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Performance evaluation of an ejector subcooled vapor-compression refrigeration cycle
Energy Conversion and Management 92 (2015) 431–436
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Performance evaluation of an ejector subcooled vapor-compression refrigeration cycle Meibo Xing, Gang Yan ⇑, Jianlin Yu Department of Refrigeration & Cryogenic Engineering, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, China
a r t i c l e
i n f o
Article history: Received 22 October 2014 Accepted 29 December 2014 Available online 16 January 2015 Keywords: Refrigerating system Ejector Subcooling Performance
a b s t r a c t In this study, a novel vapor-compression refrigeration cycle with mechanical subcooling using an ejector is proposed to improve the performance of a conventional single-stage vapor-compression refrigeration cycle. In the theoretical study, a mathematical model is developed to predict the performance of the cycle by using R404A and R290, and then compared with that of the conventional refrigeration cycle. The simulation results show that the performance of the ejector subcooled cycle is better than that of the conventional cycle. When the evaporator temperature ranges from 40 to 10 °C and the condenser temperature is 45 °C, the novel cycle displays volumetric refrigeration capacity improvements of 11.7% with R404A and 7.2% with R290. And the novel cycle achieves COP improvements of 9.5% with R404A and 7.0% with R290. In addition, the improvement of the COP and cooling capacity of this novel cycle largely depends on the operation pressures of the ejector. The potential practical advantages offered by the cycle may be worth further attention in future studies. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Vapor-compression refrigeration cycles are widely used in a variety of refrigerator, air conditioner and heat pump systems [1,2]. They contribute an appreciable part of energy consumption in energy use, and thus improving their energy efficiency is of paramount importance. In practical applications, nevertheless, various thermodynamic losses in the cycles make the system performance to degrade, Ahamed et al. [3]. Therefore, many attempts to improve the vapor-compression refrigeration cycle efficiency by using cycle modifications have been made over the past decades. Typically, using mechanical subcooling for COP increase in vapor-compression refrigeration cycles is a known method [4,5]. For examples, Qureshi et al. [6,7] proposed a dedicated mechanical subcooling cycle to enhance the system COP and confirmed that the secondlaw efficiency of the cycle could be significantly increased through the subcooling. Yang and Zhang [8,9] applied the mechanical subcooling for an integrated supermarket refrigeration system and demonstrated the merits of the mechanical subcooling method with appropriate subcooler designs. Torrella et al. [10] carried out an experimental analysis on a two-stage refrigerating cycle with a subcooler system and showed a gain in COP of the two-stage cycle. ⇑ Corresponding author. Tel.: +86 29 82668738; fax: +86 29 82668725. E-mail address: [email protected] (G. Yan). http://dx.doi.org/10.1016/j.enconman.2014.12.091 0196-8904/Ó 2015 Elsevier Ltd. All rights reserved.
In addition, other subcooling methods have been also proposed in the literatures to improve vapor compression refrigeration cycles. These methods include the air conditioning system utilizing a cold storage unit as a subcooler, the heat pump system with ice storage subcooler and the refrigeration system subcooled by liquid desiccant dehumidification as well as the CO2 transcritical vapor compression cycle with thermoelectric subcooler [11–15]. Relevant analytical or experimental results in the literatures all confirmed improvements in performance when using the new subcooling methods in vapor compression refrigeration cycles. In general, various active subcooling methods could provide potential way of increasing refrigerating effect and cycle efficiency for conventional vapor-compression refrigeration cycles. In this paper, a mechanical subcooling method using an ejector is proposed to improve the performance of a single-stage vaporcompression refrigeration cycle. With the use of the ejector and an additional mechanical pump, the equivalent subcooling effect in the cycle can be realized, resulting in the improved cooling capacity and coefficient of performance (COP) of the cycle system. In fact, ejector refrigeration and its combined system are widely investigated by many researchers [16,17]. In the normal ejector refrigeration, ejectors have since been used in steam ejector refrigeration applications driven by heat, where a boiler, an ejector, and a pump are used to replace the mechanical compressor of a conventional vapor compression refrigeration system [18–22]. In a steam ejector refrigeration system, the ejector utilizes the motive
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Nomenclature COP h _ m P q Q_ c t v w _ W
coefficient of performance specific enthalpy (kJ kg1) mass flow rate (kg s1) pressure (kPa) volumetric capacity (kJ m3) refrigeration capacity (kW) temperature (°C) specific volume (m3 kg1) velocity (m s1) input power (kW)
Greek letters l entrainment ratio p pressure ratio
g
isentropic efficiency
Subscripts c condenser, compressor d diffuser e evaporator m mixed flow n nozzle p primary flow, pump s secondary flow, isentropic process 1 inlet, the first nozzle 2 outlet, the second nozzle 1–10, 90 , 100 state points of refrigerant
stream generated from the boiler to draw low pressure refrigerant from the evaporator, resulting in the refrigerant to evaporate at low pressure and produce the useful refrigeration. Thus, the ejector plays a vapor compression role in the ejector refrigeration system. It is worth noting that ejector applications in vapor-compression refrigeration cycles have also been concerned in the past. Previous researches have been mainly concentrated on the ejector as an expansion device in vapor compression refrigeration cycles [23–25]. These research revealed the advantage of using the ejector for enhancing the efficiency and cooling capacity of the cycles. This study is to further the ejector application in a single-stage vaporcompression refrigeration cycle for refrigeration, and provide a description of the ejector subcooling method as well as benefits from the use of the technology. The purpose of this work aims at the development of a novel vapor compression cycle for refrigeration with use of the ejector and an additional mechanical pump. The major improvement proposed hereby is a mechanical subcooling method by an ejector device.
2. Cycle description and modeling The cycle configuration for the proposed vapor-compression refrigeration cycle with mechanical subcooling is shown in Fig. 1(a). Compared to a conventional vapor-compression refrigeration cycle (CVRC), the cycle has a main circuit coupled with a subcooling circuit. The main circuit consists of a compressor, a condenser, a flash tank, an evaporator and two expansion valves (EVs). The subcooling circuit includes a mechanical pump and an ejector. In the cycle, the refrigerant liquid leaving the condenser (point 4) is separated into two streams, and the main refrigerant flow expands through the first expansion valve and enters the flash tank (point 5). The saturated refrigerant liquid coming from the flash tank (point 7) is diverted through the second expansion valve to the evaporator (point 8) to produce refrigerating effect. The refrigerant vapor from the evaporator (point 1) is compressed by the compressor, and then is discharged to the condenser (point 2). Another portion of refrigerant liquid leaving the condenser is driven through the compression process of the mechanical pump to the ejector (point 9) to entrain the saturated refrigerant vapor from the flash tank (point 6), and the two-phase mixed refrigerant fluid (point 10) returns to the condenser. Then, the two fluids from both the compressor and ejector are again mixed and diverted to the condenser (point 3). The corresponding P–h (pressure–specific enthalpy) diagram for the cycle is shown in Fig. 1(b). In the P–h diagram, process paths 9–90 , 90 –100 and 100 –10 are the expansion process, mixing process and compression
Fig. 1. The ESVRC cycle: (a) schematic system; (b) P–h diagram.
