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Thermodynamic analysis of a novel energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation
Energy Conversion and Management 78 (2014) 286–296
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Thermodynamic analysis of a novel energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation Xiaohui She, Yonggao Yin ⇑, Xiaosong Zhang School of Energy and Environment, Southeast University, Nanjing 210096, China Ministry of Education of Key Laboratory of Energy Thermal Conversion and Control, Southeast University, Nanjing 210096, China
a r t i c l e
i n f o
Article history: Received 20 July 2013 Accepted 24 October 2013 Available online 21 November 2013 Keywords: Refrigeration Liquid desiccant Subcooling Condensation heat utilization Energy efficiency improvement
a b s t r a c t A new energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation was proposed in this paper. In the system, liquid desiccant system could produce very dry air for an indirect evaporative cooler, which would subcool the vapor compression refrigeration system to get higher COP than conventional refrigeration system. The desiccant cooling system can use the condensation heat for the desiccant regeneration. Thermodynamic analysis is made to discuss the effects of operation parameters (condensing temperature, liquid desiccant concentration, ambient air temperature and relative humidity) on the system performance. Results show that the proposed hybrid vapor compression refrigeration system achieves significantly higher COP than conventional vapor compression refrigeration system, and even higher than the reverse Carnot cycle at the same operation conditions. The maximum COPs of the hybrid systems using hot air and ambient air are 18.8% and 16.3% higher than that of the conventional vapor compression refrigeration system under varied conditions, respectively. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent years, refrigeration systems consumed a large amount of energy in maintaining thermal comfort for occupants and suitable climatic conditions for cooling cases, which made up 50% of building energy consumption [1]. The utilization of evaporative condensing indicated energy efficient potential in reduction of power requirements [2,3]. In a traditional refrigeration system, a great deal of condensation heat, which could be used for other purposes, is directly dissipated to the environment. The dissipated heat not only wastes energy, but also causes severe heat island effect in the surrounding areas. Many methods have been attempted to tackle these problems. Some researchers utilized the condensation heat from air conditioners to preheat domestic hot water [4–7], leading to claims that water heating in summer, primarily for bathing. It can be made available virtually free whenever space cooling is required, and is considered one of the most cost effective energy conservation measures. Several other researchers tried to add a heat recovery system on the refrigeration system. For instance, a condensing heat recovery with thermal storage of phase change material (paraffin wax as PCM) was designed and analyzed by Zhang et al. [8]. In addition, Kaushik et al. [9] introduced a canopus heat exchanger ⇑ Corresponding author at: School of Energy and Environment, Southeast University, Nanjing 210096, China. Tel.: +86 25 83792722. E-mail address: [email protected] (Y. Yin). 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.10.057
for heat recovery for a refrigeration system. A drawback with heat recovery system is the high condensing temperature that increases the energy consumption of the refrigeration system. To solve this problem, Arias and Lundqvist [10] proposed an alternative to heat recovery, namely floating condensing pressure, which improved the coefficient of performance and decreased the energy consumption of the refrigeration system at lower outdoor temperature. From thermodynamic standpoint, further cooling of liquid refrigerant leaving condenser can significantly increase the cooling capacity and improve the system COP. The regenerative refrigeration cycle adopted a liquid-suction heat exchanger as a sub-cooler to improve the system coefficient of performance (COP) [11–13]. The liquid-suction heat exchanger achieves condensate subcooling by transferring the refrigerant heat from the condenser outlet to the compressor inlet. This subcooling method may result in the superheated degree existing at the compressor inlet and reduce the system efficiency. Therefore the liquid-suction heat exchanger is not suitable for all air conditioning systems. In order to eliminate the drawbacks of the liquid-suction heat exchanger, Khan et al. [14] and Qureshi et al. [15,16] used integrated and dedicated mechanical-subcooling methods to enhance the system COP and also to remove the superheated degree, respectively. Thermoeconomic considerations were given to heat exchanger inventory allocation in vapor compression cycles with mechanical subcooling by Qureshi and Zubair [17], and it was concluded that the cost optimization of the integrated mechanical subcooling system was qualitatively the same as the dedicated subcooling system. Since the
X. She et al. / Energy Conversion and Management 78 (2014) 286–296
287
Nomenclature COP Cp d G h H hC hD L Le m NTU P Q q0 RH T DTsc DTsh w W X
coefficient of performance specific heat capacity (kJ kg1 K1) differential mass flow rate (kg s1) enthalpy (kJ kg1) height (m) heat transfer coefficient (W m2 K1) mass transfer coefficient (kg m2 s1) length (m) Lewis number adiabatic index of refrigerant number of mass transfer unit pressure (Pa) refrigerating capacity (kW) specific refrigerating capacity (kJ kg1) relative humidity (%) temperature (°C) subcooling degree (°C) superheat degree (°C) specific power consumption (kJ kg1) width (m) Concentration (%)
mechanical-subcooling required an additional compressor to provide the driving force of subcooling, the electricity consumption of the compressor would carry the undesirable effect on the system COP. Besides, the expensive initial cost is also one of the drawbacks in the mechanical-subcooling method. To avoid the demerits of mechanical-subcooling method, many studies utilized phase change material and heat pipes in cold storage unit as a subcooler [18–20]. Chieh et al. [21] developed a thermal battery composed of the phase change material and the heat pipes to store energy in air conditioning application. Huang et al. [22] experimentally investigated the performance of the cold storage air conditioning system utilizing a thermal battery as a subcooler. Because the charge heat exchanger of the thermal battery had the larger thermal resistances, it induced the lower ice stored ability. For eliminating the disadvantages discussed in previous studies, Hsiao et al. [23,24] used an ice storage tank as a subcooler to enhance the system capabilities and reduces the superheated degree without inputting the supplemental electricity. By the innovative design of the charge heat exchanger, the overall thermal resistances can be decreased. The liquid desiccant system using low-grade heat resource was proposed by Lof [25], and its application in air-conditioning systems has been widely investigated [26–28]. In this paper, a novel energy-efficient vapor compression refrigeration system combined with a liquid desiccant cooling system is proposed. The vapor compression system is subcooled by the liquid desiccant cooling cycle driven by condensing heat to achieve more subcooling degree of the refrigerant than the conventional system. This paper will discuss the potential of the performance improvement and the effect of different climatic and operating conditions on the performance of the system, and a comparative study was made on the effect of different utilization ways of condensation heat.
2. System description Fig. 1 illustrates the novel refrigeration cycle subcooled by the liquid desiccant dehumidification and evaporation, which is
Greek symbols e effectiveness x humidity ratio (kg kg1) Subscripts a air AHE air-to-air heat exchanger amb ambient com compressor con condenser eva evaporator IEC indirect evaporative cooler in inlet out outlet max maximum min minimum ref refrigerant reg regenerator s solution sc subcooling sh superheat SHE solution-to-solution heat exchanger
composed of refrigeration cycle, closed air cycle and liquid desiccant cycle. The refrigeration cycle includes an evaporator, a compressor, a solution-to-refrigerant heat exchanger and a condenser. Refrigerant R-22 is chosen to be the working fluid inside the refrigeration cycle, and the refrigerating capacity of the baseline system is 30 kW. The closed air cycle consists of an indirect evaporative cooler, an air-to-air heat exchanger, a dehumidifier and an air cooler, while the liquid desiccant cycle is made up of a dehumidifier, a solution cooler, a solution-to-solution heat exchanger, a regenerator and a solution-to-refrigerant heat exchanger. The dehumidifier is internally cooled by cooling water and its physical size is assumed large enough to make the air outlet humidity ratio be in equilibrium with the solution inlet humidity ratio above the surface, which can be achieved as long as the cooling water is sufficient. The regenerator is adiabatic, and its physical scale at several different operating conditions is shown in Table 1. The physical scale of other components of the system is believed large enough to meet our needs. Fig. 2(a) shows the ideal pressure enthalpy diagram of refrigeration cycle. As indicated in Fig. 2(a), the refrigerant at state 4 is subcooled to state 5 by the indirect evaporative cooler, using the dry closed air absorbing moisture from water. A conceptual schematic diagram for the processes of the liquid desiccant cycle and closed air cycle is shown in Fig. 2(b). There is sensible heat recovery of the closed air from state 7 to state 8. The moist air (state 8) is dehumidified in the dehumidifier to go to state 9. The dehumidified air is precooled by an air cooler and then cooled to state 11 by an air-to-air heat exchanger. In indirect evaporative cooler, the cold dry air is humidified by water evaporation and reaches state 7. In the liquid desiccant cycle, state 12 and state 14 are in the isoconcentration line of liquid desiccant solution. The diluted solution is preheated by hot concentrated solution leaving the regenerator in solution-to-solution heat exchanger. Prior to entering regenerator, the liquid desiccant is heated to the specified set point temperature (state 14) in the solution-to-refrigerant heat exchanger. Then the hot diluted solution enters into the regenerator and finally reaches state 15. State 15 and state 17 are also in iso-concentration
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X. She et al. / Energy Conversion and Management 78 (2014) 286–296 Solution cooler
17
Air cooler
9
16
Solution-tosolution heat exchanger
20
15
Dehumidifier
10
Regenerator 13
12
8
19
14
Air-to-air heat exchanger 4
7
11
3 Solution-to-refrigerant heat exchanger
Condenser
2
18
Indirect evaporative cooler
Compressor
Fan
5 Pump
Throttle valve Evaporator
6
1
Fig. 1. Schematic diagram of proposed energy-efficient refrigeration system.
Table 1 The physical scale of regenerator at several different operation conditions. Tcon (°C)
Tamb (°C)
RHamb (%)
Xs (%)
H (m)
W (m)
L (m)
50 50 50 50 50 52.5 55
31 35 35 35 35 35 35
60 30 40 50 60 60 60
33 33 33 33 31 33 33
0.65 0.34 0.43 0.64 0.57 1.02 0.46
0.65 0.34 0.43 0.64 0.57 1.02 0.46
0.65 0.34 0.43 0.64 0.57 1.02 0.46
line of liquid desiccant solution. The hot concentrated solution leaving regenerator is precooled by solution-to-solution heat exchanger, and it is cooled to state 17 by solution cooler. The concentrated solution goes to the dehumidifier and reaches state 12. Effective use of the condensation heat is quite vital to the performance of the desiccant cycle. In the following research, condensation sensible heat is employed to heat the solution before it enters regenerator. Different air is used in the regeneration process to make a comparative study, namely hot air from condenser and
the ambient air. The solution mass flow rate (Gs) and air mass flow rate (Gamb) should be chosen reasonably in order to achieve the best regeneration effect. As far as one certain regeneration process (i.e. the inlet parameters are known) is concerned, there exists minimum regeneration temperature (Ts,min). If the air inlet temperature of regenerator (Tamb,reg,in) is much higher than Ts,min, then the more the Gamb, the better the regeneration effect. Whereas if Tamb,reg,in is below Ts,min, Gamb should be lower than Gs and be chosen properly, as shown in Fig. 3. This is due to the fact that if Tamb,reg,in is beyond Ts,min, then the air is able to restrain heat transfer from solution to air and improve the regeneration effect significantly. When Tamb,reg,in is lower than Ts,min, the solution will transfer heat to the air, and its temperature would decrease considerably, which will reduce the vapor pressure above solution and be not helpful for the regeneration process.
