二级圆锥齿轮减速器设计说明书

word文档可自由复制编辑一¸设计任务书····················································································2二¸电机的选择计算1.择电机的转速·················································································22.工作机的有效功率···········································································23.选择电动机的型号···········································································3三¸运动和动力参数的计算分配传动比·············································································3各轴的转速·············································································3各轴的功率·············································································44.各轴的转矩····················································································4四¸传动零件的设计计算1.闭式直齿轮圆锥齿轮传动的设计计算·················································42.闭式直齿轮圆柱齿轮传动的设计计算·················································6五¸轴的设计计算减速器高速轴I的设计·······························································9减速器低速轴II的设计····························································113.减速器低速轴III的设计··································································14六¸滚动轴承的选择与寿命计算1.减速器高速I轴滚动轴承的选择与寿命计算·········································162.减速器低速II轴滚动轴承的选择与寿命计算········································173.减速器低速III轴滚动轴承的选择与寿命计算·····································18七¸键联接的选择和验算1.联轴器与高速轴轴伸的键联接·························································192.大圆锥齿轮与低速轴II的的键联接···················································19大圆柱齿轮与低速轴III的的键联接···········································20八¸润滑油的选择与热平衡计算1.减速器的热平衡计算·······································································212.润滑油的选择···············································································22九¸参考文献···················································································23计算内容计算结果一对圆锥滚子轴承的效率η3=0.98一对球轴承的效率η4=0.99闭式直齿圆锥齿传动效率η5=0.95闭式直齿圆柱齿传动效率η6=0.97b.总效率η=η1η22η33η4η5η6=0.96×0.992×0.983×0.99×0.95×0.97=0.808c.所需电动机的输出功率Pr=Pw/η=2.4/0.808=3kw选择电动机的型号查参考文献[1]表4-12.2得表1.1方案号电机类型额定功率同步转速满载转速总传动比1Y100L2-431500142022.2942Y132S-63100096015.072根据以上两种可行同步转速电机对比可见，方案2传动比小且质量价格也比较合理，所以选择Y132S-6型电动机。