吉利帝豪轿车离合器设计
吉利帝豪轿车离合器设计
摘
汽车机构里边离合器特部位作通分离接合适量动力传递汽车零部件离合器控制着电力传送命脉发动机直接联系文介绍吉利帝豪轿车中离合器参数计算结构选取材料选择必强度校核
关键词：离合器选型计算
Clutch design of Geely Dihao car
Abstract
In the body of the car the clutch is a special part its role is through separation and joint to transfer the right amount of power to other automotive parts The clutch also controls the lifeblood of power transmission and is in direct contact with the engine This paper mainly introduces the calculation of various parameters the selection of structure the selection of materials and the necessary strength check of the clutch in Geely Dihao car
Keywords Spring clutch Type Selection calculation
目录
1前言 1
11离合器概述 1
12离合器结构 2
13离合器工作原理 2
2离合器结构形式选择 4
21离合器种类选择 4
22选离合器结构形式 4
3选择离合器基参数 5
31离合器备系数β 5
32单位压力P0 6
33摩擦片外径D径d厚度b 7
34确定离合器基参数 8
4确定零件结构计算 9
41动盘设计 9
42离合器盖设计 12
421设计离合器盖 13
422压盘设计 13
43膜片弹簧设计 18
431确定膜片弹簧重参数 19
432载荷F1变形λ间关系 22
433膜片弹簧工作位置确认 24
434离合器完全停止工作分离轴承作载荷F2 25
435分离轴承行程λ2 25
436校核膜片弹簧 26
437制造膜片弹簧工艺 26
5结 27
参考文献 28
谢辞 29
附录 30
1前言
离合器特部位汽车机构里边作通分离接合适量动力传递汽车零部件离合器传递动力变速箱汽车系统动力汽车条件获适量驱动力离合器转换输入输出功率问题速度差半连杆机构够实现
作汽车必备零件离合器存抹杀离合器利分离接合汽车城市里坦路面郊区泥泞凹凸复杂道路够协调车辆系统保护汽车受损坏离合器仅仅感觉更位驾驶员首先掌握技巧利身离合器认知解决种离合器发生情况
11离合器概述
离合器汽车重部件通分离接合程动力传递变速器致 汽车系统间充足动力协调配合汽车运作起汽车工作运行期间离合器操作暂时分离接合离合器分离接合长时间做动作否汽车变速器会出现极问题离合器作汽车起步时够缓进行会汽车系统发生强烈碰撞作汽车遇种状况需换挡时候够减少换挡时零件发生摩擦间接保护零件损耗
离合器动部分动部分传动时候通种方式变速器传递动力普遍利两接触时产生摩擦传递者通特殊方式达传递动力目
汽车遇紧急情况制动时候汽车传动系统会遭受惯性击时候离合器会发生特限制作滑会零件接收较转矩防止损坏
12离合器结构
离合器作极重部位里面许零件存飞轮压盘离合器盖动盘动轴压紧弹簧分离杠杆支承柱摆动销分离套筒分离轴承离合器踏板等零件
13离合器工作原理
离合器手动挡汽车中缺部位驾驶员通离合器切换档位汽车泥泞者凹凸道路会汽车复杂行驶条件道路抛锚驾驶员合理运离合器够带更驾驶感受坐车体验想离合器运臂挥指首先解拥工作原理谓离合器控制引擎轮毂间动力传递作发动机变速器间传递扭矩样汽车获动力离合器般三种工作状态第种全联动第二种部分半联动第三种完全半联动状态汽车处起步时工作状态驾驶员踩离合器通踏板带动里面压盘压盘摩擦片分开压盘飞轮处分开状态样存摩擦力情况正常工作状态汽车离合器里压盘飞轮摩擦片紧密间连起时候摩擦力
图11 离合器结构图
1飞轮 2动盘 3压盘 4离合器盖 5分离杠杆
6浮动销 7调节拨叉 8分离拨叉 9拉杆
10离合器踏板 11分离轴承 12分离轴承座 13回位弹簧
14变速器第轴轴承盖 15离合器轴 16轴承
17压紧弹簧
2离合器结构形式选择
21离合器种类选择
离合器生产现广泛断进步着摩擦离合器现没遗弃相反种促离合器保持接合状态促离合器保持分离状态两部分组成保持接合部分动部分动部分压紧机构保持分离状态操机构动部件惯性非常锥形鼓形缺陷传动装置难换档组合够柔软容易卡
选择盘式摩擦离合器
22选离合器结构形式
正前面描述次毕业设计选离合器结构形式呼欲出压紧弹簧选膜片弹簧动盘考虑选择带扭转减震器
3选择离合器基参数
表31 吉利帝豪轿车整车参数
整备质量
滚动半径
减速器传动
变速器传动
高车速
发动机转矩
发动机功率
起步时发动机转速
1255kg
0303m
5614
4379
170kmh
140N2·m
80KW
2000rmin
31离合器备系数β
离合器设计时候考虑重参数—备系数β重性体现离合器工作时接收传递发动机扭矩否行参数保证离合器够安全执行身务时缓汽车起步时磨损样子更离合器时间延长离合器受损失降低选择β时注意点：
(1)离合器频繁会导致摩擦片受磨损种情况然需够传递发动机扭矩
(2)滑动磨损避免
(3)传动系超载会汽车带危险量避免
发现β果选取太够传递发动机扭矩控制滑动摩损增果β选取太离合器尺寸控制较操作起没灵敏β取值需情况定β取情况：①备动力足状况良发动机②气缸数发动机β取情况：①需考虑滑动磨损起步时②质量较③般柴油机工作时间较长转矩没汽油机稳需β
般说备系数β没固定取值会相应推荐值轿车β取值范围1213间国外推荐值12轿车β取值中符合面取值情况备动力足状况良次设计中吉利帝豪轿车离合器备系数β取值
32单位压力P0
作汽车重零部件摩擦必然存决定着耐磨性单位压力P0关重环节仅影响工作时优劣会影响离合器时间长短选取单位压力P0时候密切关注离合器项特征离合器工作状态
离合器汽车正常行驶时较频繁选取单位压力P0时应该取样减低摩擦片周围热负荷然果特殊状况合适范围增单位压力P0
材料摩擦片单位压力取值范围见表3—2
表3—2 单位压力取值范围
摩擦片材料
单位压力M
石绵基材料
编织
015—025
模压
025—035
粉末冶金材料
铜基
035—050
铁基
金属陶瓷材料
070—150
机材料
025—050
33摩擦片外径D径d厚度b
离合器结构基础形式取决摩擦片外径离合器终成正关系够决定离合器轻重会影响离合器时间需转矩少然尺寸更转矩摩擦片外径D需发动机转矩Temax
Temax发动机转矩
A够影响D条件
轿车选A＝47
发动机转矩Temax摩擦片外径D计算出需利摩擦片参数统表33国摩擦片项尺寸统参数
表33离合器摩擦片项参数
外径
160
180
200
220
240
280
300
320
340
径
110
125
140
150
155
165
175
190
195
厚度
32
35
35
35
35
35
35
35
4
CdD
0687
0694
0700
0682
0646
0589
0583
0594
0574
1c2
0676
0667
0657
0535
0583
0796
0802
0647
0671
c范围0507间摩擦片径外径值摩擦片厚度市场规定三规格分32毫米35毫米4毫米
34确定离合器基参数
次毕业设计选离合器基参数确定：
吉利帝豪发动机扭矩Temax140N2m根验公式： 初选摩擦片外径D轿车A47 173mm
通表32离合器摩擦片标准表次设计选摩擦片尺寸D＝180mmd＝125mmh＝35mmc＝0694mm
采状况良膜片弹簧离合器备系数β应该取值β＝12采机材料摩擦片单位压力P0范围025～05MPa
公式验算单位压力p0
f摩擦面间静摩擦数取f＝03
Z摩擦面数单片离合器Z＝2
P003Mpa终计算出单位压力P003MPa合适范围离合器项基参数计算出合理
4确定零件结构计算
41动盘设计
够较影响离合器性零部件动盘动盘没复杂零件组成仅仅动片摩擦片动盘毂动盘结构部位相显简易质量较需注意点：
(1)需更转动惯量减少齿轮间碰撞
(2)达摩擦损耗目必须轴弹性离合器接合畅易启动摩擦力均匀分布摩擦面
(3)缓离合器受碰撞应装备扭转减震器时防止传动系振
扭转减震器安装动盘里会发挥较作般轿车恰需种动盘减震弹簧动片花键毂连接重物件种特特性—切弹性样掉高频振动具减缓击力作减震摩擦片垫片夹起系统工作时消耗会太够降低低频振动根发动机传动系统特殊设计通项检验知道扭转减震器达佳减震降噪效果汽油发动机汽车常带线性弹性扭转减震器汽车更起减震降噪发动机空挡情况运转时候果柴油机旋转时候够均匀会离合器连接变速器齿轮啮合部分发生碰撞时候果想减轻齿轮间发生碰撞降低碰撞发生噪音采具非线性特性扭转减震器
离合器接合滑非常重时候具轴弹性动片成佳选择做呢？较实方法开T形槽动片然会产生弧形通工具变成波形呈弯曲波形形状相互分离弧形片安装摩擦片离合器接合时受压动片中弯曲波形逐渐扁增加动盘压力扭矩接合变更加畅种做法然说减动片弯曲利刚度方面难做相整体式弹性动片样状况
压成型刚度致波形弹簧压成型动片安装起分开式结构波形弹簧较薄钢板制成容易产生较转动惯量整体式结构方面相分开式结构会显较差点
离合器中动盘毂承受载荷功臣部分载荷动盘毂接收保证离合器部分司职承受载荷时断裂常采材料锻钢硬度较高承受载荷意味着位置较重花键变速器输入轴前端花键花键轴需动盘配合花键轴存偏差时没完全分离滑需动盘轴长度变更长镀铬表面硬度耐磨性提高动盘部零件消耗磨损降低花键孔考虑镀铬
减震弹簧弹簧钢丝60Si2MnA50CrVA65Mn等类型
选取花键结构时国际标准选取
表41 找出动盘毂花键参数
动盘外径Dmm
花键齿数
n
花键外径
D′mm
花键径
d′mm
齿厚
bmm
效齿长
lmm
挤压应力
σ
180
10
27
22
3
20
118
选择完花键必须进行强度校核花键常表面受压力太遭破坏花键普遍受损坏原花键必须计算挤压应力实现强度校核果挤压应力话增加花键毂长度
挤压应力 （41）
花键齿侧面压力
P花键齿侧面压力N
D′花键径m
d＇花键外径m
Z动盘毂数目
Temax发动机转矩N 2·m
n花键齿数
h花键齿工作高度m
l花键效长度m
24490MPa＜[σ挤压]30MPa
42离合器盖设计
离合器盖部分总成包含东西较顾名思义离合器盖外然压盘压紧弹簧传动片部分设计出
421设计离合器盖
离合器盖作壳体然起保护支撑作起两种作必须强刚度条件避免离合器盖工作中种条件产生变形时防止整离合器工作中离合器盖刚度足离合器分崩离析预防离合器讲转动效率摩擦片损耗引起零部件必磨损重点离合器盖必须搞飞轮间联系足够中防止影响离合器甚整轿车系统正常运行驾驶员带更麻烦离合器盖中没麻烦相反种方式定位销定位螺栓者止口中中未尝离合器汽车运行程中较频繁容易发热致温度升高带危险必须考虑冷效果离合器盖装风扇开窗口利离合器中散发产生热量
422压盘设计
压盘出台式键式销式等种连接方式会存缺口目前说传动片较种连接方式作仅结构简单许确定压盘位置起重作传动片身言点受力差异会带更加严重磨损优化种状况根正反特性差异圆周切线方安装分三组组四片组成传动片方式连接方式中佳选择
汽车刚开始工作时需发动机分配动力压盘作通完全压榨离合器片致发动机动力输入传动系统样汽车防冻剂动力驱动驾驶压盘飞轮相连起紧紧相连避免压力两零件连接稳散开
(1)压盘传动方式选择
传动片传动方式较选择作选择然着作离合器接合时候传动片带动压盘起转动作离合器分离时通特弹性带动压盘分离降低控制力传动片两边分固定离合器盖压盘样够发挥作传动片弹簧钢带材料做
传动片压盘中分布通圆周方切线方非常匀称散布压盘样仅压盘结构精练许装配没复杂更利整离合器工作中衡感
(2)确定压盘
灰铸铁HT200制成压盘开始先选择200mm120mm15mm分作压盘外径径厚度
完成选择压盘项参数必须做次温升校核利温升校核公式检验离合器接合时温度升应该5℃10℃间利增加压盘厚度应付温升超范围导致温升高
校核计算公式：
t—温升℃
W—滑磨功N2m
—分配压盘滑磨功占百分单片离合器压盘050
c—热容里指压盘热容铸铁制成铸铁压盘热容c54428J（kg2℃）
m压盘—压盘质量kg
压盘质量：
算压盘质量m压盘236kg接25kg取m压盘25kg
吉利帝豪轿车整备质量ma1255kg车轮滚动半径R0303m汽车刚开始启动时候发动机转速ne2000rmin减速器传动i05614变速器传动ig4379
滑磨功
温升
温升校核结果合适范围次压盘设计合理
(3)设计传动片传动片强度校核
开始先确定选择4组组4片传动片然选定传动片宽度b厚度h孔距l分15mm15 mm40mm传动片布置方式着切线反方布置圆周半径R选取80mm制成传动片材料弹性模量E2×105MPa样初步完成离合器压盘传动片设计参数确定
①计算传动片效长度l1：
l1l15d4015×5325mm
②计算传动片弯曲总刚度：
③通面计算结果推算出σmax驱动应力：
Ⅰ完全分开时 (43)
设计求f0Te0知道σ样值σ0
Ⅱ离合器压盘安装起知道Te0然推出fmax38mm
Ⅲ离合器分正驱反驱传递转矩时候离合器摩擦片损耗极致时候fmax会产生利尺寸链计算出fmax24mm
(Ⅰ)正驱(发动机车轮)：
(Ⅱ)反驱(车轮发动机)：
通计算结果发现正驱反驱两相反驱较危险选取数进行计算时候较谨慎传动片许应力选极限80号钢作传动片材料合适
④动盘没受磨损时候传动片弹性恢复力已出现时候时弹性弯曲变形量离合器接合时候f确定f087mm出K∑017MNm(传动片弯曲总刚度)f 087mm代入弹性恢复力公式求弹性恢复力：
认定行
43膜片弹簧设计
膜片弹簧设计较复杂必须通样品事先选膜片弹簧分析项参数计算项参数数值优胜劣汰制定出合适结构应该非常解关膜片弹簧种属性仅包括固特性知道结构失效样式重膜片弹簧项参数进行检验校核
膜片弹簧设计复杂时带类型弹簧更优异特点：
(1)膜片弹簧身带理想非线性特性种特性仅膜片弹簧离合器接合时保证身压力变整离合器程中传递扭矩力保持变离合器分离时身压力降缓踏板压力
(2)膜片弹簧压紧弹簧分离杠杆结合体样离合器结构方面部零件总体质量方面较缓空间
(3)压缩力高速转动时候没发生变化性方面较稳定
(4) 冷效果明显寿命长
(5) 规模生产降低制造成
膜片弹簧制造难题固特性—非线性弹性日常生产中容易失控断裂磨损频繁科学发展技提升机器进步材料提高膜片弹簧制造变容易许受限次设计选膜片弹簧离合器
431确定膜片弹簧重参数
(1)确定Hh值
Hh指高度截锥高度弹簧片高度厚度值反映膜片弹簧式样弹性特性样弹性特性曲线利值找出够表达佳效果特性曲线
