中型普通車床主軸箱設(shè)計【Dmax=350mm Nmin=212r-minN=5.5KW φ=1.26 Z=8】
中型普通車床主軸箱設(shè)計【Dmax=350mm Nmin=212r-minN=5.5KW =1.26 Z=8】,Dmax=350mm Nmin=212r-min N=5.5KW =1.26 Z=8,中型普通車床主軸箱設(shè)計【Dmax=350mm,Nmin=212r-min,N=5.5KW,=1.26,Z=8】,中型
寧XX大學課程設(shè)計(論文)中型普通車床主軸箱設(shè)計(題目14)所在學院專 業(yè)班 級姓 名學 號指導老師 年 月 日5摘 要本設(shè)計著重研究機床主傳動系統(tǒng)的設(shè)計步驟和設(shè)計方法,根據(jù)已確定的運動參數(shù)以變速箱展開圖的總中心距最小為目標,擬定變速系統(tǒng)的變速方案,以獲得最優(yōu)方案以及較高的設(shè)計效率。在機床主傳動系統(tǒng)中,為減少齒輪數(shù)目,簡化結(jié)構(gòu),縮短軸向尺寸,用齒輪齒數(shù)的設(shè)計方法是試算,湊算法,計算麻煩且不易找出合理的設(shè)計方案。本文通過對主傳動系統(tǒng)中三聯(lián)滑移齒輪傳動特點的分析與研究,繪制零件工作圖與主軸箱展開圖及剖視圖。關(guān)鍵詞:傳動系統(tǒng)設(shè)計,傳動副,結(jié)構(gòu)網(wǎng),結(jié)構(gòu)式,目 錄摘 要2目 錄4第1章 緒論61.1 課程設(shè)計的目的61.2課程設(shè)計的內(nèi)容61.2.1 理論分析與設(shè)計計算61.2.2 圖樣技術(shù)設(shè)計61.2.3編制技術(shù)文件61.3 課程設(shè)計題目、主要技術(shù)參數(shù)和技術(shù)要求6第2章 車床參數(shù)的擬定82.1車床主參數(shù)和基本參數(shù)82.2擬定參數(shù)的步驟和方法82.2.1 極限切削速度Vmax、Vmin82.2.2 主軸的極限最低轉(zhuǎn)速82.2.3 主電機功率動力參數(shù)的確定92.2.4確定結(jié)構(gòu)式92.2.5確定結(jié)構(gòu)網(wǎng)92.2.6繪制轉(zhuǎn)速圖和傳動系統(tǒng)圖102.3 確定各變速組此論傳動副齒數(shù)102.4 核算主軸轉(zhuǎn)速誤差11第3章 傳動件的計算123.1 帶傳動設(shè)計123.2選擇帶型133.3確定帶輪的基準直徑并驗證帶速133.4確定中心距離、帶的基準長度并驗算小輪包角143.5確定帶的根數(shù)z153.6確定帶輪的結(jié)構(gòu)和尺寸153.7確定帶的張緊裝置153.8計算壓軸力163.2 計算轉(zhuǎn)速的計算173.3 齒輪模數(shù)計算及驗算183.4 傳動軸最小軸徑的初定213.5 主軸合理跨距的計算22第4章 主要零部件的選擇234.1 軸承的選擇234.2 鍵的規(guī)格234.3 主軸彎曲剛度校核244.4.軸承校核244.5 潤滑與密封24第5章 摩擦離合器(多片式)的計算25第6章 主要零部件的選擇276.1電動機的選擇276.2 軸承的選擇276.3變速操縱機構(gòu)的選擇276.4 軸的校核276.5 軸承壽命校核29第7章 主軸箱結(jié)構(gòu)設(shè)計及說明307.1 結(jié)構(gòu)設(shè)計的內(nèi)容、技術(shù)要求和方案307.2 展開圖及其布置31結(jié)束語32參考文獻33 第1章 緒論1.1 課程設(shè)計的目的課程設(shè)計是在學完本課程后,進行一次學習設(shè)計的綜合性練習。通過課程設(shè)計,使學生能夠運用所學過的基礎(chǔ)課、技術(shù)基礎(chǔ)課和專業(yè)課的有關(guān)理論知識,及生產(chǎn)實習等實踐技能,達到鞏固、加深和拓展所學知識的目的。通過課程設(shè)計,分析比較機械系統(tǒng)中的某些典型機構(gòu),進行選擇和改進;結(jié)合結(jié)構(gòu)設(shè)計,進行設(shè)計計算并編寫技術(shù)文件;完成系統(tǒng)主傳動設(shè)計,達到學習設(shè)計步驟和方法的目的。通過設(shè)計,掌握查閱相關(guān)工程設(shè)計手冊、設(shè)計標準和資料的方法,達到積累設(shè)計知識和設(shè)計技巧,提高學生設(shè)計能力的目的。通過設(shè)計,使學生獲得機械系統(tǒng)基本設(shè)計技能的訓練,提高分析和解決工程技術(shù)問題的能力,并為進行機械系統(tǒng)設(shè)計創(chuàng)造一定的條件。1.2課程設(shè)計的內(nèi)容機械系統(tǒng)設(shè)計課程設(shè)計內(nèi)容由理論分析與設(shè)計計算、圖樣技術(shù)設(shè)計和技術(shù)文件編制三部分組成。1.2.1 理論分析與設(shè)計計算(1)機械系統(tǒng)的方案設(shè)計。設(shè)計方案的分析,最佳功能原理方案的確定。(2)根據(jù)總體設(shè)計參數(shù),進行傳動系統(tǒng)運動設(shè)計和計算。(3)根據(jù)設(shè)計方案和零部件選擇情況,進行有關(guān)動力計算和校核。1.2.2 圖樣技術(shù)設(shè)計(1)選擇系統(tǒng)中的主要機件。(2)工程技術(shù)圖樣的設(shè)計與繪制。1.2.3編制技術(shù)文件(1)對于課程設(shè)計內(nèi)容進行自我經(jīng)濟技術(shù)評價。(2)編制設(shè)計計算說明書。1.3 課程設(shè)計題目、主要技術(shù)參數(shù)和技術(shù)要求題目:中型普通車床主軸箱設(shè)計題目14 車床的主參數(shù)(規(guī)格尺寸)和基本參數(shù)如下:工件最大回轉(zhuǎn)直徑D(mm)正轉(zhuǎn)最低轉(zhuǎn)速nmin( )電機功率N(kw)公比轉(zhuǎn)速級數(shù)Z3502125.51.26833第2章 車床參數(shù)的擬定2.1車床主參數(shù)和基本參數(shù)車床的主參數(shù)(規(guī)格尺寸)和基本參數(shù)如下:工件最大回轉(zhuǎn)直徑D(mm)正轉(zhuǎn)最低轉(zhuǎn)速nmin( )電機功率N(kw)公比轉(zhuǎn)速級數(shù)Z3502125.51.2682.2擬定參數(shù)的步驟和方法2.2.1 極限切削速度Vmax、Vmin根據(jù)典型的和可能的工藝選取極限切削速度要考慮:允許的切速極限參考值如下:表 1.1加 工 條 件 Vmax(m/min)Vmin(m/min)硬質(zhì)合金刀具粗加工鑄鐵工件 3050硬質(zhì)合金刀具半精或精加工碳鋼工件150300螺紋加工和鉸孔382.2.2 主軸的極限最低轉(zhuǎn)速計算車床主軸極限轉(zhuǎn)速時的加工直徑,則主軸極限轉(zhuǎn)速應(yīng)為:結(jié)合題目條件,取標準數(shù)列數(shù)值,即=212r/min取依據(jù)題目要求選級數(shù)Z=8, =1.26=1.064考慮到設(shè)計的結(jié)構(gòu)復(fù)雜程度要適中,故采用常規(guī)的擴大傳動。各級轉(zhuǎn)速數(shù)列可直接從標準的數(shù)列表中查出,按標準轉(zhuǎn)速數(shù)列為:212,265,335,425,530,670,850,10602.2.3 主電機功率動力參數(shù)的確定合理地確定電機功率N,使機床既能充分發(fā)揮其性能,滿足生產(chǎn)需要,又不致使電機經(jīng)常輕載而降低功率因素。根據(jù)題設(shè)條件電機功率為5.5KW可選取電機為:Y132S-4額定功率為5.5KW,滿載轉(zhuǎn)速為1440r/min.2.2.4確定結(jié)構(gòu)式已知Z=x3ba、b為正整數(shù),即Z應(yīng)可以分解為2和3的因子,以便用2、3聯(lián)滑移齒輪實現(xiàn)變速。取Z=8級 則Z=22對于Z=8可分解為:Z=212224。綜合上述可得:主傳動部件的運動參數(shù) =212 Z=8 =1.262.2.5確定結(jié)構(gòu)網(wǎng)根據(jù)“前多后少” , “先降后升” , 前密后疏,結(jié)構(gòu)緊湊的原則,選取傳動方案 Z=212224,易知第二擴大組的變速范圍r=(P3-1)x=1.264=3.958 滿足要求,其結(jié)構(gòu)網(wǎng)如圖2-1。 圖2-1結(jié)構(gòu)網(wǎng) Z=2122242.2.6繪制轉(zhuǎn)速圖和傳動系統(tǒng)圖(1)選擇電動機:采用Y系列封閉自扇冷式鼠籠型三相異步電動機。(2)繪制轉(zhuǎn)速圖:(3)畫主傳動系統(tǒng)圖。根據(jù)系統(tǒng)轉(zhuǎn)速圖及已知的技術(shù)參數(shù),畫主傳動系統(tǒng)圖如圖2-3:1-2軸最小中心距:A1_2min1/2(Zmaxm+2m+D)軸最小齒數(shù)和:Szmin(Zmax+2+D/m)2.3 確定各變速組此論傳動副齒數(shù)(1)Sz100-120,中型機床Sz=70-100(2)直齒圓柱齒輪Zmin18-20,m4 圖2-3 主傳動系統(tǒng)圖(7)齒輪齒數(shù)的確定。