动和动力参数的计算分配传动比总传动比i=15.072各级传动比：直齿轮圆锥齿轮传动比i12=3.762，直齿轮圆柱齿轮传动比i23=4实际总传动比i实=i12i34=3.762×4=15.048,∵Δi=0.021﹤0.05，故传动比满足要求满足要求。各轴的转速（各轴的标号均已在图1.1中标出）n0=960r/min，n1=n0=960r/min，n2=n1/i12=303.673r/min，n3=n2/i34=63.829r/min，n4=n3=63.829r/min各轴的功率η=0.808Pr=3kw选用三相异步电动机Y132S-6p=3kwn=960r/mini=15.072i12=3.762i23=4n0=960r/minn1=960r/minn2=303.67r/minn3=63.829r/minn4=63.829r/min3.计算内容计算结果p0=pr=3kw，p1=p0η2=2.970kw,p2=p1η4η3=2.965kw,p3=p2η5η3=2.628kw,p4=p3η2η3=2.550kw4.各轴的转矩，由式：T=9.55Pi/ni可得：T0=29.844N·m,T1=29.545N·m,T2=86.955N·m,T3=393.197N·m,T4=381.527N·m四，传动零件的设计计算1.闭式直齿轮圆锥齿轮传动的设计计算a．选材：小齿轮材料选用45号钢，调质处理，HB=217~255，σHP1=580Mpa，σFmin1=220Mpa大齿轮材料选用45号钢，正火处理，HB=162~217，σHP2=560Mpa，σFmin2=210Mpab.由参考文献[2]（以下简称[2]）式（5—33），计算应力循环次数N：N1=60njL=60×960×1×8×11×250=1.267×10eq\o(\s\up5(9),\s\do2())N2=N1/i2=1.267×10/3=2.522×10eq\o(\s\up5(8),\s\do2())查图5—17得ZN1=1.0，ZN2=1.12，由式（5—29）得ZX1=ZX2=1.0，取SHmin=1.0，ZW=1.0，ZLVR=0.92，∴[σH]1=σHP1ZLVRZWZX1ZN1/SHmin=580×0.92=533.6Mpa，[σH]2=σHP2ZN2ZX2ZWZLVR/SHmin=560×1.12×0.92=577Mpa∵[σH]1>[σH]2，∴计算取[σH]=[σH]2=533.6Mpac．按齿面接触强度设计小齿轮大端模数（由于小齿轮更容易失效故按小齿轮设计）：取齿数Z1=21，则Z2=Z1i12=3.762×32=79，取Z2=79∵实际传动比u=Z2/Z1=79/21=3.762，且u=tanδ2=cotδ1，∴δ2=72.2965eq\o(\s\up5(o),\s\do2())=72eq\o(\s\up5(o),\s\do2())1635，δ1=17.7035eq\o(\s\up5(o),\s\do2())=17eq\o(\s\up5(o),\s\do2())4212，则小圆锥齿轮的当量齿数zm1=z1/cosδ1=21/cos17.7035eq\o(\s\up5(o),\s\do2())=23，zm2=z2/cosδ2=79/cos72.2965eq\o(\s\up5(o),\s\do2())=259.79p0=3kwp1=2.970kwp2=2.965kwp3=2.628kwp4=2.550kwT0=29.844N·mT1=29.545N·mT2=86.955N·mT3=393.197N·mT=381.527N·mσHP1=580Mpa，σFmin1=220MpaσHP2=560Mpa，σFmin2=210Mpa[σH]=533.6Mpa圆锥齿轮参数Z1=21Z2=79δ1=17eq\o(\s\up5(o),\s\do2())4212δ2=72eq\o(\s\up5(o),\s\do2())16354.计算内容计算结果此资料来自举哥设计工作室，淘宝网搜索店铺举哥设计工作室进入店铺搜索二级圆锥齿轮减速器可获得此设计的整套资料，下面为此套资料的文件列表截图由[2]图5-14，5-15得YFa=2.8，Ysa=1.55，YFa2=2.23，Ysa2=1.81ZH=√2/cosα×sinα=√2/cos20eq\o(\s\up5(o),\s\do2())×sin20eq\o(\s\up5(o),\s\do2())=2.5由[2]表11-5有ZE=189.8，取Kt·Zeq\o(\s\up5(2),\s\do2(εt))=1.1，由[2]取K=1.4又∵T1=28.381N·m，u=3.762，фR=0.3由[2]式5-56计算小齿轮大端模数：m≥√4KT1YFaYsa/{фRZeq\o(\s\up7(2),\s\do3(1))[σF]（1-0.5фR）2√u2+1}将各值代得m≥1.498由[2]表5-9取m=3㎜d．