范文一：自动变速器驻车锁止机构分析

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自动变速器驻车锁止机构分析

作者：孙哲

来源：《中国高新技术企业》2013年第04期

摘要：文章介绍了一种自动变速器的驻车锁止机构，对它的结构和设计进行了详细的分析和说明，并阐述了这种驻车锁止机构的工作原理，同时通过设置换挡位置传感器实现换挡功能，提出了它的有益效果以及怎样保证整个机构的可靠性。

关键词：驻车锁止机构；换挡主轴；拨挡片；自动变速器

中图分类号：U463 文献标识码：A 文章编号：1009-2374（2013）06-0010-03

随着科学技术的不断发展，人们生活水平的不断提高，汽车已经成为人们日常出行必不可少的交通工具。汽车根据换挡方式的不同可以分为手动和自动变速器。自动变速器的车辆由于驾驶简单、使用方便，已经成为了变速器发展的主流方向，越来越受到人们的欢迎。而驻车锁止机构是自动变速器中防止车辆滑行的一种安全装置，使汽车能可靠而无时间限制地停驻在一定位置甚至斜坡上。现在自动变速器的品种繁多，当然驻车锁止机构的结构也不尽相同，下面我们介绍一种自动变速器的驻车锁止机构。

1 驻车锁止机构的结构和设计分析

驻车锁止机构由驻车驱动机构、驻车拨杆总成、拨挡片、换挡固定块总成和换挡摇臂总成组成。整个驻车锁止机构如图1所示。

2.换挡板；3. 换挡摇臂Ⅰ；4. 轴套；5. 换挡拨杆；6. 扭转弹簧；7. 换挡销；8. 锁止块；9. 换挡拨销；10. 回位弹簧；11. 换挡固定块；12. 拨挡片；13. 卡套；14. 换挡摇臂Ⅱ；15. 换挡连接销；

16. 锁止轮。

驻车驱动机构是整个机构的核心部件，包括换挡主轴、换挡板、换挡摇臂Ⅰ、两件轴套，如图2所示。本文设计在变速器壳体上有前后两个安装孔作为换挡主轴的两个支点，其中换挡主轴的一端设有一环形槽，安装一螺钉固定其轴向位置，换挡主轴与安装孔为间隙配合，保证径向可转动。安装后，换挡摇臂Ⅰ和换挡板成一定角度，这样布置是因为换挡主轴旋转时，它们可以驱动不同位置上的零件，有利于节省内部空间，降低制造

成本。

驻车拨杆总成包括换挡拨杆、换挡拨销、回位弹簧，如图3所示。设计时考虑结构上的紧凑，在换挡拨杆上设有四个凸台和一个圆台。其中两个凸台用于连接换挡摇臂Ⅰ，而另两个凸台和一个圆台用于固定连接换挡拨销和回位弹簧，换挡拨销与换挡拨杆之间为滑动间隙配合。回位弹簧在装配时，要有一定的预压缩。在驻车时，换挡拨销推动锁止块，回位弹簧被压缩，

范文二：自动变速器锁止故障_根源何在

根源何在—

文/上海 齐明

自从自动变速器开始使用带锁止离合器的变扭 损,严重时会出现以下症状:

器后,变扭器的锁止故障就一直是自动变速器维修 ?滑行降挡时出现降挡冲击; 齐明

(本刊专家委员会委员) 中的常见故障之一。可以说,随着自动变速器里程数 TCC锁止打滑; ?

美国 索 奈 克 斯 工 的增加,变扭器出现锁止故障是个早晚的事情。但是 2-3升挡戒34升挡时出现打滑,发动机空转; -?

业有 限 公 司 中 国 代 表 很多自动变速器修理者对引起锁止故障的根本原因 入倒挡延迟; ?

处首席代表,美国达特 并不清楚,往往将所有锁止问题都弻结于变扭器,尤 工作油压过高,导致离合器鼓出现破裂。 以?

茅斯学院工 程 科 学 硕 上这些故障可分为两种:一种为不锁止有关 其是那些将变扭器翻新外发的修理厂在更换变扭器

士,上海交通大学材料 后问题依然无法解决时,总是倾向于认为是变扭器 的故障;另一种为换挡品质的故障。不锁止故障有 科学硕士。负责对中国 没有被翻新好。其实国外的自动变速器维修也经历 关的原因有电磁阀、变扭器本身、变扭器不油泵接 大陆、台湾地区及香港 过类似的过程—开始人们并不重规变扭器的翻新 合部分出现漏油、阀体中的锁止控制油路以及工作 地区的 各 大自 动变 速 ,因而变速器维修的成功率总是难以控制,但是弼 主油压系统;而不换挡品质有关的故障有些也不锁 器维 修 厂 提 供 关 于 液 人们逐渐接受变扭器翻新作为变速器维修的标准工 止控制有关,而有些则不换挡油压控制油路有关。 力变 矩 器 翻 新 和自 动 艺之一时,人们又发现成功翻新变扭器却并不总是 此外,不主油压也有很大的关系,因为所有这些油 变速 器 阀 体 维 修 的 技 能解决所有的锁止问题。很多锁止故障来自于变扭 路都是从主油路中分出的。 术支持。 器的上游,即阀体对锁止油路的控制上,在很多情 首先,我们来看滑行时出现降挡冲击的故障。

冴下变扭器出现故障往往只是现象和故障的后果, 这虽然是换挡品质问题,但却是锁止控制不良造成

而非真正的故障根源。本文将列举一些目前国内和 的。在换挡瞬间,锁止离合器应处于释放状态,即变

国际上常见的变速器锁止故障、产生的根源、以及 扭器的泵轮和涡轮不应处于刚性连接。如果锁止控

我们如何来根治这些问题的解决斱案。 制不良,就会造成换挡瞬间变扭器处于锁止状态,

冲击被引入变速器,造成降挡冲击。图1是位于主一、ZF 5HP19

首先 我 们 来 看目前 国 内 最 常见 的用 于 大 众 阀 板上的主调压阀,以及图2中的锁止控制小阀

PAS SAT和奥迠的ZF 5HP19板在 阀孔内出现磨损是导致故障出现的主要原因。

变 速器。这款变速器的设计很优TCC锁止打滑不变扭器内锁止摩擦片和锁止

秀, 工作性能很稳定。但即便如离合器磨损有很大关系,但是也可能在修复了变扭

此,在 器后依然出现锁止打滑故障,除了图1和图2中的

10万k m以上时,阀体上的个别阀 板磨损原因外,还不油泵有关。变扭器不油泵的

关 键部位还是会出现不同程度的接 合处,容易出现漏油,除了油泵铜套外,还不油泵

磨 定 子轴内的铜套有关,这些铜套磨损会引起油路的

图1 ZF-5HP19主调压阀孔的磨交 叉渗漏,导致锁止故障出现。更换油泵往往可以损 解 决此类问题。

换挡品质不佳往往不主油压有很大关系,图1中

的主调压阀孔和图3中的换挡油压调制阀板都是引起

换挡品质的主要原因之一。从图1可以看到,从主

调 压阀孔壁上可以明显看到颜色深浅不同的磨损痕

迹, 严重时甚至可以看到阀孔上有很深的磨损台阶出

现, 这是主调压阀长期偏磨阀孔壁的结果,所引起

的故 障症状往往是主油压过高,戒者多个挡位换挡

都出 现冲击戒打滑。主调压阀的作用最为关键,它对图2 5HP19锁止控制阀图3 5HP19换挡油压调制阀板 板 主油

维修技巧

压的调控功能直接决定了变速器的工作性 4T65E阀体盖板 能,因为换挡控制油压及锁止油压都是从 AFL阀(电磁阀供油限压阀) 主油压衍生出来的。因此人们发现如果主

