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編號 無錫太湖學院 畢業(yè)設計（論文） 相關資料 題目： 餃子機及傳動系統(tǒng)設計 信機 系 機械工程及自動化專業(yè) 學 號： 　　　 0923039　　　 學生姓名： 湯東鵬 指導教師： 　 戴寧 （職稱：副教授 ） （職稱： ） 2013年5月25日 目 錄 一、畢業(yè)設計（論文）開題報告 二、畢業(yè)設計（論文）外文資料翻譯及原文 三、學生“畢業(yè)論文（論文）計劃、進度、檢查及落實表” 四、實習鑒定表 無錫太湖學院 畢業(yè)設計（論文） 開題報告 題目： 餃子機及傳動系統(tǒng)設計 信機 系 機械工程及自動化 專業(yè) 學 號： 0923039 學生姓名： 湯東鵬 指導教師： 戴寧 （職稱：副教授 ） （職稱： ） 2012年11月25日 課題來源 自擬題目 科學依據(jù)（包括課題的科學意義；國內外研究概況、水平和發(fā)展趨勢；應用前景等） （1）課題科學意義 餃子食品機械的應用前景和發(fā)展現(xiàn)狀 餃子食品在我國歷史悠久，伴隨著幾千年的文明的發(fā)展已經(jīng)成為我國食品文化中的代表，如餃子、包子、餛沌是主食的一部分；湯圓、月餅、粽子是傳統(tǒng)節(jié)日中必不可缺的食物。如今，經(jīng)濟的迅速增長、人民生活水平的提高和生活節(jié)奏的加快，對食品行業(yè)提出了新的要求。而本人認為這些要求可以歸納為兩大類： 其一是食品的質量：如食用口感、衛(wèi)生狀況、營養(yǎng)含量等。 其二便是食品供應的速度。 而解決這兩個矛盾要求的辦法便是實現(xiàn)食品生產(chǎn)的機械化和自動化， 通過機械動作可以極大程度的提高食品的生產(chǎn)率； 采用環(huán)保的機械材料和嚴格的密封技術可以很好的保證食品衛(wèi)生；而合理的工藝編排更能改善食品的口感。 （2）餃子機的研究狀況及其發(fā)展前景 目前國內外廠家在包餡夾餡食品機械化上的研究已經(jīng)取得了一定的成果成功研發(fā)了餃子機、包子機、餛沌機、湯圓機、月餅機以及自動化程度更高的全自動萬能包餡機。 因東西方飲食文化的差異， 目前國外包餡成型類機械主要為日本所生產(chǎn)，如日產(chǎn)的自動萬能包餡機，其最大生產(chǎn)能力可達每小時 8000 個，且加工范圍極廣，能生產(chǎn)各式饅頭、包子、餃子、夾餡餅干、壽司、等等近百種產(chǎn)品，采用可拆卸料斗能實現(xiàn)快速更換餡料，內置的無級變速調控裝置可以實現(xiàn)皮和餡的任意配比。廣泛用于各種帶餡食品的加工。 而國內相關機械雖然在自動化和多功能方面較之日本產(chǎn)品還有一定的差距， 但是通過改革開放以后二十余年的發(fā)展亦取得了很大的進步。 以上海滬信飲料食品機械有限公司生產(chǎn)的水餃機為例：配備 1.1Kw 的電動機，生產(chǎn)效率達每小時 7000 個。已相當接近日產(chǎn)餃子機的生產(chǎn)水平。 每逢過時過節(jié)現(xiàn)做現(xiàn)賣餃子往往出現(xiàn)供不應求的現(xiàn)象。當然也有很多人選擇在家里自己做， 卻需要提前半天甚至一天進行準備，而包餃子的時候更是要叫上好幾個親朋過來幫忙方可。 因此如果能研究開發(fā)一種能夠以機械動作代替人工勞動的機器， 那么除了可以節(jié)約大量的時間、降低餃子的生產(chǎn)成本、提高利潤之外，更可以免除人們冬日里冒寒排隊購物之苦，一舉多得。餃子生產(chǎn)機的初步目標確定為能夠實現(xiàn)餃子包餡成型工藝的機械化。 未來可在此基礎上加以改進和擴展，以實現(xiàn)橫縱兩方向發(fā)展。即餃子生產(chǎn)全過程的無人干預自動化與多功能化。 研究內容 ① 熟悉餃子機的工作原理與結構； ② 熟悉餃子機傳動系統(tǒng)的布置與結構； ③ 熟練掌握傳動系統(tǒng)的設計計算方法； ④ 掌握ＣＡＤ的使用方法； ⑤ 能夠熟練使用ＵＧ進行三維的畫圖設計。 擬采取的研究方法、技術路線、實驗方案及可行性分析 （1）實驗方案 對餃子機整體設計，擬定其傳動部分的結構、轉速等，使其能夠半自動的進行加工。 （2）研究方法 ①用ＣＡＤ進行二維畫圖，對餃子機結構有個全面的了解。 ② 對餃子的傳動部分進行計算與結構設計，使其提供合適的動力。 研究計劃及預期成果 研究計劃： 2012年10月12日-2012年12月31日：按照任務書要求查閱論文相關參考資料，完成畢業(yè)設計開題報告書。 2013年1月1日-2013年1月27日：學習并翻譯一篇與畢業(yè)設計相關的英文材料。 2013年1月28日-2013年3月3日：畢業(yè)實習。 2013年3月4日-2013年3月31日：餃子機傳動系統(tǒng)計算和總體結構設計。 2013年4月1日-2013年4月14日：傳動箱設計。 2013年4月15日-2013年4月28日：零件圖及三維畫圖設計。 2013年4月29日-2013年5月21日：畢業(yè)論文撰寫和修改工作。 預期成果： 達到預期的畢業(yè)設計要求，設計出的餃子機可以進行半自動加工，可以快速美觀的加工出餃子，并且傳動簡單緊湊、滿足工作要求。 特色或創(chuàng)新之處 ① 餃子機可以無需手工進行制作。 ② 餃子制作過程安全，方便，快速，可以批量生產(chǎn)。 ③ 傳動路線簡單、緊湊，滿足餃子加工的要求。 已具備的條件和尚需解決的問題 ① 設計方案思路已經(jīng)明確，已經(jīng)具備機械設計能力和餃子機方面的知識。 ② 進行結構設計的能力尚需加強。 指導教師意見 指導教師簽名： 年 月 日 教研室（學科組、研究所）意見 教研室主任簽名： 年 月 日 系意見 主管領導簽名： 年 月 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 ↓ Wm=f（t） Wf =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 FVA, 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 FVA3, 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. Phyd+Pfric-Qcomp-Qfluid=0 (1) Qfluid=mcfluid△Tfluid (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 i