M. Xing et al. / Energy Conversion and Management 92 (2015) 431–436
process in an ejector, respectively, which represent the ejector working process. From the operation of cycle, it can be known that the second expansion valve receives the saturated liquid refrigerant from the flash tank and provides lower quality refrigerant through an expansion process to the evaporator. This could help to decrease the enthalpy of the refrigerant entering the evaporator, resulting in an increase in the cooling capacity. Compared to the CVRC, the cycle obtain an equivalent subcooling effect (from point 4 to 7 as shown in Fig. 1(b)), and thus it is referred to as an ejector subcooled vapor-compression refrigeration cycle (ESVRC). On the other hand, the saturated refrigerant vapor in the flash tank has to be removed by the ejector, where the ejector is driven by the mechanical pump with an amount of power consumption. In this case, total power consumption for the ESVRC would be increased, which cause a change in COP for the cycle. However, the additional power supplied to the mechanical pump is relatively small due to compressing the liquid refrigerant. Thus, it is possible to improve the COP with the use of ejector in the ESVRC. Overall, this equivalent subcooling method by using an ejector is similar to other mechanical subcooling methods, which may allow the performance improvement of the ESVRC. Besides, the proposed scheme ESVRC with using an ejector and pump is also relatively simple as compared to other mechanical subcooling schemes, such as the dedicated mechanical sub-cooling scheme as shown in Ref. [4]. The scheme of a pump working downstream of the condenser is also possible because it is widely used in ejector refrigeration cycle [19]. However, it should be noted that for the actual application of ESVRC, there must be a liquid subcooling at the condenser outlet for avoiding the cavitation of the refrigerant circulating pump. The theoretical modeling of the ESVRC can be performed based on the first law of thermodynamics. For modeling the ejector, the 1-D constant–pressure mixing model and the homogeneous flow model are carried out [26]. Additionally, the following common assumptions have been made in the thermodynamic analysis: (1) All components are assumed to be a steady-state and steady-flow process. (2) The throttling process in the expansion valve is isenthalpic. (3) Refrigerant pressure losses in heat exchangers are neglected. (4) Refrigerant heat losses in the ejector are neglected. (5) The inlet velocities of the primary fluid and secondary fluid of ejector are neglected. (6) The compressor and the pump have given isentropic efficiencies, and the efficiencies of the ejector are also considered as constant values.
The velocity of the mixed refrigerant flow in the mixing chamber can be found by:
wm;100 ¼
wp;90 pffiffiffiffiffiffi gm 1þl
ð3Þ
where gm is the mixing efficiency accounting for the frictional loss in the mixing chamber. And the enthalpy of the refrigerant flow is, 2
hm;100 ¼
hp;9 þ lhs;6 wm;100 1þl 2
_ m
l¼ _6 m9
ð1Þ
_ 6 is the mass flow rate of secondary fluid (i.e. the saturated where m _ 9 is the mass flow rate refrigerant vapor from the flash tank), and m of primary fluid coming from the condenser. Based on the energy equations in the ejector, the velocity of primary refrigerant flow leaving the nozzle can be determined as:
wp;90 ¼
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2gn ðhp;9 hp;90 s Þ
ð2Þ
where hp,9 is the inlet specific enthalpy of the primary refrigerant flow, hp;90 s is the exit enthalpy through an isentropic expansion process in the nozzle, and gn is the nozzle isentropic efficiency.
ð4Þ
where hs,6 is the inlet specific enthalpy of the secondary fluid. The exit enthalpy of the mixed refrigerant flow at the diffuser outlet can be obtained by:
hd;10 ¼ hm;100 þ
hd;10s hm;100
gd
ð5Þ
where hd,10s is the exit enthalpy of the diffuser under an isentropic compression condition for the same exit pressure, gd is the diffuser isentropic efficiency. From the above equations and neglecting the outlet velocity of the diffuser, the entrainment ratio can be summarized by the following simplified equation,
l¼
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi gn gm gd ðhp;9 hp;90 s Þ 1: ðhd;10s hm;100 Þ
ð6Þ
For the compressor and the pump, the input powers can be written as, respectively,
_ C ¼m _ 1 ðh2 h1 Þ ¼ m _ 1 ðh2s h1 Þ=gcs W
ð7Þ
_ p¼m _ 9 ðh9 h4 Þ ¼ m _ 9 ðh9s h4 Þ=gps W
ð8Þ
_ 1 and m _ 9 are the refrigerant mass flow rates through the where m compressor and pump, h2s and h9s are the refrigerant specific enthalpies of the compressor and the pump outlets under the isentropic compression processes, h2 and h9 are the refrigerant specific enthalpies of the compressor and the pump outlets under the actual compression processes, gcs and gps are the compressor and the pump isentropic efficiencies, respectively. Mass balances for ejectors and flash tank can be written as:
_ 10 ¼ m _ 9þm _6 m
ð9Þ
_5¼m _ 6þm _7 m
ð10Þ
Based on the mass and energy conservation, the enthalpy of the mixed fluid at the condenser inlet can be obtained as:
_ 2 þ h10 m _ 10 Þ=m _3 h3 ¼ ðh2 m Based on the above assumptions, the following equations for the main components of the ESVRC cycle can be obtained in terms of the mass, momentum and energy conservation. As known, the entrainment ratio l is the most important parameter for evaluating the performance of the ejector. Thus, it is defined as:
433
ð11Þ
The refrigeration capacity and volumetric refrigeration capacity can be expressed as, respectively,
_ 1 ðh1 h8 Þ Q_ c ¼ m
ð12Þ
qcv ¼ ðh1 h8 Þ=v 1
ð13Þ
where v1 is suction specific volume of the compressor. The cycle heating COP can be calculated by:
_ C þW _ PÞ COP ¼ Q_ c =ðW
ð14Þ
Thus, the simulation of the novel cycle can be performed by using the above equations as well as a set of equations of state for the refrigerant thermodynamic properties. Note that the relevant set of equations of state is not presented in the above modeling for simplicity, whereas the refrigerant thermodynamic properties are calculated by using NIST database and subroutines [27]. The simulation program is written in Fortran Language. It
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on the cycle performance need to be investigated in detail. Here, their effects are described by the pressure ratios, ppc = Pp/Pc and pcs = Pc/Ps, and will be discussed later. In addition, the compressor has a given constant isentropic efficiency of gcs = 0.75 [31]. The isentropic efficiency of a liquid pump is usually in the range of 0.6–0.85 according to some literatures [32,33], and thus the median value of 0.75 is assumed in this study. The ejector is assumed to have the following efficiencies: gn = 0.85, gd = 0.85 and gm = 0.95 [34–36]. It should be noted that in the following simulations the cycle performance are evaluated on per unit of mass flow rate at _ 1 ¼ 1 kg s1. the compressor inlet, i.e. m To indicate the performance enhancement of the ESVRC, the comparisons between the ESVRC and CVRC are shown in Figs. 3 and 4 with the various condenser and evaporator temperatures. The condenser temperature is assumed to be 25 °C and 45 °C, and accordingly the pcs is 1.14 and 1.12. In this case, the Pp is set to be 2056 and 3221 kPa for R404A as the condenser temperature of 25 and 45 °C, respectively. For the R290, the Pp is 1907 and 2849 kPa at the condenser temperature of 25 and 45 °C, respectively. Fig. 3 shows the variations of volumetric refrigeration capacity at the different condenser temperature with the evaporator temperature for two refrigerants. It can be observed that the volumetric refrigeration capacity of ESVRC cycle is much higher than that of the CVRC cycle on the whole. At the given operating
Fig. 2. Flowchart for the simulation of ESVRC cycle.
