機(jī)械設(shè)計外文資料翻譯--活塞油膜和組件變形在往復(fù)壓縮機(jī)中相互作用的數(shù)值分析（英文）

1、A numerical analysis of the interaction between the piston oil film and the component deformation in a reciprocating compressorJ.R. Cho*, S.J. MoonSchool of Mechanical Engineering, Pusan National University, Jangjeon-Don

2、g, Kumjung-Ku, Busan 609-735, South KoreaReceived 27 December 2003; received in revised form 10 May 2004; accepted 6 October 2004 Available online 23 November 2004AbstractThe piston secondary motion significantly influen

3、ces the major characteristics of lubrication in a reciprocating compressor, such as the oil leakage, the piston slap phenomenon and the frictional power loss. Therefore, the design parameters governing piston dynamics sh

4、ould be carefully determined based upon a reliable dynamic characteristic investigation. As a preliminary research step, this paper is concerned with the finite element analysis for the piston dynamic response. By coupli

5、ng FDM for the lubricating pressure field with FEM for the piston dynamic motion, we numerically approximate the lubricant–structure interaction in a reciprocating compressor. Numerical results illustrating the theoretic

6、al work are presented. q 2004 Elsevier Ltd. All rights reserved.Keywords: Ringless small reciprocating compressor; Lubricant–structure interaction; Piston secondary motion; Eccentricity and tilt; FEM and FDM1. Introducti

7、onRingless small reciprocating compressors are widely used to compress coolant gas in household refrigerators and air-conditioners. In such devices the piston becomes a key component influencing all the major performance

8、s of reciprocating compressor, such as pumping efficiency, noise, power consumption, anti-wear, and so on. It is because piston dynamics characterizes the oil leakage, the piston slap phenomenon [1] and the frictional lo

9、ss, which determine such major performances. While moving up and down along the lubricated cylinder wall, a piston displays oscillatory radial translation and rotation within the oil film clearance owing to the unbalance

10、 in dynamic forces acting on it. This secondary motion in piston dynamics has become a crucial research subject in order to improve the performance and stability of reciprocating compressor [2,3]. The piston secondary mo

11、tion is associated with several design parameters, such as the radial clearance, the lubricant viscosity, the wrist-pin location, the crankshaft eccentricity, and the piston skirt profile [4,5]. As well, a piston in ring

12、lessreciprocating compressors is subject to lubricating pressure and frictional force, besides the primary coolant pressure and the motor-driven force. So, the piston dynamic motion is strongly influenced by the geometri

13、c structure and the oil film pressure. Therefore, the above-mentioned design parameters should be carefully determined, in order to maximize the dynamic performance and stability, based upon a parametric dynamic investig

14、ation, which would be achieved by an appropriate coupled numerical analysis, such as one for general fluid–structure interaction problems [6,7]. According to our brief literature survey, Li et al. [8]investigated the lub

15、rication characteristics theoretically and experimentally, and found the effect of the wrist-pin location on the frictional force. Traditionally, the lubricating pressure has been modeled by Reynolds equation by assuming

16、 the oil lubricant be a Newtonian iso-viscous fluid. On the other hand, the piston dynamics has been mostly described by the particle dynamics equations for the piston mass center by replacing external surface tractions

17、with equivalent resultant forces and moments [9–11]. Furthermore, both piston and cylinder were assumed as rigid bodies. Even though this particle dynamics approach makes the problem simpler and reduces the computation t

18、ime,0301-679X/$ - see front matter q 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.triboint.2004.10.002Tribology International 38 (2005) 459–468www.elsevier.com/locate/triboint* Corresponding author. Tel.: C82 51

19、 510 2467; fax: C82 51 514 7640. E-mail address: jrcho@hyowon.pusan.ac.kr (J.R. Cho).On the other hand, b0 indicates the inclined angle of the connecting rod force to the vertical axis owing to the piston eccentricity in

20、 the x-direction. We note that the effect of the piston tilt about the wrist pin is neglected. Then, the connecting-rod force components and the inclined angle are determined asFyðfÞ Z mpap CpR2ðpg Kpa

21、2; CFpf (4)FxðfÞ Z Fpx (5)b0 Z tanK1ðFx=FyÞ (6)where mp denotes the total piston mass.3. Displacement and lubricating pressure fieldsLet u(x;t) be the displacement field of the piston and cylinder, th

22、e structural dynamic response is governed bysijðuÞ;j Cfi Z r€ ui (7)with initial and boundary conditionsuiðx; 0Þ Z 0 (8)uiðx; tÞ Z 0; sijnj Z ti (9)where r denotes the mass density of the st

23、ructural components and ti the traction components by the coolant gas and lubricating oil pressures. In order to describe the lubricating pressure field within the radial clearance, we introduce a cylindrical co-ordinate

24、 attached to the center of the piston top surface, as depicted in Fig. 3. The y-axis directs to the same direction as one in the previous fixed Cartesian co-ordinate system. Referring to Fig. 1(b), the piston is allowed

25、to move only in the x-direction and to tilt about the wrist-pin axis. The piston axis eccentricity is denoted by e while the tilting angle by g. We assume that the oil film is always 100% full within the radial clearance

26、 over whole 3608 and axial length ofthe piston. The lubricating oil flow is assumed to be incompressible laminar because two characteristic lengths H and R are significantly larger than the flow thickness. And, the press

27、ure variation in the r-direction is ignored because the radial clearance is much smaller than the piston radius. By neglecting the body and inertia forces of lubricating oil, the lubricating pressure field p(y,q) is gove

28、rned by Reynolds’s equation which is based upon the incompressible Navier–Stokes equations and the continuity conditionvvy h3 vpvy? ?C vR2vq h3 vpvq? ?Z 6Vpm vhvy (10)equipped with the boundary conditions given byp Z pg

29、at y Z 0; p Z pa at y Z H (11)vpvq Z 0 at q Z 0 and p (12)where m refers to the oil viscosity. By denoting yw be the wrist-pin location, the actual oil thickness h is expressed byhðy; qÞ Z c K½e Cðyw

30、KyÞg?cos q (13)Referring to the sign convention for forces shown inFig. 2, two resultant forces Fpx and Fpf are calculated such thatFpx ZðH0ð2p0 pðy; qÞcos qR dqdy (14)Fpf ZðH0ð2p0mVp h

31、 C h2vpvy? ?R dqdy (15)On the other hand, the instantaneous volumetric oil leakage Qf(f) throughout the radial clearance is calculated byQfðfÞ Zð2p0hVp 2 K h312mvpvy? ?yZH R dq (16)And, the cycle-averaged

32、power consumption P is calculated according toP Z 12pð2p0 PfðfÞdf (17)Fig. 3. Lubrication within the radial clearance between piston and cylinder.Fig. 2. Free body diagram of the piston.J.R. Cho, S.J. Moon