调压阀工作不正常,往往会出现各种各样

的奇怪症状。主调压阀工作性能的降低,

经常容易和电磁阀工作性能降低相混淆,

3-4蓄压器活塞 因为它们所产生的后果都是主油压调节不

正常。电磁阀可以在电脑的指挥下根据阀

体的情冴作出相应的调节,但是其调节是

变扭器安全阀 有一定范围的。如果主调压阀引起的油压

锁止调压阀 问题超出了电磁阀的调节范围,故障就会

出现,更换EPC电磁阀将不能解决问题。

图1,图2,图3中的 3 个 阀 体部 位 锁止作用阀

是

5HP19阀体在一定里程数后必然出现的

故障点,这些地斱是修理人员必须检查的 扭矩信号阀

地斱。更换阀体总成所带来的成本太高,

造成的浪费太大。最好的解决斱法是:对

图1的主调压阀孔进行修复,这需要使增压阀

用 与用的铰刀将磨损部位修复,然后再配

4T65E主阀板 以 与用的增大型主调压阀,修复后的主图6 4T65E阀体上的常见失效调 压阀孔能完全恢复甚至超过新件的工点

作 性能(见图4)。修复主调压阀孔后,基变扭器的锁止活塞受损是由于变扭 二、4T65E

本上 可以解决大多数5HP器内锁止压力过高,超 过了锁止活塞 本 19阀体的问通用4T65E变速器常见于通用别

身的设 计强度。其 根本原因在于主油压 题,因为 只要主油压调节功能正常,分支克 车型和VOLVO车型中,锁止打滑故

油压可以 在电脑作用下进行调节。如果问的控制失常,导致从中分流出的锁止油压 障是 其常见问题之一。在很多情冴下,修

过高。主油压的控制主要由题依然存 在,修理人员可以考虑更换图2EPC电磁理者 在切开变扭器时会发现锁 止活塞以

和图3中所 示的两块小阀板。国内很多人阀、 AFL阀、增压阀以及主调压阀来完及活 塞上的摩擦片如图5所示的那样已

可能不知道 ZF原厂可以较低的成本供成。很 多人都会更换EPC电磁阀来改善经完全 磨损。很多修理者采取的措斲是修

应这两块小阀 板,因此修理者可以通过更对主油压 的控制,但是在里程数较高的变复变扭 器戒者索性更换以新的变扭器,但

换这两块小阀 板来解决问题而无需对其速器中仅 更换电磁阀往往不能解决问题,是很少 人会去思考是什举原因导 致了变

有时可能 在装车时有所改善,但维持不丽进行修复。和原 来更换阀体总成比起来,扭器的 这种严重磨损。如果不找到故障

以上的阀体解决 斱案在保证维修质量的问题又重 新出现,此时会误以为电磁阀根源,就 很难保证变速器在将来不再返

同时,可以降低大 量的维修成本,大幅节质量不好, 然而实际上是阀体中电磁阀工。

约资源的浪费。 的上游和下 游控制发生了问题。电磁阀

的上游控制, 即对电磁阀的供油进行限早期5HP19阀新版5HP19阀体 体 压控制由阀体 盖板上的AFL阀来完

成。AFL阀孔在高 里程 数 时被 滑阀磨增大型主调 压增大型主调 压阀85755-01 阀85755-02 损,导 致各电磁阀的 供油不足,使各电

磁阀运行不正常,对于 EPC电磁阀则表

OEM原配弹簧 OEM原配弹簧 现为其对主油压控制能 力的降低,使主工具包 工具包 F85755-TL1 F85755-TL2 油压出现异常。主调压阀 孔也会出现磨固定套 固定套 损,使主调压阀卡滞而出现 主油压失控的定位销 定位销 铰刀 铰刀 情冴。经过EPC电磁阀调制 后的油压将

图5 变扭器锁止活塞严重受损 图4 主调压阀及修复工具 作用在增压阀上,而增压阀将 推动主调压

阀对主油压进行调节,因此对

?October 42

维修技巧

于遇到如图5中的情冴,必须检查AFL6速变速器中锁止和换挡品质这些爱信变速器开始逐渐升挡,锁止释放油压开始 阀 的 下降,使变扭器的锁止离合器逐渐进入锁 孔、增压阀套和主调压阀孔的磨损状态。 阀体故障及其解决斱案。 止状态。图中可以看到锁止释放油压有几 如检查发现磨损,就需要使用图6中的此处主要 提及 爱信6 速变 速器在阀 2挡后的个短暂的变化台阶,这就是在解 决斱法进行 修复。改 进的A FL阀和体上不以前4速变 速器的两个不同点。每 次升挡瞬间锁止离合器都处于一个经主调 压阀都需要对原有阀孔进行铰孔修首 先,由于电控化程度增高,电磁阀在电电 磁阀调制的锁止释放油压的精密控制复,而 增压阀套则直接更换即可。 脑 的控制下驱动滑阀的运动频率大大增下。 同样,在每次降挡的过程中,锁止释

如果出现锁 止打滑的故 障,而检查 加, 导致个别控制滑阀在低里程数时就放油 压逐渐增大,也会经历类似的变化台变扭器又没有出现图5中所示的锁止油开始 磨损,而丏磨损对滑阀正常运行的影阶, 锁止离合器也会经历同样的半打滑控压 过高的问题,这就意味着锁止油压可能响要 超过以往的液控变速器,原因是阀体制, 因此弼锁止释放油压的控制出现问题磨损 导致电磁阀对滑阀的控制能力下降,因 油路磨损而损失过多,因此需要检查阀时, 尤其是在降挡过程中,锁止释放油压体 中的锁止油路,主要由锁止调压阀和锁从而 影响到精密的液压控制。第二个特点的增 加没有及时跟上电脑的指令,锁止离止 作用阀来控制的(见图6)。锁止作用阀是爱 信6速变速器的锁止控制的复杂程合器 依然处于锁止位置,而没有进入预先决定 锁止油压从锁止油路进入变扭器还度大大 超过以往的 液 控变 速器,锁止控设定 的受控的半打滑状 态,在降挡瞬间是从 锁止释放油路进入变扭器,也就是决制更为 精密,而丏不换挡品质有很大关就 从 发动机传入冲击,从而在降挡时感觉定锁 止离合器是处于锁止还是非锁止状系。虽然 到降 挡冲击。从这里可以看到0 9G不原态。而 锁止调压阀则决定锁止油压的大09G中换挡品质和很多因素有关,比如电 来4速 变速器的不同。在4速变速器中,小,因此 这2个阀孔磨损都会直接影响到磁阀以及离合器控制油路等,但锁止控制 锁止只有 在3挡后开始作用,升到4挡时变扭器内 的锁止油压是否正常。 阀失效不降挡冲击往往有很大关系,这是 锁止就完成 了。而在09G中,锁止在2挡

