Particular Characteristics of Transcritical CO2 Refrigeration Cycle with an Ejector

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Applied Thermal Engineering 27 (2007) 381–388 www.elsevier.com/locate/apthermeng

Particular characteristics of transcritical CO2 refrigeration cycle with an ejector

Jian-qiang Deng a, Pei-xue Jiang b,*, Tao Lu b, Wei Lu b

a Department of Process Equipment and Control Engineering, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, China b Key Laboratory for Thermal Science and Power Engineering, Department of Thermal Engineering, Tsinghua University, Beijing 100084, China

Received 4 April 2005; accepted 16 July 2006 Available online 26 September 2006

Abstract

The present study describes a theoretical analysis of a transcritical CO2 ejector expansion refrigeration cycle (EERC) which uses an ejector as the main expansion device instead of an expansion valve. The system performance is strongly coupled to the ejector entrainment ratio which must produce the proper CO2 quality at the ejector exit. If the exit quality is not correct, either the liquid will enter the compressor or the evaporator will be filled with vapor. Thus, the ejector entrainment ratio significantly influences the refrigeration effect with an optimum ratio giving the ideal system performance. For the working conditions studied in this paper, the ejector expansion system maximum cooling COP is up to 18.6% better than the internal heat exchanger cycle (IHEC) cooling COP and 22.0% better than the conventional vapor compression refrigeration cycle (VCRC) cooling COP. At the conditions for the maximum cooling COP, the ejector expansion cycle refrigeration output is 8.2% better than the internal heat exchanger cycle refrigeration output and 11.5% better than the conventional cycle refrigeration output. An exergy analysis showed that the ejector expansion cycle greatly reduces the throttling losses. The analysis was also used to study the variations of the ejector expansion cycle cooling COP for various heat rejection pressures, refrigerant temperatures at the gas cooler exit, nozzle efficiencies and diffuser efficiencies. Ó 2006 Elsevier Ltd. All rights reserved.

Keywords: Refrigeration system; Transcritical cycle; CO2; Ejector

1. Introduction The throttling loss can be reduced by staged expansion, internal heat exchangers or a work-generating expansion. During the expansion of a refrigerant in a throttling pro- In principle, a low cost ejector with no moving parts is also cess, much friction heat is dissipated to the refrigerant due an attractive alternative for the expansion device in the to the large kinetic energy increases as the refrigerant pres- transcritical CO2 cycle. The Denso Corporation in Japan sure decreases. The process is then not isenthalpic, and this in 2004 stated that the coeﬃcient of performance (COP) throttling loss reduces the refrigeration eﬀect. In a CO2 of a CO2 transcritical automotive air conditioning with transcritical vapor compression refrigeration cycle, the an ejector was 25% better than the cooling COP of a con- supercritical CO2 is expanded to a subcritical state. The vention vapor compression refrigeration cycle in their throttling loss is greater than with conventional refriger- experiments [1]. ants owing to the higher pressure change during the Relatively little information is available on the use of expansion. ejectors as expansion devices in a vapor compression refrigeration cycle. Kornhauser [2] theoretically analyzed the * Corresponding author. Tel.: +86 010 62772661; fax: +86 010 performance of an ejector expansion refrigeration cycle 62770209. using R-12 as the refrigerant. He found a theoretical cool- E-mail address: [email protected] (P.-x. Jiang). ing COP improvement of up to 21% over the conventional

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Nomenclature

COP coefficient of performance in cooling condition Subscripts ex exergy (kJ kg1) 1,2,...,i cycle states h enthalpy (kJ kg1) 30,5x,5L cycle states m relative exhaust mass flow rate or relative mo- 12,23,... processes between two cycle states tive mass flow rate 1s,3s,4s locations downstream of the isentropic process q specific refrigeration output per unit mixture a referenced zero state in the exergy analysis mass flow rate (kW kg1) c compressor s entropy (kJ kg1 K1) d discharge pressure (high-side) of the compressor t temperature (°C) or ejector diffuser T temperature (K) e evaporator u velocity (m s1) eje ejector V compressor volume displacement ex exergy w specific power (kW kg1) exp expansion valve x vapor quality H heat sink EERC the ejector expansion refrigeration cycle loss total exergy loss VCRC the conventional vapor compression refrigera- L heat source tion cycle n ejector nozzle IHEC the internal heat exchange cycle s vapor–liquid separator