should be also noted that in the simulation procedure, the entrainment ratio of ejector are obtained by an iterative process. An outline of the calculation procedure for the solution of the above model is shown in Fig. 2. In the next section, the performance of the ESVRC cycle is simulated in detail at chosen operating conditions. In order to display the performance improvement of the cycle, the simulation result will be compared with that of the CVRC cycle, which is simulated under identical operating conditions. 3. Cycle performance analyses As known, majority of commercial refrigeration utilizes the conventional vapor compression cycle. High-GWP refrigerant R404A (GWP = 3260) as medium and low temperature refrigerant is widely used in the CVRC based commercial refrigeration. It needs to be gradually replaced by low-GWP refrigerants, such as R290 [28,29]. Therefore, in this study, the refrigerants R404A and possible alternative R290 used in small low-temperature cabinet freezers are selected as examples to analyze the performance of ESVRC cycle [30]. As known, the performance of CVRC cycle is directly dependent on the basic operating conditions: condenser temperature tc (or pressure Pc), evaporator temperature te (or pressure Pe) as well as superheating degree and subcooling degree of the cycle. Here the simulation conditions are given as follows: the condenser temperature is assumed to be 25 °C and 45 °C, while the evaporator temperature varies in the range from 40 to 10 °C. For simplicity, the degrees of superheating and subcooling are both set to zero. Since the performance of the ESVRC is directly related to the pressure of the primary fluid Pp and secondary fluid Ps (corresponding to the intermediate temperature tm in the flash tank) in the inlets of ejector, they are considered as operation parameters. Therefore, the effects of the two operation parameters
Fig. 3. Variations of volumetric refrigeration capacity with the evaporator temperature.
Fig. 4. Variations of COP with the evaporator temperature.
M. Xing et al. / Energy Conversion and Management 92 (2015) 431–436
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conditions, the refrigeration capacity enhancement of ESVRC cycle is obvious. For the R404A and R290, the volumetric refrigeration capacity of the ESVRC can be increased by 11.7% and 7.2% on average within the range of evaporator temperature, respectively, as the condenser temperature of 45 °C is specified. This is because that compared to the CVRC, the ESVRC allows reducing the enthalpy of the refrigerant entering the evaporator, and this reduction causes an increase in the refrigeration capacity of evaporator per unit of mass. On the other hand, the suction specific volume of the compressor for the ESVRC is same as that of CVRC. Consequently, the positive effect of increase in volumetric refrigeration capacity can be achieved by the ESVRC. Fig. 4 shows the variations of COP with the evaporator temperature at different condenser temperature when using R404A and R290, respectively. It can be seen that the ESVRC cycle outperforms the CVRC cycle in the COPs over the whole operating range when employing the two different refrigerants. For the R404A, the COP of the ESVRC is increased by 5.6% and 9.5% on average within the range of evaporator temperature when condenser temperature is set at 25 °C and 45 °C, respectively. Similarly, for the R290 there is a COP improvement of 4.7% and 7.0% with respect to the condenser temperature of 25 °C and 45 °C, respectively. The main reason is that although the total power consumption in the ESVRC is increased due to the additional power consumption from the mechanical pump, the benefit of the refrigeration capacity in the ESVRC is relatively higher. Accordingly, the ESVRC provides the larger COPs, where the increase of the refrigeration capacity is more significant. Moreover, the improvement of COP shows a rise trend as decreasing the evaporator temperature. When condenser temperature is set at 45 °C, the COP of ESVRC using R404A is improved by 10.94% and 8.08% as the evaporator temperature is 40 °C and 10 °C. By using R290, the improvement COP is 7.04% and 5.91% at the evaporator temperature is 40 °C and 10 °C, respectively. Therefore, the ESVRC cycle could relatively provide a better performance under the low temperature operating conditions. Besides, Fig. 4 also presents the COPs of two cycles using R290 are all larger than those of the two cycles with R404A. This also implies that from the thermodynamic point of view, R290 could be a good technical and environmental option for the ESVRC if the security problem will be carefully solved in practice. For understanding the behavior of the novel cycle, the effect of ppc and pcs on the performance of ESVRC by using R290 are evaluated at given operating conditions: tc = 45 °C and te = 30 °C. Fig. 5 shows the effects of ppc on the COP and volumetric refrigeration capacity of the ESVRC cycle with the pressure ratio pcs at a value of 1.4. It can be seen that COP increase and volumetric refrigeration capacity keep invariant with rising ppc (i.e. increasing the pump
outlet pressure). This is mainly because that for the fixed secondary fluid pressure Ps and condenser pressure Pc, the entrainment ratio of the ejector l increases when increasing the primary fluid pressure Pp. However, the mass flow rate of the saturated refrigerant liquid and vapor in the flash tank are constant. In this case, the mass flow rate of refrigerant passing though the evaporator is also constant, and thus the volumetric refrigeration capacity keeps invariant. On the other hand, the increase of entrainment ratio causes a decline of the mass flow rate of primary fluid in ejector as the mass flow rate of the secondary fluid is invariable. The input power of pump reduces with the declined mass flow rate of liquid refrigerant. For this reason, the COP increases subsequently with rising ppc. But, this trend is limited because the power consumption of mechanical pump depends on both the liquid refrigerant mass flow rate and ppc. Thus, ppc should be properly chosen in terms of the acceptable COP values. It should be noted that in the above discussion, the effect of different operation conditions on the ejector design is not analyzed here for mainly focusing on the behavior of the novel cycle. In fact, different ejector designs should be made based on different operation conditions, i.e. the fixed ejector geometry corresponds to a fixed operating condition. Hence, in practice the ejector geometry should be designed at the each operating condition in Fig. 5. Fig. 6 shows the variations of COP and volumetric refrigeration capacity with the increasing pcs, where the pressure ratio ppc is kept at 1.6. It can be observed that an increase in pcs will increase the volumetric refrigeration capacity. This trend can be mainly attributed to the fact that the flash tank pressure ps would reduce with increasing pcs when the condenser pressure Pc is fixed. This could further help to decrease the specific enthalpy of the refrigerant at the evaporator inlet, resulting in an increase in the volumetric refrigeration capacity. As shown in Fig. 6, the COP demonstrates a rise and fall trend with increasing pcs, and it can reach its maximum at an optimal pressure ratio pcs. This can be explained as follows. At the given condenser pressure Pc, the entrainment ratio of the ejector l presents a decline trend with increasing the pcs as shown in Fig. 7, which is the main operating characteristic of an ejector. Increased the pcs is associated with a decreased the flash tank pressure ps as above mentioned, which causes an increase in the mass flow rate and a lower pressure of the secondary liquid refrigerant. Consequently, a great increase in the primary fluid flow pumped into the ejector is necessary to entrain the secondary fluid (i.e. the saturated refrigerant vapor from the flash tank), which increases the input power of mechanical pump (see Fig. 7). At a small pcs, the increased refrigeration capacity plays a predominant role in the COPs. However, more mechanical pump power at a lar-
Fig. 5. Effects of the ppc on the COP and volumetric refrigeration capacity.
Fig. 6. Effects of the pcs on the COP and volumetric refrigeration capacity.
