新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu)設(shè)計

新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu)設(shè)計,新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu)設(shè)計,新型,翻轉(zhuǎn),犁掛接,機構(gòu),設(shè)計 設(shè)計 新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu) 設(shè)計說明書 學生姓名 學 號 所屬學院 機械電氣化工程學院 專 業(yè) 農(nóng)業(yè)機械化及其自動化 班 級 指導教師 日 期 前 言 隨著農(nóng)業(yè)生產(chǎn)的不斷發(fā)展，翻轉(zhuǎn)式雙向犁的應(yīng)用日益廣泛。80年代初，我國曾引進多種形式 的翻轉(zhuǎn)犁，這種犁要求： 采用三點掛接方式，翻轉(zhuǎn)犁的升降是通過上掛接點和拖拉機連接為主的液壓裝置，只適用于 小型翻轉(zhuǎn)犁的升降。現(xiàn)采用的三點掛接方式，翻轉(zhuǎn)犁升降下端兩個對稱的掛接點，以實現(xiàn)翻轉(zhuǎn)犁 的升降，掛接可靠，可提升較大重量的翻轉(zhuǎn)犁。 犁架相對于懸掛架能作180翻轉(zhuǎn)，以實現(xiàn)左、右犁體的工位變換，進行梭式耕作。翻轉(zhuǎn)機構(gòu) 作為翻轉(zhuǎn)犁的主要工作部件，直接影響整機工作性能。但在其國產(chǎn)化進程中，存在結(jié)構(gòu)復雜，斷 裂，磨損，翻轉(zhuǎn)不可靠，工作位置不穩(wěn)定等問題。為此，我用現(xiàn)在所學的農(nóng)業(yè)機械學知識和實物 對比、觀察，對其翻轉(zhuǎn)機構(gòu)進行設(shè)計。 新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu)設(shè)計 目 錄 1 緒論 .2 1.1 本課題來源及研究的目的和意義 .2 1.2 國內(nèi)外翻轉(zhuǎn)犁發(fā)展狀況 .2 1.2.1 國內(nèi)現(xiàn)狀 .2 1.2.2 行業(yè)發(fā)展尚存三大問題 .2 1.2.3 國外現(xiàn)狀 .3 1.3 方案的確定 .4 1.3.1 工作示意圖 .4 1.3.2 工作過程 .4 2 設(shè)計方案的選擇 .4 2.1 翻轉(zhuǎn)犁的掛接與牽引 .4 2.1.1 正確牽引 .4 2.1.2 掛接調(diào)整 .5 2.2 翻轉(zhuǎn)機構(gòu)的工作原理 .5 2.3 翻轉(zhuǎn)犁作業(yè) .5 2.3.1 結(jié)構(gòu)分析 .6 3 翻轉(zhuǎn)機構(gòu)的設(shè)計 .7 3.1 主要性能指標及其技術(shù)參數(shù) .7 3.2 工作阻力與配套動力的選擇 .7 3.2.1 翻轉(zhuǎn)犁的工作阻力 .7 3.2.2 翻轉(zhuǎn)犁正壓力 .8 3.2.3 液壓缸的選擇 .8 3.3 犁頭的設(shè)計 .9 3.3.1 犁頭的組成 .10 3.3.2 犁頭各個螺栓剪切力 .11 3.3.3 軸的設(shè)計 .11 3.4 銷軸的校核 .12 4 結(jié)論 .13 致 謝 .14 參考文獻 .15 摘 要 我國是一個農(nóng)業(yè)生產(chǎn)大國?，F(xiàn)階段，我國的農(nóng)業(yè)生產(chǎn)是從農(nóng)業(yè)機械化生產(chǎn)替代人力和小型農(nóng) 機具，以此來提高農(nóng)產(chǎn)品總值，并有效的改善從事年農(nóng)業(yè)生產(chǎn)者的生活條件。現(xiàn)在，農(nóng)機市場處 于一種比較活躍的狀態(tài)，各種農(nóng)機具的研發(fā)和生產(chǎn)，推動了我國農(nóng)業(yè)機械化的發(fā)展。 翻轉(zhuǎn)犁的投入和生產(chǎn)是農(nóng)機市場經(jīng)濟的重要部分。翻轉(zhuǎn)犁主要用于耕地、整地，大多數(shù)農(nóng)作 物在播種或者栽培前都要對土壤進行改善，為其提供一個良好的生長環(huán)境。 翻轉(zhuǎn)犁主要是由犁頭架、犁梁、犁體、合墑器等組成，它的掛接方式與其它農(nóng)機具是基本相 同的，但是翻轉(zhuǎn)犁在工作時不僅需要動力部分牽引懸掛，而且需要犁頭架對整個犁組翻轉(zhuǎn)。犁頭 架作為雙向翻轉(zhuǎn)犁的重要部分，既起到掛接牽引作用，又起到翻轉(zhuǎn)作用。本次設(shè)計將圍繞犁頭進 行闡述和設(shè)計。 關(guān)鍵字：掛接 牽引 翻轉(zhuǎn)機構(gòu) 犁頭架 1 緒論 1.1 本課題來源及研究的目的和意義 21 世紀，我國全面進入了全面建設(shè)小康社會 ，加快推進社會主義現(xiàn)代化的發(fā)展階段，我國 農(nóng)業(yè)和農(nóng)村經(jīng)濟也進入了新的歷史發(fā)展時期。農(nóng)業(yè)現(xiàn)代化進程加快、農(nóng)業(yè)和農(nóng)村經(jīng)濟戰(zhàn)略統(tǒng)一化， 必然要求農(nóng)業(yè)機械向更廣的領(lǐng)域，更高的水平上發(fā)展。 農(nóng)機化發(fā)展面臨新機遇，農(nóng)業(yè)現(xiàn)代化對農(nóng)業(yè)機械化形成了巨大需求。農(nóng)業(yè)機械是農(nóng)業(yè)生產(chǎn)力 的重要組成部分，而農(nóng)業(yè)整地機械不僅是農(nóng)業(yè)機械的先頭組成，更是農(nóng)業(yè)生產(chǎn)的開始，因此，在 活躍的農(nóng)業(yè)現(xiàn)代化發(fā)展上有著不可忽視的地位。 現(xiàn)在的農(nóng)業(yè)整地機械種類繁多，各種各樣層次不齊，半懸掛式、全懸掛式逐步替代了以往的 牽引犁。隨著農(nóng)業(yè)機械的發(fā)展，整地機械向著高速、高效、操縱簡單、保質(zhì)、保證安全等多方面 進行。翻轉(zhuǎn)犁是現(xiàn)代整地機械研發(fā)的結(jié)晶，但仍具有很多的發(fā)展空間，現(xiàn)就對翻轉(zhuǎn)犁的掛接與翻 轉(zhuǎn)機構(gòu)進行設(shè)計。 翻轉(zhuǎn)犁的主要特點： （1）結(jié)構(gòu)簡單、操作方便、作業(yè)效率高。 （2）翻轉(zhuǎn)犁的犁鏵，上下對稱，犁刀起到翻土、碎土作用。限深輪限定耕作深度。采用單液 壓缸帶動翻轉(zhuǎn)犁翻轉(zhuǎn)，單液壓缸是雙向 180翻轉(zhuǎn)來進行耕地作業(yè)，操作步驟有：中立-前推-后 推-中立。作業(yè)時翻轉(zhuǎn)犁從農(nóng)田的一方行走，到地頭后，拖拉機掉頭，翻轉(zhuǎn)犁翻轉(zhuǎn)即可順著翻土的 一側(cè)繼續(xù)工作。 1.2 國內(nèi)外翻轉(zhuǎn)犁發(fā)展狀況 1.2.1 國內(nèi)現(xiàn)狀 近兩年，我國在一系列惠農(nóng)政策的驅(qū)動下，作為我國農(nóng)機工業(yè)一部分的耕整機械行業(yè)得到了 快速健康發(fā)展，整個行業(yè)所呈現(xiàn)出的良好發(fā)展勢態(tài)是多年來未曾有過的。據(jù)中國農(nóng)業(yè)機械工業(yè)協(xié) 會耕整機械分會統(tǒng)計，2005年不僅分會會員企業(yè)產(chǎn)銷兩旺，經(jīng)濟效益狀況明顯改善，全行業(yè)也由 虧損轉(zhuǎn)為低水平盈利，工業(yè)產(chǎn)值與銷售分別比2004年增長了30和50以上。進入到2006年，大 中型拖拉機配套農(nóng)具市場在2005年的基礎(chǔ)上繼續(xù)平穩(wěn)發(fā)展，預(yù)計今年全行業(yè)的產(chǎn)值和銷售額仍將 有較大幅度的增長，翻轉(zhuǎn)犁的產(chǎn)銷形勢進一步趨旺，顯示出了全國農(nóng)機具市場巨大的發(fā)展?jié)摿Α?由于我國局部農(nóng)村生產(chǎn)力水平還很低, 經(jīng)濟還不富裕, 所以在今后四至五年內(nèi), 液壓翻轉(zhuǎn)犁 的市場需求還是比較樂觀的。 1.2.2 行業(yè)發(fā)展尚存三大問題 盡管當前耕整機械行業(yè)發(fā)展的勢態(tài)較好，但尚存在著一些制約因素及問題。耕整機械行業(yè)缺 少有一定經(jīng)濟實力、有知名品牌產(chǎn)品、具有較強研發(fā)能力的龍頭企業(yè)。國內(nèi)現(xiàn)有的大多數(shù)農(nóng)機具 企業(yè)經(jīng)濟實力弱、生產(chǎn)規(guī)模小，年銷售額在l億元以上的企業(yè)屈指可數(shù)。不少企業(yè)生產(chǎn)工藝和設(shè)備 落后，技術(shù)力量貧乏，發(fā)展后勁不足，難以生產(chǎn)技術(shù)含量相對較高的產(chǎn)品。尤其是一些小型企業(yè) 產(chǎn)品的質(zhì)量問題很多。 缺少自主知識產(chǎn)權(quán)的品牌產(chǎn)品，特別是技術(shù)含量相對較高和與大型拖拉機配套的農(nóng)機具產(chǎn)品。 