饺子机及传动系统设计毕业设计论文

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无锡太湖学院

毕业设计（论文）

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题目：饺子机及传动系统设计

信机系机械工程及自动化专业

学号：0923039

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指导教师：（职称：副教授）

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2013年5月25日

目录

一、毕业设计（论文）开题报告

二、毕业设计（论文）外文资料翻译及原文

三、学生“毕业论文（论文）计划、进度、检查及落实表”

四、实习鉴定表

无锡太湖学院

毕业设计（论文）

开题报告

题目：饺子机及传动系统设计

信机系机械工程及自动化专业

学号：0923039

学生姓名：

指导教师：（职称：副教授）

（职称：）

2012年11月25日

英文原文

wear 181-183 (1995) 868-875

Case Study

Theoretical and practical aspects of the wear of vane pumps

Part B. Analysis of wear behaviour in the Vickers vane pump

test

A. Kunz a, R. Gellrich b, G. Beckmann c, E. Broszeit a

a Institute of Material Science, Technical University Darmstadt, P.O. Box

11 1452, 64229 Darmstadt,Gcmb University for Technol08y, Economy and Social Science Zittau/Goditz, Facuky of Maihematics, P.O. Box 264, 02763 Zutau

cPetersiliensrr. 2d, 03044 Cottbus, Received 16 August 1994; accepted l November 1994

Abstract

The wear behaviour of the vane pump used in the standard method for indicating the wear characteristics of hydraulicfluids (ASTM D 2882/DIN 51 389) has been examined by comparison of the calculated wear and experimental data using alubricant without any additives. In addition to the test series according to DIN 51 389, temperature profiles from the pump have been analysed using the bulk temperatures of the contacting components and the temperature in the lubrication gap as input data for the wear calculation. Cartridges used in tests according to the Gennan standard have been examined extensively before and after each run to obtain input data for the mathematical model and to Jocate wear. An analysis of the :tluid properties and an investigation of the innuence of wear particles in the hydraulic circuit were performed. The experimental results were compared with the wear prediction, which was verified by the agreement in terms of load, temporal wear progress and local wear. Conclusions have been drawn with regard to the validity of the load assumptions and wear calculation, as well as to the limits of applicability of this method in the presence of additives.

Keywords: Vane pumps; Hydraulic fluids; Wear prediction; Vickers vane pump test

1. Introduction

Efforts to develop a mathematical tool for wearprediction will not be successful without considering wear and its phenomena. The task of Part B of this study is to describe the analysis of the wear behaviour in the tribo system investigated and how the knowledge achieved influences the calculations. Input data are derived from the measurement of mechanical and geometrical quantities, such as the hardness, stylus profilometry, fluid properties and contact radii. Thermal quantities are also essential for the modelling of lubrication. The calculations must be verified with wear data. Because the tribo system to be analysed is the vane pump employed in the Vickers vane pump test,which has been in use for about 40 years, several wear data can be used for comparison between calculated and measured wear results. These are the wear masses0043-1648/95/$09.50@ 1995 Elsevier Science S.A. All rights reserved SSDI 0043-1648(94)07087-3 after each tcst run, the progrcssion of wear over time and the local wear on

the inner ring surface; in combination, these enable a comprehensive statement to be made on the validity of the mathematical model described in Part A.

2. Experiments

AlI Vickers vane pump tests described were run with the same fiuid. It is a reference oil of the German Rcscarch Association for Transmission Technique (FVA), and is a mineral oil without any additives (FVA3). Thus the disturbing influences of additives can be excluded.

2./.Input data for calculation

Fig. 1 lists the input and output quantities of the calculations. Most of the input parameters were derived surface profiles contact force and contact velocity dynamic viscosity contact radiihardness values Youngs moduli, Poisson numbersand lubrication gapspecific shear energy densities* pressure exponentc,f viscosity; tlubrication gap temperature

Rough surfuce ←→ shaar energy hypot ←→ elasto liubiction

↓

W m=f（t）

W f =f（ɑ）

Fig. 1. Input parameters and output quantities of the mathematicalmodel of Part A.

Fig. 2. Cartridge V 104 C: bushing, rotor, ring, bushing (abcwe),single vane, pin (below).experimentally from all the components involved beforeand after use in the vane pump tests. The mechanical components, which must be renewed for each test run,are shown in Fig. 2. Such a cartridge kit consists of a rotor, ring, 12 vanes, bushings and pin.

