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减速器-圆锥圆柱齿轮减速器设计【链式输送机传动装置】【F=2500M V=0.67ms D=445 L=800mm】
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链式输送机传动装置
减速器-圆锥圆柱齿轮减速器设计【链式输送机传动装置】【F=2500M
V=0.67ms
D=445
L=800mm
减速器
圆锥
圆柱齿轮
设计
链式
输送
传动
装置
2500
0.67
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减速器-圆锥圆柱齿轮减速器设计【链式输送机传动装置】【F=2500M V=0.67ms D=445 L=800mm】,链式输送机传动装置,减速器-圆锥圆柱齿轮减速器设计【链式输送机传动装置】【F=2500M,V=0.67ms,D=445,L=800mm,减速器,圆锥,圆柱齿轮,设计,链式,输送,传动,装置,2500,0.67
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机械设计基础课程设计减速器-圆锥圆柱齿轮减速器设计链式输送机传动装置F=2500 V=0.67 D=445 L=800计算说明书指导教师院 系班 级学 号姓 名目 录设计任务书3传动方案的拟订及说明3电动机的选择3计算传动装置的运动和动力参数5带传动的设计计算8齿轮传动件的设计计算10轴的设计计算.16滚动轴承的选择及计算.38键联接的选择及校核计算.42联轴器的选择.43减速器附件的选择.44润滑与密封.49设计小结.50参考资料.51设计计算及说明结果一、 设计任务书设计一链式输送机传动装置的圆锥圆柱齿轮减速器,已知链式输送机传动装置输送机常温下当地室外作业,工作寿命10年(设每年工作300天),一班制,每班按6小时计算。简图如下:(图2)设计计算及说明结果圆锥圆柱齿轮减速器计算驱动卷筒的转速选用同步转速为1000r/min或1500r/min的电动机作为原动机由图可知,该设备原动机为电动机,传动装置为减速器,减速器为两级展开式三、 选择电动机1)电动机类型和结构型式按工作要求和工作条件,选用一般用途的Y(IP44)系列三相异步电动机。它为卧式封闭结构。2)电动机容量(1)卷筒的输出功率(2)电动机输出功率传动装置的总效率查表2-1,取一对轴承效率轴承=0.99,锥齿轮传动效率锥齿轮=0.96,斜齿圆柱齿轮传动效率齿轮=0.97,联轴器效率联=0.99(位于滚筒与减速器之间的),V带传动带=0.96得电动机到工作机间的总效率为故 (3)电动机额定功率由机械设计(机械设计基础)课程设计表20-1选取电动机额定功率。3)电动机的转速推算电动机转速可选范围,由机械设计(机械设计基础)课程设计表2-1查得单级圆柱齿轮传动比范围,圆锥齿轮传动比范围,V带传动一般 i=25,则电动机转速可选范围为:设计计算及说明结果其中750r/min的电动机不常用,初选同步转速分别为1000r/min和1500r/min的两种电动机进行比较,如下表:方案电动机型号额定功率(KW)电动机转速(r/min)电动机质量(kg)总传动比同步满载1Y112M-62.210009404532.672Y100L1-42.2150014203449.36两方案均可行,选定方案一 ,结构尺寸相对较小,能适合卷筒的工况,选定电动机的型号为Y112M-64)电动机的技术数据和外形,安装尺寸由机械设计(机械设计基础)课程设计表20-1、表20-2查得主要数据,并记录备用。四、计算传动装置的运动和动力参数1)传动装置总传动比2)分配各级传动比根据V带传动一般 i=25,课程设计要求:锥齿轮速比不适宜过大,圆柱齿轮速比不宜过小,带传动速比也不适宜过大。初取,那么圆锥圆柱二级减速器的传动比为因为是圆锥圆柱齿轮减速器,所以那么 设计计算及说明结果3)各轴转速(轴号见图)4)各轴输入功率按电动机所需功率计算各轴输入功率,即5)各轴转矩项目轴1轴2轴3轴4轴5转速(r/min)940303.23108.6828.7728.77功率(kw)1.971.871.7971.731.71转矩(N*m)20.0158.89157.91574.26567.62传动比13.12.793.771设计计算及说明结果五、带轮的计算3.1 带传动设计输入功率P=2.2kW,转速n1=940r/min,带传动比i=3.1表3-1 工作情况系数工作机原动机类类一天工作时间/h10161016载荷平稳液体搅拌机;离心式水泵;通风机和鼓风机();离心式压缩机;轻型运输机1.01.11.21.11.21.3载荷变动小带式运输机(运送砂石、谷物),通风机();发电机;旋转式水泵;金属切削机床;剪床;压力机;印刷机;振动筛1.11.21.31.21.31.4载荷变动较大螺旋式运输机;斗式上料机;往复式水泵和压缩机;锻锤;磨粉机;锯木机和木工机械;纺织机械1.21.31.41.41.51.6载荷变动很大破碎机(旋转式、颚式等);球磨机;棒磨机;起重机;挖掘机;橡胶辊压机1.31.41.51.51.61.8根据V带的载荷变动小,单班工作制(6小时),查机械设计P296表4,取KA1.1。即3.2选择带型普通V带的带型根据传动的设计功率Pd和小带轮的转速n1按机械设计P297图1311选取。图3-1 带型图根据算出的Pd2.42kW及小带轮转速n1940r/min ,查图得:dd=80100可知应选取A型V带。3.3确定带轮的基准直径并验证带速由机械设计P298表137查得,小带轮基准直径为80100mm则取dd1=90mm ddmin.=75 mm(dd1根据P295表13-4查得)表3-2 V带 带轮最小基准直径槽型YZABCDE205075125200355500由机械设计P295表13-4查“V带轮的基准直径”,得=280mm误差验算传动比: (为弹性滑动率)误差 符合要求 带速 满足5m/sv300mm,所以宜选用E型轮辐式带轮。总之,小带轮选H型孔板式结构,大带轮选择E型轮辐式结构。带轮的材料:选用灰铸铁,HT200。3.7确定带的张紧装置 选用结构简单,调整方便的定期调整中心距的张紧装置。3.8计算压轴力由机械设计P303表1312查得,A型带的初拉力F0133.46N,上面已得到=153.36o,z=3,则对带轮的主要要求是质量小且分布均匀、工艺性好、与带接触的工作表面加工精度要高,以减少带的磨损。转速高时要进行动平衡,对于铸造和焊接带轮的内应力要小, 带轮由轮缘、腹板(轮辐)和轮毂三部分组成。带轮的外圈环形部分称为轮缘,轮缘是带轮的工作部分,用以安装传动带,制有梯形轮槽。由于普通V带两侧面间的夹角是40,为了适应V带在带轮上弯曲时截面变形而使楔角减小,故规定普通V带轮槽角 为32、34、36、38(按带的型号及带轮直径确定),轮槽尺寸见表7-3。装在轴上的筒形部分称为轮毂,是带轮与轴的联接部分。中间部分称为轮幅(腹板),用来联接轮缘与轮毂成一整体。