3. Modeling of energy-efficient refrigeration system In this section, a thermodynamic model of the energy-efficient refrigeration system is developed for the performance analysis.
P
(b)
Subcooled 5
6
4
14
2
3
15
Temperature (oC)
(a)
1
17 9
Xs
12 8
00%
RH=1
11
7
h
Humidity ratio (kg/kg) Fig. 2. Refrigeration cycle baseline and process air line in liquid desiccant cycle: (a) pressure-enthalpy diagram of the refrigeration cycle and (b) process air line in liquid desiccant cycle.
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3.2. Model description
Solution concentration Xs
Air inlet parameters of regenerator Tamb,reg,in, RHamb,reg,in
For a given refrigerating capacity of the baseline system, the mass flow rate of refrigerant (Gref) can be obtained. The solution temperature entering regenerator (T14) is calculated through energy conservation of the solution-to-refrigerant heat exchanger, as shown in Eq. (1).
Air humidity ratio
ωamb,reg,in
T 14 ¼ T 13 þ Gref ðh2 h3 Þ=ðGs Cps Þ
ð1Þ
The ambient air mass flow rate (Gamb) is calculated using Eq. (2). For different operating conditions, air mass flow rate in the regenerator should be chosen properly to achieve the best regeneration effect. For example, in some cases, all the Gamb should be used in the regenerator, while in other occasions, only part of Gamb is chosen.
Minimum regeneration temperature Ts,min =f (ωamb,reg,in,Xs)
Gamb ¼ Gref ðh3 h4 Þ=ðh19 h18 Þ Tamb,reg,in>Ts,min
Tamb,reg,in
Case 1: using hot air Choose all the heat from condenser and calculate Gamb Case 2: using ambient air Choose the optimal Gamb to achieve the best regeneration effect
Case 1: using hot air Choose proper ratio of the condensation latent heat and calculate Gamb Case 2: using ambient air Choose proper ratio (Gamb /Gs) to achieve the best regeneration effect
ð2Þ
In the regenerator, NTU-Le model and finite element method [31] are employed to analyze internal changes of regenerator, as shown in Fig. 4. The governing equations for heat and moisture conservation in the air and solution can be written as follows.
dxa ¼ dNTU ðxs xa Þ
ð3Þ
dT a ¼ dNTU Le ðT s T a Þ
ð4Þ
dha ¼ dNTU Le ½ðhs ha Þ þ ð1=ðLe 1ÞÞ 2500 ðxs xa Þ Fig. 3. Flowchart for calculating the optimal ratio between Gamb and Gs in the regeneration process.
ð5Þ
3.1. System definitions To analyze the thermodynamic performance of this system, several assumptions are made as follows:
ð6Þ
dX s ¼ Ga dxa X s =ðGs Ga dxa Þ
ð7Þ
where, NTU is number of mass transfer unit, and Le is Lewis number. dNTU and Le are defined as Eqs. (8) and (9), respectively.
(1) The evaporating temperature of the refrigerant is set to 5 °C, and the refrigerant superheat degree at the outlet of evaporator is set at 5 °C. The process taking place in the compressor is isentropic. LiCl–H2O is used as the liquid desiccant. (2) The refrigerant temperature at the outlet of solution-torefrigerant heat exchanger is assumed to be equal to the condensing temperature. (3) The refrigerant temperature leaving condenser is assumed to 40 °C in all operating conditions. What’s more, the outlet temperature of ambient air heated by condenser is set to be 5 °C lower than the condensing temperature. (4) The power consumption including pumps and fans is 10% of that consumed by the compressor in the hybrid systems, and 5% in the baseline system [29,30]. In addition, there is no energy or pressure loss during transportation.
(a)
dT s ¼ ðCps T s Ga dxa Ga dha Þ=ðGs CPs Þ
dNTU ¼
Le ¼
hD aw L dx dy Ga
hc hD Cpa
ð9Þ
where, hC and hD represent the heat transfer coefficient and mass transfer coefficient, respectively. dx is the micro-width along x-axis, dy is the micro-height of y-axis and L is the length of regenerator. aw is defined as the ratio between mass transfer area to volume of the regenerator. The outlet temperature of diluted solution (T13) and concentrated solution (T16) in solution-to-solution heat exchanger can be calculated by Eqs. (10) and (11), respectively.
T 13 ¼ T 12 þ eSHE ðT 15 T 12 Þ
(b)
Solution
Gs
Ta H
Air
Ts
Ga Ta +dTa
dy
ωa+ dωa
ωa
L
ð10Þ
Xs
dx
Ga
x y
ð8Þ
Gs
Ts +dTs
Xs+dXs
W Fig. 4. (a) Sketch of cross flow regeneration and (b) control volume.
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T 16 ¼ T 15 eSHE ðT 15 T 12 Þ
ð11Þ
where, eSHE is the effectiveness of solution-to-solution heat exchanger. T15 is the temperature of concentrated solution leaving regenerator, and T12 is the diluted solution outlet temperature of dehumidifier. The dehumidifier is internally cooled by cooling water, and the temperature of solution and air in the dehumidifier is kept stable. When solution inlet conditions are certain, the air outlet humidity ratio (x9) can be obtained, which is in equilibrium with the solution inlet humidity ratio above surface (x17).
x17 ¼ 0:622 Ps ðX 17 ; T 17 Þ=ð101325 Ps ðX 17 ; T 17 ÞÞ
ð12Þ
where, Ps(X17,T17) is the equilibrium vapor pressure above solution, referred to the paper [32]. The closed air temperature of T8 and T11 in the air-to-air heat exchanger can be calculated by Eqs. (13) and (14), respectively.
T 8 ¼ T 7 þ eAHE ðT 10 T 7 Þ
ð13Þ
T 11 ¼ T 10 þ eAHE ðT 10 T 7 Þ
ð14Þ
where, eAHE is the effectiveness of air-to-air heat exchanger. T7 and T8 are the inlet and outlet temperature of saturated closed air, while T10 and T11 are the inlet and outlet temperature of dry closed air leaving dehumidifier. In the model of indirect evaporative cooler, the closed air is believed to be fully humidified. In ideal case (i.e. eAHE is 1.0), the process taking place in the indirect evaporative cooler is isothermal. The closed air mass flow rate (Ga) can be obtained through mass conservation, as shown in Eq. (15).
Ga ¼ Gamb ðx20 x19 Þ=ðx7 x11 Þ
ð15Þ
The refrigerant outlet enthalpy (h5) in the indirect evaporative cooler is calculated by Eq. (16), based on the energy conservation.
h5 ¼ h4 Ga ðh7 h11 Þ=Gref
be verified through the experimental data in the literature [33,34] where the differences between the inlet and outlet air temperature are within 10%, and there is little fluctuation of the spray water temperature along the height of IEC. So the fluctuation of the spray water enthalpy can be neglected, which is often taken into account in previous models of IEC [35]. Based on the two assumptions, the evaporation latent heat of spray water only comes from the refrigerant sensible heat, since the refrigerant temperature is much higher than the air temperature in the IEC, which leads to the decrease of refrigerant temperature. Therefore, the ideal outlet temperature of refrigerant should be equal to the air temperature. What’s more, the air can be fully humidified only if the physical scale of the IEC is large enough to make the air and spray water contact fully. Hence the model of IEC is reliable and reasonable. The dehumidifier is internally cooled by cooling water and is simulated based on the effectiveness method [36] and the isothermal method [37]. A comparison has been made between the experimental data and the simulation results in the literature [37,38]. It can be found that the model is acceptable. For counter flow in a packed bed, the maximum achievable difference in air humidity is obtained when the outlet air is in equilibrium with the inlet desiccant solution, as long as the physical scale of the dehumidifier is large enough and the cooling water is sufficient. To see the discrepancy between theoretical value and actual value, effectiveness of heat exchangers is reconsidered to calculate the performance of the system, where the effectiveness of airto-air heat exchanger eAHE is chosen as 0.8, and the effectiveness of solution-to-solution heat exchanger eSHE is set to 0.85. The calculation results and discrepancy are shown in Table 2. As we can see, the maximum discrepancy of COP is 3.0%, and the maximum discrepancy of q0 is 2.94%. The maximum discrepancy of DTsc is 13.3%. The minimum discrepancy of the three performance indexes is 0%. Therefore, the assumption of the effectiveness for the heat exchangers to be 1.0 is acceptable.
ð16Þ
When h5 is certain, the refrigerant outlet temperature (T5) will be obtained. If the heat and mass transfer process in the indirect evaporative cooler is fully developed, T5 should be equal to the closed air outlet temperature (T7).
3.4. Performance indexes In this study, three performance indexes are used to investigate the influences of the studied parameters on the proposed energyefficient refrigeration system. These indexes are as follows:
3.3. Model validation
(1) Subcooling degree (DTsc), which is defined as the temperature difference between condensing temperature and the refrigerant temperature leaving indirect evaporative cooler.
