双柱液压式汽车举升机【含外文翻译+任务书+3A0图纸量】

资源目录里展示的全都有预览可以查看的噢，，下载就有,,请放心下载，原稿可自行编辑修改=【QQ：11970985 可咨询交流】====================喜欢就充值下载吧。。。资源目录里展示的全都有，，下载后全都有,,请放心下载，原稿可自行编辑修改=【QQ：197216396 可咨询交流】==================== 英文原文 Foundations of machine design Bearings with rolling contact In bearings with rolling contact, the shaft is directly or indirectly supported by rolling elements, such as balls, cylindrical or conical rollers, or needles (ie , cylindrical rollers with a high l/D ratio).the occasionally encountered name ” antifriction bearings” suggests that this type has litter or no friction. This is erroneous, since the friction is merely of another nature than in journal bearings. In bearing with rolling contact, friction losses are caused by the elastic deformation of the surface in rolling contact , sliding friction of rolling elements with cages, retaining rings and seals, or with one another ,and also by some shear of lubricant. Characteristic 1. Ball bearings have rolling elements in the form of balls, which in all but the most inexpensive types are held in cages, separators, or retainers, and inner and outer grooved races. 2. Roller bearings have mainly cylindrical, conical, or barrel-shaped rollers instead of balls, but are otherwise quite similar to ball bearings. 3. Needle bearings usually have neither an inner race nor a cage .The needles are retained by integral flanges on the outer race. Transitional types between roller and needle bearings are found in many catalogs. Bearings with rolling contact have no slipstick effect, low starting torque and running friction, and unlike as in journal bearings, the coefficient of friction varies litter with load or speed. Low starting torque is of great advantage in railroad cars, and the railroad industry has given the main impetus for the development of mass-produced roller bearings in the past, mainly for this reason. These bearings may take both radial and axial loads (depending upon the type), and need less space axially but more radially(except needle bearings), and less lubrication and maintenance, than journal bearings .they are also noisier and more expensive, and cannot be repaired easily. Since rolling elements and raceways are theoretically in point or line contact and thus highly stressed in the contact area by typically cyclical loads, all bearings with rolling contact will eventually fail by fatigue when operated at their rated load, metal failure is typically by shear, just below the surface of raceways or rolling elements—a result of the three principal stresses in the x, y and z axis, all compressive .the phenomenon is popularly called spalling and is characterized by the presence of metal flakes in the grease. Even a small bearing load will induce contact stresses above the yield point (up to 500000psi; called hertz stresses after the mathematician who analyzed them). As a result, the contact surfaces are under residual compressive stresses, the phenomenon is popularly called spalling and is characterized by the presence of metal flakes in the grease. Even a small bearing load will induce contact stresses above the yield point (up to 500000psi; called hertz stresses after the mathematician who analyzed them). As a result, the contact surfaces are under residual compressive stresses, balanced