1、High-Efficiency Compressor Design for HC-600a Refrigerant Using Energy-Saving Household Refrigerators Ichiro Kita Makoto Katayama Akira Nakano ABSTRACT This paper reports environmentally friendly technologies for compressors and refrigerators. For household refrigerators, HC-600a isobutane refrigera
2、nt is used worldwide. First, we compare its characteristics to HFC-134a refrigerant and discuss the technologies regarding the higher volumetric efficiency of compressors. Second, we discuss technologies for improving the volumetric efficiency at the piston and the cylinder bore portions. The piston
3、 groove outperforms oil seals. We also report the optimum groove design and fluid analysis using CFD at the groove portion. INTRODUCTION In recent years, refrigeration and air-conditioning systems using natural refrigerants instead of fluorocarbon refrigerants have been researched and developed. In
4、refrigerators, this includes HC refrigerants, especially isobutane, which is environmentally friendly and demonstrates good performance. Table 1 compares the basic properties of HC-600a and HFC-134a. Table 1 Basic properties of HC-600a and HFC-134a Refrigerant HC-600a HFC-134a GWP 3 1300 ASHARE cond
5、ition properties (theoretical) Specific work kJ/kg(BTU/lb) 115.78 (0.04977) 67.91 (0.02919) Refrigeration effect kJ/kg(BTU/lb) 185.58 (0.07978) 335.14 (0.1441) COP (theoretical) ( EER ) W/W(BTU/Wh) 2.895 (9.89) 2.733 (9.33) Rate 1.059 Base (=1) Inlet vapor volume m3/kg(ft3/lb) 0.063506 (1.017) 0.019
6、212 (0.3078) Piston displacement rate 1.83 Base (=1) Ichiro Kita , Makoto Katayama , Akira Nakano are senior staff engineer , Home Appliances Company, Panasonic Corporation, Kusatsu, Shiga, Japan. LV-11-C024 2011 ASHRAE 1952011. American Society of Heating, Refrigerating and Air-Conditioning Enginee
7、rs, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Volume 117, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAES prior written permission.Even though HC-600a has about a 6% better theoretic
8、al COP than HFC-134a, the HC-600a compressor requires almost double displacement compared to HFC-134a. In compressor design, larger displacement needs a larger bore and/or piston stroke and a larger compressor. When we consider HC-600as characteristics, higher volumetric efficiency reduces its relat
9、ive size and increases its efficiency. Using an inverter compressor and expanding the motors higher revolutions also reduce its relative size. On the other hand, to attain significant energy savings of such refrigerating systems as household refrigerators, the compressors COP must be maximally adjus
10、ted under actual operating conditions. In this paper, the compressor designs for higher volumetric efficiency are discussed by focusing on the piston and the cylinder bore portions. The technologies of optimized piston groove design that achieve higher volumetric efficiency are also discussed. Furth
11、ermore, a trial of CFD applications is discussed to analyze oil flow. MAXIMUM THEORETICAL VOLUMETRIC EFFICIENCY Since the required compressor displacement for HC-600a refrigerant is almost double that of HFC-134a to achieve identical capacity and since the larger bore size increases piston-cylinder
12、clearance and a small displacement compressor increases re-expansion volume, raising volumetric efficiency in small refrigerating-capacity compressors is difficult. First, we estimated the maximum volumetric efficiency without gas leakage from the piston-cylinder portion with refrigerant gas heating
13、. We estimated under the ASHRAE condition (Td/Te = 54.4 / -23.3C (130 / -10 F) and used the following equation: , where vcis the theoretical volumetric efficiency by re-expansion, Pd is the discharge pressure, Ps is the suction pressure, c is a clearance ratio, and n is a polytropic exponent. Next w
14、e estimated the gas heating effect to the volumetric efficiency with the following equation: , 196 ASHRAE Transactionswhere vc(t)is the volumetric efficiency by refrigerant gas heating during the compression process, ts is the compressed refrigerant temperature, and t1 is the intake refrigerant temp
15、erature; vc(shell to suction)= compressor suction inlet/ calorie suction temperature ,(3)where vc(shell to suction)is the volumetric efficiency by the refrigerant gas heating from the shell suction port to the suction intake and is the refrigerant density in each portion; expected maximum= vc vc（t）
16、vc（shell to suction）,(4) where expected maximumis the total expected maximum volumetric efficiency for a tested 6-cm3(0.37 in3) displacement compressor without piston-cylinder bore leakage loss. We propose total expected maximum volumetric efficiency and applied total expected maximum volumetric eff
17、iciency expected maximum. From temperature experiments and calculations, the expected maximum volumetric efficiency is almost 80.2% with HC-600a. Higher volumetric efficiency is used to obtain higher COP, which is effective when volumetric efficiency is close to the expected maximum. Table 2 shows t
18、he major specifications using the theoretical volumetric efficiency calculations. Table 2 Major specifications for calculations Conditions Discharge pressure 762 (110.5) kPa(psi) Condensing temperature 54.4 (130) C(F) Suction pressure 62.6 (9.079) kPa(psi) Evaporating temperature -23.3 (-10) C(F) Co
19、mpressor dimensions Bore diameter 21 (0.827) mm(in) Bore clearance 7 (0.00028) (in) Piston stroke 17.5 (0.689) mm(in) Top clearance 80 (0.0031) (in) Refrigerant properties Polytropic exponent: n 1.15 - Compressed gas temp: t1 102.6 (217) C(F) Density at calorie suction: calorie suction temp1.34 (48.