表3-5 普通V带轮的轮槽尺寸(摘自GB/T13575.1-92)项目 符号 槽型 Y Z A B C D E 基准宽度 b p 5.3 8.5 11.0 14.0 19.0 27.0 32.0 基准线上槽深 h amin 1.6 2.0 2.75 3.5 4.8 8.1 9.6 基准线下槽深 h fmin 4.7 7.0 8.7 10.8 14.3 19.9 23.4 槽间距 e 8 0.3 12 0.3 15 0.3 19 0.4 25.5 0.5 37 0.6 44.5 0.7 第一槽对称面至端面的距离 f min 6 7 9 11.5 16 23 28 最小轮缘厚 5 5.5 67.5 10 12 15 带轮宽 B B =( z -1) e + 2 f z 轮槽数 外径 d a 轮 槽 角 32 对应的基准直径 d d 60 - - - - - - 34 - 80 118 190 315 - - 36 60 - - - - 475 600 38 - 80 118 190 315 475 600 极限偏差 1 0.5 V带轮按腹板(轮辐)结构的不同分为以下几种型式: (1) 实心带轮:用于尺寸较小的带轮(dd(2.53)d时),如图3-2a。 (2) 腹板带轮:用于中小尺寸的带轮(dd 300mm 时),如图3-2b。 (3) 孔板带轮:用于尺寸较大的带轮(ddd) 100 mm 时),如图3-2c 。 (4) 椭圆轮辐带轮:用于尺寸大的带轮(dd 500mm 时),如图3-2d。(a) (b) (c) (d)图3-2 带轮结构类型根据设计结果,可以得出结论:小带轮选择实心带轮,如图(a),大带轮选择孔板带轮如图(b)五、传动件的设计计算圆锥直齿轮设计已知输入功率,小齿轮转速940r/min,齿数比u=2.79,由电动机驱动,工作寿命10年(设每年工作300天),一班制,输送机工作经常满载,空载起动,工作平稳。选定齿轮精度等级、材料及齿数圆锥圆柱齿轮减速器为通用减速器,速度不高,故选用7级精度(GB10095-88)材料选择 由机械设计(第八版)表10-1选择小齿轮材料为(调质),硬度为280HBS,大齿轮材料为45钢(调质),硬度为240HBS。选小齿轮齿数,大齿轮齿数,取整。则按齿面接触强度设计由设计计算公式进行试算,即确定公式内的各计算数值试选载荷系数计算小齿轮的转矩选齿宽系数4)由机械设计(第八版)图10-21d按齿面硬度查得小齿轮的接触疲劳强度极限,大齿轮的接触疲劳强度极限5)由机械设计(第八版)表10-6查得材料的弹性影响系数6) 计算应力循环次数设计计算及说明结果7) 由机械设计(第八版)图10-19取接触疲劳寿命系数8) 计算接触疲劳许用应力取失效概率为1%,安全系数S=1,得(2) 计算1) 试算小齿轮分度圆直径,代入中较小的值2) 计算圆周速度v设计计算及说明结果3) 计算载荷系数根据,7级精度,由机械设计(第八版)图10-8查得动载系数直齿轮由机械设计(第八版)表10-2查得使用系数根据大齿轮两端支撑,小齿轮作悬臂布置,查机械设计(第八版)表得轴承系数,则接触强度载荷系数4) 按实际的载荷系数校正所算得的分度圆直径,得5) 计算模数m取标准值6) 计算齿轮相关参数7) 圆整并确定齿宽圆整取,设计计算及说明结果1、 校核齿根弯曲疲劳强度1) 确定弯曲强度载荷系数2) 计算当量齿数3) 由机械设计(第八版)表10-5查得齿形系数应力校正系数4) 由机械设计(第八版)图20-20c查得小齿轮的弯曲疲劳强度极限,大齿轮的弯曲疲劳强度极限5) 由机械设计(第八版)图10-18取弯曲疲劳寿命系数6) 计算弯曲疲劳许用应力取弯曲疲劳安全系数,得7)校核弯曲强度设计计算及说明结果根据弯曲强度条件公式进行校核满足弯曲强度,所选参数合适。圆柱斜齿轮设计已知输入功率,小齿轮转速108.68r/min,齿数比u=3.77,由电动机驱动,工作寿命10年(设每年工作300天),一班制,带式输送机工作经常满载,空载起动,工作平稳。1、 选定齿轮精度等级、材料及齿数1) 圆锥圆柱齿轮减速器为通用减速器,速度不高,故选用7级精度(GB10095-88)2) 材料选择 由机械设计(第八版)表10-1选择大小齿轮材料均为45钢(调质),小齿轮齿面硬度为250HBS,大齿轮齿面硬度为220HBS。3) 选小齿轮齿数,大齿轮齿数,大小齿轮一般取互质数,故取4) 选取螺旋角。初选螺旋角 2、按齿面接触强度设计,设计计算及说明结果由设计计算公式进行试算,即(1) 确定公式内的各计算数值1) 试选载荷系数2) 计算小齿轮的转矩3) 选齿宽系数4) 由机械设计(第八版)图10-30选取区域系数5) 由机械设计(第八版)图10-26查得,则6) 由机械设计(第八版)表10-6查得材料的弹性影响系数7) 计算应力循环次数8) 由机械设计(第八版)图10-21d按齿面硬度查得小齿轮的接触疲劳强度极限,大齿轮的接触疲劳强度极限9) 由机械设计(第八版)图10-19取接触疲劳寿命系数设计计算及说明结果10)计算接触疲劳许用应力取失效概率为1%,安全系数S=1,得(2)计算1)试算小齿轮分度圆直径,由计算公式得2) 计算圆周速度v3) 计算齿宽b及模数4) 计算纵向重合度5)计算载荷系数设计计算及说明结果根据,7级精度,由机械设计(第八版)图10-8查得动载系数由机械设计(第八版)表10-3查得由机械设计(第八版)表10-2查得使用系数由机械设计(第八版)表10-13查得 由机械设计(第八版)表10-4查得接触强度载荷系数6)按实际的载荷系数校正所算得的分度圆直径,得7) 计算模数取8) 几何尺寸计算(1) 计算中心距取得(2) 按圆整后的中心距修正螺旋角因值改变不多,故参数、等不必修正(3)计算大小齿轮的分度圆直径设计计算及说明结果(4)计算齿轮宽度圆整后取 3、 校核齿根弯曲疲劳强度1) 确定弯曲强度载荷系数2) 根据重合度,由机械设计(第八版)图10-28查得螺旋角影响系数3) 计算当量齿数4)由机械设计(第八版)表10-5查得齿形系数应力校正系数5) 由机械设计(第八版)图20-20c查得小齿轮的弯曲疲劳强度极限,大齿轮的弯曲疲劳强度极限6)由机械设计(第八版)图10-18取弯曲疲劳寿命系数 设计计算及说明结果7) 计算弯曲疲劳许用应力取弯曲疲劳安全系数,得8) 校核弯曲强度根据弯曲强度条件公式进行校核满足弯曲强度,所选参数合适。六、轴的设计计算输入轴设计1、求输入轴上的功率、转速和转矩 2、求作用在齿轮上的力已知高速级小圆锥齿轮的分度圆半径为设计计算及说明结果而圆周力、径向力及轴向力的方向如图二所示图二设计计算及说明结果3、 初步确定轴的最小直径先初步估算轴的最小直径。选取轴的材料为45钢(调质),根据机械设计(第八版)表15-3,取,得,输入轴的最小直径为安装联轴器的直径,为了使所选的轴直径与联轴器的孔径相适应,故需同时选取联轴器型号。联轴器的计算转矩,查机械设计(第八版)表14-1,由于转矩变化很小,故取,则查机械设计(机械设计基础)课程设计表17-4,选HL1型弹性柱销联轴器,其公称转矩为160000,半联轴器的孔径,故取,半联轴器长度,半联轴器与轴配合的毂孔长度为38mm。4、 轴的结构设计(1) 拟定轴上零件的装配方案(见图三)图三设计计算及说明结果(2) 根据轴向定位的要求确定轴的各段直径和长度1) 为了满足半联轴器的轴向定位,1-2轴段右端需制出一轴肩,故取2-3段的直径2) 初步选择滚动轴承。