Modeling the regeneration processes in the liquid desiccant cycle is critical for the performance simulation of the whole system. NTU-Le model and finite element method have been validated satisfactorily by the experimental data in the literature [31]. Hence, the model of regenerator is reliable enough to be used in the performance evaluation of the whole system. When eAHE is set to 1.0, the closed air inlet and outlet temperature of indirect evaporative cooler (IEC) are equal, so the closed air experiences an isothermal process in the ideal situation. This can
DT sc ¼ T con T 5
ð17Þ
(2) Specific refrigerating capacity (q0), which is calculated from the enthalpy difference between outlet and inlet refrigerant of the evaporator.
q0 ¼ h1 h6
ð18Þ
Table 2 Discrepancy between the theoretical value and actual value at different operation conditions. Tcon (°C)
Tamb (°C)
RHamb (%)
Xs (%)
COP Theoretical value
Actual value
Discrepancy (%)
q0 (kJ/kg) Theoretical value
Actual value
Discrepancy (%)
DTsc (°C) Theoretical value
Actual value
Discrepancy (%)
50 52.5 55 50 50 50 50 50
35 35 35 31 35 35 35 35
60 60 60 60 30 40 50 60
33 33 33 33 33 33 33 31
5.61 5.82 5.65 6.11 6.19 6.13 6.05 5.94
5.58 5.65 5.58 5.96 6.19 6.11 5.95 5.85
0.54 3.00 1.25 2.52 0 0.33 1.68 1.54
183.78 196.9 196.90 200 203 200.80 198.25 194.50
182.01 191.27 194.30 195.27 203 200.30 195.08 191.76
0.97 2.94 1.34 2.40 0 0.25 1.62 1.43
26.1 39.2 41.7 39.3 41.7 39.9 37.8 34.7
24.7 34.6 39.6 35.4 41.7 39.5 35.2 32.5
5.6 13.3 5.3 11 0 1 7.4 6.7
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(3) Coefficient of performance (COP), which is defined as the ratio between the specific refrigerating capacity and the total specific power consumption including compressor, pumps and fans in the energy-efficient refrigeration system.
COP ¼ q0 =ð1:1 wcom Þ
dehumidification and evaporation is shown in Fig. 5. For a given operating condition, the solution state and air state entering regenerator are determined first based on the solution-to-refrigeration heat exchanger model and the condenser model. Through regenerator model and iterative computations, the maximum regeneration mass flow of moisture and the outlet parameters of solution can be achieved. Since the closed air outlet humidity ratio of dehumidifier is same as that of concentrated solution entering dehumidifier, the closed air humidity ratio leaving dehumidifier can also be obtained. By assumption of closed air temperature leaving the indirect evaporative cooler, the subcooled refrigerating capacity and cooling load of indirect evaporative cooler are achieved based on the air-to-air heat exchanger model and indirect evaporative
ð19Þ
where, wcom is the specific power consumption of the compressor (kJ kg1). 3.5. Flowchart of the energy-efficient refrigeration system The flowchart for calculating the performance of the energyefficient refrigeration system subcooled by liquid desiccant
Refrigerant state T2,Tcon
Assume solution outlet temperature of regenerator T15
Model of solution-torefrigerant heat exchanger
Model of solution-tosolution heat exchanger
Solution state T14
Solution state T13
Solution mass flow rate Gs Air inlet parameters of regenerator T19, ω 19
Model of condenser Solution inlet parameters of regenerator T14, Xs
Choose optimal ratio Gamb /Gs
Model of regenerator
Maximum regeneration mass flow: Greg,max Outlet parameters of solution: X15,Ts,reg,out Assume closed air outlet temperature of indirect evaporative cooler (IEC) T7
NO
T15=Ts,reg,out YES
Model of indirect evaporative cooler
Model of dehumidifier
Closed air humidity ratio: ω 7
Closed air humidity ratio:ω 9
Mass conservation GIEC =Greg,max
Closed air mass flow rate:Ga
Closed air parameters T7,ω 7,T11,ω 11
Outlet temperature of refrigerant: T5 Subcooled refrigeration capacity: Qref,sc
NO
291
Cooling load of IEC: QIEC
Qref,sc=QIEC YES Calculate performance indexes Δ Tsc, q0, COP Fig. 5. Procedure for calculating the system performance.
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Table 3 The reference values and variation ranges of the studied parameters.
Tamb RHamb (%) Teva 4Tsh Tcon Xs (%)
Reference value
Minimum value
Maximum value
Increment
Unit
35 60 5 5 50 33
31 30 – – 45 31
39 70 – – 55 35
2 10 – – 2.5 1
°C – °C °C °C –
cooler model. Through iterative computations, the performance indexes can be obtained. Table 4 Description of different systems involved in the energy-efficient refrigeration system.
4. Results and discussion The effects of key parameters on performance of the proposed system are presented in this section. The main parameters of the system are as follows: condensing temperature (Tcon), ambient air temperature (Tamb), ambient air relative humidity (RHamb) and liquid desiccant concentration in regeneration (Xs). In each case, only one parameter is changed, while other parameters are kept constant at the reference values. The reference value for each parameter and the variation range are shown in Table 3. To make comparative studies, four different systems are employed, as listed in Table 4. 4.1. Effect of the condensing temperature on the system performance Theoretical analysis results of the proposed system at five different condensing temperatures are shown in Table 5. As we can see, with the increase of condensing temperature, refrigerant outlet temperature of IEC (T5) decreases at first and then remains stable at 13.3 °C, which is restrained by the solution concentration at the outlet of the dehumidifier and cooling water temperature in the dehumidifier. Fig. 6 shows the effects of condensing temperature Tcon on the subcooling degree of the liquid refrigerant (DTsc), the specific refrigerating capacity (q0) and the COP. When the condensing temperature (Tcon) rises from 45 °C to 55 °C, the hybrid systems (system 1 and system 2) could achieve more subcooling degree of the liquid refrigerant (DTsc), compared with the baseline refrigeration system (system 3), as shown in Fig. 6(a). The results also indicate that DTsc of system 1 is significantly higher than that of system 2 when Tcon is higher than 50 °C. The higher the condensing temperature (Tcon), the higher the solution inlet temperature of regenerator (Ts,in), which would provide more heat for the desiccant cycle and enhance the desiccant regeneration. For system 1, when condensing temperature (Tcon) is beyond 50 °C, the air inlet temperature of regenerator (Tamb,reg,in) is higher than minimum regeneration temperature (Ts,min) and the hot air from condenser can improve the regeneration effect significantly. Therefore, there is a considerable increase in DTsc, whereas when Tcon is above 52.5 °C, q0 increases generally, which is restrained by the dew point temperature of closed air inlet state of indirect evaporative cooler.
Systems
Description
System 1
The energy-efficient refrigeration system which uses hot air from the condenser in the desiccant regeneration The energy-efficient refrigeration system which uses ambient air in the desiccant regeneration Baseline refrigeration system without the desiccant cycle which keeps the refrigerant temperature leaving condenser at 40 °C Reverse Carnot cycle
System 2 System 3
System 4
Fig. 6(b) shows that specific refrigerating capacity (q0) of the hybrid systems is much higher than that of baseline refrigeration system. With the increase of Tcon, q0 of system 2 increases gradually from 178.7 kJ/kg to 185.7 kJ/kg, while q0 of system 1 rises generally at first, increases significantly as Tcon rises from 50 °C to 52.5 °C, and remains stable when Tcon is higher than 52.5 °C. All these changes result from the similar variation of DTsc. Fig. 6(c) shows the effect of condensing temperature on the COPs of four different systems. As shown in Fig. 6(c), with the increase of Tcon, the COPs of system 3 and system 4 decrease linearly due to the fact that higher condensing temperature would result in larger increase of compressor power consumption wcom. However, the situation for system 1 is not different since it drops slowly when Tcon is below 50 °C, increases generally when Tcon rises from 50 °C to 52.5 °C and then decreases slightly as Tcon rises above 52.5 °C. It can be observed that the COPs of the hybrid systems are higher than that of the baseline conventional vapor compression system. When Tcon is above 52.5 °C, system 1 shows higher COP than the Carnot cycle (system 4), since the condensing heat utilization through desiccant cooling cycle enhances the vapor compression system. There is an apparent difference between system 1 and system 2. For system 1, it seems that there exists a maximum COP while Tcon is around 52.5 °C.
