外文翻譯--基于有限元法旋耕機傳動齒輪應力分析

基于有限元法，旋耕機傳動齒輪應力分析 穆罕默德 托帕克 庫薩特 西里克 丹妮資 耶爾馬茲 易卜拉欣 阿辛琪 阿克登尼基大學農(nóng)學院，安塔利亞，土耳其 農(nóng)業(yè)機械部 2008年 8月 12日 摘要 旋耕機的耕作工具，獲取自己的運動由拖拉機動力起飛（ PTO）有被設計為混合土。降低土壤交通在很大程度上與此工具混合土。使用旋耕機是提高我國由于其許多優(yōu)點。旋耕機結構具有一個齒輪箱，改變運動方向由拖拉機動力輸出軸 90度，旋轉速度傳動齒輪和轉子軸放置在水平的土壤混合。有刀片在進入轉子軸件和混合土。特別是，在刀片和傳動 齒輪，變形發(fā)生由于高無振動，高功率，土壤的部分影響，使用條件設計制造誤差和錯誤。特別是用于建筑和傳動部件的應力分布，為理解好的確定失敗的原因。在這項研究中，傳動的旋耕機而設計制造的一種本地制造商被建模為三維參數(shù)化設計軟件和結構應力在根據(jù)其工作使用有限元軟件模擬了在傳動齒輪的分布條件。后仿真結果評價，對齒輪應力分布表明，齒輪工作無故障根據(jù)齒輪的材料應力屈服。此外，計算的參考齒輪工作安全系數(shù)仿真結果。 關鍵詞： 旋耕機，應力分析，有限元法 1. 引言 旋耕機耕作機， 適 用于農(nóng)田、果園 ， 在農(nóng)業(yè)。旋耕機有削 減巨大的能力， 混合土和苗床準備直接制備。此外，旋耕機有更多的混合能力是比犁的七倍。 旋耕機是連接到拖拉機三點聯(lián)動系統(tǒng)，它是由拖拉機動力輸出軸驅(qū)動（ PTO）。運動的方向改變 90 度，從拖拉機動力輸出第二齒輪箱的水平軸。轉子軸與第二齒輪箱的運動。 旋耕機的元素在 其他作用力下 由于高無振動，高功率，土壤的部分影響，設計制造錯誤和錯誤的使用條件，耕作。因此，不需要的應力在它的元素分布。如果元素不能補償?shù)牟倏v力，這些元素變得毫無用處，因為打破或大變形破壞。特別是葉片及傳動元件必須耐用 于操縱力下 。應力分布的預測是非常重要 的無故障產(chǎn)生的良好工作的設計和產(chǎn)品設計師和制造商。機的廠家，要為自己的機器，可能的錯誤，防止使用的材料，具有很高的安全系數(shù)，或者他們使用高質(zhì)量的機械元件。雖然這些措施可以安全，產(chǎn)品的重量和成本的上升。 幫助開發(fā)的技術和設計軟件，集成在新的代計算機，設計變得更加方便和可靠的。設計師可以設計在虛擬屏幕上自己的產(chǎn)品和他們可以利用計算機仿真技術，評價產(chǎn)品的工作條件。今天的三維（ 3D）與有限元法的應用中越來越廣泛的建模工業(yè)。許多三維建模和有限元的應用實例可以在不同的工程學科（ 谷內(nèi) ，見1993）。 在這項研究中， 一種旋耕機傳動齒輪火車，這是由本地制造商制造，采用Solid Works三維參數(shù)化建模設計軟件。三維建模后的程序，進行了模擬研究 在利用 COSMOS Works 有限元軟件傳動齒輪火車。旋耕機和第二齒輪箱傳動輪系及其三維模型在圖 1中給出的。此外，圖 2 顯示了一個架構，是屬于旋耕機傳動系統(tǒng)（不同等人。， 2005）。如下圖所示架構的運動和動力的傳遞與拖拉機動力輸出萬向節(jié)連接到第一齒輪箱有 2個螺旋錐齒輪的齒數(shù)的 10和 23，然后 到 第二齒輪箱軸。 2。材料和方法 2.1三維建模及應力分析 傳動齒輪 根據(jù)齒輪傳動齒輪 的原尺寸模型 ， 然后他們聚集。通過圖 3可以看到他們的 3D模型和它的值在表 1中給出的。開始的應力分析，我們認為，在正常工作條件下工作的齒輪。 在耕作，與旋耕機的操作，所需的拖拉機動力輸出功率為 49.5千瓦拖拉機動力輸出革命是 540分鐘，根據(jù)拖拉機動力輸出功率和傳動比的齒輪，時刻 齒輪已經(jīng)占用 。 表 1。傳動齒輪的值 傳動齒輪 的值 齒輪 I 齒輪 II 齒輪 III 模塊 mm 齒數(shù) - 面寬度 - 軸直徑 mm 力矩 Nm 6 6 6 31 43 38 38 38 38 55 82 55 373.00 263.56 292.41 在模擬中，兩種分析各齒輪副齒輪產(chǎn)生（ I-II和齒輪 II-III）對工作條件。分析了已生成的三維，靜態(tài)和線性 COSMOS Works有限元軟件的假設。各向同性材料屬性中使用的齒輪材料的模擬和性能的了表 2（ 庫塔 ， 2003）。裝配時，值得注意的是，在接觸工作齒輪的齒，配對就在彼此接觸條件下的單。因為，實驗結果表明，對齒輪的表面發(fā)生的最大應力和失效對齒輪接觸區(qū)和齒根單接觸條件（ 庫股 ， 1993）。 表 2。齒輪的材料特性 材料 DIN C45 彈性模量 GPa 拉伸強度 MPa 屈服 強度 h MPa 泊松比 - 密度 kg/m3 211 700 500 0.30 7850 2.2齒輪 1和齒輪 II之間 的 應力分析 齒輪和齒輪 II 我組裝后，施加邊界條件。齒輪 II固定于其軸軸承。占矩值法旋轉軸方向的網(wǎng)格構造齒輪我在圖 4中可以看到。 COSMOS Works嚙合的功能已被用于地圖網(wǎng)格。高階（二級）的拋物型固體四面體單元具有四個角節(jié)點， 六中間節(jié)點，六邊的高質(zhì)量的網(wǎng)格劃分功能（ COSMOS工程建設， 2006） 。 在嚙合操作，共 342160個元素和獲得 489339 個節(jié)點的包含，總共 對 于 嚙合的齒輪 1和齒輪 II 來說 。 在求解過程中，應力分布如圖 5所示的了 ， 對齒輪和齒輪 II。作為一個結果的最大等效應力（ von米塞斯） ， 確定對齒輪工作齒的接觸面為我和 123.59 MPa，73.98 MPa時最大等效應力 由 齒輪工作齒 II 確定。 2.3 齒輪 II 和 齒輪 III 的應力分析 在這一部分中，同樣的必要程序，應用應力分析齒輪 II和齒輪 III施加邊界條 件，生成的網(wǎng)格劃分和求解程序。齒輪 III被固定在軸承和占力矩值應用于齒輪 II。在嚙合操作模型，總共有 326600 個元素 468512總節(jié)點嚙合齒輪 II和 III總齒輪（圖 6）。 結果圖顯示對齒輪 II和 III在圖 7齒輪。分析結果表明，最大等效應力發(fā)生在接觸表面加工齒輪 III為 47.13 M Pa。根據(jù)施加的力矩 46.37 M Pa的等效應力值對齒輪 II齒面接觸區(qū)發(fā)生工作。 得到的仿真結果表明，我們是如何工作的分布應力傳動齒輪齒。根據(jù)仿真結果和齒輪材料的屈服應力，工作安全系數(shù)占傳動齒輪（表 3）。 表 3。傳 動齒輪的工作安全系數(shù) 傳動齒輪 屈服應力產(chǎn)量范圍 M Pa 馮米塞斯 von M Pa 安全系數(shù)。 K coeff. = yield / von 齒輪 I 500 123.59 4.05 齒輪 II 500 73.98 6.76 齒輪 III 500 47.13 10.60 3。結論 在這項研究中，對一種旋耕機，由本地制造商制造的傳動齒輪的應力分布進行了模擬。為了這個目的，傳動齒輪進行了建模和結構應力分析的產(chǎn)生 利用 Solid Works 三維參數(shù)化軟件 COSMOS Works有限 元軟件。 根據(jù)仿真結果，以下的信息可以說； 1.當傳動齒輪進行了仿真結果屈服應力的材料的齒輪，齒輪無故障檢測。齒輪工作在正常條件下。 2.在應力分析齒輪 I和齒輪 II之間，最大等效應力在確定齒輪齒的接觸面為工作 123.59 M Pa。在相同的齒輪工作齒結果 II 73.98兆帕的應力值對接觸表面。 3。在應力分析齒輪 II和 III之間的齒輪，最大等效應力確定對齒輪 工作齒接觸面為 47.13 M Pa。