ASHRAE IJHVAC 15-4-2009 HVAC&R Research《《HVAC&R研究》第15卷 4号》.pdf

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1、 VOLUME 15, NUMBER 4 HVAC accepted March 8, 2009This paper presents the experimental investigation of a double acting rotary vane expander for work recovery in the transcritical CO2cycle and focuses on the design improvements for leak-age and friction within the expander. The bench tests were carrie

2、d out to investigate the leakage and friction distributions within the expander. The test results showed that the end gaps caused the major leakage within the improved expander prototype, while the friction losses associated with the vanesespecially due to the springs in the slotswere dominant and a

3、ccounted for about 70% of the total friction losses. By comparing the pressure-rotation angle diagrams of the improved prototype with the original one, the effects of adding springs in the slots and arrang-ing sealing vanes at the sealing arc on the thermodynamic processes were analyzed. It was show

4、n that the tight contact between the vanes and cylinder wall owing to the springs had a sig-nificant improvement on the thermodynamic processes in that the cycle duration resumed being normal and the expander demonstrated a reasonable expansion process. By putting springs in the vane slots and arran

5、ging the sealing vane in the cylinder at the sealing arc, the volumetric efficiency increased from 17% to 30%, and the isentropic efficiency improved from 9% to 23%, resulting in a maximum coefficient of performance COP improvement of 14.2% compared with the throttling cycle under the same test cond

6、itions. INTRODUCTIONCarbon dioxide (CO2) is considered a potential substitute for CFC/HCFC refrigerants owing to its environmentally benign feature. Due to the low critical temperature of CO2, the refrigera-tion system with CO2as working fluid has to be operated in transcritical cycle under typical

7、ambient conditions for refrigeration and air-conditioning applications. The large entropy genera-tion during the throttling process, which accounts for 30% to 40% of the total entropy genera-tion, is thought to be the major reason for comparatively low energy efficiency of the transcritical CO2cycle

8、 (Lorentzen and Pettersen 1993). Replacing the throttling valve with a work-recovering expander in the CO2refrigeration system is considered to have a great improvement on the coefficient of performance (COP) of the cycle. For the working conditions in which the evaporating temperature is 0C and the

9、 heat rejection temperature is 40C, an expander with an efficiency of 70% can improve the COP by 40% to 60% if the isentropic effi-ciency of compressor is 80% (Nickl et al. 2005).Several kinds of CO2expanders have been investigated in the last decade. Huff and Radermacher (2003) modified a scroll-re

10、frigeration compressor into an expander to replace the throttling valve in the CO2air-conditioning system, and the isentropic efficiency was reported to increase from 30% to Bingchun Yang is a doctoral student, Xueyuan Peng is a professor and the corresponding author, Shaoyi Sun is a master student,

11、 Bei Guo is an associate professor, and Ziwen Xing is a professor in the School of Energy and Power Engineering, Xian Jiaotong University, Xian, China. 2009, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in HVAC two springs with a rigidit

12、y of 2000 N/m (11.43 lbf/in.) and an initial deflection of 6 mm (0.2362 in.) were put underneath the vane in each slot to help the vane contact the cylinder wall tightly, thus preventing the inter-nal leakage through the gap between the vane and cylinder inner wall. EXPERIMENTAL INVESTIGATIONPerform

13、ance Test of the Expander in the CO2System Figure 3 shows the experimental system for the expander performance test, which was built on the basis of a transcritical CO2refrigeration cycle. The high-pressure gas from the gas cooler of the CO2refrigeration system flows through a Coriolis mass flowmete

14、r, by which the mass flow rate is measured, and enters the expander. Both pressure and temperature at the expander inlet/outlet ports are measured so that the specific enthalpy ( ) and density ( )in the inlet state is calculated. The specific enthalpy ( ) in the outlet state is determined by the Fig

15、ure 1. Structure of the rotary vane expander prototype.mrealhininhout676 HVAC the larger pressure difference pushed the rubber seal to contact the rotor outer surface more tightly. In order to validate the sealing improvement measures, the leakage rate within the expander with-out any sealing measur

16、es was measured to be 0.084 kg/s (0.038lb/s) under the pressure differ-ence of 3.5 MPa (507.6 psi), while the leakage rate was only 0.013 kg/s (0.006lb/s) as the sealing measure were taken. Friction DistributionAs shown in Figure 8, all the friction losses tended to increase with the rotational spee

17、d. For the expander with springs in the slots, the friction loss associated with the vanes increased almost linearly from 6 W (20.472 Btu/h) at 100 rpm to 212 W (723.34 Btu/h) at 1800 rpm. How-ever, after the springs were taken out of the slots, the friction loss associated with the vanes had a much