目前大多數(shù)企業(yè)仍缺乏自主研發(fā)能力，農(nóng)機科研院所在轉(zhuǎn)型后研發(fā)能力也有所減弱，特別是對基 礎(chǔ)部件的研究不夠，新產(chǎn)品的研發(fā)也以模仿為主。 中小型產(chǎn)品同質(zhì)化、市場無序低價競爭等問題嚴重。在一些地區(qū)同類產(chǎn)品生產(chǎn)廠家高度集中， 產(chǎn)品在功能和結(jié)構(gòu)上幾乎沒有區(qū)別，如河南商丘市附近的機引犁廠，在全國農(nóng)機展上就可看到每 個地區(qū)都有十幾家企業(yè)在展示同一類的產(chǎn)品。特別是個別企業(yè)以代用材料和簡化制造工藝來降低 生產(chǎn)成本，在市場進行無序低價競爭，其對行業(yè)發(fā)展的負面影響不可低估。因此，要使耕整機械 行業(yè)保持穩(wěn)定、健康發(fā)展，行業(yè)中有實力的企業(yè)亟待強化品牌意識，增強自主創(chuàng)新能力，并在產(chǎn) 品的多樣化、增加附加價值等方面下功夫。而小企業(yè)則應(yīng)通過保證產(chǎn)品質(zhì)量，以發(fā)展自己的特色 產(chǎn)品求生存。 1.2.3 國外現(xiàn)狀 國外的農(nóng)業(yè)機械化已經(jīng)達到了一個很高的程度。 從農(nóng)業(yè)作業(yè)程度上來說，國外的翻轉(zhuǎn)犁農(nóng)業(yè)生產(chǎn)效率很高，各種大功率的拖拉機，帶動不同 犁鏵數(shù)量的翻轉(zhuǎn)犁，6 鏵、7 鏵的翻轉(zhuǎn)犁在國外很普及，但是在我國，用的還是 4 鏵犁、5 鏵犁， 有些地區(qū)還是引進的高幅寬作業(yè)的翻轉(zhuǎn)犁。 現(xiàn)在的國外的翻轉(zhuǎn)犁產(chǎn)品在同一系列上規(guī)格多樣，適用于不同的大功率拖拉機。 系列化是農(nóng)業(yè)機械發(fā)展的重要趨勢。國外著名大公司逐步實現(xiàn)其產(chǎn)品系列化進程，形成了大型 不同規(guī)格的產(chǎn)品。與此同時，產(chǎn)品更新?lián)Q代的周期明顯縮短。利用 GPS（全球定位系統(tǒng)） 、GIS 和 GSM 技術(shù)，實現(xiàn)了精準農(nóng)業(yè)，如無線電數(shù)據(jù)通信、機器監(jiān)測、診斷、工作與業(yè)務(wù)管理軟件和機器 控制等裝置。精準農(nóng)業(yè)由以下 3 部分組成： （1）計算機輔助農(nóng)田作業(yè)系統(tǒng)。其包括機載計算、GPS 微波定位和高速無線電通信 3 項技術(shù)。 在運行中，機載系統(tǒng)通過無線電接收整個無線網(wǎng)絡(luò)中的農(nóng)業(yè)耕作數(shù)據(jù)、土壤數(shù)據(jù)或現(xiàn)場規(guī)劃數(shù)據(jù)。 這些數(shù)據(jù)都顯示在駕駛室內(nèi)的一個屏幕上，司機在駕駛室內(nèi)能夠直觀地了解機器的作業(yè)位置，并 準確地判斷需要的精準耕地位置及其耕深。 （2）關(guān)鍵信息管理系統(tǒng)，監(jiān)測機器中極其關(guān)鍵的性能與作業(yè)參數(shù)，并且通過無線電將數(shù)據(jù)從 該機器傳送到業(yè)主辦公室。業(yè)主可立即分析數(shù)據(jù)以便估量機器的當前狀態(tài)，或加以收集和整理， 以便顯示機器的作業(yè)趨勢。 （3）把即時數(shù)據(jù)監(jiān)控機器的數(shù)據(jù)及時傳送給業(yè)主，產(chǎn)生一個集成的現(xiàn)時作業(yè)模型，使業(yè)主能 在接近實時條件下對現(xiàn)場或遠處監(jiān)控各種作業(yè)。GPS 導航系統(tǒng)與電子地圖、無線電通信網(wǎng)絡(luò)及計 算機車輛管理信息系統(tǒng)相結(jié)合，可以實現(xiàn)車輛跟蹤和交通管理等許多功能。這些功能包括：利用 GPS 和電子地圖可以實時顯示出車輛的實際位置，并任意放大、縮小、還原、換圖；可以隨目標 移動，使目標始終保持在屏幕上。 1.3 方案的確定 1.3.1 工作示意圖 圖 1-1 工作示意圖 1-牽引架 2-調(diào)節(jié)螺桿 3-牽引調(diào)節(jié)座 4-升降油缸 5-犁架 6-翻轉(zhuǎn)液壓缸 7-犁體 8-陷深輪 1.3.2 工作過程： 翻轉(zhuǎn)犁在工作前由拖拉機和行走輪將翻轉(zhuǎn)犁懸空，帶入農(nóng)田，進入農(nóng)田后，可卸下行走輪。 工作時拖拉機通過后懸掛連接牽引架使翻轉(zhuǎn)犁處于懸空狀態(tài)，牽引架可通過調(diào)節(jié)螺桿改變與 犁架的相對角度，拖拉機也可根據(jù)情況調(diào)整后懸掛牽引點的高度，使翻轉(zhuǎn)犁在任何情況下都能保 持最 佳的入土角度。三點掛接是拖拉機牽引、懸掛的主要方式。 翻轉(zhuǎn)犁由液壓缸提供翻轉(zhuǎn)動力，手動拉桿調(diào)幅裝置，限深輪控制前后耕深，根據(jù)農(nóng)田性質(zhì)， 選擇合適的幅寬。耕地時，犁鏵將土壤翻向一側(cè)，行進到地頭后，拖拉機提升翻轉(zhuǎn)犁，同時，翻 轉(zhuǎn)犁翻轉(zhuǎn) 180，拖拉機掉頭，順著剛犁過的地方繼續(xù)工作，土壤翻垡的方向一致。 農(nóng)田耕地完成后，翻轉(zhuǎn)液壓缸收縮，使兩排翻轉(zhuǎn)犁同時懸空，犁頭后擋板有一個銷孔，固定 翻轉(zhuǎn)液壓缸，安裝行走輪。即可拖帶翻轉(zhuǎn)犁離開農(nóng)田。 2 設(shè)計方案的選擇 2.1 翻轉(zhuǎn)犁的掛接與牽引 全懸掛翻轉(zhuǎn)犁由懸掛裝置(犁頭、翻轉(zhuǎn)軸、油缸、轉(zhuǎn)向閥)、橫向水平調(diào)節(jié)螺栓、犁架(梁)、 犁體(犁柱、犁壁、犁脛、犁鏟、犁側(cè)板)、限深輪、調(diào)幅斜撐桿等組成。 2.1.1 正確牽引 懸掛翻轉(zhuǎn)犁正確牽引標志是拖拉機在正常耕地作業(yè)時，中央拉桿在拖拉機上應(yīng)掛接固定在高 位，與犁頭掛接應(yīng)在長孔中，處于中間位置，從而可前后擺動。左右提升拉桿的長度一定要調(diào)節(jié) 一致、適宜、以確保犁地時能達到最大耕深，并保證犁提升翻轉(zhuǎn)時不碰地面。左右下拉桿處于對 稱位置并在上提升桿的圓孔中，下拉桿的左右限位桿銷放在長孔位置上，要讓懸掛架有 35cm 的 活動量。使犁架縱梁與拖拉機前進方向一致，保持平行。前犁鏟翼偏過拖拉機右輪內(nèi)側(cè) 1025lnm，即單體犁耕寬的重疊量。 輪式拖拉機配全懸掛翻轉(zhuǎn)犁時，拖拉機的輪距必須與犁的總耕幅相適應(yīng)，才能保證正確牽引， 也有利于在采用內(nèi)翻法耕地時耕到邊。拖拉機輪距的調(diào)整：輪式拖拉機輪距=犁總耕幅+12 單體 犁寬+輪胎寬度(cm)；前輪比后輪輪距寬 10-15cm(拖拉機行走在溝里)。后輪輪距相當于 3 個犁垡 的寬度。 2.1.2 掛接調(diào)整 （1）犁的前后水平調(diào)整 調(diào)整方法：伸長或縮短中央拉桿，直至犁架前后水平。 （2）犁架左右水平調(diào)整 通過水平調(diào)節(jié)限位螺栓，使犁工作時左右水平(從后面看犁柱與地面垂直，也可向已翻地方向 稍有傾斜，翻土效果會更好)。左右限位螺栓要調(diào)節(jié)一致。 （3）正位調(diào)整 在耕地過程中若出現(xiàn)犁架偏斜、前體犁漏耕、接垡不平、左右下拉桿向未耕地方向偏斜、拖 拉機操向困難等，應(yīng)調(diào)整懸掛左右拉鏈或限位桿銷，使后體犁向未耕地方向偏擺，后犁犁側(cè)板壓 向溝壁，這樣犁在耕地過程中借助溝壁的反作用力而擺正。若犁架向相反方向偏擺，前體犁重耕， 接垡不平，左右下拉桿向已耕地方向偏擺，調(diào)整方法與上述方法相反。犁 擺正后，看前體犁的寬 度是否合適，若不合適，則在犁上左右移動調(diào)節(jié)懸掛軸。前犁耕寬過大， 應(yīng)向未耕地方向移動懸掛左右拉鏈或限位桿銷；反之則向已耕地方向移動。經(jīng)過上述調(diào)整后， 若前體犁的耕寬仍不合適，則只有通過調(diào)整拖拉機的輪距才能克服。 (1)耕深調(diào)整 懸掛犁的耕深調(diào)整隨著拖拉機液壓系統(tǒng)的不同而不同。分置式液壓系統(tǒng)通過改變犁的限深輪 高低位置來調(diào)節(jié)耕深；半分置式和整體式液壓系統(tǒng)一般都采用力調(diào)節(jié)和位調(diào)節(jié)方法，通過改變犁 的限深輪高低位置以及由拖拉機上液壓提升操縱系統(tǒng)來調(diào)節(jié)耕深。 (2)入土性能調(diào)整 懸掛連接點高度可以調(diào)節(jié)的懸掛犁在干硬土地不易人土時，可以通過降低下拉桿在拖拉機上 的連接高度或提高下拉桿在犁上的連接高度來調(diào)節(jié)，也可以提高上拉桿在拖拉機上的連接高度或 降低上拉桿在犁上的連接高度來調(diào)節(jié)。耕濕、軟地時犁容易下鉆，調(diào)節(jié)方法與上相反。 (3)偏牽引的調(diào)整 犁的總耕幅減少，與拖拉機輪距不配套時就產(chǎn)生偏牽引。調(diào)整方法是移動懸掛左右下拉鏈或 限位桿銷，調(diào)好犁與拖拉機的正確相對位置，使后犁向未耕地方向偏過一些，讓犁在作業(yè)時犁架 稍有偏斜，以增大犁側(cè)板對溝壁反力的平衡能力。