變速組內(nèi)取模數(shù)相等,據(jù)設(shè)計要求Zmin1820,齒數(shù)和Sz100120,由表4.1,根據(jù)各變速組公比,可得各傳動比和齒輪齒數(shù),各齒輪齒數(shù)如表2-2。 表2-2 齒輪齒數(shù)傳動比基本組第一擴大組第二擴大組1:11:1.581:11:1.261.26:11:2代號ZZZZZZZZZ5Z5ZZ齒數(shù)4747 36 58 4242 3747 49 3929592.4 核算主軸轉(zhuǎn)速誤差實際傳動比所造成的主軸轉(zhuǎn)速誤差,一般不應(yīng)超過10(-1),即10(-1)=2.6各級轉(zhuǎn)速誤差n 1060850670530425335265212n1064.2837.8672.1524.5426.05329.6266.2208.6誤差0.41.40.41.40.41.40.41.4轉(zhuǎn)速誤差小于2.6,因此不需要修改齒數(shù)。第3章 傳動件的計算3.1 帶傳動設(shè)計輸出功率P=5.5kW,轉(zhuǎn)速n1=1440r/min,n2=850r/min3.1.1計算設(shè)計功率Pd表4 工作情況系數(shù)工作機原動機類類一天工作時間/h10161016載荷平穩(wěn)液體攪拌機;離心式水泵;通風機和鼓風機();離心式壓縮機;輕型運輸機1.01.11.21.11.21.3載荷變動小帶式運輸機(運送砂石、谷物),通風機();發(fā)電機;旋轉(zhuǎn)式水泵;金屬切削機床;剪床;壓力機;印刷機;振動篩1.11.21.31.21.31.4載荷變動較大螺旋式運輸機;斗式上料機;往復(fù)式水泵和壓縮機;鍛錘;磨粉機;鋸木機和木工機械;紡織機械1.21.31.41.41.51.6載荷變動很大破碎機(旋轉(zhuǎn)式、顎式等);球磨機;棒磨機;起重機;挖掘機;橡膠輥壓機1.31.41.51.51.61.8根據(jù)V帶的載荷平穩(wěn),兩班工作制(16小時),查機械設(shè)計P296表4,取KA1.1。即3.2選擇帶型普通V帶的帶型根據(jù)傳動的設(shè)計功率Pd和小帶輪的轉(zhuǎn)速n1按機械設(shè)計P297圖1311選取。根據(jù)算出的Pd6.05kW及小帶輪轉(zhuǎn)速n11440r/min ,查圖得:dd=80100可知應(yīng)選取A型V帶。3.3確定帶輪的基準直徑并驗證帶速由機械設(shè)計P298表137查得,小帶輪基準直徑為80100mm則取dd1=100mm ddmin.=75 mm(dd1根據(jù)P295表13-4查得)表3 V帶帶輪最小基準直徑槽型YZABCDE205075125200355500由機械設(shè)計P295表13-4查“V帶輪的基準直徑”,得=170mm 誤差驗算傳動比: (為彈性滑動率)誤差 符合要求 帶速 滿足5m/sv300mm,所以宜選用E型輪輻式帶輪??傊?,小帶輪選H型孔板式結(jié)構(gòu),大帶輪選擇E型輪輻式結(jié)構(gòu)。帶輪的材料:選用灰鑄鐵,HT200。3.7確定帶的張緊裝置 選用結(jié)構(gòu)簡單,調(diào)整方便的定期調(diào)整中心距的張緊裝置。3.8計算壓軸力 由機械設(shè)計P303表1312查得,A型帶的初拉力F0130.59N,上面已得到=153.36o,z=6,則對帶輪的主要要求是質(zhì)量小且分布均勻、工藝性好、與帶接觸的工作表面加工精度要高,以減少帶的磨損。轉(zhuǎn)速高時要進行動平衡,對于鑄造和焊接帶輪的內(nèi)應(yīng)力要小, 帶輪由輪緣、腹板(輪輻)和輪轂三部分組成。帶輪的外圈環(huán)形部分稱為輪緣,輪緣是帶輪的工作部分,用以安裝傳動帶,制有梯形輪槽。由于普通V帶兩側(cè)面間的夾角是40,為了適應(yīng)V帶在帶輪上彎曲時截面變形而使楔角減小,故規(guī)定普通V帶輪槽角 為32、34、36、38(按帶的型號及帶輪直徑確定),輪槽尺寸見表7-3。裝在軸上的筒形部分稱為輪轂,是帶輪與軸的聯(lián)接部分。中間部分稱為輪幅(腹板),用來聯(lián)接輪緣與輪轂成一整體。表 普通V帶輪的輪槽尺寸(摘自GB/T13575.1-92) 項目 符號 槽型 Y Z A B C D E 基準寬度 b p 5.3 8.5 11.0 14.0 19.0 27.0 32.0 基準線上槽深 h amin 1.6 2.0 2.75 3.5 4.8 8.1 9.6 基準線下槽深 h fmin 4.7 7.0 8.7 10.8 14.3 19.9 23.4 槽間距 e 8 0.3 12 0.3 15 0.3 19 0.4 25.5 0.5 37 0.6 44.5 0.7 第一槽對稱面至端面的距離 f min 6 7 9 11.5 16 23 28 最小輪緣厚 5 5.5 6 7.5 10 12 15 帶輪寬 B B =( z -1) e + 2 f z 輪槽數(shù) 外徑 d a 輪 槽 角 32 對應(yīng)的基準直徑 d d 60 - - - - - - 34 - 80 118 190 315 - - 36 60 - - - - 475 600 38 - 80 118 190 315 475 600 極限偏差 1 0.5 V帶輪按腹板(輪輻)結(jié)構(gòu)的不同分為以下幾種型式: (1) 實心帶輪:用于尺寸較小的帶輪(dd(2.53)d時),如圖7 -6a。 (2) 腹板帶輪:用于中小尺寸的帶輪(dd 300mm 時),如圖7-6b。 (3) 孔板帶輪:用于尺寸較大的帶輪(ddd) 100 mm 時),如圖7 -6c 。 (4) 橢圓輪輻帶輪:用于尺寸大的帶輪(dd 500mm 時),如圖7-6d。(a) (b) (c) (d)圖7-6 帶輪結(jié)構(gòu)類型根據(jù)設(shè)計結(jié)果,可以得出結(jié)論:小帶輪選擇實心帶輪,如圖(a),大帶輪選擇腹板帶輪如圖(b)3.2 計算轉(zhuǎn)速的計算(1)主軸的計算轉(zhuǎn)速nj,由公式n=n得,主軸的計算轉(zhuǎn)速nj=311.85r/min,取335 r/min。(2). 傳動軸的計算轉(zhuǎn)速 軸2=670 r/min,軸1=850r/min。(2)確定各傳動軸的計算轉(zhuǎn)速。表3-1 各軸計算轉(zhuǎn)速軸 號 軸 軸 軸計算轉(zhuǎn)速 r/min 850850670(3) 確定齒輪副的計算轉(zhuǎn)速。3-2。 表3-2 齒輪副計算轉(zhuǎn)速序號ZZZZZn8508508508506703.3 齒輪模數(shù)計算及驗算(1)模數(shù)計算。一般同一變速組內(nèi)的齒輪取同一模數(shù),選取負荷最重的小齒輪,按簡化的接觸疲勞強度公式進行計算,即mj=16338可得各組的模數(shù),如表3-3所示。表3-3 模數(shù)組號基本組第一擴大組模數(shù) mm 2.52.5(2)基本組齒輪計算。 基本組齒輪幾何尺寸見下表齒輪Z1Z1 Z2Z2齒數(shù)47473658分度圓直徑117.5117.590145齒頂圓直徑122.5122.595150齒根圓直徑111.25111.2583.75138.75 齒寬20202020按基本組最小齒輪計算。小齒輪用40Cr,調(diào)質(zhì)處理,硬度241HB286HB,平均取260HB,大齒輪用45鋼,調(diào)質(zhì)處理,硬度229HB286HB,平均取240HB。計算如下: 齒面接觸疲勞強度計算: 接觸應(yīng)力驗算公式為 彎曲應(yīng)力驗算公式為: 式中 N-傳遞的額定功率(kW),這里取N為電動機功率,N=5kW; -計算轉(zhuǎn)速(r/min). =335(r/min); m-初算的齒輪模數(shù)(mm), m=2.5(mm); B-齒寬(mm);B=20(mm); z-小齒輪齒數(shù);z=36; u-小齒輪齒數(shù)與大齒輪齒數(shù)之比,u=1.6; -壽命系數(shù); = -工作期限系數(shù); T-齒輪工作期限,這里取T=15000h.; -齒輪的最低轉(zhuǎn)速(r/min), =500(r/min) -基準循環(huán)次數(shù),接觸載荷取=,彎曲載荷取= m-疲勞曲線指數(shù),接觸載荷取m=3;彎曲載荷取m=6; -轉(zhuǎn)速變化系數(shù),查【5】2上,取=0.60 -功率利用系數(shù),查【5】2上,取=0.78 -材料強化系數(shù),查【5】2上, =0.60 -工作狀況系數(shù),取=1.1 -動載荷系數(shù),查【5】2上,取=1 -齒向載荷分布系數(shù),查【5】2上,=1 Y-齒形系數(shù),查【5】2上,Y=0.386;-許用接觸應(yīng)力(MPa),查【4】,表4-7,取=650 Mpa;-許用彎曲應(yīng)力(MPa),查【4】,表4-7,取=275 Mpa;根據(jù)上述公式,可求得及查取值可求得:=635 Mpa =78 Mpa(3)擴大組齒輪計算。第一擴大組齒輪幾何尺寸見下表齒輪Z3Z3 Z4Z4齒數(shù)42423747分度圓直徑10510592.5117.5齒頂圓直徑11011097.5122.5齒根圓直徑98.7598.7586.25111.25 齒寬20202020第二擴大組齒輪幾何尺寸見下表 齒輪Z5Z5Z6Z6齒數(shù)49392959分度圓直徑14711787117齒頂圓直徑15312393183齒根圓直徑139.5109.579.5169.5齒寬24242424按擴大組最小齒輪計算。