图41载荷F变形λ间关系①Hh＜时载荷增速较快变形λ然増加确极缓慢速度增加着②Hh＝时清楚变形λ段相似速度增加期间载荷没发生变化③时条曲线里段载荷逐渐变变形λ然增加载荷变回增加状态恢复性种特性想找非常适合膜片弹簧④Hh＝时时载荷落差更⑤时载荷呈现负数形式合适
截锥高度H膜片弹簧厚度h值佳范围13次设计选Hh154
载荷F
变形λ
1
2
3
4
5
1 Hh< 2Hh 3
图41 膜片弹簧弹性特性曲线
(2)选择RRr值
弹簧必须定柔软性然会较误差影响汽车Rr值太佳范围115135间较误差影响左结果次设计Rr1255摩擦片均半径R>Rc取R80mm
(3)膜片弹簧起始圆锥底角α
9°15°汽车膜片弹簧起始圆锥底角α佳范围设计选α12°
(4)rarb选择
ra膜片弹簧端半径rb分离轴承作半径膜片弹簧端半径ra值应变速器第轴花键外径取ra16mmrb18mm
(5)分离指数目n切槽宽δ1δ2半径re
膜片弹簧选取n时候注意膜片弹簧结构较n选24结构较膜片弹簧n选12次设计分离指数目选择18δ1范围32～35mm间δ2范围9～10mm间re选择值(rre) ＞δ2规定次设计δ1选择32mmδ2选择10mm
(6)lL
l支承圈均半径L膜片弹簧压盘接触半径均半径lr相稍微r接触半径LR相稍微R次设计选取L78mml64mm
432载荷F1变形λ间关系
载荷F1变形λ间关系式：
(44)
设
(45)
已知E20×105MPaμ03数值代入
F111413F1＇
λ121λ1＇
λ1＇代入式中计算出数结果列表
表42 载荷F变形λ间关系
λ1＇
01
02
04
06
08
1026
12
14
16
1896
20
F1＇
0071
0131
0220
0274
0302
0310
0307
0299
0295
0312
0328
λ1
021
042
084
126
168
215
252
294
336
398
420
F1
812
1494
2508
3129
3443
3539
3499
3414
3367
3559
3743
F1—λ1特性曲线图42
0
500
1000
1500
2000
2500
3000
3500
4000
4500
5000
0
1
2
3
4
5
变形λmm
图42 膜片弹簧佳弹性特性曲线
433膜片弹簧工作位置确认
膜片弹簧进行工作位置确认首先做确认膜片弹簧端变形量膜片弹簧端变形量离合器工作时候选取特性曲线图膜片弹簧压紧力：
校核备系数：
离合器彻底分离时膜片弹簧端变形量
压盘行程
膜片弹簧端变形量离合器刚开始停止工作时压盘行程Δf＇1mm出
摩擦片磨损磨损量
434离合器完全停止工作分离轴承作载荷F2
变形λ1载荷F2关系式：
(46)
取
435分离轴承行程λ2
λ2＇端分离轴承处轴变形量λ1λ2＇关系式：取
宽度系数
分离轴承行程λ2力作分成λ2′λ2＇＇λ2＇＇指弯曲变形量
(47)
关数值代入式
436校核膜片弹簧
离合器完全变形膜片弹簧端变形量
(48)
代入相关数值
计算出数合适范围
437制造膜片弹簧工艺
现国生产膜片弹簧采精度较高钢板材料材料项参数达指标继续通热处理形式继续加工膜片弹簧抗压力想提高进行特殊强压处理增强膜片弹簧抗压承受力然膜片弹簧进行特措施增加膜片弹簧时间保护膜片弹簧种措施喷丸处理膜片弹簧耐磨性重镀铬端部进行高频感应加热淬火应时裂纹处进行挤压处理目消应力源
膜片弹簧表面必须高度精密够存必缺陷裂缝断痕万万够存
5结
次设计根吉利帝豪轿车原始设计参数出设计求通理解离合器工作状态注意事项根已知基公式计算出参数选择零部件材料选择种结构方案离合器中零部件结构选定出相关零部件材料次设计里部分设计迹循循序渐进进行次设计遵循着出设计求实事求计算项数利次计算出数画出相关零件CAD图水限知识储备够次设计难免出现漏洞希老师够加指导
参考文献
(1)徐石安江发潮 编汽年离合器汽车设计书 清华学出版
(2)马明星 编汽车设计课程设计指导书 中国电力出版社
(3)王林超 编汽年构造 中国水利水电出版社
(4)许洪国 编汽车理 民交通出版社
(6)王予 编汽车发动机原理 机械工业出版社
(7)钱川 编新型联轴器离合器选型设计制造工艺实手期北京工业学出版社
(8)骆素君朱诗 编 机械课程设计简明手册 机械工业出版社
(9)濮良贵纪名刚 编 机械设计 东北学出版社
附录
附录1
英文文献
Optimum Design of Automobile Diaphragm Spring Clutch
Zhao Lijun * Liu Tao ** and Song Baoyu ***
* Automotive Engineering Department Harbin Institute of Technology Weihai China Email lijun7422@263net
** Automotive Engineering Department Harbin Institute of Technology Weihai China
*** School of Mechanical and electrical Engineering Harbin Institute of Technology Harbin China
Abstract—In order to meet the design requirements and characters the optimum mathematical model of a light carry truck’s diaphragm spring clutch is established to decrease its size as far as the engine torque is reliably transferred The truck’s clutch design is optimized by MATLAB optimum toolbox The 3D model of the optimized results is established by ProE and the Finite Element Analysis is carried out It’s proved that the design period can be shortened the clutch size can be decreased and the production cost can be lowered
Keywords — Clutch Optimum design MATLAB Finite element
I I NTRODUCTION
As an important part of modern design theories and methods optimum design has become a new effective way for engineering design since 1960s when computer technique was introduced into engineering design It may shorten the design period improve the product performances and lower the production cost [1] [2] [3] [4] FMINCON provided by MATLAB optimum toolbox may effectively solve the problem of optimization Then the optimized results can be analyzed with a 3D model of PROE and a model of finite element analysis in order to prove the correctness of the optimized mathematical model and the rationality of the optimized results Currently optimum methods have been widely applied in automobile design [5] [6] [7] An ideal automobile clutch should transmit the engine torque reliably and save labor in operation In addition its size should be minimized and the cost should be lowered
II THE OPTIMUM MATHEMATICAL MODE L FOR BASIC CLUTH PARAMETERS
The basic clutch performance parameters mainly include the reserve factor β and the unit pressure p The main size parameters include the outside diameter of the friction plate D and the inside diameter d and the thickness of the friction plate b The variation of these parameters influences the structure size and working
performances of the clutch
A The Design of Variables
The reserve factor β can be calculated by formula (1)dependent on the clutch working pressure F the outside diameter of the friction plate D and the inside diameter d
(1)
Where Tc is the static friction moment (N·m) f is thestatic friction factor of friction surfaces F is the working pressure of the platen on the friction surface (N)R is the average friction diameter of the friction plates (mm) Z is the number of friction surfaces c is the ratio between the inside diameter and the outside diameter of the friction plate β is the reserve factor of the clutch
The unit pressure p can be calculated by formula (1) and it is also dependent on F D and d Thus the following variables are chosen for the optimum design of basic clutch parameters
(2)
B The Objective Function
The objective of the optimum design of basic clutch parameters is the minimum of the structure size as far as the clutch performance requirements are met The objective function is
(3)
C The Constraint Conditions
(1) The outside diameter of the friction plate D should confine the maximal circumference velocity D v within 65~70ms ie
(4)
Where nemax stands for the maximal rotate speed (rmin)
(2) The ratio between the inside diameter and the outsidediameter of the friction plate c should be within053~070 ie
(5)
(3) To ensure that the clutch can transfer the torque reliably and prevent the overload of transmitting systemβ should be chosen within 12~40 ie
(6)
(4) In order to install the torsion absorber the inside diameter of the friction plate d should be 50 mm greater than the absorber spring’s position diameter 2 R ie
(7)
(5) In order to reflect the capability of protection against overload while the clutch transmits torch the transferred torch of a unit area should be less than the permitted value ie
(8)
Where C0T is the transferred torch of a unit area of friction (Nmmm 2 ) [] C0 T is the permitted value of C0T (N·mmm 2 )
(6) In order to lower the heat load of the clutch friction and prevent the abrasion of the friction plate the unit pressure should be decided according to different types of the trucks and according to the materials of frictionwithin a maximal range between 010~150MPa ie