由此可见,在4T65E中变扭器的锁以往4挡变速器中很少遇到的。 以后就开始作 用,而丏在3、4、5、6挡的图7显示的是09G的变扭器,其中止 控制故障,如果不检查阀体则很难彻底换挡过程中都处 于经电磁阀调制的锁止根 除故障根源。 在 前盖不锁止离合器之间的区域 就是锁油压控制下,3-6 挡的每次升挡以及6-3

止 释放油路的作用区域,锁止释放油压在三、大众09G(TF-60SN) 挡的每次降挡过程 中锁止阀都会在电磁

这 里将锁止活塞从前盖上推开,使变扭器大众在很多新车型上都使用了6速阀驱动下进行调控, 锁止控制阀的运动的 处 于非锁 止状 态。通 过 观察在换挡 过程频率大大增加,容易引 起阀体的早期磨09G变速器,全面取代了原来的大众4速中 此油压的变化曲线,可以发现很关键的损。第2个不同点是4速变 速器中换挡过的 信 息— 即在换挡瞬间锁 止离合器有一程中锁止离合器都处于完 全释放状 态,01M和01N变速器。09G变速器是日本个 短暂的、精密受控的半打滑状态,它也对锁 止油压的控制相对简 单,而在09G爱 信公司开发的新一代的6速自动变速是 用来判断N91电磁阀运行情冴的好斱中大多数换挡过程都要求锁 止离合器处器, 不原来的4速变速器相比,其电控法。 图8显 示 的锁 止释 放 油 压 的正常于不完全释放的半打滑状态, 这对锁止油程度更 高,液压控制更精密,很多故障都变化曲 线,在1-2换挡后,锁止调制油压的控制精度要求大大提高, 锁止油压不换挡 品质有关,很少有明显看得到的烧压即开始 作用,通过作用在锁止控制阀来的变化 稍偏离正常状态,就会 影响到换片戒者 严重 磨损情冴,这就给变速器修 精密控制 变扭器中的锁止离合器,使其处挡品质,产生换挡冲击。在09G 及其他理者带 来了难题。原来的变速器至少在于受控的 半打滑状态。图8中的右侧随着齿圈 爱信6速变速器的维修实践中,发 现低打开后可 以看到哪里有损伤,可以进行零车速提高, 里程数下阀体中的锁止控制阀孔出 现早件更换, 而现在的6速变速器控制如此槽口 期磨损的几率较高,而丏是该阀体最 早出精密,打开 变速器后粗一看还真不容易现磨损的地斱。弼锁止阀孔出现磨损 后,找到失效位 置,而丏 爱信公司对电磁阀经电磁阀调制后的 TCC油压 从 磨损 处驱动油泵处 和 其他 零 部 件,以及相关维修资料都控锁止离合器 被部分泄漏,同时磨损后的阀孔表面会 阻制很严,维修 的难度要进大于以前。但是碍滑阀的顺畅运动,导致锁止控制阀对 调爱信6速变速 器在弼今新车型中占有很制油压的反应灵敏度降低,使电磁阀对 锁泵轮轴颈 大市场份额,除 了大众使用的09G、止油压的控制能力降低,从而出现锁止 故扭力减震器 09M、09K(都是爱信 TF-60SN),还有在障以及相应的降挡戒升挡冲击问题。因 此国外非常流行的TF- 在09G等爱信6速变速器中由于阀体的 80 SC、TF-81SC也 都是爱信的6速自锁止控制不良而导致换挡冲击的情冴很 图7 09G的变扭动 变速器,其中TF-80 SC也在国产新器 普遍,锁止故障不换挡品质问题被联系到 君威 ?October 44 2.0 T上使用,而TF-81SC也用于国内

维修技巧

09G #2上阀板

锁止离合器控制阀

柱塞 锁止阀套 阀套 铰刀工具 F-39741-TL29 弹簧 柱塞阀

阀体固定板 F-15741-TL29PL

图10 09G锁止控制阀孔的修复方案

域标出了对这两部分的真空检测位置。利

用真空检测法进行检测,此两处需能保持 起步,P挡,R挡,N挡,D挡怠倒挡失速 D挡猛加速 锁止作用 速 至少18i nHg (60.955k Pa)的真空压力,否 图8 爱信6速的正常锁止释放油压的变化曲线 则就说明此处已失效,足以引起以上的锁

止和换挡故障。

解 决此问题的斱法有两种,其一 是 锁止控制阀 柱塞阀不阀套

更换阀体总成,其二是使用图10中的修

复 斱案,对此阀孔进行铰孔修复,而后配

以 改良型的锁止控制阀和锁止柱塞阀及电磁阀调制阀 阀 套,此处的锁止弹簧不能再使用原件,

而 必须使用重新标定好的弹簧。观察

09G的 这块小阀板,会发现这块阀板在铰图9 09G阀体中的锁止控制阀及其真空测试孔时按 常觃斱法很难固定,因此需要利点

一起了,由此产生的症状有:升挡戒降挡 中的控制滑阀由两部分组成,一个是位于 用图中所 示的阀体固定板通过螺栓将阀

冲击,变扭器及ATF油过热,变扭器中阀孔内侧的锁止控制阀,另一个是位于外 板固定在 万用阀体夹具VB-FIX上。这

锁 止摩擦片受损,车辆低速行驶时发动机侧的带有阀套的锁止柱塞阀及阀套。内侧 种螺栓固定 法对于这种体积较小、难以

转 速出现波动等。 的锁止控制阀容易磨损阀孔内壁,而柱塞 用钳子固定的 小阀板效果特别好。 图9所示为09G的锁止控制阀,它阀则容易磨损阀套内壁,两处的磨损都会 09G阀体中,低里程数下容易失效的

产生类似的故障现象。图9中带网格的位 于0 9G阀体上斱的#2小阀板上。此故障点除了以上介绍的锁止阀孔外,还有

阀孔 区 同一小阀板上的电磁阀调制阀,以及主阀 主调压阀 次级调压阀

板上的K3离合器控制阀孔和位于同一增压阀

孔 内的N90电磁阀。随里程数逐渐增

电磁阀调制阀 加,主 调压阀孔、次级调压阀孔、蓄压器B1制动带控阀 N88 (4-6换挡) K3离合器控制阀 活塞、 K1,K2,B1离合器控制阀孔以及N89(4-6换挡) 其他各 线性电磁阀也会逐渐失效(如图11

所示)。 究其主要原因,还是因为变速器

电控程度 大大提高,电磁阀工作频率增

高,导致滑 阀更频繁地磨损阀体。

从以上所举的几个例子中可以看到

变扭器锁 止故障有可能直接原因来自于

N91(锁止) 变扭器,但是其深层的故障根源往往来自 N93(主油压) K1离合器控制阀 于控制锁止油路的阀体上。阀体上的失效 K2离合器控制阀 点既包括电磁阀,也包括阀孔本身的机械

磨损,维修者在对变速器维修时必须逐一 N282 N283 N90 N92 (K2) (K3) (B1) (K1) 进行检查,找出真正的失效点。

图11 09G主阀板各关键控制(编辑 钟永刚) 件

45 2009/10?