小齒輪用40Cr,調(diào)質(zhì)處理,硬度241HB286HB,平均取260HB,大齒輪用45鋼,調(diào)質(zhì)處理,硬度229HB286HB,平均取240HB。 同理根據(jù)基本組的計算,查文獻【6】,可得 =0.62, =0.77,=0.60,=1.1,=1,=1,m=3.5,=355;可求得:=619 Mpa =135Mpa 3.4 傳動軸最小軸徑的初定由【5】式6,傳動軸直徑按扭轉(zhuǎn)剛度用下式計算: d=1.64(mm) 或 d=91(mm)式中 d-傳動軸直徑(mm) Tn-該軸傳遞的額定扭矩(N*mm) T=9550000; N-該軸傳遞的功率(KW) -該軸的計算轉(zhuǎn)速 -該軸每米長度的允許扭轉(zhuǎn)角,=。各軸最小軸徑如表3-3。 表3-3 最小軸徑軸 號 軸 軸最小軸徑mm 3540 3.5 主軸合理跨距的計算由于電動機功率P=5.5kw,根據(jù)【1】表3.20,前軸徑應(yīng)為6090mm。初步選取d1=80mm。后軸徑的d2=(0.70.9)d1,取d2=60mm。根據(jù)設(shè)計方案,前軸承為NN3016K型,后軸承為圓錐滾子軸承。定懸伸量a=120mm,主軸孔徑為30mm。軸承剛度,主軸最大輸出轉(zhuǎn)矩T=9550=9550=424.44N.m設(shè)該機床為車床的最大加工直徑為350mm。床身上最常用的最大加工直徑,即經(jīng)濟加工直徑約為最大回轉(zhuǎn)直徑的50%,這里取60%,即180mm,故半徑為0.09m;切削力(沿y軸) Fc=4716N背向力(沿x軸) Fp=0.5 Fc=2358N總作用力 F=5272.65N此力作用于工件上,主軸端受力為F=5272.65N。先假設(shè)l/a=2,l=3a=240mm。前后支承反力RA和RB分別為RA=F=5272.65=7908.97NRB=F=5272.65=2636.325N根據(jù) 文獻【1】式3.7 得:Kr=3.39得前支承的剛度:KA= 1689.69 N/ ;KB= 785.57 N/;=2.15 主軸的當量外徑de=(80+60)/2=70mm,故慣性矩為 I=113.810-8m4 =0.14查【1】圖3-38 得 =2.0,與原假設(shè)接近,所以最佳跨距=1202.0=240mm合理跨距為(0.75-1.5),取合理跨距l(xiāng)=360mm。 根據(jù)結(jié)構(gòu)的需要,主軸的實際跨距大于合理跨距,因此需要采取措施增加主軸的剛度,增大軸徑:前軸徑D=100mm,后軸徑d=80mm。前軸承采用雙列圓柱滾子軸承,后支承采用背對背安裝的角接觸球軸承。第4章 主要零部件的選擇 4.1 軸承的選擇I軸:與帶輪靠近段安裝雙列角接觸球軸承代號7007C 另一安裝深溝球軸承6012II軸:對稱布置深溝球軸承6009III軸:后端安裝雙列角接觸球軸承代號7015C 另一安裝端角接觸球軸承代號7010C中間布置角接觸球軸承代號7012C4.2 鍵的規(guī)格 I軸安裝帶輪處選擇普通平鍵規(guī)格:BXL=10X56 II軸選擇花鍵規(guī)格:N dDB =8X36X40X7 III軸選擇鍵規(guī)格:BXL=14X90 4.3 主軸彎曲剛度校核(1)主軸剛度符合要求的條件如下:a主軸的前端部撓度b主軸在前軸承處的傾角c在安裝齒輪處的傾角(2)計算如下:前支撐為雙列圓柱滾子軸承,后支撐為角接觸軸承架立放圓柱滾子軸承跨距L=450mm.當量外徑 de=主軸剛度:因為di/de=25/285=0.0880.7,所以孔對剛度的影響可忽略;ks=2kN/mm剛度要求:主軸的剛度可根據(jù)機床的穩(wěn)定性和精度要求來評定4.4.軸承校核 4.5 潤滑與密封 主軸轉(zhuǎn)速高,必須保證充分潤滑,一般常用單獨的油管將油引到軸承處。 主軸是兩端外伸的軸,防止漏油更為重要而困難。防漏的措施有兩種: 1)密封圈加密封裝置防止油外流。 2)疏導在適當?shù)牡胤阶龀龌赜吐罚褂湍茼樌亓骰氐接拖?。?章 摩擦離合器(多片式)的計算設(shè)計多片式摩擦離合器時,首先根據(jù)機床結(jié)構(gòu)確定離合器的尺寸,如為軸裝式時,外摩擦片的內(nèi)徑d應(yīng)比花鍵軸大26mm,內(nèi)摩擦片的外徑D的確定,直接影響離合器的徑向和軸向尺寸,甚至影響主軸箱內(nèi)部結(jié)構(gòu)布局,故應(yīng)合理選擇。摩擦片對數(shù)可按下式計算 Z2MnK/fbp式中 Mn摩擦離合器所傳遞的扭矩(Nmm); Mn955/955110.98/8001.28(Nmm); Nd電動機的額定功率(kW); 安裝離合器的傳動軸的計算轉(zhuǎn)速(r/min); 從電動機到離合器軸的傳動效率; K安全系數(shù),一般取1.31.5; f摩擦片間的摩擦系數(shù),由于磨擦片為淬火鋼,查機床設(shè)計指導表2-15,取f=0.08; 摩擦片的平均直徑(mm); =(D+d)/267mm; b內(nèi)外摩擦片的接觸寬度(mm); b=(D-d)/2=23mm; 摩擦片的許用壓強(N/);1.11.001.000.760.836 基本許用壓強(MPa),查機床設(shè)計指導表2-15,取1.1; 速度修正系數(shù) n/6=2.5(m/s) 根據(jù)平均圓周速度查機床設(shè)計指導表2-16,取1.00; 接合次數(shù)修正系數(shù),查機床設(shè)計指導表2-17,取1.00; 摩擦結(jié)合面數(shù)修正系數(shù),查機床設(shè)計指導表2-18,取0.76。所以 Z2MnK/fbp21.281.4/(3.140.08230.83611 臥式車床反向離合器所傳遞的扭矩可按空載功率損耗確定,一般取0.40.4114.4 最后確定摩擦離合器的軸向壓緊力Q,可按下式計算:Q=b(N)1.13.14231.003.57式中各符號意義同前述。摩擦片的厚度一般取1、1.5、1.75、2(mm),內(nèi)外層分離時的最大間隙為0.20.4(mm),摩擦片的材料應(yīng)具有較高的耐磨性、摩擦系數(shù)大、耐高溫、抗膠合性好等特點,常用10或15鋼,表面滲碳0.30.5(mm),淬火硬度達HRC5262。第6章 主要零部件的選擇 6.1電動機的選擇轉(zhuǎn)速n1440r/min,功率P5.5kW選用Y系列三相異步電動機 6.2 軸承的選擇I軸:與帶輪靠近段安裝雙列角接觸球軸承代號7007C 另一安裝深溝球軸承6012II軸:對稱布置深溝球軸承6009III軸:后端安裝雙列角接觸球軸承代號7015C 另一安裝端角接觸球軸承代號7010C中間布置角接觸球軸承代號7012C6.3變速操縱機構(gòu)的選擇選用左右擺動的操縱桿使其通過桿的推力來控制II軸上的三聯(lián)滑移齒輪和二聯(lián)滑移齒輪。6.4 軸的校核(a) 主軸的前端部撓度(b) 主軸在前軸承處的傾角(c) 在安裝齒輪處的傾角E取為,由于小齒輪的傳動力大,這里以小齒輪來進行計算將其分解為垂直分力和水平分力由公式可得主軸載荷圖如下所示:由上圖可知如下數(shù)據(jù):a=364mm,b=161mm,l=525mm,c=87mm計算(在垂直平面),,計算(在水平面),,合成:6.5 軸承壽命校核由軸最小軸徑可取軸承為7008C角接觸球軸承,=3;P=XFr+YFaX=1,Y=0。對軸受力分析得:前支承的徑向力Fr=2642.32N。 由軸承壽命的計算公式:預(yù)期的使用壽命 L10h=15000hL10h=hL10h=15000h 軸承壽命滿足要求。第7章 主軸箱結(jié)構(gòu)設(shè)計及說明7.1 結(jié)構(gòu)設(shè)計的內(nèi)容、技術(shù)要求和方案設(shè)計主軸變速箱的結(jié)構(gòu)包括傳動件(傳動軸、軸承、帶輪、齒輪、離合器和制動器等)、主軸組件、操縱機構(gòu)、潤滑密封系統(tǒng)和箱體及其聯(lián)結(jié)件的結(jié)構(gòu)設(shè)計與布置,用一張展開圖和若干張橫截面圖表示。課程設(shè)計由于時間的限制,一0般只畫展開圖。主軸變速箱是機床的重要部件。設(shè)計時除考慮一般機械傳動的有關(guān)要求外,著重考慮以下幾個方面的問題。精度方面的要求,剛度和抗震性的要求,傳動效率要求,主軸前軸承處溫度和溫升的控制,結(jié)構(gòu)工藝性,操作方便、安全、可靠原則,遵循標準化和通用化的原則。主軸變速箱結(jié)構(gòu)設(shè)計時整個機床設(shè)計的重點,由于結(jié)構(gòu)復(fù)雜,設(shè)計中不可避免要經(jīng)過反復(fù)思考和多次修改。在正式畫圖前應(yīng)該先畫草圖。目的是:1 布置傳動件及選擇結(jié)構(gòu)方案。2 檢驗傳動設(shè)計的結(jié)果中有無干涉、碰撞或其他不合理的情況,以便及時改正。3 確定傳動軸的支承跨距、齒輪在軸上的位置以及各軸的相對位置,以確定各軸的受力點和受力方向,為軸和軸承的驗算提供必要的數(shù)據(jù)。7.2 展開圖及其布置展開圖就是按照傳動軸傳遞運動的先后順序,假想將各軸沿其軸線剖開并將這些剖切面平整展開在同一個平面上。I軸上裝的摩擦離合器和變速齒輪。有兩種布置方案,一是將兩級變速齒輪和離合器做成一體。