齿轮参数计算：大端分度圆直径d1=mz1=3×21=63㎜，d2=mz2=3×79=237㎜齿顶圆直径da1=d1+2mcosδ1=63+6cos17.7035=68.715㎜，da2=d2+2mcosδ2=237+6cos72.2965eq\o(\s\up5(o),\s\do2())=238.827㎜齿根圆直径df1=d1-2.4mcosδ1=63-7.2cos17.7035eq\o(\s\up5(o),\s\do2())=56.142㎜df2=d2-2.4mcosδ2=237-7.2×cos72.2965eq\o(\s\up5(o),\s\do2())=231.808㎜齿轮锥距R=√d1+d2/2=122.615㎜，大端圆周速度v=∏d1n1/60000=3.14×63×960/60000=3.165m/s，齿宽b=RфR=0.3×122.615=36.78㎜由[2]表5-6，选齿轮精度为8级由[1]表4.10-2得Δ1=（0.1～0.2）R=（0.1～0.2）305.500=30.05～60.1㎜取Δ1=10㎜，Δ2=14㎜,c=10㎜轮宽L1=（0.1～0.2）d1=（0.1～0.2）93=12.4㎜L2=（0.1～0.2）d2=（0.1～0.2）×291=39㎜e．验算齿面接触疲劳强度：按[2]式5-53σH=ZHZE√2KT1√u+1/[bdeq\o(\s\up7(2),\s\do3(1))u（1-0.5фR）2]，代入各值得圆锥齿轮参数m=3㎜d1=63㎜d2=237㎜da1=68.715㎜da2=238.827㎜df1=56.142㎜df2=231.808㎜R=122.615㎜v=3.165m/sb=36.78㎜Δ1=10㎜Δ2=14㎜c=10㎜L1=12.4㎜L2=39㎜5.计算内容计算结果σH=470.899﹤[σH]=533.6Mpa∴小齿轮满足接触疲劳强度，且大齿轮比小齿轮接触强度高，故齿轮满足接触强度条件f．齿轮弯曲疲劳强度校核：按[2]式5-55由[2]图5-19得YN1=YN2=1.0，由[2]式5-32及m=2﹤5㎜，得YX1=YX2=1.0取YST=2.0，SFmin=1.4，由[2]式5-31计算许用弯曲应力：[σF1]=σFmin1YFa1Ysa1YST/SFmin=220×2.0/1.4=314.29Mpa[σF2]=σFmin2YFa2Ysa2YST/SFmin=210×2.0/1.4=300Mpa∵[σF1]﹥[σF2]，∴[σF]=[σF2]=300Mpa由[2]式5-24计算齿跟弯曲应力：σF1=2KT1YFa1Ysa1/[b1md1（1-0.5фR）]=2×1.4×80070×2.8×1.55/0.85×2×28.935×62=181.59﹤300MpaσF2=σF1YFa2Ysa2/（YFa1Ysa1）=181.59×1.81×2.23/（2.8×1.55）=178.28﹤300Mpa∴两齿轮满足齿跟弯曲疲劳强度2.闭式直齿轮圆柱齿轮传动的设计计算a．选材：小齿轮材料选用45号钢，调质处理，HB=217~255，σHP1=580Mpa，σFmin1=220Mpa大齿轮材料选用45号钢，正火处理，HB=162~217，σHP2=560Mpa，σFmin2=210Mpab.由参考文献[2]（以下简称[2]）式（5—33），计算应力循环次数N：N1=60njL=60×960×1×8×11×250=1.267×10eq\o(\s\up5(9),\s\do2())，N2=N1/i23=1.267×10/3=2.522×10eq\o(\s\up5(8),\s\do2())查图5—17得ZN1=1.05，ZN2=1.16，由式（5—29）得ZX1=ZX2=1.0，取SHmin=1.0，ZW=1.0，ZLVR=0.92，[σH]1=σHP1ZLVRZWZX1ZN1/SHmin=580×1.05×0.92=560.28MPa[σH]=533.6Mpa[σF]=300MpaσHP1=580MpaσFmin1=220MpaσHP2=560MpaσFmin2=210Mpa6.计算内容计算结果[σH]2=σHP2ZN2ZX2ZWZLVR/SHmin=560×1.16×0.92=597.63MPa∵[σH]1>[σH]2，∴计算取[σH]=[σH]2=560.28Mpac.按齿面接触强度计算中心距（由于小齿轮更容易失效故按小齿轮设计）：∵u=i34=4，фa=0.4，ZH=√2/cosα·sinα=√2/cos200·sin200=2.5且由[2]表11-5有ZE=189.8，取Kt·Zeq\o(\s\up5(2),\s\do2(εt))=1.1∴[2]式5-18计算中心距：a≥（1+u）√KT1（ZEZHZε/[σH]）2/（2uφa）=5×√1.1×86955×2.5×189.8/（2×4×0.4×560.28）=147.61㎜由[1]表4.2-10圆整取a=160㎜d．