by tensile stresses under the surface. Friction between rolling elements and races causes tensile stresses in the surface, believed to cause pitting. The pits, in turn, are the starting point of fatigue failure cracks. Life expectancy The life expectancy of ball bearings is not uniform and is base upon statistics. Usually, rated load values in catalogs give the minimum life (revolutions or hours at a given rpm) for 90% of a group of bearings, the so-called B-10 life . Mean or average life is about three times rated life, while median or 50% survival life is about five times rated life. The considerable spread of the expectancy curve is due to the many possible forms of favorable or unfavorable interaction between the different bearing elements on the basis of dimensional tolerances, surface finish and varying structural properties, as well as conditions of service. For plain ball bearing, radial load ratings in pounds (as given in catalogs) are based upon a given rpm and a given number of hours of life with 90% survival. The product rpm 60 hr life is then equal to . Other manufactures give rated loads on the basis of revolutions instead of hours of life (again, 90%survival). In that case, the number of hours of life can be found by dividing by rpmx60. it will be clear from the foregoing that ,for a given load, changes in rpm(within the design limits as given by the manufacturer)affect life in hours inversely; thus where n=rpm and H=life(hr.).changes in load, on the other hand, affect life in revolutions exponentially; thus = where F=load(lb) and B=life(revolutions).values of k range between 3 and 4 (3 for ball bearings;3.3for most roller and needle bearings).note that for k=3 ,doubling the load means reducing the life in revolutions by a factor of 8. The relation between B(life in revolutions)and H(life in hours),is as follows: B=Hrpm60 Other factors affecting bearing life The rated life of any bearing with rolling contact is based upon proper application conditions, such as adequate lubrication, good alignment, and adherence to recommended interference fit values for the races, if applicable. Any negative variance from these conditions will affect fatigue life unfavorably. On the other hand, life can be extend considerably by materials, the manufacture can produce a bearing with great life expectancy by increasing the accuracy of manufacture i.e. by reducing manufacture tolerances Static and dynamic load In tables giving permissible load data for ball dynamic loads, i.e, with the bearing stationary or rotating. The permissible static load is smaller than the permissible dynamic load due to the possibility of permanent indentation (flattening of the contact areas when the bearing is at rest). This phenomenon is called Brinelling .Some permanent indentation is almost unavoidable due to the high contact stresses, however, it has been found that permanent deformations smaller than times the element diameter usually do not unfavorably affect the bearingˊs operation. Rotation factor When the outer race rotates and the inner race is stationary (the reversal of common practice), the rated load may be increased by a factor of 1.2,since a ball rolls further per revolution on the outer race when the outer race rotates (the ball rotates in