因轴承同时受有径向力和轴向力,故选用单列圆锥滚子轴承,参照工作要求并根据,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30306,其尺寸为,而。这对轴承均采用轴肩进行轴向定位,由机械设计(机械设计基础)课程设计表15-7查得30306型轴承的定位轴肩高度,因此取3)取安装齿轮处的轴段6-7的直径;为使套筒可靠地压紧轴承, 5-6段应略短于轴承宽度,故取。4)轴承端盖的总宽度为20mm。根据轴承端盖的装拆及便于对轴承添加润滑油 的要求,求得端盖外端面与半联轴器右端面间的距离,故取 5)锥齿轮轮毂宽度为64.86mm,为使套筒端面可靠地压紧齿轮取。6) 由于,故取(3) 轴上的周向定位圆锥齿轮的周向定位采用平键连接,按由机械设计(第八版)表6-1查得平键截面,键槽用键槽铣刀加工,长为50mm,同时为保设计计算及说明结果证齿轮与轴配合有良好的对中性,故选择齿轮轮毂与轴的配合为;滚动轴承与轴的周向定位是由过渡配合来保证的,此处选轴的尺寸公差为k6。(4) 确定轴上圆角和倒角尺寸取轴端倒角为5、 求轴上的载荷载荷水平面H垂直面V支反力F弯矩M 总弯矩扭矩T6、按弯扭合成应力校核轴的强度根据上表中的数据及轴的单向旋转,扭转切应力为脉动循环变应力,取,轴的计算应力前已选定轴的材料为45钢(调质),由机械设计(第八版)表15-1查得,故安全。6、 精确校核轴的疲劳强度(1) 判断危险截面截面5右侧受应力最大(2)截面5右侧设计计算及说明结果抗弯截面系数抗扭截面系数截面5右侧弯矩M为截面5上的扭矩为截面上的弯曲应力截面上的扭转切应力轴的材料为45钢,调质处理。由表15-1查得。截面上由于轴肩而形成的理论应力集中系数及按机械设计(第八版)附表3-2查取。因,经插值后查得又由机械设计(第八版)附图3-2可得轴的材料敏感系数为故有效应力集中系数为设计计算及说明结果由机械设计(第八版)附图3-2的尺寸系数,扭转尺寸系数。轴按磨削加工,由机械设计(第八版)附图3-4得表面质量系数为轴未经表面强化处理,即,则综合系数为又取碳钢的特性系数计算安全系数值故可知安全。中间轴设计1、求中间轴上的功率、转速和转矩 设计计算及说明结果2、求作用在齿轮上的力已知圆柱斜齿轮的分度圆半径而已知圆锥直齿轮的平均分度圆半径而圆周力、,径向力、及轴向力、的方向如图四所示设计计算及说明结果图四3、初步确定轴的最小直径先初步估算轴的最小直径。选取轴的材料为(调质),根据机械设计(第八版)表15-3,取,得,中间轴最小直径显然是安装滚动轴承的直径和设计计算及说明结果4、 轴的结构设计(1) 拟定轴上零件的装配方案(见下图图五)(2)根据轴向定位的要求确定轴的各段直径和长度1)初步选择滚动轴承。因轴承同时受有径向力和轴向力,故选用单列圆锥滚子轴承,参照工作要求并根据,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30306,其尺寸为,。 这对轴承均采用套筒进行轴向定位,由机械设计(机械设计基础)课程设计表15-7查得30306型轴承的定位轴肩高度,因此取套筒直径。2)取安装齿轮的轴段,锥齿轮左端与左轴承之间采用设计计算及说明结果套筒定位,已知锥齿轮轮毂长,为了使套筒端面可靠地压紧端面,此轴段应略短于轮毂长,故取,齿轮的右端采用轴肩定位,轴肩高度,故取,则轴环处的直径为。3) 已知圆柱直齿轮齿宽,为了使套筒端面可靠地压紧端面,此轴段应略短于轮毂长,故取。4)箱体一小圆锥齿轮中心线为对称轴,则取。(3)轴上的周向定位圆锥齿轮的周向定位采用平键连接,按由机械设计(第八版)表6-1查得平键截面,键槽用键槽铣刀加工,长为22mm,同时为保证齿轮与轴配合有良好的对中性,故选择齿轮轮毂与轴的配合为;圆柱齿轮的周向定位采用平键连接,按由机械设计(第八版)表6-1查得平键截面,键槽用键槽铣刀加工,长为56mm,同时为保证齿轮与轴配合有良好的对中性,故选择齿轮轮毂与轴的配合为;滚动轴承与轴的周向定位是由过渡配合来保证的,此处选轴的尺寸公差为m6。(4)确定轴上圆角和倒角尺寸取轴端倒角为5、 求轴上的载荷设计计算及说明结果载荷水平面H垂直面V支反力F弯矩M 总弯矩扭矩T6、按弯扭合成应力校核轴的强度根据上表中的数据及轴的单向旋转,扭转切应力为脉动循环变应力,取,轴的计算应力前已选定轴的材料为(调质),由机械设计(第八版)表15-1查得,故安全。7、精确校核轴的疲劳强度(1)判断危险截面截面5左右侧受应力最大(2)截面5右侧抗弯截面系数抗扭截面系数设计计算及说明结果截面5右侧弯矩M为截面5上的扭矩为截面上的弯曲应力截面上的扭转切应力轴的材料为,调质处理。由表15-1查得。截面上由于轴肩而形成的理论应力集中系数及按机械设计(第八版)附表3-2查取。因,经插值后查得又由机械设计(第八版)附图3-2可得轴的材料敏感系数为故有效应力集中系数为由机械设计(第八版)附图3-2的尺寸系数,扭转尺寸系数。轴按磨削加工,由机械设计(第八版)附图3-4得表面质量系数为设计计算及说明结果轴未经表面强化处理,即,则综合系数为又取合金钢的特性系数计算安全系数值故可知安全。(3)截面5左侧抗弯截面系数抗扭截面系数截面5左侧弯矩M为设计计算及说明结果截面5上的扭矩为截面上的弯曲应力截面上的扭转切应力过盈配合处的,由机械设计(第八版)附表3-8用插值法求出,并取,于是得轴按磨削加工,由机械设计(第八版)附图3-4得表面质量系数为故得综合系数为计算安全系数值设计计算及说明结果故可知安全。输出轴设计1、求输出轴上的功率、转速和转矩 2、求作用在齿轮上的力已知圆柱斜齿轮的分度圆半径而圆周力、径向力及轴向力的方向如图六所示设计计算及说明结果图六设计计算及说明结果3、初步确定轴的最小直径先初步估算轴的最小直径。选取轴的材料为45钢(调质),根据机械设计(第八版)表15-3,取,得,输出轴的最小直径为安装联轴器的直径,为了使所选的轴直径与联轴器的孔径相适应,故需同时选取联轴器型号。联轴器的计算转矩,查机械设计(第八版)表14-1,由于转矩变化很小,故取,则查机械设计(机械设计基础)课程设计表17-4,选HL3型弹性柱销联轴器,其公称转矩为630000,半联轴器的孔径,故取,半联轴器长度,半联轴器与轴配合的毂孔长度为84mm。2、 轴的结构设计(1) 拟定轴上零件的装配方案(见图六)设计计算及说明结果图六(2)根据轴向定位的要求确定轴的各段直径和长度1)为了满足半联轴器的轴向定位,1-2轴段右端需制出一轴肩,故取2-3段的 直径,左端用轴端挡圈定位,按轴端挡圈直径, 半联轴器与轴配合的毂孔长度,为了保证轴端挡圈只压在半联 轴器上而不压在轴的端面上,故1-2段的长度应比略短些,现取 。2) 初步选择滚动轴承。因轴承同时受有径向力和轴向力,故选用单列圆锥滚子轴承,参照工作要求并根据,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30310,其尺寸为,而。左端轴承采用轴肩进行轴向定位,由机械设计(机械设计基础)课程设计计算及说明结果表15-7查得30310型轴承的定位轴肩高度,因此取;齿轮右端和右轴承之间采用套筒定位,已知齿轮轮毂的宽度为71mm,为了使套筒端面可靠地压紧齿轮,此轴段应略短于轮毂宽度,故取。