车辆设计外文翻译

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624 Race Car Vehicle Dynamics
Besides providing linear motion the restraint device must provide lateral force reactions between the sprung and unsprung masses. In so doing it is most desirable that the forces are transmitted only in a purely lateral sense, with no vertical componen t. For example, if a Panhard bar (track bar) is utilized, the slope of the bar in the rear view dictates the
force coupling. If it is horizontal the coupling is zero. If it is sloped the coupling will either lift the sprung mass or pull it down depending on the direction of cornering and the slope of the bar. Some considerations in the design of a track bar include whether it is normally in tension or compression. For circle track racing, always turning left means
that the bar should be attached to the body on the right side and to the axle on the left as suring that it is always in tension when cornering.
The pair of arms, the A-arm, and the sliding pin tend to have very low
vertical-lateral coupling. The Panhard bar and the Watts linkage will always have some, but with attention to detail it can be minimized. This coupling effect must be controlled as much as possìble because it tends to vary with the ride height and roll angle positions of the suspensìon.
By keeping the angle small, the changes will be small and the effect minimized.
17.5 Front Suspensions
lntroduction
Many types of front suspensions have been used over the years. They include various be am type axles with steering via kingpins at each end of the axle, the parallel trailing arm type such as the VW, the Morgan sliding pillar type, and the Chevrolet Dubonne t. In recent history, passenger car designs have come down to basically two types: the
MacPherson Strut and the SLA (Short-Long-Arm).
This chapter will deal only with the last two mentioned as these make up the majority of front suspensions that will be encountered. The other types suffer from either high bending loads, poor geometry, high friction，or a combination of these problems. The best way to discuss each type is to go through the design process step by step. For each step a
decision has to be made that is often a compromise. By discussing these decisions, hopefully a feeling for the limitations of the design will develop.
Front Suspension Design lssues-General
The frist task in designing a front suspension of any type is to establish the packaging pa rameters that are fix ed，or absolutely cannot be changed for whatever reason (see Figure17.17). These should be listed so that they are not ove rl ooked. The next task is to package
the wheel, tire, brakes, and bearings. This is done in car position, so the track width has to be known. If it is not yet established, it should be made as wide as practica l. This sounds evasive, but there are trade-offs in everything, even things as simple as choosing the track width. For ex am ple, what do the rules allow? Wh at is the predominant race
track type on which the car will run? Is top speed, thus low frontal area important? Are slow-speed tight street circuits of concern? All these issues can affect the decision on the basic track width!
Tire size and rim di am eter and width must be settled. The specific wheel manufacturer needs to be known and a cross section of the wheel is desirable for optimizing the use of that whee l. Tire sizes are usually limited by the sanctioning body rules. In general, use a ll the tire they will Iet you get away with. Another point is to always design for the latest
sizes being developed by the suppliers; this guarantees that the latest compounds and constructions will fit your ca r. Remember, the tire is the single most important chassis component on the car.
The wheel offset is worked out in conjunction with fitting the brake caliper to clear the inside surface of the whee l. Once the caliper is located, this automatically locates the brake rotor. With the rotor location comes the absolute farthest outboard location for the lower ba ll join t. Wheel bearings need to be looked at soon, as ideally they should be located such that the tire center is between the two rows of balls or rollers (to minimize loads on the bearings).
Now that the lower ball joint cross car boundary (l ateral position) has been set, the height of the lower ball joint comes nex t. In production cars it must be above a
5-in. wash rack clearance requirement, but on race cars it should be made as low as possible for structural reasons. Usually there is no rule but some practical considerations such as de fl ated tire ground clearance might be in order. If it is totally inside the wheel all it has to do is clear the wheel and the brake rotor under all travel and load conditions.
The decision about the kingpin angle in the front view is the next order ofbusiness. The issues here become scrub radius, spindle length, and kingpin angle. They are interrelated and a compromise is needed. If you want a certain scrub radius you now have two points established, i.e., the lower ball joint and the ground contact point of the kingpin (set by the scrub radius)-一the kingpin angle becomes fix ed automatically. If you want a certain kingpin angle then the scrub radius will not necessarily be what you wan t. Basically, on rear-wheel-drive cars push the lower ball joint out as far as possible and run a fairly low kingpin angle, less than 80 , and accept the scrub radius that results. If you are dealing with a front-wheel-drive car you must minimize the spindle length and have a negative scrub radius. This may result in a kingpin angle as high as 160 , but you will have to accept it or find another clever way around i t.