犁架偏斜程度視土壤和耕深等具體情況而定， 以保證耕地作業(yè)質(zhì)量又能直線行駛為標準。 2.2 翻轉(zhuǎn)機構(gòu)的工作原理 翻轉(zhuǎn)機構(gòu)是一個搖桿滑塊機構(gòu)，它的翻轉(zhuǎn)角度范圍在 0180之間，依靠液壓缸的伸長和 收 縮來實現(xiàn)翻轉(zhuǎn)的。翻轉(zhuǎn)原理入如下所示： 圖 2-1 翻轉(zhuǎn)機構(gòu)工作原理圖 設(shè)計要求： 液壓翻轉(zhuǎn)犁的翻轉(zhuǎn)機構(gòu)主要在于犁頭部分，液壓缸給轉(zhuǎn)動軸提供一個動力，使連接轉(zhuǎn)動軸的 犁架整體在 180內(nèi)翻轉(zhuǎn)。 此次設(shè)計的重點部分是犁頭，根據(jù)翻轉(zhuǎn)犁總質(zhì)量選擇合適的液壓缸，并對翻轉(zhuǎn)軸軸向拉力、 抗拉強度、抗彎強度等進行校核。 2.3 翻轉(zhuǎn)犁作業(yè) 液壓翻轉(zhuǎn)犁是近幾年推廣應(yīng)用的一種新型農(nóng)業(yè)機械。用翻轉(zhuǎn)犁進行耕翻作業(yè)具有無開閉壟， 生產(chǎn)率高，節(jié)能等優(yōu)點。翻轉(zhuǎn)機構(gòu)是翻轉(zhuǎn)犁的重要組成部分，其性能好壞直接影響翻轉(zhuǎn)犁的性能 和可靠性。 2.3.1 結(jié)構(gòu)分析 （1） “A ”形犁頭 “A ”形犁頭上部為 1 點連接，掛在拖拉機的后懸掛（或固定）牽引銷孔，下部為 2 點連接， 與犁頭形成平面的剛性聯(lián)接。拖拉機通過后懸掛連接牽引桿拖拽翻轉(zhuǎn)犁，拖拉機可根據(jù)情況調(diào)整 后懸掛牽引桿的高度，使犁體在任何情況下都能保持最佳的入土角度。由于地面不平， A ”形 犁頭上部牽引點銷軸處應(yīng)可順牽引方向的縱向軸轉(zhuǎn)動。 支撐調(diào)節(jié)桿連接“A ”形牽引架和犁頭部分，正反螺桿，可調(diào)節(jié)長度。翻轉(zhuǎn)犁作業(yè)時，應(yīng)根 據(jù)土壤性質(zhì)和切削阻力大小適時調(diào)整鏟犁體切入角。可通過調(diào)節(jié)支撐調(diào)節(jié)桿的螺桿改變與犁體的 相對角度，以獲得最佳的耕地效果。 （2）液壓油缸 目前，國內(nèi)外市場上銷售的液壓翻轉(zhuǎn)犁，都是采用單油缸控制翻轉(zhuǎn)來完成翻轉(zhuǎn)操作的。而現(xiàn) 階 段研究的新型的雙油缸控制翻轉(zhuǎn) 180 裝置，其功能上要比單液壓缸的穩(wěn)定、可靠。但是雙液壓 缸不僅結(jié)構(gòu)復雜、成本也比較高，而且還增加了額外的負載。本次設(shè)計采用立式油缸翻轉(zhuǎn)機構(gòu)。 液壓缸是一種配置靈活、設(shè)計制造比較容易而應(yīng)用廣泛的液壓執(zhí)行元件。液壓缸有系列化標 準的產(chǎn)品和專業(yè)系列產(chǎn)品。液壓油缸是由缸體、柱塞、缸蓋等組成。立式翻轉(zhuǎn)液壓缸的選擇依據(jù)： 1)了解和掌握液壓缸在機器上的用途和工作要求。 2)了解液壓缸工作環(huán)境條件。 3)了解外部負載情況。 4)了解液壓缸運動形態(tài)及安裝約來條件。 5)了解液壓系統(tǒng)的情況。 （3）犁架 犁架是犁的重要支撐部件和傳力部件，其強度和剛度對犁的工作性能具有重要影響。尤其對 翻轉(zhuǎn)犁來說，如果犁架的剛度差，會造成左右犁體的耕深不一致，影響作業(yè)質(zhì)量。液壓翻轉(zhuǎn)犁的 犁架采用矩形管和板材組成的剛架。從力學原理來看，這種剛架屬于超靜定結(jié)構(gòu)，用平衡方程確 定各構(gòu)件的內(nèi)力及變形困難較大。 液壓翻轉(zhuǎn)犁犁架，主梁采用斷面為 140mm140mm12mm 的矩形鋼管，與牽引梁組成銷連接式 結(jié)構(gòu)。 （4）犁體 犁體是鏵式犁的主要工作部件，一般由犁鏵、犁壁、犁側(cè)板、犁柱以及犁托等組成。犁鏵、 犁壁、犁托等部件組成一個整體，通過犁柱安裝在犁架上。犁體的功用是切土、破碎和翻轉(zhuǎn)土壤， 達到覆蓋雜草、殘茬和疏松土壤的目的。 犁鏵主要起入土、切土作用。常用的有鑿形、梯形和三角形犁鏵。液壓翻轉(zhuǎn)犁采用三角形犁 鏵。 犁壁與犁鏵一起構(gòu)成犁體曲面，將犁鏵移來的土壤加以破碎和翻轉(zhuǎn)。 犁側(cè)板位于犁鏵后上方，耕地時緊貼溝壁，承受并平衡耕作時產(chǎn)生的側(cè)向力和部分垂直壓力。 （5）限深輪 一種翻轉(zhuǎn)犁的限深裝置。本實用新型包括限深輪、限深輪臂，其特征在于限深輪臂一端設(shè)有 與犁架副梁鉸接的鉸接軸，限深輪臂上鉸接有限位限深調(diào)節(jié)裝置，所說的限位限深調(diào)節(jié)裝置包括 調(diào)節(jié)絲桿、調(diào)節(jié)絲套、限位擋塊、限位掛鉤、鎖緊螺帽和固定套，限位掛鉤鉸接在固定套上，固 定套上設(shè)有與犁架付梁鉸接的鉸接軸；本實用新型與已有技術(shù)相比，具有能雙向限位、耕深一致、 工作可靠、操作簡單、方便的優(yōu)點。 設(shè)計要求： （1）為了保證耕地質(zhì)量，翻轉(zhuǎn)犁需要選擇與之相適應(yīng)的大馬力拖拉機。 （2）為保證翻轉(zhuǎn)犁犁體不受到運輸過程中的磨損，拖拉機的懸掛裝置在升降翻轉(zhuǎn)犁時要滿足 升降拉力。選擇翻轉(zhuǎn)犁的材料應(yīng)安全、可靠。翻轉(zhuǎn)結(jié)構(gòu)應(yīng)盡量簡單化。 （3）為了滿足耕地作業(yè)率，保證耕地過程中不出現(xiàn)犁架斷裂或者翻轉(zhuǎn)軸彎曲等故障。 3 翻轉(zhuǎn)機構(gòu)的設(shè)計 3.1 主要性能指標及其技術(shù)參數(shù) 連接方式:懸掛翻轉(zhuǎn)式 配套機具: 120140 匹馬力的拖拉機帶動 1LCHT- 5 型翻轉(zhuǎn)犁 工作速度：79km/ h 耕深：350400mm 耕深穩(wěn)定性變異系數(shù)： %10 犁體間距：825mm 單寬幅：250380mm 總寬幅：12131928mm 生產(chǎn)率：1518（hm2/h） 犁架高度：750mm 整機質(zhì)量：1200kg 配套拖拉機額定功率：50kW 外形尺寸：479217751710 3.2 工作阻力與配套動力的選擇 3.2.1 翻轉(zhuǎn)犁的工作阻力 工作阻力主要來自以下兩個方面： 翻垡土壤阻力 1T =KF （3-1T 1） 式中：K單位切削力，對于輕質(zhì)壤土，取 NF4105 F切下的土塊的斷面積, 其中：b 為平均工作深度取 h=0.35m；aF k土垡翻轉(zhuǎn)比，k=1.27 k=b/a,所以 a=3.5/1.27=2.76 所以： =KF =5 0.350.2765=24150N （1TK 3-2） 3.2.2 翻轉(zhuǎn)犁正壓力 2 =G 式中：2T1f 土壤之間的摩擦系數(shù)，對于砂土， =0.530.75, 取 =0.75；G= 1f1f1mf 所以 15009.80.75=11025N 12gGfT （3-3） 由以上可知，翻轉(zhuǎn)犁的總阻力 A 為： =24150N=11025N=35175N1TA2 考慮沖擊、慣性等動載荷因素，計算工作阻力 P 為： ADPK 式中： 動載系數(shù)， =1.3；D 所以 （3-4NTAP 5.472351. ） 選擇配套動力時，其牽引力應(yīng)大于 本機選用的配套動力為東方紅 1604，可滿足要求，其技術(shù)參數(shù)如下： 外形尺寸：539026963450 前輪輪距（可調(diào)）：1580-2176mm 后輪輪距（可調(diào)）：1700-2395mm 最小離地間隙：495mm 最小使用質(zhì)量：7350kg 速度范圍：前進 2.35-30.69km/h 后退 4.44-11.69km/h 發(fā)動機標定功率：132kW 最大提升力，懸掛點后 610mm 處：36kN 最大配重質(zhì)量前/后：810/450kg 最大牽引力：58.5kN 3.2.3 液壓缸的選擇 液壓工作原理圖 圖 3-1 液壓翻轉(zhuǎn)犁的工作原理圖 按照動力相似的原則，考慮配套拖拉機液壓系統(tǒng)和活塞移動速度的要求，在國家標準范 圍內(nèi)，選擇合適的油缸型號。 動力相似應(yīng)滿足以下條件： （1） 要求犁架的運動規(guī)律相似，從而保證機構(gòu)的力傳遞相似。 （2） 要求機構(gòu)對液流量的適應(yīng)范圍寬，以降低使用要求，便于操作。 油缸由拖拉機的液壓系統(tǒng)控制。犁處于工作狀態(tài)時，油缸處于最大的伸長狀態(tài)。翻轉(zhuǎn)換向時， 操縱分配器使油缸縮短，帶動犁梁向上翻轉(zhuǎn)。當犁梁旋轉(zhuǎn)到接近垂直位置時再操縱分配器使其 處于浮動狀態(tài)此時油缸處于卸荷狀態(tài)失去了對犁梁的控制。犁梁在慣性作用下越過“死點” 位置然后在慣性及重力的作用下繼續(xù)旋轉(zhuǎn)。直到另一側(cè)工作位置時停止。 翻轉(zhuǎn)速度控制分析： 1)在翻轉(zhuǎn)五鏵犁液壓回路中增設(shè)緩沖閥以增加回油阻力 。緩沖閥體、閥芯、管接頭和密F 封圈組成，其中閥芯可以在閥體內(nèi)自由滑動。