3491 四速电动葫芦机械系统的设计
3491 四速电动葫芦机械系统的设计,电动葫芦,机械,系统,设计
1河南理工大学万方科技学院本科毕业设计(论文)中期检查表指导教师: 王 志 阳 职称: 副 教 授 所在院(系):机械与动力工程学院 教研室(研究室):机械设计教研室题 目 四速电动葫芦机械系统的设计学生姓名 王彦婷 专业班级 机设 10 升(一)班 学号 0816101040一、选题质量:(主要从以下四个方面填写:1、选题是否符合专业培养目标,能否体现综合训练要求;2、题目难易程度;3、题目工作量;4、题目与生产、科研、经济、社会、文化及实验室建设等实际的结合程度)选题符合机械设计专业的培养目标,能够体现综合训练的要求。设计任务难易程度和工作量适中。该毕业设计可以训练机械专业学生按要求进行计算机绘图和手工绘图能力,训练学生方案设计、结构设计和工艺设计的能力。对以往所学知识进行总结和应用,能够满足综合训练的要求。所选题目与实际贴合比较紧密,在实际的应用中比较广泛。二、开题报告完成情况:按时按要求完成了开题报告。2三、阶段性成果:1通过各种渠道查阅了相关资料,并提交了开题报告和毕业实习报告。2掌握了电动葫芦的理论基础,并初步确定了四速电动葫芦的研究方案。3完成了一篇与本课题有关的文献翻译。四、存在主要问题:1对电动葫芦的减速器设计中齿轮的选择不太合理,技术方面还有一些问题需要解决;决;2说明书有文字错误,通过目前设计知此电动葫芦丰富的功能只能通过机械装置完成 成;3局部结构设计思路不是很清晰;五、指导教师对学生在毕业实习中,劳动、学习纪律及毕业设计(论文)进展等方面的评语指导教师: (签名)年 月 日四速电动葫芦机械系统的设计目录前言 .11 电动葫芦简介 .21.1 电动葫芦的原理 .21.2 发展前景 .31.3 电动葫芦在使用时应该注意的事项 .41.4 设计要求 .52 四速电动葫芦的结构分析与设计 .62.1 电动葫芦的结构分析 .62.2 电动葫芦的设计方案 .63 电动葫芦起升机构部件的设计 .83.1 起升机构的工作分析 .83.2 电动机的选择 .93.3 滑轮组的选择 .93.4 钢丝绳的选择和校核 .103.4.1 钢丝绳的选择 .103.4.2 计算钢丝绳所承受的最大静拉力 .113.4.3 计算钢丝绳破断拉力 .113.5 吊钩的设计 .113.5.1 吊钩的选择 .123.5.2 吊钩的尺寸设计 .123.6 卷筒装置的设计 .143.6.1 卷筒直径的确定 .153.6.2 卷筒长度的确定 .153.6.3 卷筒厚度的计算 .164 同轴式三级齿轮传动减速器的设计 .174.1 确定传动装置的总传动比和分配转动比 .174.2 计算各轴的转速和转矩和功率 .174.3 传动零件的设计计算 .194.3.1 第一轴齿轮的设计计算 .194.3.2 第二轴齿轮的设计计算 .254.3.3 第三轴齿轮的设计计算 .294.4 轴的设计 .354.4.1 第一根轴的设计计算 .354.4.2 初步估算轴的最小直径 .364.4.3 第二根轴的设计计算 .394.4.3 第三根轴的设计计算 .425 轴的校核 .455.1 第一根轴的校核 .455.1.1 求支反力 .455.1.2 求弯矩 .455.2 第三根轴的校核 .475.2.1 求支反力 .475.2.2 求弯矩 .475.3 中间轴的校核 .495.3.1 求支反力 .495.3.2 求弯距 .495.3.3 总弯距的计算 .506 轴承的校核 .526.1 计算轴承的支撑反力 .526.2 轴承的当量动载荷 .536.3 轴承的寿命 .537 减速器箱体结构的设计 .548 减速器润滑密封设计 .579 运行机构外壳的选择 .589.1 行走机构电动机及车轮的选取 .589.2 行走机构减速比的确定 .5810 结论 .6011 致谢 .61参考文献 .62前言起重机械广泛应用于各种物料的起重、运输、装卸等作业中,可以减轻劳动强度,提高生产效率。如在工厂、矿山、车站、港口、建筑工地、水电站、仓库等生产部门中得到应用。随着我国经济改革的不断深入,一些老的工业基地逐渐复苏,大量冶炼、铸造和机加工行业出现增长势头,引发市场对起重机械需求量的不断增加。有关调查资料表明,65%的起重机械用户主要是为了提高生产率、减少劳动工资、降低职工劳动强度。因而用户对起重机械的安全性、先进性、适用性和自动化程度就提出了更高的要求,使起重机械的制造厂家面临更加严峻的挑战。起重机械制造行业的发展趋势为设计、制作的计算机化、自动化近年来,随着电子计算机的广泛应用,许多国外起重机制造商从应用计算机辅助设计系统,提高到应用计算机进行起重机的模块化设计。起重机采用模块单元化设计,不仅是一种设计方法的改革,而且将影响起重机行业的技术、生产和管理水平,老产品的更新换代,新产品的研制速度都将大大加快。对起重机的改进,只需针对几个需要修改的模块;设计新的起重机只需选用不同的模块重新进行组合,提高通用化程度,可使单件小批量的产品改换成相对大批量的模块生产。亦能以较少的模块形式,组合成不同功能和不同规格的起重机,满足了市场的需求,提高了竞争能力。作为起重设备中轻便灵活的电动葫芦作业范围是有点、线为主、自重轻、构造紧凑、体积小、维修方便、经久耐用等特点。目前起重设备较多,如单、双梁桥式起重机、门式起重机等,但结构体积庞大,一次性投资与运行成本较高,就是不能良好的满足生产现场的要求,急需技术经济性能价格良好的起重设备,电动葫芦在此方面具有优势,但目前电动葫芦多以为单速、双速为主,多速电动葫芦极少,特别是四速电动葫芦。作为起重基地的新乡,研究开发四速电动葫芦,是很有前景的。1 电动葫芦简介1.1 电动葫芦的原理电动葫芦,简称电葫芦。又名电动提升机。它保留了手拉葫芦轻巧方便的特点,由电动机、传动机构和卷筒或链轮组成,分钢丝绳电动葫芦和环链电动葫芦两种。通常用自带制动器的鼠笼型锥形转子电动机(或另配电磁制动器的圆柱形转子电动机)驱动,起重量一般为 0.180 吨,起升高度为 330 米。多数电动葫芦由人用按纽在地面跟随操纵,也可在司机室内操纵或采用有线(无线)远距离控制。电动葫芦除可单独使用外,还可同手动、链动或电动小车装配在一起,悬挂在建筑物的顶棚或起重机的梁上使用。手拉葫芦和手扳葫芦都叫做手动葫芦,是用人力来提升重物的。电动葫芦是一种用途十分广泛的轻小型起重设备。其特点是体积小,重量轻,承载能力大,常被安装在电动单梁桥门起重机和悬挂式起重机上 ,用来升降和运移物品。 电动葫芦的各类较多电动葫芦主要有钢丝绳电动葫芦,环链电动葫芦,微型电动葫芦和防爆电动葫芦几种型号。电动葫芦又改进了手拉葫芦人工操作、提升速度慢等不足,它集电动葫芦和手拉葫芦的优点于一身。采用盘式制动电机作用力,行星减速器减速,具有结构紧凑、体积小、重量轻、效率高、使用方便,制动可靠维护简单等特点。适用于低速小行程的、物料装卸、设备安装、矿山及工程建筑等方面,价廉物美,安全可靠,为您的工作带来便利。本设计是钢丝绳电动葫芦,因为钢丝绳电动葫芦有它特有的特点。下面就来和大家看一看钢丝绳电动葫芦的结构原理。减速器:采用三级定轴斜齿轮转动机构,齿轮和齿轮轴用经过热处理的合金钢制成,箱体,箱盖由优质铸铁制成,装配严密,密封良好。减速器自成一个部件,装卸极为方便。控制箱:采用能在紧急情况下切断主电路,并带有上下行程保护断火限位器的装置。确保了电动葫芦的安全运行。电器元件寿命长,使用可靠。钢丝绳:采用 GB1102-74(6*37+1)X 型起重钢丝绳,它保证了经久耐用。锥行电动机:起升电机采用较大起动力矩锥形转子制动异步电动机,无须外加制动器。电机负载持续率为 25%,电机采用 B 级或 F 级绝缘,电机防护等级 IP44/IP54。按钮开关:手操作轻巧灵便,分有绳操纵和无线遥控两种方式.钢丝绳电动葫芦的结构原理就决定它的优点,在市场上也有很好的反映。从深层次了解钢丝绳电动葫芦,可以让你在它的维护保养中做得更好,也更能让钢丝绳电动葫芦在工作中发挥更大的作用。提高它的工作效率,也就提高了相对的收入。1.2 发展前景目前,国内外电动葫芦产品在构造特征、性能配置等方面仍存在一定差异,通过对国内外该类产品的比较,明示了其差异情况。1964 年联合设计的 CDMD 葫芦,在 1975 年设计改进之后,虽经各制造企业不同程度的改进,并未吸收世界进程中的任何技术发展。包括 1983 年引进德国 Stahl 公司的 AS 钢丝绳葫芦,距离当代发达国家的产品水平,仍有数十年差距。而随着科技的不断发展与进步新一代多速电动葫芦有着跟多的发展趋势:向大型化、高效化、无保养化合节能化发展。向智能化、集成化合信息化发展。向成套化、系统化、综合化和规模化发展。向模块化、组合化、系列化和通用化发展。向小型化、轻型化、简易化和多样化发展。所以,新型电动葫芦的开发研究对于我国的起重行业还是很有实际意义的。而这个设计题目这样不但可以是我们和社会科技环境接轨。虽然我们的水平有限,但是可以借此更加全面的了解起重器材的性能和工作环境,为将来的起重行业的工作做一个铺垫。同时可以把以前学过的知识巩固一下,把以往不太注意的基础知识更加熟悉起来,为以后的工作打下坚实的基础。所以,在设计中,我们应该采用新理论、新方法、新技术和新手段来提高我们的的设计质量。电动葫芦种类一般分为几种:环链电动葫芦,钢丝绳电动葫芦,防爆电动葫芦,气动葫芦,微型电动葫芦,舞台专用电动葫芦,还有台湾进口的小金刚。按照能否运行来分,又分为固定式与运行式两种,按照起升速度来分,分为单速与双速两种。电动葫芦使用非常简单,有操作手柄,运行式电动葫芦手柄上一般有上下左右四个按扭,固定式有上下两个按扭,特殊的也有其他设置,根据您的需要从下面调节手柄即可操作电动葫芦。一般电动葫芦都配有说明书,按照说明书上来按装即可。1.3 电动葫芦在使用时应该注意的事项(1)电动葫芦在使用前,应进行静负荷和动负荷试验。 (2)检查电动葫芦制动器的制动片上是否粘有油污,各触点均不能涂润滑油或用锉刀挫平。 (3)严禁超负荷使用。不允许倾斜起吊或作为拖拉工具使用。 (4)操作人员操作时,应随时注意并及时消除钢丝绳在卷筒上脱槽或绕有两层的不正常情况。 (5)电动葫芦盘式制动器要用弹簧调整至是物件能容易处于悬空状态,其制动距离在最大负荷时不得超过 80mm。 (6)电动葫芦应有足够的润滑油,并保持干净。 (7)电动葫芦不工作时,禁止把重物悬于空中,以防零件产生永久变形1.4 设计要求本设计的四速电动葫芦机械系统的根据现有普通电动葫芦的应用情况提出要求是:(1) 四速电动葫芦的最大载重为 10 吨,最大起升高度为 9 米。(2) 四速电动葫芦的强度等级为 M,工作级别为 M5。(3) 通过电机的变速实现在一个电机带动下输出 4 种速度。2 四速电动葫芦的结构分析与设计2.1 电动葫芦的结构分析电动葫芦主要由起升机构和运行机构组成。起升机构包括吊钩、钢丝绳、滑轮组、电机、卷筒和减速器组成;它的运行机构为小车。电动机的总体结构如图 2-1 所示 图 2-1 电动葫芦总体结构简图电葫芦中间是钢丝绳卷筒,用小车将其悬挂于工字钢锻造的天车大梁上,一端用法兰固定一台可制动的锥形转子电动机,用传动轴将动力传递到另一端的减速机。经过减速的动力传递给钢丝绳卷筒,带动吊钩起重。2.2 电动葫芦的设计方案电动葫芦起升的结构主要为电动机、减速器和卷筒装置 3 个部件。排列方式主要有平行轴和同轴两种方式排列形式,如图 2-2 所示 a b图 2-2 起升机构部件排列图1 电动机 2 减速器 3 卷筒装置经过分析这里优先选用 b 方案,其方案的电机、减速器、卷筒布置较为合理。减速器中的大齿轮和卷筒连在一起,起吊产生的转距经大齿轮可以直接传给卷筒,使得卷筒只受弯距而不受扭距。其优点是结构紧凑,传动稳定,安全系数高。减速器用斜齿轮传动,载荷方向不变和齿轮传动的脉动循环,对电动机产生一个除弹簧制动的轴向力以外的载荷制动轴向力。当斜齿轮倾斜角一定时,轴向的力大小与载荷成正比,起吊载荷越大,该轴向力也越大,产生的制动力矩也越大;反之亦然。它可以减小制动弹簧的轴受力,制动瞬间产生的冲击减小,电动机轴受扭转的冲击也将减小,尤其表现在起吊轻载荷时,从而提高了电动机轴的安全性。因此,选择 b 方案。a 方案中结构电机与卷筒布置不再同一平面上通过减速器相连,使得减速器转距增大。3 电动葫芦起升机构部件的设计电动葫芦式起升机构用来实现物料垂直升降,是任何起重机不可缺少的部分,因而也是起重机最主要、也是最基本的机构。起升机构的安全状态得好坏将直接地关系到起重作业的安全,是防止起重事故的关键。电动葫芦的机构主要包括:起升用锥形转子制动电动机、减速器、卷筒装置和吊钩装置等 4 个动力和传动部件。起升电机、减速器、和卷筒装置构筑成一个革命性紧凑又坚固的结构,使起重机能更有效的利用厂房空间,增加了起升高度。平稳安静的运行延长起升机构的寿命。起升电机处于大直径卷筒内使电动葫芦具有较小的外形尺寸且起升电机具有良好的冷却性能。所有起升电机都装有盘式直流电磁制动器,自动监控间隙。电器和制动器和谐工作保证吊钩任何时候都不打滑。制动器为长闭设计防止失电事故,制动摩擦片不含石棉。卷筒由高强度无缝钢管制成,两端轴承支撑,钢丝绳由压板固定。卷筒最少有 2 圈安全绳槽,标准钢丝绳为刚强度钢丝制成并镀锌,重级制导绳器由耐磨的球墨铸铁制成,防止乱绳。大直径滑轮由球磨铸铁制成,防止跳绳。3.1 起升机构的工作分析电动机通过联轴器与减速器的中间轴连接,而中间轴又通过齿轮连接与减速器的高速轴相连,用减速器的低速轴带动卷筒,吊钩等钩取物装置与并卷绕在卷筒上的钢丝绳滑轮组连接起来。当电动机正反两个方向的运动经联轴器和减速器传递给卷筒时,通过卷筒不同方向的旋转将钢丝绳卷入或放出,从而使吊钩与吊挂在其上的物料实现升降运动,这样,就将电动机输出的旋转运动转化为吊钩的垂直上下的直线运动。用常闭式制动器空竹起重机机构的运转。通电时松闸,使机构运转;在失电情况下制动,使吊钩连同货物停止升降,并在指定位置上保持静止状态,当于与自锁。当滑轮组升到最高极限位置时,上升极限位置限制器被触碰面动作,使吊钩停止上升,即起到了限位开关的作用。当吊载接近额定起重量时,起重量限制器及时检测出来,并给予显示,同时发出警示信号,一旦超过额定值及时切断电源,使起升机构停止运行,以此来保证生产安全。3.2 电动机的选择本次设计为 10 吨四速电动葫芦,电动机采用 YZR 系列起重用三相一步电动机用电动机。由公式得:P=FV/1000=GV/1000=10000(4/60)/1000=0.67kW (3-1)滚筒传动的效率取:0.96联轴器的效率取: 0.99电机轴的效率取: 0.98(3-2) 与 电 机 与 与 输 出 轴 与 筒 与 输 出 轴总 =0.96(0.990.99)(0.990.99)(0.990.99)0.98=0.8857电动机功率:= / =0.67/0.8857=0.75266kW (3-3)dpw总由于钢丝绳电葫芦起吊和停止时有一些冲击,根据冲击程度一般使用系数 =1.4 故 1.4 =1.0537kWAkpd电机转速取:n 电 =930r/min故选电动机的电动的额定功率为 8.5kw,转速为 930/min3.3 滑轮组的选择滑轮组由动滑轮和定滑轮组成,其上缠绕钢丝绳,此方法可以减小起重所须的力还可以达到增速的目的。其中通过滑轮可以改变钢丝绳的运动方向。平衡滑轮还可以均衡张力。四速电动葫芦选用的滑轮组倍率由 1查得 m2。滑轮组效率 0.99z3.4 钢丝绳的选择和校核本次设计选用的钢丝绳主要依据其工作环境及工作强度及使用特点及重要性选用。柔韧性好、钢丝绳强度高、耐冲击、安全可靠。虽然在正常情况下使用的钢丝绳不会发生突然破断,但是钢丝绳广泛应用在起重机上,可能会因为承受的载荷超过其极限载荷而破坏。而钢丝绳的破坏是有前兆的,总是从断丝开始,极少发生整条绳的突然断裂。钢丝绳的破坏会导致严重的后果,所以钢丝绳既是起重机械的重要零件之一,也是保证起重作业安全的关键环节。 3.4.1 钢丝绳的选择钢丝绳是起重机械中最常用的构件之一,根据其本身的结构特点及工作环境的需要选择。查得钢丝绳型号选为 6X37-15-1550-I-右。(1)根据设计要求起重重量为 10 吨,按照构造易紧凑的原则,选用滑轮倍数为:a=2F=58800NK:安全系数;取 1.2(2)钢丝绳的直径 dd=21mmC=0.898为选择系数查得钢丝绳型号选为 6X37-15-1550-I-右。3.4.2 计算钢丝绳所承受的最大静拉力钢丝绳所承受的最大静拉力(即钢丝绳分支的最大静拉力)为:(3-4)kQZmPSax式中: -额定起升载荷,指所有起升质量的重力,包括允许起升的QP最大有效物品、取物装置(如下滑轮组吊钩、吊梁、抓斗、容器、起重电磁铁等) 、悬挂挠性件以及其 他在升降中的设备的质量的重力;Z-绕上卷筒的钢丝绳分支数,单联滑轮组 Z=1,双联滑轮组Z=2,根据要求 Z=1;m-滑轮组倍率;-滑轮组的机械效率。h其中 810000N ,m2, 0.99QPh所以 29.7Naxs3.4.3 计算钢丝绳破断拉力计算钢丝绳破断拉力为:=n (3-5)psmax式中:n-安全系数,根据机构工作级别查表确定,n5;=150 =136psax所以钢丝绳满足要求。3.5 吊钩的设计吊钩在起重装置中属于取物装置,用于提取物料。既是起重机械的重要零件之一,也是保证起重作业安全的关键环节3.5.1 吊钩的选择吊钩按形状分为单钩和双钩,按制造方法分为锻造吊钩叠片吊钩。单钩制造简单、使用方便,但受力情况不好。大多用在起重量为 80吨以下的场合;起重量大时常采用受力对称的双钩。叠片式吊钩由数片切割成形的钢板铆接而成,个别板材出现裂纹时整个吊钩不会破坏,安全性较好,单自重较大,大多用在大起重量或吊运钢水盛桶的起重机上。吊钩在作业过程中常受冲击,需采用韧性好的优质碳素钢制造。吊钩分类极广,一般包括:卸扣、吊环、圆环、梨形环、长吊环、组哈吊环,S 钩、鼻吊钩、美式吊钩、羊角吊钩、眼形滑钩、带保险卡吊环螺钉、链条卸扣,居于独特、新颖、质优安全的特点,适用于工厂、矿山、石油、化工及船舶码头等。确保安全,质量安全系数高,静载荷达到 3 倍,起重量从 5 吨到 150 吨。吊钩是起重机械常见的一种吊具,吊钩常借助滑轮组等部件悬挂在起重机构的钢丝绳上,还适用于工厂、矿山、石油、化工和船舶码头等吊运重物的场所。锻造吊钩分为单钩和双钩。单钩一般用于小起重量,双钩多用于较大的起重量。锻造吊钩材料采用优质低碳镇静钢或低碳合金钢,如 20 优质低碳钢、16Mn、20MnSi、36MnSi。这次设计的是 5 吨的葫芦,属于小起重量,结合电葫芦的生产现状,选用锻造单钩。3.5.2 吊钩的尺寸设计吊钩钩孔直径与起重能力有一定关系:单钩: tQD350钩身各部分尺寸(见图 3)间的关系如下:hLS75.021(2.)/05Lh图 3-1 锻造单钩计算得: D=80 S=60 H=96 =184 =482L23-2 吊钩的三维效果图由于负载属于轻型因此吊钩的各部位直径选择按照起重设计手册的常规数据选取完全可以满足工作要求,但注意的是吊钩的前端尖嘴部分应有一定的扬角避免磨损后起吊容易脱钩。在参考常规设计的基础上进行设计的已满足设计要求,故在次不与校核。3.6 卷筒装置的设计卷筒是用来卷绕钢丝绳的部件,它承载了起升载荷,收放钢丝绳,实现勾取物装置的升降,是实现四速电动葫芦机械系统满足要求的装置。(1)电动葫芦卷筒的种类电动葫芦按卷筒的筒体形状,可分为长轴卷筒和短筒卷筒;按制造方式,可分为铸造卷筒和焊接卷筒;按卷筒表面是否有绳槽,可分为光面卷筒和螺旋槽面卷筒;按钢丝绳在卷筒表面卷绕层数,可分为单层缠绕卷筒和多层缠绕卷筒,多层缠绕卷筒用于起升高度特大,或要求机构紧凑的起重机上。(2)电动葫芦卷筒的结构电动葫芦的卷筒是由筒体、连接盘、卷筒轴以及轴承支架等组成。单层缠绕的卷筒的筒体表面切有弧形螺旋槽,以增大钢丝绳与筒体的接触面积,避免相邻绳之间摩擦,并使钢丝绳在卷筒上缠绕位置固定。其缺点是筒体体积较大。多层缠绕卷筒的筒体表面直接采用光面,筒体两端有凸缘,以防止钢丝绳滑出。其缺点是钢丝绳排列紧密产生摩擦,各层互相叠加,对钢丝绳的寿命影响很大。