4.2. Effect of the liquid desiccant concentration on the system performance Theoretical analysis results of the proposed system at five different solution concentrations are shown in Table 6. As we can see, with the increase of solution concentration, refrigerant outlet
Table 5 Theoretical analysis results of the proposed system at different condensing temperature. Tcon (°C)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
x19
T20 (°C)
(g/kg)
x20
45 47.5 50 52.5 55
27.0 25.4 23.9 13.3 13.3
23 20.9 19 9.7 9.7
9.0 8.98 8.9 8.4 8.4
56.3 59.2 62.2 65.1 68
33 33 33 33 33
47 47.9 48.5 46.6 47.2
33.2 33.3 33.4 34.1 34.2
40 42.5 45 47.5 50
21.8 21.8 21.8 21.8 21.8
56 59 61.8 46.8 49.1
38.1 44.6 51.5 23.2 23.4
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(b)
System 1
System 1 System 2
System 2 System 3
System 3
q0 (kJ/kg)
Tsc (oC)
(a)
Tcon (oC)
Tcon (oC)
(c)
System 1 System 2
COP
System 3 System 4
Tcon (oC) Fig. 6. Effect of condensing temperature (Tcon) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
temperature of IEC (T5) decreases at first and then increases. There exists a minimum value when Xs is 0.32. Fig. 7 shows the effects of the desiccant concentration Xs on the system performance indexes. As shown in Fig. 7(a), in general, DTsc of the hybrid system, including system 1 and system 2, is at least 137% higher than that of the baseline system (system 3) in the variation range. DTsc of system 1 is higher than that of system 2, especially when Xs is below 0.33. It can be observed that DTsc of system 1 increases slightly from 34.7 °C to 35.1 °C when Xs increases from 0.31 to 0.32, and drops significantly as Xs is beyond 0.33, however DTsc of system 2 decreases slightly always with the increase of Xs. The reasons can be explained as follows: the higher the Xs, the lower the desiccant regeneration thermal efficiency when Ts,in remains stable. Consequently the desiccant cycle would provide less evaporative cooling capacity and the relatively less DTsc in system 1 and system 2. Fig. 7(b) shows q0 of system 1 increases slightly when Xs increases from 0.31 to 0.32, and drops significantly from 194.96 kJ/ kg to 183.78 kJ/kg as Xs increases from 0.32 to 0.33, however q0 of system 2 decreases only by 1.47% in the variation range and is lower than that of system 1. All the changes are due to the similar changes of DTsc with the increase of Xs when Tcon remains stable. As shown in Fig. 7(c), COPs of the hybrid systems (system 1 and system 2) are much higher than that of the baseline system (system 3), which indicates that the proposed system is quite an effective way to improve the performance of conventional vapor compression refrigeration system. The COP of system 1 increases slightly with lower desiccant concentration, and then drops
significantly when the Xs is more than 0.32. Whereas, COP of system 2 is lower than that of system 1 and decreases slightly from 5.59 to 5.52 as Xs increases from 0.31 to 0.35, which results from the corresponding changes of q0 with the increase of Xs when wcom is kept constant. Suitable concentration of the desiccant solution should be considered carefully for higher performance of the hybrid system. In the typical case, the suggested concentration of the LiCl–H2O solution is around 0.32 for the hybrid vapor compression system. 4.3. Effect of the ambient air temperature on the system performance Theoretical analysis results of the proposed system at five different ambient air temperatures are shown in Table 7. As we can see, with the increase of ambient air temperature, refrigerant outlet temperature of IEC (T5) increases. This is due to the fact that higher ambient air temperature leads to the worse regeneration efficiency and consequently the lower cooling capacity in the IEC. The parameters of the ambient air would impact the desiccant cooling system evidently. In the analysis, the temperature of the air varies from 31 °C to 39 °C while the relative humidity of the air is kept constant at 60%. Fig. 8 shows the effects of Tamb on the performance indexes. The hybrid systems (system 1 and system 2) achieve much more subcooling degree (DTsc) than the baseline system (system 3). As Tamb increases, DTsc of the hybrid system decreases, as shown in Fig. 8(a). It is clear that DTsc of system 1 is significantly influenced, while DTsc of system 2 decreases slightly in the studied range. Due to the same relative humidity of the air in
Table 6 Theoretical analysis results of the proposed system at different solution concentration. Xs (%)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
x19
T20 (°C)
(g/kg)
x20
31 32 33 34 35
15.3 14.9 23.9 24.6 24.9
11 10.7 19 19.9 20.3
9.9 9.3 8.9 8.2 7.7
62.2 62.2 62.2 62.2 62.2
31 32 33 34 35
43.9 44.7 48.5 50 50.6
32.0 32.8 33.4 34.4 35.3
45 45 45 45 45
21.8 21.8 21.8 21.8 21.8
44.5 45 61.8 62 62
22.7 22.7 51.5 48.5 45
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(a)
(b)
System 1 System 2 System 3
q0 (kJ/kg)
Tsc (o C)
System 1 System 2 System 3
Xs
Xs
(c)
System 1 System 2 System 3
COP
System 4
Xs Fig. 7. Effect of liquid desiccant concentration (Xs) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
Table 7 Theoretical analysis results of the proposed system at different ambient air temperature. Tamb (°C)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
T20 (°C)
(g/kg)
31 33 35 37 39
10.7 12 23.9 24.2 24.9
8.1 8.9 19 19.4 20.3
6.8 7.7 8.9 9.9 10.8
62.2 62.2 62.2 62.2 62.2
33 33 33 33 33
43.1 44.3 48.5 50.1 51
34.3 34 33.4 33.3 33.3
45 45 45 45 45
17.3 19.4 21.8 24.4 27.3
43.7 44.5 61.8 62 62
18.9 20.6 51.5 52 52.3
the analysis, the air humidity ratio increases significantly with the increase of Tamb, which would reduce the regeneration effect. For the system 1, the general decrease of DTsc is due to the fact that higher air temperature determines higher dehumidification temperature, which results in the increase of dew point temperature of closed air entering the indirect evaporative cooler. Fig. 8(b) shows that q0 of system 1 decreases generally from 200 kJ/kg to 183.78 kJ/kg as Tamb increases from 31 °C to 35 °C, while q0 of system 2 decreases slightly in the variation range, which is due to the corresponding changes of DTsc when Tcon is kept constant. Fig. 8(c) shows the comparisons of COP of different systems involved in the energy-efficient refrigeration system. It can be observed that COP of the hybrid systems (system 1 and system 2) is at least 6.14% higher than that of the baseline system (system 3). The COP of system 1 is higher than that of system 2, especially when Tamb is lower than 35 °C. 4.4. Effect of the ambient air relative humidity on the system performance Theoretical analysis results of the proposed system at different ambient air relative humidity are shown in Table 8. As we can see, with the increase of ambient air relative humidity, refrigerant outlet temperature of IEC (T5) increases. In this analysis, the ambient air temperature is constant at 35 °C. Fig. 9 shows the effects of the ambient air relative humidity
x19
x20
RHamb on subcooling degree (DTsc), specific refrigerating capacity (q0) and the COP. DTsc of the hybrid systems (system 1 and system 2) is much higher than that of baseline system (system 3). As shown in Fig. 9(a), DTsc of system 1 decreases gradually from 41.7 °C to 37.8 °C as RHamb increases from 30% to 50%, drops significantly by 30.9%, and decreases slightly when RHamb increases beyond 0.6, whereas DTsc of system 2 drops considerably by 31.5% as RHamb increases from 30% to 40%. The ambient air humidity ratio increases with the increase of RHamb and constant air temperature, which reduces its potential for the moisture transfer from solution to air in the desiccant regeneration, and consequently both curves of system 1 and system 2 decrease with the increase of RHamb. Fig. 9(b) shows the similar trends of q0 and DTsc. For system 1, q0 decreases gradually when RHamb is below 50% and reduces significantly from 198.25 kJ/kg to 183.78 kJ/kg. In terms of system 2, q0 decreases considerably from 198.6 kJ/kg to 183.78 kJ/kg when RHamb increases from 30% to 40%. All the changes of q0 result from the corresponding changes of DTsc with the increase of RHamb, when Tcon is kept constant. Fig. 9(c) shows the comparisons of COP. It is observed that COPs of system 1 and system 2 are at least 6% higher than that of the conventional vapor compression system (system 3) with different relative humidity of the ambient air. COP of system 1 decreases from 6.19 to 6.05 linearly when RHamb rises from 30% to 50%. Moreover, COP of system 1 is even higher than that of the ideal Carnot cycle (system 4) when RHamb is below 40%.
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(a)
(b)
System 1 System 2 System 3
q0 (kJ/kg)
Tsc (oC)
System 1 System 2 System 3
Tamb (oC)
Tamb (oC)
(c)
System 1 System 2
COP
System 3 System 4
Tamb (oC) Fig. 8. Effect of ambient air temperature (Tamb) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
Table 8 Theoretical analysis results of the proposed system at different ambient air relative humidity. RHamb (%)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
T20 (°C)
(g/kg)
30 40 50 60 70
8.3 10.1 12.2 23.9 24.6
6.9 7.8 9 19 19.9
5.5 6.5 7.7 8.9 10.7
62.2 62.2 62.2 62.2 62.2
33 33 33 33 33
38.5 41.2 43.6 48.5 50.9
34.7 34.4 34.1 33.4 33.3
45 45 45 45 45
10.7 14.3 18 21.8 25.6
43.8 44 44.2 61.8 62
11.9 15.6 19.2 51.5 52.3
Tsc (oC)
System 1 System 2 System 3
(b)
RHamb
x20
System 1 System 2 System 3
q0 (kJ/kg)
(a)
x19
RHamb
(c)
System 1 System 2
COP
System 3 System 4
RH amb Fig. 9. Effect of ambient air relative humidity (RHamb) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
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5. Conclusion This paper presents the thermodynamic analysis of the energyefficient refrigeration system subcooled by the liquid desiccant dehumidification and indirect evaporation. The indirect evaporative cooler with dry air leaving from dehumidifier is employed to subcool the liquid refrigerant at the outlet of condenser, while the compensation heat of the desiccant regeneration is provided by the condensation heat. Main conclusions can be drawn as follows: (1) Generally, the proposed hybrid vapor compression refrigeration system shows significantly higher COP than conventional vapor compression refrigeration system with the maximum COPs using hot air and ambient air increasing by 18.8% and 16.3% respectively, and even higher than the reverse Carnot cycle at the same operation conditions, which indicates that the subcooling method using liquid desiccant and evaporation is quite an efficient way to improve the performance of conventional refrigeration system. (2) Effects of important operation parameters on the hybrid vapor compression refrigeration system are disclosed. The system would show more obvious potential compared with traditional vapor compression refrigeration system at higher condensing temperature. Suitable concentration of the desiccant solution in the desiccant cycle should be considered carefully for higher performance of the hybrid system. In regard to LiCl–H2O as the desiccant, the suggested mass concentration is around 0.32. (3) Theoretical calculations using hot air from condenser and ambient air for desiccant regeneration indicate that the performance of system 1 is higher than that of system 2, especially when the air inlet temperature of regenerator is higher than the minimum regeneration temperature. The maximum difference percentage between system 1 and system 2 is 9.5% under studied conditions. Since the theoretical analysis indicates that the new subcooling method for vapor compression refrigeration systems is very effective, an experimental setup will be established in an environment chamber with regulable air temperature and humidity, combining the liquid desiccant system with the indirect evaporative cooling system available at the moment. Acknowledgements This work was partially supported by Natural Science Foundation of China (No. 51006022) and the 12th Five Year Science and Technology Support Key Project of China (No. 2011BAJ03B14). References [1] Perez-Lombard L, Ortiz J, Pout C. A review on buildings energy consumption information. Energy Buildings 2008;40:394–8. [2] Thu HTM, Sato H. Proposal of an eco-friendly high-performance airconditioning system part 2. Application of evapo-transpiration condenser to residential air-conditioning system. Int J Refrig 2013;36:1596–601. [3] Youbi-Idrissi M, Macchi-Tejeda H, Fournaison L, Guilpart J. Numerical model of sprayed air cooled condenser coupled to refrigerating system. Energy Convers Manage 2007;48(7):1943–51. [4] Gong GC, Chen FH, Su H, Zhou JY. Thermodynamic simulation of condensation heat recovery characteristics of a single stage centrifugal chiller in a hotel. Appl Energy 2012;91:326–33. [5] Yi XW, Lee WL. The use of helical heat exchanger for heat recovery domestic water-cooled air-conditioners. Energy Convers Manage 2009;50(2):240–6. [6] Jiang ML, Wu JY, Wang RZ. Research on refrigerant flow characteristics and performance of a multi-functional heat pump system. Energy Convers Manage 2011;52(6):2323–8.
