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Investigation of potential benefits of compressor cooling
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Applied Thermal Engineering 28 (2008) 1791–1797 www.elsevier.com/locate/apthermeng
Investigation of potential benefits of compressor cooling Xudong Wang, Yunho Hwang *, Reinhard Radermacher Center for Environmental Energy Engineering, Department of Mechanical Engineering, University of Maryland, College Park, MD 20742, USA Received 22 November 2006; accepted 12 November 2007 Available online 22 November 2007
Abstract The compressor is certainly the largest power consumer in a vapor compression system. To reduce the power consumption of the compressors two performance improving options are investigated theoretically for refrigerants R22, R134a, R410A and R744 as working fluids. The first option is cooling the motor by external means other than using the suction gas. Analysis results for this option show that R22, R410A and R744 have larger potential benefits than R134a. In low temperature refrigeration applications larger improvements are achievable than in air conditioning applications for all four refrigerants. The second option is to make the compression process isothermal by transferring heat from the compression chamber. To approach the isothermal compression process while avoiding wet compression, two cases of combining isothermal and isentropic compression processes are analyzed. The analysis results show that this strategy can reduce the compression work up to 14% as compared to the isentropic compression process for the R22 refrigeration system at ASHRAE T condition. The second analysis on the ideal vapor compression cycles using such compressor cooling strategy show that the compression power of the system can be reduced by up to about 16% depending on operating conditions and fluid choice. Ó 2007 Elsevier Ltd. All rights reserved. Keywords: Compressor; Cooling; Isothermal compression; Isentropic compression
1. Introduction Compressor performance significantly affects refrigeration system performance. Heat transfer to the suction gas through the suction line, compressor shell and compressor motor eventually degrades the performance of the compressor [1]. Effective cooling of the compressor motor may reduce this adverse effect. Dreiman [2] reported that about 2% coefficient of performance (COP) increase was achieved by utilizing the vortex tube phenomena to cool the motor stator winding. Nowadays, for conventional compressor technology, the accepted motor efficiencies range from approximately 88% for a 2 kW motor to 95% for a 75 kW motor and the mechanical efficiency is in the same order of magnitude [3]. Opportunities for significant improvements are possible only when the compression process is changed, as an example, to isothermal compression.
*
Corresponding author. Tel.: +1 301 405 5247; fax: +1 301 405 2025. E-mail address: [email protected] (Y. Hwang).
1359-4311/$ - see front matter Ó 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2007.11.010
Isothermal compression is a process in which the temperature of the working fluid remains constant during the compression process. However, it is impossible for refrigerants with a high specific heat capacity to be compressed isothermally over the entire compression process [4]. It is because refrigerants with a large molecular mass and therefore, large specific heat capacity, becomes two-phase state before they reach the condensing pressure [5], and leads to wet compression. Furthermore, a large amount of heat must be removed during the compression process by some means to maintain a constant temperature of the working fluid. Coney et al. [6] developed an air compressor which could approach quasi-isothermal compression via water injection, and reported that 28% savings of compression work could be achieved practically at a pressure ratio of 25. However, water injection is not favored option to current commercial compressors being utilized in the field of air conditioning and refrigeration. A technique similar to the water injection and applicable to current commercial compressors is refrigerant injection. The basic idea is to inject the refrigerant in either liquid [7–13] or vapor phase
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X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
Nomenclature ARI COP h m P Q RPM v Vdis
Air Conditioning and Refrigeration Institute coefficient of performance enthalpy (kJ/kg) refrigerant mass flow rate (kg/s) pressure (kPa) heat transfer amount per unit mass (kJ/kg) revolutions per minute specific volume (m3/kg) displacement volume (m3)
[10,14–23] to the sealed compression chamber to reduce the discharge temperature and to improve the system performance. However, both liquid and vapor refrigerant injection eventually increase the compression work since an extra amount of refrigerant is injected and compressed, which requires additional compression work. Since the existing technologies of refrigerant injection to the compressor are not able to achieve an isothermal compression process, it is warranted to further investigate isothermal compression process and alternative cooling concepts. The purpose of this paper is to outline the potential benefits of compressor cooling in an ideal vapor compression cycle through theoretical analysis. While the aforementioned research focuses on the compression process with refrigerant injection to the sealed compression chamber, the current work is focused on the compression process without any refrigerant injection to the sealed compression chamber so that the compressor does not need extra power consumption associated with refrigerant injection. The benefit of changing the compression process from isentropic to isothermal and the heat transfer amount required for such change are presented to provide a quantified guideline for alternative compressor cooling concepts.