Greek symbols l ejector entrainment ratio g isentropic efficiency/exergy efficiency q density (kg m3) vapor compression refrigeration cycle for an evaporator ture and independent of the heat source temperature. Liu temperature of 15 °C and a condenser temperature of et al. [11] analyzed the influence of an ejector on a transcrit- 30 °C. Harrell and Kornhauser [3] found that the cooling ical CO2 cycle to show that the new cycle effectively COP was improved by 3.9% to 7.6% with R-134a as the improved the transcritical CO2 cycle performance. The refrigerant with a two-phase ejector. Menegay and results illustrated the effects of ejector entrainment ratio Kornhauser [4] developed a bubbly flow tube installed and efficiency on the cycle performance. upstream of the nozzle to reduce the non-equilibrium ther- In the ejector expansion refrigeration cycle, the refriger- modynamic losses in the ejector nozzle. The cycle cooling ant leaving the ejector is divided into a saturated liquid COP with an ejector using the bubbly flow tube was stream and a saturated vapor stream in the vapor–liquid improved by 3.8% over the conventional cycle for standard separator. The ejector entrainment ratio is equal to the conditions with R-12 as the refrigerant. Menegay and Kor- mass ratio of the two streams in a stable system while the nhauser [4] suggested that this result was not as good as mass percent of saturated vapor is equal to the vapor qual- expected so they anticipated more studies of the ejector ity at the ejector exit. Therefore, the ejector entrainment expansion refrigeration cycle. Domanski [5] pointed out ratio is related to the vapor quality, which will be the focus that the ejector efficiency significantly influences the cool- of the analysis in this paper. The optimum ejector entrain- ing COP of the ejector expansion refrigeration cycle. Fan ment ratio will then give the optimum system performance. et al. [6,7] and Wu et al. [8] studied the modified ejector Previous studies have used both theoretical thermodynamic expansion refrigeration cycle with two heat sources. Nak- analyses and experimental research but have not clearly agawa and Takeuchi [9] showed that a longer diverging sec- established the relationship between the ejector entrain- tion in the nozzle increased the nozzle efficiency. Disawas ment ratio and the vapor quality. and Wongwises [10] experimentally investigated the perfor- This work presents a theoretical analysis of the performance of the ejector expansion refrigeration cycle without mance of the CO2 transcritical ejector expansion refrigera- the expansion valve upstream of the evaporator so that the tion cycle (EERC) to identify the thermodynamic evaporator is flooded with the refrigerant. Their tests relationship between the ejector entrainment ratio and showed an improved cooling COP at low heat sink temper- the vapor quality at the ejector exit. The overall perfor- atures relative to the convention cycle with R-134 a as the mance of the ejector expansion cycle is then compared with refrigerant. The motive mass flow rate in the ejector, which that of the conventional vapor compression refrigeration is the flow rate entering the ejector from the gas cooler, was cycle (VCRC) and the internal heat exchange cycle (IHEC) found to be strongly dependent on the heat sink tempera- as a function of the significant parameters. 中国科技论文在线 http://www.paper.edu.cn

