ASHRAE LV-11-C024-2011 High-Efficiency Compressor Design for HC-600a Refrigerant Using Energy-Saving Household Refrigerators.pdf
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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
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