W Wisen Wiso
W gm gvol q
compression work per unit mass (kJ/kg) isentropic compression work per unit mass (kJ/ kg) isothermal compression work per unit mass (kJ/ kg) power consumption (kW) motor efficiency volumetric efficiency density (kg/m3)
air or by other means instead of the suction gas, the performance of the compressor would be improved. To demonstrate the effect of eliminating suction gas preheating, the performance of ideal cycles with and without suction gas preheating is carried out. The two cycles, superimposed on a pressure–enthalpy (P–h) diagram, are shown in Fig. 1. In this analysis, it is assumed for the base cycle (cycle 10 –20 –3–4) that the suction gas is only preheated by the motor loss from state point 1 to state point 10 , and the compression process is isentropic process. In the second cycle (cycle 1–2–3–4), the motor is cooled presumably by external means other than using the suction gas. Eq. (1) is used to calculate the refrigerant mass flow rate from the predetermined variables: volumetric efficiency, suction density, displacemental volume, and compressor rotational speed. The suction gas state after preheating (state point 10 ) can be found by combining the Eqs. (2) and (3) from the predetermined variables: motor efficiency, suction enthalpy before preheating (h1), and discharge pressure. Then Eq. (2) is used to calculate the compression work. The heat absorbed by the suction gas is calculated by Eq. (3)
m ¼ gvol q1 V dis RPM=60
W gm ¼ mðh20 h10 Þ 2. Analysis of compressor cooling options 2.1. External motor cooling Once the suction gas leaving the evaporator enters the compressor shell, it is heated by various sources since the compressor housing temperature is generally much higher than the suction gas temperature. The suction gas preheating is caused by the electrical loss from the motor, the mechanical loss from the frictional parts, and the thermal losses caused by the heat transfer from the compression chamber and the internal refrigerant leakage. Suction gas preheating eventually results in a decrease of the suction gas density. Consequently, the refrigerant mass flow rate and the volumetric efficiency are reduced. Therefore, energy required to maintain the same mass flow rate is also increased due to the suction gas preheating [1]. If the motor could be effectively cooled by the forced convention with
W ð1 gm Þ ¼ mðh10 h1 Þ
ð1Þ ð2Þ ð3Þ
Other conditions used in the analysis are as follows: a 54.2 cc displacement volume, a volumetric efficiency of 90%, a motor efficiency of 90%, and a compressor RPM
Fig. 1. Effects of external motor cooling shown in P–h diagram.
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
of 3600. Refrigerants R22, R134a, R410A and R744 are studied as examples for typical refrigerants. The system performance is evaluated under Air Conditioning and Refrigeration Institute (ARI) conditions for air conditioning and low temperature refrigeration applications [24] as summarized in Table 1. Since R744 cycle is a transcritical cycle, the compressor discharge pressure is higher than its critical pressure, the discharge pressures of 9 MPa and 10 MPa are chosen as an example for low temperature refrigeration application and air conditioning application, respectively. The performance enhancement by the external motor cooling is illustrated in Fig. 2. The results show that the system capacity and COP can be improved when the motor is cooled by the external means instead of the suction gas. The refrigerants R22, R410A and R744 have better potential benefits than R134a. The low temperature refrigeration application can achieve more improvement than the air conditioning application. The largest COP improvement of 5.4% is observed for an R744 system at ARI low temperature refrigeration application. The capacity of the R744 system improved up to 5.5% for the same condition. However, cooling motor has little effect in reducing the power consumption. 2.2. Isothermal compression with active compressor cooling To reduce the compressor power consumption a more active compressor cooling is investigated. In Fig. 3, two ideal compression processes, isentropic process (1–2) and isothermal process (1–20 ), are shown in a pressure–specific volume (P–v) diagram. A closed area 1–2–3–4–1 represents the isentropic compression work. Similarly, the enclosed area 1–20 –3–4–1 represents the isothermal compression
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Fig. 3. Isentropic and isothermal compression processes in P–v diagram.
work. Isothermal compression requires less work input than the isentropic compression for a given pressure ratio. The shaded area 1–2–20 –1 represents the work saving by changing the compression process from isentropic to an isothermal process. To approach the isothermal compression a certain amount of heat must be removed from the compression process. Such amount of heat can be calculated based on the energy balance of the compression process. When the isothermal compression is approached the wet compression could happen. In order to approach isothermal compression while avoiding the wet compression, a combination of partial isothermal compression and partial isentropic compression is proposed by applying effective cooling techniques to different compression stages. Two possible options under this strategy are illustrated in Fig. 4. It shows only the compression process of the refrigeration cycle in P–h diagram. As an example, refrigerant R22 is selected as the working fluid and the cycle is assumed to be operating at ASHRAE T condition [25], which is referred as 7.2 °C, 54.4 °C, 35 °C, and 8.3 °C for saturated evaporating temperature, saturated condensing
Table 1 ARI conditions [24] Conditions
Temperature (°C)
Air conditioning Low temperature refrigeration
Ambient
Evaporating
Suction
Condensing
Liquid
35 35
7.2 23.3
18.3 4.4
54.4 48.9
46.1 40.6
ARI-Air Conditioning Application Capacity
ARI-Refrigeration Application Electric Power
COP
Capacity
COP
3.0% 2.5% 2.0% 1.5% 1.0% 0.5% 0.0% -0.5%
R22
R134a
R410A
Refrigerants
R744
Change of Performance
Change of Performance
Electric Power
5.5% 4.5% 3.5% 2.5% 1.5% 0.5% -0.5%
R22
R134a
R410A
Refrigerants
Fig. 2. Effects of external motor cooling on system performance.