緩沖閥結(jié)構(gòu)示意圖見圖 6。閥芯右端有 1 個錐面、1 個軸向小孔和 4 個徑向通油孔。當犁梁向上翻轉(zhuǎn)時，壓力油從緩沖閥的右端流向左端閥芯在壓 力作用下滑向左端液壓油可通過 1 個軸向孔和 4 個徑向孔流動。當犁向下翻轉(zhuǎn)時液壓油從左 端向右端流動，閥芯在液壓油的作用下滑向右端，其右端的錐面封閉了 4 個徑向孔油液只能從 1 個孔中流動。因為軸向小孔的直徑很小對液壓油的流動有節(jié)流作用所以限制了犁粱向下旋 轉(zhuǎn)的速度，減小了換向時的沖擊現(xiàn)象。翻轉(zhuǎn)五鏵犁的緩沖閥節(jié)流孔直徑為 3mm。此時犁梁向下旋 轉(zhuǎn)的角速度可限制在 0305rs 范圍內(nèi)。 圖 3-2 緩沖閥結(jié)構(gòu)示意圖 2）拖托機駕駛員在實際作業(yè)時，可通過增加切換油路時的提前角 來減小犁梁在“死點”位 置時的角速度 ，最終減小翻轉(zhuǎn)終了時的動能，減小換向沖擊。但必須保證 0。因為 。 （3-F510.2TP.427 11） 銷的剪切力達到條件，可以選取，該銷安全。 上掛接點銷的剪切力 。 （3-t8.2 P5. 12） 銷的剪切力達到條件，可以選取，該銷安全。 軸選用 40Cr。 40Cr 的許用應(yīng)力 ，軸的彎曲應(yīng)力 ， 。Mpa401 pa.301Mpa8.256 工作狀態(tài)下 。 5.31 懸掛狀態(tài)下 。826 軸的彎曲應(yīng)力在許用范圍內(nèi)，可以選取,該軸安全。 4 結(jié)論 在設(shè)計的過程中，我收集了一些文獻資料，再以實踐為主，積累了豐富的第一手材料，在翻 轉(zhuǎn)犁翻轉(zhuǎn)機構(gòu)設(shè)計、配套設(shè)備選擇、材料選擇等具體設(shè)計任務(wù)中進行了比較、計算。有效的培養(yǎng) 了自己分析問題、解決問題的能力，并使專業(yè)知識得到鞏固和升華。在設(shè)計工程中，因為時間緊、 任務(wù)重，特別是CAD制圖難度比較大，經(jīng)常是通宵達旦的計算、繪圖，十分辛苦。這使我深深感受 到了奮戰(zhàn)在我國農(nóng)業(yè)機械設(shè)計第一線的專家、工程師和技術(shù)人員的辛勞，對他們?yōu)槲覈r(nóng)業(yè)事業(yè) 所付出的汗水所做出的貢獻表示深深的敬意。所以為什么資料那么難找，有權(quán)威性的資料是要錢 的，沒有權(quán)威性的資料有些胡謅也就理所當然了。 在以后的學習和工作中，我將繼續(xù)發(fā)揚這種能吃苦的精神，為我國農(nóng)機事業(yè)發(fā)展做出應(yīng)有的 貢獻。 在本次設(shè)計中仍有不足與疏漏。在設(shè)計過程中，雖然有老師的耐心講解，但對于一些具體問 題，比如關(guān)鍵部位的校核，液壓缸工作能否達到靈敏度等問題，仍感覺吃不透，我將在以后的工 作、學習中揚長避短，發(fā)揚嚴謹?shù)目茖W態(tài)度，使所到的知識不斷升華。 致 謝 經(jīng)過近三個月的努力，在老師的指導下、同學們的幫助下，新型翻轉(zhuǎn)犁掛接與翻轉(zhuǎn)機構(gòu)的設(shè) 計終于設(shè)計完成了，在此對老師和同學們給予我的幫助表示衷心的感謝。 在畢業(yè)設(shè)計過程中，感謝安靜老師的督促和幫助，也感謝在本次設(shè)計中馬少輝老師的指導， 馬老師在百忙之中對我的設(shè)計給予了細致的指導和建議,對我的輔導耐心認真，百問不煩，給我清 晰的設(shè)計思路，使我的這次設(shè)計能順利完成。他那嚴謹求實的教學作風、誨人不倦的耐心，給我 留下了難以磨滅的印象。同時感謝班級同學為我提供的幫助，我還要感謝塔里木大學帶過我們的 所有老師，你們對待知識嚴謹求實的態(tài)度、為人師表的工作作風，使我受益匪淺。 在此，我對你們表示最衷心的感謝，我將在今后的工作中不斷追求新知識、繼續(xù)努力，不辜 負老師們對我們悉心的培養(yǎng)。 即將面臨畢業(yè)，我真心的祝愿我的老師們身體健康，工作順利，祝愿我的母校越辦越好。 參考文獻 1 楊俊江.懸掛翻轉(zhuǎn)犁的正確掛接與調(diào)整J.農(nóng)機化,2008( 3). 2 孫殿軍.M-S950型翻轉(zhuǎn)五鏵犁的正確使用與調(diào)整J.農(nóng)機化,2001(1). 3 趙永滿, 梅衛(wèi)江.垂直換向懸掛翻轉(zhuǎn)五鏵犁的虛擬設(shè)計J.整地機械,2001(3). 4 鄭德聰,王玉順,吳海平等.雙向犁翻轉(zhuǎn)機構(gòu)反求設(shè)計J.農(nóng)業(yè)工程學報.1999(1). 5 顧海英,周小偉.現(xiàn)代都市農(nóng)業(yè)可持續(xù)發(fā)展的意義及內(nèi)涵J.農(nóng)業(yè)現(xiàn)代化研究,2002(1). 6 徐福玲，陳曉明.液壓與氣壓傳動.北京：機械工業(yè)出版社.2007,(5):21-22. 7 王忠，鄭明新.機械工程材料.北京：清華大學出版社.2005,(10).45-47. 8 袁佳平.農(nóng)業(yè)機械的設(shè)計和計算.北京：中國農(nóng)業(yè)出版社.2012(1)12-15. 9 東北工學院.機械零件設(shè)計手冊.北京：冶金工業(yè)出版社.1980.(3)155-157. 10 西北工業(yè)大學.機械設(shè)計(第七版).北京：高等教育出版社.78-79. 11 單輝祖.材料力學.(第二版).北京：高等教育出版社.2004（8）65-67. Korea Soil fax: +82 42 823 6246. E-mail address: sochungcnu.ac.kr (S.-O. Chung). 0167-1987/$ see front matter C223 2013 Elsevier B.V. All rights reserved. http:/dx.doi.org/10.1016/j.still.2013.07.013 Effects of gear selection of an agricultural PTO load during rotary tillage Yong-Joo Kim a , Sun-Ok Chung b, *, Chang-Hyun Choi a Machinery Technology Group, Advanced R Van et al., 2009) for efficient and optimum design of a tractor (Han et al., 1999). Most studies on the load analysis have focused on the transmission since it makes up about 30% of the total tractor costs (e.g., Kim, 1998). For analysis of the transmission load, researchers analyzed torque load acting on the transmission input shaft and the driving axle shafts of the tractor during field operations such as plow tillage (Kim et al., 2001; Nahmgung, 2001). The load on the transmission input shaft and the driving axle shafts increased with plowing speed in most of the field conditions. Some research considered load on PTO shafts during rotary tillage and baling operations. Kim et al. (2011b) analyzed power consumption of a tractor with a rated engine power of 75 kW during baler operation and reported that ratios of the power consumption to the engine power were 5075% for all PTO gear levels. Also, Kim et al. (2011a) analyzed power requirement of major components (driving axles, PTO shaft, and hydraulic pumps) of a 30 kW agricultural tractor during plow tillage, rotary tillage, and loader operations. Rotary tillage required the greatest power, and the PTO shaft experienced the greatest amount of the power among the components during the process. Summarizing the findings above, considerable amount of the load was applied on PTO shaft during rotary tillage. However, research related to the effects of transmission (i.e., operation speed) and PTO gear selection on the tractor load during field operations has not been reported. This study was an attempt to provide guidelines for the optimum gear setting, considering both field efficiency and load severeness of the major power transmission parts. The purpose of this study was to analyze effects of gear selection on loads acting on the transmission and PTO shafts of a 75 kW agricultural tractor during rotary tillage. 