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M. Xing et al. / Energy Conversion and Management 92 (2015) 431–436
Fig. 7. Variations of the entrainment ratio and input power of pump with pcs.
ger pcs is necessary for the ESVRC cycle. This results in considerable degradation for the COPs. Thus, it is evident that for the ESVRC cycle, an optimal pressure ratio pcs will allow a greater COP improvement. This also means that from the COP analysis, an optimal intermediate temperature tm in the flash tank should be considered in the configuration of the ESVRC cycle. 4. Conclusions A novel ejector subcooled vapor-compression refrigeration cycle is presented and its features are preliminarily investigated using a mathematical model in this paper. The performance of the novel cycle by employing refrigerants R404A and R290 are studied, and then compared with that of the conventional vaporcompression refrigeration cycle. From the theoretical results, it can be concluded that the novel cycle can improve the cycle COP and volumetric refrigeration capacity by using an ejector to increase the cycle subcooling. The improvement of the COP and cooling capacity of this novel cycle largely depends on the operation pressures of the ejector. The novel cycle bring about the certain benefits for refrigeration application especially at low ambient temperature conditions. It is worth noting that in the present paper, we mainly focus on the preliminary theoretical study on the proposed ESVRC cycle. Hence, the results are estimated by the simulation with considering some parameters such as the efficiencies of ejector, pump and compressor that remain to be verified against real data. Although the initial theoretical study demonstrates the potentials of the proposed ESVRC cycle for performance enhancement, more intensive experimental study on the cycle performance is still required in the next step. References [1] Cecchinato L, Chiarello M, Corradi M. Design and experimental analysis of a carbon dioxide transcritical chiller for commercial refrigeration. Appl Energy 2010;87(750–762):2095–101. [2] Ko Y, Park S, Jin S, Kim B, Jeong JH. The selection of volume ratio of two-stage rotary compressor and its effects on air-to-water heat pump with flash tank cycle. Appl Energy 2013;104:187–96. [3] Ahamed JU, Saidur R, Masjuki HH. A review on exergy analysis of vapor compression refrigeration system. Renew Sust Energy Rev 2011;15:1593–600. [4] Qureshi BA, Zubair SM. Mechanical sub-cooling vapor compression systems: current status and future directions. Int J Refrig 2013;36:2097–110. [5] Qureshi BA, Zubair SM. The impact of fouling on performance of a vapor compression refrigeration system with integrated mechanical sub-cooling system. Appl Energy 2012;92:750–62.
[6] 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. [7] 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. [8] Yang L, Zhang CL. Analysis on energy saving potential of integrated supermarket HVAC and refrigeration systems using multiple subcoolers. Energy Build 2010;42:251–8. [9] Yang L, Zhang CL. On subcooler design for integrated two-temperature supermarket refrigeration system. Energy Build 2011;43:224–31. [10] Torrellaa E, Llopisb R, Cabello R, Sánchez D. Experimental energetic analysis of the subcooler system in a two-stage refrigeration facility driven by a compound compressor. HVAC&R Res 2009;15:583–96. [11] Hsiao MJ, Cheng CH, Huang MC, Chen SL. Performance enhancement of a subcooled cold storage air conditioning system. Energy Convers Manage 2009;50:2992–8. [12] Hsiao MJ, Kuo YF, Shen CC, Cheng CH, Chen SL. Performance enhancement of a heat pump system with ice storage subcooler. Int J Refrig 2010;33:251–8. [13] Schoenfield J, Hwang Y, Radermacher R. CO2 transcritical vapor compression cycle with thermoelectric subcooler. HVAC&R Res 2012;18:297–311. [14] Sarkar J. Performance optimization of transcritical CO2 refrigeration cycle with thermoelectric subcooler. Int J Energy Res 2013;37:121–8. [15] She X, Yin Y, Zhang X. Thermodynamic analysis of a novel energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation. Energy Convers Manage 2014;78:286–96. [16] Farshi LG, Mosaffa AH, Infante Ferreirac CA, Rosen MA. Thermodynamic analysis and comparison of combined ejector–absorption and single effect absorption refrigeration systems. Appl Energy 2014;133:335–46. [17] Chen X, Worall M, Omer S, Su Y, Riffat S. Theoretical studies of a hybrid ejector CO2 compression cooling system for vehicles and preliminary experimental investigations of an ejector cycle. Appl Energy 2013;102:931–42. [18] Chen X, Omer S, Worall M, Riffat S. Recent developments in ejector refrigeration technologies. Renew Sust Energy Rev 2013;19:629–51. [19] Elbel S. Historical and present developments of ejector refrigeration systems with emphasis on transcritical carbon dioxide air-conditioning applications. Int J Refrig 2011;34:1545–61. [20] Khalil A, Fatouh M, Elgendy E. Ejector design and theoretical study of R134a ejector refrigeration cycle. Int J Refrig 2011;34:1684–98. [21] Grazzini G, Milazzo A, Paganini D. Design of an ejector cycle refrigeration system. Energy Convers Manage 2012;54:38–46. [22] Chen Jianyong, Havtun Hans, Palm Björn. Investigation of ejectors in refrigeration system: optimum performance evaluation and ejector area ratios perspectives. Appl Therm Eng 2014;64:182–91. [23] Banasiaka K, Hafnerb A. 1D Computational model of a two-phase R744 ejector for expansion work recovery. Int J Therm Sci 2011;50:2235–47. [24] Sumeru K, Nasution H, Ani FN. A review on two-phase ejector as an expansion device in vapor compression refrigeration cycle. Renew Sust Energy Rev 2012;16:4927–37. [25] Lia H, Caoc F, Bua X, Wanga L, Wang X. Performance characteristics of R1234yf ejector-expansion refrigeration cycle. Appl Energy 2014;121:96–103. [26] He S, Li Y, Wang RZ. Progress of mathematical modeling on ejectors. Renew Sust Energy Rev 2009;13:1760–80. [27] Lemmon EW, Huber ML, McLinden MO. NIST Thermodynamic and transport properties of refrigerants and refrigerant mixtures (REFPROP) version 8.0. NIST 2007. [28] Sarbu I. A review on substitution strategy of non-ecological refrigerants from vapour compression-based refrigeration, air-conditioning and heat pump systems. Int J Refrig 2014;46:123–41. [29] Mota-Babiloni A, Navarro-Esbrı J, Barragan A, Moles F, Peris B. Theoretical comparison of low GWP alternatives for different refrigeration configurations taking R404A as baseline. Int J Refrig 2014;44:81–90. [30] Mastrullo R, Mauro AW, Menna L, Vanoli GP. Replacement of R404A with propane in a light commercial vertical freezer: a parametric study of performances for different system architectures. Energy Convers Manage 2014;82:54–60. [31] Shilliday JA, Tassou SA, Shilliday N. Comparative energy and exergy analysis of R744, R404A and R290 refrigeration cycles. Int J Low-Carbon Technol 2009;4:104–11. [32] Dai Y, Wang J, Gao L. Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Convers Manage 2009;50:576–82. [33] Mago PJ, Chamra LM, Srinivasan K, Somayaji C. An examination of regenerative organic Rankine cycles using dry fluids. Appl Therm Eng 2008;28:998–1007. [34] Yu J, Ren Y, Chen H, Li Y. Applying mechanical subcooling to ejector refrigeration cycle for improving the coefficient of performance. Energy Convers Manage 2007;48:1193–9. [35] Elbel S, Hrnjak P. Experimental validation of a prototype ejector designed to reduce throttling losses encountered in transcritical R744 system operation. Int J Refrig 2008;31:411–22. [36] Liu F, Groll EA. Study of ejector efficiencies in refrigeration cycles. Appl Therm Eng 2013;52:360–70.