Stylus profilometry was performed on the inner surface of the ring and on the tips of two vanes of the cartridge before and after each test run. Earlier investigations have shown that ten parallel sections in the sliding direction on each body are sufficient to describe the surface topography in a statistically satisfactory manner as a two-dimensionalisotropic gaussianfield according to Ref. [1]. Only the high pass filtered components of the profile (sampling length, 1.5 mm; cut o五0.25 mm) were used to determine the spectral moments mo, m2, m4 and the parameter of roughness a. According to the partition of the contact force into different loading zones, the topographic data of the new surfaces were used for zone IV (low level load, see Part A). For the other zones with higher contact forces, the profiles of the surfaces in the final condition were used, which corresponds to the appearance of the inner ring surface after the test runs.

The contact force and contact velocity were calculated with different fluid pressures and dynamic forces acting on the vanes, revolution number and ring radu, whereas the change in contact radius was documented with a profile projector. Because the ring radii are much larger thar) the radii of the vanes in the contact zone, the vanes can be assumed to be hertzian cylinders sliding along a plane surface and the contact radii are simply the radii of the vane tips. Each vane tip was twice drawn up at magnifications of 100 : 1 and the contact radii and contact locations were measured with a stenciLMean values of the contact radii were transferred to the calculation, which

is based (similar to the surfaceprofiles) on vanes in both conditions.

The Vickers hardness HVlO was measured on thering and three vanes of each cartridge. This hardnessleads to a better reproducibility than microhardness values, but due to the large indenter load, it couldonly be taken after the test runs. Therefore changes in hardness values could not be registered.

The Young's moduli, Poisson numbers and densities of the ring (AISI 52100) and vane materials (M2 reg C) are the first input parameters in the shear energy hypothesis and were obtained from the literature. The specific shear energy densities (see Part A) are materialspecific constants [2l. The fluid properties (Fig. 1) were measured, derived from the literature or calculated. To obtain the dynamicviscosity, the densities and kinematic viscosities at 20,40 and 80 0C were measured. Because the fluid is a reference oil of FV A, the pressure exponent of the viscosity is given [3]. The temperature in the lubrication gap between the ring and vanes was approx:imated by measurements and calculations described below.

2.2. Temperature profiles

Temperature measurement was performed to obtain information on how a heatable tribometer must be controlled to simulate the wear behaviour of the vane pump. Therefore shortened test runs were carried out until temperatures were stabilized. These 10 h vane pump tests delivered the input data for the approximation of the lubrication gap temperature in the ring-vane contact, as well as additional wear masses to be compared with the calculated progressiort of wear in time. The sampling principles for acquiring the temperature profiles of the vane pump are illustrated in Fig.

3.

The temperature of the lubricant in the gap between the ring and vanes was estimated to be equal to or greater than the bulk temperature on the inner ring surface. Following the first main statement of thermodynamics, the heat flux Q mp into the components of the pump can be derived from with the fluid as the medium for energy transport.Qa,mp can only be transferred to the components shownin Fig. 2. For the same temperature differences and materials, this heat nUX can be divided into single component fluxes ac cording to the relation of masses. The derived flux Qring is the heat which flows in a certain time period in a radial direction through the ring. With the known temperatures on the outer ring surface, the bulk temperatures on the inner ring surface can be calculated and transferred to the model of elastohydrodynamic lubrication.

All test runs with the Vickers vane pump V 104 C were performed on a test rig according to ASTM D2882/DIN 51 389, which is shown schematically in Fig.

4. These standards describe the procedure for testingthe anti-wear properties of hydraulic fiuids. To start the Vickers vane pump test according to the German standard, the system pressure must be raised in steps of 2 MPa every 10 min, beginning at 2 MPa, until a final pressure of 14 MPa is reached. At this stage, the fluid temperature measurcd bcfore the pump (see Fig.4) must be controlled to guarantee a kinematic viscosity of 13 mm2 S-i at the inlet for every :tluid tested. These conditions must be maintained until the test is aborted normally after 250 h by opening the bypass of the pressure control valve before the motor is stopped. By a comparison of the wear achieved on the ring and vanes with the upper wear limits, the anti-wear properties of the fluid

tested can be derived.