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2. Model and (5). The saturated liquid enters the expansion valve and expands to pressure Pe at state (6). The saturated Figs. 1 and 2 present a schematic of the ejector expan- vapor (5) is superheated, then enters the compressor. sion refrigeration cycle with a P–h diagram illustrating The ejector nozzle efficiency and the diffuser efficiency the transcritical cycle. The ejector expansion system were given by Alexis and Rogdakis [12] who assumed that includes a compressor, gas cooler, ejector, vapor–liquid the mixture pressure in the ejector at state (4) differed from separator, expansion valve and evaporator. the evaporator pressure Pe. The present analysis assumes The subcritical CO2 enters the compressor at pressure Ps that the mixture pressure in the ejector is equal to the evap- at state (1) and is compressed isentropically to the high-side orator pressure Pe since this small pressure difference can pressure Pd at state (1s). The real CO2 compression process be neglected in the transcritical CO2 cycle. to the high-side pressure Pd with an isentropic efficiency, gc, The three cycles are compared based on the following ends at supercritical state (2). The supercritical CO2 is then assumptions: cooled in the gas cooler to temperature T3 at state (3). The flow at state (3) enters the ejector nozzle and expands 1. Kinetic energies of the refrigerant at the ejector inlet and 0 to a mixture at state (3 ) with a nozzle efficiency of gn = 0.7, outlet are negligible. with the corresponding isentropic state (3s). The saturated 2. Flow losses in the pipes and heat exchangers are secondary vapor stream enters the ejector at pressure Pe negligible. corresponding to state (7). The two streams mix at constant 3. Unless stated otherwise, the evaporating temperature is pressure in the ejector with the final state of the mixture cor- 5 °C. The refrigerant leaves the evaporator as saturated responding to state (4). The mixture then flows through the vapor. ejector diffuser where it recovers to pressure Ps at state (5x). 4. Unless stated otherwise, the refrigerant temperature at The diffuser is assumed to have a diffuser efficiency of the gas cooler exit, t3,is36°C. gd = 0.8 with the isentropic outlet at state (4s). 5. Mixing occurs at constant pressure in the ejector mixing The stream leaving the ejector flows into the vapor– region with the assumption that the fluid momentum is liquid separator where it is divided into saturated liquid conserved. The pressure is assumed to be equal to the and saturated vapor streams corresponding to states (5L) evaporator pressure, Pe. 6. The heat sink temperature is 35 °C, while the heat source temperature is 27 °C. 7. The system is in thermodynamic equilibrium. All flow processes are analyzed based on their average velocities and temperatures. 8. The vapor is superheated 15 °C in the internal heat exchanger before entering the compressor in the internal heat exchanger cycle. 9. The vapor is not superheated before entering the compressor in the ejector expansion cycle and the conventional cycle.

3. Computer modeling

Fig. 1. Schematic of the ejector expansion system. 3.1. Thermodynamic modeling

The modeling of the ejector expansion cycle is based on P one unit of mixture refrigerant mass in the ejector at state (5x). The ejector entrainment ratio, l, is defined as the ejec- P 3 1s 2 d tor suction mass flow rate at (7) divided by the motive mass flow rate at (3). Therefore, for 1 kg refrigerant mixture in the ejector, the suction mass flow rate is l/(1 + l) kg and the motive mass flow rate is 1/(1 + l) kg. The motive stream enters the ejector and expands to 5L 4s 5x 5 evaporator pressure Pe with a nozzle efficiency defined as: Ps 1 P g ¼ðh h 0 Þ=ðh h Þð1Þ e 6 3s 3' 4 7 n 3 3 3 3s The energy balance between states (3) and (30) is: 0 h 1 2 u 0 ¼ h3 h30 ð2Þ Fig. 2. Schematic of the ejector expansion system P–h diagram. 2 3 中国科技论文在线 http://www.paper.edu.cn