2. Materials and methods 2.1. Measurement system A 75-kW agricultural tractor (L7040, LS Mtron Ltd., Korea) was used in this study. The tractor had a total mass of 3260 kg and dimensions of 4077 mm C2 2000 mm C2 2640 mm (length C2 width C2 height). The rated engine power and the PTO power of the tractor at an engine revolution speed of 2300 rpm were 75 kW and 65 kW, respectively. The tractor was equipped with a Synchro-mesh type manual transmission composed of two direction-gears, four main-gears, and four sub-gears. The 16 forward and 16 backward ground speeds of the tractor were determined by combination of the gear settings. The PTO rotational speeds of the tractor at P1, P2, and P3 settings were 540 rpm, 750 rpm, and 1000 rpm, respectively. Fig. 1 shows the torque transducers and radio telemetry systems fitted on the transmission and PTO input shafts for load measurement. The transmission and PTO input shafts were connected directly to the engine crankshaft; therefore, the speed ratio between the engine crankshaft and the shaft was 1:1. The load measurement system was installed inside of the clutch housing. The load measurement system was constructed with strain-gauge sensors (CEA-06-250US-350, MicroMeasurement Co., USA) to measure torque, radio telemetry I/O interfaces (R2, Manner, Germany) to acquire the sensor signals, and an embedded system to analyze the load. For load measure- ment of the transmission, a strain-gauge with a rotor antenna was installed on the transmission input shaft, and a stator antenna was installed on the shaft case. For PTO load measurement, a strain- gauge was installed on the flywheel-sleeve, and a rotor antenna and a stator antenna were installed on the flywheel and engine case, respectively. The embedded system had the maximum resolution of 24 bits. Load signal from the strain gauges of the calibrated torque transducers was digitized with a sampling rate of 19.2 kHz at a 24-bit resolution and stored in the embedded system (MGC, HMB, Germany). A program to measure the load signal was developed based on Labview software (version 2009, National Instrument, USA). 2.2. Experimental methods The load acting on the tractor during field operation depends on many factors such as soil condition and driver skill. Because taking all of these factors into consideration was not practical (Nahm- gung, 2001), effects of these factors were minimized in the study to focus on effects of ground speed and PTO rotational speed on the load through gear selection. Rotary tillage was conducted at three ground speeds and three PTO rotational speeds in upland field sites located at 35859 0 23 00 and 35859 0 26 00 North and 127812 0 56 00 and 127813 0 3 00 East. The soil type was sandy, the average water content was 22.3%, and the average cone index was 1236 kPa, over the depths of 0250 mm. The tillage depth was set as 20 cm. The transmission gear was set to at L1, L2, and L3 gears to match with PTO gears of P1, P2, and P3, respectively. The gear settings were selected based on the results of a survey for the annual usage ratio of tractor reported by Kim et al. (2011a). The ground speeds of the tractor at L1, L2, and L3 were 1.87 km h C01 , 2.64 km h C01 , and 3.77 km h C01 and the PTO rotational speeds at P1, P2, and P3 were 540 rpm, 750 rpm, and 1000 rpm, respectively. The rotary tillage tool was a heavy duty rotavator (WJ220E, WOONGJIN, Korea), and the required rated power, total mass, tillage width, and dimensions were 75 kW, 750 kg, 2220 mm, and 1050 mm C2 2390 mm C2 1380 mm (length C2 width C2 height), respectively. 