20、4)kg/cm3 (lb/in3) Density at suction inlet: compressor suction inlet 1.46 (52.7)kg/cm3 (lb/in3) LOSS FROM PISTONCYLINDER BORE CLEARANCE We used the expected maximum volumetric efficiency in Eq. 4 and estimated the expected maximum volumetric efficiency at almost 80% in an ASHRAE condition with a 6-c
21、m3 (0.37 in3) displacement 2011 ASHRAE 197compressor at 50-Hz operation. Gas leakage is one reason that volumetric efficiency is reduced if the compressor fails to reach 80% efficiency. The leakage of refrigerant gas occurs through the piston-cylinder bore clearance during the compression process. T
22、he leakage calculation was presented by Imasu et al. as follows: pPtPlpt 22=1)(1)( 02, ( 5 ) where (t) is the leakage velocity, is the viscosity of the refrigerant gas, lp is the seal length, P is the pressure, pis the piston velocity, and is the half clearance (gap) between the piston-cylinder bore
23、 diameters. Leakage experiments were conducted using a calorimeter in the ASHRAE condition by comparing the pistons with and without oil grooves. This calculation has also been done using Eq. (5). Table 3 Piston-cylinder bore leakage ratio comparison Calculation Experimental Without piston groove Fl
24、uid formula between parallel board (piston-cylinder leakage) Without piston grooveWith piston groove Piston-cylinder clearance 2 Leakage loss-ratio Leakage loss-ratio difference (vs7) = -3.5exp-6Leakage loss-ratio experiment (vs7) =-3.5exp-6=1.85 Capacity (Leakage) loss-ratio experiment (vs7) Capaci
25、ty (Leakage) loss-ratio experiment (vs7) Capacity (Leakage) loss-ratio experiment (vs8.5) (in) - - - - - - 7 (0.00028) 3.49% 0.00% 0.00% 0.00% 0.00% 1.61% 8.5 (0.00033) 4.24% 0.75% 1.38% 1.38% -1.64% 0.00% 9.5 (0.00037) 4.74% 1.25% 2.31% - -0.36% 1.25% 11 (0.00043) 5.49% 2.00% 3.69% 3.85% - - Fig. 2
26、 Fig. 1 Table 3 shows the piston-cylinder bore leakage ratio at piston-cylinder bore clearances of 7, 8.5, 9.5, and 11 micrometers (0.00028, 0.00033, 0.00037, and 0.00043 inches). Numerical calculations and experiments were done using an actual reciprocating compressor around a 100 W (400BTU) capaci
27、ty model. Figures 1 and 2 show pistons without and with the piston oil 198 ASHRAE Transactionsgrooves used for calorimeter measurements. We did the measurements in the ASHRAE condition shown in Table 2. Table 3 shows the experimental data of the capacity ratio that compared the 7 (0.00028) or the 8.
28、5 micron (0.00033 inch) capacity of the right-side end column. The capacity loss ratio of the experimental results is calculated as the capacity loss ratio. Calculation was also done using Eq. (5), and the results are listed for the leakage loss ratio, shown in the three Calculation columns of Table
29、 3. To adjust the experimental results and Eq. (5), we added the half clearance revision coefficient. Revised and were estimated by Eqs. (6) and (7): = -3.5 10exp-6 (6) =-3.5 10exp-6 =1.85 (7) and are the following corrections: 1) the differences among the theoretical piston positions, (which refer
30、to the theoretical center with uniform clearance completely around the pistons position in the bore compared to its actual position in the cylinder bore), 2) the different piston-cylinder surface conditions, 3) and the lubrication oil conditions. From the Table 3 results, modified Eq. (5) using Eq.
31、(7) provides good agreement with the experimental result without a piston groove (See Figs 1, 2, and Table 3). When comparing the experimental results for pistons with and without oil grooves, we discovered significant improvement of the leakage ratio. The piston with an oil groove (Fig. 2) had no i
32、ncrease of capacity loss ratio (i.e., leakage ratio), although the piston-cylinder bore clearance increased from 7 to 9.5 microns (0.00028 to 0.00037 inch). From the actual capacity measurement, 8.5 micron (0.00033 inch) clearance had the best capacity, which means the best volumetric Fig. 4 0200400
33、6008000 30 60 90 120 150 180 210 240 270 300 330 360Rotational angle ( )Cylinder Pressure (KPa)-10123456ReedLift(mm)Cylinder pressureDischarge reed liftSuction pressure levelDischarge pressure levelefficiency. 2011 ASHRAE 199The experimental and calculation results show that the piston oil groove al
34、most reduces the leakage to zero. Table 4 shows the major specifications using the theoretical leakage speed and capacity loss calculations, and Fig. 4 shows the act cycle. The l kage ti 30 thro one ioual pressure curve in the cylinder bore through one rotating ea me was almost ugh -cycle operat n.