齿轮参数设计：m=（0.007～0.02）a=180（0.007～0.02）=1.26～3.6㎜查[2]表5-7取m=2㎜齿数Z1=2a/m（1+u）=2×160/2（1+4）=32Z2=uZ1=4×32=128取Z2=128则实际传动比i=149/31=4分度圆直径d1=mz1=2×32=64㎜，d2=mz2=2×128=256㎜齿顶圆直径da1=d1+2m=68㎜，da2=d2+2m=260㎜齿基圆直径db1=d1cosα=64×cos20o=60.14㎜db2=d2cosα=256×cos20o=240.56㎜齿根圆直径df1=d1-2.5m=64-2.5×2=59㎜df2=d2-2.5m=256-2.5×2=251㎜圆周速度v=∏d1n2/60×103=3.14×256×63.829/60×103=1.113m/s，中心距a=（d1+d2）/2=160㎜齿宽b=aΦa=0.4×160=64㎜由[2]表5-6，选齿轮精度为8级[σH]=560.28Mpa圆柱齿轮参数m=2㎜Z1=32Z2=128d1=64㎜d2=256㎜da1=8㎜da2=260㎜db1=60.14㎜db2=240.56㎜df1=59㎜df2=251㎜v=1.113m/sa=160㎜b=64㎜7.计算内容计算结果e.验算齿面接触疲劳强度按电机驱动，载荷平稳，由[2]表5-3，取KA=1.0；由[2]图5-4（d），按8级精度和VZ/100=∏dn/60000/100=0.30144，得Kv=1.03；由[2]表5-3得Ka=1.2；由[2]图5-7和b/d1=72/60=1.2，得KB=1.13；∴K=KvKaKAKB=1.03×1.2×1.0×1.13=1.397又∵ɑa1=arccosdb1/da1=arccos（60.14/68）=28.0268eq\o(\s\up5(o),\s\do2())=28eq\o(\s\up5(o),\s\do2())136；ɑa2=arccosdb2/da2=arccos（2240.56/260）=22.0061eq\o(\s\up5(o),\s\do2())=22eq\o(\s\up5(o),\s\do2())017∴重合度εa=[z（tanɑa1-tanɑ）+z（tanɑa1-tanɑ）]/2∏=[32（tan28.0268eq\o(\s\up5(o),\s\do2())-tan20）+128（tan22.0061eq\o(\s\up5(o),\s\do2())-tan20）]=1.773即Zε=√（4-εa）/3=0.862，且ZE=189.8，ZH=2.5∴σH=ZHZEZε√2KT1（u+1）/bd21PAGE\#"'页：'#'

'"u=2.5×189.8×0.862√2×1.397×83510×5.8065/（72×622×5.024）=240.63﹤[σH]=560.28PAGE\#"'页：'#'

'"∴小齿轮满足接触疲劳强度，且大齿轮比小齿轮接触强度高，故齿轮满足接触强度条件f．齿轮弯曲疲劳强度校核：按Z1=32，Z2=128，由[2]图5-14得YFa1=2.56，YFa2=2.18；由[2]图5-15得Ysa1=1.65，Ysa2=1.84由[2]式5-23计算Y=0.25+0.75/εa=02.5+0.75/1.773=0.673由[2]图5-19得YN1=YN2=1.0，由[2]式5-32切m=2﹤5㎜，得YX1=YX2=1.0取YST=2.0，Sfmin=1.4，由[2]式5-31计算许用弯曲应力：[σF1]=σFmin1YFa1Ysa1YST/Sfmin=220×2.0/1.4=314.29Mpa[σF2]=σFmin2YFa2Ysa2YST/Sfmin=210×2.0/1.4=300Mpa[σF1]=314.29Mpa[σF2]=300Mpa8.计算内容计算结果∵[σF1]﹥[σF2]，∴[σF]=[σF2]=300Mpa由[2]式5-24计算齿跟弯曲应力：σF1=2KT1YFa1Ysa1Y/bd1m=2×1.397×83510×2.56×1.65×0.673/（2×64×64）=71.233﹤300MpaσF2=σF1YFa2Ysa2/YFa1Ysa1=71.233×1.84×2.18/（2.56×1.65）=67.644﹤300Mpa∴两齿轮满足齿跟弯曲疲劳强度五，轴的设计计算减速器高速轴I的设计a.选择材料：由于传递中小功率，转速不太高，故选用45优质碳素结构钢，调质处理，按[2]表8-3查得σB=637Mpa，[σb]-1=59Mpab.由扭矩初算轴伸直径：按参考文献[2]有d≥A√p/n∵n0=960r/min，p1=2.97kw，且A=0.11~0.16∴d1≥16~23㎜取d1=20㎜c.考虑I轴与电机伸轴用联轴器联接。并考虑用柱销联轴器，因为电机的轴伸直径为dD=38㎜，查[1]表4.7-1选取联轴器规格HL3（Y38×82，Y30×60），根据轴上零件布置，装拆和定位需要该轴各段尺寸如图1.2a所示该轴受力计算简图如图1.2b，齿轮1受力：（1）圆周力Ft1=2T1/dm1=2×29.545/（64×10-3）=915.52N，（2）径向力Fr1=Ft1·tanα·cosδ1=915.52×tan200·cos17.70350=317.44N，（3）轴向力Fa1=Ft1·tanα·sinδ1=915.52×tan200·sin17.70350=101.33N，e.