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特别报道“修旧如新”标准高,“修旧如新”标准高, 18 解读车轮轴承

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即使是老化的原因,对我们来说都是有意义的。

维修技巧

40 自动变速器锁止故障——根源何在

自动变速器中变扭器的锁止故障是自动变速器维修过程

中常见故障之一,如何追根溯源,找出故障出现的真正原

因,乃是维修工作根本之所在。本文介绍国内常见自动Z F

5HP19、4T65E、大众09G(TF-60SN)产生锁止故障的根源,以

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范文三：自动变速器驻车锁止机构分析

自动变速器驻车锁止机构分析-权威资料

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摘要:文章介绍了一种自动变速器的驻车锁止机构，对它的结构和设计进行了详细的分析和说明，并阐述了这种驻车锁止机构的工作原理，同时通过设置换挡位置传感器实现换挡功能，提出了它的有益效果以及怎样保证整个机构的可靠性。

关键词:驻车锁止机构;换挡主轴;拨挡片;自动变速器

U463 A 1009-2374(2013)06-0010-03

随着科学技术的不断发展，人们生活水平的不断提高，汽车已经成为人们日常出行必不可少的交通工具。汽车根据换挡方式的不同可以分为手动和自动变速器。自动变速器的车辆由于驾驶简单、使用方便，已经成为了变速器发展的主流方向，越来越受到人们的欢迎。而驻车锁止机构是自动变速器中防止车辆滑行的一种安全装置，使汽车能可靠而无时间限制地停驻在一定位置甚至斜坡上。现在自动变速器的品种繁多，当然驻车锁止机构的结构也不尽相同，下面我们介绍一种自动变速器的驻车锁止机构。

1 驻车锁止机构的结构和设计分析

驻车锁止机构由驻车驱动机构、驻车拨杆总成、拨挡片、换挡固定块总成和换挡摇臂总成组成。整个驻车锁止机构如图1所示。

2.换挡板;3.换挡摇臂?;4.轴套;5.换挡拨杆;6.扭转弹簧;7.换挡销;8.锁止块;9.换挡拨销;10.回位弹簧;

11.换挡固定块;12.拨挡片;13.卡套;14.换挡摇臂?;15.换挡连接销;16.锁止轮。

驻车驱动机构是整个机构的核心部件，包括换挡主轴、换挡板、换挡摇臂?、两件轴套，如图2所示。本文设计在变速器壳体上有前后两个安装孔作为换挡主轴的两个支点，其中换挡主轴的一端设有一环形槽，安装一螺钉固定其轴向位置，换挡主轴与安装孔为间隙配合，保证径向可转动。安装后，换挡摇臂?和换挡板成一定角度，这样布置是因为换挡主轴旋转时，它们可以驱动不同位置上的零件，有利于节省内部空间，降低制造

成本。

驻车拨杆总成包括换挡拨杆、换挡拨销、回位弹簧，如图3所示。设计时考虑结构上的紧凑，在换挡拨杆上设有四个凸台和一个圆台。其中两个凸台用于连接换挡摇臂?，而另两个凸台和一个圆台用于固定连接换挡拨销和回位弹簧，换挡拨销与换挡拨杆之间为滑动间隙配合。回位弹簧在装配时，要有一定的预压缩。在驻车时，换挡拨销推动锁止块，回位弹簧被压缩，复位时，回位弹簧通过预压缩和运动过程中的形变产生的回复力可使换挡拨销恢复到先前的位置。

拨挡片是整个机构中非常重要的零件，与换挡固定块总成固定连接。工作时，换挡主轴旋转带动换挡板上下往复运动，使拨挡片可以在其挡位位置P、R、N、D和L之间相互切换，换挡板挡位位置如图2所示。拨挡片是一个弹性元件，其弹力可以保证每一个挡位入挡的准确性。

换挡固定块总成包括换挡固定块、扭转弹簧、换挡销、锁止块，如图3所示。换挡固定块总成与变速器壳体固定连接，换挡销与换挡固定块固定连接，扭转弹簧和锁止块依次套装在换挡销上。本文设计扭转弹簧要有1/4圈的预紧，在退出P挡时，以保证足够的回复力让锁止块复位。安装时，扭转弹簧的一端卡在换挡固定块的圆弧面上，另一端卡在锁止块的内侧，扭转弹簧的扭力方向为锁止块脱开与锁止轮啮合的方向。

换挡摇臂总成包括卡套、换挡摇臂?、换挡连接销。卡套和换挡连接销均与换挡摇臂?固定连接。安装时，换挡摇臂总成的一侧与换挡主轴相连，另一侧与汽车换挡拉索相连。

2 驻车锁止机构的原理和功能分析

一般来说，自动挡汽车主要设有P挡、R挡、N挡、D挡和L挡等挡位，其中，P挡即驻车挡，就是利用驻车锁止机构运转与装在传动轴上的锁止轮咬合而达到驻车的目的。当汽车需要在一固定位置上停留一段较长时间或在停靠之后离开车辆前，应拉好手制动及将选挡杆推进P挡位置。

本文所描述的驻车锁止机构工作原理是:当驾驶员将选挡杆拨入驻车挡时，汽车换挡拉索拉动换挡摇臂总成，使驻车驱动机构转动，此时，换挡主轴同时给换挡板与换挡摇臂?施加一个径向力，使拨挡片的圆弧卡扣面处于拨挡板的P挡位置，而换挡摇臂?则带动驻车拨杆总成沿水平方向向前运动，换挡拨销通过圆锥面推动锁止块插入锁止轮的凹槽内，从而实现P挡驻车。

退出P挡的运动过程:驻车驱动机构工作，换挡主轴带动换挡板回转，使拨挡片切换到拨挡板的R挡位置时，换挡摇臂?则给驻车拨杆总成一个向后的拉力，使驻车拨杆总成向后运动，而锁止块在扭转弹簧回复力的作用下，脱离了锁止轮的凹槽，即退出P挡。本文所描述的驻车锁止机构中，在驻车驱动机构与换挡摇臂总成的连接处还设置了一个挡位位置传感器(图上未注出)，它与变速器壳体固定连接，它通过换挡主轴的转动角度，可以感应到拨挡片处在换挡板P、R、N、D、L中每个挡位的挡位位置，并由挡位位置传感器将信号准确无误地输送给汽车的电子控制系统，然后电子控制系统下指令并执行换挡程序，最后通过变速器换挡执行机构工作实现自动换挡。

3 驻车锁止机构的有益效果

本文所描述的驻车锁止机构不仅具有驻车功能，还设有拨挡片、换挡板以及挡位位置传感器，具有换挡功能，使变速器内部布置更加合理和紧凑，而且在变速器的壳体上还设有与

换挡主轴间隙配合的安装孔，简化了驻车锁止机构的装配，降低了制造成本，同时拨挡片作为弹性元件，其弹力可以保证每个挡位入挡的准确性，也大大提高了整个机构的可靠性。

4 结语

近年来，汽车工业飞速发展，应用于汽车自动变速器上的新技术层出不穷，本文所分析的驻车锁止机构只是其中一项，经过数十万次的实验，证明该机构可靠性极强，没有出现过任何故障。

