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Epicyclic and pre-selector gearboxes
Chapter 27
Epicyclic and pre -selector gearboxes In an epicyclic gear train, Fig. 27.1, two gears rotate about a common axis XX: that in the centre is termed the sun gear S, while the internally toothed outer one is the annulus gear A. Interposed between and meshing with them is what is termed the planet gear P, so called because it not only rotates about its own axis YY but, at the same time, orbits around the axis XX of the sun gear. In the illustration, the planet gear spindle is carried in a bearing on the end of a lever C, termed the planet carrier. This planet carrier, which is rigidly secured to the output shaft, may be a simple lever, as shown in Fig. 27.1, or a spider or a disc. In this instance, the annulus is fixed, the sun wheel is the driving gear, and the planet carrier the driven component. The larger the number of planet gears the lower is the loading on the meshing teeth. In most such gear trains, there are two or more planet gears equally spaced around the sun gear, so that the output shaft is not subjected to radial loading, and the radial thrusts due to their meshing loads are in balance. Among the major advantages of planetary gearing is that it is axially compact. Furthermore, all the components rotate around a common axis, instead of two shafts, so the gear casing can be of a relatively simple cylindrical shape. Another is that the gear ratio can be altered by using either a brake or
Fig. 27.1 792
Epicyclic and pre-selector gearboxes
793
a clutch, to lock one relative to the other two elements (gears or planet carders). Moreover, either to obtain higher overall gear ratios or more ratios within a casing of any given diameter, planetary gear trains can be compounded, or stacked coaxially in line, as in Figs 27.7 and 27.8. For road vehicle gearboxes, a major advantage of the epicyclic arrangement is that the load is spread over several gears instead of just one pair, as in conventional two-shaft gear boxes. Moreover, all the gears are constantly in mesh, so neither physical movement of gears nor the use of sliding clutches and synchronisers for shifting is necessary. On the other hand, the bandbrakes and clutches needed for changing the ratios in an epicyclic gear train are subject to wear, and obtaining smooth changes demands a high level of skill and experience on the part of the designer. A disadvantage, too, is a tendency to be noisy because of the number of gears meshing together, especially if the loading and overall gear ratio are high. In any case, an epicyclic transmission, because of its larger number of components, tends to be more costly than a simple two-shaft gearbox.
27.1
A simple epicyclic gear train
From the simplified diagram in Fig. 27.2(a) and (b) can be seen how a planetary gear train functions. For the purpose of this explanation, a lever is drawn in black on the planetary gear in both (a) and (b), to define the portion of the gear that is effective when any one set of teeth is in mesh at P and Q. The point R in this illustration is the axis YY of Fig. 27.1. In Fig. 27.2(b) the annulus is stationary, and the planet pinion has been displaced orbitally around axis XX relative to its position in Fig. 27.2(a). Because R is nearer than Q to P which, in this context, is the fulcrum about which the lever is actuated, it will not move as far as point Q. Consequently, the rotational displacement of the planet carrier C is smaller than that of the sun wheel S, but both are displaced in the same direction. To complete the picture, we now
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have to look upon the planet gear as comprising a succession of such levers, one for each meshing point, all of which are integral.