35、Tabl cifications an la for calce 4 Major spe d formu ulations Piston length lp 25 (0.984) mm (in) Discharge time 30 degree 30 of 360 (=Leakage time) (one-cycle operation) Gas velocity m/s ( ft/s ) from calculations (t) 13.7 (44.9) Condition and major dimensions Identical as Table 1 Capacity loss rat
36、io (=Leakage ratio) = Volume of leakage / piston displacement MEith the oil groove piston. NexWe ume and the lubricant oil flowed from the opposite side of the piston by cylinder force and shell pressure difference. With this initial condition, the calculation failed to conv . CHANISM ANALYSIS USING
37、 CFD WITH DUAPrimary approach of CFD (VOF) We significantly reduced the gas leakage between the piston-cylinder bores wL FLUID GAS AND OIL t we determined the mechanism of the oil sealing effect using CFD analysis. One model of CFD analysis is shown in Fig. 5.used a volume-of-fluid (VOF) method for
38、a two-phase flow, refrigerant gas, and lubricant oil mixture. In the analysis, we simultaneously investigated the oil volume rate and the speed for each mesh to find the oil behavior with the piston goove when the refrigerant gas was compressed. In the first trial, we experienced difficulties when w
39、e started the condition where the gas existed at the cylinder volCylinder head SideStandstill cylinderMoving wall (Piston surface)erge when 100% gas flow occurredFig. 5 Piston-cylinderhalf clearnceMoving lengthPiston roove200 ASHRAE TransactionsModified CFD (VOF) approach Therefore, in the second tr
40、ial, we changed the initial condition where the oil filled the piston-cylinder bore clearance and the grooves. Fig. 6 shows the result of the oil volume rate from the start of the comThe oil wall makes an oil labyrinth with gas and reduces the gas velocity in and location are key factors. pression t
41、o when the gas velocity reached about 10 m/s (32.8 ft/s). From Fig. 6, the oil that remained in the groove works the oil wall in the clearance and was raised by the gas flow into the clearance. the clearance. The oil kept in the groove supplies oil to the clearance and also the refrigerant gas-oil l
42、abyrinth seal effect. To reduce the gas leakage and/or the zero leakage to increase the volumetric efficiency to the expected maximum, oil groove design Compression startOil volume rate velocity (m/s)gas velocity : maxabout 0m/secVelocity vectorCompression +45Compression +90Oil volume rateVelocity v
43、ectorOil volume rateVelocity vectorOil volume rateFig. 6 Table 5 Major specification or calculat ormation s f ion and solver infRefrigerant gas properties Lubricant oil properties Density ideal gas kg/m3(lb/ft3) Density 8 .2) 53 (53 kg/m3(lb/ft3) Cp 2 )150 (0.514 J/kgK(BTU/lbR) Cp 1 6) 950 (0.46 J/k
44、gK(BTU/lbR)Thermal conductivity 0.02 (0.0035) W/mK (BTU/fthF) Thermal conductivity 0.14 (0.025) W/mK (BTU/fthF) Viscosity 8 exp -6 (1.1 xp -9)9 ePa-s(lbf s/in2) Viscosity 0.0055 Pa(8.21 exp ) -7-s(lbf s/in2) Molecular weight 58 kg/kg-mol Surface tension 306 (677) N/m(lb/s2) Solver and code Segregate
45、 GAMBIT Ver. 2.04 d PISO-method Unsteady FLUENT Ver. 6.1.18 2011 ASHRAE 201Table 5 shows the major specifications for theanalysis is useful for grasping the oil behavior inprocess. Fig. 7 shows the two patterns of oil grooveFrom several analyses of oil groove patterns and locations, we determined th
46、at pattern As shape apprcalculation, the CFD solver, and the solver code. CFD the piston-cylinder bore clearance in the compression design with different locations. opriately seals the clearance with oil. Fig. 7 shows that the oil volume rate is kept at the compression stage compared with pattern B.
47、 OPTIMrooves. ctors as Multiple regression anal r the ctual compressors. To ressors. Figures 10(a), (b), and (c) show the multiple regression Pattern ACompression start Oil volume rateIZATION OF GROOVE SHAPE EXPERIMENTAL APPROACH Factor of oil groove shape The basic design of the oil groove was dete
48、rmined by CFD analysis (Fig. 7, PattFig. 8 dual oil grooves): the dual oil gDual oil grooves include such design fagroove locations L1 and, L2, depth, width, etc. ern A, ysis was experimentally used fodesign factor optimization to finalize the oil groove design. Experiments for groove design Experim
49、ents were done using a calorimeter with aclarify the results, we used a lower revolution speed and selected lower capacity models. Fig. 9 shows the major design of the piston and the cylinder block portions of the tested compFig. 9 Pattern ACompression +30 Oil volume ratePattern BCompression +30 Oil volume ratePattern BCompression start O