the same sense as the race) than on the inner race (where ball rotation is opposite to that of the race), and life is limited by ball revolutions. Equivalent dynamic load A plain ball bearing is designed to take radial loads. However, due to the curvature of the tracks in an axial plane of both inner and outer races, some axial load may be imposed on a radial ball bearing as well. Note that an axial load is taken by all balls , at least in theory, while a radial load is carried by less than half their number. Manufacturesˊcatalogs give simple conversion methods to and axial loads occur simultaneously in a rotating ball bearing. One manufacture uses a load conversion facture Fc which, when multiplied with the radial component R, gives the equivalent radial load P, which is then used to select a properly size bearing. These load conversion factors depend upon the ratio of the thrust load and the radial load(T/R)and vary for different types of bearings。 Gearing A friction drive consists of two cylinders rolling together under some pressure. When no slipping occurs, the tangential velocity V at the line of contact of the two cylinders is, of course, the same for each. From physics we know that V=w*r., in which V is the tangential velocity (in./sec), W is the rotational velocity (rad /sec), and r is the radius of rotation. Since V=w1r1 and also V=w2r2, we can equate both equations. Thus, dividing both sides by we get where is the speed ratio (Mw). When w is divided by 2π and multiplied by 60,we obtain rpm(n).Therefore, Mw= That is, the speed ratio of two cylinders in rolling contact is equal to the inverse ratio of their radii or diameters, in consistent units. The drawback inherent to a friction drive, such as that described above , is that it is liable to slip when power of any consequence is transmitted. It can be used only for very small torque applications, such as phonograph turntable drives and the like. The positive prevention of slippage in the transmission of large quantities of power requires the use of teeth, penetrating into the surface of each cylinder of the friction drive. Mating cylinders provided with teeth are called gears. The diameter of each of the original rolling cylinders is called the pitch diameter, and the sectional outline is called the pitch circle. The curved shape of the tooth outline must be such that no change in speed ratio occurs during the passing contact of each tooth with its mating tooth on the other gear. This is a basic requirement for all gearing. Curves that satisfy this requirement are called conjugate curves. While several of such curves exist, the one almost universally used at this time is the involute. This curve is that described by a point on a string as it is being unwound from a cylinder. In spur gears, the simplest of all gears, the teeth are straight lengthwise and parallel to the axis of rotation. The spur gear nomenclature which is mostly applicable to all other types of gears as well. Nomenclature and definitions The pitch point P is the point of the pitch circle. The pitch diameter Dp is the diameter of the pitch circle, and is equal to twice the pitch radius. The addendum (i.e., that which is added) is the radial distance from the pitch circle to the top of tooth (the crest) .The dedendum (i.e, that which is deducted ) is the radial distance from the pitch circle to the bottom of the groove between adjacent teeth (the root). Clearance is the difference between addendum and dedendum in mating gears. Clearance prevents binding caused by any possible eccentricity. The circle pitch Pc is the distance between corresponding sides of neighboring teeth, measured along the pitch circle. The diametral pitch Pd is the number of teeth of a gear for each inch of pitch diameter.