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The analysis further assumes that ﬂuid momentum is Evaporation conserved in the mixing section, ex67 ¼ ex6 ex7 ðT a=T L 1Þqe ð17Þ u30 =ðl þ 1Þ¼u4 ð3Þ The total exergy loss for a unit refrigerant mixture mass The overall energy balance equation can be written as: in the ejector was calculated using:

exloss ¼ ex12 þ ex23 þ exeje þ exexp þ ex67 ð18Þ h3=ð1 þ lÞþh7l=ð1 þ lÞ¼h5x ð4Þ The ejector expansion cycle exergy efficiency was calcu- The energy balance equation between states (4) and (5x) lated using: is: g ¼ 1 ex =w ð19Þ 1 ex loss c u2 ¼ h h ð5Þ 2 4 5x 4 4. Results and discussion The refrigerant mixture recovers pressure in the ejector diffuser with a diffuser efficiency of: 4.1. Matching of the ejector entrainment ratio g ¼ðh h Þ=ðh h Þð6Þ d 4s 4 5x 4 When the ejector expansion cycle is running at steady The adiabatic compressor efficiency is [13]: state, the mass flow rates of the two refrigerant streams flowing into the ejector do not vary. The refrigerant leaving g 1:003 0:121 p =p 7 c ¼ ð d sÞðÞ the ejector is divided into a saturated liquid stream and a h1s h1 saturated vapor stream in the vapor–liquid separator. gc ¼ ð8Þ h2 h1 Therefore, the mass flow ratio of the two streams must be equal to the ejector entrainment ratio in a stable operat- The compressor power consumption per unit mixture ing system and a given set of system working conditions flow mass is: will have a unique ejector entrainment ratio. If the ejector wc ¼ðh2 h1Þ=ð1 þ lÞð9Þ does not have an entrainment ratio that produces the The refrigeration output per unit mixture flow mass is: proper refrigerant quality at the ejector exit, then an unsteady system results and either liquid will accumulate l qe ¼ðh7 h5LÞ ð10Þ with liquid entering the compressor or the separator will 1 þ l be full of refrigerant vapor. In this study, the ejector The cooling coefficient of performance is: entrainment ratio is always specified as the ratio giving a stable running ejector expansion system. COP ¼ q =w ð11Þ e c The computer model predicts a unique ejector entrainment ratio for the specified conditions. The variation of 3.2. Exergy efficiency modeling the ejector entrainment ratio and the ejector exit quality with changes in the high-side pressure for three evaporator The reference zero state (a) is defined as the environment temperatures is shown in Fig. 3. The ejector entrainment temperature, 35 °C, and the high-side pressure. The exergy ratio initially rapidly increases and then levels off with values of all states were calculated based on one unit refrigerant mixture mass in the ejector: 0.8 exi ¼½ðhi haÞT aðsi saÞ mi ð12Þ 0.6 0 while i = 1,2,3,3 ,5, mi = 1/(1 + l) while i = 5L,6,7, mi = l/(1 + l) 0.4 ° Ratio T = 0 C while i = 4,5x, m =1 e i ° The exergy losses in the processes were calculated using: 0.2 Te = 5 C ° Compression Ejector Entrainment 0.9 Te = 10 C ex12 ¼ðex1 ex2Þþws ð13Þ 0.8 Heat rejection 0.7 ex23 ¼ ex2 ex3 ð14Þ

Vapor Quality 0.6 Ejection 8 9 10 11 12 exeje ¼ ex3 þ ex7 ex5x ð15Þ High-side Pressure (MPa) Throttling Fig. 3. Entrainment ratio and vapor quality variations for various compressor discharge pressures and evaporator temperatures in the exexp ¼ ex5L ex6 ð16Þ ejector expansion cycle. 中国科技论文在线 http://www.paper.edu.cn

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recovery process. Disawas and Wongwises [10] found the 5.6 same phenomenon in their tests for compressor speeds over (MPa) 5.2

5x 450 rpm which increased the high-side pressures resulting 4.8 in a higher ejector entrainment ratio, Fig. 3. This higher 4.4 ratio then leads to a lower vapor–liquid separator pressure, 4.0 Fig. 4. 3.6 ° Te = 0 C Fig. 5 shows the variation of the cooling COP with the