確定了 46.37 M Pa的接觸面齒輪 II工作牙最大應力值。 4。根據(jù)模擬結果對齒輪的最大應力，工作安全系數(shù)占齒輪 齒輪齒輪 II和III，如表 3。 使用具有安全系數(shù)高的材料要容易應用設計者。但這種方式去過多的成本上升，重量和時間。避免這些結果，仿真技術和計算機軟件的使用準備好的設計師，是如此有用的工具和應用程序，以獲得時間和制造成本。此外，它是可能增加的素質(zhì)和能力最佳的機械和工具在農(nóng)業(yè)機械化系統(tǒng)的設計。 參考文獻 耶爾馬茲，博士， 卡納克基先生 ， 2005。一種旋耕機齒輪失效。工程失效分析，12（ 3）： 400 404。 2006軟件 COSMOS Works幫助文件， 2006。 COSMOS Works用戶指南。 庫股 ， 1993。機械元件?？瀑Z埃利 大學 出版社，第二卷，科賈埃利（土耳其）。 庫塔 ， M.G.， 2003。指導制造商。 比爾深 出版社，伊斯坦布爾（土耳其）。 谷內(nèi) ， D.， 1993。有限元原理方法工程師。（翻譯），薩卡里亞大學出版社， No.03，薩卡里亞（土耳其）。 奧美資 ， A.， 2001。園林植物的機械化。阿克登尼基大學出版社： No.76，安塔利亞， （土耳其）。 STRESS ANALYSIS ON TRANSMISSION GEARS OF A ROTARY TILLER USING FINITE ELEMENT METHOD Mehmet TOPAKCI a H.Kursat CELIK Deniz YILMAZ Ibrahim AKINCI Akdeniz University, Faculty of Agriculture, Department of Agricultural Machinery, Antalya, Turkey Accepted 12 August 2008 Abstract: Rotary tiller is one of the tillage tools which gets own motion from tractor power take off (PTO) and it had been designed for blend to soil. Soil traffic is decreased to great extent with this tool by blending the soil. Using of rotary tiller is increasing nowadays in our country because of its many benefits. Rotary tiller construction has a gear box that changes motion direction with 90 degrees from tractor PTO, transmission gears for rotation velocity and a rotor shaft which placed as horizontal to soil for blending. There are cutter blades on rotor shaft for breaking into pieces and blend to soil. Especially, on cutter blade and transmission gears, deformations occur because of high vibration, pointless high power, impact effect of soil parts, design-manufacturing error and wrong using conditions. Especially for construction and transmission parts, stress distributions should be determined well for understand failure reasons. In this study, transmission gear train of a rotary tiller which was designed and manufactured by a local manufacturer was modeled as three-dimensional in a parametric design software and structural stress distributions on transmission gears were simulated using a finite element method software according to its operating condition. After evaluating of simulation results, stress distributions on gears show that gears working without failure according to yield stress of gears materials. Additionally, working safety coefficient of gears calculated by reference simulation results. Keywords: Rotary Tiller, Stress Analysis, Finite Elements Method 1. Introduction Rotary tiller is a tillage machine which is used in arable field and fruit gardening agriculture. Rotary tiller has a huge capacity for cutting, mixing to topsoil and preparing the seedbed preparation directly. Additionally, a rotary tiller has more mixing capacity seven times than a plough ( Ozmerzi , 2002). The rotary tiller is attached to three point linkage system of a tractor and it is driven by the tractor PTO (Power Take Off). The motion direction is changed as 90 degrees from tractor PTO to second gear box by horizontal shaft. The rotor shaft gets its motion from second gear box. Rotary tillers elements work under miscellaneous forces because of high vibration, pointless high power, impact effect of soil parts, design-manufacturing errors and wrong using conditions in tillage operation. Therefore, undesired stress distributions occur on its elements. If the elements cannot