电动葫芦的卷筒结构尺寸中,影响钢丝绳寿命的关键尺寸是按钢丝绳中心算起的计算直径,卷筒允许的最小卷绕直径必须满足所在机构工作级别所要求的规定值。3.6.1 卷筒直径的确定卷筒的直径式卷筒集合尺寸中最关键的尺寸,其名义直径 D 是指光面卷筒的卷筒外包直径尺寸,有槽卷筒取槽底直径,大小按下式确定。mm (3-6)28917)()1(dhD式中 D-按钢丝绳中心计算的最小卷筒直径,mmh-与机构工作级别和钢丝绳有关的系数,查表为 18d-钢丝绳的直径,mm计算得 289mm,取 290mmminD3.6.2 卷筒长度的确定由表查得卷筒几何尺寸的计算: (3-7)210L(3-8)PZDmH)(1ax式中 L-卷筒长度;-卷筒上螺旋绳槽部分的长度;0L-无绳槽卷筒端部尺寸的长度,由结构需要决定;1-卷筒两端多余部分的长度 ;2 23LPP-绳槽节距;-最大起升高度;maxHm-滑轮组倍率;-卷筒的计算直径。1D其中 720mm , 83mm, 32mm,L835mm0L1L23.6.3 卷筒厚度的计算对于铸钢卷筒, 式中 -卷筒壁厚;d-钢丝绳直径。所以 15mm4 同轴式三级齿轮传动减速器的设计电动葫芦减速器是起升机构中传动的重要组成部分,也是本次设计的重要部分,所以单独进行计算。其传动关系如图 4-1 所示a b图 4-1 同轴式三级传动减速器示意图4.1 确定传动装置的总传动比和分配转动比(1) 总传动比: = = (4-1)ainm3.651980(2)分配减速器的各级传动比:按同轴式布置。由 2表 15-1-3 三级圆柱齿轮减速器分配传动比。 由图查的 =5.7, =3.6。1i2i则低速级传动比 = = 3.23i5.26814.2 计算各轴的转速和转矩和功率(1) 各轴转速: n =n =nm =980 rinn = 1980r7.min5.min 2r4.i3.6in 275r1.92in.in =n(2)各轴输入转矩: 40.8295.095mddpNMT =Td 01.1.7d联 轴 器T 21.6 花 键T 21.48.75.09420.85i齿 轮 滚 动 轴 承T = 34323.7391NM齿 轮 滚 动 轴 承T =T .5.12.ii 齿 轮 滚 动 轴 承T = 6908.64卷 筒 滚 动 轴 承(3)各轴入输功率: .5kWdPP =Pd Pd.0180.92.43k联 轴 器P =P . P =2、681P =P P3、.8.109.87.kP =P P23、.7.4W P =P P34、.7.4098.7.1kWP =P P34、.15.6.各轴的运动和动力参数如表 4-1 所示表 4-1 运动和动力参数4.3 传动零件的设计计算4.3.1 第一轴齿轮的设计计算(1)选择齿轮材料,查表选小择齿轮材料为 40cr,调质和表面淬火处理或氮化 4855 HRC。(2)按齿面接触疲劳强度设计选择齿数取 z 1=13, z 2=i1 z1=5.7 13=74齿宽系数 由表 14-1-79,选 =0.8dd初选螺旋角 =14轴号 输出功率P(kW)转速 n(r/min) 输出转矩 T/(Nm)轴 8.432 980 81.74轴 8.11 980 78.47轴 7.78 171.9 420.85轴 7.47 47.75 395.51V 轴 7.18 14.92 1190.83VI 轴 6.68 14.92 1131.14初选载荷系数 按齿轮非对称布置速度中等冲击载荷不大来选择Kt=1.3转距 T 18.401.NmT弹性系数 ZE 由表查的 Z E=187.7 MPa确定变位系数 z1=12 z2=68 a=20 h*an=h*acos 由图查的x1=0.39x2=-0.38节点区域系数 ZH X =0 =8 查图得 Z H=2.43重合度系数 Z纵向重合度 35.014tan8.013.tan138.0sin zmbd端面重合度 42.9.1nxz由机械设计手册图 14-1-7 查的重合度则 78.01a87.02a 61.870)39.1(78.0)39.1()()(11 nnxx由图 14-1-19 查得 5.螺旋角系数 .4cos许用接触应力 接触疲劳极限 由机械设计手册图 14-1-24 查得大小齿轮的接触lim疲劳极限为 Hlim1= Hlim2=1160MPa应力循环次数 N 1=60 n1 Lh=60 980 1 6300=3.70 108N2= (4-2)71049.67.53i接触疲劳寿命系数由机械设计手册图 6.4-10 查得KHN1=1.08 KHN2=1.14计算接触疲劳许用应力取失效概率为 1 安全系数 S1 1lim.0816253MPaHN 2= SK1li.4则 1253128PaH(3)计算小齿轮分度圆直径 d1t小齿轮分度圆直径 d1t= (4-3)23 )(12HEadt ZuTK34 2.805.189.46()m3.4516验算圆周速度 132.01.ss6tdnV选择精度等级 根据圆周速度由机械设计手册 6.4-19、6.4-20选择齿轮精度等级为 7 级精度(4)计算齿宽 b 和模数 mntb= 10.83245.96mdt1coscos12.4tntzfah 46.28.1)3.04cos1()(* nanmxh 56.18)30cos5tc2.56.2fa4.0.9hb(5) 计算载荷系数 K使用系数 由表查的K A=1.25动载系数 KV 根据圆周速度 v=1.66 由图查的 K V1.1sm齿间载荷分配系数 根据 由图查得Ha ar =1.20HaF齿间载荷分配系数 K 由表查的齿轮装配时检验调整HK 1.05+0.26 (1+0.6 ) +0.16 10-3b=1.29H2d载荷系数 K KK A KV K =1.25 1.1 1.20 1.29=2.12Ha修正小齿轮直径 1d331 2.1.45m35.8ttd计算模数 mn mn= 1cos.8cos2.6z(6)按齿根弯曲疲劳强度校核(4-4)cos221Fsaadn YzKTY计算载荷载荷系数 K 由 K 1.29 由图查得 =1.2746.hbHFKK= KA KV =1.25 1.1 1.20 1.27=1.74aF齿轮的弯曲疲劳强度极 由图查得EMPaFE89021齿形系数 FaY由当量齿数 z 80.134cos221znz 76.522n由图查的 78.1FaY90.12FaY应力修正系数 S由图查的 56.1a 85.2Sa重合度系数 anY7.0.由表查得 2)cos(i1arcosnb= cos08n 2)cos(in1nba= 94.2.051.cos2baan72.08.075.0anY螺旋角系数 由图根据 查得 0.98Y尺寸系数 由表的公式 5 时,取X nXm01.5.=5 =2nmY弯曲寿命系数 根据 N1=5.29 108 N2=9.35 107由图查得N8.011Y计算许用弯曲疲劳应力 取弯曲疲劳安全系数 S=1.4F1=F1620.870.5MPa4ENXYS2 21.计算大、小齿轮的 并加以比较YSaF=1FSa0745.5.789231.32SaY小齿轮的数值较大设计计算42322.17508.9cos80.745m2.3815m对比计算结果,由齿面接触疲劳强度计算的法面模数 mn与由齿根弯曲疲劳强度计算的法面模数相差不大,取标准值 mn2.5 ,取分度圆直径 d1=30.54(4-4)09.125.8cos430cos1nmdz则 ,取 13z 76.512u742z(7)几何尺寸计算计算中心距(4-5)12()(374)2.5m1.82coscos8nzma将中心距圆整为 120 。按圆整后的中心距修正螺旋角(4-6)512902.)7413()(arcos21 mzn因 值改变不多,故参数 等不必修正。HaZK、计算大、小齿轮的分度圆直径123.538.26mcos8647.179.nzdm计算齿轮宽度(4-7) 0.38.01db圆整后取 ; 。23B44.3.2 第二轴齿轮的设计计算(1) 按齿面接触疲劳强度设计选择齿数取 z 1=24, z 2=i1 z1=3.6 24=84确定变位系数 z 1=24 z2=84 a=20 h*an=h*acos 由图 查得 x1=0.38 x2=-0.38重合度系数 Z纵向重合度 037.8tan24.0138.tan138.0sin zmbd端面重合度 . 54.183.021nxz查得重合度则 则71.0a86.02a58.1a应力循环次数 N1=60.n 1.Lh=60 247.35 1 6300=9.35 107N2= (4-8)771 0.26.39i接触疲劳寿命系数由图查得KHN1=1.19 KHN2=1.15计算接触疲劳许用应力取失效概率为 1 安全系数 S1 1 1.19 1160=1380HNlim 2= =1.15 1160=1344SK1li则 1238041357MPaH(2)计算小齿轮分度圆直径 d1t小齿轮分度圆直径 d1t= (4-9)23 )(12HEadt ZuTK=35 2.6403.6189.46()m7.5881357(3)计算载荷系数 K齿间载荷分配系数 根据 由图查得查得 Ha ar HaK=1.20FaK齿间载荷分配系数 K 由表查的 齿轮装配时检验调整得 HK 1.30H载荷系数 K KK A KV K =1.25 1.05 1.20 1.30=2.05Ha修正小齿轮直径 1d331 2.0547.8m49.116ttd计算模数 mnt 1cos9.cos2.0nt(4)按齿根弯曲疲劳强度设计(4-10)cos212Fsaadn YzKTY计算载荷载荷系数 K 由图 查得 =1.25K= KA KV =1.25 1.05 1.20 1.25=1.97HaF齿形系数 FaY由当量齿数 z 16.8cos221zvz 0.7422v由 4图 14-1-47 59.1FaY21.FaY应力修正系数 Sa由图查得 6.1 74.2Sa重合度系数 anY7502.