( do not confuse Pd with Dp). The circle pitch and the diametral pitch are related as follows: The line of centers is a line passing through the centers of two mating gears. The center distance C (measured along the line of centers) equals the sum of the pitch radii of pinion and gear () or half the sum of the pitch diameters: Tooth width is the width of the tooth, measured along the Space width is the distance between facing sides of adjacent teeth, measured along the pitch circle. Tooth width plus space width equals the circular pitch. Face width measures tooth width in an axial direction. The circle from which the involute is generated is called the base circle. Backlash is the space width minus the tooth. It is necessitated by the tolerance of the manufacturing processes used. The face of the tooth is the active surface of the tooth outside the pitch cylinder. The flank of the tooth is the active surface inside the pitch cylinder. The fillet is the rounded corner at the base of the tooth. The working depth is the sum of the addendum of a gear and the addendum of its mating gearing The base pitch is similar to the circular pitch but is measured along the base circle instead of along the pitch circle. It can easily be seen that the base radius equals the pitch radius times the cosine of the pressure angle. Since, for a given angle, the ratio between any subtended arc and its radius is constant, it is also true that the base pitch equals the circle pitch times the cosine of the pressure angleΨ.the pressure angleΨ is the angle between the common tangent to the base circles, and the common tangent to the pitch circles at the pitch point. At present, the preferred pressure angle for spur angle gears is . In newer designs this angle replace the value of formerly used. The AGMA (American gear manufactures association) has establish the following proportions for pressure angle standard spur gears: Addendum= Dedendum= Minimum clearance=dedendum-addendum= In order to mate properly, gears running together must have :(1) the same pitch, (2) the same pressure angle, and (3) the same addendum and dedendum .The last requirement is valid for standard only. Since the number of teeth of each of two mating gears is proportional to its respective pitch diameter (prove), it is easier and thus customary to express the speed ratio in tooth numbers rather than pitch diameters; thus, Mw= Where subscript p=pinion, and g=gear. Dpg=pitch diameter of gear, and Dpp=pitch diameter of pinion; n=rpm, and N=tooth number. Recall that the involute is the curve almost universally used at present for shaping the outline of gear teeth, let us now analyze the action of a pair of mating involute teeth. 中文译文 机 械 设 计 基 础 滚动接触轴承 在滚动接触轴承中，轴直接或间接的由滚动体支撑着。如球滚子，圆柱滚子和圆锥滚子，还有滚针（一种长径比较大的圆柱滚子）。