(9)
(7) In order to reduce the slippery friction when the automobile starts and prevent the burning of friction plate due to a toohigh surface temperature the friction work of a unit friction area for each joint should be less than the permitted value ie
(10)
where ω is the slippery friction work of a unit friction area (Jmm 2 ) ω is the permitted value of ω (Jmm 2 ) W is the total slippery friction work for one clutch joint when the automobile starts (J)
III THE OPTIMUM MATH MATICAL MODEL OF THE DIAPHRAGM SPRING
The optimum design of the diaphragm spring means deciding a group of fundamental parameters of the springmaking its torsion features with load meet the requirements of the clutch The spring intensity should also meet the design requirements so that the optimal integrating results can be reached [5]
A The Objective Function
The optimum design of the diaphragm spring tries to make the capabilities of the spring meet the requirements of the clutch (transfer the torch reliably and prevent the overload of transmitting system) This means that the working pressing force of the diaphragm spring should
not fluctuate too much with the abrasion of the friction plate The objective function of the optimum diaphragm spring design is the minimum average variation of the diaphragm spring’s working pressing force within the limit of utmost abrasion That is
(11)
Where P is the working pressing force of the new friction plate’s diaphragm spring (N)pi is the working pressing force of the diaphragm spring in the process of abrasion (N) N is the number of equal divisions within the limit of utmost abrasion
B The Design of Variables
The relation between the load of diaphragm spring F and its torsion λ is
(12)
where E is the modulus of elasticity (MPa) μ is the Poisson ratio H is the height of the plate spring’s inner frustum in free state (mm) h is the thickness of the spring armor plate (mm) R is the diameter of the plate spring at the main aspect in free state (mm) r is the diameter of the plate spring at the little end in free state (mm) R is the diameter of the platen at the load point (mm) r is the diameter of the supporting ring at the load point (mm)
From formula (12) it can be concluded that the six sizeparameters H h R r R and 1 r should be chosen as the variables for optimum design as well as the deformation of the main aspect Bλ in accordance with the spring’s working pressing force 1B F at the joint working point ie
(13)
C The Constraint Conditions
(1) To ensure that the clutch transfer the torch reliablythe working pressing force of the diaphragm spring F should be equal to the required working pressing force YF when the clutch rubs
(14)
(2) To ensure that the operating points A B C have their proper positions (A should be to the left of protruding point M B should be close to the inflexion point H C should be close to the concave point N as is shown by Figure 1) the position of 1B λ relative to1Hλ should be correctly chosen Usually 1B 1H 08 ~10 λ λ ie
(15)
(3) To ensure that the clutch can transfer the torch reliably after the friction plate abrades and considering the drop of friction factor the working pressing force of the spring after the friction plate abrades1AF should be greater than or equal to the pressing force of the new friction plate 1BF ie
(16)
(4) In order to meet the requirements of clutch performances h H of the spring and the initial base cone angle ( ) r R H − ≈ α should be within a certain range ie
(17)
(18)
(5) In order to make the friction plate’s working pressing force equally distributed the diameter of the platen at the load point1R should be between the average radius of the friction plate and the outer radius ie
(19)
(6) The separated finger of the diaphragm spring can separate the levers thus the lever ration should be chosen within a certain range ie
(20)
(7) The maximal point of the diaphragm spring pressing force σtBmax should not be greater than the permitted value ie
(21)
(8) To meet the requirements of fatigue intensity the maximal pressing force of point A during the process of spring working σtAmax should not be greater than the permitted value ie
(22)
IV THE INSTANCE OF CALCULATION
The optimum calculation of HB1030G light carry truck’s diaphragm spring clutch is carried out with FMINCON provided by MATLAB optimum toolbox [7]The optimized results are compared with the original parameters The optimized results of basic clutch parameters are shown in Table 1 and the optimized results of basic diaphragm spring parameters are shown in Table 2 The optimized results have been rounded By calculation it’s concluded that the round meets the constraint conditions Thus the optimized results can reduce the total size of the clutch
TABLE I THE OPTIMIZED RESULTS OF BASIC CLUTH PARAMETERS
The Parameters
Before Optimized
After
Optimized
Round
Working Pressure of the
Friction Plate (N)
Outside Diameter of the
FrictionPlate (mm)
Inside Diameter of the Friction Plate (mm)
11932
250
150
12988
2235
1338
12706
225
135
TABLE II THE OPTIMIZED RESULTS OF BASIC DIAPHRAGM SPRING PARAMETERS
The Parameters
Before
Optimized
After
Optimized
Round
Height of the Inside Frustum (mm)
Thickness of the Armor Plate (mm)
Diameter of the Plate Spring at the Main Aspect in Free State (mm)
Diameter of the Plate Spring at the Little End in Free State (mm)
Diameter of the Platen at the Load Point (mm)
Diameter of the Supporting Ring at the Load Point (mm)
Deformation at the Working Point (mm)
5
2
205
150
200
155
368
44
21
1786
1227
1753
1278
35
45
2
180
123
176
128
35
V THE FINITE ELEMENT ANALYSIS
The Finite Element Analysis (FEA) can analyze the structure intensity and dynamic features of the components As a modern design method it may overcome the problems of traditional design methods The efficiency and precision are greatly improved It may even solve the issues which traditional methods are at wit’s end A special advantage of FEA is that it can solve
the mechanical issues with arbitrary structures and boundary conditions It has been widely used in mechanical structure design and calculation and is proved to be a reliable theoretical calculating method With the prevalence and development of computers FEA has become a crucial part of CAD and CAE It is an indispensable tool for designing almost every machine and automobile