compensate to the operating forces, these elements become useless because of breaking or high deformation failure. Especially blades and transmission elements have to be durable against to operating forces. Predicting to stress distributions is so important for the designers and manufacturers to generate good working designs and products without failure. Machine manufacturers, which want to prevent for probable errors of their own machines, use materials, which have high safety coefficient, or they use high weight machine elements. Although these prevention methods can be safety, weight and cost of products rise. Helping with developed technologies and design software which integrated in new generation computers, designs are getting easier and reliable. Designers can design own products in virtual screen and they can evaluate working condition of the products by simulating techniques using the computers. Today three-dimensional (3D) modeling and finite elements method applications are getting so widespread in the industry. Many of 3D modeling and finite elements application samples can be seen on different engineering disciplines (Gunay, 1993). In this study, transmission gear train of a rotary tiller, which was designed and manufactured by a local manufacturer, was modeled using Solid works 3D parametric design software. After 3D modeling procedure, a simulation study was carried out on the transmission gear train using Cosmos works finite elements software. Rotary tiller and its second gear box transmission gear train and its 3D model were given in Figure 1. Additionally, Figure 2 shows a schema that is belong to transmission system of rotary tiller (Akinci et al., 2005). As shown in the schema that motion and power transmit with universal joint from tractor PTO output to first gear box that has 2 helical bevel gears which have 10 and 23 number of teeth and then goes to second gear box to rotor shaft. 2. Materials and Methods 2.1 3D Modeling and Stress Analysis of Transmission Gears Transmission gears were modeled according to original dimensions of gears then they were assembled. It can be seen in Figure 3 their 3D model and its values were given in Table 1. Getting started stress analysis, we assumed that gears are working in normal working condition. In the tillage operation with rotary tiller, required tractor PTO power was taken as 49.5 kW and tractor PTO revolution was 540 min According to tractor PTO power and transmission ratios, moments of gears have been accounted. Table 1. Values of Transmission Gears Values of Transmission Gears GEAR I GEARII GEARIII Module mm Number of teeth - Face width - Axel diameter mm Moments Nm 6 6 6 31 43 38 38 38 38 55 82 55 373.00 263.56 292.41 In simulation, two analyses generated for each two gear pairs (Gear I-II and Gear II-III) on working condition. Analyses have been generated in3D, static and linear assumptions in Cosmos works finite elements software. Isotropic material properties were used in simulation and properties of gears material was given at Table 2 (Kutay, 2003). While assembling, it was noted that working gears tooth in contact, paired just at single contact condition with each others. Because, experiments show that maximum stresses and failures on gears occur on gears surface contact zone and tooth root on single contact