求垂直面内的支撑反力：∵ΣMB=0，∴Rcy=Ft1（L2+L3）/L2=915.52（74+55）/74=1595.97.97N∵ΣY=0，∴RBY=Ft1-Rcy=915.52-1595.97=-680.45N，[σF]=300MpaσB=637Mpa，[σb]-1=59Mpad1=20㎜选用柱销联轴器HL3（Y38×82，Y30×60）Ft1=915.52NFr1=317.44NFa1=101.33NRcy=1595.97NRBY=-680.45N9.计算内容计算结果∴垂直面内D点弯矩Mdy=0，Meq\o(\s\up5(1),\s\do2(dy))=RcyL3+RBY（L2+L3）=1595.97×55-680.45×129=3662.14N·㎜=3.662N·mf.水平面内的支撑反力：∵ΣMB=0，∴RCz=[Fr1(L3+L2)-Fa1dm1/2]/L2=[317.44（74+55）-680.45×64]/74=419.07N，∵ΣZ=0，∴RBz=Fr1-RCz=317.44-419.07=-101.63N，∵水平面内D点弯矩MDz=0，Meq\o(\s\up5(1),\s\do2(Dz))=RCzL3+RBz(L3+L2)=419.07×55-101.63×129=-7.095N·m合成弯矩：MD=√Meq\o(\s\up5(2),\s\do2(Dz))+Meq\o(\s\up5(2),\s\do2(Dy))=0N·m，Meq\o(\s\up5(1),\s\do2(D))=√Meq\o(\s\up5(12),\s\do2(Dy))+Meq\o(\s\up5(12),\s\do2(Dz))=7.98N·m作轴的扭矩图如图1.2c所示，计算扭矩：T=T1=29.545N·mI.校核高速轴I：根据参考文献[3]第三强度理论进行校核：由图1.2可知，D点弯矩最大，故先验算D处的强度，∵MD＜Meq\o(\s\up5(1),\s\do2(D))，∴取M=Meq\o(\s\up5(1),\s\do2(D))=7.98N·m，又∵抗弯截面系数：w=∏d3min/32=3.14×203/32=1.045×10eq\o(\s\up5(-6),\s\do2())meq\o(\s\up5(3),\s\do2())∴σ=√Meq\o(\s\up5(2),\s\do2())+Teq\o(\s\up5(2),\s\do2())/w=√7.98eq\o(\s\up5(2),\s\do2())+29.545eq\o(\s\up5(2),\s\do2())/1.045×10eq\o(\s\up5(-6),\s\do2())=39.132≤[σb]-1=59Mpa故该轴满足强度要求。2.减速器低速轴II的设计选择材料：因为直齿圆柱齿轮的小轮直径较小（齿跟圆直径db1=62㎜）需制成齿轮轴结构，故与齿轮的材料和热处理应该一致，即为45优质碳素结构钢，调质处理按[2]表8-3查得σb=637Mpa，[σb]-1=59Mpa该轴结构如图1.3a，受力计算简图如图1.3b齿轮2受力（与齿轮1大小相等方向相反）：Ft2=915.52N，Fr2=317.44N，Fa2=101.33N，齿轮3受力：Mdy=0Meq\o(\s\up5(1),\s\do2(dy))=3.662N·mRCz=419.07NRBz=-101.63NMDz=0Meq\o(\s\up5(1),\s\do2(Dz))=-7.095N·mMD=0N·m，Meq\o(\s\up5(1),\s\do2(D))=7.98N·mT=29.545N·mM=7.98N·mσb=637Mpa，[σb]-1=59MpaFt2=915.52NFr2=317.44NFa2=101.33N10.计算内容计算结果（1）圆周力Ft3=2T2/dm3=2×86.955/（64×10-3）=2693.87N（2）径向力Fr3=Ft2·tanα=2693.87×tan200=980.49Nc.求垂直面内的支撑反力：∵ΣMB=0，∴RAy=[Ft2（L2+L3）+Ft3L3]/（L1+L2+L3）=[915.52（70+63）+2693.87×63]/183=1919.26N∵ΣY=0，∴RBY=Ft2+Ft3-Rcy=915.52+2693.87-1919.26=1690.13N∴垂直面内C点弯矩：MCy=RAyL1=1919.26×21.5=41.26N·m，Meq\o(\s\up7(1),\s\do3(Cy))=RBY（L2+L3）-Ft3L2=1690.13×133-2693.87×70=41.26N·m，D点弯矩：MDy=RBYL3=1690.13×63=92.96N·m，Meq\o(\s\up7(1),\s\do3(Dy))=Ray（L1+L2）-Ft2L2=1919.26×120-915.52×70=92.96N·md.