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Performance evaluation of an ejector subcooled vapor-compression refrigeration cycle Meibo Xing, Gang Yan ⇑, Jianlin Yu Department of Refrigeration & Cryogenic Engineering, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, China
a r t i c l e
i n f o
Article history: Received 22 October 2014 Accepted 29 December 2014 Available online 16 January 2015 Keywords: Refrigerating system Ejector Subcooling Performance
a b s t r a c t In this study, a novel vapor-compression refrigeration cycle with mechanical subcooling using an ejector is proposed to improve the performance of a conventional single-stage vapor-compression refrigeration cycle. In the theoretical study, a mathematical model is developed to predict the performance of the cycle by using R404A and R290, and then compared with that of the conventional refrigeration cycle. The simulation results show that the performance of the ejector subcooled cycle is better than that of the conventional cycle. When the evaporator temperature ranges from 40 to 10 °C and the condenser temperature is 45 °C, the novel cycle displays volumetric refrigeration capacity improvements of 11.7% with R404A and 7.2% with R290. And the novel cycle achieves COP improvements of 9.5% with R404A and 7.0% with R290. In addition, the improvement of the COP and cooling capacity of this novel cycle largely depends on the operation pressures of the ejector. The potential practical advantages offered by the cycle may be worth further attention in future studies. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Vapor-compression refrigeration cycles are widely used in a variety of refrigerator, air conditioner and heat pump systems [1,2]. They contribute an appreciable part of energy consumption in energy use, and thus improving their energy efficiency is of paramount importance. In practical applications, nevertheless, various thermodynamic losses in the cycles make the system performance to degrade, Ahamed et al. [3]. Therefore, many attempts to improve the vapor-compression refrigeration cycle efficiency by using cycle modifications have been made over the past decades. Typically, using mechanical subcooling for COP increase in vapor-compression refrigeration cycles is a known method [4,5]. For examples, Qureshi et al. [6,7] proposed a dedicated mechanical subcooling cycle to enhance the system COP and confirmed that the secondlaw efficiency of the cycle could be significantly increased through the subcooling. Yang and Zhang [8,9] applied the mechanical subcooling for an integrated supermarket refrigeration system and demonstrated the merits of the mechanical subcooling method with appropriate subcooler designs. Torrella et al. [10] carried out an experimental analysis on a two-stage refrigerating cycle with a subcooler system and showed a gain in COP of the two-stage cycle. ⇑ Corresponding author. Tel.: +86 29 82668738; fax: +86 29 82668725. E-mail address: [email protected] (G. Yan). http://dx.doi.org/10.1016/j.enconman.2014.12.091 0196-8904/Ó 2015 Elsevier Ltd. All rights reserved.
In addition, other subcooling methods have been also proposed in the literatures to improve vapor compression refrigeration cycles. These methods include the air conditioning system utilizing a cold storage unit as a subcooler, the heat pump system with ice storage subcooler and the refrigeration system subcooled by liquid desiccant dehumidification as well as the CO2 transcritical vapor compression cycle with thermoelectric subcooler [11–15]. Relevant analytical or experimental results in the literatures all confirmed improvements in performance when using the new subcooling methods in vapor compression refrigeration cycles. In general, various active subcooling methods could provide potential way of increasing refrigerating effect and cycle efficiency for conventional vapor-compression refrigeration cycles. In this paper, a mechanical subcooling method using an ejector is proposed to improve the performance of a single-stage vaporcompression refrigeration cycle. With the use of the ejector and an additional mechanical pump, the equivalent subcooling effect in the cycle can be realized, resulting in the improved cooling capacity and coefficient of performance (COP) of the cycle system. In fact, ejector refrigeration and its combined system are widely investigated by many researchers [16,17]. In the normal ejector refrigeration, ejectors have since been used in steam ejector refrigeration applications driven by heat, where a boiler, an ejector, and a pump are used to replace the mechanical compressor of a conventional vapor compression refrigeration system [18–22]. In a steam ejector refrigeration system, the ejector utilizes the motive
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Nomenclature COP h _ m P q Q_ c t v w _ W
coefficient of performance specific enthalpy (kJ kg1) mass flow rate (kg s1) pressure (kPa) volumetric capacity (kJ m3) refrigeration capacity (kW) temperature (°C) specific volume (m3 kg1) velocity (m s1) input power (kW)
Greek letters l entrainment ratio p pressure ratio
g
isentropic efficiency
Subscripts c condenser, compressor d diffuser e evaporator m mixed flow n nozzle p primary flow, pump s secondary flow, isentropic process 1 inlet, the first nozzle 2 outlet, the second nozzle 1–10, 90 , 100 state points of refrigerant
stream generated from the boiler to draw low pressure refrigerant from the evaporator, resulting in the refrigerant to evaporate at low pressure and produce the useful refrigeration. Thus, the ejector plays a vapor compression role in the ejector refrigeration system. It is worth noting that ejector applications in vapor-compression refrigeration cycles have also been concerned in the past. Previous researches have been mainly concentrated on the ejector as an expansion device in vapor compression refrigeration cycles [23–25]. These research revealed the advantage of using the ejector for enhancing the efficiency and cooling capacity of the cycles. This study is to further the ejector application in a single-stage vaporcompression refrigeration cycle for refrigeration, and provide a description of the ejector subcooling method as well as benefits from the use of the technology. The purpose of this work aims at the development of a novel vapor compression cycle for refrigeration with use of the ejector and an additional mechanical pump. The major improvement proposed hereby is a mechanical subcooling method by an ejector device.
2. Cycle description and modeling The cycle configuration for the proposed vapor-compression refrigeration cycle with mechanical subcooling is shown in Fig. 1(a). Compared to a conventional vapor-compression refrigeration cycle (CVRC), the cycle has a main circuit coupled with a subcooling circuit. The main circuit consists of a compressor, a condenser, a flash tank, an evaporator and two expansion valves (EVs). The subcooling circuit includes a mechanical pump and an ejector. In the cycle, the refrigerant liquid leaving the condenser (point 4) is separated into two streams, and the main refrigerant flow expands through the first expansion valve and enters the flash tank (point 5). The saturated refrigerant liquid coming from the flash tank (point 7) is diverted through the second expansion valve to the evaporator (point 8) to produce refrigerating effect. The refrigerant vapor from the evaporator (point 1) is compressed by the compressor, and then is discharged to the condenser (point 2). Another portion of refrigerant liquid leaving the condenser is driven through the compression process of the mechanical pump to the ejector (point 9) to entrain the saturated refrigerant vapor from the flash tank (point 6), and the two-phase mixed refrigerant fluid (point 10) returns to the condenser. Then, the two fluids from both the compressor and ejector are again mixed and diverted to the condenser (point 3). The corresponding P–h (pressure–specific enthalpy) diagram for the cycle is shown in Fig. 1(b). In the P–h diagram, process paths 9–90 , 90 –100 and 100 –10 are the expansion process, mixing process and compression
Fig. 1. The ESVRC cycle: (a) schematic system; (b) P–h diagram.
