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Application of an ejector in autocascade refrigeration cycle for the performance improvement
international journal of refrigeration 31 (2008) 279–286
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Application of an ejector in autocascade refrigeration cycle for the performance improvement Jianlin Yu*, Hua Zhao, Yanzhong Li Department of Refrigeration and Cryogenics Engineering, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, China
article info
abstract
Article history:
This paper presented a novel autocascade refrigeration cycle (NARC) with an ejector. In the
Received 2 November 2006
NARC, the ejector is used to recover some available work to increase the compressor suc-
Received in revised form 5 May 2007
tion pressure. The NARC enables the compressor to operate at lower pressure ratio, which
Accepted 11 May 2007
in turn improves the cycle performance. Theoretical computation model based on the con-
Published online 18 May 2007
stant pressure-mixing model for the ejector is used to perform a thermodynamic cycle analysis for the NARC with the refrigerant mixture of R23/R134a. The effects of some
Keywords:
main parameters on cycle performance were investigated. The results show the NARC
Refrigeration system
has an outstanding merit in decreasing the pressure ratio of compressor as well as increas-
Compression system
ing the COP. For NARC operated at the condenser outlet temperature of 40 C, the evapora-
Sub-cooling
tor inlet temperature of 40.3 C, and the mass fraction of R23 is 0.15, the pressure ratio of
Liquid
the ejector reaches to 1.35, the pressure ratio of compressor is reduced by 25.8% and the
Ejector
COP is improved by 19.1% over the conventional autocascade refrigeration cycle. ª 2007 Elsevier Ltd and IIR. All rights reserved.
Modelling Thermodynamic cycle Enhancement Performance
Application d’un e´jecteur dans un cycle frigorifique a` cascade avec comme objectif l’ame´lioration de la performance Mots cle´s : Syste`me frigorifique ; Syste`me a` compression ; Sous-refroidissement ; Liquide ; E´jecteur ; Mode´lisation ; Cycle thermodynamique ; Ame´lioration ; Performance
1.
Introduction
Autocascade refrigeration can use only one compressor to obtain lower refrigerating temperature between 40 C and 180 C. Especially, the use of one compressor makes the autocascade refrigeration systems to have greater advantages
such as the simple configuration, high reliability and the low cost. Therefore, autocascade refrigeration obtained more practical applications recently in commercial refrigerator, cryogenic coolers and natural gas liquefaction plants. Missimer (1997) discussed the refrigerant conversion of autocascade refrigeration systems and described selecting
* Corresponding author. Tel.: þ86 29 82668738; fax: þ86 29 82668725. E-mail address: [email protected] (J. Yu). 0140-7007/$ – see front matter ª 2007 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2007.05.008
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Nomenclature COP h _ m Q qo t u W z rp
coefficient of performance specific enthalpy, kJ kg1 K1 mass flow rate, kg s1 refrigerating capacity, kW refrigerating capacity per unit mass flow of refrigerant mixture in the compressor, kJ kg1 temperature, C velocity, m s1 compressor power, kW mass fraction pressure ratio
Greek letters m entrainment ratio h efficiency
CFC substitutes for autocascade refrigeration systems. HFC candidates like R32, R134a, R152a, and R23 can be a choice in a range of temperature. Naer and Rozhentsev (2002) investigated the application of hydrocarbon mixtures in small refrigerating and cryogenic machines, one of which employed the autocascade refrigeration cycle with one liquid phase separator. Their research results demonstrate that small, single stage, sealed, lubricated compressors can be applied to microcoolers for the temperature range of 73 C to 183 C. Kim and Kim (2002) investigated the performance of an autocascade refrigeration system using zeotropic refrigerant mixtures of R744/R134a and R744/R290. Their experimental results show that as the composition of R744 in the refrigerant mixture increases, cooling capacity is enhanced, but the coefficient of performance (COP) tends to decrease while the system pressure rises. It is known that even though the advantage of autocascade refrigeration cycle is to use only one compressor to produce low temperature over a wide span, the COP of the cycle becomes lower and higher pressure ratio risk appears for the compressor with the decreases of refrigerating temperature. In a conventional autocascade refrigeration cycle with a two-compound mixed refrigerant, the throttling process in throttling devices generates irreversible loss and reduces the usable work potential. Clearly, this loss is one important reason that causes the relative low COP of the cycle. In addition, there also exists heat transfer irreversible loss in the evaporative condenser because of the larger temperature difference across it. The lower the refrigerating temperature, the higher the heat transfer irreversible loss. This loss also reduces the COP of the cycle. The overall cycle performance can be improved if these individual losses are reduced. For reducing the throttling loss, the best ideal method is to use expander instead of the throttling device. However, additional expander actually makes the system more complicated and higher cost due to expander itself. In order to recover the potential kinetic energy in the expansion process, some researchers have attempted to use other expansion engine rather than the expander. Due to the simple configuration, the low cost and ability to handle twophase flow without damage, an ejector could be used instead
Subscripts c condenser com compressor d diffuser e evaporator ec evaporative condenser eje ejector m mixing flow n nozzle p primary flow s secondary flow, isentropic process 1 inlet in the ejector 2 outlet in the ejector 1–8 state points in Fig. 1
of a throttling device as a feasible method to recover some available work of the expansion process in the refrigeration cycle. Through the function of the ejector the compressor suction pressure can be increased, and thus resulting in less compression work and improving cycle performance. Kornhauser (1990) firstly proposed the idea to employ an ejector in vaporcompression refrigeration cycle and analyzed the thermodynamic performance of the cycle. He found a theoretical COP improvement of up to 21% over the standard cycle under standard conditions, 15 C and 30 C for evaporator and condenser temperatures, respectively. Tomasek and Radermacher (1995) analyzed a domestic refrigerator cycle with an ejector, which consists of two evaporators that operate at different pressure and temperature levels. The analysis results show that the ejector cycle gives an increase of up to 12.4% in the COP compared to that of a standard refrigerator-freezer refrigeration cycle. Disawas and Wongwises (2004) experimentally investigated the performance of the refrigeration cycle using a two-phase ejector as an expansion device. The research results show that the coefficient of performance of the cycle with the ejector is higher than that of the conventional refrigeration cycle over the whole range of experimental conditions. Li and Groll (2005) proposed transcritical CO2 refrigeration cycle with ejector expansion device and studied the effect of different operating conditions on the relative performance of the ejector expansion transcritical CO2 cycle. Their research results found that the COP of the ejector expansion transcritical CO2 cycle can be improved by more than 16% over the basic transcritical CO2 cycle for typical air conditioning operation conditions. In this paper, it is also proposed to apply an ejector in autocascade refrigeration cycle to improve the cycle performance. The present study mainly focuses on a theoretical investigation on the performance of this novel autocascade refrigeration cycle using two-compound mixed refrigerant. The cycle will be simulated at typical operating conditions to investigate the effects of main parameters, such as the composition of the used refrigerant mixtures, the evaporator outlet dew point temperature and primary pressure of the ejector on its performance. In addition, the performance comparison between the novel autocascade refrigeration
international journal of refrigeration 31 (2008) 279–286
Fig. 1 – The conventional autocascade refrigeration cycle.
cycle and the conventional autocascade refrigeration cycle will also be presented.