已知 914.0b 58.1cos2baan72.0.07.0anY弯曲寿命系数 根据 N1=9.35 108 N2=2.67 107由图查得N.1Y计算许用弯曲疲劳应力 取弯曲疲劳安全系数 S=1.4F1= 160.8253.7MPa4ENXS2F21.Y计算大、小齿轮的 并加以比较SaF=1FSaY0761.583972492.2Sa小齿轮的数值较大设计计算 523221.970.98cos140.681m.3.m对比计算结果,由齿面接触疲劳强度计算的法面模数 mn与由齿根弯曲疲劳强度计算的法面模数相差不大,取标准值 mn2.5 ,取分度圆直径 d1=48.9096.15.24cos908cos1 nmdz则 ,则12z 4.68127.51uz692z(7)几何尺寸计算计算中心距(4-11)12()(69)2.5m104.3coscos4nzma将中心距圆整为 105 。按圆整后的中心距修正螺旋角(4-12)021 36.150.2)69()(arcoszn因 值改变不多,因此参数中 等不须要修正。HaZK、 计算大、小齿轮的分度圆直径102.5m31.cos69.78.9nzd计算齿轮宽度.241.3801db圆整后取 ; 。25mB图 4-2 齿轮的三维效果图4.3.3 第三轴齿轮的设计计算(1)按齿面接触疲劳强度设计选择齿数取 z 1=12, z 2=i z1=3.2 11=35.23转距 T 5T.20Nm确定变位系数 z1=12 z2=45 a=20 h*an=h*acos 由机械设计手册图 14-1-4 查的 x1=0.35 x2=-0.35节点区域系数 ZH X =0 =8 查由机械设计手册图 14-1-16ZH=2.46重合度系数 Z纵向重合度 17.08tan1.038.tan138.0sin zmbd端面重合度 . 15.83.01nxz查得重合度则65.01a87.02a43.1a应力循环次数 N1=60.n1.Lh=60 70.67 1 6300=2.67 107N2= 671 05.09.46i接触疲劳寿命系数由由机械设计手册图查KHN1=1.20 KHN2=1.15计算接触疲劳许用应力取失效概率为 1 安全系数 S1 1 1.23 1160=1427HNlim 2 =1.39 1160=1612MPaSK1li 124716250MPaH(2)计算小齿轮分度圆直径 d1t小齿轮分度圆直径 d1t= (4-13)23 )(12HEadt ZuTK=35 2.6904.918.46()m3.0781350(3)计算载荷系数 K齿间载荷分配系数 根据 由图查得HaKar =1.10HaKF齿间载荷分配系数 K 由机械设计得 设计手册齿轮装配时检H验调整K 1.05+0.26 (1+0.6 ) +0.16 10-3b=1.292d载荷系数 K KK A KV K =1.25 1.05 1.10 1.29=1.86Ha修正小齿轮直径 1d(4-14)331 1.866.07m.2tt计算模数 mnt (4-15)1cos.2cos85.961ntd(4)按齿根弯曲疲劳强度设计(4-16)cos212Fsaadn YzKTYm计算载荷载荷系数 K K= KA KV =1.25 1.05 1.10 1.25=1.80 HaF齿形系数 aY由当量齿数 z 2.18cos221znz 9.4522n由图 14-1-47 31.FaY0.FaY应力修正系数 SaY由图 50.176.12Sa重合度系数 an50.47.19.3cos22ban6.0.5.07河南理工大学万方科技学院毕业设计外文资料与中文翻译1外文资料与中文翻译外文资料:The Effect of a Viscous Coupling Used as a Front-Wheel Drive Limited-Slip Differential on Vehicle Traction and Handling1 ABCTRACTThe viscous coupling is known mainly as a driveline component in four wheel drive vehicles. Developments in recent years, however, point toward the probability that this device will become a major player in mainstream front-wheel drive application. Production application in European and Japanese front-wheel drive cars have demonstrated that viscous couplings provide substantial improvements not only in traction on slippery surfaces but also in handing and stability even under normal driving conditions.This paper presents a serious of proving ground tests which investigate the effects of a viscous coupling in a front-wheel drive vehicle on traction and handing. Testing demonstrates substantial traction improvements while only slightly influencing steering torque. Factors affecting this steering torque in front-wheel drive vehicles during straight line driving are described. Key vehicle design parameters are identified which greatly influence the compatibility of limited-slip differentials in front-wheel drive vehicles.Cornering tests show the influence of the viscous coupling on the self steering behavior of a front-wheel drive vehicle. Further testing demonstrates that a vehicle with a viscous limited-slip differential exhibits an improved stability under acceleration and throttle-off maneuvers during cornering.2 THE VISCOUS COUPLINGThe viscous coupling is a well known component in drivetrains. In this 河南理工大学万方科技学院毕业设计外文资料与中文翻译2paper only a short summary of its basic function and principle shall be given.The viscous coupling operates according to the principle of fluid friction, and is thus dependent on speed difference. As shown in Figure 1 the viscous coupling has slip controlling properties in contrast to torque sensing systems.This means that the drive torque which is transmitted to the front wheels is automatically controlled in the sense of an optimized torque distribution.In a front-wheel drive vehicle the viscous coupling can be installed inside the differential or externally on an intermediate shaft. The external solution is shown in Figure 2.This layout has some significant advantages over the internal solution. First, there is usually enough space available in the area of the intermediate shaft to provide the required viscous characteristic. This is in contrast to the limited space left in todays front-axle differentials. Further, only minimal modification to the differential carrier and transmission case is required. In-house production of differentials is thus only slightly affected. Introduction as an option can be made easily especially when