齒輪的直徑受到離合器內(nèi)徑的約束,齒根圓的直徑必須大于離合器的外徑,負責齒輪無法加工。這樣軸的間距加大。另一種布置方案是離合器的左右部分分別裝在同軸線的軸上,左邊部分接通,得到一級反向轉(zhuǎn)動,右邊接通得到三級反向轉(zhuǎn)動。這種齒輪尺寸小但軸向尺寸大。我們采用第一種方案,通過空心軸中的拉桿來操縱離合器的結(jié)構(gòu)??偛贾脮r需要考慮制動器的位置。制動器可以布置在背輪軸上也可以放在其他軸上。制動器不要放在轉(zhuǎn)速太低軸上,以免制動扭矩太大,是制動尺寸增大。齒輪在軸上布置很重要,關(guān)系到變速箱的軸向尺寸,減少軸向尺寸有利于提高剛度和減小體積。結(jié)束語1、本次課程設(shè)計是針對專業(yè)課程基礎(chǔ)知識的一次綜合性應(yīng)用設(shè)計,設(shè)計過程應(yīng)用了機械制圖、機械原理、工程力學等。2、本次課程設(shè)計充分應(yīng)用了以前所學習的知識,并應(yīng)用這些知識來分析和解決實際問題。3、本次課程設(shè)計進一步掌握了一般設(shè)計的設(shè)計思路和設(shè)計切入點,同時對機械部件的傳動設(shè)計和動力計算也提高了應(yīng)用各種資料和實際動手的能力。4、本次課程設(shè)計進一步規(guī)范了制圖要求,掌握了機械設(shè)計的基本技能。5、本次課程設(shè)計由于學習知識面的狹窄和對一些概念的理解不夠深刻,以及缺乏實際設(shè)計經(jīng)驗,使得設(shè)計黨中出現(xiàn)了許多不妥和錯誤之處,誠請老師給予指正和教導。參考文獻【1】、段鐵群 主編 機械系統(tǒng)設(shè)計 科學出版社 第一版【2】、于惠力 主編 機械設(shè)計 科學出版社 第一版【3】、戴 曙 主編 金屬切削機床設(shè)計 機械工業(yè)出版社【4】、戴 曙 主編 金屬切削機床 機械工業(yè)出版社 第一版【4】、趙九江 主編 材料力學 哈爾濱工業(yè)大學出版社 第一版【6】、鄭文經(jīng) 主編 機械原理 高等教育出版社 第七版【7】、于惠力 主編 機械設(shè)計課程設(shè)計 科學出版社 Bebek, Bearing load Bending stress beam is rate, parameter with the most important influence on design of the crankshaft. Results of bearing loads and web bending stresses are tabulated. must overall systems on parameters of the crankshaft system. Studies on crankshaft of internal combustion engines mainly fo- cus on vibration and stress analyses 19. Although stress analy- ses of crankshafts are available in literature, there are few studies on the effect of counterweight configuration on main bear- ing loads and crankshaft stresses. Sharpe et al. 10 studied balanc- ing of the crankshaft of a V-8 engine using a rigid crankshaft model tions are carried out at engine speed range of 10002000 rpm. Bending stresses at the centres of each web are also calculated. 2. Engine specifications The specifications of in-line six-cylinder diesel engine are given in Table 1. The 9.0 L engine crankshaft has eight counterweights at crank webs 1, 2, 5, 6, 7, 8, 11 and 12. 3D solid model of the crank- shaft is obtained using Pro/Engineer and is shown in Fig. 1. Sche- matic representation of the crankshaft is given in Fig. 2. Static * Corresponding author. Tel.: +90 212 359 7534; fax: +90 212 287 2456. Advances in Engineering Software 40 (2009) 95104 Contents lists available E-mail address: yasin.yilmazboun.edu.tr (Y. Yilmaz). being the main part responsible for power production. Crankshaft system mainly consists of piston, piston pin, con- necting rod, crankshaft, torsional vibration (TV) damper and fly- wheel. Counterweights are placed on the opposite side of each crank to balance rotating inertia forces. In general, counterweights are designed for balancing rates between 50% and 100%. For acceptable maximum and average main bearing loads, mass of counterweights and their positions are important. Maximum and average main bearing loads of an engine depend on cylinder pres- sure, counterweight mass, engine speed and other geometric study on effect of counterweight configuration on main bearing loads and crankshaft stresses is still needed. In this study, counterweight positions and masses of an in-line six-cylinder diesel engine crankshaft system are studied. Maxi- mum and average main bearing forces and crankshaft bending stresses are calculated for 12-counterweight configurations with a zero degree counterweight angle, and for eight-counterweight configurations with 30C176 counterweight angle for 0%, 50% and 100% counterweight balancing rates. Analyses are carried out using Multibody System Simulation Program, ADAMS/Engine. Simula- 1. Introduction New internal combustion engines power, good fuel economy, small engine harmless as possible to the environment. each component of the engine on its be investigated in detail. Crankshaft tion engines have important influence 0965-9978/$ - see front matter C211 2008 Elsevier Ltd. All doi:10.1016/j.advengsoft.2008.03.009 C211 2008 Elsevier Ltd. All rights reserved. have high engine size, and should be as Therefore, the effect of performance should of internal combus- engine performance and optimized counterweights to minimize main bearing loads. Stanley and Taraza 11 obtained maximum and