2. The autocascade refrigeration cycle with an ejector The conventional autocascade refrigeration cycle (CARC) using two-compound mixed refrigerant is shown in Fig. 1. The CARC system usually consists of a compressor, a condenser, two throttling devices, an evaporator, an evaporative condenser, a phase separator, and connecting tubes. The compressor is used to pump the refrigerant mixtures. The condenser rejects a heat from the inside condensate to the environment cooling media. The evaporator absorbs a refrigerated load from the cold refrigerated space. The evaporative condenser is a counter-flow heat exchanger, which acts as an evaporator for the liquid phase stream from the phase separator and a condenser for the vapor phase stream from the phase separator. The two throttling devices are used to accomplish the expansion process for the cycle, which can be expansion valves or capillary tubes.
The novel autocascade refrigeration cycle (NARC) with an ejector is shown in Fig. 2. Similar to the CARC using twocompound mixed refrigerant, the main components but the ejector in NARC are a compressor, a condenser, a phase separator, an evaporative condenser, an evaporator and two throttling devices. The ejector is set between the evaporative condenser and the evaporator. The basic working process of NARC is same as that of CARC. The refrigerant mixture vapor with initial charge composition (at point 1) is sucked by compressor. After compression to point 2, the refrigerant mixture is partially condensed to point 3 in the condenser. This twophase stream flows into the phase separator, in which the liquid phase at point 30 will have a different composition from the vapor phase at point 300 . Liquid phase has higher mass fraction of the pure refrigerant with higher boiling point and the vapor phase has higher mass fraction of the pure refrigerant with lower boiling point. The liquid phase stream will flow through the throttling device 1, expanding from point 30 to point 7 and then entering the evaporative condenser at a lower temperature (at point 7) to cool the vapor phase stream from the phase separator. After vaporizing in the evaporative condenser, this stream flows further into the ejector as primary flow slightly superheated at point 8 to entrain the vapor (secondary flow) at point 6 from the evaporator. The vapor phase stream from the phase separator condenses on the other side of the evaporative condenser, leaving as saturated liquid or as slightly subcooled liquid at point 4. This fraction of the refrigerant mixture is then expanded through the throttling device 2 to a state with lower pressure and temperature (at point 5). This stream flows into evaporator, vaporizing to realize the refrigerating effect and leaving as saturated or slightly superheated vapor at point 6. Then, the vapor is sucked by the ejector and then the mixed vapor is discharged at high pressure. Eventually, the mixture vapor from the ejector is sucked by the compressor, and it forms a complete cycle. Since the mixture vapor from the evaporative condenser is the primary flow of the ejector, the exit pressure of evaporative condenser should be higher than the exit pressure of evaporator. This also means the pressure drop through the throttling device 1 should be lower than the pressure drop in the throttling device 2. For this reason, the irreversible loss of throttling process in this throttling device will be reduced and the partial usable work can be recovered to lift the evaporator outlet pressure, i.e. increase the suction pressure of the compressor. In addition to this, the heat transfer loss in the evaporative condenser can also be reduced due to the decrease of temperature difference in the NARC compared to that in the CARC. It should be mentioned that the pressure drop in the throttling device 1 has to ensure there exists minimum temperature difference in the evaporative condenser. In general, the cycle performance of NARC can be improved by applying an ejector in it.
3.
Fig. 2 – The autocascade refrigeration cycle with an ejector.
281
Theoretical computation model
The ejector working process is shown in Fig. 3. The primary flow expands in the nozzle from high pressure to low pressure and then sucks the secondary flow. The two flows mix in the mixing section and become one mixed flow. This mixed flow
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where hp;n2;s is the ideal exit enthalpy of the primary flow under the isentropic expansion, hn is the nozzle efficiency. In the mixing section, the momentum conservation equation and energy conservation equation for the ideal mixing process according to constant pressure-mixing process are given as _ s us;n2 ¼ m _pþm _ u _ p up;n2 þ m (3) m s m;m2;s _ s hs;n2 þ u2s;n2 =2 _ p hp;n2 þ u2p;n2 =2 þ m m _ s Þðhm;m2;s þ u2m;m2;s =2Þ _pþm ¼ ðm
ð4Þ
where um;m2;s and hm;m2;s are the ideal exit velocity and enthalpy of the mixed flow under the constant pressure mixing. If neglecting the velocity of secondary flow us;n2 compared with the primary flow velocity up;n2 and considering the energy loss of the mixing process, the actual exit velocity um;m2 and enthalpy hm;m2 of the mixed flow then can be derived from Eqs. (1)–(4) as Fig. 3 – The structure and working process of ejector.
um;m2 ¼ up;n2 ðhm Þ1=2 =ð1 þ mÞ
(5)
hm;m2 ¼ hp;n1 þ mhs;n2 ð1 þ mÞ u2m;m2 =2
(6)
where the mixing efficiency hm is defined as further increases its pressure to the back pressure in the diffuser. In this paper, the ejector performance computation is carried out based on the one-dimensional constant pressure flow model used by most researchers in studying ejector refrigeration cycle. The basic principle of the model was introduced by Keenan et al. (1950) based on gas dynamics, and then developed by Huang et al. (1999) and Ouzzane and Aidoun (2003). Usually, the computation model is set up based on the following main assumptions: (1) The flow inside the ejector is steady and one dimensional. (2) The kinetic energy at the inlets of primary and secondary flow and the outlet of diffuser are negligible. (3) For simplicity, the effects of frictional and mixing losses in the nozzle, diffuser and mixing section are taken into account by using these efficiencies, the nozzle efficiency hn , the mixing efficiency hm and the diffuser efficiency hd . (4) Mixing in the mixing section of ejectors occurs at constant pressure and complying with the conversion of energy and momentum. (5) The inner walls of ejectors are adiabatic, namely no heat losses. (6) Normal transverse shocks may occur at any plane in the throat section. (7) A homogenous two-phase flow is idealized for the ejector when it involves the two-phase flow inside. Using the above assumptions, the equations for the ejector are given based on the mass, momentum and energy conservation as follows. In the nozzle section, if the inlet velocity of primary flow up;n1 is neglected, the exit velocity of primary flow up;n2 and enthalpy hp;n2 can be derived as (1) up;n2 ¼ 2hn ðhp;n1 hp;n2;s ÞÞ1=2 hp;n2 ¼ hp;n1 u2p;n2 =2
(2)
hm ¼ u2m;m2 =u2m;m2;s
(7)
and the entrainment ratio of the ejector m is defined as _p _ s =m m¼m
(8)
In diffuser section, if neglecting the exit velocity of the mixed flow and taking diffuser efficiency into account, the actual exit enthalpy of the mixed flow can be expressed according to energy balance equation as (9) hm;d2 ¼ hm;m2 þ ðhm;d2;s hm;m2 Þ hd where hm;d2;s is the ideal exit enthalpy of the mixed flow under the isentropic compression, hd is the diffuser efficiency. According to the above equations, the entrainment ratio m can be found. qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi (10) m ¼ hn hm hd hp;n1 hp;n2;s ðhm;d2;s hm;m2;s Þ 1 For the whole cycle performance calculation, the following assumptions are also made: (1) Neglect the pressure loss and the heat losses to the environment in the condenser, phase separator, evaporative condenser and evaporator. (2) The vapor stream from the separator is saturated vapor and the liquid stream from the separator is saturated liquid. (3) The flow across the throttling devices is isenthalpic. (4) Take the isentropic efficiency into account for the compressor. Therefore, the equations for the cycle performance can be found based on the energy conservation. Refrigerating capacity is given as _6 Qe ¼ ðh6 h5 Þm
(11)
Refrigerating capacity per unit mass flow of refrigerant mixture in the compressor is given as _1 qo ¼ Qe =m
(12)
international journal of refrigeration 31 (2008) 279–286
The compressor power of the compressor is given as _ 1 =hcom Wcom ¼ h2 h1 m
4.1. (13)
The coefficient of performance (COP) of the cycle can be determined by COP ¼ Qe =Wcom
(14)
_1 m _ 8 are the mass flow where h1–h8 are the enthalpies and m rates of state points 1–8 in the cycle. The corresponding thermodynamic properties of the states of the cycle can be determined from the equation of state of refrigerant mixture. In the calculating program written with Fortran Language, the NIST routines were used to calculate these properties, which is currently an industry standard (NIST Standard Reference Database, 1998).
