1、MANUAL GEARBOXES 9.1 MANUAL GEARBOX CLASSIFICATION Gearboxes are normally classied according to the number of toothed wheel couples (stages) involved in the transmission of motion at a given speed； in the case of manual vehicle transmissions, the number to be taken into account is that of the forwar
2、d speeds only, without consideration of the nal gear, even if included in the gearbox. Therefore there are: Single stage gearboxes Dual stage or countershaft gearboxes Multi stage gearboxes Figure 9.1 shows the three congurations for a four speed gearbox. It is useful to comment on the generally ado
3、pted rules of these schemes. Each wheel is represented by a segment whose length is proportional to the pitch diameter of the gear； the segment is ended by horizontal strokes, representing the tooth width. If the segment is interrupted where crossing the shaft, the gear wheel is idle； the opposite o
4、ccurs if the segment crosses the line of the shaft without interruption. Then the wheel rotates with the shaft. Hubs are represented according to the same rules, while sleeves are represented with a pair of horizontal strokes. Arrows show the input and output shafts. Single stage gearboxes are prima
5、rily applied to front wheel driven vehicles, because in these it is useful that the input and the output shaft are oset； in G. Genta and L. Morello, The Automotive Chassis, Volume 1: Components Design, 425 Mechanical Engineering Series, c Springer Science+Business Media B.V. 2009426 9. MANUAL GEARBO
6、XES FIGURE 9.1. Schemes for a four speed gearbox shown in three dierent congurations: a: single stage, b: double stage and c: triple stage. conventional vehicles, on the other hand, it is better that input and output shafts are aligned. This is why rear wheel driven vehicles usually adopt a double s
7、tage gearbox. The multi-stage conguration is sometime adopted on front wheel driven vehicles with transversal engine, because the transversal length of the gearbox can be shortened； it is used when the number of speeds or the width of the gears do not allow a single stage transmission to be used. It
8、 should be noted that on a front wheel driven vehicle with transversal engine, having decided on the value of the front track and the size of the tire, the length of the gearbox has a direct impact on the maximum steering angle of the wheel and therefore on the minimum turning radius. The positive r
9、esult on the transversal dimension of multi-stage gearboxes is oset by higher mechanical losses, due to the increased number of engaged gear wheels. It should be noted that in triple stage gearboxes, shown in the picture, the axes of the three shafts do not lie in the same plane, as the scheme seems
10、 to show. In a lateral view, the outline of the three shafts should be represented as the vertices of a triangle； this lay-out reduces the transversal dimension of the gearbox. In this case and others, as we will show later, the drawing is represented by turning the plane of the input shaft and of t
11、he counter shaft on the plane of the counter shaft and of the output shaft. Gear trains used in reverse speed are classied separately. The inversion of speed is achieved by using an additional gear. As a matter of fact, in a train of three gears, the output speed has the same direction as the input
12、speed, while the other trains of two gears only have an output speed in the opposite direction； the added gear is usually called idler. The main congurations are reported in Fig. 9.2. In scheme a, an added countershaft shows a sliding idler, which can match two close gears that are not in contact, a
13、s, for example, the input gear of the rst speed and the output gear of the second speed. It should be noted that, in this scheme, the drawing does not preserve the actual dimension of the parts.9.1 Manual gearbox classication 427 FIGURE 9.2. Schemes used for reverse speed； such schemes t every type
14、of gearbox lay-out. Scheme b shows instead two sliding idlers, rotating together； this arrange- ment oers additional freedom in obtaining a given transmission ratio. The coun- tershaft is oset from the drawing plane； arrows show the gear wheels that match when the reverse speed is engaged. Scheme c
15、is similar to a in relation to the idler； it pairs an added specic wheel on the output shaft with a gear wheel cut on the shifting sleeve of the rst and second speed, when it is in idle position. Conguration d shows a dedicated pair of gears, with a xed idler and a shifting sleeve. The following are
16、 the advantages and disadvantages of the congurations shown in the gure. Schemes a, b and c are simpler, but preclude the application of synchro- nizers (because couples are not always engaged), nor do they allow the use of helical gears (because wheels must be shifted by sliding). Scheme d is more
17、complex but can include a synchronizer and can adopt helical gears. Schemes a, b and c do not increase gearbox length.428 9. MANUAL GEARBOXES 9.2 MECHANICAL EFFICIENCY The mechanical eciency of an automotive gear wheel transmission is high com- pared to other mechanisms performing the same function；