齿轮的左端采用轴肩定位,轴肩高度,故取,则轴环处的直径为。轴环宽度,取。4)轴承端盖的总宽度为20mm,根据轴承端盖的装拆及便于对轴承添加润滑油的要求,求得端盖外端面与半联轴器右端面间的距离,故取 5)箱体一小圆锥齿轮中心线为对称轴,则取。(3)轴上的周向定位齿轮、半联轴器的周向定位均采用平键连接,按由机械设计(第八版)表6-1查得平键截面,键槽用键槽铣刀加工,长为50mm,同时为保证齿轮与轴配合有良好的对中性,故选择齿轮轮毂与轴的配合为;同样,半联轴器与轴的连接,选用平键,半联轴器与轴的配合为,滚动轴承与轴的周向定位是由过渡配合来保证的,此处选轴的尺寸公差为k6。(4)确定轴上圆角和倒角尺寸取轴端倒角为5、求轴上的载荷设计计算及说明结果载荷水平面H垂直面V支反力F弯矩M 总弯矩扭矩T6、按弯扭合成应力校核轴的强度根据上表中的数据及轴的单向旋转,扭转切应力为脉动循环变应力,取,轴的计算应力前已选定轴的材料为45钢(调质),由机械设计(第八版)表15-1查得,故安全。7、精确校核轴的疲劳强度(1)判断危险截面截面7右侧受应力最大(2)截面7右侧抗弯截面系数抗扭截面系数设计计算及说明结果截面7右侧弯矩M为截面7上的扭矩为截面上的弯曲应力截面上的扭转切应力轴的材料为45钢,调质处理。由表15-1查得。截面上由于轴肩而形成的理论应力集中系数及按机械设计(第八版)附表3-2查取。因,经插值后查得又由机械设计(第八版)附图3-2可得轴的材料敏感系数为故有效应力集中系数为由机械设计(第八版)附图3-2的尺寸系数,扭转尺寸系数。轴按磨削加工,由机械设计(第八版)附图3-4得表面质量系数为设计计算及说明结果轴未经表面强化处理,即,则综合系数为又取碳钢的特性系数计算安全系数值故可知安全。七、滚动轴承的选择及计算输入轴滚动轴承计算初步选择滚动轴承,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30306,其尺寸为, ,载荷水平面H垂直面V支反力F则设计计算及说明结果则则则,则 则故合格。中间轴滚动轴承计算初步选择滚动轴承,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30306,其尺寸为,载荷水平面H垂直面V支反力F设计计算及说明结果则则则则,则 则故合格。输出轴轴滚动轴承计算初步选择滚动轴承,由机械设计(机械设计基础)课程设计表15-7中初步选取0基本游隙组,标准精度级的单列圆锥滚子轴承30310,其尺寸为,设计计算及说明结果载荷水平面H垂直面V支反力F则则则则,则 则故合格设计计算及说明结果八、键联接的选择及校核计算输入轴键计算1、 校核联轴器处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:,故单键即可。2、 校核圆锥齿轮处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:,故单键即可。中间轴键计算1、 校核圆锥齿轮处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:,故单键即可。2、 校核圆柱齿轮处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:设计计算及说明结果,故单键即可。输出轴键计算1、 校核联轴器处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:,故单键即可。2、 校核圆柱齿轮处的键连接该处选用普通平键尺寸为,接触长度,则键联接所能传递的转矩为:,故单键即可。九、联轴器的选择在轴的计算中已选定联轴器型号。输入轴选HL1型弹性柱销联轴器,其公称转矩为160000,半联轴器的孔径,故取,半联轴器长度,半联轴器与轴配合的毂孔长度为38mm。输出轴选选HL3型弹性柱销联轴器,其公称转矩为630000,半联轴器的孔径,故取,半联轴器长度,半联轴器与轴配合的毂孔长度为84mm。设计计算及说明结果十、减速器附件的选择由机械设计(机械设计基础)课程设计选定通气帽,A型压配式圆形油标A20(GB1160.1-89),外六角油塞及封油垫,箱座吊耳,吊环螺钉M12(GB825-88),启盖螺钉M8。十一、润滑与密封齿轮采用浸油润滑,由机械设计(机械设计基础)课程设计表16-1查得选用N220中负荷工业齿轮油(GB5903-86)。当齿轮圆周速度时,圆锥齿轮浸入油的深度约一个齿高,三分之一齿轮半径,大齿轮的齿顶到油底面的距离3060mm。由于大圆锥齿轮,可以利用齿轮飞溅的油润滑轴承,并通过油槽润滑其他轴上的轴承,且有散热作用,效果较好。密封防止外界的灰尘、水分等侵入轴承,并阻止润滑剂的漏失。十二、设计小结这次关于带式运输机上的两级圆锥圆柱齿轮减速器的课程设计是我们真正理论联系实际、深入了解设计概念和设计过程的实践考验,对于提高我们机械设计的综合素质大有用处。通过两个星期的设计实践,使我对机械设计有了更多的了解和认识.为我们以后的工作打下了坚实的基础.机械设计是机械工业的基础,是一门综合性相当强的技术课程,它融机械原理、机械设计、理论力学、材料力学、互换性与技术测量、工程材料、机械设计(机械设计基础)课程设计等于一体。这次的课程设计,对于培养我们理论联系实际的设计思想、训练综合运用机械设计和有关先修课程的理论,结合生产实际反应和解决工程实际问题的能力,巩固、加深和扩展有关机械设计方面的知识等方面有重要的作用。本次设计得到了指导老师的细心帮助和支持。衷心的感谢老师的指导和帮助。设计计算及说明结果设计中还存在不少错误和缺点,需要继续努力学习和掌握有关机械设计的知识,继续培养设计习惯和思维从而提高设计实践操作能力。十三、参考文献1、机械设计(第八版)高等教育出版社2、机械设计(机械设计基础)课程设计高等教育出版社5112th IFToMM World Congress, Besanon, June 18-21, 2007 Review of Researches on Ring-plate Gear Reducers with Small Tooth Number Difference Ce Zhang* Yimin Song + Jun Zhang School of Mechanical Engineering, Tianjin University, Tianjin, 300072, P.R. China AbstractRing-plate gear reducer with small tooth number difference has drawn great attention from academic and industrial fields in China due to its special characteristics such as high transmission ratio, high load capacity, simple structure and low manufacturing