average main bearing loads of four and six-cylinder symmetric in-line engines using a rigid crankshaft model and estimated ideal counterweight mass that resulted in acceptable maximum bearing load. Rigid crankshaft models that are used in counterweight analyses do not consider the effect of crankshaft flexibility on main bearing loads and can lead to considerable errors. Therefore, an extensive Crankshaft models Balancing rate Both configurations show the same trend. The load from gas pressure rather than inertia forces is the An investigation of the effect of counterweight load and crankshaft bending stress Yasin Yilmaz * , Gunay Anlas Department of Mechanical Engineering, Faculty of Engineering, Bogazici University, 34342 article info Article history: Received 11 February 2008 Received in revised form 17 March 2008 Accepted 24 March 2008 Available online 6 May 2008 Keywords: Counterweight configuration abstract In this study, effects of counterweight stress of an in-line six-cylinder ADAMS. In the analysis, rigid, rigid, beam and 3D solid models analyses. Twelve-counterweight terweight configurations with ing rates, are considered. It with increasing balancing Advances in Engineering journal homepage: rights reserved. configuration on main bearing Istanbul, Turkey mass and position on main bearing load and crankshaft bending diesel engine is investigated using Multibody System Simulation Program, and 3D solid crankshaft models are used. Main bearing load results of are compared and beam model is used in counterweight configuration configurations with a zero degree counterweight angle and eight-coun- 30C176 counterweight angle, each for 0%, 50% and 100% counterweight balanc- found that maximum main bearing load and web bending stress increase and average main bearing load decreases with increasing balancing rate. at ScienceDirect Software cate/advengsoft unbalance of each crank throw (with and w/o counterweights) is determined using Pro/Engineer and is given in Table 2. The balanc- ing system data for the crank train are given in Table 3. 3. Modeling of crankshaft system Using ADAMS/Engine, a crankshaft can be modeled in four dif- ferent ways: rigid crankshaft, torsionalflexible crankshaft, beam crankshaft and 3D solid crankshaft. Rigid crankshaft model is mainly used to obtain free forces and torques, and for balancing purposes. Torsionalflexible crankshaft model is used to investi- gate torsional vibrations where each throw is modeled as one rigid part, and springs are used between each throw to represent tor- sional stiffness. Beam crankshaft model is used to represent the torsional and bending stiffness of the crankshaft. Using beam mod- el bending stresses at the webs can be calculated 12. Table 1 Engine specifications Unit 9.0 L engine Bore diameter mm 115 Stroke mm 144 Axial cylinder distance mm 134 Peak firing pressure MPa 19 Rated power at speed kW/rpm 295/2200 Max. torque at speed Nm/rpm 1600/12001700 Main journal/pin diameter mm 95/81 Firing order 1-5-3-6-2-4 Flywheel mass kg 47.84 Flywheel moment of inertia kg mm 2 1.57E+9 Mass of TV damper ring kg 4.94 Mass of TV damper housing kg 6.86 Moment of inertia of the ring kg mm 2 1.27E+5 Moment of inertia of the housing kg mm 2 0.56E+5 Main Bearing #1 Main Bearing #2 Main Bearing #3 Main Bearing #4 Main Bearing #5 Main Bearing #6 Main Bearing #7 Counterweights Fig. 1. 3D solid model of the crankshaft. C3, C4, C5, C6 C1, C2, C7, C8 1, 6 3, 4 2, 5 C1 C2 C3 C4 C5 C6 1 2 Fig. 2. Eight-counterweight arrangement Table 2 Properties of the crank throws Throw 1 Throw 2 Mass (kg) 12.50 9.25 CG position from crank rotation axis (mm) 12.423 31.435 Static unbalance (kg mm) 155.265 290.767 96 Y. Yilmaz, G. Anlas/Advances in Engineering Software 40 (2009) 95104 C7 C8 3 4 5 6 of the 9.0 L engine crankshaft. Throw 3 Throw 4 Throw 5 Throw 6 12.50 12.50 9.28 12.55 11.967 11.966 31.027 11.702 149.734 149.734 287.871 146.856 Elastic 3D solid model of the crankshaft can be obtained using an additional finite element program. The procedure is lengthy and time consuming and usually one ends