the shaft and the viscous unit is supplied as a complete unit. Finally, the intermediate shaft makes it possible to provide for sideshafts of equal length with transversely installed engines which is important to reduce torque steer (shown later in section 4).This special design also gives a good possibility for significant weight and cost reductions of the viscous unit. GKN Viscodrive is developing a low weight and cost viscous coupling. By using only two standardized outer diameters, standardized plates, plastic hubs and extruded material for the housing which can easily be cut to different lengths, it is possible to utilize a wide range of viscous characteristics. An example of this development is shown in Figure 3.3 TRACTION EFFECTSAs a torque balancing device, an open differential provides equal tractive effort to both driving wheels. It allows each wheel to rotate at different speeds 河南理工大学万方科技学院毕业设计外文资料与中文翻译3during cornering without torsional wind-up. These characteristics, however, can be disadvantageous when adhesion variations between the left and right sides of the road surface (split-) limits the torque transmitted for both wheels to that which can be supported by the low- wheel.With a viscous limited-slip differential, it is possible to utilize the higher adhesion potential of the wheel on the high-surface. This is schematically shown in Figure 4.When for example, the maximum transmittable torque for one wheel is exceeded on a split-surface or during cornering with high lateral acceleration, a speed difference between the two driving wheels occurs. The resulting self-locking torque in the viscous coupling resists any further increase in speed difference and transmits the appropriate torque to the wheel with the better traction potential.It can be seen in Figure 4 that the difference in the tractive forces results in a yawing moment which tries to turn the vehicle in to the low-side, To keep the vehicle in a straight line the driver has to compensate this with opposite steering input. Though the fluid-friction principle of the viscous coupling and the resulting soft transition from open to locking action, this is easily possible, The appropriate results obtained from vehicle tests are shown in Figure 5.Reported are the average steering-wheel torque Ts and the average corrective opposite steering input required to maintain a straight course during acceleration on a split-track with an open and a viscous differential. The differences between the values with the open differential and those with the viscous coupling are relatively large in comparison to each other. However, they are small in absolute terms. Subjectively, the steering influence is nearly unnoticeable. The torque steer is also influenced by several kinematic parameters which will be explained in the next section of this paper.河南理工大学万方科技学院毕业设计外文资料与中文翻译44 FACTORS AFFECTING STEERING TORQUEAs shown in Figure 6 the tractive forces lead to an increase in the toe-in response per wheel. For differing tractive forces, Which appear when accelerating on split-with limited-slip differentials, the toe-in response changes per wheel are also different.Unfortunately, this effect leads to an undesirable turn-in response to the low- side, i.e. the same yaw direction as caused by the difference in the tractive forces.Reduced toe-in elasticity is thus an essential requirement for the successful front-axle application of a viscous limited-slip differential as well as any other type of limited-slip differential.Generally the following equations apply to the driving forces on a wheelVTFWith Tractive ForceVertical Wheel LoadVUtilized Adhesion CoefficientThese driving forces result in steering torque at each wheel via the wheel disturbance level arm “e” and a steering torque difference between the wheels given by the equation: =eTloHhiFcosWhere Steering Torque DifferenceeTe=Wheel Disturbance Level ArmKing Pin Anglehi=high-side subscriptlo=low-side subscriptIn the case of front-wheel drive vehicles with open differentials, Ts is 河南理工大学万方科技学院毕业设计外文资料与中文翻译5almost unnoticeable, since the torque bias ( ) is no more than 1.35.loHhiTF/For applications with limited-slip differentials, however, the influence is significant. Thus the wheel