Based on the former optimized results a 3D model of the clutch is established with ProE and the finite element analysis is carried out When analyzing first of all the clutch cover and diaphragm spring are assembled together then all DOFs of the clutch cover are restricted at the face where the flywheel combines the clutch coverAt last the evenly distributed pressure is loaded on the clutch cover and the diaphragm spring respectively along the circumference With the modules of solid164 and the aptitude mesh the clutch cover is divided into 10287 units and 10012 nodes and the diaphragm spring is divided into 8849 units and 8512 nodes After analysis
the stress and strain of the clutch cover are shown in Figure 2 and the stress and strain of the diaphragm spring are shown in Figure 3 and their maximum stresses respectively are 73MPa and 1367Mpa The maximum reactive forces are smaller than the safe reactive force It can be proved that the optimum mathematical model is correct and the optimized results are reasonable
VI CONCLUSION
The optimum design and finite element analysis have been carried out in order to improve the design of clutch and increase the efficiency of design The optimum mathematical model of the clutch has been established according to the design requirements and characters so that the weight and size can be reduced as far as the intensity rigidity and life span of the spare parts are
ensured The optimum design is based on MATLAB toolbox A 3D model of the clutch has been established with ProE according to the optimized results and the finite element analysis has been carried out to prove the correctness of the mathematical model and the rationality of the optimized results It has been proved that the optimum design is significant for increasing the design efficiency reduce the size and weight and lower the production cost
REFERENCES
[1] He Xianzhong Li Ping Optimum Technology and Its ApplicationBeijing Institute of Technology Press 1995 pp 2528
[2] Liu Weixin Mechanical Optimum Design Tsinghua University Press 1997 pp 5056
[3] Jiang Dazhi Chi Guotai and Lin Jianhua Decisionmaking model for optimization of loan＇s portfolio based on efficient boundary Journal of Harbin Institute of Technology 5th ed 2002pp 614617
[4] Zhu Xuejun Wang Anlin and Zhang Huiqiao Robust Design for MechanicalStructural System by Using Non Stationary Penalty Function with GA Mechanical Science and Technology Vol 192000 pp 4951
[5] Xia Changgao Zhu Mangmao A Study on Mathematic Model of Optimum Design for Automotive Clutch Diaphragm SpringAutomobile Technology Vol 12 1996 pp 1215
[6] Cui Shengmin Yang Zhanchun Optimum Design of Automobile Wheels Based on Definite Element Analysis Machine DesignVol 9 2001 pp 4142
[7] Su Jinming A Guidance of MATLAB61 Practical ApplicationPublishing House of Electronics Industry2002 pp 4855
中文翻译
汽车膜片弹簧离合器优化设计
赵丽君刘涛宋宝玉
威海市哈尔滨工业学汽车工程部电子邮件 lijun7422@263 Net
威海哈尔滨工业学汽车工程系
哈尔滨哈尔滨工业学机电工程学院
摘满足轻型载货汽车膜片弹簧离合器设计求特点建立膜片弹簧离合器优化数学模型便传递发动机扭矩情况减膜片弹簧离合器体积 利 matlab 优化工具箱卡车离合器进行优化设计 利 pro e 软件建立优化结果三维模型进行限元分析分析 实践证明缩短设计周期减离合器尺寸降低生产成
I引
优化设计作现代设计理方法重组成部分20世纪60年代计算机技术引入工程设计已成工程设计种新效方法 缩短设计周期提高产品性降低生产成[1] [2] [3] [4] 利 matlab 优化工具箱提供 fmincon 算法效解决优化问题 然利 pro e 三维模型限元分析模型优化结果进行分析验证优化数学模型正确性优化结果合理性 目前优化方法已广泛应汽车设计[5][6][7] 理想汽车离合器应传递发动机扭矩节省操作力 外应量减少成应降低
II基线索参数优数学模型
离合器基性参数包括备系数单位压力 p 尺寸参数包括摩擦片外径 d摩擦片径 d 摩擦片厚度 b 参数变化影响着离合器结构尺寸工作性
A 变量设计
根离合器工作压力 f摩擦片外径 d 摩擦片径 d公式(1)计算备系数
(1)
中 tc 摩擦片静摩擦力矩(n · m) f 摩擦片静摩擦系数f 摩擦片摩擦片工作压力(n) r 摩擦片均摩擦直径(mm) z 摩擦片数量c 摩擦片径外径离合器备系数 基离合器参数优化设计中选择变量
(2)
B目标函数
基离合器参数优化设计目标满足离合器性求前提离合器结构尺寸 目标函数
(3)
C 约束条件
(1)摩擦片外径 d 应周速度 d v 限制65ー70m s
(4)
Nemax 代表转速(r min)
(2)摩擦片径外径 c 应053ー070间
(5)
(3)保证离合器传递扭矩防止传递系统载应选择12ー40范围
(6)
(4)安装扭转吸收器摩擦片 d 径应吸收器弹簧位置直径2r 50毫米
(7)
(5)反映离合器传递火炬时载保护力单位面积传递火炬应允许值
(8)
中 单位摩擦面积传递火炬(n m mm 2) (n m mm 2)允许值
(6)降低离合器摩擦热负荷防止摩擦片磨损应根汽车类型摩擦材料确定单位压力范围010ー150 mpa
(9)
(7)减少汽车起动时滑动摩擦防止摩擦片表面温度高烧毁接头单位摩擦面积摩擦功应允许值
(10)
中单位摩擦面积滑摩擦功(j mm2) 允许值(j mm2) w 汽车启动时离合器接头总滑摩擦功
III膜片弹簧佳数学模型
膜片弹簧优化设计指确定组弹簧基参数负荷扭转特性满足离合器求 弹簧强度应满足设计求便达佳综合效果[5]
A 目标函数
膜片弹簧优化设计弹簧性满足离合器求(传递火炬防止传递系统载) 意味着膜片弹簧工作压力会着摩擦片磨损产生太波动 膜片弹簧优化设计目标函数膜片弹簧工作压力极限磨损范围均变化量
(11)
中 p 新型摩擦板膜片弹簧工作压力(n) pi 膜片弹簧磨损程中工作压力(n) n 磨损极限等分数
B 变量设计
膜片弹簧 f 载荷扭矩关系
(12)
中 e 弹性模量(mpa) 泊松 h 板弹簧状态锥体高度(mm) h 弹簧装甲板厚度(mm) r 板弹簧状态部位直径(mm) r 板弹簧状态端部直径(mm) r 板弹簧加载点压板直径(mm) r 加载点压板直径(mm) r 加载点支撑环直径(mm)
公式(12)中出结弹簧接头工作点工作压力1b f选择六尺寸参数 h h r r 1r 作优化设计变量
(13)
C 约束条件
(1) 保证离合器传递火炬性膜片弹簧工作压紧力 f 应等离合器摩擦时需工作压紧力 f
(14)
(2)确保操作点 a b c 适位置(a 应该凸点 m 左边b 应该接拐点 hc 应该接凹点 n图1示) 应该正确选择1b 相 to1h 位置 通常1b1h 08 ~ 10
(15)
(3)确保离合器摩擦片磨损传递摩擦炬考虑摩擦数降低摩擦片磨损弹簧工作压力应等新摩擦片磨损弹簧压力
(16)
(4)满足离合器性求弹簧 h h 初始基座锥角() r h≈ 应定范围
(17)
(18)
(5)摩擦板工作压力分布均匀加载点1r 处压板直径应该摩擦板均半径外半径间
(19)
(6)膜片弹簧分离指分离杠杆杠杆应选择定范围
(20)
(7)膜片弹簧压紧力点应允许值
(21)
(8)满足疲劳强度求弹簧加工 tamax 程中 a 点压力应允许值
(22)
IV计算实例
利 matlab 优化工具箱提供 fmincon hb1030g 轻型载货汽车膜片弹簧离合器进行优化计算[7] 优化结果原参数进行较 基离合器参数优化结果见表1基膜片弹簧参数优化结果见表2 优化结果已圆满完成 通计算出该轮满足约束条件 优化结果减离合器总体尺寸
表基离散参数优化结果
参数
优化前
优化
工作点变形(毫米)
摩擦片工作压力(n)
摩擦片外径(毫米)
摩擦板径(毫米)
11932
250
150
12988
2235
1338
12706
225
135
表二膜片弹簧基参数优化结果
参数
优化前
优化
工作点变形(毫米)
锥体高度(毫米)
钢板厚度(毫米)
板弹簧状态方面直径(毫米)
板弹簧端状态直径(毫米)
压板加载点直径(毫米)
支撑环负载点直径(毫米)
工作点变形(毫米)
5
2
205
150
200
155
368
44
21
1786
1227
1753
1278
35
45
2
180
123
176
128
35
V限元分析
限元分析分析构件结构强度动态特性 作种现代设计方法克服传统设计方法存问题 提高效率精度 甚解决传统方法法解决问题 限元法特殊优点解决意结构边界条件力学问题 机械结构设计计算中广泛应种理计算方法 着计算机技术普发展限元分析已成计算机辅助设计计算机辅助工程重组成部分
优化结果基础利 pro e 软件建立离合器三维模型进行离合器限元分析优化设计 分析时首先离合器盖膜片弹簧装配起然离合器盖度限制飞轮离合器盖结合面压力均匀分布离合器盖膜片弹簧分圆周加载 采 solid164模块 aptitude 啮合离合器盖分10287单元10012节点膜片弹簧分8849单元8512节点 分析离合器盖应力应变图2示膜片弹簧应力应变图3示应力分73mpa 1367mpa 反作力安全反作力 证明优化数学模型正确性优化结果合理性
VI结
改进离合器设计提高设计效率进行优化设计限元分析 根离合器设计求特点建立离合器优化数学模型保证零件强度刚度寿命前提减离合器重量尺寸 基 matlab 工具箱进行优化设计 根优化结果 proe 软件建立离合器三维模型进行限元分析试验验证数学模型正确性优化结果合理性 实践证明优化设计提高设计效率减尺寸重量降低生产成具重意义
参考文献
[1]贤忠李佳技术应 北京理工学出版社1995年第2528页
[2]刘伟新机械优化设计 清华学出版社1997年版第5056页
[3]江志池国泰林建华基效边界贷款组合优化决策模型 哈尔滨工业学 2002年第5版第614617页
[4]朱学军王安林张慧乔利 ga 非稳惩罚函数机械 结构系统进行稳健设计 机械科学技术 第卷 192000pp 4951
[5]夏昌高朱满茂汽车离合器膜片弹簧优化设计数学模型研究 汽车工艺 第卷 121996pp 1215
[6]崔胜民杨占春基限元分析汽车车轮优化设计 机械设计 第卷 92001pp 4142
[7]苏金明实际应指南 电子工业出版社2002年第4855页
附录2
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吉利帝豪轿车离合器设计
