【机械类毕业论文中英文对照文献翻译】弹性运动
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机械类毕业论文中英文对照文献翻译
机械类
毕业论文
中英文
对照
文献
翻译
弹性
运动
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【机械类毕业论文中英文对照文献翻译】弹性运动,机械类毕业论文中英文对照文献翻译,机械类,毕业论文,中英文,对照,文献,翻译,弹性,运动
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Error! Reference source not found.Spring into Action.The recent development of the Lotus active suspension system has proved that car body movements such as pitch and roll can be precisely controlled. Unfortunately it will take some time before we can all take advantage of these developments, and in the meantime we are stuck with the conventional springs and suspension that seems to have changed little since the days of the horse drawn carriage. Every enthusiast knows that the roll characteristics of a car have considerable influence on its dynamic behaviour as well as the comfort of its occupants. The parameters that control these characteristics are numerous, terms such as roll centre, roll stiffness, roll axis and roll couple are often bandied about but seldom adequately explained. I have often seen, in books and magazine articles, geometric constructions that can be used to determine roll centres and axis for different suspension layouts, but I have never seen an explanation of why these methods actually provide this information. It seems to be taken for granted that the reader will just accept it on face value, or is it because the writer is unsure himself and is just repeating that which he has read elsewhere. Anyway lets break with tradition and see if we can untangle the subject. Firstly, we must understand the terminology. The roll axis is an imaginary line running through the car from end to end; when the car rolls, whilst cornering on a smooth road, it rotates about this axis. Any part of the vehicle not on this axis bodily moves, either up or down, side to side or both. Fig 1 shows the motion, as well as the longitudinal position of the roll centres, these usually being points on the roll axis in line with the wheels. Note that theres no theoretical reason why the roll centre heights should be at the same level at each end, and indeed they rarely are. Now before going into the details of determining the roll centre and axis location, it must be understood that these parameters are not fixed in relation to the cars chassis, but move about depending on the deflection of the suspension and therefore vary depending on the roll angle which is influenced by the cornering force. When the car rolls the suspension on one side is compressed and on the other it becomes extended, for the purposes of analysis it matters not whether we consider the wheels as fixed and the body as capable of movement relative to them, or the body as fixed with wheels capable of moving. But as it is easier to visualise the motion with a fixed chassis we will go with that. Lets now consider the case of a double wishbone suspension system as in Fig 2. If we allow a very small vertical wheel displacement to take place, then the path of this movement (at the wheel end of the wish- bones) is at right angles to the wishbones, therefore the length of the wishbone does not affect the geometry of movement of the upright and wheel (for small displacements). So if the wishbones were extended in length until their inner pivots coincide then the motion of the wheel would be unaffected, but as the two wishbones now pivot around the same axis, we could replace them with a single arm fixed to the wheel axle, thus in effect creating an equivalent swing axle suspension system.Because this new swing axle is only a figment of our imagination, lets call it a virtual swing axle and its pivot a virtual pivot. Now, we need to consider the motion of the wheel at the tyre contact patch because this is our only connection with terra-firma, the reference from which we measure the roll. The diagram shows that this motion is at right angles to the line connecting the virtual pivot and the contact patch. Again for small deflections, this motion is unchanged as long as the effective pivot is located anywhere along this line. A mirror image of this