18、 different tendency, increasing more slowly until the rotational speed reached 800 rpm and then increasing quickly almost with the same rate as the expander with springs in the slots. This difference may be caused by the fact that at lower speed the vanes did not contact the cylin-der wall at all an

19、d began to contact the wall as the speed reached 800 rpm. The friction losses caused by the shaft seals and sealing vanes were relatively small compared with the friction loss associated with the vanes. For the expander with springs in the slots, the friction losses caused by the vanes, sealing vane

20、s, and shaft seal accounted for 74%, 9%, and 16%, respectively at 800 rpm and 71%, 10%, and 19%, respectively at 1800 rpm. This result means that further improvement efforts on the friction issue should be focused on reducing the friction caused by the vanes. For example, the method to introduce the

21、 high-pressure gas into the slots is being investigated.Figure 7. Leakage distributions under different pressure differences.VOLUME 15, NUMBER 4, JULY 2009 683Effects of Sealing Improvements on Thermodynamic Processes The P- diagram recorded from the original expander prototype is shown in Figure 9a

22、. Com-pared with the ideal case (Figure 5b), two typical differences are found. First, the duration for one cycle is twice that of the ideali.e., it changes from 51 to 102. Secondly, the two pressure sensors both recorded the whole working processes; though, theoretically, they have to combine their

23、 signals to compose the whole working process. These two phenomena suggest that the vane did not contact the cylinder wall at most of the rotation angles. Through the gap between the vane tip and the cylinder wall, large internal leakage made the two adjacent chambers behave as if they were connecte

24、d into one chamber, which made the duration double, and two pressure sensors recorded almost the same pressure in most of the rotational angles. After adding the springs underneath the vane in each slot, an obvious change in the expansion process occurred, as shown in Figure 9b. In addition, the dur

25、ation of the pressure history changed to 51. Note that the decrease in the pressure during expansion becomes slower after the rotating angle passes 70, which could be attributed to the larger leakage from the adjacent high-pressure chamber. At the angle of 86, pressure sensor 1 begins to collect the

26、 pressure sig-nals of the next chamber but pressure sensor 2 continues recording the pressure signals of this chamber. Therefore, further drop in the pressure is observed from Figure 9b at the angle of 95 until it approaches the outlet pressure. Due to the leakage from the neighboring high-pressure

27、chamber, the pressure increases later. When the rotation angle reaches the outlet angle (110), the pressure decreases to the outlet pressure. The fluctuation in the pressure for both the inlet and outlet processes is about 0.3 MPa (43.5 psi). Owing to better sealing and less leakage, better internal

28、 working process was recorded in the expander with seal at the sealing arc. As shown in Figure 9c, the pressure at the inlet port increased from 6.5 MPa (942.7 psi) to 8.5 MPa (1232.8 psi), and the pressure difference across the expander increased from 2.0 MPa (290.1 psi) to 2.7 MPa (391.6 psi). It

29、is seen clearly that the expansion process began from the supercritical region, and the pressure decreased quickly before it reached the critical pressure. Then the pressure decreased slowly to about 6.5 MPa Figure 8. Friction loss distributions within the expander.684 HVAC one vane moved just like

30、in Figure 10a, and another moved as in Figure 10b. This phenomenon could be explained as follows: the (a)(a)Figure 10. Trajectories of the vane in the slot.686 HVAC the maximal volumetric efficiency increased from 17% to 30%, and maximal isentropic efficiency increased from 9% to 23%. The COP of the

31、 expander-based system was improved by 14.2% at maximum under the test conditions.2. The springs installed in the vane slots significantly improved the sealing and, thereby, improved the thermodynamic processes of the expander by helping the vane keep tight con-tact with the cylinder inner wall. Wit

32、hout the springs, the adjacent vanes demonstrated differ-ent behaviors, and one of the neighboring vanes did not contact the cylinder wall throughout most of the cycle, resulting in poor efficiency.3. For the expander prototype with springs in the slots, the leakage rate through the end gaps account

33、ed for 43% under the pressure difference of 2.3 MPa (333.6 psi), which was about the same as that through the sealing arc. Decreasing the end gap clearance further or designing new seal structures is being considered.4. The friction loss associated with the vanes for the expander with springs in the

34、 slots was dominant, accounting for about 70% of the total friction losses, which indicated that better measurement should be taken instead of using springs to maintain contact between vanes and the cylinder wall.VOLUME 15, NUMBER 4, JULY 2009 6875. The visualization of the vane movement showed that

35、 the inlet gas flow pushed the vanes downward toward the slot bottom and caused the loss of contact between the vanes and the cylinder wall at the early stage of the inlet process. 6. Arrangement of a seal at the sealing arc was verified to be helpful to reduce the gas leakage directly from the inle