R744
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X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
Fig. 4. Combinations of isothermal and isentropic compressions in P–v diagram.
temperature, return gas temperature, and degree of subcooling, respectively. The base compression process for both cases is the isentropic compression 1–20 . In case I, the effective cooling is applied to the compressed R22 only during the first compression process 1–2. Once the R22 reaches its saturated vapor state isothermally, the cooling is stopped. Then, the second compression process follows the isentropic process 2–3. This case requires for the suction gas to have some degree of superheating. In case II, the R22 is isentropically compressed first to an intermediate pressure from state 1 to 200 . Then it undergoes an isothermal compression process from state 200 to 300 by an external cooling. The two compression cases and the baseline, isentropic compression, are illustrated in P–v diagram (Fig. 5). The isothermal compression work is calculated by integrating vdP along the isotherm in the P–v diagram. For
2250
2147 kPa 3" 3
2'
Pressure (kPa)
2000 1750 Baseline case: isentropic process
example, the compression work and the heat transfer amount in case I are calculated by Eqs. (4)–(7) Z P2 W iso ¼ v dP ð4Þ P1 Z P2 v dP ð5Þ Q ¼ ðh1 h2 Þ þ P1
W isen;2–3 ¼ h3 h2 W isen;1–20 ¼ h20 h1
ð6Þ ð7Þ
The comparison of compression work and heat transfer amounts for these two cases and isentropic compression process is summarized in Table 2. The results show that the case I and the case II can reduce the compression work up to 14.0% and 8.5%, respectively, as compared to the isentropic compression process at ASHRAE T condition. Since the case I needs to remove less heat, and saves more work than the case II, it is preferred to the case II in the performance enhancement point of view. However, in the case I it must be assured that the temperature of cooling sources is below the suction temperature, which may be a considerable challenge. For the case II, part of the condenser can possibly be used as the heat sink, or an additional heat exchanger may be applied to achieve the process.
1500 1355 kPa 2
1250
Table 2 Comparison of three compression cases at ASHRAE T condition
Case I 1000
917 kPa
Case II 2"
750
1
626 kPa
Compression process
Base case (1–20 )
Case I (1–2–3)
Compression work (kJ/kg)
35.6
1–2 2–3 Total
19.7 10.9 30.6
1–2 2–3 Total
10.6 22.0 32.6
Removed heat (kJ/kg)
0.0
1–2 2–3 Total
32.9 0.0 32.9
1–2 2–3 Total
0.0 43.4 43.4
500 0
0.01
0.02
0.03
Specific volume
0.04
(m3/kg)
Fig. 5. Three compression processes in P–v diagram.
0.05
Case II (1–200 –300 )
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
The external motor cooling is found to be not a very effective means to reduce the power consumption. This is because the increase of the mass flow rate balances out the decrease of the power consumption per unit of mass flow by avoiding the suction gas preheating. However, the increase of the mass flow rate results in an increase in the capacity. Then the improvement of the COP is dominated by the improvement of the system capacity instead of the reduction of the power consumption. Contrastingly, the isothermal compression with active compressor cooling appears to be an effective means to reduce the compressor power consumption, and the system COP improvement comes from the power reduction. To further demonstrate the benefit of the isothermal compression with active compressor cooling, the compression process of case II in Figs. 4 and 5 are analyzed for multiple refrigerants, R22, R134a, R410A and R744, for both an air conditioning application and a low temperature refrigeration application under ARI conditions [24]. The discharge pressures of 9 MPa and 10 MPa are chosen for R744 system as an example for low temperature refrigeration application and air conditioning application, respectively. The discharge temperature of isothermal compression varies depending upon the temperature and capacity of the external cooling sources. Fig. 6 illustrates how the degree of external cooling affects the cycles and discharge temperature in P–h and P–v diagrams. The discharge temperature affects the compressor performance. The power savings and heat transfer amounts needed for various discharge temperatures are illustrated in Figs. 7 and 8, respectively for case II under ARI conditions. From Fig. 7, one can observe that the combined isentropic and isothermal compression can reduce the compression power of the system by up to about 16% for the case of R744 at low temperature refrigeration application. For all
18
Power Reduction [%]
3. Results and discussion
four refrigerants investigated the reduced discharge temperature results in increased power reduction. This means that the earlier the isothermal compression starts, the more energy savings one can actually obtain from this process. Since the discharge temperatures for all four refrigerants at low temperature refrigeration application is much higher than those at air conditioning application, a better performance improvement in low temperature refrigeration application can be obtained than in air conditioning application for all four refrigerants. Moreover, there is a significant
R134a-Air conditioning
R134a-Refrigeration
16
R22-Air conditioning
R22-Refrigeration
14
R410A-Air conditioning
R410A-Refrigeration
R744-Air conditioning
R744-Refrigeration
12 10 8 6 4 2 0 45
55
65
75
85
95
105
115
Discharge Temperature [°C] Fig. 7. Power reduction vs. discharge temperature for case II.
140 120
Heat Transfer [kJ/kg]
Since the current study is focused on the theoretical analysis, the uncertainty source only comes from the uncertainties of the refrigerants properties. The uncertainties of the refrigerants enthalpy and specific volume are less than 1% and 0.5%, respectively [26].
1795
100
R134a-Air conditioning
R134a-Refrigeration
R22-Air conditioning
R22-Refrigeration
R410A-Air conditioning
R410A-Refrigeration
R744-Air conditioning
R744-Refrigeration
80 60 40 20 0 45
55
65
75
85
95
105
115
Discharge Temperature [°C] Fig. 8. Heat transfer amount vs. discharge temperature for case II.
Fig. 6. Compression processes with different discharge temperatures for case II.
