2021-11-02 10:11

Suspension principles

Obviously, if the loads applied to the rolling wheels of a vehicle were transmitted directly to the chassis, not only would its occupants suffer severely but also its structure would be subjected to an excessive degree of fatigue loading. The primary function of the suspension system, therefore, is to isolate the structure, so far as is practicable, from shock loading and vibration due to irregularities of the road surface. Secondly, it must do this without impairing the stability, steering or general handling qualities of the vehicle. The primary requirement is met by the use of flexible elements and dampers, while the second is achieved by controlling, by the use of mechanical linkages, the relative motions between the unsprung masses – wheel-and-axle assemblies – and the sprung mass. These linkages may be either as simple as a semi- elliptic spring and shackle or as complex as a double transverse link and anti-roll bar or some other such combination of mechanisms.

42.1 Road irregularities and human susceptibility

Some indication of the magnitudes of the disturbances caused by road irregularities can be gained from Surface Irregularity of Roads, DSIR Road Research Board Report, 1936–7. From this report it appears that surface undulations on medium-quality roads have amplitudes in general of 0.013 m or less, while amplitudes of 0.005 m are characteristic of very good roads. The average pitch of these undulations is under 4 m while most road vehicle wheels roll forwards at about 2 m/rev. In addition to the conventional tarmac roads, there are pavé and washboard surfaces, the latter occurring largely on unsurfced roads and tracks. Representative replicas of these two types of surface are described in The MIRA Proving Ground, by A. Fogg, Proc. A.D. Inst. Mech. Engrs. 1955–65.

Obviously the diameter of the tyre, size of contact patch between tyre and road, the rate of the tyre acting as a spring, and weight of wheel and axle assembly affect the magnitude of the shock transmitted to the axle, while the amplitude of wheel motion is influenced by all these factors plus the rate of the suspension springs, damping effect of the shock absorbers, and the weights of the unsprung and sprung masses. The unsprung mass can be loosely defined as that between the road and the main suspension springs, while the sprung mass is that supported on these suspension springs, though both may also include the weights of parts of the springs and linkages.


1110 The Motor Vehicle

Two entirely different types of shock are applied to the wheel: that due to the wheelrsquo;s striking a bump, and that caused by the wheelrsquo;s falling into a pot- hole. The former will be influenced to a major extent by the geometry of the bump and the speed of the vehicle, while the major influence on the latter, apart from the geometry of the hole, is the unsprung masses and spring rates, speed being an incidental influencing factor.

Human sensitivity to these disturbances is very complex, and a more detailed discussion can be found in Car Suspension and Handling by Donald Bastow, Pentech Press, London, 1980. It is widely held that vertical frequencies associated with walking speeds between 2.5 and 4 mph – that is, 1.5 to 2.3 Hz – are comfortable, and that fore-and-aft or lateral frequencies of the head should be less than 1.5 Hz. Dizziness and sickness are liable to be experienced if the inner ear is subjected to frequencies between 0.5 and about 0.75 Hz. Serious discomfort may be felt in other important organs at frequencies between 5 and 7 Hz.

42.2 Suspension system

A suspension system can be represented, in simplified form, as illustrated in Fig. 42.1. The natural frequency of the sprung mass – that at which it would bounce up and down if momentarily disturbed and left to bounce freely on its springs – is determined by the combined rate of the tyres and the suspension springs in series, which is—

1=1 1 R Rs Rt

where R is the overall suspension rate Rs is the suspension spring rate Rt is the tyre rate

In Fig. 42.1, the shock absorber is the hydraulic damper at D. Any friction in the suspension system will be additional to the hydraulic damping. However, whereas the hydraulic damping force of the shock absorber can be taken as proportional to the square of the vertical velocity of the sprung mass relative to that of the unsprung mass, the dynamic friction damping force is, in effect, constant regardless of velocity. It follows that while small amplitude, small


Rs D



Fig. 42.1

Suspension principles 1111

velocity movements of the suspension are virtually unaffected by the hydraulic damping, the force applied by the friction damping is the same for these small movements as it is for large ones.

With, for example, a new multi-leaf semi-elliptic spring, there is only a small difference between the static and dynamic interleaf friction forces – sometimes differentiated by calling them respectively stiction and friction. When, however, the same spring becomes rusty and dirty, this difference can become considerable, with the result that mean value of the work done by the friction damping during high frequency, small amplitude motions becomes excessive. Indeed, in extreme cases, the spring may become so stiff that, for small amplitude disturbances, it in effect does not deflect at all. This can of course lead to a harsh uncomfortable ride. It must be borne in mind, too, that even the hydraulic forces are in any case transmitted derectly – that is, uncushioned by the suspension springs – to the sprung mass.