up with degrees of free- dom in order of millions. To simplify the finite element model, modal superposition technique is used. The elastic deformation of the structure is approximated by linear combination of suitable modes which can be shown as follows: u Uq 1 where q is the vector of modal coordinates andUis the shape func- tion matrix. Table 3 Crankshaft system data Crank radius (mm) 72 Connecting rod length (mm) 239 Mass of complete piston (kg) 3.42 Connecting rod reciprocating mass (kg) 0.92 Reciprocating mass (total per cylinder) (kg) 4.32 Connecting rod rotating mass (kg) 2.01 Y. Yilmaz, G. Anlas/Advances in Engineering An elastic body contains two types of nodes, interface nodes where forces and boundary conditions interact with the structure during multibody system simulation (MSS), and interior nodes. In MSS the position of the elastic body is computed by superposing its rigid body motion and elastic deformation. In ADAMS, this is performed using Component Mode Synthesis” technique based on CraigBampton method 13,14. The component modes contain static and dynamic behavior of the structure. These modes are con- straint modes which are static deformation shapes obtained by giving a unit displacement to each interface degree of freedom (DOF) while keeping all other interface DOFs fixed, and fixed boundary normal modes which are the solution of eigenvalue problem by fixing the entire interface DOFs. The modal transforma- tion between the physical DOF and the CraigBampton modes and their modal coordinates is described by 15 u u B u I C26C27 I0 U C U N C20C21 q C q N C26C27 2 where u B and u I are column vectors and represent boundary DOF and interior DOF, respectively. I, 0 are identity and zero matrices, respectively. U C is the matrix of physical displacements of the inte- rior DOF in the constraint modes. U N is the matrix of physical dis- Fig. 3. Model of the crankshaft system. placements of the interior DOF in the normal modes. q C is the column vector of modal coordinates of the constraint modes. q N is the column vector of modal coordinates of the fixed boundary nor- mal modes. To obtain decoupled set of modes, constrained modes and normal modes are orthogonalized. Elastic 3D solid crankshaft model of the 9.0 L engine is obtained in MSC.Nastran using modal superposition technique. First, 3D so- lid model of the crankshaft that is shown in Fig. 1 is exported to MSC.Nastran and finite element model of the crankshaft, which is characterized by approximately 300,000 ten-node tetrahedral ele- ments and 500,000 nodes is obtained. The modal model of the crankshaft is developed with 32 boundary DOFs associated with 16 interface nodes. Constrained modes obtained from static analy- sis correspond to these DOFs. Flexible crankshaft model is obtained through modal synthesis considering the first 40 fixed boundary normal modes. Therefore flexible crankshaft model is character- ized by a total of 72 DOFs. This model is exported to ADAMS/En- gine and crankshaft system model that is shown in Fig. 3 is obtained. 3D finite element model is run with ADAMS. 4. Forces acting on crankshaft system and balancing Forces in an internal combustion engine may be divided into inertia forces and pressure forces. Inertia forces are further divided into two main categories: rotating inertia forces and reciprocating inertia forces. The rotating inertia force for each cylinder can be written as shown below: F iR;j m R C1 r R C1 x 2 C1C0sinh j j cosh j k3 where m R is the rotating mass that consists of the mass of crank pin, crank webs and mass of rotating portion of the connecting rod; r R is the distance from the crankshaft centre of rotation to the centre of gravity of the rotating mass, x is angular velocity of the crankshaft, and h j is the angular position of each crank throw with respect to Top Dead Centre” (TDC). If there are two counterweights per crank throw, each counterweight force is given by 11 F CWi;j C0m CWi;j C1 r CWi;j C1 x 2 C1C0sinh j c i;j j cosh j c i;j k hi ; i 1;2 j 1;2;.;6 4 where c i,j is the offset angle of counterweight