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X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
improvement potential for R744 in low temperature refrigeration application since it has a high discharge temperature and small molecular mass. The refrigerants with smaller molecular mass benefit more from compressor cooling than those with larger molecular mass. This is because the saturated vapor line of the refrigerant with smaller molecular mass has a steeper slope in the P–h diagram because of their smaller specific heat capacity. Therefore, a larger portion of the isentropic process can possibly be switched to an isothermal compression process at the latter part of the compression process while avoiding wet compression. R134a has the least benefit among four refrigerants investigated. This result is not surprising since R134a has the lowest discharge temperature among those refrigerants. Therefore, the potential for performance improvement is more limited for R134a as compared to other three refrigerants. Applying isothermal compression either in the case I or II does not make any difference in the system capacity since the suction condition is assumed to be the same. The COP improvement comes from the power reduction. Therefore, the COP changes follow the same magnitude and trend as the power does. From an energy balance point of view, increased power savings mean more heat needs to be removed from the compression process. Consequently, the amount of heat transferred from the compressor cylinder follows the same trend as the power savings. When the two cases are compared, the case II is more practical than case I. It is because case I requires for the temperature of the cooling source to be below the suction temperature. For case II, one design example is illustrated in Fig. 9. A heat sink can be installed on the top of the compression chamber to enhance the heat transfer from the compression chamber to ambient. Furthermore, a separate space can be created on the top of the compression cylinder to allow refrigerants to flow through the heat sink and to absorb heat. In this case, an additional heat exchanger is needed. After cooling down the compressor head, the evaporated refrigerant vapor can release the heat to ambient air via the additional heat exchanger. Then the condensed liquid refrigerant can flow back to the compressor head by gravity. Since a separate space is created on the top of the compressor head, the refrigerant flowing through
the space may, or may not be the same refrigerant as the one in the main loop. Another design option for the hermetic compressor is to expose the compression cylinder head to the environment to take advantage of a relatively large temperature difference and heat transfer area to enhance the heat transfer effect. 4. Conclusions The current study focused on a theoretical analysis of benefits from the various options of active compressor cooling for refrigerants R22, R134a, R410A and R744 as working fluids. The external motor cooling improves the system COP and capacity up to approximately 4% for an R22 system at ARI low temperature refrigeration application, but it has little effect in reducing the power consumption. Contrastingly, the isothermal compression with active compressor cooling can effectively reduce the compressor power consumption. Ideal vapor compression cycles using combined isentropic and isothermal compression process are analyzed. The results show that the low temperature refrigeration application, which has large pressure ratios and high discharge temperatures, can potentially obtain more benefits than the air conditioning application by applying isothermal compression. The potential COP improvement for R134a system is the least among four refrigerants investigated since R134a has a relatively high molecular mass and low discharge temperature. Correspondingly, R744 has the largest improvement due to the lowest molecular mass among those four refrigerants. Combined isentropic and isothermal compression can reduce the compression power of the system by up to about 16% depending on operating conditions and fluid choice. Acknowledgements The support of this effort through the sponsors of the Alternative Cooling Technologies and Applications Consortium and the Center for Environmental Energy Engineering (CEEE) at the University of Maryland is gratefully acknowledged.
Fig. 9. Design example for the compressor cooling case II.
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
References [1] P. Prakash, S. Werner, On suction gas heating in hermetic compressors (A technical note), in: Proceedings of the 1978 Purdue Compressor Technology Conference, 1978, pp. 144–147. [2] N. Dreiman, Hermetic compressor with improved motor cooling, in: Proceedings of the 16th International Compressor Engineering Conference at Purdue, 2002, pp. C12–2. [3] ASHRAE Handbook, HVAC Systems and Equipment, I-P Edition, ASHRAE, Atlanta, 2000. [4] R. Radermacher, Y. Hwang, Vapor Compression Heat Pumps: With Refrigerant Mixtures, Taylor & Francis, 2005. [5] I.B. Vaisman, COP analysis of carbon dioxide cycles, ASHRAE Transactions 108 (Part 2) (2002) 252–262. [6] M.W. Coney, P. Stephenson, A. Malmgren, C. Linnemann, R.E. Morgan, R.A. Richards, R. Huxley, H. Abdallah, Development of a reciprocating compressor using water injection to achieve quasiisothermal compression, in: Proceedings of the 16th International Compressor Engineering Conference at Purdue, 2002, pp. C4–3. [7] T. Afjei, P. Suter, Experimental analysis of an inverter-driven scroll compressor with liquid injection, in: Proceedings of the 1992 International Compressor Engineering Conference at Purdue, 1992, pp. 541–550. [8] S. Ayub, J.W. Bush, D.K. Haller, Liquid refrigerant injection in scroll compressors operating at high compression ratios, in: Proceedings of the 1992 International Compressor Engineering Conference at Purdue, 1992, pp. 561–567. [9] A.K. Dutta, T. Yanagisawa, M. Fukuta, An investigation of the performance of a scroll compressor under liquid refrigerant injection, International Journal of Refrigeration 24 (2001) 577–587. [10] E.L. Winandy, J. Lebrun, Scroll compressor using gas and liquid injection: experimental analysis and modeling, International Journal of Refrigeration 25 (2002) 1143–1156. [11] Y.C. Park, Y. Kim, H. Cho, Thermodynamic analysis on the performance of a variable speed scroll compressor with refrigerant injection, International Journal of Refrigeration 25 (2002) 1072–1082. [12] H. Yamazaki, T. Itoh, K. Sato, H. Kobayashi, High performance scroll compressor with liquid refrigerant injection, in: Proceedings of the 16th International Compressor Engineering Conference at Purdue, 2002, pp. C22–1. [13] H. Cho, J.T. Chung, Y. Kim, Influence of liquid refrigerant injection on the performance of an inverter-driven scroll compressor, International Journal of Refrigeration 26 (2003) 87–94.