Dampers have a two-fold function. First, they are for reducing the tendency for t


显然,如果施加到车辆的滚轮上的载荷直接传递到底盘,不仅其乘员遭受严重损坏,而且其结构也会受到过度的疲劳载荷。因此,悬架系统的主要功能是在可行的范围内将结构与由于路面不规则引起的冲击载荷和振动隔离开。其次,它必须在不损害车辆的稳定性,转向或一般操纵性质的情况下这样做。通过使用柔性元件和阻尼器满足了主要要求,而第二个是通过使用机械连杆控制非簧载质量 - 车轮 - 车轴组件 - 和簧载质量之间的相对运动来实现的。这些连杆可以像半椭圆形弹簧和钩环一样简单,也可以像双横向连杆和防倾杆那样复杂,或者其他一些这样的机构组合.


道路不平顺引起的扰动幅度的一些迹象可以从道路表面不规则性获得,DSIR道路研究委员会报告,1936-7。从这份报告中可以看出,中等质量道路上的表面起伏一般为0.013米或更小,而0.005米的幅度是非常好道路的特征。这些起伏的平均间距小于4米,而大多数公路车辆车轮以2米/转的速度向前滚动。除了传统的柏油碎石路面之外,还有密镶和搓板表面,后者主要出现在无瑕疵的道路和轨道上。这两种类型表面的代表性复制品描述于A.Fogg,Proc。的MIRA Proving Ground中。1955年至1965年。



人类对这些干扰的敏感性非常复杂,更详细的讨论可以在Donald Bastow,Pentech Press,London,1980中的汽车悬架和搬运中找到。人们普遍认为垂直频率与步行速度在2.5到4英里/小时之间有关 - 也就是说,1.5到2.3赫兹 - 很舒服,头部的前后或横向频率应小于1.5赫兹。如果内耳受到0.5至约0.75Hz之间的频率,则容易出现头晕和疾病。在频率在5到7赫兹之间的其他重要器官中可能会感觉到严重的不适。


悬架系统可以以简化的形式表示,如图42.1所示。簧载质量的固有频率 - 如果瞬间受到干扰而在弹簧上自由弹跳时会上下弹跳 - 取决于轮胎和悬架弹簧串联的速率,





例如,使用新的多叶半椭圆弹簧,静态和动态夹层摩擦力之间只有很小的差异 - 有时通过分别称它们的静摩擦力和摩擦力来区分。然而,当相同的弹簧变得生锈和变脏时,这种差异可能变得相当大,结果是在高频率,小振幅运动期间由摩擦阻尼完成的工作的平均值变得过大。实际上,在极端情况下,弹簧可能变得如此僵硬,以至于对于小振幅干扰,它实际上根本不会偏转。这当然可以导致严酷的不舒适乘坐。还必须记住,即使液压力在任何情况下都是直接传递 - 即,由悬架弹簧缓冲 - 传递到簧载质量。

阻尼器具有双重功能。首先,它们是为了减少在导致初始运动的干扰停止后,滑架单元继续在其弹簧上上下弹跳的趋势。其次,它们防止由于在与簧载质量系统的固有振动频率相同的频率下的周期性激励而导致的反弹振幅的过度累积。该固有频率是簧载质量和弹簧刚度的函数,实际上可以显示与1 /radic;delta;成正比,其中delta;是弹簧的静态偏转。

两种形式的干扰中的每一种都可以引起两种完全不同的共振。一种这样的形式是车轮以一定速度通过一系列等距凸块,使得它们产生的扰动频率与悬架系统的固有频率一致。第二个是车轮的不平衡,其不平衡力将随着旋转速度的平方而增加。在两个不同的频率中:一个是悬架弹簧系统上的簧载质量,第二个是轮胎上的非簧载质量 - 车轮和车轴组件。显然后者受到悬架弹簧刚度的影响,但只是略有下降。前者将经历相对较低的频率 - 可能约1至1.5赫兹 - 运输单元的弹跳,而后者是轮跳,频率较高 - 通常为10至15赫兹 - 并且几乎完全独立地产生运输单位的议案。为了最小化轮跳的幅度 - 不仅是共振跳跃而且是隔离的跳跃 - 必须保持非簧载重量尽可能小。簧载或非簧载质量的共振可能对车辆的操纵特性产生不利影响,实际上甚至危险程度。因此,显然,将阻尼器或减震器保持在良好的工作状态是很重要的。

这些相同的干扰也会引起车辆的俯仰或滚动振动。在这些情况下,固有频率分别是滚动和俯仰模式下的弹簧刚度和簧载质量绕横向轴和纵向轴的惯性矩的函数。单独的轴的滚动振荡,即围绕平行于车辆的纵向轴线的轴,通常被称为流浪,因为该效果类似于人的踩踏运动 - 一次前进一步。

通过使用扭杆弹簧,可以获得辅助侧倾刚度,而不会影响两轮弹跳刚度 - 完全轴的垂直运动阻力,或者同时具有独立悬架的两个车轮。这种弹簧通常被称为防倾杆,它横向安装在车辆下方的两个轴承中 - 通常是橡胶衬套 - 其端部通过杠杆连接,有时还通过轴环连接到车轴。如果使用简单的枢轴而不是钩环将杠杆连接到车轴上,这些杠杆可以用作半径杆,用于在车轴上下移动时引导车轴,如图43.10所示的麦弗逊式悬架。由于使用防倾杆不可避免地意味着在滚动期间在轮胎上施加额外的垂直载荷,因此它对转向和操纵有影响。这是因为外轮的轮胎在其与地面接触的区域中的额外垂直偏转使得当车辆转弯时它更容易受到横向偏转 - 增加滑移角。当设计师决定是在前部还是后部安装防倾杆时,或者两者都要考虑这种效果。