cost. However, its application is limited because of the lack in systematic analysis method and design theory. Researches on ring-plate reducer are reviewed from aspects of force analysis, load capacity, balance and vibration reduction, mechanical efficiency and serial design of products. In addition, some questions to be solved are discussed. Key words: ring-plate gear reducer, force analysis, load capacity, balance, vibration reduction, efficiency IIntroduction The predecessor of the ring-plate transmission with small tooth number difference is the K-H-V coaxial planetary gear train with small tooth number difference, which has two shortcomings: the limitation of the size of planetary bearing as well as its short service life; and absolute necessity of the output connection mechanism, as shown in Fig.1. Fig. 1. Sketch of K-H-V transmission In 1985, Chen Zongyuan, a Chinese engineer, invented the three-ring-plate transmission, which ingeniously overcame the above-mentioned problems of K-H-V transmission 1. *E-mail: ce_zhang E-mail: ymsong E-mail: zhang_jun The three-ring-plate transmission is shown in Fig 2. Fig. 2. Three-ring-plate transmission Input shaft 1 and support shaft 2, which are the cranks in the parallelogram mechanism, share one eccentric part. The coupler 3 of the parallelogram mechanism, namely the ring-plate, is also an internal gear meshing with the external gear mounted on the output shaft 4 and rotating around a fixed axis. The ring-plate 3 driven by input shaft moves translationally and does not rotate. Three parallelograms are arranged to avoid the uncertainty when the crank and coupler are aligned. The phase angle is 120 degree. The transmission ratio is in the range of 11-99 for a single-stage drive, and can reach 9801 for a double-stage one, and even higher when needed. For the visualization purpose, a CAD solid model of the symmetrical three-ring gear reducer is shown in Fig. 3. Compared with other transmissions, the ring-plate transmission with such characteristics as higher transmission ratio, higher load capacity, simpler structure and lower production cost has drawn academic and industrial attentions as well. In China, this transmission has already been applied in metallurgical industry, mining, transportation, building and light industry. However, due to limitation in research and development and flaws in earlier design, some problems such as vibration, ring-plate break and short service life of the bearings arise. To some degree, its image has been undermined and further application has been limited. In this paper, a review is presented from aspects of force analysis, load capacity, balance and vibration reduction, mechanical efficiency and serial design of products. Fig. 3. Three-ring gear reducer(symmetrical) IIForce Analysis The ring-plate transmission is a kind of over-constraint mechanism, and its force analysis is complicated due to its sensitivity to elastic deformations and manufacturing errors. The compatibility equations of deformations are necessary in the formulation. In 