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[14] P.A. Domanski, Theoretical evaluation of the vapor compression cycle with a liquid-line/suction-line heat exchanger, economizer, and ejector, NIST, NISTIR 5606, 1995. [15] I.B. Vaisman, Economizer cycle in air conditioning systems with rotary vane compressors, in: Proceedings of the 8th International Refrigeration Conference at Purdue University, 2000, pp. 513– 520. [16] J. Siddharth, J. Gauray, B. Clark, Vapor injection in scroll compressors, in: Proceedings of the 2004 International Compressor Engineering conference at Purdue, 2004, p. C081. [17] M.M. Perevozchikov, H.M. Pham, Scroll compressor for mobile HVAC/R application, in: Proceedings of the 17th International Compressor Engineering conference at Purdue, 2004, p. C095. [18] M.F. Taras, Is economizer cycle justified for AC applications, ASHRAE Journal (July) (2005) 38–44. [19] D. Shapiro, C. Rohrer, Two-stage compressor with economizer cycle where piston(s) stroke(s) are varied to optimize energy efficiency, in: Proceedings of the International Compressor Engineering Conference at Purdue, July 17–20, 2006, p. C007. [20] B. Wang, X. Li, W. Shi, Q. Yan, Effects of refrigerant injection on the scroll compressor, in: Proceedings of the International Compressor Engineering Conference at Purdue, July 17–20, 2006, p. C091. [21] S. He, Northern China heat pump application with the digital heating scroll compressor, in: Proceedings of the International Compressor Engineering Conference at Purdue, July 17–20, 2006, p. R116. [22] B. Wang, X. Li, W. Shi, Q. Yan, Design of experimental bench and internal pressure measurement of scroll compressor with refrigerant injection, International Journal of Refrigeration 30 (2007) 179– 186. [23] Y. Hwang, X. Wang, R. Radermacher, Two-stage cycle with vapor injection compressor, in: Proceedings of the International Congress of Refrigeration, Beijing, 2007. [24] Air Conditioning and Refrigeration Institute (ARI), ANSI/ARI Standard 540, Positive displacement refrigerant compressors and compressor units, 1999. [25] ASHRAE, ASHRAE Standard 23-1993, Methods of testing for rating positive displacement refrigerant compressors and condensing units, 1993. [26] E.W. Lemmon, M.O. McLinden, M.L. Huber, REFPROP Reference fluid thermodynamic and transport properties, NIST Standard Reference Database 23, Version 7.0, 2002.
Applied Thermal Engineering 28 (2008) 1791–1797 www.elsevier.com/locate/apthermeng
Investigation of potential benefits of compressor cooling Xudong Wang, Yunho Hwang *, Reinhard Radermacher Center for Environmental Energy Engineering, Department of Mechanical Engineering, University of Maryland, College Park, MD 20742, USA Received 22 November 2006; accepted 12 November 2007 Available online 22 November 2007
Abstract The compressor is certainly the largest power consumer in a vapor compression system. To reduce the power consumption of the compressors two performance improving options are investigated theoretically for refrigerants R22, R134a, R410A and R744 as working fluids. The first option is cooling the motor by external means other than using the suction gas. Analysis results for this option show that R22, R410A and R744 have larger potential benefits than R134a. In low temperature refrigeration applications larger improvements are achievable than in air conditioning applications for all four refrigerants. The second option is to make the compression process isothermal by transferring heat from the compression chamber. To approach the isothermal compression process while avoiding wet compression, two cases of combining isothermal and isentropic compression processes are analyzed. The analysis results show that this strategy can reduce the compression work up to 14% as compared to the isentropic compression process for the R22 refrigeration system at ASHRAE T condition. The second analysis on the ideal vapor compression cycles using such compressor cooling strategy show that the compression power of the system can be reduced by up to about 16% depending on operating conditions and fluid choice. Ó 2007 Elsevier Ltd. All rights reserved. Keywords: Compressor; Cooling; Isothermal compression; Isentropic compression
1. Introduction Compressor performance significantly affects refrigeration system performance. Heat transfer to the suction gas through the suction line, compressor shell and compressor motor eventually degrades the performance of the compressor [1]. Effective cooling of the compressor motor may reduce this adverse effect. Dreiman [2] reported that about 2% coefficient of performance (COP) increase was achieved by utilizing the vortex tube phenomena to cool the motor stator winding. Nowadays, for conventional compressor technology, the accepted motor efficiencies range from approximately 88% for a 2 kW motor to 95% for a 75 kW motor and the mechanical efficiency is in the same order of magnitude [3]. Opportunities for significant improvements are possible only when the compression process is changed, as an example, to isothermal compression.
*
Corresponding author. Tel.: +1 301 405 5247; fax: +1 301 405 2025. E-mail address: [email protected] (Y. Hwang).