applies to the wheel at the other side also. Thus the only common pivot point that satisfies both sides is the one at the junction of these lines, through the contact patches to the respective virtual pivots. So if there is one pivot point that can relate the motion of both tyre contact patches relative to the chassis, then that same point locates the chassis relative to the wheels. It is the point about which the body will pivot should the suspension on one side be compressed by the same amount as the suspension on the other side be extended, in other words it is the roll centre. I have emphasised the point about small wheel displacements, this is very important because the roll centre may vary its position enormously throughout the range of normal wheel movement. As the car rolls, its roll centres may change not only in height but also from side to side, as Fig 3, demonstrates.Even though we have used a double wishbone system to explain the method of determining the roll centre position, it is very easy to apply the method to any other suspension configuration. It is only necessary to determine the directions of movement of the contact patches, and draw lines at right angles to these through the contact patches, the point at which these last two lines cross is the roll centre. Fig 4, shows the method for several different systems.All of the above makes one big assumption (anyone spot it?) that the effective spring rates at the wheels are equal side to side. But arent they, I hear you ask. No, not always: what if you have progressive rate springing? The effective spring rate will be increasing on the side that is compressing and will be reducing on the other. To understand the effect that this may have, lets look at the extreme case of a car with a rigid suspension on one side only and with a normal spring on the other. The chassis is thus completely tied to one contact patch and so this is the only point about which it can roll. Thus the roll centre is at ground level directly under the wheel with the infinitely stiff springing. Obviously this situation is unreal but demonstrates how the actual roll centre moves away from the geometrically constructed one, if the springing is not symmetrical. It is all very well knowing where the roll centres and thus the roll axis are but what use is that knowledge, how can we use it and where should they be anyway? To answer this, we need to look at the superficially obvious question, Just what causes roll?. As we negotiate a curve the car is subject to centrifugal force, which is equal to the lateral acceleration multiplied by the mass of the machine (for a 1000Kg car cornering at 0.5g. The centrifugal force is 500Kg). This force is distributed throughout the car but for most analysis purposes can be considered to be acting only through the centre of gravity (C of G). Fig 5, shows that unless the C of G is level with the roll axis, a torque or couple (the roll couple) will be created, tending to make the machine roll about the roll axis. There is another equally valid way of considering the roll mechanism. The centrifugal force acting through the C of G produces a torque about ground level and is resisted by weight transfer to the outside wheels, that is, the outside wheels support a greater proportion of the cars weight and the inside wheels a lesser proportion. This change of load on each wheel normally causes the suspension to adopt a new position, or to put it another way the car rolls. It may have occurred to some of you that if the roll axis is made to coincide with the C of G then there will be no roll couple and hence no roll or, to take things a step further, if the roll axis is above the C of G then the roll couple will be in a direction which makes the car lean inward like a motorbike. Indeed it is quite possible to design the suspension layout to achieve these effects. If this is so, why do we need active suspension to do it for us? Well, because if we use the high roll axis necessary, then the suspension layout will cause a jacking up effect under cornering conditions, a phenomenon experienced with some swing axle designs in the