C) Pressure P 24

° ° Te = 5 C ejector entrainment ratio for the various evaporator tem- ( 20 5x ° peratures. The results show an optimum value of the cool- Te = 10 C 16 ing COP at an ejector entrainment ratio between 0.5 and 12 0.6. If the entrainment ratio is too low, the COP will 8 decrease to near 1.0 even though the system is stable. The 4 results in Fig. 5 show that the ejector must be carefully Temperature t 0.1 0.2 0.3 0.4 0.5 0.6 0.7 matched with the ejector expansion system. For a given Ejector Entrainment Ratio ejector with a fixed entrainment ratio, the system parameters must be adjusted to fit the ejector characteristics which Fig. 4. Influence of ejector entrainment ratio on the pressures and temperatures at the ejector exit in an ejector expansion cycle. may not give the optimum COP. Disawas and Wongwises [10] adjusted the compressor speed to fit the ejector. He maintained that the appropriate compressor speed is increasing high-side pressure. The quality moves in the 450 rpm by referring to former operating experience on opposite direction due to the relation: the ejector expansion cycle. At compressor speeds lower 1 than 450 rpm, the amount of liquid in the separator grad- x ¼ ð20Þ l þ 1 ually increased. Eventually, the liquid refrigerant flooded the separator outlet and flowed to the compressor, which The variations of the pressure and temperature in the results in compressor failure. vapor–liquid separator for various ejector entrainment ratios for three evaporator temperatures are shown in 4.2. Comparison of the three refrigeration cycles Fig. 4. The pressure and temperature in the vapor–liquid separator decrease slowly with increasing ejector entrain- The performances of the CO transcritical ejector expan- ment ratio to minimums and then slowly increase for ejec- 2 sion cycle, internal heat exchanger cycle and a conventional tor entrainment ratios greater than about 0.6. Therefore, cycle were compared using the model described in Section 3. considering the results in both Figs. 3 and 4, for higher Fig. 6 shows the variation of the COP with high-side pres- high-side pressures, lower pressures and temperatures can sure in the three cycles. The three cycles all have an be more easily maintained in the vapor–liquid separator optimum high-side pressure corresponding to a maximum because a higher ejector entrainment ratio means more sat- COP. However, the maximum COP for the ejector expan- urated vapor flow at the evaporator pressure and tempera- sion system is 18.6% higher than the maximum COP for the ture exhausted into the ejector to facilitate the mixing and internal heat exchanger system and 22.0% higher than that

5 4.0 COPmax EERC IHEC 3.5 4 VCRC

3.0 3 2.5

2 2.0 ° Te = 0 C T = 5°C e 1.5 Coefficient of Performance 1 ° Coefficient of Performance Te = 10 C 1.0 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 8 9 10 11 12 Ejector Entrainment Ratio High-side Pressure (MPa)

Fig. 5. Cooling COP for various entrainment ratios and evaporator Fig. 6. Cooling COP of the ejector expansion cycle, internal heat temperatures. exchanger cycle and conventional cycle for various high-side pressures. 中国科技论文在线 http://www.paper.edu.cn

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0.11 1.4

0.10 EERC IHEC 1.2 0.09 VCRC

0.08 1.0

0.07 0.8 0.06 0.6 EERC 0.05 Exergy Efficiency IHEC 0.04 Refrigeration Output Ratio 0.4 VCRC

0.03 0.2 0.02 8 9 10 11 12 8 9 10 11 12 High-side Pressure (MPa) High-side Pressure (MPa)