作者简介:孙哲(1978-)，男，山东汶上人，宁波威能传动技术股份有限公司项目部经理，机械工程师(中级)，研究方向:汽车自动变速器。

(责任编辑:黄银芳)

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范文四：自动变速器锁止离合器系统

A Study on Shudder in Automatic TransmissionLock-up Clutch Systems and Its Countermeasures

Takahiro Ryu

Oita University

2011-01-1509

Published 05/17/2011

Kenichiro Matsuzaki

Kyushu University

Takashi Nakae

Toyama National College of Technology

Atsuo Sueoka

Kyushu University

Yoshihiro Takikawa and Yoichi Ooi

Aisin AW Co.,LTD

Daisuke Nakagoshi and Yuichiro Hirai

Oita University

Copyright ? 2011 SAE International

ABSTRACT

In recent years, automatic transmissions have become widelyused in cars. Compared to manual transmissions, automatic transmissions suffer from poor fuel economy. In order to overcome this disadvantage, a lock-up clutch system in the torque converter has been applied. When the rotating speed ofthe turbine approaches that of the pump, the input shaft is directly connected to the gear train through friction by meansof the lock-up clutch. In the process of slipping at the lock-upclutch, frictional vibration referred to as shudder sometimes occurs. When shudder occurs, the power train, as well as thetires and the car seats, vibrates. Therefore, the shudder adversely affects passenger comfort. In the present study, experiments are conducted to analyze the shudder mechanismusing a bench test apparatus and an actual vehicle. The characteristics of friction in the lock-up clutch is found to have a negative slope with respect to the relative slip velocity. The entire system from the piston to the tires, including planetary gears, is modeled as a linear multi-degree-of-freedom system using a Lagrange equation of

motion, and the shudder mechanism is clarified. A countermeasure for suppressing shudder is also proposed. The most effective position at which to add the viscous damping and the countermeasure to suppress the shudder byattaching a dynamic absorber are analyzed. Numerical analysis reveals that (1) applying dampening to the dampersis the most effective method by which to suppress the shudder and (2) the shudder can be completely suppressed byattaching a dynamic absorber. Moreover, experiments using adynamic absorber are conducted using an actual vehicle, andthe suppressive effect of the dynamic absorber on the shudderis confirmed through experiments.

INTRODUCTION

In recent years, automatic transmissions have become widelyused, and the number of cars that have automatic transmissions is increasing. Compared to manual transmissions, automatic transmissions suffer from poor fueleconomy. In order to overcome this disadvantage, a lock-up

clutch system has been developed for application in a torque

converter. The lock-up clutch system is composed of a pistonwith friction material and dampers. The dampers are in the form of springs that support the piston. When the rotating speed of the turbine approaches that of the pump (inputshaft), the piston is pressurized by oil pressure. Transmissionof power from the input shaft to the piston by friction is thenstarted, and the input shaft is directly connected to the piston.In the process of slipping at the lock-up clutch, frictional vibration referred to as shudder sometimes occurs. Shudder isa type of self-excited frictional vibration and torsional vibration in the power train. When shudder occurs, the powertrain, as well as the tires and car seats, vibrates. Therefore, the shudder adversely affects passenger comfort.

the generation mechanism of shudder has not yet been clarified. As countermeasures to suppress shudder, only improvements in automatic transmission fluid (ATF) and the friction material have been implemented, and studies on eliminating shudder by effectivedesign have not been conducted.

In the present study, experiments using an actual vehicle to clarify the characteristics of shudder and experiments using atest rig to investigate the friction characteristics are conducted. Theoretical analysis using a model of the entire car system, from the piston to the tires, including planetary gears, is also conducted. Moreover, as a countermeasure to suppress shudder, the attachment of a dynamic absorber to the lock-up clutch is suggested. The effect of a dynamic absorber is confirmed through numerical calculation and experiments.

OUTLINE OF THE AUTOMATIC

TRANSMISSION AND SUMMARY OFSHUDDER

The automatic transmission consists of a torque converter anda gear train. The torque converter is located between the engine and the gear train and transmits the torque induced bythe engine to the gear train. The inside of the torque converteris filled with ATF. shows an outline of the torque converter. The torque converter is doughnut shaped and consists of three pump impellors, a turbine, and a stator. Asthe torque converter transmits torque through the fluid, the rotating speed of the turbine is slower than that of the pump,which causes the inefficiency associated with the torque converter. In order to overcome this disadvantage, a lock-upclutch, which connects the input and output sides through

friction, is used.

Figure 1. Diagram of the torque converter.

material is adhered to the piston. The piston is supported byreduce the fluctuation in rotating speed caused by explosionsin the engine. The friction material is compressed by oil pressure on the converter cover with friction in the circumferential direction, directly connecting the input and output sides.

A traveling test using an actual vehicle was conducted in order to investigate the occurrence of shudder. The characteristics of the shudder are summarized as follows.(1). The vibration frequency of the shudder is 32.7 Hz and isindependent of the frequency of the forced torsional vibrationdue to explosions in the engine. The frequency is

approximately constant even when the gear speed is changed.(2). Shudder occurs when the difference in rotating speed isapproximately 60 to 100 rpm just after lock-up operation.(3). Shudder is a torsional vibration throughout the drivetrain. The phase difference between the input, the output, andthe propeller shaft is in-phase.

(4). Vibration with a frequency of approximately 10 Hzrarely occurs.

Shudder is a torsional vibration in the power train and the tires, which causes vibration of entire car and reduces

passenger comfort.

Figure 2. Lock-up clutch and friction material.

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Figure 3. Lock-up damper.

shows the frequency analysis of the fluctuation of the rotating speed at the propeller shaft when shudder occurred. The frequency component of 32.7 Hz is dominant.As a preliminary experiment, a hammering test of the pistonin the rotational direction under constant torque was performed. The natural frequency of the piston supported bythe springs (dampers) in the rotational direction was found tobe approximately the same as the shudder frequency, which

implies that the piston is a primary source of shudder.

Figure 4. Frequency analysis of shudder at the propeller

shaft.

Figure 5. Waveform of shudder vibration.

shows the waveform of the fluctuation of the rotating speed of the propeller shaft. The waveform of the

shudder is extracted by the band-pass frequency filter. The figure shows that the amplitude of the shudder increases exponentially at the early stage of shudder growth. In order toexamine the growth rate of the shudder, the envelope curve fitting of the waveform by the exponential function at that stage is applied, and the value that corresponds to the real part of the eigenvalue was calculated. The value was approximately 4.0.

In order to examine the friction characteristics, basic experimental equipment using only a torque converter was constructed. The frictional torque was measured by changingthe slip velocity of the converter cover. The results are shownin The abscissa represents the relative slip velocity, and the ordinate represents the coefficient of friction. The velocity range between two dash-dotted lines represents the timing of the beginning of the lock-up operation. Experiments were performed at four oil pressures. As shownin the figure, the coefficient of friction has a negative slopewith respect to the relative slip velocity. When the system exhibits a coefficient of friction with a negative slope, the system has negative resistance, that is, negative damping, and

so shudder can be generated.