水平面内的支撑反力：∵ΣMB=0，∴RAz=[Fr2(L3+L2)+Fr3L3-Fa2dm2/2]/（L1+L2+L3）=[317.44×133＋980.49×63-101.33×238.827/2]/128=750.70N∵ΣZ=0，∴RBz=Fr2+Fr3-RAz=317.44+980.49-750.70=547.23N，∵水平面内C点弯矩：MCz=RAzL1=750.70×50=23.65N·m，M1Cz=RBz(L3+L2)-Fr3L2=547.23×133-980.49×70=-10.55N·m，D点弯矩：MDz=RBzL3=547.23×63=30.10N·m，M1Dz=RAz（L1+L2）-Fa2dm2/2-Fr2L2=750.70×120-101.33×164.9/2-317.44×70=29.92N·me.合成弯矩：MC=√Meq\o(\s\up5(2),\s\do2(Cz))+Meq\o(\s\up5(2),\s\do2(Cy))=47.56N·mMeq\o(\s\up5(1),\s\do2(C))=√Meq\o(\s\up5(12),\s\do2(Cy))+Meq\o(\s\up5(12),\s\do2(Cy))=42.59N·mFt3=2693.87NFr3=980.49NRAy=1919.26NRBY=1690.13NMCy=41.26N·mMeq\o(\s\up7(1),\s\do3(Cy))=41.26N·mMDy=92.96N·mMeq\o(\s\up7(1),\s\do3(Dy))=92.96N·mRAz=750.70NRBz=547.23NMCz=23.65N·mM1Cz=-10.55N·mMDz=30.10N·mM1Dz=29.92N·mMC=47.56N·mMeq\o(\s\up5(1),\s\do2(C))=42.59N·m11.计算内容计算结果MD=√Meq\o(\s\up5(2),\s\do2(Dz))+Meq\o(\s\up5(2),\s\do2(Dy))=97.71N·m，Meq\o(\s\up5(1),\s\do2(D))=√Meq\o(\s\up5(12),\s\do2(Dy))+Meq\o(\s\up5(12),\s\do2(Dz))=97.66N·mf.作轴的扭矩图如图1.3c所，计算扭矩：T=T2=86.955N·mg.校核低速轴II强度，由参考文献[3]第三强度理论进行校核：由图1.3可知，D点弯矩最大，故先验算D处的强度，∵MD＞Meq\o(\s\up5(1),\s\do2(D))，∴取M=Meq\o(\s\up5(1),\s\do2(D))=97.71N·m，∵抗弯截面系数：w=∏d3min/32=3.14×303/32=2.65×10-6m3∴σ=√M2+T2/w=√97.712+86.9552/2.65×10-3=44.27≤[σb]-1=59Mpa（2）.由于C点轴径较小故也应进行校核：∵MC＞Meq\o(\s\up5(1),\s\do2(C))，∴取M=Meq\o(\s\up5(1),\s\do2(C))=47.56N·m，∵抗扭截面系数：w=∏d3min/32=3.14×303/32=2.65×10-6m3∴σ=√M2+T2/w=√47.562+86.9552/2.65×10-6=35.14≤[σb]-1=59Mpa故该轴满足强度要求3.减速器低速轴III的设计a.选择材料：由于传递中小功率，转速不太高，故选用45优质碳素结构钢，调质处理，按[2]表8-3查得σB=637Mpa，[σb]-1=59Mpab.该轴受力计算简图如图1.2b齿轮4受力（与齿轮1大小相等方向相反）：圆周力Ft4=2693.87N，径向力Fr4=980.49Nc.求垂直面内的支撑反力：∵ΣMC=0，∴RBY=Ft4L1/（L1+L2）=2693.87×71/（125+71）=1157.52N∵ΣY=0，∴Rcy=Ft4-RBY=2693.87-1157.52=1536.35N，∴垂直面内D点弯矩MDy=RcyL1=1536.35×55=84.50N·m，Meq\o(\s\up5(1),\s\do2(Dy))=RBYL2=1157.52×125=84.50N·md.水平面内的支撑反力：MD=97.71N·mMeq\o(\s\up5(1),\s\do2(D))=97.66N·mT=86.955N·mM=47.56N·mσB=637Mpa[σb]-1=59MpaFt4=2693.87NFr4=980.49NRBY=1157.52NRcy=1536.35NMDy=84.50N·mMeq\o(\s\up5(1),\s\do2(Dy))=84.50N·m12.计算内容计算结果 ∵ΣMC=0，∴RBz=Fr4L1/（L1+L2）=980.49×70/196=421.31N∵ΣZ=0，∴RCz=Fr4-RBz=980.49-421.31=559.18N，∵水平面内D点弯矩MDz=RCzL1=559.18×71=30.75N·m，Meq\o(\s\up5(1),\s\do2(Dz))=RBzL2=421.31×125=30.76N·m合成弯矩：MD=√Meq\o(\s\up5(2),\s\do2(Dz))+Meq\o(\s\up5(2),\s\do2(Dy))=90.20N·m，Meq\o(\s\up5(1),\s\do2(D))=√Meq\o(\s\up5(12),\s\do2(Dy))+Meq\o(\s\up5(12),\s\do2(Dz))=89.92N·m作轴的扭矩图如图1.2c所，计算扭矩：T=T3=393.197N·mg.