1990s, a few works discussed force analysis of this transmission, but the basic assumptions in the works may deviate from practical application. The results of the analysis, therefore, could be open to discussion. To improve the dynamic performance of the ring-plate reducers, an in-depth analysis on dynamics and vibration of the reducer is necessary. Yang and Zhang 2 established elastodynamic equations of this transmission, and analyzed the natural frequencies, vibration modes and dynamic response. But the influence of ring-plates deformation on dynamic behavior of the system has been ignored. By use of finite element modeling, the stresses, displacements and deformations of the ring-plates have been analyzed in reference 3,4. It has been revealed that tension and bending deformation of the plates have an unnegligible influence on the systems kinematic and dynamic performance. So these two kinds of deformation of the plates cannot be overlooked and are included as one of the parameters of compatibility equations in 4. A force analysis program of the transmission, in which many factors have been taken into consideration, could already be put into actual use 5. The time history of the loads in gears and bearings can be obtained through this program. IIILoad Capacity The researches on load capacity focus on the effect of so-called multi-teeth elastic meshing, which is a special phenomenon in internal gearing with small tooth number difference. As shown in Fig. 4, when the No.0 teeth pair is in conjugation, the clearances between working surfaces of the neighboring teeth pairs (No.1-4) are very small. When the teeth deform under load, the clearances may disappear and multi-teeth meshing takes place. Fig.4. Multi-teeth elastic meshing Early in 1965 6 predicted the existence of this effect and proposed a formula to calculate the clearances value. Later Sunaga 7 further proved its actual existence through an experiment. Based on tooth stiffness, clearances and manufacturing error, Shu 8, 9 deduced the number of simultaneously meshing teeth pairs and measured out the load distribution on the teeth pairs through photo-elastic experiment. According to Sunaga and Shus works, the number of simultaneously meshing teeth pairs cannot exceed 3-5. But in their experiments they failed to give the measured values of errors of the experimented gears. Since the base-pitch errors are random and have an enormous influence upon the load distribution, the results of their experiments cannot be taken as proofs of the actual effect of multi-teeth meshing. Later, the reference 10 made similar effort. Because the base-pitch errors are much greater than the clearances in No. 3 and No.4 teeth pairs, their approaches to the gear errors are open to discussion. The reference 11 precisely deduced the theoretical value of the clearances, calculated the elastic deformations and obtained the number of the simultaneously meshing teeth pair and the load distribution factors. The conclusion is that normally the number of working teeth pairs is