2.3. Load analysis Procedures to analyze the tractor load would be different, depending on the purpose. Many researchers have used simple statistics such as average, maximum, and minimum values in order to represent the load. This method extracts representative values to show the difference of amplitude, but this simplification prohibits characterizing the whole load profiles since the field load is irregular. Effects of gear setting on the transmission and PTO load, One-Way ANOVA and the least significant difference test (LSD) were conducted with SAS (version 9.1, SAS Institute, Cary, USA). Also, it is difficult to represent effects of the load on the tractor since the load causes damages to the tractor, and fatigue of the tractor parts also needs to be investigated. Fatigue of a tractor is defined as the damage sum from repeated loadings (Lampman, 1997). Severeness, another method of load representation proposed by Kim et al. (1998, 2000), is defined as the ratio of the damage sum at each operation to the minimum damage sum from all the operations. Severeness would be inversely proportional to fatigue life. When load severeness is greater, fatigue life would be shorter. Kim et al. (1998) measured the loads acting on the transmission input shaft and analyzed the load severeness during plow tillage, rotary tillage, and transportation operations. They found that the severeness during transportation operation was similar to that during plow tillage, but the severeness during rotary tillage was about 63 times greater than that during the transportation operation. Later, Kim et al. (2000) analyzed the severeness of the transmission input shaft during rotary tillage at four speed combinations of the tractor ground speeds (2.9 km h C01 and 4.1 km h C01 ) and PTO rotational speeds (588 and 704 rpm) using a tractor with a rated engine power of 30 kW. The load severeness increased by 2.32.6 times when the PTO speed increased with the Y.-J. Kim et al. / Soil Nguyen et al., 2011). The SN curve was obtained for the material of the input shaft, SCM 420H, using the ASTM standard (2004) in Eq. (3). The ASTM standard has been widely used for fatigue analysis of materials (Wannenburg et al., 2009; Mao, 2010). N 10 6C06:097logS=223 (3) where N is the number of cycles, S is shear stress (MPa). To calculate damage sum, equivalent torque of load spectrum was converted to stress (Rahama and Chancellor, procedures of severeness evaluation. input Y.-J. Kim et al. / Soil Petracconi et al., 2010). The diameters of the transmission and the PTO input shafts were 28 mm and 26.5 mm, respectively. S 16T pd 3 (4) where S is stress (MPa), T is equivalent torque (Nm), and d is diameter of the shaft (mm). The damage sum was calculated based on the Miners rule (Miner, 1945) in Eq. (5). Miners rule is a procedure for estimating the number of cycles of loading to failure (Miner, 1945; Robson, 1964; Renius, 1977). The number of cycles (n) was derived from an equivalent torque of the load spectrum. The fatigue life cycles (N) was derived from the SN of SCM 420H. The damage (D) was calculated by dividing the number of the fatigue life cycles into the number of cycles. D t X k i1 n i N i (5) where D t is damage sum, n i is number of cycles, and N i is fatigue life (cycles). 3. Results and discussion 3.1. Transmission and PTO loads by gear selection Fig. 3 shows examples of torque loads on the transmission and PTO input shafts for the ground speed at L1 and the PTO rotational speed at P2 during the rotary tillage operation. The rotary tillage operation consisted of a preparing period to descend the 3-point Fig. 3. Example of torque loads on the transmission and PTO hitch, an operating period to till the soil, and a completion period to ascend the 3-point hitch. The measured torque on the transmission and PTO input shafts showed steep increasing in the preparing period and decreasing in the completion period, and torque on these components showed irregular fluctuation patterns in the Table 1 Average torque (Nm) on the transmission and PTO input shafts by gear setting during Transmission input shaft P1 P2 P3 L1 29.8 C6 2.8 Aa 1,2,3 35.9 C6 4.2 Ba 38.7 C6 4.7 Ba L2 43.5 C6 4.7 Ab 51.3 C6 3.3 Bb 58.5 C6 3.1 Cb L3 64.9 C6 3.2 Ac 71.5 C6 3.5 Bc 82.1 C6 4.9 Cc 1 Average C6 standard deviation. 