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Energetic evaluation of an internal heat exchanger in a CO2 transcritical refrigeration plant using experimental data. Int J Refrig 2011;34(1):40–9. [14] Khan JR, Zubair SM. Design and rating of an integrated mechanical-subcooling vapor-compression refrigeration system. Energy Convers Manage 2000;41:1201–22. [15] Qureshi BA, Zubair SM. The effect of refrigerant combinations on performance of a vapor compression refrigeration system with dedicated mechanical subcooling. Int J Refrig 2012;35:47–57. [16] Qureshi BA, Inam M, Antar MA, Zubair SM. Experimental energetic analysis of a vapor compression refrigeration system with dedicated mechanical subcooling. Appl Energy 2013;102:1035–41. [17] Qureshi BA, Zubair SM. Cost optimization of heat exchanger inventory for mechanical subcooling refrigeration cycles. Int J Refrig 2013;36:1243–53. [18] Liu ZL, Wang ZY, Ma CF. An experimental study on heat characteristics of heat pipe heat exchanger with latent heat storage. 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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Thermodynamic analysis of a novel energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation Xiaohui She, Yonggao Yin ⇑, Xiaosong Zhang School of Energy and Environment, Southeast University, Nanjing 210096, China Ministry of Education of Key Laboratory of Energy Thermal Conversion and Control, Southeast University, Nanjing 210096, China
a r t i c l e
i n f o
Article history: Received 20 July 2013 Accepted 24 October 2013 Available online 21 November 2013 Keywords: Refrigeration Liquid desiccant Subcooling Condensation heat utilization Energy efficiency improvement
a b s t r a c t A new energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation was proposed in this paper. In the system, liquid desiccant system could produce very dry air for an indirect evaporative cooler, which would subcool the vapor compression refrigeration system to get higher COP than conventional refrigeration system. The desiccant cooling system can use the condensation heat for the desiccant regeneration. Thermodynamic analysis is made to discuss the effects of operation parameters (condensing temperature, liquid desiccant concentration, ambient air temperature and relative humidity) on the system performance. Results show that the proposed hybrid vapor compression refrigeration system achieves significantly higher COP than conventional vapor compression refrigeration system, and even higher than the reverse Carnot cycle at the same operation conditions. The maximum COPs of the hybrid systems using hot air and ambient air are 18.8% and 16.3% higher than that of the conventional vapor compression refrigeration system under varied conditions, respectively. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent years, refrigeration systems consumed a large amount of energy in maintaining thermal comfort for occupants and suitable climatic conditions for cooling cases, which made up 50% of building energy consumption [1]. The utilization of evaporative condensing indicated energy efficient potential in reduction of power requirements [2,3]. In a traditional refrigeration system, a great deal of condensation heat, which could be used for other purposes, is directly dissipated to the environment. The dissipated heat not only wastes energy, but also causes severe heat island effect in the surrounding areas. Many methods have been attempted to tackle these problems. Some researchers utilized the condensation heat from air conditioners to preheat domestic hot water [4–7], leading to claims that water heating in summer, primarily for bathing. It can be made available virtually free whenever space cooling is required, and is considered one of the most cost effective energy conservation measures. Several other researchers tried to add a heat recovery system on the refrigeration system. For instance, a condensing heat recovery with thermal storage of phase change material (paraffin wax as PCM) was designed and analyzed by Zhang et al. [8]. In addition, Kaushik et al. [9] introduced a canopus heat exchanger ⇑ Corresponding author at: School of Energy and Environment, Southeast University, Nanjing 210096, China. Tel.: +86 25 83792722. E-mail address: [email protected] (Y. Yin). 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.10.057
for heat recovery for a refrigeration system. A drawback with heat recovery system is the high condensing temperature that increases the energy consumption of the refrigeration system. To solve this problem, Arias and Lundqvist [10] proposed an alternative to heat recovery, namely floating condensing pressure, which improved the coefficient of performance and decreased the energy consumption of the refrigeration system at lower outdoor temperature. From thermodynamic standpoint, further cooling of liquid refrigerant leaving condenser can significantly increase the cooling capacity and improve the system COP. The regenerative refrigeration cycle adopted a liquid-suction heat exchanger as a sub-cooler to improve the system coefficient of performance (COP) [11–13]. The liquid-suction heat exchanger achieves condensate subcooling by transferring the refrigerant heat from the condenser outlet to the compressor inlet. This subcooling method may result in the superheated degree existing at the compressor inlet and reduce the system efficiency. Therefore the liquid-suction heat exchanger is not suitable for all air conditioning systems. In order to eliminate the drawbacks of the liquid-suction heat exchanger, Khan et al. [14] and Qureshi et al. [15,16] used integrated and dedicated mechanical-subcooling methods to enhance the system COP and also to remove the superheated degree, respectively. Thermoeconomic considerations were given to heat exchanger inventory allocation in vapor compression cycles with mechanical subcooling by Qureshi and Zubair [17], and it was concluded that the cost optimization of the integrated mechanical subcooling system was qualitatively the same as the dedicated subcooling system. Since the
X. She et al. / Energy Conversion and Management 78 (2014) 286–296
287
Nomenclature COP Cp d G h H hC hD L Le m NTU P Q q0 RH T DTsc DTsh w W X
coefficient of performance specific heat capacity (kJ kg1 K1) differential mass flow rate (kg s1) enthalpy (kJ kg1) height (m) heat transfer coefficient (W m2 K1) mass transfer coefficient (kg m2 s1) length (m) Lewis number adiabatic index of refrigerant number of mass transfer unit pressure (Pa) refrigerating capacity (kW) specific refrigerating capacity (kJ kg1) relative humidity (%) temperature (°C) subcooling degree (°C) superheat degree (°C) specific power consumption (kJ kg1) width (m) Concentration (%)
mechanical-subcooling required an additional compressor to provide the driving force of subcooling, the electricity consumption of the compressor would carry the undesirable effect on the system COP. Besides, the expensive initial cost is also one of the drawbacks in the mechanical-subcooling method. To avoid the demerits of mechanical-subcooling method, many studies utilized phase change material and heat pipes in cold storage unit as a subcooler [18–20]. Chieh et al. [21] developed a thermal battery composed of the phase change material and the heat pipes to store energy in air conditioning application. Huang et al. [22] experimentally investigated the performance of the cold storage air conditioning system utilizing a thermal battery as a subcooler. Because the charge heat exchanger of the thermal battery had the larger thermal resistances, it induced the lower ice stored ability. For eliminating the disadvantages discussed in previous studies, Hsiao et al. [23,24] used an ice storage tank as a subcooler to enhance the system capabilities and reduces the superheated degree without inputting the supplemental electricity. By the innovative design of the charge heat exchanger, the overall thermal resistances can be decreased. The liquid desiccant system using low-grade heat resource was proposed by Lof [25], and its application in air-conditioning systems has been widely investigated [26–28]. In this paper, a novel energy-efficient vapor compression refrigeration system combined with a liquid desiccant cooling system is proposed. The vapor compression system is subcooled by the liquid desiccant cooling cycle driven by condensing heat to achieve more subcooling degree of the refrigerant than the conventional system. This paper will discuss the potential of the performance improvement and the effect of different climatic and operating conditions on the performance of the system, and a comparative study was made on the effect of different utilization ways of condensation heat.
2. System description Fig. 1 illustrates the novel refrigeration cycle subcooled by the liquid desiccant dehumidification and evaporation, which is
Greek symbols e effectiveness x humidity ratio (kg kg1) Subscripts a air AHE air-to-air heat exchanger amb ambient com compressor con condenser eva evaporator IEC indirect evaporative cooler in inlet out outlet max maximum min minimum ref refrigerant reg regenerator s solution sc subcooling sh superheat SHE solution-to-solution heat exchanger
composed of refrigeration cycle, closed air cycle and liquid desiccant cycle. The refrigeration cycle includes an evaporator, a compressor, a solution-to-refrigerant heat exchanger and a condenser. Refrigerant R-22 is chosen to be the working fluid inside the refrigeration cycle, and the refrigerating capacity of the baseline system is 30 kW. The closed air cycle consists of an indirect evaporative cooler, an air-to-air heat exchanger, a dehumidifier and an air cooler, while the liquid desiccant cycle is made up of a dehumidifier, a solution cooler, a solution-to-solution heat exchanger, a regenerator and a solution-to-refrigerant heat exchanger. The dehumidifier is internally cooled by cooling water and its physical size is assumed large enough to make the air outlet humidity ratio be in equilibrium with the solution inlet humidity ratio above the surface, which can be achieved as long as the cooling water is sufficient. The regenerator is adiabatic, and its physical scale at several different operating conditions is shown in Table 1. The physical scale of other components of the system is believed large enough to meet our needs. Fig. 2(a) shows the ideal pressure enthalpy diagram of refrigeration cycle. As indicated in Fig. 2(a), the refrigerant at state 4 is subcooled to state 5 by the indirect evaporative cooler, using the dry closed air absorbing moisture from water. A conceptual schematic diagram for the processes of the liquid desiccant cycle and closed air cycle is shown in Fig. 2(b). There is sensible heat recovery of the closed air from state 7 to state 8. The moist air (state 8) is dehumidified in the dehumidifier to go to state 9. The dehumidified air is precooled by an air cooler and then cooled to state 11 by an air-to-air heat exchanger. In indirect evaporative cooler, the cold dry air is humidified by water evaporation and reaches state 7. In the liquid desiccant cycle, state 12 and state 14 are in the isoconcentration line of liquid desiccant solution. The diluted solution is preheated by hot concentrated solution leaving the regenerator in solution-to-solution heat exchanger. Prior to entering regenerator, the liquid desiccant is heated to the specified set point temperature (state 14) in the solution-to-refrigerant heat exchanger. Then the hot diluted solution enters into the regenerator and finally reaches state 15. State 15 and state 17 are also in iso-concentration
288
X. She et al. / Energy Conversion and Management 78 (2014) 286–296 Solution cooler
17
Air cooler
9
16
Solution-tosolution heat exchanger
20
15
Dehumidifier
10
Regenerator 13
12
8
19
14
Air-to-air heat exchanger 4
7
11
3 Solution-to-refrigerant heat exchanger
Condenser
2
18
Indirect evaporative cooler
Compressor
Fan
5 Pump
Throttle valve Evaporator
6
1
Fig. 1. Schematic diagram of proposed energy-efficient refrigeration system.