past. One important point to note, one which is often misunderstood, is that regardless of the amount of roll allowed by the suspension design, the actual degree of weight transfer remains unchanged. This is only affected by the track, C of G height and cornering acceleration. So, as with most design features in anything mechanical, the selection of roll axis position is a compromise: too low and we get excessive roll, too high and other undesirable handling traits surface. In practice, the com- promise varies with different types of car but always such that some roll occurs. Lowering the C of G is another technically possible way of reducing the roll couple, but this can only be done to a certain extent, due to the boring necessity of leaving comfortable space for the occupants.The roll couple that such compromise leaves must be resisted by the cars springs, which leads us to roll stiffness. This term is defined such that the degree of roll is equal to the roll couple divided by the roll stiffness. Stiff springing obviously reduces roll and hence increases the roll stiffness, but if this is the criterion for selecting spring rates we will usually end up with an uncomfortable ride over normal road irregularities, so the anti-roll bar was developed to ease the situation. The anti-roll bar is a torsion bar (torsion spring) connecting the suspension systems on each side of the vehicle in such a way as to allow both wheels to respond unhindered to two-wheel bumps, such as a ridge across the road. But if the wheels try to move independently, as with a single-wheel bump, or in opposite directions when the car rolls then the anti-roll bar resists this tendency. Roll is reduced as intended but comfort suffers as the effective spring rate of each wheel is increased in the individual single-wheel bumps, although the combined spring rate of the two wheels is unchanged over joint disturbances. Again a compromise must be reached between the requirements of minimum roll and good response to road shocks. Roll bars, unlike the springs, are undamped (theoretically damping could be incorporated, although the manufacturers have generally concluded that it is not worthwhile). This is another reason for limiting the influence of the anti-roll bar, as oscillations might occur if the undamped bar is too stiff. It is well known that the under/over steering characteristics of a car can be substantially modified by tuning the springing and anti-roll bar stiffnesses, altering the roll stiffnesses of each end. While the vehicle as a whole has a certain roll stiffness, this is made up of the separate roll stiffnesses at the front and back, which may be quite different. For example, lets consider the case of a beam axle pivotted on the chassis at its mid point, and devoid of any form of springing. As unlikely as this layout seems, you may see it fitted to the front of some tractors, because it has good terrain-following properties. Now, because the chassis is completely free to rotate about the pivot point of this beam axle, then no roll stiffness is provided at this end of the machine and so all the stiffness needed must be available from the other end. This lack of any roll stiffness means that body roll cannot cause any weight transfer to the outside wheel, and hence as the total weight transfer must be the same anyway, the other end must obviously be subjected to proportionally more. Tyres have the interesting property that although they are capable of supporting higher cornering forces when subject to higher vertical loading, this does not go up in proportion. In other words, the co-efficient of friction is reduced as more weight is placed on them. In practice this means that weight transfer reduces the combined cornering force capable of being developed by the pair of tyres at one end of the car. Now as we have seen, the weight transfer at either end of the vehicle can be controlled to some extent by altering the roll stiffnesses of one or other, or both ends. Therefore, the tyre slip angles needed to produce the required cornering forces can be adjusted by modifications to the wheel springing, thus giving us