吉利帝豪轿车离合器设计
摘
汽车机构里边离合器特部位作通分离接合适量动力传递汽车零部件离合器控制着电力传送命脉发动机直接联系文介绍吉利帝豪轿车中离合器参数计算结构选取材料选择必强度校核
关键词：离合器选型计算
Clutch design of Geely Dihao car
Abstract
In the body of the car the clutch is a special part its role is through separation and joint to transfer the right amount of power to other automotive parts The clutch also controls the lifeblood of power transmission and is in direct contact with the engine This paper mainly introduces the calculation of various parameters the selection of structure the selection of materials and the necessary strength check of the clutch in Geely Dihao car
Keywords Spring clutch Type Selection calculation
目录
1前言 1
11离合器概述 1
12离合器结构 2
13离合器工作原理 2
2离合器结构形式选择 4
21离合器种类选择 4
22选离合器结构形式 4
3选择离合器基参数 5
31离合器备系数β 5
32单位压力P0 6
33摩擦片外径D径d厚度b 7
34确定离合器基参数 8
4确定零件结构计算 9
41动盘设计 9
42离合器盖设计 12
421设计离合器盖 13
422压盘设计 13
43膜片弹簧设计 18
431确定膜片弹簧重参数 19
432载荷F1变形λ间关系 22
433膜片弹簧工作位置确认 24
434离合器完全停止工作分离轴承作载荷F2 25
435分离轴承行程λ2 25
436校核膜片弹簧 26
437制造膜片弹簧工艺 26
5结 27
参考文献 28
谢辞 29
附录 30
1前言
离合器特部位汽车机构里边作通分离接合适量动力传递汽车零部件离合器传递动力变速箱汽车系统动力汽车条件获适量驱动力离合器转换输入输出功率问题速度差半连杆机构够实现
作汽车必备零件离合器存抹杀离合器利分离接合汽车城市里坦路面郊区泥泞凹凸复杂道路够协调车辆系统保护汽车受损坏离合器仅仅感觉更位驾驶员首先掌握技巧利身离合器认知解决种离合器发生情况
11离合器概述
离合器汽车重部件通分离接合程动力传递变速器致 汽车系统间充足动力协调配合汽车运作起汽车工作运行期间离合器操作暂时分离接合离合器分离接合长时间做动作否汽车变速器会出现极问题离合器作汽车起步时够缓进行会汽车系统发生强烈碰撞作汽车遇种状况需换挡时候够减少换挡时零件发生摩擦间接保护零件损耗
离合器动部分动部分传动时候通种方式变速器传递动力普遍利两接触时产生摩擦传递者通特殊方式达传递动力目
汽车遇紧急情况制动时候汽车传动系统会遭受惯性击时候离合器会发生特限制作滑会零件接收较转矩防止损坏
12离合器结构
离合器作极重部位里面许零件存飞轮压盘离合器盖动盘动轴压紧弹簧分离杠杆支承柱摆动销分离套筒分离轴承离合器踏板等零件
13离合器工作原理
离合器手动挡汽车中缺部位驾驶员通离合器切换档位汽车泥泞者凹凸道路会汽车复杂行驶条件道路抛锚驾驶员合理运离合器够带更驾驶感受坐车体验想离合器运臂挥指首先解拥工作原理谓离合器控制引擎轮毂间动力传递作发动机变速器间传递扭矩样汽车获动力离合器般三种工作状态第种全联动第二种部分半联动第三种完全半联动状态汽车处起步时工作状态驾驶员踩离合器通踏板带动里面压盘压盘摩擦片分开压盘飞轮处分开状态样存摩擦力情况正常工作状态汽车离合器里压盘飞轮摩擦片紧密间连起时候摩擦力
图11 离合器结构图
1飞轮 2动盘 3压盘 4离合器盖 5分离杠杆
6浮动销 7调节拨叉 8分离拨叉 9拉杆
10离合器踏板 11分离轴承 12分离轴承座 13回位弹簧
14变速器第轴轴承盖 15离合器轴 16轴承
17压紧弹簧
2离合器结构形式选择
21离合器种类选择
离合器生产现广泛断进步着摩擦离合器现没遗弃相反种促离合器保持接合状态促离合器保持分离状态两部分组成保持接合部分动部分动部分压紧机构保持分离状态操机构动部件惯性非常锥形鼓形缺陷传动装置难换档组合够柔软容易卡
选择盘式摩擦离合器
22选离合器结构形式
正前面描述次毕业设计选离合器结构形式呼欲出压紧弹簧选膜片弹簧动盘考虑选择带扭转减震器
3选择离合器基参数
表31 吉利帝豪轿车整车参数
整备质量
滚动半径
减速器传动
变速器传动
高车速
发动机转矩
发动机功率
起步时发动机转速
1255kg
0303m
5614
4379
170kmh
140N2·m
80KW
2000rmin
31离合器备系数β
离合器设计时候考虑重参数—备系数β重性体现离合器工作时接收传递发动机扭矩否行参数保证离合器够安全执行身务时缓汽车起步时磨损样子更离合器时间延长离合器受损失降低选择β时注意点：
(1)离合器频繁会导致摩擦片受磨损种情况然需够传递发动机扭矩
(2)滑动磨损避免
(3)传动系超载会汽车带危险量避免
发现β果选取太够传递发动机扭矩控制滑动摩损增果β选取太离合器尺寸控制较操作起没灵敏β取值需情况定β取情况：①备动力足状况良发动机②气缸数发动机β取情况：①需考虑滑动磨损起步时②质量较③般柴油机工作时间较长转矩没汽油机稳需β
般说备系数β没固定取值会相应推荐值轿车β取值范围1213间国外推荐值12轿车β取值中符合面取值情况备动力足状况良次设计中吉利帝豪轿车离合器备系数β取值
32单位压力P0
作汽车重零部件摩擦必然存决定着耐磨性单位压力P0关重环节仅影响工作时优劣会影响离合器时间长短选取单位压力P0时候密切关注离合器项特征离合器工作状态
离合器汽车正常行驶时较频繁选取单位压力P0时应该取样减低摩擦片周围热负荷然果特殊状况合适范围增单位压力P0
材料摩擦片单位压力取值范围见表3—2
表3—2 单位压力取值范围
摩擦片材料
单位压力M
石绵基材料
编织
015—025
模压
025—035
粉末冶金材料
铜基
035—050
铁基
金属陶瓷材料
070—150
机材料
025—050
33摩擦片外径D径d厚度b
离合器结构基础形式取决摩擦片外径离合器终成正关系够决定离合器轻重会影响离合器时间需转矩少然尺寸更转矩摩擦片外径D需发动机转矩Temax
Temax发动机转矩
A够影响D条件
轿车选A＝47
发动机转矩Temax摩擦片外径D计算出需利摩擦片参数统表33国摩擦片项尺寸统参数
表33离合器摩擦片项参数
外径
160
180
200
220
240
280
300
320
340
径
110
125
140
150
155
165
175
190
195
厚度
32
35
35
35
35
35
35
35
4
CdD
0687
0694
0700
0682
0646
0589
0583
0594
0574
1c2
0676
0667
0657
0535
0583
0796
0802
0647
0671
c范围0507间摩擦片径外径值摩擦片厚度市场规定三规格分32毫米35毫米4毫米
34确定离合器基参数
次毕业设计选离合器基参数确定：
吉利帝豪发动机扭矩Temax140N2m根验公式： 初选摩擦片外径D轿车A47 173mm
通表32离合器摩擦片标准表次设计选摩擦片尺寸D＝180mmd＝125mmh＝35mmc＝0694mm
采状况良膜片弹簧离合器备系数β应该取值β＝12采机材料摩擦片单位压力P0范围025～05MPa
公式验算单位压力p0
f摩擦面间静摩擦数取f＝03
Z摩擦面数单片离合器Z＝2
P003Mpa终计算出单位压力P003MPa合适范围离合器项基参数计算出合理
4确定零件结构计算
41动盘设计
够较影响离合器性零部件动盘动盘没复杂零件组成仅仅动片摩擦片动盘毂动盘结构部位相显简易质量较需注意点：
(1)需更转动惯量减少齿轮间碰撞
(2)达摩擦损耗目必须轴弹性离合器接合畅易启动摩擦力均匀分布摩擦面
(3)缓离合器受碰撞应装备扭转减震器时防止传动系振
扭转减震器安装动盘里会发挥较作般轿车恰需种动盘减震弹簧动片花键毂连接重物件种特特性—切弹性样掉高频振动具减缓击力作减震摩擦片垫片夹起系统工作时消耗会太够降低低频振动根发动机传动系统特殊设计通项检验知道扭转减震器达佳减震降噪效果汽油发动机汽车常带线性弹性扭转减震器汽车更起减震降噪发动机空挡情况运转时候果柴油机旋转时候够均匀会离合器连接变速器齿轮啮合部分发生碰撞时候果想减轻齿轮间发生碰撞降低碰撞发生噪音采具非线性特性扭转减震器
离合器接合滑非常重时候具轴弹性动片成佳选择做呢？较实方法开T形槽动片然会产生弧形通工具变成波形呈弯曲波形形状相互分离弧形片安装摩擦片离合器接合时受压动片中弯曲波形逐渐扁增加动盘压力扭矩接合变更加畅种做法然说减动片弯曲利刚度方面难做相整体式弹性动片样状况
压成型刚度致波形弹簧压成型动片安装起分开式结构波形弹簧较薄钢板制成容易产生较转动惯量整体式结构方面相分开式结构会显较差点
离合器中动盘毂承受载荷功臣部分载荷动盘毂接收保证离合器部分司职承受载荷时断裂常采材料锻钢硬度较高承受载荷意味着位置较重花键变速器输入轴前端花键花键轴需动盘配合花键轴存偏差时没完全分离滑需动盘轴长度变更长镀铬表面硬度耐磨性提高动盘部零件消耗磨损降低花键孔考虑镀铬
减震弹簧弹簧钢丝60Si2MnA50CrVA65Mn等类型
选取花键结构时国际标准选取
表41 找出动盘毂花键参数
动盘外径Dmm
花键齿数
n
花键外径
D′mm
花键径
d′mm
齿厚
bmm
效齿长
lmm
挤压应力
σ
180
10
27
22
3
20
118
选择完花键必须进行强度校核花键常表面受压力太遭破坏花键普遍受损坏原花键必须计算挤压应力实现强度校核果挤压应力话增加花键毂长度
挤压应力 （41）
花键齿侧面压力
P花键齿侧面压力N
D′花键径m
d＇花键外径m
Z动盘毂数目
Temax发动机转矩N 2·m
n花键齿数
h花键齿工作高度m
l花键效长度m
24490MPa＜[σ挤压]30MPa
42离合器盖设计
离合器盖部分总成包含东西较顾名思义离合器盖外然压盘压紧弹簧传动片部分设计出
421设计离合器盖
离合器盖作壳体然起保护支撑作起两种作必须强刚度条件避免离合器盖工作中种条件产生变形时防止整离合器工作中离合器盖刚度足离合器分崩离析预防离合器讲转动效率摩擦片损耗引起零部件必磨损重点离合器盖必须搞飞轮间联系足够中防止影响离合器甚整轿车系统正常运行驾驶员带更麻烦离合器盖中没麻烦相反种方式定位销定位螺栓者止口中中未尝离合器汽车运行程中较频繁容易发热致温度升高带危险必须考虑冷效果离合器盖装风扇开窗口利离合器中散发产生热量
422压盘设计
压盘出台式键式销式等种连接方式会存缺口目前说传动片较种连接方式作仅结构简单许确定压盘位置起重作传动片身言点受力差异会带更加严重磨损优化种状况根正反特性差异圆周切线方安装分三组组四片组成传动片方式连接方式中佳选择
汽车刚开始工作时需发动机分配动力压盘作通完全压榨离合器片致发动机动力输入传动系统样汽车防冻剂动力驱动驾驶压盘飞轮相连起紧紧相连避免压力两零件连接稳散开
(1)压盘传动方式选择
传动片传动方式较选择作选择然着作离合器接合时候传动片带动压盘起转动作离合器分离时通特弹性带动压盘分离降低控制力传动片两边分固定离合器盖压盘样够发挥作传动片弹簧钢带材料做
传动片压盘中分布通圆周方切线方非常匀称散布压盘样仅压盘结构精练许装配没复杂更利整离合器工作中衡感
(2)确定压盘
灰铸铁HT200制成压盘开始先选择200mm120mm15mm分作压盘外径径厚度
完成选择压盘项参数必须做次温升校核利温升校核公式检验离合器接合时温度升应该5℃10℃间利增加压盘厚度应付温升超范围导致温升高
校核计算公式：
t—温升℃
W—滑磨功N2m
—分配压盘滑磨功占百分单片离合器压盘050
c—热容里指压盘热容铸铁制成铸铁压盘热容c54428J（kg2℃）
m压盘—压盘质量kg
压盘质量：
算压盘质量m压盘236kg接25kg取m压盘25kg
吉利帝豪轿车整备质量ma1255kg车轮滚动半径R0303m汽车刚开始启动时候发动机转速ne2000rmin减速器传动i05614变速器传动ig4379
滑磨功
温升
温升校核结果合适范围次压盘设计合理
(3)设计传动片传动片强度校核
开始先确定选择4组组4片传动片然选定传动片宽度b厚度h孔距l分15mm15 mm40mm传动片布置方式着切线反方布置圆周半径R选取80mm制成传动片材料弹性模量E2×105MPa样初步完成离合器压盘传动片设计参数确定
①计算传动片效长度l1：
l1l15d4015×5325mm
②计算传动片弯曲总刚度：
③通面计算结果推算出σmax驱动应力：
Ⅰ完全分开时 (43)
设计求f0Te0知道σ样值σ0
Ⅱ离合器压盘安装起知道Te0然推出fmax38mm
Ⅲ离合器分正驱反驱传递转矩时候离合器摩擦片损耗极致时候fmax会产生利尺寸链计算出fmax24mm
(Ⅰ)正驱(发动机车轮)：
(Ⅱ)反驱(车轮发动机)：
通计算结果发现正驱反驱两相反驱较危险选取数进行计算时候较谨慎传动片许应力选极限80号钢作传动片材料合适
④动盘没受磨损时候传动片弹性恢复力已出现时候时弹性弯曲变形量离合器接合时候f确定f087mm出K∑017MNm(传动片弯曲总刚度)f 087mm代入弹性恢复力公式求弹性恢复力：
认定行
43膜片弹簧设计
膜片弹簧设计较复杂必须通样品事先选膜片弹簧分析项参数计算项参数数值优胜劣汰制定出合适结构应该非常解关膜片弹簧种属性仅包括固特性知道结构失效样式重膜片弹簧项参数进行检验校核
膜片弹簧设计复杂时带类型弹簧更优异特点：
(1)膜片弹簧身带理想非线性特性种特性仅膜片弹簧离合器接合时保证身压力变整离合器程中传递扭矩力保持变离合器分离时身压力降缓踏板压力
(2)膜片弹簧压紧弹簧分离杠杆结合体样离合器结构方面部零件总体质量方面较缓空间
(3)压缩力高速转动时候没发生变化性方面较稳定
(4) 冷效果明显寿命长
(5) 规模生产降低制造成
膜片弹簧制造难题固特性—非线性弹性日常生产中容易失控断裂磨损频繁科学发展技提升机器进步材料提高膜片弹簧制造变容易许受限次设计选膜片弹簧离合器
431确定膜片弹簧重参数
(1)确定Hh值
Hh指高度截锥高度弹簧片高度厚度值反映膜片弹簧式样弹性特性样弹性特性曲线利值找出够表达佳效果特性曲线