condition (Curgul, 1993). Table 2. Material Properties of Gears kMaterial DIN C45 Elastic modulus G Pa Tensile strength M Pa Yield strength M Pa Poissons ratio - Density kg/m3 211 700 500 0.30 7850 2.2 Stress Analysis Between on Gear I and Gear II After assembling of Gear I and Gear II, boundary condition was applied. Gear II fixed from bearing of its shaft. Accounted moment value was applied at direction of rotation axis to Gear I and its mesh construction can be seen in Figure 4. Cosmos works meshing functions have been used to map the meshing. Higher-order (Second-order) parabolic solid tetrahedral element which has four corner nodes, six mid-side nodes, and six edges attached by meshing function for high quality mesh construction (Cosmos Works, 2006). After meshing operation, 342160 total elements and 489339 total nodes obtained for meshed Gear I and Gear II in total. After solve process, stress distributions has been shown in Figure 5 for pairs of Gear I and Gear II. As a result maximum equivalent stress (Von Mises) determined on the contact surface of working teeth of Gear I as 123.59 M Pa and 73.98 M Pa maximum equivalent stresses determined on working teeth of Gear II. 2.3 Stress Analysis Between on Gear II and Gear III In this section, same necessary procedures are applied for stress analysis of Gear II and Gear III. Boundary conditions are applied, generated meshing and solve procedure. Gear III has been fixed on bearing and accounted moment value is applied to Gear II. After meshing operation models have 326600 total elements and 468512 total nodes for meshed Gear II and Gear III in total (Figure 6). Result plots were showed for pairs of Gear II and Gear III in Figure 7. Analysis results show that maximum equivalent stress occurred on contact surface working teeth of Gear III as 47.13 M Pa. According to applied moment 46.37 M Pa equivalent stress value occurred on contact zone of working teeth of Gear II. Obtained simulation results show us to how is distributing stresses on working teeth of transmission gears. According to simulation results and yield stress of gears material, working safety coefficient accounted for transmission gears (Table 3). Table 3. Working Safety Coefficient for Transmission Gears TRANSMISSION GEARS YIELD STRESS yield MPa VON MISES von MPa SAFETY COEFF. K coeff. = yield / von GEAR I 500 123.59 4.05 GEAR II 500 73.98 6.76 GEAR III 500 47.13 10.60 3. Conclusions In this study, stress distributions were simulated on transmission gears of a rotary tiller which designed and manufactured by local manufacturer. For this aim, transmission gears were modeled and structural stress analysis was generated using Solid works 3D parametric software and Cosmos works finite elements software. According to simulation results, following notes can be said； 1. When transmission gears were evaluated in the simulation results according to yield stress of gears material, no failure was detected on gears. Gears are working on normal condition. 2. In stress analysis between Gear I and Gear II, maximum equivalent stress was determined on contact surface of working teeth of Gear I as 123.59 M Pa. In same results plot of Gear II working teeth has 73.98 M Pa stress value on contact surface. 3. In stress analysis between Gear II and Gear III, maximum equivalent stre