mass from 180C176 oppo- site of crank throw j”. There are two counterweights per throw. i” denotes the counterweight number. The counterweight size that is required to accomplish an assessed balancing rate is U CW K C1U Crank throw m cr-r C1 rC1cosc 2 5 where U CW is the static unbalance of each counterweight, U Crank_throw is the static unbalance of each crank throw, m cr-r is the mass of connecting rod rotating portion, r is the crank radius and K is the balancing rate of the internal couple due to rotating forces. From this formula follows the balancing rate for a given crankshaft and a given counterweight size: K 2 C1 U CW U Crank throw m cr-r C1 rC1cosc 6 For a standard in-line six-cylinder engine crankshaft with three pairs of crank throws disposed at angles of 120C176 that are arranged symmetrical to the crankshaft centre, rotating forces, and first and second order reciprocating forces are naturally balanced. This can be explained by the first and second order vector stars shown in Fig. 4. The six-cylinder crankshaft generates rotating and first Software 40 (2009) 95104 97 and second order reciprocating couples in each crankshaft half that balance each other but which result in internal bending moment. At high speeds, the two equally directed crank throws, 3 and 4 yield a high rotating load on centre main bearing. The rotating inertia force of each cylinder is usually offset at least partially by counterweights placed on the opposite side of each crank. In gen- eral, the counterweights are designed for balancing rates between 50% and 100% of the internal couple. Gas forces in cylinders are acting on piston head, cylinder head and on side walls of the cylinder. These forces are equal to F p;j C0 pD 2 4 C1P cyl;j hC0P cc;j hC138 k; j 1;2;.;6 7 1, 6 2, 5 3, 4 3, 4 1, 6 2, 5 Fig. 4. First and second order vector stars. 0 20 40 60 80 100 120 140 160 180 200 0 90 180 270 360 450 540 630 720 Crank Angle (degree) Pressure (bar) 1000rpm 1200rpm 1350rpm 1675rpm 2000rpm Fig. 5. Gas pressure values at different engine speeds for the 9.0 L engine. Bearing #1 0 25 50 75 100 125 150 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 6. Forces acting on main bearing #1 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #2 0 25 50 75 100 125 150 175 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 7. Forces acting on main bearing #2 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #3 0 25 50 75 100 125 150 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 8. Forces acting on main bearing #3 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #4 0 25 50 75 100 125 150 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 9. Forces acting on main bearing #4 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #5 125 150 Rigid Bam 3D solid 98 Y. Yilmaz, G. Anlas/Advances in Engineering Software 40 (2009) 95104 0 25 50 75 100 0 120 240 360 480 600 720 Crank Angle deg Force kN Fig. 10. Forces acting on main bearing #5 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. where D is cylinder diameter, P cyl is the gas pressure in the cylinder and P cc is the pressure in the crankcase. The gas forces are transmit- ted to the crankshaft through the piston and connecting rod. Cylin- der pressure curves for the 9.0 L engine studied under full load at different engine speeds are given in Fig. 5. Pressure curves are ob- tained using AVL/Boost engine cycle calculation program which simulates thermodynamic processes in the engine taking into ac- count one dimensional gas dynamics in the intake and exhaust sys- tems 16. 