4.
Results and discussion
In the present study, the refrigerant mixture of R23/R134a is selected as typical working fluid for simulating the cycle performance of NARC. To investigate the characteristics of the NARC, the following basic operating conditions are firstly assumed. The efficiencies of the ejector are assumed to be hn ¼ 0:90, hm ¼ 0:85, and hd ¼ 0:85 by making reference to literatures (Li and Groll, 2005; Huang et al., 1999; Cizungu et al., 2001). The compressor is assumed to have an isentropic efficiency hcom ¼ 0:65, which is a constant and does not vary by the pressure ratio in all cases. In the simulation for the cycle, we selected the temperature, not the pressure, as the given operating parameters for condenser, evaporator and evaporative condenser. For the condenser, it is the condenser outlet temperature tc . The tc used in the simulation is fixed at 40 C. For the evaporator, the evaporator outlet dew point temperature te is used as operating parameter because there is the temperature glide in it. Considering the different refrigeration request, the variable te over a wide range of 20 C to 35 C was performed in the simulation. Similarly, the evaporative condenser outlet dew point temperature tec is also selected as an analytical parameter. Actually, it is undetermined design parameter for the cycle design when the tc and te are determined, which has effects on both the temperature differences in the evaporative condenser and the performance of the ejector. The parametric study must be conducted to determine the optimum value in the simulation for the cycle design. It is known that the composition of refrigerant mixture is an important variable, which has more important effect on the cycle performance. Therefore, the mass fraction of R23 z as the initial charged refrigerant mixture at point 1 is also treated as a variable operating condition. In the simulation, the variation of the cycle performance with the change of z is represented. In addition, in order to investigate the effects of the variation of those main variables on the cycle performance, the degree of superheating and subcooling at evaporative condenser and evaporator outlet is held zero. It is noted that a performance comparison for the NARC and CARC is made under the same operating conditions, which is shown in the following corresponding figures.
283
Effects of the tec on cycle performance
The following results are obtained when the condenser outlet temperature tc ¼ 40 C, the evaporator outlet dew point temperature te ¼ 30 C, the mass fraction of R23 z ¼ 0:35. In the NARC, for a given operating conditions, the entrainment ratio of the ejector m is dependent on the primary pressure, secondary pressure and outlet pressure of ejector. The primary pressure and secondary pressure are, respectively, relevant to tec and te . The outlet pressure of ejector is determined based on meeting the mass conservation constraint for the cycle and the heat balance constraint for the evaporative condenser, which can be evaluated by iteration in the simulation. Fig. 4 shows the variation of m and pressure ratio rpeje of the ejector with the tec , respectively. In this figure, we can see that both the m and rpeje increase with the increases of the tec . The primary pressure of ejector is corresponding to the tec , which increases with the increases of tec . Consequently, this results in an increase of the entrainment ratio and outlet pressure of the ejector. For the given conditions, the rpeje can reach 1.15– 1.31. In addition, it can be found that the m is in the range of 0.80–0.88 in the present NARC, which is close to 1 in values. This is because, as this ejector operates under the given operating conditions in the NARC, the pressure ratios across the ejector are maintained at low level, which are much smaller compared to that of the ejector operated in the conventional ejector refrigeration cycle. Usually, the entrainment ratio of the ejector depends on the primary pressure, secondary pressure and back pressure of ejector. When the primary pressure and secondary pressure are fixed, the value of the entrainment ratio for the ejector increases as the value of the pressure ratio decreases, i.e. the back pressure decreases. Therefore, the ejector can attain a higher entrainment ratio in the NARC, which provides an increase of its performance. It should be noted that the high value of the entrainment ratio in present theoretical simulation was obtained after taking the efficiencies of the ejector into account in the modeling of ejector. Of course, this value could be slightly smaller than it is in real refrigerating machines due to other practical factors.
Fig. 4 – Variation of m and rpeje with the tec .
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Fig. 5 – Variation of rpcom with the tec .
improvement in COP of NARC, which is shown in Fig. 6. It can be seen that for the given conditions, NARC has a 8.7– 17.7% improvement in COP over CARC. This is just because the drop of pressure through the throttling device 1 in the NARC can be reduced compared to the previous throttling device in CARC. For this reason, the ejector can recover some available work of the previous expansion process in the CARC to increase the compressor suction pressure. As viewed from total throttling loss, the throttling loss can be reduced when the ejector is used in the NARC compared to only using the throttling device in CARC. However, since there is a small heat transfer temperature difference across the evaporative condenser of the NARC, the heat transfer area of the evaporative condenser will become larger. Therefore, the limit in the parameter of tec should be moderately considered in the design of the NARC. In addition, Fig. 7 shows the variations of refrigerating capacity per unit mass flow of refrigerant mixture in the compressor with respect to tec . The figure shows the refrigerating capacity of the NARC qo is greater than that of CARC, which is because the refrigerating capacity is very much dependent on the mass flow rate of refrigerant mixture, and that the mass flow rate of refrigerant mixture in the NARC is greater than that in CARC at the same operating conditions. As the tec increases, the refrigerating capacity of the NARC increases owing to the increase of the mass flow rate of refrigerant mixture. Therefore, NARC also represents a merit of performance improvement in refrigerating capacity per unit mass flow of refrigerant mixture in the compressor over CARC.