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process in an ejector, respectively, which represent the ejector working process. From the operation of cycle, it can be known that the second expansion valve receives the saturated liquid refrigerant from the flash tank and provides lower quality refrigerant through an expansion process to the evaporator. This could help to decrease the enthalpy of the refrigerant entering the evaporator, resulting in an increase in the cooling capacity. Compared to the CVRC, the cycle obtain an equivalent subcooling effect (from point 4 to 7 as shown in Fig. 1(b)), and thus it is referred to as an ejector subcooled vapor-compression refrigeration cycle (ESVRC). On the other hand, the saturated refrigerant vapor in the flash tank has to be removed by the ejector, where the ejector is driven by the mechanical pump with an amount of power consumption. In this case, total power consumption for the ESVRC would be increased, which cause a change in COP for the cycle. However, the additional power supplied to the mechanical pump is relatively small due to compressing the liquid refrigerant. Thus, it is possible to improve the COP with the use of ejector in the ESVRC. Overall, this equivalent subcooling method by using an ejector is similar to other mechanical subcooling methods, which may allow the performance improvement of the ESVRC. Besides, the proposed scheme ESVRC with using an ejector and pump is also relatively simple as compared to other mechanical subcooling schemes, such as the dedicated mechanical sub-cooling scheme as shown in Ref. [4]. The scheme of a pump working downstream of the condenser is also possible because it is widely used in ejector refrigeration cycle [19]. However, it should be noted that for the actual application of ESVRC, there must be a liquid subcooling at the condenser outlet for avoiding the cavitation of the refrigerant circulating pump. The theoretical modeling of the ESVRC can be performed based on the first law of thermodynamics. For modeling the ejector, the 1-D constant–pressure mixing model and the homogeneous flow model are carried out [26]. Additionally, the following common assumptions have been made in the thermodynamic analysis: (1) All components are assumed to be a steady-state and steady-flow process. (2) The throttling process in the expansion valve is isenthalpic. (3) Refrigerant pressure losses in heat exchangers are neglected. (4) Refrigerant heat losses in the ejector are neglected. (5) The inlet velocities of the primary fluid and secondary fluid of ejector are neglected. (6) The compressor and the pump have given isentropic efficiencies, and the efficiencies of the ejector are also considered as constant values.
The velocity of the mixed refrigerant flow in the mixing chamber can be found by:
wm;100 ¼
wp;90 pffiffiffiffiffiffi gm 1þl
ð3Þ
where gm is the mixing efficiency accounting for the frictional loss in the mixing chamber. And the enthalpy of the refrigerant flow is, 2
hm;100 ¼
hp;9 þ lhs;6 wm;100 1þl 2
_ m
l¼ _6 m9
ð1Þ
_ 6 is the mass flow rate of secondary fluid (i.e. the saturated where m _ 9 is the mass flow rate refrigerant vapor from the flash tank), and m of primary fluid coming from the condenser. Based on the energy equations in the ejector, the velocity of primary refrigerant flow leaving the nozzle can be determined as:
wp;90 ¼
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2gn ðhp;9 hp;90 s Þ
ð2Þ
where hp,9 is the inlet specific enthalpy of the primary refrigerant flow, hp;90 s is the exit enthalpy through an isentropic expansion process in the nozzle, and gn is the nozzle isentropic efficiency.
ð4Þ
where hs,6 is the inlet specific enthalpy of the secondary fluid. The exit enthalpy of the mixed refrigerant flow at the diffuser outlet can be obtained by:
hd;10 ¼ hm;100 þ
hd;10s hm;100
gd
ð5Þ
where hd,10s is the exit enthalpy of the diffuser under an isentropic compression condition for the same exit pressure, gd is the diffuser isentropic efficiency. From the above equations and neglecting the outlet velocity of the diffuser, the entrainment ratio can be summarized by the following simplified equation,
l¼
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi gn gm gd ðhp;9 hp;90 s Þ 1: ðhd;10s hm;100 Þ
ð6Þ
For the compressor and the pump, the input powers can be written as, respectively,
_ C ¼m _ 1 ðh2 h1 Þ ¼ m _ 1 ðh2s h1 Þ=gcs W
ð7Þ
_ p¼m _ 9 ðh9 h4 Þ ¼ m _ 9 ðh9s h4 Þ=gps W
ð8Þ
_ 1 and m _ 9 are the refrigerant mass flow rates through the where m compressor and pump, h2s and h9s are the refrigerant specific enthalpies of the compressor and the pump outlets under the isentropic compression processes, h2 and h9 are the refrigerant specific enthalpies of the compressor and the pump outlets under the actual compression processes, gcs and gps are the compressor and the pump isentropic efficiencies, respectively. Mass balances for ejectors and flash tank can be written as:
_ 10 ¼ m _ 9þm _6 m
ð9Þ
_5¼m _ 6þm _7 m
ð10Þ
Based on the mass and energy conservation, the enthalpy of the mixed fluid at the condenser inlet can be obtained as:
_ 2 þ h10 m _ 10 Þ=m _3 h3 ¼ ðh2 m Based on the above assumptions, the following equations for the main components of the ESVRC cycle can be obtained in terms of the mass, momentum and energy conservation. As known, the entrainment ratio l is the most important parameter for evaluating the performance of the ejector. Thus, it is defined as:
433
ð11Þ
The refrigeration capacity and volumetric refrigeration capacity can be expressed as, respectively,
_ 1 ðh1 h8 Þ Q_ c ¼ m
ð12Þ
qcv ¼ ðh1 h8 Þ=v 1
ð13Þ
where v1 is suction specific volume of the compressor. The cycle heating COP can be calculated by:
_ C þW _ PÞ COP ¼ Q_ c =ðW
ð14Þ
Thus, the simulation of the novel cycle can be performed by using the above equations as well as a set of equations of state for the refrigerant thermodynamic properties. Note that the relevant set of equations of state is not presented in the above modeling for simplicity, whereas the refrigerant thermodynamic properties are calculated by using NIST database and subroutines [27]. The simulation program is written in Fortran Language. It
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on the cycle performance need to be investigated in detail. Here, their effects are described by the pressure ratios, ppc = Pp/Pc and pcs = Pc/Ps, and will be discussed later. In addition, the compressor has a given constant isentropic efficiency of gcs = 0.75 [31]. The isentropic efficiency of a liquid pump is usually in the range of 0.6–0.85 according to some literatures [32,33], and thus the median value of 0.75 is assumed in this study. The ejector is assumed to have the following efficiencies: gn = 0.85, gd = 0.85 and gm = 0.95 [34–36]. It should be noted that in the following simulations the cycle performance are evaluated on per unit of mass flow rate at _ 1 ¼ 1 kg s1. the compressor inlet, i.e. m To indicate the performance enhancement of the ESVRC, the comparisons between the ESVRC and CVRC are shown in Figs. 3 and 4 with the various condenser and evaporator temperatures. The condenser temperature is assumed to be 25 °C and 45 °C, and accordingly the pcs is 1.14 and 1.12. In this case, the Pp is set to be 2056 and 3221 kPa for R404A as the condenser temperature of 25 and 45 °C, respectively. For the R290, the Pp is 1907 and 2849 kPa at the condenser temperature of 25 and 45 °C, respectively. Fig. 3 shows the variations of volumetric refrigeration capacity at the different condenser temperature with the evaporator temperature for two refrigerants. It can be observed that the volumetric refrigeration capacity of ESVRC cycle is much higher than that of the CVRC cycle on the whole. At the given operating
Fig. 2. Flowchart for the simulation of ESVRC cycle.