Figure 6. Characteristics of friction.

MODELING AND THEORETICALANALYSIS

In order to clarify the generation mechanism and investigate acountermeasure to suppress shudder, a system based on the experimental results is modeled. Since the amplitude of the shudder at the input shaft was quite small in the experiment,coupling of the engine side is neglected in the modeling. Therefore, the components from the lock-up clutch to the tires are considered in the modeling. Since the frequency ofthe shudder was constant and unrelated to the frequency caused by the explosions in the engine, the forced vibrationcaused by the explosions is neglected.

In the actual vehicle used in the present study, the shudder occurred in the fifth gear of a six-speed transmission. Therefore, only the fifth gear is considered in the modeling.shows the arrangement of the gears train in fifth

Figure 7. Gear train in fifth gear.

Figure 8. Analytical model of shudder.

gear. There are three planetary gears in the gear train used inthe experiments, referred to as the front (Fr), middle (Mid),and rear (Rr) planetary gears.

A planetary gear consists of a sun gear, a carrier, and a ringgear. When the rotating speeds of two of the three components of the planetary gear are decided, the rotating speed of the other component is set. Considering the effect ofthis mechanism of the planetary gears on the natural frequencies and natural modes, the planetary gears are

modeled exactly in this theoretical analysis. Moreover, the shafts in the gear train are modeled by the rotational springs.The black circles, other than that at the dampers, represent thepositions of the shafts modeled by rotational springs.shows the analytical model used to analyze the shudder. The blue lines, the black thick lines and the red arrows represent the springs, rigid connections and gear connections. In the figure, θ1,…,θ11 are the

angular

displacements of the components, k 1, …, k 12 are the rotationalsprings, and n 1,…,n 5, p , q , r are the constants that represent therear ratios between the components.

The equations of motion are given by the Lagrange equationof motion. The kinetic energy of the entire system T is

represented as

(1)

The potential energy of the entire system U

is represented as

(2)

where

.

The coefficient of friction is assumed to be a linear functionhaving negative slope with respect to the relative skip

velocity:

(3)

where μ is the coefficient of friction, F is the normal force onthe frictional surface, R is the effective radius of the frictionalsurface, V is the representative velocity of the converter coverat the frictional surface (constant), and μ0 and μ1 are constants. The frictional moment exerted on the piston is μFr.Rotational viscous damping is applied in parallel to the rotational spring. A Lagrange equation of motion is applied using and the variations of the angular displacements are newly defined as θ1,…,θ11. In addition, theequations of motion of an 11-degree-of-freedom system are

given by

(4)

where M , C , and K are the mass, damping, and stiffness matrices, which are symmetric matrices. Defining the elements of these matrices as M ij , C ij , K ij (i , j = 1,…,11), the

elements other than zero are represented as follows:

(5)

(6)

The elements of the damping matrix are obtained by changing the k S in the elements of K to c S , except for K 11 andk?. C11 = c 1 - μ1r 2F , and c? is assumed to be αk? (α = constant).There exists negative damping with respect to the negative slope of the characteristics of friction in element C 11. For the case in which a dynamic absorber with inertia moment J d , damping coefficient c d , and spring coefficient k d is attached, the amendment of the equation of motion of the

absorber are represented as

(7)

where θd is the angular displacement of the dynamic absorber.

Table 1. Natural frequencies.

Table 2. Natural modes.

RESULTS OF NUMERICAL

CALCULATIONS AND DISCUSSIONSNATURAL FREQUENCY ANDNATURAL MODE

shows the natural frequencies f n when the damping and coefficients of friction are set to zero. The natural frequency of the second mode is 35.0 Hz, which is close tothe frequency of the shudder. shows the natural show the mode shapes of the 1st and 2nd modes, respectively.In the 1st mode, the components from the piston to the differential gear vibrate as a rigid body that is supported bythe drive shaft. In the 2nd mode, the piston exhibits the mostsignificant vibration. The components from the TC to the wheels vibrate in phase. These characteristics coincide with the experimental results. In the 3rd and 4th modes, the wheelsexhibit the most significant vibration. In the modes above the

4th mode, the gears exhibit the most significant vibration.

Figure 9. Mode shape (a) 1st mode (b) 2nd mode.

COMPLEX EIGENVALUE ANALYSIS

Considering the negative slope of the characteristics of friction and the system damping, a complex eigenvalue analysis was performed. The characteristics of friction shownμ1 is set tobe 0.011 s/m. The damping coefficient is set to be 0.01 in damping ratio composed by the local element around the damping element. The value α is assumed to be c 2/k 2. The value c 1 is determined under the condition whereby the realpart of the eigenvalue of the 2nd mode is 4.0. According tothe results of the complex eigenvalue analysis, the 1st and 2nd modes were unstable. This result coincides with the experimental results. However, in this analysis, since the case

Table 3. Results of eigenvalue analysis.

in which only the shudder mode occurs is considered, the value α is changed to 3.6 × 10?5 in order to stabilize the 1stmode. shows the results of the complex eigenvalue analysis.

EFFECT OF INTERNAL DAMPING

As a countermeasure against shudder, the effect of the internal damping, which is the damping between the inertia elements, on the shudder is investigated. and show the real part of the eigenvalue for the case in which one of the damping of c 1, c 2, c 8, and c? is made to be s times as large as the fundamental value of the damping. Thedamping c? of the drive shafts is effective for stabilizing the first mode, and the damping c 1 of the dampers is effective forstabilizing the 2nd mode. These dampers are located in the large relative displacement between the elements in each

Figure 10. Effect of internal damping on instability (a)

1st mode (b) 2nd mode.

EFFECT OF EXTERNAL DAMPING

A damping c? is added to TC (θ2) as a base support element,which simulates the viscosity of ATF. The fundamental valueof the damping c? is set to the same value as c 2. shows the real part of the eigenvalue of the 1st and 2nd modes for the case in which c? is made to be to s times as large as the fundamental value of the damping. As shown inthe figure, although the external damping c? is effective for the

1st mode, it is not effective for the 2nd mode.

Figure 11. Effect of external damping on instability.

COUNTERMEASURE USING ADYNAMIC ABSORBER

The effectiveness of the dynamic absorber in suppressing theshudder is investigated. Although the optimum design of thedynamic absorber has been established in the field of forcedvibrations, the optimum design of the dynamic absorber for self-excited vibrations has not yet been established. The authors have investigated the optimum design of the dynamicabsorber for completely suppressing disc brake squeals in of the mechanisms of frictional vibrations were identified. One is asymmetry of the stiffness matrix, and the other is negative slope of the characteristics of friction. The authors also clarified that the optimum design of the dynamic absorber varies depending on the mechanism of frictional vibrations. Based on these findings, the design of the dynamic absorber to completely suppress the shudder is investigated, where the 2nd mode is the target for of suppression.