27.2
An alternative epicyclic gear train
Many alternative epicyclic gear layouts are possible. One is illustrated in Fig. 27.3. Instead of meshing with an annulus, the planet in this gear set meshes with sun wheel S1. The relatively large diameter sun wheel S1 is integral with its shaft and free to rotate about its axis XX carded in the frame F. It meshes with the teeth of P1, which is the smaller of the two wheels of a compound, or integral, planet gear pair P1 and P2. This pair rotates about the axis YY of a spindle carried in a bearing on the arm A of the planet carder the shaft of which, rotating about the axis XX, is also carded in the frame F. The larger diameter planet pinion P2, integral with and to the fight of the smaller P1, meshes with a smaller sun wheel $2. This sun wheel, integral with the frame F, is fixed. If the planet carrier arm A is rotated anti-clockwise, the planet wheel cluster Pl+2 orbits in the same direction, as demonstrated in Fig. 27.4(a) and (b), but rotates at a lower speed. If, on the other hand, the fixed wheel $2 were to be larger than S1, as shown at (c), the planet cluster would, for a reason to be explained in the next two paragraphs, rotate wheel S1 in the direction opposite to that of the orbiting planet carder. The lever representing the smaller of the two planet wheels, which meshes with the large sun wheel S1, is shown black in Fig. 27.4, while that representing the larger planet wheel, meshing with fixed, smaller, sun wheel $2, is hatched. At (a) and (b), therefore, P is the meshing point of the longer lever with the fixed smaller sun wheel S1, and point Q is the meshing point of the shorter lever with sun wheel $2. If the planet carrier is rotated anti-clockwise, the short lever pivots about point P and therefore tends to displace points Q and R to the left relative to P. Since point P is fixed, both the planet carrier and the sun wheel S1 have to rotate anti-clockwise. However, because of the differences of their radii, the angular displacement of the planet carrier is larger than that of the sun wheel. On the other hand, as can be seen at (c), if the fixed sun wheel $2 is larger
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Epicyclic and pre-selector gearboxes
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than the rotating sun wheel, the situation is similar, but P becomes the fulcrum so, although the planet carrier still rotates anti-clockwise, the sun wheel S1 rotates in the opposite direction. Although there are many other forms of planetary gearing, the two described here are those most commonly used in automotive applications. The broad principles already explained apply to them all. Heavy duty epicyclic gearing is to be found mostly in pre-selector and automatic gearboxes and back axle differentials. In the latter, bevel type planetary gears are usually used for turning the drive through 90~ they have also the advantage of compactness. Examples of bevel gear types are illustrated in Chapters 31, 33, and 34, the clearest being Fig. 34.7. As is explained in those chapters, the terminology for differentials is different: the planets are housed in a carrier called the differential cage bolted to the crown wheel. The latter, in effect, is substituted for the annulus. There are two sun wheels, termed differential gears, one on the inner end of each halfshaft. These mesh with the planet, or differential, pinions.
27.3
Epicyclic gear ratios
For ease of recognition of the various gears in all that follows, the annulus gear is called A, the planet carrier C, the sun gear S, and the planet gear P, and it is assumed that the gears have respectively A, S and P teeth. If either end of the planet carrier arm is integral with a tubular shaft and coaxial with the gear shaft at that end, the arm can be used as either the input or output. First, take the case of the planet being fixed and the cartier arm, serving as the input, turned through one revolution, clockwise about the axis of S, Fig. 27.4(a). If the wheel S were not in mesh with P and were fixed to the planet arm, it too would have orbited through 360 ~. However, since it is in mesh with P, it will also have rotated P/S times about its own axis, so its total rotation will be 1 + P/S revolutions. Therefore, the planet wheel has 20 and the sun wheel 40 teeth, the planet wheel will have turned 11/2 revolutions for
796
The Motor Vehicle
1 revolution of the carrier arm, so the gear ratio of this arrangement is 1.5:1. Now consider the situation if the sun wheel S is fixed, with the cartier, still the input, rotating around it, Fig. 27.4(b). In this case, the planet wheel P will have turned 1 + S/P, giving a gear ratio of 1 + 40/20 = 3 : 1. Finally, if the carrier arm is fixed and P is rotated, the ratio is P/S, giving a gear ratio of 0.5 9 1, but if S is rotated, the ratio is A/P giving a ratio of 2 9 1. Note that, in both instances, the driven and driving gears rotate in opposite directions. With such an epicyclic arrangement, therefore, it is possible to obtain four different ratios, simply by applying a brake or clutch to stop one of the elements in the gear train. This, however, is not of much significance for automotive transmissions because the space it takes up is similar to that obtained with a two shaft gearbox, some of the ratios are unsuitable and, in any case, it becomes more complex if a reverse gear is to be provided.