显然,前悬架和后悬架的运动之间必须存在相互作用 - 振动耦合效应,这必然会影响俯仰倾向。相互作用的大小将取决于扰动的频率,或汽车滚动的颠簸,以及前悬架和后悬架的固有频率。显然,强迫频率取决于凸起的间距,车辆的速度以及围绕俯仰轴的质量惯性矩,而车辆俯仰的响应幅度将不仅取决于这两个因素而且它的轴距也是如此。

在可以示出的情况下,如果后悬架具有比前部更低的固有频率,则俯仰运动倾向于比前部具有较低频率的情况持续更长时间。此外,如果后悬架具有较高的固有频率,则初始俯仰运动不太严重。因此,后悬架的固有频率通常高于前悬架的固有频率。车辆的速度越高,初始俯仰运动越不严重。产生这种效果的原因在于,随着速度的增加,撞击凸起的前轮和后轮之间的时间变为车辆在其固有频率下俯仰运动的周期时间的较小比例,并且理论上最终可能变为零。该原理类似于振动隔离的原理 - 例如发动机在具有低固有频率的橡胶安装系统上以相对高的频率自由振动。



现代阻尼器几乎总是伸缩式液压支柱插入簧载和非簧载质量之间 - 滑架单元和轴 - 或者不常见的是杠杆类型,它们也是液压单元。带有液压缸的杠杆式阻尼器的主体通常安装在滑架单元上,其致动杆连接到轴上。如果车身安装在车轴上,那么它所承受的高频高速运动可能会导致液压油的通气,从而对车辆的阻尼能力产生不利影响。在动态条件下轴的最小垂直加速度可以是20至30g的量级。


理想情况下,阻尼器设计的目的是为任何给定尺寸获得最大可能的能量吸收潜力,这意味着对凸起和回弹冲程的阻尼相等。然而,由于冲击行程通常是猛烈的强制运动,并且不希望将这种力直接通过阻尼器传递到簧载质量,因此冲击行程的阻尼通常小于回弹冲程上的阻尼。当然是由于轴的重量和悬架弹簧施加的力而产生的更温和的运动。为了减轻所有阻尼的冲击行程,从而减轻所有直接传递的冲击的滑架单元是不切实际的,原因有二:首先,它会使阻尼系统的能量吸收能力减半;其次,仅在回弹时阻尼将倾向于将滑架单元向下推至低于其在弹簧上的静态偏转的平均水平。通过在它们的端部配件和它们在轴和滑架单元上的固定装置之间插入橡胶衬套或块来避免通过阻尼器直接传递到滑架单元的高频小振幅振动 - 即,弹簧和簧下群众。





双管设计如图42.2所示。它由一个圆柱体A组成,焊接在一个圆柱体B上。后者用螺钉固定在外管C上,在该外管上焊接一个压制的钢盖和孔眼D,通过它将气缸A固定在轴或轮组件上。 。气缸A中的活塞E固定在活塞杆F上,活塞杆F在其上端有一个焊接在其上的孔眼,活塞杆F安装在车辆的车架上。活塞杆从气缸中出来的部分由焊接在固定孔上的盖子保护。压盖G防止活塞路穿过头部B的泄漏;由压盖填料刮掉的任何液体沿着排放孔向下流到气缸A和外管C之间的储存空间。气缸A的底部是一个底阀组件L.

活塞E有两个通过它钻出的同心环孔:外环由圆盘阀H覆盖,圆盘阀H由星形碟形弹簧I压住,而内圈由盘簧弹簧支撑的盘形阀J覆盖K.脚中的阀组件类似于活塞中的阀组件,除了覆盖孔的内环的下盘阀由盘簧而不是螺旋弹簧保持。这是为了减少减震器的死角 - 即工作行程不可用的长度。

气缸A的两端完全充满液体,但空间A和C之间只是部分填充。如果眼睛D向上移动,则流体必须从活塞E的下方移动到活塞E的上方。通过将阀门H提升到弹簧I上,该流体将穿过孔的外环。但是,由于活塞E的体积增加。气缸上端较小 - 通过活塞杆进入气缸的部分的体积 - 比下端的体积减小,流体也将通过底阀内孔和水平位移储层空间中的流体将上升。


双管阻尼器的一个优点是从主管移入外管的油带有热量,然后容易被导走。这倾向于将所有流体保持在适中的温度。显然,外管中的流体水平越高,传热效果越大。另一个优点是外管中的损坏 - 例如凹痕 - 不会干扰活塞的工作。



在单管阻尼器中,气体的压力 - 通常是氮气 - 必须高于主活塞下方的流体中的最大工作压力,并且可以是2.5MN / m 2的量级。这当然增加了悬架的总弹簧率,其量等于气体压力乘以活塞杆的有效横截面积。



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