Kingpin ang1e affects the performance of the car when the wheels are steered. One concept that should be understood is that the more the kingpin angle the more the car is lifted when it is steered. This is one source of steering returnability, the weight of the car returns the steering to center. The amount the car is lifted is also a function of the spindle length where a longer spindle means more lif t.
The camber of the wheels when steered is a function of the kingpin angle and the caster angle. With no kingpin angle (and no caster angle) there is no camber change with steer lock. As kingpin is added (but st ill no caster) the wheel w ill "lose" camber with steer lock, or in other words it will change in a direction giving positive camber on the outside
wheel.60 As caster is added this modifies the effect of kingpin. With positive caster and no kingpin angle, the whee1 gains negative camber on the outside wheel and positive camber on the inside whee l. Thus caster can add favorable camber angle to the effects of kingpin angle. In other words, the reason that low kingpin angles are desirable is that kingpin angle subtracts from the negative camber gain due to caster on the outside whee l.
The decision on a rack location depends on several packaging factors such as engine location and orientation，front-wheel drive vs. rear-wheel drive, whether it is to be high- or low-mounted, etc. In addition there are performance reasons for choosing the rack location.
First we must assume that every structure is a spring and should be treated as such. As an example the rack mounting stiffness versus the upper or lower control arm mounting stiffness to the chassis will not necessarily be the same. Therefore, when a cornering force is applied, any difference in the lateral displacement of the ball joints in relation to
the tie rod outer pivot w ill cause a steer angle. To assure stability it is better to have lateral force deflection toe-out (l ateral force understeer) rather than toe-in. We can assure that this happens by the proper location of the rack. If a high-mounted rack is required it must be behind wheel center and if it is low-mounted then it must be ahead of wheel cen ter as shown by the shaded areas in Figure 17.17.
Structural requirements for the suspension design must always be considered when packaging each element of the total system. Control arms that have one leg
straight across from the ball joint are superior in system stiffness to arms th at are splayed. Establishing linkage ratios for the spring, shock, and stabilizer bar as close to 1: 1 as possible will provide more direct load paths thus improving system stiffness while providing a lighter overall design.
This is the most common form of front suspension and can be used as a rear suspension very sily.The toe link is grounded to the chassis(instead of attaching to a rack) as shown in Figure17.30(b).If a front suspension were to be used as the rear suspension this style is easy to adapt with one one important consideration.The left front components should be installed as the right rear corner and the right front as the left rear.The reason these are “turned around”is that geometric toe considerations for front suspensions are exactly opposite of what is required for a rear suspension to give roll understeer.
A minor variation of this design is the case where the toe link does not attach to the chassis;rather it is attached to either the upper or the lower control arm (shown to the lower arm in Figure 17.18(c)).This can be contemplated only when the toe link outer pivot is very close to the same height as the ball joint.This is called”an ungrounded”toe link.The ride toe characteristics are not always obvious on this type,especially when there is significant caster change present.For best results a computer simulation should be used.
Upper A-arm with Three Links
Sometimes a lower A-arm is not practical for packaging reasons.An equivalent suspension geometry using three individual straight links in place of a lower A-arm and a toe link along with an upper A-arm can provide very good geometry.The links can be arranged basically in two nontrailing basically lateral/diagonal links and another lateral link acting as a toe link.The trailing link arrangement will have more of a problem getting the toe curve to be linear,and will have other parameters controlled by the side view swing arm that will change with ride travel.This may not be desirable.When the two lateral diagonal links are used,a virtual center (point “A”in Figure17.31(b))is created farther outboard in the wheel than a ball joint can physiically be located.This feature can be an advantage in achieving a negative scrub radius and/or a short spindle length.
Lower A-arm and Three links
A system basically opposite of that described above is also a viable suspension in the SLA family.In this case the upper arm is formed by two of the links.Again a virtual center may be created to achieve a particular feature that is unobtainable with a ball jointed A-arm.The use of one link trailing and the other straight lateral is also possible but not generally recommended.