should be also noted that in the simulation procedure, the entrainment ratio of ejector are obtained by an iterative process. An outline of the calculation procedure for the solution of the above model is shown in Fig. 2. In the next section, the performance of the ESVRC cycle is simulated in detail at chosen operating conditions. In order to display the performance improvement of the cycle, the simulation result will be compared with that of the CVRC cycle, which is simulated under identical operating conditions. 3. Cycle performance analyses As known, majority of commercial refrigeration utilizes the conventional vapor compression cycle. High-GWP refrigerant R404A (GWP = 3260) as medium and low temperature refrigerant is widely used in the CVRC based commercial refrigeration. It needs to be gradually replaced by low-GWP refrigerants, such as R290 [28,29]. Therefore, in this study, the refrigerants R404A and possible alternative R290 used in small low-temperature cabinet freezers are selected as examples to analyze the performance of ESVRC cycle [30]. As known, the performance of CVRC cycle is directly dependent on the basic operating conditions: condenser temperature tc (or pressure Pc), evaporator temperature te (or pressure Pe) as well as superheating degree and subcooling degree of the cycle. Here the simulation conditions are given as follows: the condenser temperature is assumed to be 25 °C and 45 °C, while the evaporator temperature varies in the range from 40 to 10 °C. For simplicity, the degrees of superheating and subcooling are both set to zero. Since the performance of the ESVRC is directly related to the pressure of the primary fluid Pp and secondary fluid Ps (corresponding to the intermediate temperature tm in the flash tank) in the inlets of ejector, they are considered as operation parameters. Therefore, the effects of the two operation parameters
Fig. 3. Variations of volumetric refrigeration capacity with the evaporator temperature.
Fig. 4. Variations of COP with the evaporator temperature.
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conditions, the refrigeration capacity enhancement of ESVRC cycle is obvious. For the R404A and R290, the volumetric refrigeration capacity of the ESVRC can be increased by 11.7% and 7.2% on average within the range of evaporator temperature, respectively, as the condenser temperature of 45 °C is specified. This is because that compared to the CVRC, the ESVRC allows reducing the enthalpy of the refrigerant entering the evaporator, and this reduction causes an increase in the refrigeration capacity of evaporator per unit of mass. On the other hand, the suction specific volume of the compressor for the ESVRC is same as that of CVRC. Consequently, the positive effect of increase in volumetric refrigeration capacity can be achieved by the ESVRC. Fig. 4 shows the variations of COP with the evaporator temperature at different condenser temperature when using R404A and R290, respectively. It can be seen that the ESVRC cycle outperforms the CVRC cycle in the COPs over the whole operating range when employing the two different refrigerants. For the R404A, the COP of the ESVRC is increased by 5.6% and 9.5% on average within the range of evaporator temperature when condenser temperature is set at 25 °C and 45 °C, respectively. Similarly, for the R290 there is a COP improvement of 4.7% and 7.0% with respect to the condenser temperature of 25 °C and 45 °C, respectively. The main reason is that although the total power consumption in the ESVRC is increased due to the additional power consumption from the mechanical pump, the benefit of the refrigeration capacity in the ESVRC is relatively higher. Accordingly, the ESVRC provides the larger COPs, where the increase of the refrigeration capacity is more significant. Moreover, the improvement of COP shows a rise trend as decreasing the evaporator temperature. When condenser temperature is set at 45 °C, the COP of ESVRC using R404A is improved by 10.94% and 8.08% as the evaporator temperature is 40 °C and 10 °C. By using R290, the improvement COP is 7.04% and 5.91% at the evaporator temperature is 40 °C and 10 °C, respectively. Therefore, the ESVRC cycle could relatively provide a better performance under the low temperature operating conditions. Besides, Fig. 4 also presents the COPs of two cycles using R290 are all larger than those of the two cycles with R404A. This also implies that from the thermodynamic point of view, R290 could be a good technical and environmental option for the ESVRC if the security problem will be carefully solved in practice. For understanding the behavior of the novel cycle, the effect of ppc and pcs on the performance of ESVRC by using R290 are evaluated at given operating conditions: tc = 45 °C and te = 30 °C. Fig. 5 shows the effects of ppc on the COP and volumetric refrigeration capacity of the ESVRC cycle with the pressure ratio pcs at a value of 1.4. It can be seen that COP increase and volumetric refrigeration capacity keep invariant with rising ppc (i.e. increasing the pump
outlet pressure). This is mainly because that for the fixed secondary fluid pressure Ps and condenser pressure Pc, the entrainment ratio of the ejector l increases when increasing the primary fluid pressure Pp. However, the mass flow rate of the saturated refrigerant liquid and vapor in the flash tank are constant. In this case, the mass flow rate of refrigerant passing though the evaporator is also constant, and thus the volumetric refrigeration capacity keeps invariant. On the other hand, the increase of entrainment ratio causes a decline of the mass flow rate of primary fluid in ejector as the mass flow rate of the secondary fluid is invariable. The input power of pump reduces with the declined mass flow rate of liquid refrigerant. For this reason, the COP increases subsequently with rising ppc. But, this trend is limited because the power consumption of mechanical pump depends on both the liquid refrigerant mass flow rate and ppc. Thus, ppc should be properly chosen in terms of the acceptable COP values. It should be noted that in the above discussion, the effect of different operation conditions on the ejector design is not analyzed here for mainly focusing on the behavior of the novel cycle. In fact, different ejector designs should be made based on different operation conditions, i.e. the fixed ejector geometry corresponds to a fixed operating condition. Hence, in practice the ejector geometry should be designed at the each operating condition in Fig. 5. Fig. 6 shows the variations of COP and volumetric refrigeration capacity with the increasing pcs, where the pressure ratio ppc is kept at 1.6. It can be observed that an increase in pcs will increase the volumetric refrigeration capacity. This trend can be mainly attributed to the fact that the flash tank pressure ps would reduce with increasing pcs when the condenser pressure Pc is fixed. This could further help to decrease the specific enthalpy of the refrigerant at the evaporator inlet, resulting in an increase in the volumetric refrigeration capacity. As shown in Fig. 6, the COP demonstrates a rise and fall trend with increasing pcs, and it can reach its maximum at an optimal pressure ratio pcs. This can be explained as follows. At the given condenser pressure Pc, the entrainment ratio of the ejector l presents a decline trend with increasing the pcs as shown in Fig. 7, which is the main operating characteristic of an ejector. Increased the pcs is associated with a decreased the flash tank pressure ps as above mentioned, which causes an increase in the mass flow rate and a lower pressure of the secondary liquid refrigerant. Consequently, a great increase in the primary fluid flow pumped into the ejector is necessary to entrain the secondary fluid (i.e. the saturated refrigerant vapor from the flash tank), which increases the input power of mechanical pump (see Fig. 7). At a small pcs, the increased refrigeration capacity plays a predominant role in the COPs. However, more mechanical pump power at a lar-
Fig. 5. Effects of the ppc on the COP and volumetric refrigeration capacity.
Fig. 6. Effects of the pcs on the COP and volumetric refrigeration capacity.