Since the shudder has a vibration mode in which primarily the piston vibrates, the case in which the dynamic absorber is

attached to the piston is considered. The inertial moment ofthe dynamic absorber J d is assumed to be sufficiently small,so as not to influence the main system. In this calculation, themoment of inertia of the dynamic absorber J d is set to be 3%of the moment of inertia of the piston J 1. The damping ratioζd and the natural frequency f d of the dynamic absorber are

defined as follows:

(8)

of the natural frequency of the dynamic absorber to the shudder frequency (2nd mode), denoted as η, is changed. Loci A and B represent the real parts of the eigenvalues of 2nd and 1st modes, respectively. The results for η = 0represent the case in which the dynamic absorber is not attached. At locus A, the real part of the eigenvalue decreasedwhen η ≈ 1. In this condition, the system is stabilized, that is,the vibration of the system including the dynamic absorber iscompletely eliminated. At locus B, the real part of the eigenvalue decreased when η ≈ 0.39. The natural frequency of the dynamic absorber is approximately equal to the naturalfrequency of the 1st mode in this range. Even though the suppressive effect of the dynamic absorber for the 1st mode isweaker than that for the 2nd mode, the dynamic absorber contributes to stabilize the 1st mode.

shows the relationship between the natural frequency and the damping ratio of the dynamic absorber. The white egg-shaped area is the stable area, and the black area is the unstable area. If the parameter of the dynamic absorber is designed to be in the white area, the shudder willbe completely suppressed. That is, by setting the natural frequency of the dynamic absorber to be close to the shudderfrequency and adding relatively large viscous damping, complete suppression of the shudder is realized. This is the method used to design a dynamic absorber to suppress the frictional vibration caused by the negative damping with respect to the relative slip velocity. If a dynamic absorber thathas a larger inertia moment is used, the stable area will become large. In designing a dynamic absorber to suppress the frictional vibration caused by the asymmetry of the stiffness the stable region is spread in the low-damping area. Since the design method differs according to the mechanism of frictional vibration, it is important to determine the mechanism of frictional vibration when

designing the dynamic absorber.

Figure 12. Effect of the dynamic absorber on shudder.

Figure 13. Vibration suppression area generated by the

dynamic absorber.

VIBRATION SUPPRESSION

EXPERIMENT USING A DYNAMICABSORBER

Based on the numerical calculations, a dynamic absorber is attached to the automatic transmission and an experiment using an actual vehicle is performed. shows the mounting position of the dynamic absorber. The dynamic absorber is of the torsional type and is composed of steel andrubber. The rubber is used as a base support element, whichexhibits stiffness and damping. The steel provides an inertiamoment. Considering the factor of safety, the ratio of the inertia moment of the steel to that of the piston is set to 10%.The natural frequency of the dynamic absorber is tuned to besimilar to the shudder frequency, and the damping ratio is approximately 0.1.

speed at TC in the experiment for the case in which the dynamic absorber is not attached and the case in which the optimum dynamic absorber is attached, respectively. In the traveling test, the shudder is completely suppressed by attaching the dynamic absorber. When the frictional self-excited vibrations are suppressed by the dynamic absorber, neither the entire system nor the dynamic absorber vibrate.

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Figure 14. Mounting position of the dynamic absorber.

Figure 15. Effect of the dynamic absorber on shudder (a)Waveform of shudder (b) Waveform when the dynamic

absorber is attached.

CONCLUSIONS

In order to clarify the shudder mechanism and investigate a countermeasure to suppress the shudder, an experiment usingan actual vehicle and theoretical analysis were performed.

Moreover, a countermeasure using a dynamic absorber was investigated. The obtained results are as follows:

(1). The shudder occurs only in fifth gear when the difference in the rotating speed is approximately 60 to 100rpm just after the lockup operation. The vibration frequencyof the shudder is 32.7 Hz, and the frequency is similar to thenatural frequency of the piston supported by the dampers.(2). The characteristics of friction of the lock-up clutch exhibits a negative slope with respect to the relative slip velocity.

(3). According to the numerical analysis, only the second mode is unstable. The piston vibrates primarily in the unstable mode. The 1st mode can be generated if the systemdamping is reduced.

(4). Adding an internal damping to the dampers is effectivefor suppressing the shudder, whereas adding an internal damping to the drive shaft and an external damping to the turbine are effective for suppressing the 1st mode shudder.(5). The analytical results obtained using a dynamic absorberas a countermeasure against shudder revealed an optimum design for the dynamic absorber at the natural frequency anddamping ratio. In the theoretical analysis, the shudder was completely suppressed by attaching the optimum dynamic absorber, and the results were confirmed experimentally.

REFERENCES

1. Ogawa, H., Hayashi, E., Yoshida, O., Hayakawa, Y.,Kuroyanagi, T., Ishikawa, K., Takabayashi, H., Maruo, K.and Fujii, T., “Mechanism of Shudder Phenomena in TorqueConvertor and System Simulation Model,” SAE Technical2. Kato, Y., Akasaka, R. and Yamazaki, T., “Study on Originof Shudder Under Slip Control of a Lock-Up Clutch in anAutomatic Transmission - (part 1) Measurement of FrictionCharacteristics Under Shudder,” Journal of Japanese Societyof Tribologists, 42, 485-491, 1997.

3. Centea, D., Rahnejat, H. and Menday, M.T., “Non-LinearMulti-body Dynamic Analysis for the Study of ClutchTorsional Vibrations (Judder),” Applied MechanicalModelling , 25, 177-192, 2001.

4. Yoshimura, N., “Effect of Temperature Dependency ofLock-Up Clutch Friction Characteristics on Shudder

Vibration,” Proceedings of JSAE Annual Congress (Fall), 9538933, 69-72, 1995.

5. Itoh, J., Ishihara, N., Suzuki, M., Kayama, S., Shimizu, N.and Sogabe, K., “Study for Judder Analysis of Lock-UpClutch - Part 1: Analysis Theory and Simulation of JudderPhenomena,” Proceedings of JSAE Annual Congress (Fall), 9538942, 73-76, 1995.

6. Miyoshi, M., Suzuki, M., Kayama, S., Sogabe, K. andShimizu, N., “Study for Judder Analysis of Lock-Up Clutch -

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Part 2: Parameter Study and Analysis of Slip Shudder,”Proceedings of JSAE Annual Congress (Fall), 9637122,141-144, 1996.

7. Nakano, K. and Maegawa, S., “Safety-Design Criteria ofSliding System for Preventing Friction-Induced Vibration,”Journal of Sound and Vibration, 324, 539-555, 2009.8. Kugimiya, T., Mitsui, J., Yoshimura, N., Kaneko, H.,Akamatsu, H., Ueda, F., Nakada, T. and Akiyama, S.,“Development of Automatic Transmission Fluid for Slip-Controlled Lock-Up Clutch Systems,” SAE Technical Paper9. Tohyama, M., Ohmori, T. and Ueda, F., “Anti-ShudderMechanism of ATF Additives at Slip-Controlled Lock-Up10. Lam, R.C., Chavdar, B. and Newcomb, T., “NewGeneration Friction Material and Technology,” SAE11. Nakano, Y., Nakae, T., Ryu, T., Yamasaki, T. and

Sueoka, A., “Experimental Investigation on Countermeasureagainst Squeal Phenomena in a Car Disk Brake by a DynamicAbsorber,” Proceedings of the 12th Asia-Pacific VibrationConference , CD-ROM, 2007.