图41载荷F变形λ间关系①Hh＜时载荷增速较快变形λ然増加确极缓慢速度增加着②Hh＝时清楚变形λ段相似速度增加期间载荷没发生变化③时条曲线里段载荷逐渐变变形λ然增加载荷变回增加状态恢复性种特性想找非常适合膜片弹簧④Hh＝时时载荷落差更⑤时载荷呈现负数形式合适
截锥高度H膜片弹簧厚度h值佳范围13次设计选Hh154
载荷F
变形λ
1
2
3
4
5
1 Hh< 2Hh 3
图41 膜片弹簧弹性特性曲线
(2)选择RRr值
弹簧必须定柔软性然会较误差影响汽车Rr值太佳范围115135间较误差影响左结果次设计Rr1255摩擦片均半径R>Rc取R80mm
(3)膜片弹簧起始圆锥底角α
9°15°汽车膜片弹簧起始圆锥底角α佳范围设计选α12°
(4)rarb选择
ra膜片弹簧端半径rb分离轴承作半径膜片弹簧端半径ra值应变速器第轴花键外径取ra16mmrb18mm
(5)分离指数目n切槽宽δ1δ2半径re
膜片弹簧选取n时候注意膜片弹簧结构较n选24结构较膜片弹簧n选12次设计分离指数目选择18δ1范围32～35mm间δ2范围9～10mm间re选择值(rre) ＞δ2规定次设计δ1选择32mmδ2选择10mm
(6)lL
l支承圈均半径L膜片弹簧压盘接触半径均半径lr相稍微r接触半径LR相稍微R次设计选取L78mml64mm
432载荷F1变形λ间关系
载荷F1变形λ间关系式：
(44)
设
(45)
已知E20×105MPaμ03数值代入
F111413F1＇
λ121λ1＇
λ1＇代入式中计算出数结果列表
表42 载荷F变形λ间关系
λ1＇
01
02
04
06
08
1026
12
14
16
1896
20
F1＇
0071
0131
0220
0274
0302
0310
0307
0299
0295
0312
0328
λ1
021
042
084
126
168
215
252
294
336
398
420
F1
812
1494
2508
3129
3443
3539
3499
3414
3367
3559
3743
F1—λ1特性曲线图42
0
500
1000
1500
2000
2500
3000
3500
4000
4500
5000
0
1
2
3
4
5
变形λmm
图42 膜片弹簧佳弹性特性曲线
433膜片弹簧工作位置确认
膜片弹簧进行工作位置确认首先做确认膜片弹簧端变形量膜片弹簧端变形量离合器工作时候选取特性曲线图膜片弹簧压紧力：
校核备系数：
离合器彻底分离时膜片弹簧端变形量
压盘行程
膜片弹簧端变形量离合器刚开始停止工作时压盘行程Δf＇1mm出
摩擦片磨损磨损量
434离合器完全停止工作分离轴承作载荷F2
变形λ1载荷F2关系式：
(46)
取
435分离轴承行程λ2
λ2＇端分离轴承处轴变形量λ1λ2＇关系式：取
宽度系数
分离轴承行程λ2力作分成λ2′λ2＇＇λ2＇＇指弯曲变形量
(47)
关数值代入式
436校核膜片弹簧
离合器完全变形膜片弹簧端变形量
(48)
代入相关数值
计算出数合适范围
437制造膜片弹簧工艺
现国生产膜片弹簧采精度较高钢板材料材料项参数达指标继续通热处理形式继续加工膜片弹簧抗压力想提高进行特殊强压处理增强膜片弹簧抗压承受力然膜片弹簧进行特措施增加膜片弹簧时间保护膜片弹簧种措施喷丸处理膜片弹簧耐磨性重镀铬端部进行高频感应加热淬火应时裂纹处进行挤压处理目消应力源
膜片弹簧表面必须高度精密够存必缺陷裂缝断痕万万够存
5结
次设计根吉利帝豪轿车原始设计参数出设计求通理解离合器工作状态注意事项根已知基公式计算出参数选择零部件材料选择种结构方案离合器中零部件结构选定出相关零部件材料次设计里部分设计迹循循序渐进进行次设计遵循着出设计求实事求计算项数利次计算出数画出相关零件CAD图水限知识储备够次设计难免出现漏洞希老师够加指导
参考文献
(1)徐石安江发潮 编汽年离合器汽车设计书 清华学出版
(2)马明星 编汽车设计课程设计指导书 中国电力出版社
(3)王林超 编汽年构造 中国水利水电出版社
(4)许洪国 编汽车理 民交通出版社
(6)王予 编汽车发动机原理 机械工业出版社
(7)钱川 编新型联轴器离合器选型设计制造工艺实手期北京工业学出版社
(8)骆素君朱诗 编 机械课程设计简明手册 机械工业出版社
(9)濮良贵纪名刚 编 机械设计 东北学出版社
附录
附录1
英文文献
Optimum Design of Automobile Diaphragm Spring Clutch
Zhao Lijun * Liu Tao ** and Song Baoyu ***
* Automotive Engineering Department Harbin Institute of Technology Weihai China Email lijun7422@263net
** Automotive Engineering Department Harbin Institute of Technology Weihai China
*** School of Mechanical and electrical Engineering Harbin Institute of Technology Harbin China
Abstract—In order to meet the design requirements and characters the optimum mathematical model of a light carry truck’s diaphragm spring clutch is established to decrease its size as far as the engine torque is reliably transferred The truck’s clutch design is optimized by MATLAB optimum toolbox The 3D model of the optimized results is established by ProE and the Finite Element Analysis is carried out It’s proved that the design period can be shortened the clutch size can be decreased and the production cost can be lowered
Keywords — Clutch Optimum design MATLAB Finite element
I I NTRODUCTION
As an important part of modern design theories and methods optimum design has become a new effective way for engineering design since 1960s when computer technique was introduced into engineering design It may shorten the design period improve the product performances and lower the production cost [1] [2] [3] [4] FMINCON provided by MATLAB optimum toolbox may effectively solve the problem of optimization Then the optimized results can be analyzed with a 3D model of PROE and a model of finite element analysis in order to prove the correctness of the optimized mathematical model and the rationality of the optimized results Currently optimum methods have been widely applied in automobile design [5] [6] [7] An ideal automobile clutch should transmit the engine torque reliably and save labor in operation In addition its size should be minimized and the cost should be lowered
II THE OPTIMUM MATHEMATICAL MODE L FOR BASIC CLUTH PARAMETERS
The basic clutch performance parameters mainly include the reserve factor β and the unit pressure p The main size parameters include the outside diameter of the friction plate D and the inside diameter d and the thickness of the friction plate b The variation of these parameters influences the structure size and working
performances of the clutch
A The Design of Variables
The reserve factor β can be calculated by formula (1)dependent on the clutch working pressure F the outside diameter of the friction plate D and the inside diameter d
(1)
Where Tc is the static friction moment (N·m) f is thestatic friction factor of friction surfaces F is the working pressure of the platen on the friction surface (N)R is the average friction diameter of the friction plates (mm) Z is the number of friction surfaces c is the ratio between the inside diameter and the outside diameter of the friction plate β is the reserve factor of the clutch
The unit pressure p can be calculated by formula (1) and it is also dependent on F D and d Thus the following variables are chosen for the optimum design of basic clutch parameters
(2)
B The Objective Function
The objective of the optimum design of basic clutch parameters is the minimum of the structure size as far as the clutch performance requirements are met The objective function is
(3)
C The Constraint Conditions
(1) The outside diameter of the friction plate D should confine the maximal circumference velocity D v within 65~70ms ie
(4)
Where nemax stands for the maximal rotate speed (rmin)
(2) The ratio between the inside diameter and the outsidediameter of the friction plate c should be within053~070 ie
(5)
(3) To ensure that the clutch can transfer the torque reliably and prevent the overload of transmitting systemβ should be chosen within 12~40 ie
(6)
(4) In order to install the torsion absorber the inside diameter of the friction plate d should be 50 mm greater than the absorber spring’s position diameter 2 R ie
(7)
(5) In order to reflect the capability of protection against overload while the clutch transmits torch the transferred torch of a unit area should be less than the permitted value ie
(8)
Where C0T is the transferred torch of a unit area of friction (Nmmm 2 ) [] C0 T is the permitted value of C0T (N·mmm 2 )
(6) In order to lower the heat load of the clutch friction and prevent the abrasion of the friction plate the unit pressure should be decided according to different types of the trucks and according to the materials of frictionwithin a maximal range between 010~150MPa ie