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Fig. 7. Variations of the entrainment ratio and input power of pump with pcs.
ger pcs is necessary for the ESVRC cycle. This results in considerable degradation for the COPs. Thus, it is evident that for the ESVRC cycle, an optimal pressure ratio pcs will allow a greater COP improvement. This also means that from the COP analysis, an optimal intermediate temperature tm in the flash tank should be considered in the configuration of the ESVRC cycle. 4. Conclusions A novel ejector subcooled vapor-compression refrigeration cycle is presented and its features are preliminarily investigated using a mathematical model in this paper. The performance of the novel cycle by employing refrigerants R404A and R290 are studied, and then compared with that of the conventional vaporcompression refrigeration cycle. From the theoretical results, it can be concluded that the novel cycle can improve the cycle COP and volumetric refrigeration capacity by using an ejector to increase the cycle subcooling. The improvement of the COP and cooling capacity of this novel cycle largely depends on the operation pressures of the ejector. The novel cycle bring about the certain benefits for refrigeration application especially at low ambient temperature conditions. It is worth noting that in the present paper, we mainly focus on the preliminary theoretical study on the proposed ESVRC cycle. Hence, the results are estimated by the simulation with considering some parameters such as the efficiencies of ejector, pump and compressor that remain to be verified against real data. Although the initial theoretical study demonstrates the potentials of the proposed ESVRC cycle for performance enhancement, more intensive experimental study on the cycle performance is still required in the next step. References [1] Cecchinato L, Chiarello M, Corradi M. Design and experimental analysis of a carbon dioxide transcritical chiller for commercial refrigeration. Appl Energy 2010;87(750–762):2095–101. [2] Ko Y, Park S, Jin S, Kim B, Jeong JH. The selection of volume ratio of two-stage rotary compressor and its effects on air-to-water heat pump with flash tank cycle. Appl Energy 2013;104:187–96. [3] Ahamed JU, Saidur R, Masjuki HH. A review on exergy analysis of vapor compression refrigeration system. Renew Sust Energy Rev 2011;15:1593–600. [4] Qureshi BA, Zubair SM. Mechanical sub-cooling vapor compression systems: current status and future directions. Int J Refrig 2013;36:2097–110. [5] Qureshi BA, Zubair SM. The impact of fouling on performance of a vapor compression refrigeration system with integrated mechanical sub-cooling system. Appl Energy 2012;92:750–62.
[6] 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. [7] 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. [8] Yang L, Zhang CL. Analysis on energy saving potential of integrated supermarket HVAC and refrigeration systems using multiple subcoolers. Energy Build 2010;42:251–8. [9] Yang L, Zhang CL. On subcooler design for integrated two-temperature supermarket refrigeration system. Energy Build 2011;43:224–31. [10] Torrellaa E, Llopisb R, Cabello R, Sánchez D. Experimental energetic analysis of the subcooler system in a two-stage refrigeration facility driven by a compound compressor. HVAC&R Res 2009;15:583–96. [11] Hsiao MJ, Cheng CH, Huang MC, Chen SL. Performance enhancement of a subcooled cold storage air conditioning system. Energy Convers Manage 2009;50:2992–8. [12] Hsiao MJ, Kuo YF, Shen CC, Cheng CH, Chen SL. Performance enhancement of a heat pump system with ice storage subcooler. Int J Refrig 2010;33:251–8. [13] Schoenfield J, Hwang Y, Radermacher R. CO2 transcritical vapor compression cycle with thermoelectric subcooler. HVAC&R Res 2012;18:297–311. [14] Sarkar J. Performance optimization of transcritical CO2 refrigeration cycle with thermoelectric subcooler. Int J Energy Res 2013;37:121–8. [15] She X, Yin Y, Zhang X. Thermodynamic analysis of a novel energy-efficient refrigeration system subcooled by liquid desiccant dehumidification and evaporation. Energy Convers Manage 2014;78:286–96. [16] Farshi LG, Mosaffa AH, Infante Ferreirac CA, Rosen MA. Thermodynamic analysis and comparison of combined ejector–absorption and single effect absorption refrigeration systems. Appl Energy 2014;133:335–46. [17] Chen X, Worall M, Omer S, Su Y, Riffat S. Theoretical studies of a hybrid ejector CO2 compression cooling system for vehicles and preliminary experimental investigations of an ejector cycle. Appl Energy 2013;102:931–42. [18] Chen X, Omer S, Worall M, Riffat S. Recent developments in ejector refrigeration technologies. Renew Sust Energy Rev 2013;19:629–51. [19] Elbel S. Historical and present developments of ejector refrigeration systems with emphasis on transcritical carbon dioxide air-conditioning applications. Int J Refrig 2011;34:1545–61. [20] Khalil A, Fatouh M, Elgendy E. Ejector design and theoretical study of R134a ejector refrigeration cycle. Int J Refrig 2011;34:1684–98. [21] Grazzini G, Milazzo A, Paganini D. Design of an ejector cycle refrigeration system. Energy Convers Manage 2012;54:38–46. [22] Chen Jianyong, Havtun Hans, Palm Björn. Investigation of ejectors in refrigeration system: optimum performance evaluation and ejector area ratios perspectives. Appl Therm Eng 2014;64:182–91. [23] Banasiaka K, Hafnerb A. 1D Computational model of a two-phase R744 ejector for expansion work recovery. Int J Therm Sci 2011;50:2235–47. [24] Sumeru K, Nasution H, Ani FN. A review on two-phase ejector as an expansion device in vapor compression refrigeration cycle. Renew Sust Energy Rev 2012;16:4927–37. [25] Lia H, Caoc F, Bua X, Wanga L, Wang X. Performance characteristics of R1234yf ejector-expansion refrigeration cycle. Appl Energy 2014;121:96–103. [26] He S, Li Y, Wang RZ. Progress of mathematical modeling on ejectors. Renew Sust Energy Rev 2009;13:1760–80. [27] Lemmon EW, Huber ML, McLinden MO. NIST Thermodynamic and transport properties of refrigerants and refrigerant mixtures (REFPROP) version 8.0. NIST 2007. [28] Sarbu I. A review on substitution strategy of non-ecological refrigerants from vapour compression-based refrigeration, air-conditioning and heat pump systems. Int J Refrig 2014;46:123–41. [29] Mota-Babiloni A, Navarro-Esbrı J, Barragan A, Moles F, Peris B. Theoretical comparison of low GWP alternatives for different refrigeration configurations taking R404A as baseline. Int J Refrig 2014;44:81–90. [30] Mastrullo R, Mauro AW, Menna L, Vanoli GP. Replacement of R404A with propane in a light commercial vertical freezer: a parametric study of performances for different system architectures. Energy Convers Manage 2014;82:54–60. [31] Shilliday JA, Tassou SA, Shilliday N. Comparative energy and exergy analysis of R744, R404A and R290 refrigeration cycles. Int J Low-Carbon Technol 2009;4:104–11. [32] Dai Y, Wang J, Gao L. Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Convers Manage 2009;50:576–82. [33] Mago PJ, Chamra LM, Srinivasan K, Somayaji C. An examination of regenerative organic Rankine cycles using dry fluids. Appl Therm Eng 2008;28:998–1007. [34] Yu J, Ren Y, Chen H, Li Y. Applying mechanical subcooling to ejector refrigeration cycle for improving the coefficient of performance. Energy Convers Manage 2007;48:1193–9. [35] Elbel S, Hrnjak P. Experimental validation of a prototype ejector designed to reduce throttling losses encountered in transcritical R744 system operation. Int J Refrig 2008;31:411–22. [36] Liu F, Groll EA. Study of ejector efficiencies in refrigeration cycles. Appl Therm Eng 2013;52:360–70.