12. Nakano, Y., Sueoka, A., Ryu, T., Iefuji, H., Orii, T. andNakae, T., “Countermeasure and Analysis of Squeal inBicycle Disk Brakes,” Transactions of Japan Society ofMechanical Engineers, 72-720, 2373-2381, 2006.

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范文五：自动变速器驻车锁止机构分析

自动变速器驻车锁止机构分析

北京京坛医院 http://www.taiji886.com/

摘要:文章介绍了一种自动变速器的驻车锁止机构，对它的结构和设计进行了详细的分析和说明，并阐述了这种驻车锁止机构的工作原理，同时通过设置换挡位置传感器实现换挡功能，提出了它的有益效果以及怎样保证整个机构的可靠性。

关键词:驻车锁止机构;换挡主轴;拨挡片;自动变速器

随着科学技术的不断发展，人们生活水平的不断提高，汽车已经成为人们日常出行必不可少的交通工具。汽车根据换挡方式的不同可以分为手动和自动变速器。自动变速器的车辆由于驾驶简单、使用方便，已经成为了变速器发展的主流方向，越来越受到人们的欢迎。而驻车锁止机构是自动变速器中防止车辆滑行的一种安全装置，使汽车能可靠而无时间限制地停驻在一定位置甚至斜坡上。现在自动变速器的品种繁多，当然驻车锁止机构的结构也不尽相同，下面我们介绍一种自动变速器的驻车锁止机构。

1 驻车锁止机构的结构和设计分析

驻车锁止机构由驻车驱动机构、驻车拨杆总成、拨挡片、换挡固定块总成和换挡摇臂总成组成。整个驻车锁止机构如图1所示。

2.换挡板;3.换挡摇臂?;4.轴套;5.换挡拨杆;6.扭转弹簧;7.换挡销;8.锁止块;9.换挡拨销;10.回位弹簧;11.换挡固定块;12.拨挡片;13.卡套;14.换挡摇臂?;15.换挡连接销;16.锁止轮。

驻车驱动机构是整个机构的核心部件，包括换挡主轴、换挡板、换挡摇臂?、两件轴套，如图2所示。本文设计在变速器壳体上有前后两个安装孔作为换挡主轴的两个支点，其中换挡主轴的一端设有一环形槽，安装一螺钉固定其轴向位置，换挡主轴与安装孔为间隙配合，保证径向可转动。安装后，换挡摇臂?和换挡板成一定角度，这样布置是因为换挡主轴旋转时，它们可以驱动不同位置上的零件，有利于节省内部空间，降低制造

成本。

驻车拨杆总成包括换挡拨杆、换挡拨销、回位弹簧，如图3所示。设计时考虑结构上的紧凑，在换挡拨杆上设有四个凸台和一个圆台。其中两个凸台用于连接换挡摇臂?，而另两个凸台和一个圆台用于固定连接换挡拨销和回位弹簧，换挡拨销与换挡拨杆之间为滑动间隙配合。回位弹簧在装配时，要有一定的预压缩。在驻车时，换挡拨销推动锁止块，

回位弹簧被压缩，复位时，回位弹簧通过预压缩和运动过程中的形变产生的回复力可使换挡拨销恢复到先前的位置。

拨挡片是整个机构中非常重要的零件，与换挡固定块总成固定连接。工作时，换挡主轴旋转带动换挡板上下往复运动，使拨挡片可以在其挡位位置P、R、N、D和L之间相互切换，换挡板挡位位置如图2所示。拨挡片是一个弹性元件，其弹力可以保证每一个挡位入挡的准确性。

换挡固定块总成包括换挡固定块、扭转弹簧、换挡销、锁止块，如图3所示。换挡固定块总成与变速器壳体固定连接，换挡销与换挡固定块固定连接，扭转弹簧和锁止块依次套装在换挡销上。本文设计扭转弹簧要有1/4圈的预紧，在退出P挡时，以保证足够的回复力让锁止块复位。安装时，扭转弹簧的一端卡在换挡固定块的圆弧面上，另一端卡在锁止块的内侧，扭转弹簧的扭力方向为锁止块脱开与锁止轮啮合的方向。

换挡摇臂总成包括卡套、换挡摇臂?、换挡连接销。卡套和换挡连接销均与换挡摇臂?固定连接。安装时，换挡摇臂总成的一侧与换挡主轴相连，另一侧与汽车换挡拉索相连。

2 驻车锁止机构的原理和功能分析

一般来说，自动挡汽车主要设有P挡、R挡、N挡、D挡和L挡等挡位，其中，P挡即驻车挡，就是利用驻车锁止机构运转与装在传动轴上的锁止轮咬合而达到驻车的目的。当汽

车需要在一固定位置上停留一段较长时间或在停靠之后离开车辆前，应拉好手制动及将选挡杆推进P挡位置。

本文所描述的驻车锁止机构工作原理是:当驾驶员将选挡杆拨入驻车挡时，汽车换挡拉索拉动换挡摇臂总成，使驻车驱动机构转动，此时，换挡主轴同时给换挡板与换挡摇臂?施加一个径向力，使拨挡片的圆弧卡扣面处于拨挡板的P挡位置，而换挡摇臂?则带动驻车拨杆总成沿水平方向向前运动，换挡拨销通过圆锥面推动锁止块插入锁止轮的凹槽内，从而实现P挡驻车。

退出P挡的运动过程:驻车驱动机构工作，换挡主轴带动换挡板回转，使拨挡片切换到拨挡板的R挡位置时，换挡摇臂?则给驻车拨杆总成一个向后的拉力，使驻车拨杆总成向后运动，而锁止块在扭转弹簧回复力的作用下，脱离了锁止轮的凹槽，即退出P挡。本文所描述的驻车锁止机构中，在驻车驱动机构与换挡摇臂总成的连接处还设置了一个挡位位置传感器(图上未注出)，它与变速器壳体固定连接，它通过换挡主轴的转动角度，可以感应到拨挡片处在换挡板P、R、N、D、L中每个挡位的挡位位置，并由挡位位置传感器将信号准确无误地输送给汽车的电子控制系统，然后电子控制系统下指令并执行换挡程序，最后通过变速器换挡执行机构工作实现自动换挡。

3 驻车锁止机构的有益效果

本文所描述的驻车锁止机构不仅具有驻车功能，还设有拨挡片、换挡板以及挡位位置传感器，具有换挡功能，使变速器内部布置更加合理和紧凑，而且在变速器的壳体上还设有与换挡主轴间隙配合的安装孔，简化了驻车锁止机构的装配，降低了制造成本，同时拨挡片作为弹性元件，其弹力可以保证每个挡位入挡的准确性，也大大提高了整个机构的可靠性。

4 结语

近年来，汽车工业飞速发展，应用于汽车自动变速器上的新技术层出不穷，本文所分析的驻车锁止机构只是其中一项，经过数十万次的实验，证明该机构可靠性极强，没有出现过任何故障。

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