Moreover, the high value of the entrainment ratio also means that the mass flow rates of refrigerant mixture flowing through the evaporative condenser at both high pressure side and low pressure side are relatively matched. Fig. 5 shows the effect of tec on the pressure ratio with respect to the compressor, respectively, in NARC and CARC. It shows clearly that the pressure ratio of compressor rpcom in NARC deceases with an increase of tec . This results from an increase of the outlet pressure of the ejector, i.e. compressor inlet pressure. Therefore, the rpcom is in turn decreased with an increase of tec . However, the decrease of rpcom is limited by the minimum temperature differences in the evaporative condenser. The minimum temperature difference ðt300 t8 Þ tends to decrease with increasing tec at the given condenser outlet temperature. In Fig. 5, the corresponding ðt300 t8 Þ will exceed 5 C when tec is close to 30 C. From the results shown in Fig. 5, the rpcom of NARC is reduced by 13.1–23.5% compared with that of the CARC. As seen above, the pressure elevation of the ejector causes a decrease of the pressure ratio across the compressor and thus, decreases the compression work and increases the
As we know, the rpcom has a direct relation to the possible lower te when other operating conditions are given. In the cycle, the rpcom will increase with the decreases of the te . This not only gives high pressure ratio risk for compressor, but also significantly increases the loss of the cycle at the same time. Fig. 8 shows the effect of the ejector on the rpcom when decreasing the te . In this simulation, the tc is 40 C and the z is 0.35. The te is kept in a range of 20 C to 35 C. Due to the
Fig. 6 – Variation of COP with the tec .
Fig. 7 – Variation of qo with the tec .
4.2.
Effects of the te on cycle performance
international journal of refrigeration 31 (2008) 279–286
285
Fig. 10 – Variation of qo with the te . Fig. 8 – Variation of rpcom with the te .
4.3. Effect of the refrigerant mixture composition on cycle performance temperature glide in the evaporator, the corresponding evaporator inlet temperature is in a range of 39.4 C to 52.9 C, which is evaluated by simulation. It can be clearly seen from the figure that the rpcom of the NARC is obviously smaller than that of CARC. As the te further decreases, the rpcom is also decreased more compared to CARC. The reason for this is that when the te becomes lower, the pressure ratio of the ejector increases. Compared with CARC, the rpcom of the NARC decreases by 6.2–24.4% in the above range of te . Therefore, it can be concluded that for the NARC, the large improvement in COP can be obtained with decreasing te . Fig. 9 shows the variation of the COP with the te in the NARC and CARC for the different tec . From this figure, it was found that the NARC improves the COP by 3.9–17.4% compared to the CARC for the above operating conditions. Similar to the situation of performance improvement in COP, it can also be seen as shown in Fig. 10 that the NARC has an improvement in refrigerating capacity over the CARC. Moreover, NARC represents the more significant potential for increasing the refrigerating capacity especially at lower te .
In autocascade refrigeration cycle, the pressure–temperature relationship of the refrigerant mixture in the condenser or evaporator is dependent on the composition. A pressure change at the given temperatures is accompanied by a change of the composition, which affects the ejector performance as well as cycle performance. Therefore, in this case the selection of composition should be more appropriate. As for composition of the refrigerant mixture, the z was selected in a range of 0.15–0.45, so that the maximum operating pressure was below 3 MPa. In the following simulation, the tc and te were, respectively, 40 C and 30 C. The variations of the rpcom in the NARC and CARC with respect to z for different tec are displayed in Fig. 11. It can be seen that for the given conditions, the rpcom in the CARC increases with a decrease of the composition. For the NARC, however, the rpcom decreases with a decrease of the composition. The reason for this case is that the pressure ratio of ejector tends to increase with a decrease of the z, i.e. raise the compressor suction pressure. This in turn leads to a decrease of rpcom in the NARC relative to the total
Fig. 9 – Variation of COP with the te .
Fig. 11 – Variation of rpcom with the z.
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Fig. 12 – Variation of COP with the z.
cycle pressure lift (from evaporator pressure to condenser pressure). It can also be seen that the change tendencies of the rpcom in the NARC with decreasing the composition are same at different tec . However, the possible minimum rpcom accomplished by decreasing the composition and increasing the tec is limited as the constraint of the minimum temperature difference for the evaporative condenser. Compared with CARC, the pressure lift ratio of the ejector reaches 1.13– 1.35, the rpcom of the NARC decreases by 11.4–25.8% in the above range of the z and tec . When the rpcom is decreased, this causes the compression work to decrease. As a result, the COP increases while refrigeration request was kept constant, which can be seen in Fig. 12. Fig. 12 also shows the variation of COP with respect to the z for various tec . It can be derived from Fig. 12 that COP of NARC has higher values than that of CARC because of the function of ejector. It can also be seen that the COP has a large trend of increase for NARC as the z decreases and the tec also increases. Compared with that of CARC, the COP of NARC increases by 7.7–19.1% as shown in Fig. 12.
5.
Conclusion
This paper proposed that an ejector is employed in the autocascade refrigeration cycle to improve the cycle performance. Theoretical computation model based on the constant pressuremixing model for the ejector was used to perform a thermodynamic cycle analysis for this novel autocascade refrigeration cycle with the refrigerant mixture of R23/R134a. The effects of main parameters on the performance for the novel cycle were investigated using the theoretical model. The performance of the novel cycle is then compared with that of the conventional autocascade refrigeration cycle at the same operating conditions. From the above theoretical computation results, the following conclusions are drawn for the novel cycle: (1) On the whole, application of an ejector in autocascade refrigeration cycle can be considered an attractive option for improving cycle performance. Especially, it demonstrates an outstanding merit in decreasing the rpcom as well as
in increasing the COP when operated at lower refrigerating temperature. For example, in the NARC the pressure lift ratio of the ejector reaches to 1.35, the rpcom is reduced by 25.8% and the COP is improved by 19.1% over the CARC for operating at the condenser outlet temperature of 40 C, the evaporator inlet temperature of 40.3 C, the evaporative condenser outlet dew point temperature of 20 C, and the mass fraction of R23 is 0.15. (2) Increasing the tec is always beneficial to performance improvement of the NARC, but is limited by the minimum temperature differences of the evaporative condenser. Obviously, the performance improvement is actually at a cost of increasing the heat transfer area of the evaporative condenser. Therefore, the parameter tec should be synthetically considered in the design of NARC. (3) In NARC, both the compressor discharge pressure and the pressure ratio can be decreased as decreasing the composition of low-temperature refrigerant in two-compound mixture. But the constraint for the minimum temperature difference in the evaporative condenser should be taken into consideration when determining the appropriate composition.