H-arm and A-arm
An H-arm functions as three links in this example and is similar to the A-arm with an ungrounded toe link(seeFigure 17.30(c)).It can be used either as an upper arm or a lower arm in a rear suspension.When it is used,the inner pivot axis of the H-arm must be parallel to the inner pivot axis of the A-arm or else binding will occur with suspension travel due to the attempt to rotate the knuckle(caster change) thus twisting the H –arm.This means the side view swing arm instant center must be at infinity,thus
making it difficult to obtain high values for anti features.This is a limiting factor for the use of H-arms in general.A potential cost advantage of the H-arm is that it requires only two chassis reaction points instead of three to perform its controlling functions.
H-arm lower and camber link
The use of an H-type lower arm and a single lateral upper link is a special case where the H-arm is being asked to perform the function of four links instead of just three.It accomplishes this by reacting all braking loads as a torsional input .Reactions to all wheel center inputs in the longitudinal direction end up putting the H-arm in torsion over its length.The structural requirements of the arm become quite different in this case.The single upper link acts only in the front view and contributes to the determination of the front view instant center.The side view geometry is controlled totally by the H-arm inner pivot axis.It is determined in the same manner as the
semi-trailing arm.A version of this rear suspension is used in the 1989
Thunderbird(see Figure 17.32).
Five-link
A system utilizing five individual links can make a very satisfactory suspension.The placement and orientation is similar to the three-link and A-arm mentioned above with the A-arm now formed by two links;two variations are shown in Figure 17.33 and 17.34.The kinematics are very flexible with this type of design where the issue to get the front view kinematics desired without compromising the side view geometry.Another major reason for doing a five-link is to obtain a short spindle length and a small negative scrub radius.This is done by aiming the links to virtual centers instead of having to package physical ball joints at the desired centers.
A caution in using this design is that the lateral diagonal pairing of upper and lower links to simulate A-arms is preferred to pure lateral and pure trailing links to form the arms.This is because the rate of change of the side view geometry is generally too high with the trailing arm concept.Also a trailing arm must be packaged inboard of the wheel edge which results in a spindle length at least as large as half the tire section width.See Figure 17.34 for an explanation of this concept.Examples that demonstrate the differences between the trailing arm and the lateral diagonal style are the ’84 Corvette versus the Mercedes 190 five-link.The Corvette uses the half shaft and the camber link in the front view along with the two trailing links in the side view to form the upper and lower arms.A tie rod is the fifth link.
17.7Beam Axle Rear Suspensions
As mentioned earlier in the general section on beam-type axles there are five basic kinematic properties that need to be controlled by the suspension links.The wheel path,anti-lift,and anti-squat are controlled by the side view instant center.The roll center height and roll steer are controlled by the roll axis.All loads that occur between the sprung and unsprung masses are reacted by the suspension links.The fore and aft loads such as braking and acceleration are coupled through the side view instant center and the lateral loads are coupled through the roll axis.So the basis for understanding any beam axle rear suspension is to understand how the instant centers and roll axis are defined.This will be covered for some typical rear suspension types that are commonly used.
Four bar links
The very first section of this chapter discussed degrees of freedom and motion path and it was shown that a beam axle suspension has two motion paths,therefore,the motion can be controlled by four links.There are various ways to arrange those four links to achieve a workable suspensions.The list below includes some of the most popular.
Basic four bar link
Figure 17.35 is representative of this style of geometry.The side view instant center is determined by projecting or extending a line through the ends of both the upper and lower arms until they intersect.Normally the intersection will occur 100 in.or so ahead of the wheel center and will have a height somewhere between ground level and the wheel center.This part of the analysis is simple and straightforward and can be done in one view of the design.
The roll axis determination is a little more complicated because it involves looking at things in both the side view and the plan (or top) view.The roll axis is a line connecting the two lateral restraint points.These lateral restraint points were discussed in the general section on beam axles.For the basic four bar link the pair of upper control arms are angled in the plan view and therefore have an intersecting point which is a lateral restraint point ,marked as “A”in the figure.The same is done for the lower pair of control arms,marked as”B”.Now we connect the two lateral restraint point where the roll axis crosses a vertical transverse plane through the wheel centers.The amount of is the slope of the roll axis in the side view.