disturbance lever arm e should be as small as possible. Differing wheel loads also lead to an increase in Te so the difference should also be as small as possible.When torque is transmitted by an articulated CV-Joint, on the drive side (subscript 1) and the driven side (subscript 2),differing secondary moments are produced that must have a reaction in a vertical plane relative to the plane of articulation. The magnitude and direction of the secondary moments (M) are calculated as follows (see Figure 8):Drive side M1 = vvTTtan/)2/tan(Driven side M2 = With T2 = dynTrF=stemJofi,2Where Vertical Articulation AnglevResulting Articulation AngleDynamic Wheel RadiusdynrAverage Torque LossTThe component acts around the king-pin axis (see figure 7) as cos2Ma steering torque per wheel and as a steering torque difference between the wheels as follows: )tan/2/tan()sin/2/tan(cos 22 liwhiwTTT where Steering Torque Difference河南理工大学万方科技学院毕业设计外文资料与中文翻译6W Wheel side subscriptIt is therefore apparent that not only differing driving torque but also differing articulations caused by various driveshaft lengths are also a factor. Referring to the moment-polygon in Figure 7, the rotational direction of M2 or respectively change, depending on the position of the wheel-center to the Tgearbox output.For the normal position of the halfshaft shown in Figure 7(wheel-center below the gearbox output joint) the secondary moments work in the same rotational direction as the driving forces. For a modified suspension layout (wheel-center above gearbox output joint, i.e. negative) the secondary vmoments counteract the moments caused by the driving forces. Thus for good compatibility of the front axle with a limited-slip differential, the design requires: 1) vertical bending angles which are centered around or 0vnegative ( ) with same values of on both left and right sides; and 2) 0vvsideshafts of equal length.The influence of the secondary moments on the steering is not only limited to the direct reactions described above. Indirect reactions from the connection shaft between the wheel-side and the gearbox-side joint can also arise, as shown below:Figure 9: Indirect Reactions Generated by Halfshaft Articulation in the Vertical PlaneFor transmission of torque without loss and both of the vdwsecondary moments acting on the connection shaft compensate each other. In reality (with torque loss), however, a secondary moment difference appears: WDWM12With T2河南理工大学万方科技学院毕业设计外文资料与中文翻译7The secondary moment difference is:DWM VWVWVDV TTwT tan/2/tansi/tan22/2 For reasons of simplification it apply that and VDto giveWD VVVTtan/1si/2tan requires opposing reaction forces on both joints where DM. Due to the joint disturbance lever arm f, a further steering LFW/torque also acts around the king-pin axis:fTDf /cosloDWhiWf LM/Where Steering Torque per WheelfSteering Torque DifferencefTJoint Disturbance LeverConnection shaft (halfshaft) LengthLFor small values of f, which should be ideally zero, is of minor fTinfluence.5EFFECT ON CORNERINGViscous couplings also provide a self-locking torque when cornering, due to speed differences between the driving wheels. During steady state cornering, as shown in figure 10, the slower inside wheel tends to be additionally driven through the viscous coupling by the outside wheel.Figure 10: Tractive forces for a front-wheel drive vehicle during steady state cornering 河南理工大学万方科技学院毕业设计外文资料与中文翻译8The difference between the Tractive forces Dfr and Dfl results in a yaw moment MCOG, which has to be compensated by a higher lateral force, and hence a larger slip angle af at the front axle. Thus the influence of a viscous coupling in a front-wheel drive vehicle on self-steering tends towards an understeering characteristic. This behavior is totally consistent with the handling bias of modern vehicles which all under steer during steady state cornering maneuvers. Appropriate test results are shown in figure 11.Figure 11: comparison between vehicles fitted with an open differential and viscous coupling during steady state cornering.The asymmetric distribution of the tractive forces during cornering as shown in figure 10 improves also the straight-line running. Every deviation from the straight-line position causes the