the means to alter the under/over steering properties. Dampers, too, have their part to play in the extremely complex interrelations between the various forces acting on the car. During the transitional period between initiating a turn and it becoming fully established, the dampers will affect the dynamic roll stiffnesses. Because the dampers only contribute whilst the suspension is actually moving, however, they have no effect once the car has settled down to a steady state turn. So now you know why racers spend so much time setting up the suspension rates on their machines, when at first sight it would appear that the suspension is there only to cushion the bumps. Many competition cars have facilities for altering roll bar stiffness whilst on the move, making it quicker to achieve the desired performance but also allowing adjustments during a race as tyres wear and fuel weight reduces. So how will active suspension improve the situation? The computer program controlling the system could, for example, be set to apportion the front/rear roll stiffness. Assuming, that the system is set to give zero roll (this is by no means certain to be the aim of future manufacturers) then it is very interesting to follow through the implications for geometric roll centres, etc. Basically if we have no roll then the term roll axis becomes irrelevant, as does the need to have suspension link layouts that try to keep the outside wheels upright under cornering roll rather than under all conditions. Perhaps active suspension means a return to parallel equal length wishbones or perhaps better still, true trailing or leading arm systems. Normally, these designs suffer because of vast changes in roll centre positions, and because the camber angle varies with the roll angle of the body, the wheels always being held parallel to it. Straight line stability would be improved as bump induced wheel deflections would not cause any of the bump steer which comes from changing wheel camber. Thats a desirable state of affairs that is hard, if not impossible to achieve with conventional springs and dampers as we have seen. 弹性运动近来莲花主动悬架系统的发展证明了汽车车身像前倾后仰和反转这样的运动时是可以准确的控制的。令人遗憾的是在我们能够利用所有的这些成果之前还将花费一些时间的。与此同时我们无法摆脱从四轮马车时就有的只是看起来有些很小的变化的弹簧和悬架。每一个爱好者都知道汽车的侧倾性对汽车的动力性表现和乘坐舒适性有很大影响。这些参数控制这些特性是很多方面的，例如侧倾中心、侧倾刚度、侧倾轴线和侧倾力矩是经常被讨论的但是很少给出合适的解释。我曾经经常在书中和杂志文章中看到运用几何构造学来决定对于不同悬架安排的侧倾中心和侧倾轴线，但是我从来没有看见一种解释为什么这种方法真正可以提供这些信息。这似乎被认为假定读者只是表面认同这种方法，或者是不是因为作者本人也不确定，只是重复他从别处读来的。无论如何让我们打破传统来看看我们是否能解开这个问题。首先我们必需了解一些专用术语。侧倾轴线是虚构的一条从汽车的一端到另一端的直线；当汽车侧倾时，当在一段光滑的路面上转弯时，汽车绕着这条轴线旋转。这条轴线上的汽车任何部分都不移动，既不向上也不向下，也不从一边到另一边运动或者两者都有。图1展示了这种运动，同时也展示了侧倾中心的纵向位置，通常侧倾中心在侧倾轴线与车轮轴线的交点处。注意到没有理论原因为什么侧倾中心的高度在两头应该在统一水平线上，而且事实上它们很少在同一水平线上。现在在进入决定侧倾中心和轴线的位置的细节之前，必须明白这些参数对于汽车底盘并不是固定不变的，而是根据悬架的倾斜而运动，因此是根据由侧向力引起的侧倾角而变化的。当汽车旋转时一侧的悬架压缩另一侧的悬架伸张，分析它的目的是关系到我们应该认为车轮不动而车身相对车轮是可以运动的，还是应该认为车身不动而车轮可以运动。但是当把固定的底盘想象成运动的是比较容易的时候我们就会这样去做。现在让我们考虑如图2的双横臂悬架系统的情况。如果允许垂直车轮发生一个很小的位移，那么这个运动的路径（在横臂末端的车轮上）相对于横臂成直角，因此横臂的长度不影响垂直车轮的运动几何学（因为位移很小）。因此如果横臂延长至它们的内侧轴相交那么车轮的运动就不会受影响，但是如果两个横臂绕一个轴转动，我们就可以把它们设置成一个横臂固定到车轮轴上，因此实际上等价于一个等长的摆动轴的悬架系统。因为这种新的摆动轴只是我们设想的虚构的，我们叫它虚拟摆动轴和它的轴线叫做虚拟摆动轴线。现在我们需要考虑车轮轮胎边缘处的运动因为这是我们唯一与泥土相接触的，作为我们测量滚动的参考。上图展示了这个运动与连接虚拟轴线和接触边缘的线成直角。所以如果有一个轴线点能够使两个轮胎接地点的运动相对于悬架有关联，也就是相对于车轮在悬架上的点。这一点就是车身绕其旋转的点，悬架的一侧伸张，另一侧压缩，换句话说这就是侧倾中心。我已经强调国这一点关于车轮的小位移，这是非常重要的因为在车轮正常运动时侧倾中心可能变化很大。当车侧倾时它的侧倾中心不仅可能在垂直方向改变也可以在水平方向改变，参见图3所示。虽然我们已经用双横臂悬架解释了如何确定侧倾中心的位置，将这种方法运用到其它任何悬架结构是很简单的。只是需要确定接触边缘的运动方向，通过接触边画直角线，两条线的交点就是侧倾中心。图4展示了这种方法运用在一些不同的系统中。以上的全部做了一个大胆的假设（有人能指出来么？）那就是两边车轮的弹性是一样的。但是它们不一样么，我听见你问了。是的，不总一样：除非你有非常先进的等比弹性材料。实际的弹性特性是在被压缩的一边增加另一边减少。为了理解可能导致的后果，让我们看看极限情况，一边的悬架是刚性的另一边有正常弹簧。底盘因此完全和一边的接触边缘相连，这样这就是唯一的悬架能够绕其转动的点。因此侧倾中心在极其坚硬的刚性车轮的沿地平线的后面。显然这种情况是假的但是却论证了如果弹性不对称，侧倾中心是如何从原来的由几何学确定的位置移动的。这样就很好的了解了侧倾中心的位置也就知道了侧倾轴但是这有什么用呢,我们怎么应用它呢和在任何工况下它们的位置应该是怎样的呢？为了回答这个问题，我们需要看看一个显而易见的问题“到底是什么引起侧倾的呢”？大家都知道转弯时汽车受离心力影响，这个力等价于随着质量的增加而增加的侧向加速度（1000千克的汽车在0.5g的情况下转弯。离心力时500千克）。这个力被分解到汽车的每一个部分但是为了分析的目的可以认为这个力只作用在重心上。如图5所示，除非重心与侧倾轴平行，力矩或者转矩（侧倾转矩）就会产生，就产生汽车绕侧倾轴转动的倾向。由另一个等价的令人信服的方法可以讨论这个侧倾的机器。离心力作用通过重心产生一个关于地平线的力矩和转移到外侧车轮的重力的抵抗，这样外侧车轮支持汽车较大的重量而内侧车轮支持较小的重量。每个车轮的这种载荷的变化引起悬架需要适应一种新的位置，或者使汽车进入另一种
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