You will note that in the plan view the upper and lower arm intersection points lie on the centerline of the car.This is because the right and left sides of the suspension are exactly the same.If they were not the same you would still use the same procedures to understand the geometry,but you might end up with a roll axis off center or one that angles in the plan view.This may or may not be desirable depending on the kind of performance that is needed.
When designing a four bar link rear suspension the side view geometry is adjusted by raising or lowering the pivots of the links relative to each other to get them to aim toward the desired instant center.The absolute and relative lengths of the upper and lower arms in the side view affect the rate of change of the side view instant center as well as how the axle housing rotates with wheel travel.This rotation is important to maintaining control over the propshaft joint angles.The change in pinion shaft angle with wheel travel is similar to camber change in the front view of an SLA suspension.The arm lengths and the ratio of their lengths control this.
Front Suspension Design Issues-S LA
The Short-Long Arm (SLA) suspension is the choice of designers without question for its ability to meet desired performance objectives with minimum compromise.
The design starts with the basic package as described above. The details of the track width, the wheel size, the tire, the brakes, etc., bring about the location available for the lower ball join t. The upper ball joint is located either via kingpin angle requirements or by scrub radius requirements. There is a little more freedom with the
SLA that is not available to the strut design and that is the choice of a short knuckle or a tall knuckle.
The short knuckle means the upper ball joint is located basically within the diameter of the whee l. With high offset and large-diameter wheels the kingpin angle can be kept small (while achieving small spindle lengths and scrub radius) by tucking the upper ball joint into the whee l.
To reduce the loads on the control arms and other suspension components, it is desirable to have a long kingpin length, that is, separate the upper and lower ball joints as much as possible. Depending on deta i!s of the installation, the short knuckle may yield less than optimum kingpin length. The other alternative is the tall knuckle where the upper ball
joint is above the tire. In the tall knuckle design the ba ll joints naturally have a very large span and thus reduce reaction loads. This option also allows reasonable kingpin angles while achieving desired spindle length and scrub radius. Another advantage for the tall knuckle is that build errors w ill result in smaller geometry errors than with short knuckle designs. Some negatives to the tall knuckle, of course, are the added structural requirements of the knuckle, and the limitation of never changing tire size or width without widening the track and increasing the spindle length and scrub radius after the design is completed.
With the upper and lower ball joint locations established, the tie rod outer point should also be set per the requirements established in Chapter 19 on steering geometry.
Front View Geometry
The front view geometry can now star t. The front view swing arm instant center is uniquely deternmined by the desired roll center height and roll camber (see Figure 17.18). The desired roll camber sets the front view swing arm length (l ocation of line A-A) as follows:
The front view instant center height is set by projecting a line from the tire center ground contact patch through the desired roll center heigh t. The instant center must lie on this line. Now we can project lines from both ball joints to the instant center. These become the centerlines of the upper and lower control arm planes as projected into the vertical plane through the wheel cente r. Packaging requirements w ill establish the length of the lower control arm but it should be rnade as long as possible. The length of the upper control arm in relation to the lower arm adjusts the shape of the camber curve. If they are the same length the camber versus wheel travel curve w ill be a straight line. If the upper is longer than the lower, the curve w ill be convex with its curvature toward positive camber. If the upper is shorter than the lower, the curve will be concave toward negative cambe r. As the upper is made progressively shorter, the curvature increases. The ideal curve has progressive negative camber in bump with much less camber change in droop. Some designers try to get the camber to go positive in droop and progressively negative in bump.
To finish the front view geometry，the tie rod and rack location should be
roughedin. This is done by projecting a line through the tie rod outer point
(established in Chapter 19 on steering) and the front view instant center. The correct tie rod length is then established for a linear ride toe curve. This length will be modified after the side view geometry is completed, but doing it now is a good idea to help plan a realistic rack location.
624 赛车动力特性
除了提供直线运动外，安全装置还要在安装弹簧和非安装弹簧的部件之间提供侧向反力。

