160 1.526 1.026 0.0687 97.48 10.82 LiBr-H2O 180 1.438 0.948 0.0635 105.5
586 587 588 589 590 591 592 593
5.3 Discussion of the results The results of this study show that the use of LiCl-H2O in parallel flow double-effect absorption chiller leads to higher performance than the use of LiBr-H2O in the same system. The performance enhancement is both energetic and exergetic and it is depended on the heat rejection temperature level. More specifically, the mean enhancement is about 7.5% and it is getting greater for higher heat rejection temperatures, as table 4 exhibits.
594 595 596 597 598 599
Respective results had been presented in Ref [13], where a similar single effect system had been analyzed. By combining the results of this study and of the Ref. [13] results, it is clear that LiCl-H2O is thermodynamically the most suitable working pair, compared with the LiBr-H2O. Extra advantages of LiCl-H2O are its greater long-term stability and its higher internal energy storage capacity. On the other hand, there is greater crystallization danger because the maximum allowed solution concentration of 25
600 601 602 603 604 605 606
LiCl is about 50%, lower than 75% in LiBr case [36-37]. Moreover, the generator temperature level range is lower for the LiCl-H2O case than the LiBr-H2O case, as it is shown in the results of this study. Thus, the operation with the alternative working pair (LiCl-H2O) is a promising way for designing more efficient solar cooling system but careful design is needed due to crystallization problems. Especially in the doubleeffect chiller, the increase in the cooling production is greater than in single-effect machines because the COP is greater in the double-effect systems.
607 608 609 610 611 612 613 614 615 616
Moreover, according to the results of table 4, the specific collecting area is calculated to be close to 1 m2∙kW-1 ; a low value compared to the single-effect chillers. In literature, similar values for the double-effect systems have been referred. More specifically, the specific collecting area in systems with double-effect absorption chillers operating with LiBr-H2O is found to be 1.25 and 0.94 in Refs. [40] and [41] respectively. This result is very interesting and it makes the solar cooling with high efficient machines (as the double effect case) to be a feasible solution in cases with a restricted available area for installing solar collectors. Moreover, the relative low collecting area makes the system more attractive from the financial point of view, something that has to be analyzed more in the future.
617 618 619 620 621 622 623 624 625 626 627 628 629 630 631
6. Conclusions In this study, the use of an alternative working pair (LiCl-H2O) in a double-effect absorption chiller powered by parabolic trough collectors is investigated energetically and exergetically. Different combinations of evaporating temperature levels, heat rejection temperature levels, and heat source temperature levels are examined in order to present a multilateral study. The results of the analysis are compared with the respective for operation with the conventional working pair, LiBr-H2O. Finally, it is proved that the LiCl-H2O leads to higher performance in energetic and exergetic terms. More specifically, an enhancement of 7.55% in solar coefficient of performance is proved, while a similar reduction (7.59%) in the demanded collecting area is also found. In all the examined cases, both the COP and the exergetic performance of the system are higher for the LiCl-H2O case, with the enhancement to be greater in higher heat rejection temperature levels. Another interesting result is that the optimum heat source temperature level is lower for the LiCl-H2O case, a fact that also proves its higher exergetic performance.
632 633 634 635 636 637
These results are important for establishing the use of this working pair in future installations and the solar cooling can be a more feasible solution from the financial point of view due to the reduction in the specific collecting area. On the other hand, constraints about the LiCl concentration, the generator temperature level range and the system stability have to be taken carefully into account when the LiCl-H2O is used as working pair in absorption chillers.
638 639
26
640
Nomenclature
641
Aa
Collecting area, m2
642
Ast
Storage tank area, m2
643
cp
Specific heat capacity, kJ∙kg-1∙K-1
644
COP
coefficient of performance, -
645
D
Diameter, m
646
E
Exergy flow, kW
647
Gb
Solar direct beam radiation, kW∙m-2
648
h
enthalpy, kJ∙kg-1
649
K
Incident angle modifier, -
650
L
Tank height, m
651
m
Mass flow rate, kg∙s-1
652
M
The storage thermal oil mass, kg
653
p
Pressure, kPa
654
Q
Heat rate, kW
655
SCOP Solar coefficient of performance, -
656
t
Time, s
657
T
Temperature, oC
658
UL
Tank heat loss coefficient, kW∙m-2∙K-1
659
V
Tank volume, m3
660
W
Pump work, kW
661
Xs
Mass concentration of strong mixture, -
662
Xw
Mass concentration of weak mixture, -
663
Greek symbols
664
η
efficiency, -
665
θ
Incident angle, o
666
ρ
Thermal oil density, kg∙m-3 27
667
Subscripts and superscripts
668
a
Absorber
669
am
Ambient
670
c
Condenser
671
ch
Condenser high
672
cl
Condenser low
673
col
Collector
674
e
Evaporator
675
eff
Effective
676
ex,ch Exergetic chiller
677
ex,col Exergetic collector
678
ex,sol Exergetic solar system
679
h
High
680
hex
Heat exchanger
681
in
Inlet
682
g
Generator
683
gh
High generator
684
gl
Low generator
685
l
Low
686
low
Low level
687
loss
Heat losses
688
m
Medium level
689
max
high level
690
out
Outlet
691
S
Heat source
692
Sat
saturated
693
sol
Solar energy 28
694
st1
1st zone of the storage tank
695
st2
2nd zone of the storage tank
696
st3
3rd zone of the storage tank
697
sun
Sun
698
th
Thermal
699
u
Useful
700
Abbreviations
701
EES
Engineer Equator Solver
702
ETC
Evacuated tube collector
703
LiBr
Lithium bromide
704
LiCl
Lithium chloride
705
PTC
Parabolic trough collector
706
S.P.
State point
707
SCOP Solar coefficient of performance
708 709 710 711
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