Table 1 The physical scale of regenerator at several different operation conditions. Tcon (°C)
Tamb (°C)
RHamb (%)
Xs (%)
H (m)
W (m)
L (m)
50 50 50 50 50 52.5 55
31 35 35 35 35 35 35
60 30 40 50 60 60 60
33 33 33 33 31 33 33
0.65 0.34 0.43 0.64 0.57 1.02 0.46
0.65 0.34 0.43 0.64 0.57 1.02 0.46
0.65 0.34 0.43 0.64 0.57 1.02 0.46
line of liquid desiccant solution. The hot concentrated solution leaving regenerator is precooled by solution-to-solution heat exchanger, and it is cooled to state 17 by solution cooler. The concentrated solution goes to the dehumidifier and reaches state 12. Effective use of the condensation heat is quite vital to the performance of the desiccant cycle. In the following research, condensation sensible heat is employed to heat the solution before it enters regenerator. Different air is used in the regeneration process to make a comparative study, namely hot air from condenser and
the ambient air. The solution mass flow rate (Gs) and air mass flow rate (Gamb) should be chosen reasonably in order to achieve the best regeneration effect. As far as one certain regeneration process (i.e. the inlet parameters are known) is concerned, there exists minimum regeneration temperature (Ts,min). If the air inlet temperature of regenerator (Tamb,reg,in) is much higher than Ts,min, then the more the Gamb, the better the regeneration effect. Whereas if Tamb,reg,in is below Ts,min, Gamb should be lower than Gs and be chosen properly, as shown in Fig. 3. This is due to the fact that if Tamb,reg,in is beyond Ts,min, then the air is able to restrain heat transfer from solution to air and improve the regeneration effect significantly. When Tamb,reg,in is lower than Ts,min, the solution will transfer heat to the air, and its temperature would decrease considerably, which will reduce the vapor pressure above solution and be not helpful for the regeneration process.
3. Modeling of energy-efficient refrigeration system In this section, a thermodynamic model of the energy-efficient refrigeration system is developed for the performance analysis.
P
(b)
Subcooled 5
6
4
14
2
3
15
Temperature (oC)
(a)
1
17 9
Xs
12 8
00%
RH=1
11
7
h
Humidity ratio (kg/kg) Fig. 2. Refrigeration cycle baseline and process air line in liquid desiccant cycle: (a) pressure-enthalpy diagram of the refrigeration cycle and (b) process air line in liquid desiccant cycle.
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3.2. Model description
Solution concentration Xs
Air inlet parameters of regenerator Tamb,reg,in, RHamb,reg,in
For a given refrigerating capacity of the baseline system, the mass flow rate of refrigerant (Gref) can be obtained. The solution temperature entering regenerator (T14) is calculated through energy conservation of the solution-to-refrigerant heat exchanger, as shown in Eq. (1).
Air humidity ratio
ωamb,reg,in
T 14 ¼ T 13 þ Gref ðh2 h3 Þ=ðGs Cps Þ
ð1Þ
The ambient air mass flow rate (Gamb) is calculated using Eq. (2). For different operating conditions, air mass flow rate in the regenerator should be chosen properly to achieve the best regeneration effect. For example, in some cases, all the Gamb should be used in the regenerator, while in other occasions, only part of Gamb is chosen.
Minimum regeneration temperature Ts,min =f (ωamb,reg,in,Xs)
Gamb ¼ Gref ðh3 h4 Þ=ðh19 h18 Þ Tamb,reg,in>Ts,min
Tamb,reg,in
Case 1: using hot air Choose all the heat from condenser and calculate Gamb Case 2: using ambient air Choose the optimal Gamb to achieve the best regeneration effect
Case 1: using hot air Choose proper ratio of the condensation latent heat and calculate Gamb Case 2: using ambient air Choose proper ratio (Gamb /Gs) to achieve the best regeneration effect
ð2Þ
In the regenerator, NTU-Le model and finite element method [31] are employed to analyze internal changes of regenerator, as shown in Fig. 4. The governing equations for heat and moisture conservation in the air and solution can be written as follows.
dxa ¼ dNTU ðxs xa Þ
ð3Þ
dT a ¼ dNTU Le ðT s T a Þ
ð4Þ
dha ¼ dNTU Le ½ðhs ha Þ þ ð1=ðLe 1ÞÞ 2500 ðxs xa Þ Fig. 3. Flowchart for calculating the optimal ratio between Gamb and Gs in the regeneration process.
ð5Þ
3.1. System definitions To analyze the thermodynamic performance of this system, several assumptions are made as follows:
ð6Þ
dX s ¼ Ga dxa X s =ðGs Ga dxa Þ
ð7Þ
where, NTU is number of mass transfer unit, and Le is Lewis number. dNTU and Le are defined as Eqs. (8) and (9), respectively.
(1) The evaporating temperature of the refrigerant is set to 5 °C, and the refrigerant superheat degree at the outlet of evaporator is set at 5 °C. The process taking place in the compressor is isentropic. LiCl–H2O is used as the liquid desiccant. (2) The refrigerant temperature at the outlet of solution-torefrigerant heat exchanger is assumed to be equal to the condensing temperature. (3) The refrigerant temperature leaving condenser is assumed to 40 °C in all operating conditions. What’s more, the outlet temperature of ambient air heated by condenser is set to be 5 °C lower than the condensing temperature. (4) The power consumption including pumps and fans is 10% of that consumed by the compressor in the hybrid systems, and 5% in the baseline system [29,30]. In addition, there is no energy or pressure loss during transportation.
(a)
dT s ¼ ðCps T s Ga dxa Ga dha Þ=ðGs CPs Þ
dNTU ¼
Le ¼
hD aw L dx dy Ga
hc hD Cpa
ð9Þ
where, hC and hD represent the heat transfer coefficient and mass transfer coefficient, respectively. dx is the micro-width along x-axis, dy is the micro-height of y-axis and L is the length of regenerator. aw is defined as the ratio between mass transfer area to volume of the regenerator. The outlet temperature of diluted solution (T13) and concentrated solution (T16) in solution-to-solution heat exchanger can be calculated by Eqs. (10) and (11), respectively.
T 13 ¼ T 12 þ eSHE ðT 15 T 12 Þ
(b)
Solution
Gs
Ta H
Air
Ts
Ga Ta +dTa
dy
ωa+ dωa
ωa
L
ð10Þ
Xs
dx
Ga
x y
ð8Þ
Gs
Ts +dTs
Xs+dXs
W Fig. 4. (a) Sketch of cross flow regeneration and (b) control volume.
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T 16 ¼ T 15 eSHE ðT 15 T 12 Þ
ð11Þ
where, eSHE is the effectiveness of solution-to-solution heat exchanger. T15 is the temperature of concentrated solution leaving regenerator, and T12 is the diluted solution outlet temperature of dehumidifier. The dehumidifier is internally cooled by cooling water, and the temperature of solution and air in the dehumidifier is kept stable. When solution inlet conditions are certain, the air outlet humidity ratio (x9) can be obtained, which is in equilibrium with the solution inlet humidity ratio above surface (x17).
x17 ¼ 0:622 Ps ðX 17 ; T 17 Þ=ð101325 Ps ðX 17 ; T 17 ÞÞ
ð12Þ
where, Ps(X17,T17) is the equilibrium vapor pressure above solution, referred to the paper [32]. The closed air temperature of T8 and T11 in the air-to-air heat exchanger can be calculated by Eqs. (13) and (14), respectively.
T 8 ¼ T 7 þ eAHE ðT 10 T 7 Þ
ð13Þ
T 11 ¼ T 10 þ eAHE ðT 10 T 7 Þ
ð14Þ
where, eAHE is the effectiveness of air-to-air heat exchanger. T7 and T8 are the inlet and outlet temperature of saturated closed air, while T10 and T11 are the inlet and outlet temperature of dry closed air leaving dehumidifier. In the model of indirect evaporative cooler, the closed air is believed to be fully humidified. In ideal case (i.e. eAHE is 1.0), the process taking place in the indirect evaporative cooler is isothermal. The closed air mass flow rate (Ga) can be obtained through mass conservation, as shown in Eq. (15).
Ga ¼ Gamb ðx20 x19 Þ=ðx7 x11 Þ
ð15Þ
The refrigerant outlet enthalpy (h5) in the indirect evaporative cooler is calculated by Eq. (16), based on the energy conservation.
h5 ¼ h4 Ga ðh7 h11 Þ=Gref
be verified through the experimental data in the literature [33,34] where the differences between the inlet and outlet air temperature are within 10%, and there is little fluctuation of the spray water temperature along the height of IEC. So the fluctuation of the spray water enthalpy can be neglected, which is often taken into account in previous models of IEC [35]. Based on the two assumptions, the evaporation latent heat of spray water only comes from the refrigerant sensible heat, since the refrigerant temperature is much higher than the air temperature in the IEC, which leads to the decrease of refrigerant temperature. Therefore, the ideal outlet temperature of refrigerant should be equal to the air temperature. What’s more, the air can be fully humidified only if the physical scale of the IEC is large enough to make the air and spray water contact fully. Hence the model of IEC is reliable and reasonable. The dehumidifier is internally cooled by cooling water and is simulated based on the effectiveness method [36] and the isothermal method [37]. A comparison has been made between the experimental data and the simulation results in the literature [37,38]. It can be found that the model is acceptable. For counter flow in a packed bed, the maximum achievable difference in air humidity is obtained when the outlet air is in equilibrium with the inlet desiccant solution, as long as the physical scale of the dehumidifier is large enough and the cooling water is sufficient. To see the discrepancy between theoretical value and actual value, effectiveness of heat exchangers is reconsidered to calculate the performance of the system, where the effectiveness of airto-air heat exchanger eAHE is chosen as 0.8, and the effectiveness of solution-to-solution heat exchanger eSHE is set to 0.85. The calculation results and discrepancy are shown in Table 2. As we can see, the maximum discrepancy of COP is 3.0%, and the maximum discrepancy of q0 is 2.94%. The maximum discrepancy of DTsc is 13.3%. The minimum discrepancy of the three performance indexes is 0%. Therefore, the assumption of the effectiveness for the heat exchangers to be 1.0 is acceptable.
ð16Þ
When h5 is certain, the refrigerant outlet temperature (T5) will be obtained. If the heat and mass transfer process in the indirect evaporative cooler is fully developed, T5 should be equal to the closed air outlet temperature (T7).
3.4. Performance indexes In this study, three performance indexes are used to investigate the influences of the studied parameters on the proposed energyefficient refrigeration system. These indexes are as follows:
3.3. Model validation
(1) Subcooling degree (DTsc), which is defined as the temperature difference between condensing temperature and the refrigerant temperature leaving indirect evaporative cooler.