这样做是最可取的，力只在侧面有分力，在垂直方向上没有分力。

例如，如果利用一个潘哈德杆（跟踪条），后视图中杆的倾斜度要求有力耦合。

如果是横向的耦合是零。

如果是倾斜的耦合要么解除簧载质量或压下来取决于转弯的方向和杆的倾斜度。

在一个跟踪条的设计中考虑的因素包括是否通常在拉伸或压缩状态。

对于循环的轨道赛车，总是向左转的条件意味着该杆应该连在车身的右侧和轴的左侧，确保转弯时杆始终在紧张状态。

Ａ型的一对横臂和滑销有非常小的垂直和侧向力偶。

潘哈德杆和瓦特连杆总是有一些力偶，但是注意细节可以把力偶最小化。

这种耦合作用，必须尽可能地控制，因为它往往会随着车身高度和悬架侧倾角位置变化而变化。

通过保持较小的角度，这些变化将很小，而其影响减至最低。

17.5前悬架
介绍
许多类型的前悬架已经使用了数年之久。

他们包括各种梁式轴通过方向盘凭借关键每个后面的轴，两端各轴平行拖臂式如大众,摩根滑动柱式、雪佛兰开。

在最近的历史、客车设计基本上涉及到两种类型
麦弗逊式前独立支撑和航天器(长短臂)。

本章将只处理最后提到的两个,因为这些构成了大部分的遇到的前悬架。

其他类型遭受高弯曲载荷,质量差几何学、高摩擦力,或者一些混合的问题。

最好的办法是去讨论每一个类型的设计通过循序渐进的过程。

每一步的
决定通常是一种妥协。

这些决定,希望探讨一种感觉就是设计的限制将会发展。

前悬架设计的普遍的问题
在设计的任务是任何类型的前悬架建立包装的参数是固定的,或绝对不能改变在不知什么原因的情况下(见图17.17)。

这些应该被列出来,以便他们都没有被忽略。

下一个任务是包装
车轮、轮胎、制动器、轴承。

在汽车位置这样做，所以偏斜的宽度必须认识。

如果它没有建立,应该被制造跟实际的一样。

这听起来有点儿闪烁其词,但也有权衡每一件事,即使事情像选择轨迹宽度一样简单。

例如,这个规则允许吗?什么是主要的关于汽车将要跑的比赛类型?最高的速度,因此低的前面的区域重要吗?低速的紧街电路是关心吗?所有的这些问题会影响决策的基本轨迹宽度!
轮胎的尺寸，环直径和宽度必须被解决。

具体的车轮制造商需要被认识并且轮子上的十字形部件是可取的，作为轮子的最优使用。

轮胎的尺寸通常被制定的车身规则所限制。

一般来说,用所有的轮胎会让你侥幸取走。

另一点是供应商总是设计正在发展的最新的设计尺寸
,这保证了最新的化合物和结构将更适合你的车。

记住,在这个车上，轮胎是最最重要的底盘部件。

车轮补偿被解决，一起来适应刹车测径器去清洗轮胎的内表面，一旦测径器被定位，这个会自动定位刹车转子。

与转子位置是绝对最远的外侧定位是低球接
头。

不久，转盘轴承必须被看着,作为理想,他们应该位于两个成排的球或滚筒(最小化的轴承)的轮胎之间的中心。

现在,较低的球车边(横向位置的十字架)已经确定,高度较低的球接下来被引进，在生产汽车必须位于5英寸以上.满足清洗要求,但在赛车中它应该被制造和车架上的要求一样低。