1359-4311/$ - see front matter Ó 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2007.11.010
Isothermal compression is a process in which the temperature of the working fluid remains constant during the compression process. However, it is impossible for refrigerants with a high specific heat capacity to be compressed isothermally over the entire compression process [4]. It is because refrigerants with a large molecular mass and therefore, large specific heat capacity, becomes two-phase state before they reach the condensing pressure [5], and leads to wet compression. Furthermore, a large amount of heat must be removed during the compression process by some means to maintain a constant temperature of the working fluid. Coney et al. [6] developed an air compressor which could approach quasi-isothermal compression via water injection, and reported that 28% savings of compression work could be achieved practically at a pressure ratio of 25. However, water injection is not favored option to current commercial compressors being utilized in the field of air conditioning and refrigeration. A technique similar to the water injection and applicable to current commercial compressors is refrigerant injection. The basic idea is to inject the refrigerant in either liquid [7–13] or vapor phase
1792
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
Nomenclature ARI COP h m P Q RPM v Vdis
Air Conditioning and Refrigeration Institute coefficient of performance enthalpy (kJ/kg) refrigerant mass flow rate (kg/s) pressure (kPa) heat transfer amount per unit mass (kJ/kg) revolutions per minute specific volume (m3/kg) displacement volume (m3)
[10,14–23] to the sealed compression chamber to reduce the discharge temperature and to improve the system performance. However, both liquid and vapor refrigerant injection eventually increase the compression work since an extra amount of refrigerant is injected and compressed, which requires additional compression work. Since the existing technologies of refrigerant injection to the compressor are not able to achieve an isothermal compression process, it is warranted to further investigate isothermal compression process and alternative cooling concepts. The purpose of this paper is to outline the potential benefits of compressor cooling in an ideal vapor compression cycle through theoretical analysis. While the aforementioned research focuses on the compression process with refrigerant injection to the sealed compression chamber, the current work is focused on the compression process without any refrigerant injection to the sealed compression chamber so that the compressor does not need extra power consumption associated with refrigerant injection. The benefit of changing the compression process from isentropic to isothermal and the heat transfer amount required for such change are presented to provide a quantified guideline for alternative compressor cooling concepts.
W Wisen Wiso
W gm gvol q
compression work per unit mass (kJ/kg) isentropic compression work per unit mass (kJ/ kg) isothermal compression work per unit mass (kJ/ kg) power consumption (kW) motor efficiency volumetric efficiency density (kg/m3)
air or by other means instead of the suction gas, the performance of the compressor would be improved. To demonstrate the effect of eliminating suction gas preheating, the performance of ideal cycles with and without suction gas preheating is carried out. The two cycles, superimposed on a pressure–enthalpy (P–h) diagram, are shown in Fig. 1. In this analysis, it is assumed for the base cycle (cycle 10 –20 –3–4) that the suction gas is only preheated by the motor loss from state point 1 to state point 10 , and the compression process is isentropic process. In the second cycle (cycle 1–2–3–4), the motor is cooled presumably by external means other than using the suction gas. Eq. (1) is used to calculate the refrigerant mass flow rate from the predetermined variables: volumetric efficiency, suction density, displacemental volume, and compressor rotational speed. The suction gas state after preheating (state point 10 ) can be found by combining the Eqs. (2) and (3) from the predetermined variables: motor efficiency, suction enthalpy before preheating (h1), and discharge pressure. Then Eq. (2) is used to calculate the compression work. The heat absorbed by the suction gas is calculated by Eq. (3)
m ¼ gvol q1 V dis RPM=60
W gm ¼ mðh20 h10 Þ 2. Analysis of compressor cooling options 2.1. External motor cooling Once the suction gas leaving the evaporator enters the compressor shell, it is heated by various sources since the compressor housing temperature is generally much higher than the suction gas temperature. The suction gas preheating is caused by the electrical loss from the motor, the mechanical loss from the frictional parts, and the thermal losses caused by the heat transfer from the compression chamber and the internal refrigerant leakage. Suction gas preheating eventually results in a decrease of the suction gas density. Consequently, the refrigerant mass flow rate and the volumetric efficiency are reduced. Therefore, energy required to maintain the same mass flow rate is also increased due to the suction gas preheating [1]. If the motor could be effectively cooled by the forced convention with
W ð1 gm Þ ¼ mðh10 h1 Þ
ð1Þ ð2Þ ð3Þ
Other conditions used in the analysis are as follows: a 54.2 cc displacement volume, a volumetric efficiency of 90%, a motor efficiency of 90%, and a compressor RPM
Fig. 1. Effects of external motor cooling shown in P–h diagram.
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
of 3600. Refrigerants R22, R134a, R410A and R744 are studied as examples for typical refrigerants. The system performance is evaluated under Air Conditioning and Refrigeration Institute (ARI) conditions for air conditioning and low temperature refrigeration applications [24] as summarized in Table 1. Since R744 cycle is a transcritical cycle, the compressor discharge pressure is higher than its critical pressure, the discharge pressures of 9 MPa and 10 MPa are chosen as an example for low temperature refrigeration application and air conditioning application, respectively. The performance enhancement by the external motor cooling is illustrated in Fig. 2. The results show that the system capacity and COP can be improved when the motor is cooled by the external means instead of the suction gas. The refrigerants R22, R410A and R744 have better potential benefits than R134a. The low temperature refrigeration application can achieve more improvement than the air conditioning application. The largest COP improvement of 5.4% is observed for an R744 system at ARI low temperature refrigeration application. The capacity of the R744 system improved up to 5.5% for the same condition. However, cooling motor has little effect in reducing the power consumption. 2.2. Isothermal compression with active compressor cooling To reduce the compressor power consumption a more active compressor cooling is investigated. In Fig. 3, two ideal compression processes, isentropic process (1–2) and isothermal process (1–20 ), are shown in a pressure–specific volume (P–v) diagram. A closed area 1–2–3–4–1 represents the isentropic compression work. Similarly, the enclosed area 1–20 –3–4–1 represents the isothermal compression
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Fig. 3. Isentropic and isothermal compression processes in P–v diagram.