only 3, cannot reach 9-18, as some product samples claimed. The most influential error is base-pitch error. According to references 11 base-pitch error may result in instantaneous single-tooth meshing. They advised that the load distribution factors be introduced into fatigue strength estimation of the teeth bending, and that the maximum bending stress be checked at the same time. It is necessary to estimate the influence of errors on load distribution more precisely. The first analysis shows that the positive effect of this phenomenon is too important to be neglected, even though the negative influence of errors should be taken into account. IVBalancing and Vibration Reduction The shaking forces of the three ring-plates counter- balance each other. But there exists a periodic shaking moment, which makes the reducer vibrate. In early design the unsymmetrical arrangement is employed, as shown in Fig. 2. This configuration makes the reducer vibrate strongly. Balancing is also necessary even for the symmetrical arrangement. The image of the transmission has been compromised because of the flaws in early design. Reference 12 has successfully resolved the balancing of the three-ring-plate reducers by mounting four counterweights on the input and support shafts, as shown in Fig. 6. The two-ring-plate reducers (shown in Fig.5) designed by Wang 13,14, which used a split path arrangement to avoid the uncertainty of motion, have a better dynamic performance. However, the structure is much complicated. Fig. 5. Two ring-plates reducer with split path arrangement The authors of this paper proposed a new type of two- ring-plate reducer, as shown in Fig.6, in which the phase angle between the two parallelograms is a little less than , and the split path arrangement can be removed. Four counterweights 7 are mounted on the high-speed shafts and the transmission is well balanced 15. 180 Fig. 6. Two-ring-plates reducer without split path arrangement To obtain load equilibrium among the three parallel mechanisms, the references 16-18 came up with elastic ring and oil-film floating respectively. VMechanical Efficiency The relative sliding between the teeth of internal meshing with small tooth difference is slight. The friction between the gear surfaces is not the main reason of power loss. It differs from other transmissions in this aspect. Liu 19,20 pointed out that there was almost no room left between the convex and concave gear surfaces. The heat given out due to lubricant-film being squeezed out is dominant factor affecting the mechanical efficiency. Liu proposed that solid lubricants be employed. His experiment demonstrated that the efficiency could reach 97.4% by use of solid lubrication, while the efficiency by use of liquid lubrication only 90.6%. Granted Lius point is acceptable, the efficiency of the reducer cannot be correctly evaluated by using the traditional formula of mechanical efficiency evaluation. Therefore the efficiency of the ring-plate reducer is lower than that of the 2K-H planetary reducer. When in the case