通常没有规则,但一些实际的考虑,如气胎地上间隙。

如果将它完全放在车轮中，它所有的作用就是清洗轮胎和在路况行驶下的制动轮。

这个关于止推角的决定在前视角下是商业的下一步决定。

这个问题在这里成为刷洗半径、纺锤状的长度和支撑角。

他们都是互相联系的，并且妥协是必要的。

如果你想要一个特定刷洗半径,现在你有两点需要确定,即建立低球接头和地面接触点的支撑(由刷洗半径)-这个止推角被自动的固定，如果你想要一个特定支撑角，然后刷洗半径不一定是你想要的东西。

基本上,在一个后轮驱动车上尽可能快的把低球接头推出并且以较小的角度运行，小于8度，并接受了半径的结果。

如果你正在处理前轮驱动车,你必须减少车辆轴长度和消极刷洗半径。

这可能导致主销角度高达16度,但你将不得不接受它或找到另一个聪明的办法。

主销倾角影响汽车的表现当车轮被操纵时。

这一概念的理解是,当车轮被操纵时主倾角越大，汽车被提升的角度越大。

这是一个返回操纵的来源，汽车的重量转向到中心;如果执行失败,将返回。

车的数量被提升。

在曲面的车轮驾驶时是主倾角和外倾角的作用。

没有主销倾角角度(不)是没有曲面变化与引导锁。

作为支撑加(但仍然没有脚轮)车轮将失去“曲面同锁定甚至换句话说,它将改变方向,给予积极的曲面在外轮上，正如脚轮被补充关键的作用。

用积极的脚轮和没有主倾角，车轮在外轮上获得消极的凸形和在内轮上获得积极的凸形。

因此脚轮能增加有利的外倾角去影响主倾角。

换句话说,究其原因,低的主倾角是理想的因为深入主销从负面曲面角度或得中减去由于脚轮在外面轮上。

在货架位置的这个决定取决于几个因素,如引擎的位置的包装和定位,前轮驱动装置和后轮驱动装置，不管是高-或低输送等。

此外还有选择架位置的原因。

首先,我们必须假设每个结构是一个弹簧,以这个为例,介绍了架子上安装刚度和下限控制臂安装刚度的底盘不一定会是相同的。

因此,当转弯时,任何不同的球接头侧向位移联系到的外支点将要引起一个转角。

保证稳定性最好有侧向受力变形(侧向力)而不是向内转向。

我们能保证这发生在适当的位置的架子上。

如果一个高架是必需的,它一定是车轮中心,如果它是低架，就必须将提前轮中心显示在图17.17阴影区。

悬挂结构要求设计必须要考虑当包装各要素的整个系统。

控制武器,有一支架直对面球接头具有更好的系统刚度,舒展双臂。

为弹簧、震动，稳定平衡杆建立连接比率接近1:1尽可能将会提供更多的直接加载路径以提高系统的刚度，同时提供一个较轻的整体设计。

双A型摆臂和趾型连接臂
这是前悬架最普通的形式，可以很容易用在后悬架。

趾型连接臂是接在底盘上的（代替接在车架上）表现在图17.30（b）.如果一个前悬架被用作后悬架，这种风格很容易适应一个重要的考虑。

左前方的部件应该安在右后角而右前方的安在左后角。

这样调换过来的原因是趾型前悬架的几何学考虑与后悬架的需要恰恰相反能导致侧倾不足转向。

趾型连接臂不能连接到底盘的情况，这个设计有个小变化。

然而它被连接在下面的或较低的控制臂上（如图17.18（c））表示较低控制臂。

这个可以被估计
这个趾型连接臂外部铰链是很接近那个同样高度的球铰。

这个被称为“接地”趾型连接。

这种趾型连接臂特点不总是很明显在这种类型，尤其当有重要改变。

电脑仿真最好的结果应该被使用。