2 Means with different superscript (A, B, C, D) in each row are significantly different 3 Means with different superscript (a, b, c, d) in each column are significantly different 4 Values in the parentheses are the ratios of the torque on the PTO input shaft to the operating period. The magnitude and range of the measured torque on the PTO input shaft were greater than those on the transmission input shaft in the operating period. Table 1 shows torque levels on the transmission and PTO input shafts by speed combination of the ground speeds (L1, L2, L3) and the PTO rotational speeds (P1, P2, P3). The average torque was calculated only for the operating period data, not for the preparing and completion periods. The averaged torque levels of the PTO input shaft were greater than those of the transmission input shaft at all gear levels during the rotary tillage. These results were similar to the results by Kim et al. (2011a) that the PTO required the greatest amount of power among the major components during rotary tillage. The average torque on the transmission input shaft was considerably and significantly increased as the ground speed increased from L1 to L3 at the same PTO rotational speeds. Load increases on the transmission and driving shafts with the speed increase of plow tillage were also found by Kim et al. (2011a,b) and Nahmgung (2001). Also, the averaged load on the transmission input shaft increased as the PTO rotational speed increased, except that the load values were not significantly different between L1P2 and L1P3. The average torque on the PTO input shaft increased as the ground speed and PTO rotational speed increased. The increases were statistically significant for the PTO rotational speed, but not significant for the ground speed. 3.2. Severeness evaluation Figs. 4 and 5 show load spectra of the transmission and the PTO input shafts by gear setting during the rotary tillage, respectively. shafts torque at the L1P2 selection during rotary tillage. The spectra were constructed considering the entire life of the tractor, and the numbers of cycles were in the ranges from 10 3 to 10 7 . The range of the maximum torque ratio of the transmission input shaft was 0.71.5 for the speed-combinations and the greatest torque ratio was found at L3P1 as shown in Fig. 4. In rotary tillage. PTO input shaft P1 P2 P3 105.0 C6 3.6 Aa (3.52) 4 148.7 C6 3.7 Ba (4.14) 175.8 C6 4.2 Ca (4.54) 114.7 C6 3.9 Ab (2.64) 151.7 C6 4.1 Ba (2.95) 177.7 C6 3.4 Ca (3.03) 118.7 C6 4.0 Ab (1.83) 151.9 C6 3.8 Ba (2.12) 182.9 C6 4.2 Cb (2.23) at p 0.05 by LSDs multiple range tests at p 0.05 by LSDs multiple range tests. torque on the transmission input shaft at the speed combinations. Y.-J. Kim et al. / Soil a survey of the state of the art for homogenous materials. International Journal of Fatigue 20 (1) 934. Gerlach, A., 1966. Field measurement of tractor transmission forces. Transactions of the ASAE 9 (5) 707712. Glinka, G., Kam, J.C.P., 1987. Rainflow counting algorithm for very long stress histories. International Journal of Fatigue 9 (4) 223228. Graham, J.A., Berns, D.K., Olberts, D.R., 1962. Cumulative damage used to analyze tractor final drives. Transaction of the ASAE 5 (2) 139146. Han, K.H., Kim, K.U., Wu, Y.G., 1999. Severeness of transmission loads of agricultural tractor for rotary operations in poorly drained paddy field. Journal of Biosys- tems Engineering 24 (4) 293300. Hong, N., 1991. A modified rainflow counting method. International Journal of Fatigue 13 (6) 465469. KAMICO, KSAM, 2010. Agricultural Machinery Yearbook in Republic of Korea. Korea Agricultural Machinery Industry Cooperative and Korean Society for Agricul- tural Machinery, Seoul. Kichler, C.M., Fulton, J.P., Raper, R.L., McDonald, T.P., Zech, W.C., 2011. Effects of