references
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journal homepage: www.elsevier.com/locate/ijrefrig
Application of an ejector in autocascade refrigeration cycle for the performance improvement Jianlin Yu*, Hua Zhao, Yanzhong Li Department of Refrigeration and Cryogenics Engineering, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, China
article info
abstract
Article history:
This paper presented a novel autocascade refrigeration cycle (NARC) with an ejector. In the
Received 2 November 2006
NARC, the ejector is used to recover some available work to increase the compressor suc-
Received in revised form 5 May 2007
tion pressure. The NARC enables the compressor to operate at lower pressure ratio, which
Accepted 11 May 2007
in turn improves the cycle performance. Theoretical computation model based on the con-
Published online 18 May 2007
stant pressure-mixing model for the ejector is used to perform a thermodynamic cycle analysis for the NARC with the refrigerant mixture of R23/R134a. The effects of some
Keywords:
main parameters on cycle performance were investigated. The results show the NARC
Refrigeration system
has an outstanding merit in decreasing the pressure ratio of compressor as well as increas-
Compression system
ing the COP. For NARC operated at the condenser outlet temperature of 40 C, the evapora-
Sub-cooling
tor inlet temperature of 40.3 C, and the mass fraction of R23 is 0.15, the pressure ratio of
Liquid
the ejector reaches to 1.35, the pressure ratio of compressor is reduced by 25.8% and the
Ejector
COP is improved by 19.1% over the conventional autocascade refrigeration cycle. ª 2007 Elsevier Ltd and IIR. All rights reserved.
Modelling Thermodynamic cycle Enhancement Performance
Application d’un e´jecteur dans un cycle frigorifique a` cascade avec comme objectif l’ame´lioration de la performance Mots cle´s : Syste`me frigorifique ; Syste`me a` compression ; Sous-refroidissement ; Liquide ; E´jecteur ; Mode´lisation ; Cycle thermodynamique ; Ame´lioration ; Performance
1.
Introduction
Autocascade refrigeration can use only one compressor to obtain lower refrigerating temperature between 40 C and 180 C. Especially, the use of one compressor makes the autocascade refrigeration systems to have greater advantages
such as the simple configuration, high reliability and the low cost. Therefore, autocascade refrigeration obtained more practical applications recently in commercial refrigerator, cryogenic coolers and natural gas liquefaction plants. Missimer (1997) discussed the refrigerant conversion of autocascade refrigeration systems and described selecting
* Corresponding author. Tel.: þ86 29 82668738; fax: þ86 29 82668725. E-mail address: [email protected] (J. Yu). 0140-7007/$ – see front matter ª 2007 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2007.05.008
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Nomenclature COP h _ m Q qo t u W z rp
coefficient of performance specific enthalpy, kJ kg1 K1 mass flow rate, kg s1 refrigerating capacity, kW refrigerating capacity per unit mass flow of refrigerant mixture in the compressor, kJ kg1 temperature, C velocity, m s1 compressor power, kW mass fraction pressure ratio
Greek letters m entrainment ratio h efficiency
CFC substitutes for autocascade refrigeration systems. HFC candidates like R32, R134a, R152a, and R23 can be a choice in a range of temperature. Naer and Rozhentsev (2002) investigated the application of hydrocarbon mixtures in small refrigerating and cryogenic machines, one of which employed the autocascade refrigeration cycle with one liquid phase separator. Their research results demonstrate that small, single stage, sealed, lubricated compressors can be applied to microcoolers for the temperature range of 73 C to 183 C. Kim and Kim (2002) investigated the performance of an autocascade refrigeration system using zeotropic refrigerant mixtures of R744/R134a and R744/R290. Their experimental results show that as the composition of R744 in the refrigerant mixture increases, cooling capacity is enhanced, but the coefficient of performance (COP) tends to decrease while the system pressure rises. It is known that even though the advantage of autocascade refrigeration cycle is to use only one compressor to produce low temperature over a wide span, the COP of the cycle becomes lower and higher pressure ratio risk appears for the compressor with the decreases of refrigerating temperature. In a conventional autocascade refrigeration cycle with a two-compound mixed refrigerant, the throttling process in throttling devices generates irreversible loss and reduces the usable work potential. Clearly, this loss is one important reason that causes the relative low COP of the cycle. In addition, there also exists heat transfer irreversible loss in the evaporative condenser because of the larger temperature difference across it. The lower the refrigerating temperature, the higher the heat transfer irreversible loss. This loss also reduces the COP of the cycle. The overall cycle performance can be improved if these individual losses are reduced. For reducing the throttling loss, the best ideal method is to use expander instead of the throttling device. However, additional expander actually makes the system more complicated and higher cost due to expander itself. In order to recover the potential kinetic energy in the expansion process, some researchers have attempted to use other expansion engine rather than the expander. Due to the simple configuration, the low cost and ability to handle twophase flow without damage, an ejector could be used instead
Subscripts c condenser com compressor d diffuser e evaporator ec evaporative condenser eje ejector m mixing flow n nozzle p primary flow s secondary flow, isentropic process 1 inlet in the ejector 2 outlet in the ejector 1–8 state points in Fig. 1
of a throttling device as a feasible method to recover some available work of the expansion process in the refrigeration cycle. Through the function of the ejector the compressor suction pressure can be increased, and thus resulting in less compression work and improving cycle performance. Kornhauser (1990) firstly proposed the idea to employ an ejector in vaporcompression refrigeration cycle and analyzed the thermodynamic performance of the cycle. He found a theoretical COP improvement of up to 21% over the standard cycle under standard conditions, 15 C and 30 C for evaporator and condenser temperatures, respectively. Tomasek and Radermacher (1995) analyzed a domestic refrigerator cycle with an ejector, which consists of two evaporators that operate at different pressure and temperature levels. The analysis results show that the ejector cycle gives an increase of up to 12.4% in the COP compared to that of a standard refrigerator-freezer refrigeration cycle. Disawas and Wongwises (2004) experimentally investigated the performance of the refrigeration cycle using a two-phase ejector as an expansion device. The research results show that the coefficient of performance of the cycle with the ejector is higher than that of the conventional refrigeration cycle over the whole range of experimental conditions. Li and Groll (2005) proposed transcritical CO2 refrigeration cycle with ejector expansion device and studied the effect of different operating conditions on the relative performance of the ejector expansion transcritical CO2 cycle. Their research results found that the COP of the ejector expansion transcritical CO2 cycle can be improved by more than 16% over the basic transcritical CO2 cycle for typical air conditioning operation conditions. In this paper, it is also proposed to apply an ejector in autocascade refrigeration cycle to improve the cycle performance. The present study mainly focuses on a theoretical investigation on the performance of this novel autocascade refrigeration cycle using two-compound mixed refrigerant. The cycle will be simulated at typical operating conditions to investigate the effects of main parameters, such as the composition of the used refrigerant mixtures, the evaporator outlet dew point temperature and primary pressure of the ejector on its performance. In addition, the performance comparison between the novel autocascade refrigeration
international journal of refrigeration 31 (2008) 279–286
Fig. 1 – The conventional autocascade refrigeration cycle.
cycle and the conventional autocascade refrigeration cycle will also be presented.