Fig. 7. Exergy efficiency of the ejector expansion cycle, internal heat Fig. 8. Refrigeration output of the ejector expansion cycle, internal heat exchanger cycle and convention cycle for various high-side pressures. exchanger cycle and conventional cycle for various high-side pressures. for the conventional system. Fig. 7 shows that the ejector The refrigeration output is the important performance expansion system has highest exergy efficiency for most for the refrigeration system. In the ejector expansion sys- of the high-side pressure range. tem, the refrigerant into the compressor is at Ps which is The exergy losses in each process in the three cycles are higher than the evaporator pressure, Pe, so the CO2 density listed in Table 1 for a high-side pressure of 8.7 MPa which entering the compressor is higher than that entering the results in the maximum COP for all three cycles. The other compressor in the other two cycles. Therefore, in the ejec- parameters were given in Section 2. The ejector entrain- tor expansion system, the compressor compresses more ment ratio is 0.547 for the ejector expansion cycle. The mass. More importantly, the refrigerant compressed by exergy losses in the ejector expansion cycle are based on the compressor is not directly providing cooling because a unit CO2 mass flow rate compressed by the compressor only the liquid fraction of the refrigerant flow from the (instead of the unit mass flow of mixture in the ejector). ejector flows through the evaporator. Therefore, for the As listed in Table 1, the throttling exergy loss in the con- three cycles can be compared for the same compressor conventional cycle is 12.8 kJ kg1, 34.3% of the total exergy ditions by comparing the refrigeration output: loss. However, the throttling exergy loss in the ejector Q ¼ V qw COP ð21Þ expansion system is only 0.39 kJ kg1 with the ejection exergy loss of 8.1 kJ kg1, the sum of the two losses is The equivalent refrigeration output of the three cycles 29.7% of the total system exergy loss. The exergy loss in for various high-side pressures are showed in Fig. 8 with the ejection process in the ejector expansion system is due the refrigeration output ratio defined as the refrigeration to a process that is equivalent to the throttling losses in output relative to the capacity of the conventional cycle the other cycles. The exergy losses in the compression at an evaporator temperature of 5 °C and a high-side pres- and heat rejection processes are also reduced some in the sure of 10 MPa. For refrigeration output defined at an ejector expansion system. The exergy loss in the evapora- evaporator temperature of 5 °C and a high-side pressure tion process is the largest loss in the ejector expansion of 10 MPa, where the density at the compressor inlet is system. 114.4 kg m3, the specific power is 51.1 kJ kg1, and the

Table 1 Exergy losses in all three cycles for a high-side pressure of 8.7 MPa Process VCRC IHEC EERC Loss ( kJ kg1) (%) Loss ( kJ kg1) (%) Loss ( kJ kg1) (%) Compression 9.56 25.69 10.74 24.38 6.79 23.73 Heat rejection 5.50 14.76 9.75 22.13 4.30 15.02 Ejection – – – – 8.12 28.38 Throttling 12.77 34.29 10.03 22.78 0.39 1.36 Evaporation 9.40 25.25 11.50 26.10 9.02 31.51 IHE – – 2.03 4.60 – – Total 37.229 100.0 44.04 100.0 28.64 100.0 Speciﬁc power 40.31 – 47.82 – 31.64 – 中国科技论文在线 http://www.paper.edu.cn

J.-q. Deng et al. / Applied Thermal Engineering 27 (2007) 381–388 387 cycle COP is 2.62. The compressor volume displacement is Fig. 10 shows the variation of the ejector expansion assumed to be constant. The ejector expansion cycle refrig- cooling COP with high-side pressure for various CO2 tem- eration output is distinctly greater than the other two cycles peratures at the gas cooler exit. The COP increases and the for high-side pressures greater than 8.5 MPa. For example, optimum high-side pressure corresponding to the maxi- for the high-side pressure of 8.7 MPa corresponding to the mum COP decreasing as the CO2 temperature approaches maximum COP, the ejector expansion refrigeration output the heat sink temperature. For the same evaporator tem- is 8.2% higher than the internal heat exchanger cycle refrig- perature of 5 °C, the COP at a gas cooler exit temperature eration output and 11.5% higher than the conventional of 36 °C is 23.0% more than at 40 °C. cycle refrigeration output. Fig. 11 shows the variation of the COP with the ejector diffuser efficiency for various high-side pressures. The COP 4.3. EERC performance increases with increasing diffuser efficiency, but the influence of the high-side pressure on the COP is much more Fig. 9 shows the variation of the ejector expansion cycle significant than the effect of the diffuser efficiency. Fig. 12 cooling COP for various high-side pressures and evapora- shows the variation of the COP with the ejector nozzle tor temperatures of 0, 5 and 10 °C. The COP increases with efficiency for various high-side pressures. As with the dif- evaporator temperature, with a maximum COP at 5 °C fuser efficiency, the COP increases as the nozzle efficiency that is 20.6% higher than the maximum at 0 °C. increases, but the effect is not significant.