wheels to roll on slightly different radii. The difference between the driving forces and the resulting yaw moment tries to restore the vehicle to straight-line running again (see figure 10).Although these directional deviations result in only small differences in wheel travel radii, the rotational differences especially at high speeds are large enough for a viscous coupling front differential to bring improvements in straight-line running.High powered front-wheel drive vehicles fitted with open differentials often spin their inside wheels when accelerating out of tight corners in low gear. In vehicles fitted with limited-slip viscous differentials, this spinning is limited and the torque generated by the speed difference between the wheels provides additional tractive effort for the outside driving wheel. this is shown in figure 12Figure 12: tractive forces for a front-wheel drive vehicle with viscous limited-slip differential during acceleration in a bend The acceleration capacity is thus improved, particularly when turning or accelerating out of a T-junction maneuver ( i.e. accelerating from a stopped position at a “T” intersection-right or left turn ).河南理工大学万方科技学院毕业设计外文资料与中文翻译9Figures 13 and 14 show the results of acceleration tests during steady state cornering with an open differential and with viscous limited-slip differential .Figure 13: acceleration characteristics for a front-wheel drive vehicle with an open differential on wet asphalt at a radius of 40m (fixed steering wheel angle throughout test).Figure 14: Acceleration Characteristics for a Front-Wheel Drive Vehicle with Viscous Coupling on Wet Asphalt at a Radius of 40m (Fixed steering wheel angle throughout test)The vehicle with an open differential achieves an average acceleration of 2.0 while the2/smvehicle with the viscous coupling reaches an average of 2.3 (limited 2/smby engine-power). In these tests, the maximum speed difference, caused by spinning of the inside driven wheel was reduced from 240 rpm with open differential to 100 rpm with the viscous coupling.During acceleration in a bend, front-wheel drive vehicles in general tend to understeer more than when running at a steady speed. The reason for this is the reduction of the potential to transmit lateral forces at the front-tires due to weight transfer to the rear wheels and increased longitudinal forces at the driving wheels. In an open loop control-circle-test this can be seen in the drop of the yawing speed (yaw rate) after starting to accelerate (Time 0 in Figure 13 and 14). It can also be taken from Figure 13 and Figure 14 that the yaw rate of the vehicle with the open differential falls-off more rapidly than for the vehicle with the viscous coupling starting to accelerate. Approximately 2 seconds after starting to accelerate, however, the yaw rate fall-off gradient of the viscous-coupled vehicle increases more than at the vehicle with open differential.The vehicle with the limited slip front differential thus has a more stable initial reaction under accelerating during cornering than the vehicle with the 河南理工大学万方科技学院毕业设计外文资料与中文翻译10open differential, reducing its understeer. This is due to the higher slip at the inside driving wheel causing an increase in driving force through the viscous coupling to the outside wheel, which is illustrated in Figure 12. the imbalance in the front wheel tractive forces results in a yaw moment acting in CSDMdirection of the turn, countering the understeer.When the adhesion limits of the driving wheels are exceed, the vehicle with the viscous coupling understeers more noticeably than the vehicle with the open differential (here, 2 seconds after starting to accelerate). On very low friction surfaces, such as snow or ice, stronger understeer is to be expected when accelerating in a curve with a limited slip differential because the driving wheels-connected through the viscous coupling-can be made to spin more easily (power-under-steering). This characteristic can, however, be easily controlied by the driver or by an automatic throttle modulating traction control system. Under these conditions a much easier to control than a rear-wheel drive car. Which can exhibit power-oversteering when accelerating during cornering. All