一些偶然间遇到的 名词“减摩轴承”表明这种类型（轴承）只有很少（或者没有）摩擦。其实这种认识是错误的，因为摩擦不仅仅是滑动轴承的通性。在滚动接触轴承中，摩擦损耗是由于滚动接触中的 表面塑性变形、滚动体与保持架、挡圈、密封圈或者它们相互间的滑动摩擦、以及润滑剂的剪切引起的。 特征 1.球轴承有球形滚动体，这种几乎是最便宜的滚动体有由保持架、挡圈和内外滚道固定。 2.滚动轴承主要有圆柱滚子、圆锥滚子或水桶形滚子以替代滚球，但在其他方面与球轴承非常的相似 3.滚针轴承通常既无内圈也无保持架。滚针是由整个法兰固定在外圆上。滚子和滚针轴承之间的过渡型轴承在许多种类中广泛应用。 与滑动轴承不同，滚动轴承无滑动面粘附现象的影响，其启动力矩低，运转摩擦小，摩擦系数几乎不随外载和转速变化。低启动转矩是铁路机车的主要优点之一。在过去由于大批量生产的滚子轴承的发展促进了铁路机车的发展，主要就是这个原因。 这些轴承能承受径向或轴向载荷（主要取决于轴承的类型）。与滑动轴承相比，滚动轴承占用很小的空间，但其径向空间很大。滚动轴承也仅需要很小的润滑和维护。但是滚动轴承噪声大且非常的昂贵，而且修理也不容易。 由于滚子部件在理论上是点接触或是线接触的，因而在接触区域由于循环载荷而产生很高的压力。所有的滚动接触轴承最终会由于施加在它们上的额定载荷产生疲劳而最终损坏。在滚道或滚动体的表面下由于剪切力发生金属损伤这是因为它受到x、y、三个方向的主应力。这种现象通常称为剥落。其特征是在润滑脂中出现金属薄片。即使是一个很小的轴承载荷也会导致超过屈服点的接触应力，因此，接触表面受残余压应力，而在表面下则承受拉应力。滚动体与轴承内外圈之间的摩擦产生拉应力。由此而产生凹坑，这凹坑，就是疲劳损伤裂纹的起始点。 预期寿命 所有的球轴承的预期寿命是不一样的。而是统计出来的。通常产品样本的额定载荷值给出了一组轴承内90％的轴承所能达到的最小寿命（转数或额定转速下的小时数）即所谓的B-10寿命。平均寿命大约是额定寿命的3倍。而50％的 轴承所达到的寿命通常是额定寿命的5倍。 对普通球轴承，径向载荷（以磅作单位）实基于以给定转速和一给定的轴承所达到寿命的90％的小时数的。产品的每分钟转速（rpm）60hour（小时）然后换算到。 其它生产商定的额定载荷是基于转而不是小时数。寿命（以小时计）可以由除以(rpm60)得到。通过前述，我们可以清楚的知道，对于一给定载荷，寿命转化为rpm影响与小时计的寿命是成反比关系的。即 . （ 这里n表示rpm，H表示寿命）另一方面，转化成载荷影响以转速计的寿命是指数关系的，即(F代表载荷（磅），B代表寿命(转速),K的值在3－4之间取，球轴承取3，大多数的滚子和滚针轴承取3.3)假定K=3,把载荷2，这样就意味着以转速计的寿命将为原来的1/8。 B和H之间的关系如下所示：B=Hrpm60 其它影响轴承寿命的因素 所有的球接触轴承的额定寿命基于在合适的应用条件下的。例如足够的润滑，良好的调整以及与所荐的过盈配合值相匹配。任何不符合上述使用条件的都会对疲劳寿命产生不利的影响。 从另一个方面说，如果能够给予良好的润滑，轴承寿命还是能够延长的。同时对于同一尺寸同一材料，制造商如果能够提高制造精度，如减小制造误差，就能制造出预期寿命更长的轴承。 静态载荷和动态载荷 在表中，对各种球轴承施加容许的载荷。静态载荷和动态载荷产生的差异是明显的。例如：稳定的轴承或旋转的轴承，由于永久压痕的可能性，容许的静态载荷一般都小于容许的动态载荷。这现象称之为布里涅耳现象。由于接触压力高，一些永久压痕几乎是不可避免的。然而，已证实那些小于滚动体直径的永久压痕通常不大影响轴承的运作。 旋转因素 当轴承外圈旋转而内圈静止时（与通常的情况相反）其额定载荷将增加到1.2倍，因为当外圈旋转时，球滚动体每转对外圈的作用力比对内作用力更甚。其寿命也由于球滚动体的旋转而受到限制。 当量动载荷 普通的求轴承是设计成承受径向载荷的。然而由于在轴承内圈外圈轴向面上滚道曲率影响。径向求轴承也承受了一些轴向力，轴向载荷被所有的滚球承受（至少在理论上，而径向载荷则只有少于半数的滚动球来承担。 如果一旋转轴承同时承受了径向载荷和轴向载荷，品目录手册给了一个简单方法去计算单量径向载荷。某一厂家采用了载荷转换因子FC。当它与径向分力R相乘，就得到当量径向载荷P,以此为标准去选取大小合适的轴承。这个载荷转化因子只取决于进推力与径向载荷的比，而且随不同类型轴承取值也不一样。 齿轮 摩擦驱动是在两个旋转圆柱体之间的压力下产生的。当未产生滑移时，两个圆柱齿轮啮合处的切向速度v对每个齿轮来说是相同的。从物理学中我们知道速度v=wr,在此式子中，v指切向速度，w指角速度，r指旋转半径。因为v=,同时=,所以我们得＝。两边都除以得，称为速度比（Mw）。 当w除以2，再乘以60，我们就得到了转速（n）因此Mw＝。也就是说在单位一致的情况下，两圆柱在旋转接触面上的速度比与它们的半径或直径成反比。 摩擦驱动的固有的缺点，如上述所描述的，在于它在传动功率时易产生滑移。因此它只能适用于一些小力矩方面。例如拍照的驱动等等。 在大功率传动中，在滚动体上加工齿形能有效的防止滑动。两啮合齿轮中的较小的齿轮称为小齿轮。大着称为大齿轮。每个原滚动体的直径称为节园直径。圆柱的横截面轮廓线称为节园。轮齿的外廓线应当满足以下要求：在两个啮合齿轮啮合所经过的接触面传动比不发生变化。这是对所有齿轮的一个基本要求。能满足这个要求的曲线都称为共扼曲线。 在已有的这样的曲线中，现在用的最广泛的是渐开线。这样的曲线可以如下描述：一直线绕一圆周纯滚动时直线上某点的轨迹。 在直齿圆柱齿轮（最简单的一类齿轮）中。所有的齿都是直线纵向的平行于其旋转轴的，直齿圆柱齿轮所用的一些术语大多数也适用于其它所有类型的齿轮。 术语和定义 节点P就是节园的相切点。节园直径是节园的直径，也等于节园半径的两倍。齿顶高是从节园到齿顶的径向高度。齿根高是节园与相邻近齿槽底部的径向高度。顶隙是一对啮合齿轮中，一齿轮的齿顶圆与另一齿轮的齿根圆之间的间隙。顶隙可避免各种可能的偏心所引起的相互干涉。 周节是指沿着分度圆上测量的相邻两齿同测齿廓渐的距离。径节是指节园直径单位英尺长度上齿轮的齿数。（不要混淆和）周节与径节之间的关系如下所示: =. 中心线是指穿过两啮合齿轮中心的连线。中心距C（沿中西线测量所得）等于小齿轮的节园半径与大齿轮节园半径之和，也等于两节园直径和的一半。 轮齿宽度是指在节园上测得的齿的宽度。齿宽也可称为齿厚。槽宽是指在节园上相邻两齿同侧齿廓间的弧长。齿厚与齿槽宽之和等于周节。沿轴向测得的齿宽即表面宽度。形成渐开线的圆称为基圆。侧隙是齿槽宽与齿厚的之差。齿侧间隙一般在制造过程中由公差保证。齿顶面是节圆柱外的啮合表面。齿根面是节圆柱内的啮合表面。倒角是在齿轮基部的圆角。 工作深度是齿顶高与其相啮合的另一齿轮的齿顶高（例如在标准齿轮中是2倍的齿顶高）。 基圆节距与周节相似，只是前者是在基圆上测得而不是在节圆上。我们可以知道;基圆半径等于节圆搬进乘以压力角的余弦。因为对于一个给定的齿轮，任何对应的弧长与其半径的比值都是一个常数。所以我们也可以说基圆节距也等于周节乘以压力角的余弦。压力角是指两基圆的公切线在节点处的夹角。现在，圆柱齿轮的参照压力角是。在最新的设计中，这个压力角已取代了以前常用的。 AGMA(美国齿轮制造协会)已经为压力角为的标准圆柱齿轮制定了如下的比率; 齿顶高＝ 齿根高= 最小顶隙=齿根高-齿顶高= 为了能够正确的啮合，齿轮传动必须具备下列3个条件： （1）相同的节距 （2）相同的压力角。 （3）相同的齿顶高和齿根高（适用于标准齿轮。） 因为两啮合齿轮的齿数是与他们各自的节圆直径成比例的，因此我们习惯上易用齿数而不是用节圆直径来表示速度比。即 Mw= 这里的下角p表示小齿轮，g表示大齿轮，表示大齿轮的节圆直径，是小齿轮的节圆直径。n是转速rpm，N是齿数。