5 4 ° Te = 0 C COPmax ° Te = 5 C 3 ° 4 Te = 10 C

Performance 2 Coefficient of Pd = 8.0 MPa Pd = 9.0 MPa 3 1 Pd = 10.0 MPa 0.8

0.6 2 Coefficient of Performance 0.4 Ratio

0.2

1 Ejector Entrainment 78910111213 0.5 0.6 0.7 0.8 0.9 High-side Pressure (MPa) Ejector Diffuser Efficiency

Fig. 9. Ejector expansion cycle cooling COP for various high-side Fig. 11. Ejector expansion cycle cooling COP for various ejector diﬀuser pressures and evaporator temperatures. eﬃciencies and high-side pressures.

4 COPmax 4

3 3

2 P = 8.0 MPa

Performance d Coefficient of P = 9.0 MPa 2 1 d P = 10.0 MPa 0.8 d T = 36°C 3 0.6 ° 1 T3 = 38 C

Coefficient of Performance ° 0.4 T3 = 40 C Ratio 0.2

0 Ejector Entrainment 8 9 10 11 12 13 0.5 0.6 0.7 0.8 0.9 High-side Pressure (MPa) Ejector Nozzle Efficiency

Fig. 10. Cooling COP for various high-side pressures and CO2 temper- Fig. 12. Ejector expansion cycle cooling COP for various ejector nozzle atures at the gas cooler exit. eﬃciencies and high-side pressures. 中国科技论文在线 http://www.paper.edu.cn

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Most prototypes of CO2 systems have used small diam- The ejector expansion cycle has not been extensively eter or flat multiport tubes in the evaporators to handle the studied even though the present study of the transcritical high pressure without adding heat exchanger weight. How- CO2 ejector expansion cycle showed that the system perfor- ever, these evaporators face the challenging problem of mance can be improved by 25% relative to the conventional how to distribute the developing two-phase flow from the cycle which agrees with the experimental results given by header uniformly into many tubes. In a conventional cycle, the Denso Corporation [1]. However, many other experi- the void fraction at the evaporator inlet exceeds 0.8 after mental tests have obtained no more than 8% improvement about 6 MPa of throttling. But in the ejector expansion perhaps because they did not use an optimum ejector cycle, the pressure difference due to throttling is only about design matched to the system operating conditions. Fur- 2 MPa, so the mass flashed into vapor is greatly reduced ther study of the ejector expansion refrigeration cycle is and the two-phase flow can be more easily distributed. needed to experimentally verify these results. Therefore, the evaporator tube arrangement is not as crit- ical in the ejector expansion cycle. Acknowledgement