things, considered, the advantage through the stabilized acceleration behavior of a viscous coupling equipped vehicle during acceleration the small disadvantage on slippery surfaces.Throttle-off reactions during cornering, caused by releasing the accelerator suddenly, usually result in a front-wheel drive vehicle turning into the turn (throttle-off oversteering ). High-powered modeles which can reach high lateral accelerations show the heaviest reactions. This throttle-off reaction has several causes such as kinematic influence, or as the vehicle attempting to travel on a smaller cornering radius with reducing speed. The essential reason, however, is the dynamic weight transfer from the rear to the front axle, which results in reduced slip-angles on the front and increased slip-angles on the rear wheels. Because the rear wheels are not transmitting driving torque, the influence on the rear axle in this case is greater than that of the front axle. The 河南理工大学万方科技学院毕业设计外文资料与中文翻译11driving forces on the front wheels before throttle-off (see Figure 10) become over running or braking forces afterwards, which is illustrated for the viscous equipped vehicle in Figure 15.Figure 15:Baraking Forces for a Front-Wheel Drive Vehicle with Viscous Limited-Slip Differential Immediately after a Throttle-off Maneuver While CorneringAs the inner wheel continued to turn more slowly than the outer wheel, the viscous coupling provides the outer wheel with the larger braking force . fBThe force difference between the front-wheels applied around the center of gravity of the vehicle causes a yaw moment that counteracts the normal GCM0turn-in reaction.When cornering behavior during a throttle-off maneuver is compared for vehicles with open differentials and viscous couplings, as shown in Figure 16 and 17, the speed difference between the two driving wheels is reduced with a viscous differential.Figure 16: Throttle-off Characteristics for a Front-Wheel Drive Vehicle with an open Differential on Wet Asphalt at a Radius of 40m (Open Loop)Figure 17:Throttle-off Characteristics for a Front-Wheel Drive Vehicle with Viscous Coupling on Wet Asphalt at a Radius of 40m (Open Loop)The yawing speed (yaw rate), and the relative yawing angle (in addition to the yaw angle which the vehicle would have maintained in case of continued steady state cornering) show a pronounced increase after throttle-off (Time=0 seconds in Figure 14 and 15) with the open differential. Both the sudden increase of the yaw rate after throttle-off and also the increase of the relative yaw angle are significantly reduced in the vehicle equipped with a viscous limited-slip differential.A normal driver os a front-wheel drive vehicle is usually only accustomed to neutral and understeering vehicle handing behavior, the driver 河南理工大学万方科技学院毕业设计外文资料与中文翻译12can then be surprised by sudden and forceful oversteering reaction after an abrupt release of the throttle, for example in a bend with decreasing radius. This vehicle reaction is further worsened if the driver over-corrects for the situation. Accidents where cars leave the road to the inner side of the curve is proof of this occurrence. Hence the viscous coupling improves the throttle-off behavior while remaining controllable, predictable, and safer for an average driver.6. EFFECT ON BRAKING The viscous coupling in a front-wheel drive vehicle without ABS (anti-lock braking system) has only a very small influence on the braking behavior on split- surfaces. Hence the front-wheels are connected partially via the front-wheel on the low- side is slightly higher than in an vehicle with an open differential. On the other side ,the brake pressure to lock the front-wheel on the high- side is slightly lower. These differences can be measured in an instrumented test vehicle but are hardly noticeable in a subjective assessment. The locking sequence of front and rear axle is not influenced by the viscous coupling.Most ABS offered today have individual control of each front wheel. Electronic ABS in front-wheel drive vehicles must allow for the considerable differences in effective wheel inertia between braking with the clutch engaged and disengaged.Partial couplin
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