For performing the tests safely with the fluid FV A3, it was preheated t0 40 0C and circulated in a pressurefree way. The damage which may occur during the critical first hour of the runs can be avoided using TiNcoated bushings [4]. For comparison with the results derived from computation, the wear produced in these runs must be documented as amounts, both locally and temporally. The wear masses were derived from the weight differences of the ring and vanes before and after each run. They were obtained from a sequence of four 250 h test runs and tw0 10 h runs for temperature measurement. The local linear amount of wear was documented by the differences in the inner ring radii perdegree of revolution, which were measured by surface digitization along the inner ring surface at three different positions of the ring width before and after the tesi runs.

In earlier investigations [5], the wear progression over time of the vanes was measured under identical testing conditions, except for a lower fluid temperature. For this experiment, the radiotracer technique was used. Two vane tips in the set of 12 vanes of each cartridge were radiologically activated by bombardment with protons. A detector close to the pump body allowed thedecrease in radiological activity to be monitored continuously, which was found to be reciprocally proportional to the linear amount of vane wear as a function of time [5l. Due to the good tempering properties of the vane material (M2 reg C), with a specific secondary hardness maximum between 450 and 550 0C, the infiuence of the activation process at 220 0C on the wear behaviour of the activated zone of the vane tips could be excluded.

P hyd+P fric-Q comp-Q fluid=0 (1)

Q fluid=mc fluid△T fluid (2)

Fig. 4. Hydraulic circuit of the test rig.

3 result

lines the statistical reliability of surface modelling as a two-dimensional isotropic gaussian field. Although only the filtered profiles scanned in the sliding direction are shown, a distinct change in surface roughness is obvious. A good representation of the wear phenomena (see Part A) by the input data for the wear calculation derived from these profiles can be assumed.

The change in the vane tip shape over the testing period is documented in Part A. The hardness values for the rings and vanes varied from 743 t0 769 HVlO (rings) and from 778 t0 816 HVlO (vanes). In all cases, the vanes of one cartridge had higher hardness values than the ring, but these differences varied and had a large influence on the wear calculation (see Part A).

The measurement of the fiuid properties led, in combination with the kinematic viscosity prescribed by the German standard, to a fluid temperature of 84-86oC at the pump inlet. Together with the other temperature measurements acquired in the 10 h runs, these temperature profiles are illustrated in Fig. 6.

Test Number t was found that, in about l h, all temperatures were stabilized. It should be noted that all temperatures in or on the pump components are higher than the fluid temperature measured behind the pump. The highest temperatures were found on the outer ring surface,

with significant differences depending on the location of the thermocouples.

The calculation of the bulk temperatures on the inner ring surface via the heat flux balance eliminated the infiuence of the different ring thicknesses at the scan locations. Depending on tbese different distances for heat conduction, between 4 and 7 0C must be added to the mean values of the component temperatures to obtain the surface temperatures. These values are 20c70 higher than the fluid temperature measured behind the pump, which was used as input data for the wear calculation.

During the l h starting phase of the test runs, the stepwise increase in system pressure leads to an immediate effect on the component temperatures, whereas the fluid temperature increases with a more or less constant gradient, which demonstrates the association of load and frictional heat. The four 250 h test runs caused a mixture of adhesive and abrasive wear at a high level (see Part A). The wear results achieved are shown in Fig. 7. Ring wear increased from test I to test 3. Therefore the 12 pm filter normally used was replaced after the third test by a 3 pm filter, and a pressure-free run with an additional cartridge was started as a cleaning procedure. Due to the filter change, the reservoir needed to be refilled by about lOv-/o of its content with fresh fluid before control test 4, again with a 12 ym filter, was started. In addition to these efforts to minimize possible wear particle influence, a comparison of the viscosity and neutralization number with those of fresh fluid showed only an insignificant rise in viscosity and a low neutralization number after 750 h of testing. In test 4, the highest value for ring and vane wear at a constant level was achieved. For all tests, the linear amount of wear on the ring surface showed a strong dependence on the measurement location with strictly limited areas of high and low wear.

The results of continuous vane wear monitoring are shown in Fig. 8 in addition to the principle of measurement. Degressive wear laps were found, where the stationary level was reached after 100 h.