transmission gear selection on tractor performance and fuel costs during deep tillage operations. Soil and Tillage Research 113 (2) 105111. Kim, D.C., Ryu, I.H., Kim, K.U., 2001. Analysis of tractor transmission and driving axle loads. Transactions of the ASABE 44 (4) 751757. Kim, J.H., 1998. Analysis of Mission and Transmission Loads of Agricultural Tractors. Seoul National University, Seoul, Republic of Korea (Unpublished MS Thesis). Kim, J.H., Kim, K.U., Choi, C.W., Wu, Y.G., 1998. Severeness of transmission loads of agricultural tractors. Journal of Biosystems Engineering 23 (5) 417426. Kim, J.H., Kim, K.U., Wu, Y.G., 2000. Analysis of transmission load of agricultural tractors. Journal of Terramechanics 37, 113125. Kim, Y.J., Chung, S.O., Park, S.J., Choi, C.H., 2011a. Analysis of power requirement of agricultural tractor by major field operation. Journal of Biosystems Engineering 36 (2) 7988. Kim, Y.J., Lee, D.H., Chung, S.O., Park, S.J., Choi, C.H., 2011b. Analysis of power requirement of agricultural tractor during baler operation. Journal of Biosys- tems Engineering 36 (4) 243251. Lampman, S.R., 1997. ASM Handbook 19: Fatigue and Fracture. ASM International, Ohio. Lee, D.H., 2011. Analysis of Power Requirements of Tractor for Field Operations. Sungkyunkwan University, Suwon, Republic of Korea (MS Thesis). Mao, W., 2010. Fatigue Assessment and Extreme Response Prediction of Ship Structures. Chalmers University of Technology and University of Gothenburg, Sweden (PhD Dissertation). Miner, M.A., 1945. Cumulative damage in fatigue. Journal of Applied Mechanics 12 (3) 159164. Nahmgung, M.J., 2001. Load Analysis of Driving Axles and Life Evaluation of Driving Gear of PTO on Tractors. Sungkyunkwan University, Suwon, Republic of Korea (PhD Dissertation). Nguyen, N.T., Chu, Q.T., Kim, S.E., 2011. Fatigue analysis of a pre-fabricated ortho- tropic steel deck for light-weight vehicles. Journal of Constructional Steel Research 67 (4) 647655. Park, S.H., Kim, Y.J., Im, D.H., Kim, C.K., Jang, Y., Kim, S.S., 2010a. Analysis of factors affecting fuel consumption of agricultural tractor. Journal of Biosystems Engi- neering 35 (3) 151157. Park, S.H., Kim, Y.J., Im, D.H., Kim, C.K., Jung, S.C., Kim, H.J., Jang, Y., Kim, S.S., 2010b. Development of eco driving system for agricultural tractor. Journal of Biosys- tems Engineering 35 (2) 7784. Park, S.H., Kim, Y.J., Im, D.H., Kim, C.K., Jung, S.C., Kim, H.J., Lee, J.S., Kim, S.S., 2010c. Characteristics of tractor PTO power and workloads. Journal of Biosystems Engineering 35 (1) 1520. Petracconi, C.L., Ferreira, S.E., Palma, E.S., 2010. Fatigue life simulation of a rear tow hook assembly of a passenger car. Engineering Failure Analysis 17 (2) 455463. Rahama, O.A., Chancellor, W.J., 1994. Peak and average loads on tractor structures. Transactions of the ASAE 37 (6) 17331740. Renius, K.T., 1977. Application of cumulative damage theory to agricultural tractor design elements. Konstrucktion 29 (3) 8593. Robson, J.D., 1964. An Introduction to Random Vibration. Edinburgh University Press, Edinburgh. Rotz, A., Bowers, W., 1991. Repair and Maintenance Cost Data for Agricultural Equipment. In: ASAE Report No. 911531. American Society for Agricultural Engineers, St. Joseph, MI. Simens, J.C., Bowers, W., 1999. Machinery Management: How to Select Machinery to Fit the Real Needs of Farm Managers, Farm Business Management (FMB) Series. John Deere Publishing, East Moline, IL. Van, N.N., Matsuo, T., Koumoto, T., Inaba, S., 2009. Transducers for measuring dynamic axle load of farm tractor. Bulletin of the Faculty of Agriculture, Saga University 94, 2335. Wannenburg, J., Heyns, P.S., Raath, A.D., 2009. Application of fatigue equivalent static load methodology for the numerical durability assessment of heavy vehicle structures. International Journal of Fatigue 31 (10) 15411549. Xiong, J.J., Shenoi, R.A., 2005. A reliability-based data treatment system for actual load history. Fatigue and Fracture of Engineering Materials and Structures 28 (10) 875889. Y.-J. Kim et al. / Soil & Tillage Research 134 (2013) 909696