(9)
(7) In order to reduce the slippery friction when the automobile starts and prevent the burning of friction plate due to a toohigh surface temperature the friction work of a unit friction area for each joint should be less than the permitted value ie
(10)
where ω is the slippery friction work of a unit friction area (Jmm 2 ) ω is the permitted value of ω (Jmm 2 ) W is the total slippery friction work for one clutch joint when the automobile starts (J)
III THE OPTIMUM MATH MATICAL MODEL OF THE DIAPHRAGM SPRING
The optimum design of the diaphragm spring means deciding a group of fundamental parameters of the springmaking its torsion features with load meet the requirements of the clutch The spring intensity should also meet the design requirements so that the optimal integrating results can be reached [5]
A The Objective Function
The optimum design of the diaphragm spring tries to make the capabilities of the spring meet the requirements of the clutch (transfer the torch reliably and prevent the overload of transmitting system) This means that the working pressing force of the diaphragm spring should
not fluctuate too much with the abrasion of the friction plate The objective function of the optimum diaphragm spring design is the minimum average variation of the diaphragm spring’s working pressing force within the limit of utmost abrasion That is
(11)
Where P is the working pressing force of the new friction plate’s diaphragm spring (N)pi is the working pressing force of the diaphragm spring in the process of abrasion (N) N is the number of equal divisions within the limit of utmost abrasion
B The Design of Variables
The relation between the load of diaphragm spring F and its torsion λ is
(12)
where E is the modulus of elasticity (MPa) μ is the Poisson ratio H is the height of the plate spring’s inner frustum in free state (mm) h is the thickness of the spring armor plate (mm) R is the diameter of the plate spring at the main aspect in free state (mm) r is the diameter of the plate spring at the little end in free state (mm) R is the diameter of the platen at the load point (mm) r is the diameter of the supporting ring at the load point (mm)
From formula (12) it can be concluded that the six sizeparameters H h R r R and 1 r should be chosen as the variables for optimum design as well as the deformation of the main aspect Bλ in accordance with the spring’s working pressing force 1B F at the joint working point ie
(13)
C The Constraint Conditions
(1) To ensure that the clutch transfer the torch reliablythe working pressing force of the diaphragm spring F should be equal to the required working pressing force YF when the clutch rubs
(14)
(2) To ensure that the operating points A B C have their proper positions (A should be to the left of protruding point M B should be close to the inflexion point H C should be close to the concave point N as is shown by Figure 1) the position of 1B λ relative to1Hλ should be correctly chosen Usually 1B 1H 08 ~10 λ λ ie
(15)
(3) To ensure that the clutch can transfer the torch reliably after the friction plate abrades and considering the drop of friction factor the working pressing force of the spring after the friction plate abrades1AF should be greater than or equal to the pressing force of the new friction plate 1BF ie
(16)
(4) In order to meet the requirements of clutch performances h H of the spring and the initial base cone angle ( ) r R H − ≈ α should be within a certain range ie
(17)
(18)
(5) In order to make the friction plate’s working pressing force equally distributed the diameter of the platen at the load point1R should be between the average radius of the friction plate and the outer radius ie
(19)
(6) The separated finger of the diaphragm spring can separate the levers thus the lever ration should be chosen within a certain range ie
(20)
(7) The maximal point of the diaphragm spring pressing force σtBmax should not be greater than the permitted value ie
(21)
(8) To meet the requirements of fatigue intensity the maximal pressing force of point A during the process of spring working σtAmax should not be greater than the permitted value ie
(22)
IV THE INSTANCE OF CALCULATION
The optimum calculation of HB1030G light carry truck’s diaphragm spring clutch is carried out with FMINCON provided by MATLAB optimum toolbox [7]The optimized results are compared with the original parameters The optimized results of basic clutch parameters are shown in Table 1 and the optimized results of basic diaphragm spring parameters are shown in Table 2 The optimized results have been rounded By calculation it’s concluded that the round meets the constraint conditions Thus the optimized results can reduce the total size of the clutch
TABLE I THE OPTIMIZED RESULTS OF BASIC CLUTH PARAMETERS
The Parameters
Before Optimized
After
Optimized
Round
Working Pressure of the
Friction Plate (N)
Outside Diameter of the
FrictionPlate (mm)
Inside Diameter of the Friction Plate (mm)
11932
250
150
12988
2235
1338
12706
225
135
TABLE II THE OPTIMIZED RESULTS OF BASIC DIAPHRAGM SPRING PARAMETERS
The Parameters
Before
Optimized
After
Optimized
Round
Height of the Inside Frustum (mm)
Thickness of the Armor Plate (mm)
Diameter of the Plate Spring at the Main Aspect in Free State (mm)
Diameter of the Plate Spring at the Little End in Free State (mm)
Diameter of the Platen at the Load Point (mm)
Diameter of the Supporting Ring at the Load Point (mm)
Deformation at the Working Point (mm)
5
2
205
150
200
155
368
44
21
1786
1227
1753
1278
35
45
2
180
123
176
128
35
V THE FINITE ELEMENT ANALYSIS
The Finite Element Analysis (FEA) can analyze the structure intensity and dynamic features of the components As a modern design method it may overcome the problems of traditional design methods The efficiency and precision are greatly improved It may even solve the issues which traditional methods are at wit’s end A special advantage of FEA is that it can solve
the mechanical issues with arbitrary structures and boundary conditions It has been widely used in mechanical structure design and calculation and is proved to be a reliable theoretical calculating method With the prevalence and development of computers FEA has become a crucial part of CAD and CAE It is an indispensable tool for designing almost every machine and automobile