work. Isothermal compression requires less work input than the isentropic compression for a given pressure ratio. The shaded area 1–2–20 –1 represents the work saving by changing the compression process from isentropic to an isothermal process. To approach the isothermal compression a certain amount of heat must be removed from the compression process. Such amount of heat can be calculated based on the energy balance of the compression process. When the isothermal compression is approached the wet compression could happen. In order to approach isothermal compression while avoiding the wet compression, a combination of partial isothermal compression and partial isentropic compression is proposed by applying effective cooling techniques to different compression stages. Two possible options under this strategy are illustrated in Fig. 4. It shows only the compression process of the refrigeration cycle in P–h diagram. As an example, refrigerant R22 is selected as the working fluid and the cycle is assumed to be operating at ASHRAE T condition [25], which is referred as 7.2 °C, 54.4 °C, 35 °C, and 8.3 °C for saturated evaporating temperature, saturated condensing
Table 1 ARI conditions [24] Conditions
Temperature (°C)
Air conditioning Low temperature refrigeration
Ambient
Evaporating
Suction
Condensing
Liquid
35 35
7.2 23.3
18.3 4.4
54.4 48.9
46.1 40.6
ARI-Air Conditioning Application Capacity
ARI-Refrigeration Application Electric Power
COP
Capacity
COP
3.0% 2.5% 2.0% 1.5% 1.0% 0.5% 0.0% -0.5%
R22
R134a
R410A
Refrigerants
R744
Change of Performance
Change of Performance
Electric Power
5.5% 4.5% 3.5% 2.5% 1.5% 0.5% -0.5%
R22
R134a
R410A
Refrigerants
Fig. 2. Effects of external motor cooling on system performance.
R744
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X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
Fig. 4. Combinations of isothermal and isentropic compressions in P–v diagram.
temperature, return gas temperature, and degree of subcooling, respectively. The base compression process for both cases is the isentropic compression 1–20 . In case I, the effective cooling is applied to the compressed R22 only during the first compression process 1–2. Once the R22 reaches its saturated vapor state isothermally, the cooling is stopped. Then, the second compression process follows the isentropic process 2–3. This case requires for the suction gas to have some degree of superheating. In case II, the R22 is isentropically compressed first to an intermediate pressure from state 1 to 200 . Then it undergoes an isothermal compression process from state 200 to 300 by an external cooling. The two compression cases and the baseline, isentropic compression, are illustrated in P–v diagram (Fig. 5). The isothermal compression work is calculated by integrating vdP along the isotherm in the P–v diagram. For
2250
2147 kPa 3" 3
2'
Pressure (kPa)
2000 1750 Baseline case: isentropic process
example, the compression work and the heat transfer amount in case I are calculated by Eqs. (4)–(7) Z P2 W iso ¼ v dP ð4Þ P1 Z P2 v dP ð5Þ Q ¼ ðh1 h2 Þ þ P1
W isen;2–3 ¼ h3 h2 W isen;1–20 ¼ h20 h1
ð6Þ ð7Þ
The comparison of compression work and heat transfer amounts for these two cases and isentropic compression process is summarized in Table 2. The results show that the case I and the case II can reduce the compression work up to 14.0% and 8.5%, respectively, as compared to the isentropic compression process at ASHRAE T condition. Since the case I needs to remove less heat, and saves more work than the case II, it is preferred to the case II in the performance enhancement point of view. However, in the case I it must be assured that the temperature of cooling sources is below the suction temperature, which may be a considerable challenge. For the case II, part of the condenser can possibly be used as the heat sink, or an additional heat exchanger may be applied to achieve the process.
1500 1355 kPa 2
1250
Table 2 Comparison of three compression cases at ASHRAE T condition
Case I 1000
917 kPa
Case II 2"
750
1
626 kPa
Compression process
Base case (1–20 )
Case I (1–2–3)
Compression work (kJ/kg)
35.6
1–2 2–3 Total
19.7 10.9 30.6
1–2 2–3 Total
10.6 22.0 32.6
Removed heat (kJ/kg)
0.0
1–2 2–3 Total
32.9 0.0 32.9
1–2 2–3 Total
0.0 43.4 43.4
500 0
0.01
0.02
0.03
Specific volume
0.04
(m3/kg)
Fig. 5. Three compression processes in P–v diagram.