18、 indeed, the value of this eciency should not be neglected when calculating dynamic performance and fuel consumption. The continuous eort of to limit fuel consumption justi- es the care of transmission designers in reducing mechanical losses. Total transmission losses are conveyed up by terms that a
19、re both dependent and independent of the processed power； the primary terms are: Gearing losses； these are generated by friction between engaging teeth (power dependent) and by the friction of wheels rotating in air and oil (power independent). Bearing losses； these are generated by the extension of
20、 the contact area of rolling bodies and by their deformation (partly dependent on and partly independent of power) and by their rotation in the air and oil (power independent). Sealing losses； they are generated by friction between seals and rotating shafts and are power independent. Lubrication los
21、ses； these are generated by the lubrication pump, if present, and are power independent. All these losses depend on the rotational speed of parts in contact and, therefore, on engine speed and selected transmission ratio. Table 9.1 reports the values of mechanical eciency to be adopted in calcu- lat
22、ions considering wide open throttle conditions； these values consider a pair of gearing wheels or a complete transmission with splash lubrication； in the same table we can see also the eciency of a complete powershift epicycloidal auto- matic transmission and a steel belt continuously variable trans
23、mission. For the two last transmissions, the torque converter must be considered as locked-up. TABLE 9.1. Mechanical eciency of dierent transmission mechanisms. Mechanism type Eciency (%) Complete manual gearbox with splash lubrication 9297 Complete automatic transmission (ep. gears) 9095 Complete a
24、utomatic gearbox (steel belt； without press. contr.) 7080 Complete automatic gearbox (steel belt； with press. contr.) 8086 Pair of cyl. gears 99.099.5 Pair of bevel gears 90939.2 Mechanical eciency 429 FIGURE 9.3. Contributions to total friction loss of a single stage gearbox designed for 300 Nm as
25、function of input speed. It is more correct to reference power loss measurement as a function of rotational input speed rather than eciency. Figure 9.3 shows the example of a double stage transmission, in fourth speed, at maximum power； the dierent contributions to the total are shown. This kind of
26、measurement is made by disassembling the gearbox step by step, thus eliminating the related loss. In the rst step all synchronizer rings are removed, leaving the synchronizer hubs only； mechanical losses of non-engaged synchronizers are, therefore, mea- surable. The loss is due to the relative speed
27、 of non-engaged lubricated conical surfaces； the value of this loss depends, obviously, on speed and the selected transmission ratio. In the second step all rotating seals are removed. In the third step the lubrication oil is removed, and therefore, the bulk of the lubrication losses is eliminated；
28、some oil must remain in order to leave the contact between teeth unaected. By removing those gear wheels not involved in power transmission, their mechanical losses are now measurable. The rest of the loss is due to bearings； the previous removal of parts can aect this value. A more exhaustive appro
29、ach consists in measuring the complete eciency map； the eciency can be represented as the third coordinate of a surface, where the other two coordinates are input speed and engine torque. Eciency calcu- lations can be made by comparing input and output torque of a working trans- mission. Such map ca
30、n show how eciency reaches an almost constant value at a modest value of the input torque； it must not be forgotten that standard fuel consumption evaluation cycles involve quite modest values of torque and there- fore imply values of transmission eciency that are changing with torque. Figure 9.4 sh
31、ows a qualitative cross section of the aforesaid map, cut at constant engine speed. It should be noted that eciency is also zero at input430 9. MANUAL GEARBOXES FIGURE 9.4. Mechanical eciency map, as a function of input torque at constant engine speed； the dotted line represents a reasonable approxi
32、mation of this curve, to be used on mathematical models for the prediction of performance and fuel consumption. torque values slightly greater than zero； as a matter of fact, friction implies a certain minimum value of input torque, below which motion is impossible. A good approximation to represent
33、 mechanical eciency can be made using the dotted broken line as an interpolation of the real curve. 9.3 MANUAL AUTOMOBILE GEARBOXES 9.3.1 Adopted schemes In manual gearboxes, changing speed and engaging and disengaging the clutch are performed by driver force only. This kind of gearbox is made with helical gears and each speed has a syn-