2. The autocascade refrigeration cycle with an ejector The conventional autocascade refrigeration cycle (CARC) using two-compound mixed refrigerant is shown in Fig. 1. The CARC system usually consists of a compressor, a condenser, two throttling devices, an evaporator, an evaporative condenser, a phase separator, and connecting tubes. The compressor is used to pump the refrigerant mixtures. The condenser rejects a heat from the inside condensate to the environment cooling media. The evaporator absorbs a refrigerated load from the cold refrigerated space. The evaporative condenser is a counter-flow heat exchanger, which acts as an evaporator for the liquid phase stream from the phase separator and a condenser for the vapor phase stream from the phase separator. The two throttling devices are used to accomplish the expansion process for the cycle, which can be expansion valves or capillary tubes.
The novel autocascade refrigeration cycle (NARC) with an ejector is shown in Fig. 2. Similar to the CARC using twocompound mixed refrigerant, the main components but the ejector in NARC are a compressor, a condenser, a phase separator, an evaporative condenser, an evaporator and two throttling devices. The ejector is set between the evaporative condenser and the evaporator. The basic working process of NARC is same as that of CARC. The refrigerant mixture vapor with initial charge composition (at point 1) is sucked by compressor. After compression to point 2, the refrigerant mixture is partially condensed to point 3 in the condenser. This twophase stream flows into the phase separator, in which the liquid phase at point 30 will have a different composition from the vapor phase at point 300 . Liquid phase has higher mass fraction of the pure refrigerant with higher boiling point and the vapor phase has higher mass fraction of the pure refrigerant with lower boiling point. The liquid phase stream will flow through the throttling device 1, expanding from point 30 to point 7 and then entering the evaporative condenser at a lower temperature (at point 7) to cool the vapor phase stream from the phase separator. After vaporizing in the evaporative condenser, this stream flows further into the ejector as primary flow slightly superheated at point 8 to entrain the vapor (secondary flow) at point 6 from the evaporator. The vapor phase stream from the phase separator condenses on the other side of the evaporative condenser, leaving as saturated liquid or as slightly subcooled liquid at point 4. This fraction of the refrigerant mixture is then expanded through the throttling device 2 to a state with lower pressure and temperature (at point 5). This stream flows into evaporator, vaporizing to realize the refrigerating effect and leaving as saturated or slightly superheated vapor at point 6. Then, the vapor is sucked by the ejector and then the mixed vapor is discharged at high pressure. Eventually, the mixture vapor from the ejector is sucked by the compressor, and it forms a complete cycle. Since the mixture vapor from the evaporative condenser is the primary flow of the ejector, the exit pressure of evaporative condenser should be higher than the exit pressure of evaporator. This also means the pressure drop through the throttling device 1 should be lower than the pressure drop in the throttling device 2. For this reason, the irreversible loss of throttling process in this throttling device will be reduced and the partial usable work can be recovered to lift the evaporator outlet pressure, i.e. increase the suction pressure of the compressor. In addition to this, the heat transfer loss in the evaporative condenser can also be reduced due to the decrease of temperature difference in the NARC compared to that in the CARC. It should be mentioned that the pressure drop in the throttling device 1 has to ensure there exists minimum temperature difference in the evaporative condenser. In general, the cycle performance of NARC can be improved by applying an ejector in it.
3.
Fig. 2 – The autocascade refrigeration cycle with an ejector.
281
Theoretical computation model
The ejector working process is shown in Fig. 3. The primary flow expands in the nozzle from high pressure to low pressure and then sucks the secondary flow. The two flows mix in the mixing section and become one mixed flow. This mixed flow
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where hp;n2;s is the ideal exit enthalpy of the primary flow under the isentropic expansion, hn is the nozzle efficiency. In the mixing section, the momentum conservation equation and energy conservation equation for the ideal mixing process according to constant pressure-mixing process are given as _ s us;n2 ¼ m _pþm _ u _ p up;n2 þ m (3) m s m;m2;s _ s hs;n2 þ u2s;n2 =2 _ p hp;n2 þ u2p;n2 =2 þ m m _ s Þðhm;m2;s þ u2m;m2;s =2Þ _pþm ¼ ðm
ð4Þ
where um;m2;s and hm;m2;s are the ideal exit velocity and enthalpy of the mixed flow under the constant pressure mixing. If neglecting the velocity of secondary flow us;n2 compared with the primary flow velocity up;n2 and considering the energy loss of the mixing process, the actual exit velocity um;m2 and enthalpy hm;m2 of the mixed flow then can be derived from Eqs. (1)–(4) as Fig. 3 – The structure and working process of ejector.
um;m2 ¼ up;n2 ðhm Þ1=2 =ð1 þ mÞ
(5)
hm;m2 ¼ hp;n1 þ mhs;n2 ð1 þ mÞ u2m;m2 =2
(6)
where the mixing efficiency hm is defined as further increases its pressure to the back pressure in the diffuser. In this paper, the ejector performance computation is carried out based on the one-dimensional constant pressure flow model used by most researchers in studying ejector refrigeration cycle. The basic principle of the model was introduced by Keenan et al. (1950) based on gas dynamics, and then developed by Huang et al. (1999) and Ouzzane and Aidoun (2003). Usually, the computation model is set up based on the following main assumptions: (1) The flow inside the ejector is steady and one dimensional. (2) The kinetic energy at the inlets of primary and secondary flow and the outlet of diffuser are negligible. (3) For simplicity, the effects of frictional and mixing losses in the nozzle, diffuser and mixing section are taken into account by using these efficiencies, the nozzle efficiency hn , the mixing efficiency hm and the diffuser efficiency hd . (4) Mixing in the mixing section of ejectors occurs at constant pressure and complying with the conversion of energy and momentum. (5) The inner walls of ejectors are adiabatic, namely no heat losses. (6) Normal transverse shocks may occur at any plane in the throat section. (7) A homogenous two-phase flow is idealized for the ejector when it involves the two-phase flow inside. Using the above assumptions, the equations for the ejector are given based on the mass, momentum and energy conservation as follows. In the nozzle section, if the inlet velocity of primary flow up;n1 is neglected, the exit velocity of primary flow up;n2 and enthalpy hp;n2 can be derived as (1) up;n2 ¼ 2hn ðhp;n1 hp;n2;s ÞÞ1=2 hp;n2 ¼ hp;n1 u2p;n2 =2
(2)
hm ¼ u2m;m2 =u2m;m2;s
(7)
and the entrainment ratio of the ejector m is defined as _p _ s =m m¼m
(8)
In diffuser section, if neglecting the exit velocity of the mixed flow and taking diffuser efficiency into account, the actual exit enthalpy of the mixed flow can be expressed according to energy balance equation as (9) hm;d2 ¼ hm;m2 þ ðhm;d2;s hm;m2 Þ hd where hm;d2;s is the ideal exit enthalpy of the mixed flow under the isentropic compression, hd is the diffuser efficiency. According to the above equations, the entrainment ratio m can be found. qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi (10) m ¼ hn hm hd hp;n1 hp;n2;s ðhm;d2;s hm;m2;s Þ 1 For the whole cycle performance calculation, the following assumptions are also made: (1) Neglect the pressure loss and the heat losses to the environment in the condenser, phase separator, evaporative condenser and evaporator. (2) The vapor stream from the separator is saturated vapor and the liquid stream from the separator is saturated liquid. (3) The flow across the throttling devices is isenthalpic. (4) Take the isentropic efficiency into account for the compressor. Therefore, the equations for the cycle performance can be found based on the energy conservation. Refrigerating capacity is given as _6 Qe ¼ ðh6 h5 Þm
(11)
Refrigerating capacity per unit mass flow of refrigerant mixture in the compressor is given as _1 qo ¼ Qe =m
(12)
international journal of refrigeration 31 (2008) 279–286
The compressor power of the compressor is given as _ 1 =hcom Wcom ¼ h2 h1 m
4.1. (13)
The coefficient of performance (COP) of the cycle can be determined by COP ¼ Qe =Wcom
(14)
_1 m _ 8 are the mass flow where h1–h8 are the enthalpies and m rates of state points 1–8 in the cycle. The corresponding thermodynamic properties of the states of the cycle can be determined from the equation of state of refrigerant mixture. In the calculating program written with Fortran Language, the NIST routines were used to calculate these properties, which is currently an industry standard (NIST Standard Reference Database, 1998).