Based on the former optimized results a 3D model of the clutch is established with ProE and the finite element analysis is carried out When analyzing first of all the clutch cover and diaphragm spring are assembled together then all DOFs of the clutch cover are restricted at the face where the flywheel combines the clutch coverAt last the evenly distributed pressure is loaded on the clutch cover and the diaphragm spring respectively along the circumference With the modules of solid164 and the aptitude mesh the clutch cover is divided into 10287 units and 10012 nodes and the diaphragm spring is divided into 8849 units and 8512 nodes After analysis
the stress and strain of the clutch cover are shown in Figure 2 and the stress and strain of the diaphragm spring are shown in Figure 3 and their maximum stresses respectively are 73MPa and 1367Mpa The maximum reactive forces are smaller than the safe reactive force It can be proved that the optimum mathematical model is correct and the optimized results are reasonable
VI CONCLUSION
The optimum design and finite element analysis have been carried out in order to improve the design of clutch and increase the efficiency of design The optimum mathematical model of the clutch has been established according to the design requirements and characters so that the weight and size can be reduced as far as the intensity rigidity and life span of the spare parts are
ensured The optimum design is based on MATLAB toolbox A 3D model of the clutch has been established with ProE according to the optimized results and the finite element analysis has been carried out to prove the correctness of the mathematical model and the rationality of the optimized results It has been proved that the optimum design is significant for increasing the design efficiency reduce the size and weight and lower the production cost
REFERENCES
[1] He Xianzhong Li Ping Optimum Technology and Its ApplicationBeijing Institute of Technology Press 1995 pp 2528
[2] Liu Weixin Mechanical Optimum Design Tsinghua University Press 1997 pp 5056
[3] Jiang Dazhi Chi Guotai and Lin Jianhua Decisionmaking model for optimization of loan＇s portfolio based on efficient boundary Journal of Harbin Institute of Technology 5th ed 2002pp 614617
[4] Zhu Xuejun Wang Anlin and Zhang Huiqiao Robust Design for MechanicalStructural System by Using Non Stationary Penalty Function with GA Mechanical Science and Technology Vol 192000 pp 4951
[5] Xia Changgao Zhu Mangmao A Study on Mathematic Model of Optimum Design for Automotive Clutch Diaphragm SpringAutomobile Technology Vol 12 1996 pp 1215
[6] Cui Shengmin Yang Zhanchun Optimum Design of Automobile Wheels Based on Definite Element Analysis Machine DesignVol 9 2001 pp 4142
[7] Su Jinming A Guidance of MATLAB61 Practical ApplicationPublishing House of Electronics Industry2002 pp 4855
中文翻译
汽车膜片弹簧离合器优化设计
赵丽君刘涛宋宝玉
威海市哈尔滨工业学汽车工程部电子邮件 lijun7422@263 Net
威海哈尔滨工业学汽车工程系
哈尔滨哈尔滨工业学机电工程学院
摘满足轻型载货汽车膜片弹簧离合器设计求特点建立膜片弹簧离合器优化数学模型便传递发动机扭矩情况减膜片弹簧离合器体积 利 matlab 优化工具箱卡车离合器进行优化设计 利 pro e 软件建立优化结果三维模型进行限元分析分析 实践证明缩短设计周期减离合器尺寸降低生产成
I引
优化设计作现代设计理方法重组成部分20世纪60年代计算机技术引入工程设计已成工程设计种新效方法 缩短设计周期提高产品性降低生产成[1] [2] [3] [4] 利 matlab 优化工具箱提供 fmincon 算法效解决优化问题 然利 pro e 三维模型限元分析模型优化结果进行分析验证优化数学模型正确性优化结果合理性 目前优化方法已广泛应汽车设计[5][6][7] 理想汽车离合器应传递发动机扭矩节省操作力 外应量减少成应降低
II基线索参数优数学模型
离合器基性参数包括备系数单位压力 p 尺寸参数包括摩擦片外径 d摩擦片径 d 摩擦片厚度 b 参数变化影响着离合器结构尺寸工作性
A 变量设计
根离合器工作压力 f摩擦片外径 d 摩擦片径 d公式(1)计算备系数
(1)
中 tc 摩擦片静摩擦力矩(n · m) f 摩擦片静摩擦系数f 摩擦片摩擦片工作压力(n) r 摩擦片均摩擦直径(mm) z 摩擦片数量c 摩擦片径外径离合器备系数 基离合器参数优化设计中选择变量
(2)
B目标函数
基离合器参数优化设计目标满足离合器性求前提离合器结构尺寸 目标函数
(3)
C 约束条件
(1)摩擦片外径 d 应周速度 d v 限制65ー70m s
(4)
Nemax 代表转速(r min)
(2)摩擦片径外径 c 应053ー070间
(5)
(3)保证离合器传递扭矩防止传递系统载应选择12ー40范围
(6)
(4)安装扭转吸收器摩擦片 d 径应吸收器弹簧位置直径2r 50毫米
(7)
(5)反映离合器传递火炬时载保护力单位面积传递火炬应允许值
(8)
中 单位摩擦面积传递火炬(n m mm 2) (n m mm 2)允许值
(6)降低离合器摩擦热负荷防止摩擦片磨损应根汽车类型摩擦材料确定单位压力范围010ー150 mpa
(9)
(7)减少汽车起动时滑动摩擦防止摩擦片表面温度高烧毁接头单位摩擦面积摩擦功应允许值
(10)
中单位摩擦面积滑摩擦功(j mm2) 允许值(j mm2) w 汽车启动时离合器接头总滑摩擦功
III膜片弹簧佳数学模型
膜片弹簧优化设计指确定组弹簧基参数负荷扭转特性满足离合器求 弹簧强度应满足设计求便达佳综合效果[5]
A 目标函数
膜片弹簧优化设计弹簧性满足离合器求(传递火炬防止传递系统载) 意味着膜片弹簧工作压力会着摩擦片磨损产生太波动 膜片弹簧优化设计目标函数膜片弹簧工作压力极限磨损范围均变化量
(11)
中 p 新型摩擦板膜片弹簧工作压力(n) pi 膜片弹簧磨损程中工作压力(n) n 磨损极限等分数
B 变量设计
膜片弹簧 f 载荷扭矩关系
(12)
中 e 弹性模量(mpa) 泊松 h 板弹簧状态锥体高度(mm) h 弹簧装甲板厚度(mm) r 板弹簧状态部位直径(mm) r 板弹簧状态端部直径(mm) r 板弹簧加载点压板直径(mm) r 加载点压板直径(mm) r 加载点支撑环直径(mm)
公式(12)中出结弹簧接头工作点工作压力1b f选择六尺寸参数 h h r r 1r 作优化设计变量
(13)
C 约束条件
(1) 保证离合器传递火炬性膜片弹簧工作压紧力 f 应等离合器摩擦时需工作压紧力 f
(14)
(2)确保操作点 a b c 适位置(a 应该凸点 m 左边b 应该接拐点 hc 应该接凹点 n图1示) 应该正确选择1b 相 to1h 位置 通常1b1h 08 ~ 10
(15)
(3)确保离合器摩擦片磨损传递摩擦炬考虑摩擦数降低摩擦片磨损弹簧工作压力应等新摩擦片磨损弹簧压力
(16)
(4)满足离合器性求弹簧 h h 初始基座锥角() r h≈ 应定范围
(17)
(18)
(5)摩擦板工作压力分布均匀加载点1r 处压板直径应该摩擦板均半径外半径间
(19)
(6)膜片弹簧分离指分离杠杆杠杆应选择定范围
(20)
(7)膜片弹簧压紧力点应允许值
(21)
(8)满足疲劳强度求弹簧加工 tamax 程中 a 点压力应允许值
(22)
IV计算实例
利 matlab 优化工具箱提供 fmincon hb1030g 轻型载货汽车膜片弹簧离合器进行优化计算[7] 优化结果原参数进行较 基离合器参数优化结果见表1基膜片弹簧参数优化结果见表2 优化结果已圆满完成 通计算出该轮满足约束条件 优化结果减离合器总体尺寸
表基离散参数优化结果
参数
优化前
优化
工作点变形(毫米)
摩擦片工作压力(n)
摩擦片外径(毫米)
摩擦板径(毫米)
11932
250
150
12988
2235
1338
12706
225
135
表二膜片弹簧基参数优化结果
参数
优化前
优化
工作点变形(毫米)
锥体高度(毫米)
钢板厚度(毫米)
板弹簧状态方面直径(毫米)
板弹簧端状态直径(毫米)
压板加载点直径(毫米)
支撑环负载点直径(毫米)
工作点变形(毫米)
5
2
205
150
200
155
368
44
21
1786
1227
1753
1278
35
45
2
180
123
176
128
35
V限元分析
限元分析分析构件结构强度动态特性 作种现代设计方法克服传统设计方法存问题 提高效率精度 甚解决传统方法法解决问题 限元法特殊优点解决意结构边界条件力学问题 机械结构设计计算中广泛应种理计算方法 着计算机技术普发展限元分析已成计算机辅助设计计算机辅助工程重组成部分
优化结果基础利 pro e 软件建立离合器三维模型进行离合器限元分析优化设计 分析时首先离合器盖膜片弹簧装配起然离合器盖度限制飞轮离合器盖结合面压力均匀分布离合器盖膜片弹簧分圆周加载 采 solid164模块 aptitude 啮合离合器盖分10287单元10012节点膜片弹簧分8849单元8512节点 分析离合器盖应力应变图2示膜片弹簧应力应变图3示应力分73mpa 1367mpa 反作力安全反作力 证明优化数学模型正确性优化结果合理性
VI结
改进离合器设计提高设计效率进行优化设计限元分析 根离合器设计求特点建立离合器优化数学模型保证零件强度刚度寿命前提减离合器重量尺寸 基 matlab 工具箱进行优化设计 根优化结果 proe 软件建立离合器三维模型进行限元分析试验验证数学模型正确性优化结果合理性 实践证明优化设计提高设计效率减尺寸重量降低生产成具重意义
参考文献
[1]贤忠李佳技术应 北京理工学出版社1995年第2528页
[2]刘伟新机械优化设计 清华学出版社1997年版第5056页
[3]江志池国泰林建华基效边界贷款组合优化决策模型 哈尔滨工业学 2002年第5版第614617页
[4]朱学军王安林张慧乔利 ga 非稳惩罚函数机械 结构系统进行稳健设计 机械科学技术 第卷 192000pp 4951
[5]夏昌高朱满茂汽车离合器膜片弹簧优化设计数学模型研究 汽车工艺 第卷 121996pp 1215
[6]崔胜民杨占春基限元分析汽车车轮优化设计 机械设计 第卷 92001pp 4142
[7]苏金明实际应指南 电子工业出版社2002年第4855页
附录2
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