校核低速轴III：根据参考文献[3]第三强度理论校核：由图1.2可知，D点弯矩最大，故先验算D处的强度，∵MD＞Meq\o(\s\up5(1),\s\do2(D))，∴取M=MD=90.20N·m，又∵抗弯截面系数：w=∏d3min/32=3.14×423/32=7.27×10-6m3∴σ=√M2+T2/w=√90.202+393.1972/7.27×10-6=55.73≤[σb]-1=59Mpa故该轴满足强度要求。六，滚动轴承的选择与寿命计算1.减速器高速I轴滚动轴承的选择与寿命计算高速轴的轴承既承受一定径向载荷，同时还承受轴向外载荷，选用圆锥滚子轴承，初取d=40㎜，由[1]表4.6-3选用型号为30208，其主要参数为：d=40㎜，D=80㎜，Cr=59800N，е=0.37，Y=1.6，Y0=0.9，Cr0=42800查[2]表9-6当A/R≤е时，X=1，Y=0；当A/RPAGE\#"'页：'#'

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'"b.计算轴承D的受力（图1.5），（1）支反力RB=√Req\o(\s\up5(2),\s\do2(BY))+Req\o(\s\up5(2),\s\do2(Bz))=√36.252+269.272=271.70N，RC=√Req\o(\s\up5(2),\s\do2(cy))+Req\o(\s\up5(2),\s\do2(Cz))=√1184.792+353.692=1236.46N（2）附加轴向力（对滚子轴承S=Fr/2Y）RBz=421.31NRCz=559.18NMDz=30.75N·mMeq\o(\s\up5(1),\s\do2(Dz))=30.76N·mMD=90.20N·mMeq\o(\s\up5(1),\s\do2(D))=89.92N·mT=393.197N·mM=90.20N·m选用圆锥滚子轴承30208（GB/T297-94）RB=271.70NRC=1236.46N13.计算内容计算结果∴SB=RB/2Y=271.70/3=90.57N，SC=RC/2Y=1236.46/3=412.15Nc.轴向外载荷FA=Fa1=101.33Nd.各轴承的实际轴向力AB=max（SB，FA-SC）=FA-SC=310.82N，AC=（SC，FA+SB）=SC=412.15N计算轴承当量动载由于受较小冲击查[2]表9-7fd=1.2，又轴I受较小力矩，取fm=1.5∵AB/RB=310.82/271.70=1.144＞е=0.37，取X=0.4，Y=1.6，∴PB=fdfm（XRB+YAB）=1.8×（0.4×271.7+1.6×310.82）=1090.79N∵AC/RC=412.15/1236.46=0.33＜е=0.37，取X=1，Y=0，∴PC=fdfm（XRC+YAC）=1.2×1.5×1×1236.46=2225.63N计算轴承寿命又PB＜PC，故按PC计算，查[2]表9-4得ft=1.0∴L10h=106（ftC/P）/60n1=106（59800/2225.63）10/3/（60×960）=0.12×106h，按每年250个工作日，每日一班制工作，即L1=60.26＞L=11年故该轴承满足寿命要求。2.减速器低速II轴滚动轴承的选择与寿命计算a.高速轴的轴承既承受一定径向载荷，同时还承受轴向外载荷，选用圆锥滚子轴承，初取d=35㎜，由[1]表4.6-3选用型号为30207，其主要参数为：d=35㎜，D=72㎜，Cr=51500N，е=0.37，Y=1.6，Y0=0.9，Cr0=37200查[2]表9-6当A/R≤е时，X=1，Y=0；当A/RPAGE\#"'页：'#'

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'"b.计算轴承D的受力（图1.6）1.支反力RB=√Req\o(\s\up5(2),\s\do2(BY))+Req\o(\s\up5(2),\s\do2(Bz))=√1919.262+547.232=1995.75NSB=90.57NSC=412.15NFA=101.33NAB=310.82NAC=412.15NPB=1090.79NPC=2225.63N选用圆锥滚子轴承30207（GB/T297-94）RB=1995.75N14.计算内容计算结果RA=√Req\o(\s\up5(2),\s\do2(Ay))+Req\o(\s\up5(2),\s\do2(Az))=√750.702+353.692=922.23N2.附加轴向力（对滚子轴承S=Fr/2Y）∴SB=RB/2Y=1995.75/3.2=623.67N，SA=RA/2Y=922.23/3.2=288.20Nc.轴向外载荷FA=Fa2=101.33N各轴承的实际轴向力AB=max（SB，FA+SA）=SB=623.67N，AA=（SA，FA-SB）=FA-SB=522.34Ne.