4.
Results and discussion
In the present study, the refrigerant mixture of R23/R134a is selected as typical working fluid for simulating the cycle performance of NARC. To investigate the characteristics of the NARC, the following basic operating conditions are firstly assumed. The efficiencies of the ejector are assumed to be hn ¼ 0:90, hm ¼ 0:85, and hd ¼ 0:85 by making reference to literatures (Li and Groll, 2005; Huang et al., 1999; Cizungu et al., 2001). The compressor is assumed to have an isentropic efficiency hcom ¼ 0:65, which is a constant and does not vary by the pressure ratio in all cases. In the simulation for the cycle, we selected the temperature, not the pressure, as the given operating parameters for condenser, evaporator and evaporative condenser. For the condenser, it is the condenser outlet temperature tc . The tc used in the simulation is fixed at 40 C. For the evaporator, the evaporator outlet dew point temperature te is used as operating parameter because there is the temperature glide in it. Considering the different refrigeration request, the variable te over a wide range of 20 C to 35 C was performed in the simulation. Similarly, the evaporative condenser outlet dew point temperature tec is also selected as an analytical parameter. Actually, it is undetermined design parameter for the cycle design when the tc and te are determined, which has effects on both the temperature differences in the evaporative condenser and the performance of the ejector. The parametric study must be conducted to determine the optimum value in the simulation for the cycle design. It is known that the composition of refrigerant mixture is an important variable, which has more important effect on the cycle performance. Therefore, the mass fraction of R23 z as the initial charged refrigerant mixture at point 1 is also treated as a variable operating condition. In the simulation, the variation of the cycle performance with the change of z is represented. In addition, in order to investigate the effects of the variation of those main variables on the cycle performance, the degree of superheating and subcooling at evaporative condenser and evaporator outlet is held zero. It is noted that a performance comparison for the NARC and CARC is made under the same operating conditions, which is shown in the following corresponding figures.
283
Effects of the tec on cycle performance
The following results are obtained when the condenser outlet temperature tc ¼ 40 C, the evaporator outlet dew point temperature te ¼ 30 C, the mass fraction of R23 z ¼ 0:35. In the NARC, for a given operating conditions, the entrainment ratio of the ejector m is dependent on the primary pressure, secondary pressure and outlet pressure of ejector. The primary pressure and secondary pressure are, respectively, relevant to tec and te . The outlet pressure of ejector is determined based on meeting the mass conservation constraint for the cycle and the heat balance constraint for the evaporative condenser, which can be evaluated by iteration in the simulation. Fig. 4 shows the variation of m and pressure ratio rpeje of the ejector with the tec , respectively. In this figure, we can see that both the m and rpeje increase with the increases of the tec . The primary pressure of ejector is corresponding to the tec , which increases with the increases of tec . Consequently, this results in an increase of the entrainment ratio and outlet pressure of the ejector. For the given conditions, the rpeje can reach 1.15– 1.31. In addition, it can be found that the m is in the range of 0.80–0.88 in the present NARC, which is close to 1 in values. This is because, as this ejector operates under the given operating conditions in the NARC, the pressure ratios across the ejector are maintained at low level, which are much smaller compared to that of the ejector operated in the conventional ejector refrigeration cycle. Usually, the entrainment ratio of the ejector depends on the primary pressure, secondary pressure and back pressure of ejector. When the primary pressure and secondary pressure are fixed, the value of the entrainment ratio for the ejector increases as the value of the pressure ratio decreases, i.e. the back pressure decreases. Therefore, the ejector can attain a higher entrainment ratio in the NARC, which provides an increase of its performance. It should be noted that the high value of the entrainment ratio in present theoretical simulation was obtained after taking the efficiencies of the ejector into account in the modeling of ejector. Of course, this value could be slightly smaller than it is in real refrigerating machines due to other practical factors.
Fig. 4 – Variation of m and rpeje with the tec .
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Fig. 5 – Variation of rpcom with the tec .
improvement in COP of NARC, which is shown in Fig. 6. It can be seen that for the given conditions, NARC has a 8.7– 17.7% improvement in COP over CARC. This is just because the drop of pressure through the throttling device 1 in the NARC can be reduced compared to the previous throttling device in CARC. For this reason, the ejector can recover some available work of the previous expansion process in the CARC to increase the compressor suction pressure. As viewed from total throttling loss, the throttling loss can be reduced when the ejector is used in the NARC compared to only using the throttling device in CARC. However, since there is a small heat transfer temperature difference across the evaporative condenser of the NARC, the heat transfer area of the evaporative condenser will become larger. Therefore, the limit in the parameter of tec should be moderately considered in the design of the NARC. In addition, Fig. 7 shows the variations of refrigerating capacity per unit mass flow of refrigerant mixture in the compressor with respect to tec . The figure shows the refrigerating capacity of the NARC qo is greater than that of CARC, which is because the refrigerating capacity is very much dependent on the mass flow rate of refrigerant mixture, and that the mass flow rate of refrigerant mixture in the NARC is greater than that in CARC at the same operating conditions. As the tec increases, the refrigerating capacity of the NARC increases owing to the increase of the mass flow rate of refrigerant mixture. Therefore, NARC also represents a merit of performance improvement in refrigerating capacity per unit mass flow of refrigerant mixture in the compressor over CARC.