5. Conclusions This work was supported by the Beijing City Science Plan Project in China No. H021820041420. The COP and the refrigeration output of the transcritical CO2 ejector expansion cycle were analyzed based on a References detailed thermodynamic model of the system. The ejector expansion cycle performance was then compared with that [1] H. Yuhala, J. Wang, et al., Denso Corporation in Japan, Private Communication, Tsinghua University in Beijing, China, 2004. of the internal heat exchanger system and a conventional [2] A.A. Kornhauser, The use of an ejector as a refrigerant expander, in: vapor compression system. The analyses led to the follow- Proceeding of the 1990 USNC/IIR-Purdue Refrigeration Conference, ing conclusions: 1990, pp. 10–19. [3] G.S. Harrell, A.A. Kornhauser, Performance tests of a two-phase 1. The ejector expansion cycle has a unique ejector entrain- ejector, in: Proceedings of the 30th Intersociety Energy Conversion Engineering Conference, Orlando, FL, 1995, pp. 49–53. ment ratio for a steady running system with a given set [4] P. Menegay, A.A. Kornhauser, Improvements to the ejector expan- of working conditions. The system also has an optimum sion refrigeration cycle, in: Proceedings of the 31th Intersociety ejector entrainment ratio which gives the maximum sys- Energy Conversion Engineering Conference, Washington, DC, 1996, tem cooling COP. pp. 702–706. 2. In transcritical CO refrigeration systems, the ejector [5] P.A. Domanski, Theoretical evaluation of the vapor compression 2 cycle with a liquid-line/suction-line heat exchanger, economizer, and expansion cycle cooling COP is higher than that of the ejector, Nistir-5606, National Institute of Standards and Technology, internal heat exchanger cycle and the conventional cycle. March, 1995. For the working conditions described in this paper, the [6] X.W. Fan, J.M. Yin, H.F. Liu, Z.Q. Chen, Study on a new kind of ejector improves the maximum COP by up to 18.6% vapor-compression/ejection hybrid refrigeration cycle, Journal of compared to the internal heat exchanger system and Xi’an Jiaotong University 30 (5) (1996) 5–11 (in Chinese). [7] X.W. Fan, J.M. Yin, Z.Q. Chen, Experimental study on a new kind of by 22.0% compared to the conventional system. compression/ejection refrigeration cycle, Journal of Xi’an Jiaotong 3. In the transcritical CO2 ejector expansion cycle, the University 31 (5) (1997) 6–9 (in Chinese). throttling exergy loss is much less than in the internal [8] J.T. Wu, Z.G. Liu, H. Guo, Q.S. Fu, Thermodynamic analysis of a heat exchanger cycle and in the conventional cycle. new kind vapor-compression/ejection refrigeration cycle, Fluid The exergy losses due to the compression and heat rejec- Machinery 28 (8) (2000) 47–49 (in Chinese). [9] M. Nakagawa, H. Takeuchi, Performance of two-phase ejector in tion processes are also reduced some. The evaporation refrigeration cycle, in: Proceedings of the Third International Con- exergy loss is not changed much. ference on Multiphase flow, Lyon, France, 8–12 June 1998, pp. 1–8. 4. The ejector expansion cycle refrigeration output is [10] S. Disawas, S. Wongwises, Experimental investigation on the improved by up to 8.2% compared to the internal heat performance of the refrigeration cycle using a two-phase ejector as exchanger cycle refrigeration output and by 11.5% com- an expansion device, International Journal of Refrigeration 27 (6) (2004) 587–594. pared to the conventional cycle refrigeration output for [11] J.P. Liu, J.P. Chen, Z.J. Chen, Thermodynamic analysis on CO2 the conditions used in this analysis. transcritical vapor-compression/ejection refrigeration cycle, Journal 5. The ejector expansion cycle performance is very sensi- of Shanghai Jiaotong University 38 (2) (2004) 273–275 (in Chinese). tive to the operating conditions. This paper compares [12] G.K. Alexis, E.D. Rogdakis, A verification study of steam-ejector the influences of high-side pressure, evaporator temper- refrigeration model, Applied Thermal Engineering 23 (1) (2003) 29– 36. ature, CO2 temperature at the gas cooler exit, ejector [13] S.M. Liao, T.S. Zhao, A. Jakobsen, Correlation of optimal heat entrainment ratio, nozzle efficiency and diffuser effi- rejection pressures in transcritical carbon dioxide cycles, Applied ciency on the cycle cooling COP. Thermal Engineering 20 (9) (2000) 831–841.