4. Discussion

Before the wear calculations can be verified by wear data, it must be demonstrated that the assumptions,measurements and calculations forming the input for the mathematical model correlate with the wear measured.Fig. 9 compares the calculated load n the ring-vane contact, derived from the contact force and changing shapes of the vane tips introduced in Part A, with the measured linear amounts of wear along the inner ring surface and the temperature distribution at the same place. There is qualitatively good correlation for the progression of load and wear with characteristic leaps at almost the same degree of revolution. In addition, high temperature, resisting dynamic equilibrium, is found where the load and wear are high and vice versa. Therefore it is absolutely correct to create different loading zones (according to fig. 2 in Part A) as input for the wear calculations. Although a few differences in quality can be found in the pro- gression of hertzian pressure and the linear amounts of wear, serious mistakes in the collection of input information are probably avoided, so that the verification of the calculated wear results by experimental data will show the validity of the mathematical model.

For local amounts of linear ring wear, this verification can be seen in Fig. 10. It should be noted that the calculation and experimental results are placed in the same decade, the progressions show the characteristic leaps similar to the load in Fig. 9 at almost the same degrees and the amounts are directly comparable. The loading zones are adapted to the progression of the contact force (see Part A), which the calculated linear wear must follow as well as the hertzian pressure. The different shapes of the two graphs between 300 and 700 (2100 and 2500) turn angles remain

unsatisfactory, because this shows an uncertainty in the load as sumptions. The fluid pressure in a cell formed by two vanes, rotor and ring was assumed to be segmentally constant. Therefore the contact force was determined to follow these assumptions, which need to be dempressure) in the ring-vane contact, derived from the contact force and changing shapes of the vane tips introduced in Part A, with the measured linear amounts of wear along the inner ring surface and the temperature distribution at the same place. There is qualitatively good correlation for the progression of load and wear with characteristic leaps at almost the same degree of revolution. In addition, high temperature, resisting dynamic equilibrium, is found where the load and wear are high and vice versa. Therefore it is absolutely correct to create different loading zones (according to fig. 2in Part A) as input for the wear calculations. Although a few differences in quality can be found in the progression of hertzian pressure and the linear amounts of wear, serious mistakes in the collection of input information are probably avoided, so that the verification of the calculated wear results by experimental data will show the validity of the mathematical model.

For local amounts of linear ring wear, this verification can be seen in Fig. 10. It should be noted that the calculation and experimental results are placed in the same decade, the progressions show the characteristic leaps similar to the load in Fig. 9 at almost the same degrees and the amounts are directly comparable. The loading zones are adapted to the progression of the contact force (see Part A), which the calculated linear wear must follow as well as the hertzian pressure. The different shapes of the two graphs between 300 and 700 (2100 and 2500) turn angles remain unsatisfactory, because this shows an uncertainty in the load assumptions. The fluid pressure in a cell formed by two vanes, rotor and ring was assumed to be segmentally constant. Therefore the contact force was determined to follow these assumptions, which need to be dem-

onstrated by corresponding experiments. Comparedwiththe measurement, the contact force in loading zone I was assumed to be too high and caused values above the experimental data. This was due to difficulties in modelling the large variety of vane tip geometries which can appear in one cartridge and strongly determine the contact force in this region. More information about

the reliability of load assumptions could have been obtained from a knowledge of the bulk surface temperatures, which were not measured or calculated.Despite other deviations of the two progressions, the areas below each graph are comparable, so that a good correlation between calculated and measured wear masses can be expected.

For vane wear masses after 250 h of testing, the expectations have been fulfilled in a satisfactory manner, If the progression of the vane wear mass in time (see Part A) is verified by the measured amounts of linear wear (Fig. 8), good correlation of the progressions can be found. For both values, the stationary phase is reached after 100 h. The differences in the gradients during the stationary phase may be caused by the different tcmperatures of the two test series. The dif- ferences in the ring wear masses can be interpreted as scattering (which is dependent on the testing procedure), because the thermal aging and the influence of wear particles can be neglectcd, especially if the recharge with fresh fluid is taken into consideration. Thus thesC wcar results arc also well approximated by the calculation, because the calculated average wear mass aftcr 250 h is placed in the middle of the four experimental results (see Part A). This is supported by the fact that the wear masses achieved in the short-time runs are situated close to the calculated progression

of the wear masses for that time.