Moreover, the high value of the entrainment ratio also means that the mass flow rates of refrigerant mixture flowing through the evaporative condenser at both high pressure side and low pressure side are relatively matched. Fig. 5 shows the effect of tec on the pressure ratio with respect to the compressor, respectively, in NARC and CARC. It shows clearly that the pressure ratio of compressor rpcom in NARC deceases with an increase of tec . This results from an increase of the outlet pressure of the ejector, i.e. compressor inlet pressure. Therefore, the rpcom is in turn decreased with an increase of tec . However, the decrease of rpcom is limited by the minimum temperature differences in the evaporative condenser. The minimum temperature difference ðt300 t8 Þ tends to decrease with increasing tec at the given condenser outlet temperature. In Fig. 5, the corresponding ðt300 t8 Þ will exceed 5 C when tec is close to 30 C. From the results shown in Fig. 5, the rpcom of NARC is reduced by 13.1–23.5% compared with that of the CARC. As seen above, the pressure elevation of the ejector causes a decrease of the pressure ratio across the compressor and thus, decreases the compression work and increases the
As we know, the rpcom has a direct relation to the possible lower te when other operating conditions are given. In the cycle, the rpcom will increase with the decreases of the te . This not only gives high pressure ratio risk for compressor, but also significantly increases the loss of the cycle at the same time. Fig. 8 shows the effect of the ejector on the rpcom when decreasing the te . In this simulation, the tc is 40 C and the z is 0.35. The te is kept in a range of 20 C to 35 C. Due to the
Fig. 6 – Variation of COP with the tec .
Fig. 7 – Variation of qo with the tec .
4.2.
Effects of the te on cycle performance
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285
Fig. 10 – Variation of qo with the te . Fig. 8 – Variation of rpcom with the te .
4.3. Effect of the refrigerant mixture composition on cycle performance temperature glide in the evaporator, the corresponding evaporator inlet temperature is in a range of 39.4 C to 52.9 C, which is evaluated by simulation. It can be clearly seen from the figure that the rpcom of the NARC is obviously smaller than that of CARC. As the te further decreases, the rpcom is also decreased more compared to CARC. The reason for this is that when the te becomes lower, the pressure ratio of the ejector increases. Compared with CARC, the rpcom of the NARC decreases by 6.2–24.4% in the above range of te . Therefore, it can be concluded that for the NARC, the large improvement in COP can be obtained with decreasing te . Fig. 9 shows the variation of the COP with the te in the NARC and CARC for the different tec . From this figure, it was found that the NARC improves the COP by 3.9–17.4% compared to the CARC for the above operating conditions. Similar to the situation of performance improvement in COP, it can also be seen as shown in Fig. 10 that the NARC has an improvement in refrigerating capacity over the CARC. Moreover, NARC represents the more significant potential for increasing the refrigerating capacity especially at lower te .
In autocascade refrigeration cycle, the pressure–temperature relationship of the refrigerant mixture in the condenser or evaporator is dependent on the composition. A pressure change at the given temperatures is accompanied by a change of the composition, which affects the ejector performance as well as cycle performance. Therefore, in this case the selection of composition should be more appropriate. As for composition of the refrigerant mixture, the z was selected in a range of 0.15–0.45, so that the maximum operating pressure was below 3 MPa. In the following simulation, the tc and te were, respectively, 40 C and 30 C. The variations of the rpcom in the NARC and CARC with respect to z for different tec are displayed in Fig. 11. It can be seen that for the given conditions, the rpcom in the CARC increases with a decrease of the composition. For the NARC, however, the rpcom decreases with a decrease of the composition. The reason for this case is that the pressure ratio of ejector tends to increase with a decrease of the z, i.e. raise the compressor suction pressure. This in turn leads to a decrease of rpcom in the NARC relative to the total
Fig. 9 – Variation of COP with the te .
Fig. 11 – Variation of rpcom with the z.
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Fig. 12 – Variation of COP with the z.
cycle pressure lift (from evaporator pressure to condenser pressure). It can also be seen that the change tendencies of the rpcom in the NARC with decreasing the composition are same at different tec . However, the possible minimum rpcom accomplished by decreasing the composition and increasing the tec is limited as the constraint of the minimum temperature difference for the evaporative condenser. Compared with CARC, the pressure lift ratio of the ejector reaches 1.13– 1.35, the rpcom of the NARC decreases by 11.4–25.8% in the above range of the z and tec . When the rpcom is decreased, this causes the compression work to decrease. As a result, the COP increases while refrigeration request was kept constant, which can be seen in Fig. 12. Fig. 12 also shows the variation of COP with respect to the z for various tec . It can be derived from Fig. 12 that COP of NARC has higher values than that of CARC because of the function of ejector. It can also be seen that the COP has a large trend of increase for NARC as the z decreases and the tec also increases. Compared with that of CARC, the COP of NARC increases by 7.7–19.1% as shown in Fig. 12.
5.
Conclusion
This paper proposed that an ejector is employed in the autocascade refrigeration cycle to improve the cycle performance. Theoretical computation model based on the constant pressuremixing model for the ejector was used to perform a thermodynamic cycle analysis for this novel autocascade refrigeration cycle with the refrigerant mixture of R23/R134a. The effects of main parameters on the performance for the novel cycle were investigated using the theoretical model. The performance of the novel cycle is then compared with that of the conventional autocascade refrigeration cycle at the same operating conditions. From the above theoretical computation results, the following conclusions are drawn for the novel cycle: (1) On the whole, application of an ejector in autocascade refrigeration cycle can be considered an attractive option for improving cycle performance. Especially, it demonstrates an outstanding merit in decreasing the rpcom as well as
in increasing the COP when operated at lower refrigerating temperature. For example, in the NARC the pressure lift ratio of the ejector reaches to 1.35, the rpcom is reduced by 25.8% and the COP is improved by 19.1% over the CARC for operating at the condenser outlet temperature of 40 C, the evaporator inlet temperature of 40.3 C, the evaporative condenser outlet dew point temperature of 20 C, and the mass fraction of R23 is 0.15. (2) Increasing the tec is always beneficial to performance improvement of the NARC, but is limited by the minimum temperature differences of the evaporative condenser. Obviously, the performance improvement is actually at a cost of increasing the heat transfer area of the evaporative condenser. Therefore, the parameter tec should be synthetically considered in the design of NARC. (3) In NARC, both the compressor discharge pressure and the pressure ratio can be decreased as decreasing the composition of low-temperature refrigerant in two-compound mixture. But the constraint for the minimum temperature difference in the evaporative condenser should be taken into consideration when determining the appropriate composition.
references
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