有时一个较低的A型臂为包装的原因是不实用的。

一个等臂悬架几何学用3个单独直连接代替一个较低A型臂。

一个趾型连接连在上A型臂可以提供很好的几何学。

这种连接可以基本被安排成两种结构（看17.3图（a）和(b)），一个后连接和两个侧连接，或两个非后基础的侧斜连接，另一个侧连接相当于一个趾型连接后连接将有很多问题，一个是将弯曲的趾变直，还有一些参数被随着行驶改变的摇臂角度控制，这可能不令人满意，当这两个侧斜连接被使用，一个虚拟的点（如图17.3（b）point”A”）被创造，它比实际连接的球铰更在轮子的外侧，这个特征可以有利于产生一个负面的磨砂半径或短的主轴长度。

一个系统基本是与上面描述的是相反的，也是一种可行的悬架在长短横臂家族。

在这种形式中，上臂由两个连杆形成。

再者可能创造一个虚拟的中心去获得一个特殊的特征，它不能用球铰连接A型臂，也可能用一个横臂连接和一个侧直臂，但一般不推荐。

H型臂和A型臂的作用相似都有一个不接地的趾型连接，它有三个连杆在这个例子中。

（如图17.30（c））它可以作为上臂或者纵臂用在后悬架，当它被使用时，H臂的主内销必须和A型臂的主内销相平行，否则约束将引起悬架运动时由于企图转动叉式前桥因此使得H臂旋转。

这意味着侧视摇臂的瞬时中心在无穷远处，因此获得高价值的反功能是困难的。

这一般是使用H臂的限制因素。

H臂的一个潜在成本优势是，它仅仅需要两个底盘反应点代替三个去形成它的控制功能。

H臂纵臂和一个横向连杆的使用是一个特殊的例子，这里H臂被要求执行四连杆的作用代替三连杆。

它完成那个由刹车载荷作为扭矩输入。

所有轮子纵向输入的反应以使H臂的扭矩变大告终。

所以臂的结构需要成为不同的在这个例子中。

这个上拉杆只在前视图中起作用，有利于前视图瞬时中心的测定。

侧视图的几何完全被H臂的内主轴控制。

它和半拖曳臂式用同样的方法测定。

这种后悬架的版本在1989年的雷鸟中被使用。

（如图17.32）
一个系统利用五个单独的连杆可以做一个很满意的悬架。

它的布置和方位与三连杆和上面提到的现在由两个连杆组成的A型臂相似；两个变化的连杆展示在图17.33和17.34。

这种设计类型的运动时很灵活的，这里有一个问题，希望前视图的运动学不妥协于侧视图的几何学。

另一个做成五连杆的原因是为了获得短的主轴长度和小的负面磨砂半径。

这个被做通过延长连杆到虚拟点代替在理想的中心用一些球铰。

在使用这个设计的时候有个警告，侧斜配对上连杆和下连杆模拟的A型臂更喜欢纯粹的横臂或者纯粹的纵臂来构成连接臂。

这是因为用概念横臂使得侧视图几何的变化率太高。

而且横臂一定要安在轮胎边缘的内侧，这导致主轴长度至少是轮胎截面的一半。

看图17.34有这个概念的解释。

通过一个例子证明横臂和侧斜臂的不同，就是克尔维特84和梅赛德斯190的五连杆。

克尔维特在前视图用半轴和拱形连接臂和两个横臂在侧视图，形成上下臂。

横拉杆是第五个连杆。

就像前面提到的总体剖面图的横梁型后桥，这里有五个基本的运动物件它们需要被悬架臂控制。

轮胎行径、抗升力线、抗后座被侧视图的瞬时中心控制。

侧倾中心高度和侧倾转向被滚轴控制。

全部簧载的或者非簧载的都是悬架臂起作用。

这些纵长载荷例如刹车和加速被联系通过侧视图的瞬时中心和侧向载荷被联系。