计算轴承当量动载由于受较小冲击查[2]表9-7fd=1.2，又轴I受较小力矩，取fm=1.5∵AB/RB=623.67/1995.75=0.312＜е=0.37，取X=1，Y=0∴PB=fdfm（XRB+YAB）=1.2×1.5×1995.75=3592.35N∵AA/RA=522.34/922.23=0.566＞е=0.37，取X=0.4，Y=1.6∴PA=fdfm（XRA+YAA）=1.8×（0.4×922.23+1.6×522.34）=2168.34N计算轴承寿命又PB＞PA，故按PB计算，查[2]表9-4得ft=1.0∴L10h=106（ftC/P）/60n2=106（51500/3592.35）10/3/（60×303.673）=0.1833×106h，按每年250个工作日，每日一班制工作，即L1=91.65＞L=11年故该轴承满足寿命要求。3.减速器低速III轴滚动轴承的选择与寿命计算a.高速轴的轴承只承受一定径向载荷，选用深沟球轴承，初取d=55㎜，由[1]表4.6-3选用型号为6211，其主要参数为：d=55㎜，D=100㎜，Cr=33500N，Cr0=25000b.计算轴承D的受力（图1.5）支反力RB=√Req\o(\s\up5(2),\s\do2(BY))+Req\o(\s\up5(2),\s\do2(Bz))=√1157.522+421.312=1231.81N，RC=√Req\o(\s\up5(2),\s\do2(cy))+Req\o(\s\up5(2),\s\do2(Cz))=√1536.352+559.182=1634.95Nc.轴向外载荷FA=0NRA=922.23NSB=623.67NSA=288.20NFA=101.33NAB=623.67NAA=522.34NPB=3592.35NPA=2168.34N选用深沟球轴承6211（GB/T276-94）RB=1231.81NRC=1634.95NFA=0N15.计算内容计算结果d.计算轴承当量动载由于受较小冲击查[2]表9-7fd=1.2，又轴I受较小力矩，取fm=1.5∴PB=fdfmRB=1.2×1.5×1231.8=2256.5N∴PC=fdfmRC=1.2×1.5×1×1634.95=2942.91N计算轴承寿命又PB＜PC，故按PC计算，查[2]表9-4得ft=1.0∴L10h=106（ftC/P）/60n3=106（33500/2942.91）10/3/（60×63.829）=27.41×106h，按每年250个工作日，每日一班制工作，即L1=399.45＞L=11年故该轴承满足寿命要求。七，键联接的选择和验算1.联轴器与高速轴轴伸的键联接采用圆头普通平键（GB1095-79，GB1096-79），由d=30㎜，查[1]表4.5-1得b×h=8×7，因半联轴器长为60㎜，故取键长L=50㎜，即d=30㎜，h=7㎜，L1=L-b=42㎜，T1=28.38N·m，由轻微冲击，查[2]表2-10得[σP]=100Mpa∴σP=4T/dhL1=4×29.844/（30×7×42）=12.87＜[σP]=100Mpa故此键联接强度足够。 小圆锥齿轮与高速轴I的的键联接采用圆头普通平键（GB1095-79，GB1096-79），由d=20㎜，查[1]表4.5-1得b×h=6×6，因小圆锥齿轮宽为55㎜，故取键长L=42㎜即d=20㎜，h=6㎜，L1=L-b=36㎜，T1=29.844N·m，由轻微冲击，查[2]表2-10得[σP]=100Mpa∴σP=4T/dhL1=4×29.844/（20×6×36）=27.63＜[σP]=100Mpa故此键联接强度足够。大圆锥齿轮与低速轴II的的键联接PB=2256.5NPC=2942.91NL=50㎜d=30㎜h=7㎜L1=42㎜T1=28.38N·mL=42㎜d=20㎜h=6㎜L1=36㎜T1=29.844N·m16.计算内容计算结果采用圆头普通平键（GB1095-79，GB1096-79），由d=50㎜，查[1]表4.5-1得b×h=14×9，因大圆锥齿轮宽为50㎜，故取键长L=44㎜即d=50㎜，h=9㎜，L1=L-b=30㎜，T2=86.955N·m，由轻微冲击，查[2]表2-10得[σP]=100Mpa∴σP=4T/dhL1=4×86.955/（50×9×30）=25.76＜[σP]=100Mpa故此键联接强度足够。大圆柱齿轮与低速轴III的的键联接采用圆头普通平键（GB1095-79，GB1096-79），由d=60㎜，查[1]表4.5-1得b×h=18×11，因大圆柱齿轮宽为64㎜，故取键长L=54㎜，即d=60㎜，h=11㎜，L1=L-b=36㎜，T3=393.197N·m，由轻微冲击，查[2]表2-10得[σP]=100Mpa∴σP=4T/dhL1=4×393.197/（60×11×36）=66.19＜[σP]=100Mpa故此键联接强度足够。5.低速轴III与输出联轴器的键联接采用圆头普通平键（GB1095-79，GB1096-79），由d=42㎜，查[1]表4.5-1得b×h=12×8，因半联轴器长为84㎜，故取键长L=72㎜，即d=42㎜，h=8㎜，L1=L