36、t port to the outlet port. ACKNOWLEDGMENTSThe study presented in this paper was supported by program for New Century Excellent Tal-ents in University (NCET-06-0836) and National Natural Science Foundation of China (No. 50576072).NOMENCLATURE= mass flow rate, kg/hn=rotational speed of expander, rpmP

37、= pressure, MPa= cooling capacity, kWT = temperature, CV = volume, m3v = specific volume, m3/kgW = power, kW= uncertainty= efficiency= density, kg/m3= rotation angle, Subscriptsa =adiabaticc = compressorexp = expanderthv = throttling valveevap = evaporatorin = inlet stateout = outlet state v = volum

38、etricREFERENCESBadr, O., P.W. OCallaghan, M. Hussein, and S.D. Probert. 1984. Multi-vane expanders as prime movers for low-grade energy organic rankine-cycle engines. Applied Energy 16(2):12946.Badr, O., P.W. OCallaghan, and S.D. Probert. 1985a. Multi-vane expanders: Geometry and vane kinemat-ics. A

39、pplied Energy 19(3):15982.Badr, O., P.W. OCallaghan, and S.D. Probert. 1985b. Multi-vane expander performance: Breathing char-acteristics. Applied Energy 19(4):24171.Badr, O., S.D. Probert, and P.W. OCallaghan. 1985c. Performances of multi-vane expanders. Applied Energy 20(3):20734.Badr, O., S.D. Pr

40、obert, and P.W. OCallaghan. 1985d. Multi-vane expanders: Vane dynamics and friction losses. Applied Energy 20(4):25385.Badr, O., S.D. Probert, and P.W. OCallaghan. 1985e. Multi-vane expanders: Internal-leakage losses. Applied Energy 20(1):146.Baek, J.S., E.A Groll, and P.B. Lawless. 2005a. Piston-cy

41、linder work producing expansion device in a tran-scritical carbon dioxide cyclePart I: Experimental investigation. International Journal of Refrigera-tion 28(2):14151.Baek, J.S., E.A Groll, and P.B. Lawless. 2005b. Piston-cylinder work producing expansion device in a transcritical carbon dioxide cyc

42、lePart II: Theoretical model. International Journal of Refrigeration 28(2):15264.Edwards, T.C. 1972. Analysis of mechanical friction in rotary vane machines. International Compressor Engineering Conference, Washington, DC.Fukuta, M., and T. Yanagisawa. 2003. Performance prediction of vane type expan

43、der for CO2cycle. Inter-national Congress of Refrigeration, Washington, D.C.Fukuta, M., T. Yanagisawa, O. Kosuda, and Y. Ogi. 2006. Performance of scroll expander for CO2refrig-eration cycle. International Compressor Engineering Conference, West Lafayette, IN.Guan, H., Y. Ma, and M. Li. 2006. Some d

44、esign features of CO2swing piston expander J. Applied Ther-mal Engineering 26(23):23743.mQ688 HVAC accepted February 26, 2009A mathematic model focusing on expander thermodynamics and vane dynamics was developed to investigate the major factors influencing the efficiencies of the rotary vane expande

45、r. Several fac-tors were taken into account, including the leakage through various leakage paths, friction associ-ated with the vanes, and flow through the inlet/outlet ports. The model was validated by comparing the calculated thermodynamic processes and vane movement with the experimental data, wh

46、ich showed the deviation was less than 10%. The predicted results from the model indicated that the expander would have an optimum pressure ratio of about 2.2. Although the volumetric efficiency increased with the rotational speed, the optimal rotational speed of 2300 rpm was obtained at inlet and o

47、utlet pressures 8 MPa (1160 psi) and 4 MPa (580 psi) and an inlet temperature of 40C. The leakage through both the end gaps and the sealing arc had a significant influence on the expander efficiency, accounting for 35% and 28%, respectively, of the losses in the volumetric efficiency. Among the majo

48、r geometric parameters of the expander, larger eccentricity and vane incline angle had a positive effect on the expander efficiencies, and the increase in the vane width had a negative effect, while the effect of the ratio of length to radius seemed to be insignificant.INTRODUCTIONDue to the environ

49、mental issues associated with global warming and ozone depletion, the hydro-chlorofluorocarbon (HCFC)/chlorofluorocarbon (CFC) refrigerants are being phased out. Much attention has been paid to the natural refrigerant, carbon dioxide (CO2), as a substitute refrigerant for the traditional Freon gases in air-conditioning and heat-pump applications since the 1990s. How-ever, the inherently large throttling loss and consequent low coefficient of performance (COP) of the tr