of higher transmission power and continuous operation, a heat balance analysis is indispensable. VISerial Design of Products The majority of factories in China, which produce this kind of reducers, adopt the recommended standard stipulated in 1995 by Chinese Ministry of Metallurgical Industry 21,22. Both the theoretical analysis and the practical experiences revealed that the actual load capacity is much greater than that given in the standard. Obviously, the original design did not take the effect of multi-teeth elastic meshing into account and was very conservative in load capacity evaluation. Now, having fully accomplished the theoretical analysis of the ring-plate transmission, the conditions for developing more advanced production standard are mature: 1. Kineto-elastostatic analysis method with consideration of deformation of ring-plate 4. 2. Load capacity evaluation method with consideration of the effect of multi-teeth elastic meshing 11. 3. Program package of CAD, which can automatically determine geometrical parameters of internal meshing with small difference of teeth 23. 4. Optimization method of the parameters, which can keep a compromise between the bearing service life and the gears strength 5. Having summed up all the above-mentioned work, the division of mechanical transmission of Tianjin University has already completed a serial design of products, including 400 models (20 sizes each with 20 speed ratio). The load capacity of the newly designed products is much higher than that in the old standard, some even higher than the hardened tooth surface reducers 5. VIISuggestions and Prospect This transmission has a promising prospect. In less demanding applications, many general gear and worm reducers, even two-stage planetary gear reducers should be replaced by it. In order to make the best of its advantages, following work should be carried out. 1. A reliability design method for bending strength of teeth should be established by taking manufacturing errors into account. 2. A completely elastodynamic analysis should be worked out to verify whether it can be employed in the case of higher speed (more than 1500rpm). References 1 Chen Z.Y. and others. Three-ring-plate Reducer (amplifier) Transmission. Chinese Patent No.: CN85106692A, 1985 2 Yang J. M., Zhang C., et al. Elasto-Dynamic Analysis of Three-ring-plate Reducers. Chinese Journal of Mechanical Engineering, 36 (10): 54-58, October 2000. 3 Zhang G.H., Han J. L., Long H. Stress analysis of driving ring board of three-ring type gear reducer. Chinese Journal of Mechanical Engineering, 30 (2): 58-63, April 1994. 4 Zhang Y. X. Elasto-static analysis of three-ring gear reducer in consideration of gear-coupler deformations,Thesis, Tianjin University, January 2005 5 Zhang J. A study on loading capacity of three-ring reducer and design for new products, Thesis, Tianjin University, June 2004 6 Yastrebov, V.M., Yanchenko, T.A. Inz. Vyss. Ucheb. Zaved. Mash, 23(8): 120-128, 1965. 7 Sunaga T., Nishida N.,
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