0.05
Case II (1–200 –300 )
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
The external motor cooling is found to be not a very effective means to reduce the power consumption. This is because the increase of the mass flow rate balances out the decrease of the power consumption per unit of mass flow by avoiding the suction gas preheating. However, the increase of the mass flow rate results in an increase in the capacity. Then the improvement of the COP is dominated by the improvement of the system capacity instead of the reduction of the power consumption. Contrastingly, the isothermal compression with active compressor cooling appears to be an effective means to reduce the compressor power consumption, and the system COP improvement comes from the power reduction. To further demonstrate the benefit of the isothermal compression with active compressor cooling, the compression process of case II in Figs. 4 and 5 are analyzed for multiple refrigerants, R22, R134a, R410A and R744, for both an air conditioning application and a low temperature refrigeration application under ARI conditions [24]. The discharge pressures of 9 MPa and 10 MPa are chosen for R744 system as an example for low temperature refrigeration application and air conditioning application, respectively. The discharge temperature of isothermal compression varies depending upon the temperature and capacity of the external cooling sources. Fig. 6 illustrates how the degree of external cooling affects the cycles and discharge temperature in P–h and P–v diagrams. The discharge temperature affects the compressor performance. The power savings and heat transfer amounts needed for various discharge temperatures are illustrated in Figs. 7 and 8, respectively for case II under ARI conditions. From Fig. 7, one can observe that the combined isentropic and isothermal compression can reduce the compression power of the system by up to about 16% for the case of R744 at low temperature refrigeration application. For all
18
Power Reduction [%]
3. Results and discussion
four refrigerants investigated the reduced discharge temperature results in increased power reduction. This means that the earlier the isothermal compression starts, the more energy savings one can actually obtain from this process. Since the discharge temperatures for all four refrigerants at low temperature refrigeration application is much higher than those at air conditioning application, a better performance improvement in low temperature refrigeration application can be obtained than in air conditioning application for all four refrigerants. Moreover, there is a significant
R134a-Air conditioning
R134a-Refrigeration
16
R22-Air conditioning
R22-Refrigeration
14
R410A-Air conditioning
R410A-Refrigeration
R744-Air conditioning
R744-Refrigeration
12 10 8 6 4 2 0 45
55
65
75
85
95
105
115
Discharge Temperature [°C] Fig. 7. Power reduction vs. discharge temperature for case II.
140 120
Heat Transfer [kJ/kg]
Since the current study is focused on the theoretical analysis, the uncertainty source only comes from the uncertainties of the refrigerants properties. The uncertainties of the refrigerants enthalpy and specific volume are less than 1% and 0.5%, respectively [26].
1795
100
R134a-Air conditioning
R134a-Refrigeration
R22-Air conditioning
R22-Refrigeration
R410A-Air conditioning
R410A-Refrigeration
R744-Air conditioning
R744-Refrigeration
80 60 40 20 0 45
55
65
75
85
95
105
115
Discharge Temperature [°C] Fig. 8. Heat transfer amount vs. discharge temperature for case II.
Fig. 6. Compression processes with different discharge temperatures for case II.
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X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
improvement potential for R744 in low temperature refrigeration application since it has a high discharge temperature and small molecular mass. The refrigerants with smaller molecular mass benefit more from compressor cooling than those with larger molecular mass. This is because the saturated vapor line of the refrigerant with smaller molecular mass has a steeper slope in the P–h diagram because of their smaller specific heat capacity. Therefore, a larger portion of the isentropic process can possibly be switched to an isothermal compression process at the latter part of the compression process while avoiding wet compression. R134a has the least benefit among four refrigerants investigated. This result is not surprising since R134a has the lowest discharge temperature among those refrigerants. Therefore, the potential for performance improvement is more limited for R134a as compared to other three refrigerants. Applying isothermal compression either in the case I or II does not make any difference in the system capacity since the suction condition is assumed to be the same. The COP improvement comes from the power reduction. Therefore, the COP changes follow the same magnitude and trend as the power does. From an energy balance point of view, increased power savings mean more heat needs to be removed from the compression process. Consequently, the amount of heat transferred from the compressor cylinder follows the same trend as the power savings. When the two cases are compared, the case II is more practical than case I. It is because case I requires for the temperature of the cooling source to be below the suction temperature. For case II, one design example is illustrated in Fig. 9. A heat sink can be installed on the top of the compression chamber to enhance the heat transfer from the compression chamber to ambient. Furthermore, a separate space can be created on the top of the compression cylinder to allow refrigerants to flow through the heat sink and to absorb heat. In this case, an additional heat exchanger is needed. After cooling down the compressor head, the evaporated refrigerant vapor can release the heat to ambient air via the additional heat exchanger. Then the condensed liquid refrigerant can flow back to the compressor head by gravity. Since a separate space is created on the top of the compressor head, the refrigerant flowing through
the space may, or may not be the same refrigerant as the one in the main loop. Another design option for the hermetic compressor is to expose the compression cylinder head to the environment to take advantage of a relatively large temperature difference and heat transfer area to enhance the heat transfer effect. 4. Conclusions The current study focused on a theoretical analysis of benefits from the various options of active compressor cooling for refrigerants R22, R134a, R410A and R744 as working fluids. The external motor cooling improves the system COP and capacity up to approximately 4% for an R22 system at ARI low temperature refrigeration application, but it has little effect in reducing the power consumption. Contrastingly, the isothermal compression with active compressor cooling can effectively reduce the compressor power consumption. Ideal vapor compression cycles using combined isentropic and isothermal compression process are analyzed. The results show that the low temperature refrigeration application, which has large pressure ratios and high discharge temperatures, can potentially obtain more benefits than the air conditioning application by applying isothermal compression. The potential COP improvement for R134a system is the least among four refrigerants investigated since R134a has a relatively high molecular mass and low discharge temperature. Correspondingly, R744 has the largest improvement due to the lowest molecular mass among those four refrigerants. Combined isentropic and isothermal compression can reduce the compression power of the system by up to about 16% depending on operating conditions and fluid choice. Acknowledgements The support of this effort through the sponsors of the Alternative Cooling Technologies and Applications Consortium and the Center for Environmental Energy Engineering (CEEE) at the University of Maryland is gratefully acknowledged.
Fig. 9. Design example for the compressor cooling case II.
X. Wang et al. / Applied Thermal Engineering 28 (2008) 1791–1797
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