5. Main bearing loads: comparison of crankshaft models Main bearing loads are calculated using ADAMSs rigid, beam and 3D solid crankshaft models and compared. In the rigid model, no vibration effects are considered which can lead to considerable errors if vibration effects have a major role on the system (like in multithrow crankshafts). To consider vibration effects beam crank- shaft model is used and main bearing loads and bending stresses at webs are calculated. Rigid model assumes crankshaft to be stati- cally determinate and reaction force of any given bearing depends on the load exerted on the throws adjacent to that bearing. Beam model assumes the crankshaft to be statically indeterminate and the load exerted on a throw affects all bearings. Analyses are car- ried out at an engine speed range of 10002000 rpm. A more sophisticated 3D solid hybrid model that combines FE with ADAMS is used to check the results obtained by beam model. Maximum main bearing load occurs at bearing number two at Bearing #6 0 25 50 75 100 125 150 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 11. Forces acting on main bearing #6 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #7 0 25 50 75 100 125 150 0 120 240 360 480 600 720 Crank Angle deg Force kN Rigid Beam 3D solid Fig. 12. Forces acting on main bearing #7 for rigid, beam and 3D solid crankshaft models at 1000 rpm engine speed. Bearing #1 40 50 60 70 80 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Maximum Bearing K=0% K=50% K=100% Force (kN) Fig. 13. (a) Maximum and (b) average bearing forces at Bearing #2 120 130 140 150 160 K=0% K=50% K=100% Maximum Bearing Force (kN) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Fig. 14. (a) Maximum and (b) average bearing forces at Y. Yilmaz, G. Anlas/Advances in Engineering Software 40 (2009) 95104 99 an engine speed of 1000 rpm, therefore results are plotted in Figs. 612 for 1000 rpm only. Rigid crankshaft model overestimates the maximum main bearing load at bearings 1 and 7 with respect to beam and flexible crankshaft models. However it underestimates the maximum main bearing load at other bearings. For example at bearing 2, beam model gives a maximum main bearing load that is 50% more than that of rigid models because the beam model as- sumes the crankshaft to be statically indeterminate and considers Bearing #1 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) 0 5 10 15 20 Average Bearing K=0% K=50% K=100% Force (kN) bearing #1 for 12-counterweight configurations. Bearing #2 20 25 30 35 40 K=0% K=50% K=100% 1000 1200 1400 1600 1800 2000 Average Bearing Force (kN) Crank Angular Velocity (rpm) bearing #2 for 12-counterweight configurations. bending vibrations. Maximum main bearing load difference of beam and 3D solid models is approximately 5%. Main bearing loads for beam and 3D solid crankshaft models are generally in good agreement. In bearings 3, 5 and 6, 3D solid model gives larger bear- ing loads at firing positions of the cylinders that are not adjacent to bearing. Because obtaining elastic 3D solid models for different counterweight configurations is difficult and time consuming, and beam model gives equally valid results, beam model is used Bearing #3 100 110 120 130 140 K=0% K=50% K=100% Bearing #3 20 25 30 35 40 K=0% K=50% K=100% Maximum Bearing Force (kN) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Average Bearing Force (kN) Fig. 15. (a) Maximum and (b) average bearing forces at bearing #3 for 12-counterweight configurations. Bearing #4 60 70 80 90 100 110 120 K=0% K=50% K=100% Bearing #4 10 15 20 25 30 35 40 K=0% K=50% K=100% Maximum Bearing Force (kN) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Average Bearing Force (kN) Fig. 16. (a) Maximum and (b) average bearing forces at bearing #4 for 12-counterweight configurations. Bearing #6 120 130 140 K=0% K=50% K=100% Bearing #6 35 40 45 50 K=0% K=50% K=100% Bearing #5 100 110 120 130 140 K=0% K=50% K=100% Bearing #5 20 25 30 35 40 K=0% K=50% K=100% Maximum Bearing Force (kN) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Average Bearing Force (kN) Fig. 17. (a) Maximum and (b) average bearing forces at bearing #5 for 12-counterweight configurations. 100 Y. Yilmaz, G. Anlas/Advances in Engineering Software 40 (2009) 95104 100 110 Maximum Bearing Force (kN) 1000 1200 1400 1600 1800 2000 Crank Angular Velocity (rpm) Fig. 18. (a) Maximum and (b) average bearing forces at 20 25 30 1000 1200 1400 1600 1800 2000 Average Bearing Force (kN) Crank Angular Velocity (rpm) bearing #6 for 12-counterweight con
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