Modeling the regeneration processes in the liquid desiccant cycle is critical for the performance simulation of the whole system. NTU-Le model and finite element method have been validated satisfactorily by the experimental data in the literature [31]. Hence, the model of regenerator is reliable enough to be used in the performance evaluation of the whole system. When eAHE is set to 1.0, the closed air inlet and outlet temperature of indirect evaporative cooler (IEC) are equal, so the closed air experiences an isothermal process in the ideal situation. This can
DT sc ¼ T con T 5
ð17Þ
(2) Specific refrigerating capacity (q0), which is calculated from the enthalpy difference between outlet and inlet refrigerant of the evaporator.
q0 ¼ h1 h6
ð18Þ
Table 2 Discrepancy between the theoretical value and actual value at different operation conditions. Tcon (°C)
Tamb (°C)
RHamb (%)
Xs (%)
COP Theoretical value
Actual value
Discrepancy (%)
q0 (kJ/kg) Theoretical value
Actual value
Discrepancy (%)
DTsc (°C) Theoretical value
Actual value
Discrepancy (%)
50 52.5 55 50 50 50 50 50
35 35 35 31 35 35 35 35
60 60 60 60 30 40 50 60
33 33 33 33 33 33 33 31
5.61 5.82 5.65 6.11 6.19 6.13 6.05 5.94
5.58 5.65 5.58 5.96 6.19 6.11 5.95 5.85
0.54 3.00 1.25 2.52 0 0.33 1.68 1.54
183.78 196.9 196.90 200 203 200.80 198.25 194.50
182.01 191.27 194.30 195.27 203 200.30 195.08 191.76
0.97 2.94 1.34 2.40 0 0.25 1.62 1.43
26.1 39.2 41.7 39.3 41.7 39.9 37.8 34.7
24.7 34.6 39.6 35.4 41.7 39.5 35.2 32.5
5.6 13.3 5.3 11 0 1 7.4 6.7
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(3) Coefficient of performance (COP), which is defined as the ratio between the specific refrigerating capacity and the total specific power consumption including compressor, pumps and fans in the energy-efficient refrigeration system.
COP ¼ q0 =ð1:1 wcom Þ
dehumidification and evaporation is shown in Fig. 5. For a given operating condition, the solution state and air state entering regenerator are determined first based on the solution-to-refrigeration heat exchanger model and the condenser model. Through regenerator model and iterative computations, the maximum regeneration mass flow of moisture and the outlet parameters of solution can be achieved. Since the closed air outlet humidity ratio of dehumidifier is same as that of concentrated solution entering dehumidifier, the closed air humidity ratio leaving dehumidifier can also be obtained. By assumption of closed air temperature leaving the indirect evaporative cooler, the subcooled refrigerating capacity and cooling load of indirect evaporative cooler are achieved based on the air-to-air heat exchanger model and indirect evaporative
ð19Þ
where, wcom is the specific power consumption of the compressor (kJ kg1). 3.5. Flowchart of the energy-efficient refrigeration system The flowchart for calculating the performance of the energyefficient refrigeration system subcooled by liquid desiccant
Refrigerant state T2,Tcon
Assume solution outlet temperature of regenerator T15
Model of solution-torefrigerant heat exchanger
Model of solution-tosolution heat exchanger
Solution state T14
Solution state T13
Solution mass flow rate Gs Air inlet parameters of regenerator T19, ω 19
Model of condenser Solution inlet parameters of regenerator T14, Xs
Choose optimal ratio Gamb /Gs
Model of regenerator
Maximum regeneration mass flow: Greg,max Outlet parameters of solution: X15,Ts,reg,out Assume closed air outlet temperature of indirect evaporative cooler (IEC) T7
NO
T15=Ts,reg,out YES
Model of indirect evaporative cooler
Model of dehumidifier
Closed air humidity ratio: ω 7
Closed air humidity ratio:ω 9
Mass conservation GIEC =Greg,max
Closed air mass flow rate:Ga
Closed air parameters T7,ω 7,T11,ω 11
Outlet temperature of refrigerant: T5 Subcooled refrigeration capacity: Qref,sc
NO
291
Cooling load of IEC: QIEC
Qref,sc=QIEC YES Calculate performance indexes Δ Tsc, q0, COP Fig. 5. Procedure for calculating the system performance.
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Table 3 The reference values and variation ranges of the studied parameters.
Tamb RHamb (%) Teva 4Tsh Tcon Xs (%)
Reference value
Minimum value
Maximum value
Increment
Unit
35 60 5 5 50 33
31 30 – – 45 31
39 70 – – 55 35
2 10 – – 2.5 1
°C – °C °C °C –
cooler model. Through iterative computations, the performance indexes can be obtained. Table 4 Description of different systems involved in the energy-efficient refrigeration system.
4. Results and discussion The effects of key parameters on performance of the proposed system are presented in this section. The main parameters of the system are as follows: condensing temperature (Tcon), ambient air temperature (Tamb), ambient air relative humidity (RHamb) and liquid desiccant concentration in regeneration (Xs). In each case, only one parameter is changed, while other parameters are kept constant at the reference values. The reference value for each parameter and the variation range are shown in Table 3. To make comparative studies, four different systems are employed, as listed in Table 4. 4.1. Effect of the condensing temperature on the system performance Theoretical analysis results of the proposed system at five different condensing temperatures are shown in Table 5. As we can see, with the increase of condensing temperature, refrigerant outlet temperature of IEC (T5) decreases at first and then remains stable at 13.3 °C, which is restrained by the solution concentration at the outlet of the dehumidifier and cooling water temperature in the dehumidifier. Fig. 6 shows the effects of condensing temperature Tcon on the subcooling degree of the liquid refrigerant (DTsc), the specific refrigerating capacity (q0) and the COP. When the condensing temperature (Tcon) rises from 45 °C to 55 °C, the hybrid systems (system 1 and system 2) could achieve more subcooling degree of the liquid refrigerant (DTsc), compared with the baseline refrigeration system (system 3), as shown in Fig. 6(a). The results also indicate that DTsc of system 1 is significantly higher than that of system 2 when Tcon is higher than 50 °C. The higher the condensing temperature (Tcon), the higher the solution inlet temperature of regenerator (Ts,in), which would provide more heat for the desiccant cycle and enhance the desiccant regeneration. For system 1, when condensing temperature (Tcon) is beyond 50 °C, the air inlet temperature of regenerator (Tamb,reg,in) is higher than minimum regeneration temperature (Ts,min) and the hot air from condenser can improve the regeneration effect significantly. Therefore, there is a considerable increase in DTsc, whereas when Tcon is above 52.5 °C, q0 increases generally, which is restrained by the dew point temperature of closed air inlet state of indirect evaporative cooler.
Systems
Description
System 1
The energy-efficient refrigeration system which uses hot air from the condenser in the desiccant regeneration The energy-efficient refrigeration system which uses ambient air in the desiccant regeneration Baseline refrigeration system without the desiccant cycle which keeps the refrigerant temperature leaving condenser at 40 °C Reverse Carnot cycle
System 2 System 3
System 4
Fig. 6(b) shows that specific refrigerating capacity (q0) of the hybrid systems is much higher than that of baseline refrigeration system. With the increase of Tcon, q0 of system 2 increases gradually from 178.7 kJ/kg to 185.7 kJ/kg, while q0 of system 1 rises generally at first, increases significantly as Tcon rises from 50 °C to 52.5 °C, and remains stable when Tcon is higher than 52.5 °C. All these changes result from the similar variation of DTsc. Fig. 6(c) shows the effect of condensing temperature on the COPs of four different systems. As shown in Fig. 6(c), with the increase of Tcon, the COPs of system 3 and system 4 decrease linearly due to the fact that higher condensing temperature would result in larger increase of compressor power consumption wcom. However, the situation for system 1 is not different since it drops slowly when Tcon is below 50 °C, increases generally when Tcon rises from 50 °C to 52.5 °C and then decreases slightly as Tcon rises above 52.5 °C. It can be observed that the COPs of the hybrid systems are higher than that of the baseline conventional vapor compression system. When Tcon is above 52.5 °C, system 1 shows higher COP than the Carnot cycle (system 4), since the condensing heat utilization through desiccant cooling cycle enhances the vapor compression system. There is an apparent difference between system 1 and system 2. For system 1, it seems that there exists a maximum COP while Tcon is around 52.5 °C.
4.2. Effect of the liquid desiccant concentration on the system performance Theoretical analysis results of the proposed system at five different solution concentrations are shown in Table 6. As we can see, with the increase of solution concentration, refrigerant outlet
Table 5 Theoretical analysis results of the proposed system at different condensing temperature. Tcon (°C)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
x19
T20 (°C)
(g/kg)
x20
45 47.5 50 52.5 55
27.0 25.4 23.9 13.3 13.3
23 20.9 19 9.7 9.7
9.0 8.98 8.9 8.4 8.4
56.3 59.2 62.2 65.1 68
33 33 33 33 33
47 47.9 48.5 46.6 47.2
33.2 33.3 33.4 34.1 34.2
40 42.5 45 47.5 50
21.8 21.8 21.8 21.8 21.8
56 59 61.8 46.8 49.1
38.1 44.6 51.5 23.2 23.4
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(b)
System 1
System 1 System 2
System 2 System 3
System 3
q0 (kJ/kg)
Tsc (oC)
(a)
Tcon (oC)
Tcon (oC)
(c)
System 1 System 2
COP
System 3 System 4
Tcon (oC) Fig. 6. Effect of condensing temperature (Tcon) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
temperature of IEC (T5) decreases at first and then increases. There exists a minimum value when Xs is 0.32. Fig. 7 shows the effects of the desiccant concentration Xs on the system performance indexes. As shown in Fig. 7(a), in general, DTsc of the hybrid system, including system 1 and system 2, is at least 137% higher than that of the baseline system (system 3) in the variation range. DTsc of system 1 is higher than that of system 2, especially when Xs is below 0.33. It can be observed that DTsc of system 1 increases slightly from 34.7 °C to 35.1 °C when Xs increases from 0.31 to 0.32, and drops significantly as Xs is beyond 0.33, however DTsc of system 2 decreases slightly always with the increase of Xs. The reasons can be explained as follows: the higher the Xs, the lower the desiccant regeneration thermal efficiency when Ts,in remains stable. Consequently the desiccant cycle would provide less evaporative cooling capacity and the relatively less DTsc in system 1 and system 2. Fig. 7(b) shows q0 of system 1 increases slightly when Xs increases from 0.31 to 0.32, and drops significantly from 194.96 kJ/ kg to 183.78 kJ/kg as Xs increases from 0.32 to 0.33, however q0 of system 2 decreases only by 1.47% in the variation range and is lower than that of system 1. All the changes are due to the similar changes of DTsc with the increase of Xs when Tcon remains stable. As shown in Fig. 7(c), COPs of the hybrid systems (system 1 and system 2) are much higher than that of the baseline system (system 3), which indicates that the proposed system is quite an effective way to improve the performance of conventional vapor compression refrigeration system. The COP of system 1 increases slightly with lower desiccant concentration, and then drops
significantly when the Xs is more than 0.32. Whereas, COP of system 2 is lower than that of system 1 and decreases slightly from 5.59 to 5.52 as Xs increases from 0.31 to 0.35, which results from the corresponding changes of q0 with the increase of Xs when wcom is kept constant. Suitable concentration of the desiccant solution should be considered carefully for higher performance of the hybrid system. In the typical case, the suggested concentration of the LiCl–H2O solution is around 0.32 for the hybrid vapor compression system. 4.3. Effect of the ambient air temperature on the system performance Theoretical analysis results of the proposed system at five different ambient air temperatures are shown in Table 7. As we can see, with the increase of ambient air temperature, refrigerant outlet temperature of IEC (T5) increases. This is due to the fact that higher ambient air temperature leads to the worse regeneration efficiency and consequently the lower cooling capacity in the IEC. The parameters of the ambient air would impact the desiccant cooling system evidently. In the analysis, the temperature of the air varies from 31 °C to 39 °C while the relative humidity of the air is kept constant at 60%. Fig. 8 shows the effects of Tamb on the performance indexes. The hybrid systems (system 1 and system 2) achieve much more subcooling degree (DTsc) than the baseline system (system 3). As Tamb increases, DTsc of the hybrid system decreases, as shown in Fig. 8(a). It is clear that DTsc of system 1 is significantly influenced, while DTsc of system 2 decreases slightly in the studied range. Due to the same relative humidity of the air in
Table 6 Theoretical analysis results of the proposed system at different solution concentration. Xs (%)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
x19
T20 (°C)
(g/kg)
x20
31 32 33 34 35
15.3 14.9 23.9 24.6 24.9
11 10.7 19 19.9 20.3
9.9 9.3 8.9 8.2 7.7
62.2 62.2 62.2 62.2 62.2
31 32 33 34 35
43.9 44.7 48.5 50 50.6
32.0 32.8 33.4 34.4 35.3
45 45 45 45 45
21.8 21.8 21.8 21.8 21.8
44.5 45 61.8 62 62
22.7 22.7 51.5 48.5 45
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(a)
(b)
System 1 System 2 System 3
q0 (kJ/kg)
Tsc (o C)
System 1 System 2 System 3
Xs
Xs
(c)
System 1 System 2 System 3
COP
System 4
Xs Fig. 7. Effect of liquid desiccant concentration (Xs) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
Table 7 Theoretical analysis results of the proposed system at different ambient air temperature. Tamb (°C)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
T20 (°C)
(g/kg)
31 33 35 37 39
10.7 12 23.9 24.2 24.9
8.1 8.9 19 19.4 20.3
6.8 7.7 8.9 9.9 10.8
62.2 62.2 62.2 62.2 62.2
33 33 33 33 33
43.1 44.3 48.5 50.1 51
34.3 34 33.4 33.3 33.3
45 45 45 45 45
17.3 19.4 21.8 24.4 27.3
43.7 44.5 61.8 62 62
18.9 20.6 51.5 52 52.3
the analysis, the air humidity ratio increases significantly with the increase of Tamb, which would reduce the regeneration effect. For the system 1, the general decrease of DTsc is due to the fact that higher air temperature determines higher dehumidification temperature, which results in the increase of dew point temperature of closed air entering the indirect evaporative cooler. Fig. 8(b) shows that q0 of system 1 decreases generally from 200 kJ/kg to 183.78 kJ/kg as Tamb increases from 31 °C to 35 °C, while q0 of system 2 decreases slightly in the variation range, which is due to the corresponding changes of DTsc when Tcon is kept constant. Fig. 8(c) shows the comparisons of COP of different systems involved in the energy-efficient refrigeration system. It can be observed that COP of the hybrid systems (system 1 and system 2) is at least 6.14% higher than that of the baseline system (system 3). The COP of system 1 is higher than that of system 2, especially when Tamb is lower than 35 °C. 4.4. Effect of the ambient air relative humidity on the system performance Theoretical analysis results of the proposed system at different ambient air relative humidity are shown in Table 8. As we can see, with the increase of ambient air relative humidity, refrigerant outlet temperature of IEC (T5) increases. In this analysis, the ambient air temperature is constant at 35 °C. Fig. 9 shows the effects of the ambient air relative humidity
x19
x20
RHamb on subcooling degree (DTsc), specific refrigerating capacity (q0) and the COP. DTsc of the hybrid systems (system 1 and system 2) is much higher than that of baseline system (system 3). As shown in Fig. 9(a), DTsc of system 1 decreases gradually from 41.7 °C to 37.8 °C as RHamb increases from 30% to 50%, drops significantly by 30.9%, and decreases slightly when RHamb increases beyond 0.6, whereas DTsc of system 2 drops considerably by 31.5% as RHamb increases from 30% to 40%. The ambient air humidity ratio increases with the increase of RHamb and constant air temperature, which reduces its potential for the moisture transfer from solution to air in the desiccant regeneration, and consequently both curves of system 1 and system 2 decrease with the increase of RHamb. Fig. 9(b) shows the similar trends of q0 and DTsc. For system 1, q0 decreases gradually when RHamb is below 50% and reduces significantly from 198.25 kJ/kg to 183.78 kJ/kg. In terms of system 2, q0 decreases considerably from 198.6 kJ/kg to 183.78 kJ/kg when RHamb increases from 30% to 40%. All the changes of q0 result from the corresponding changes of DTsc with the increase of RHamb, when Tcon is kept constant. Fig. 9(c) shows the comparisons of COP. It is observed that COPs of system 1 and system 2 are at least 6% higher than that of the conventional vapor compression system (system 3) with different relative humidity of the ambient air. COP of system 1 decreases from 6.19 to 6.05 linearly when RHamb rises from 30% to 50%. Moreover, COP of system 1 is even higher than that of the ideal Carnot cycle (system 4) when RHamb is below 40%.
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(a)
(b)
System 1 System 2 System 3
q0 (kJ/kg)
Tsc (oC)
System 1 System 2 System 3
Tamb (oC)
Tamb (oC)
(c)
System 1 System 2
COP
System 3 System 4
Tamb (oC) Fig. 8. Effect of ambient air temperature (Tamb) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
Table 8 Theoretical analysis results of the proposed system at different ambient air relative humidity. RHamb (%)
T5 (T7,T11) (°C)
x7
x9
(g/kg)
(g/kg)
T14 (°C)
X14 (%)
T15 (°C)
X15 (%)
T19 (°C)
(g/kg)
T20 (°C)
(g/kg)
30 40 50 60 70
8.3 10.1 12.2 23.9 24.6
6.9 7.8 9 19 19.9
5.5 6.5 7.7 8.9 10.7
62.2 62.2 62.2 62.2 62.2
33 33 33 33 33
38.5 41.2 43.6 48.5 50.9
34.7 34.4 34.1 33.4 33.3
45 45 45 45 45
10.7 14.3 18 21.8 25.6
43.8 44 44.2 61.8 62
11.9 15.6 19.2 51.5 52.3
Tsc (oC)
System 1 System 2 System 3
(b)
RHamb
x20
System 1 System 2 System 3
q0 (kJ/kg)
(a)
x19
RHamb
(c)
System 1 System 2
COP
System 3 System 4
RH amb Fig. 9. Effect of ambient air relative humidity (RHamb) on (a) subcooling degree (DTsc), (b) specific refrigerating capacity (q0) and (c) coefficient of performance (COP).
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5. Conclusion This paper presents the thermodynamic analysis of the energyefficient refrigeration system subcooled by the liquid desiccant dehumidification and indirect evaporation. The indirect evaporative cooler with dry air leaving from dehumidifier is employed to subcool the liquid refrigerant at the outlet of condenser, while the compensation heat of the desiccant regeneration is provided by the condensation heat. Main conclusions can be drawn as follows: (1) Generally, the proposed hybrid vapor compression refrigeration system shows significantly higher COP than conventional vapor compression refrigeration system with the maximum COPs using hot air and ambient air increasing by 18.8% and 16.3% respectively, and even higher than the reverse Carnot cycle at the same operation conditions, which indicates that the subcooling method using liquid desiccant and evaporation is quite an efficient way to improve the performance of conventional refrigeration system. (2) Effects of important operation parameters on the hybrid vapor compression refrigeration system are disclosed. The system would show more obvious potential compared with traditional vapor compression refrigeration system at higher condensing temperature. Suitable concentration of the desiccant solution in the desiccant cycle should be considered carefully for higher performance of the hybrid system. In regard to LiCl–H2O as the desiccant, the suggested mass concentration is around 0.32. (3) Theoretical calculations using hot air from condenser and ambient air for desiccant regeneration indicate that the performance of system 1 is higher than that of system 2, especially when the air inlet temperature of regenerator is higher than the minimum regeneration temperature. The maximum difference percentage between system 1 and system 2 is 9.5% under studied conditions. Since the theoretical analysis indicates that the new subcooling method for vapor compression refrigeration systems is very effective, an experimental setup will be established in an environment chamber with regulable air temperature and humidity, combining the liquid desiccant system with the indirect evaporative cooling system available at the moment. Acknowledgements This work was partially supported by Natural Science Foundation of China (No. 51006022) and the 12th Five Year Science and Technology Support Key Project of China (No. 2011BAJ03B14). References [1] Perez-Lombard L, Ortiz J, Pout C. A review on buildings energy consumption information. Energy Buildings 2008;40:394–8. [2] Thu HTM, Sato H. Proposal of an eco-friendly high-performance airconditioning system part 2. Application of evapo-transpiration condenser to residential air-conditioning system. Int J Refrig 2013;36:1596–601. [3] Youbi-Idrissi M, Macchi-Tejeda H, Fournaison L, Guilpart J. Numerical model of sprayed air cooled condenser coupled to refrigerating system. Energy Convers Manage 2007;48(7):1943–51. [4] Gong GC, Chen FH, Su H, Zhou JY. Thermodynamic simulation of condensation heat recovery characteristics of a single stage centrifugal chiller in a hotel. Appl Energy 2012;91:326–33. [5] Yi XW, Lee WL. The use of helical heat exchanger for heat recovery domestic water-cooled air-conditioners. Energy Convers Manage 2009;50(2):240–6. [6] Jiang ML, Wu JY, Wang RZ. Research on refrigerant flow characteristics and performance of a multi-functional heat pump system. Energy Convers Manage 2011;52(6):2323–8.
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