27.4 Simple planetary epicyclic gearing For automotive applications, a coaxial layout comprising an annulus gear with planet gear, or gears, and a sun gear is widely used. From Fig. 27.5(a), it can be seen that, if the annulus A is locked, and the input is clockwise through the sun gear S, the latter will rotate the planet gear anti-clockwise, so it will roll around the stationary annulus and therefore drive the planet cartier clockwise. The speed of the planet gear carrier will be: S/(S + A)
(1)
so, if tile sun gear has 40 and the annulus 80 teeth the gear ratio will be 40/(40 + 80) = 1/3. In other words, this arrangement will give a reduction of 3.0 : 1. However, if the drive is from A to S instead of S to A, the result will be an overdrive ratio of 0.33 9 1. Note that the planet gear, rotating in the direction opposite to that of the sun gear, is simply an idler and therefore has no influence on the gear ratio. If, on the other hand, the planet carrier C is locked, and the input is still the sun gear, Fig. 27.5(b), the output will, of course, be the annulus which is driven, through the idling planet gear, anti-clockwise. The gear ratio is" S/A
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(a) Fig. 27.5
(2)
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(d)
Epicyclic and pre-selector gearboxes
797
so S will drive P anti-clockwise and this, in turn, will drive the annulus anticlockwise at a speed of S/A times that of S, giving a ratio of 40/80 = 1/2 9 1. In other words, for one revolution of the sun gear, the annulus will rotate half a revolution (a reduction ratio of 2 : 1). With the sun gear locked, and the input from the planet carrier, Fig. 27.5(c), the annulus gear A is driven faster in the same direction. The gear ratio is:
(S + A)/A
(3)
the planet and annulus gears rotate in the same direction and the ratio is 120/80 = 1.5 : 1, so the annulus is rotated one and a half turns for every turn of the planet carrier. If the annulus gear A is locked and the planet carrier C is the input, Fig. 27.5 (d), the annulus is driven in the same direction. Again the gear ratio is: (S + A)/A
(4)
and the ratio is 1.5 : 1.
27.5 Simple planet epicyclic gearing in general From the foregoing, it can be concluded that, with an epicyclic gear train comprising an annulus, sun gear and planet carrier, any one of these can be fixed and the drive inputted through one of the other two, then the third will be driven at a different speed. The characteristics of such a gear train are as follows: 1. The output can be driven at a reduced speed relative to the input and in the same direction. 2. The output can be driven at a higher speed in the same direction. 3. The output can be driven at an alternative higher speed in the same direction. 4. The output can be driven at a lower speed than the input but in the opposite direction, to provide a reverse gear. 5. If any two of the gears locked together, the third cannot rotate relative to the others so the whole system turns as one solid mass, giving direct drive at a 1 : 1 ratio. 6. All ratios are dependent upon only the numbers of teeth on the sun and annulus gears, and are independent of the number of teeth on the planet gears. 7. With this arrangement, it is therefore possible to have direct drive, three forward gears, and one reverse gear. 27.6
Compound planet epicyclic gearing
An alternative to Fig. 27.3, in which two integral idler gears are used, is to have two independent but intermeshing 151anetary idler gears, one meshing with the sun and the other with the annulus, Fig. 27.6(a), in which case the gear ratio between these two comes into play. Consequently, if the annulus is fixed, the sun gear is the input and the planet cartier the output, the gear ratio is: P1/P2 • S/(S + A)
(5)
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Fig. 27.6 Double planet pinnions so, if A had 80 teeth, P1 10, P2 20 and S 40, the overall ratio would be that of the two planet gears times that in equation (1), which is 10/20 x 40/(40 + 80) = 0.5 x 0.33 = 0.166 : 1. Note that equation (5) is equation (1) multiplied by the planet gear ratio. It follows that to find the ratios in the other cases, all that is necessary is to multiply the ratios obtained in equations (2), (3) and (4) by the planet gear ratio. However, because there are two intermeshing planet gears, the output will rotate in the opposite direction to that previously obtained. Indeed, all epicyclic gear trains fall into two categories: given that the planet carrier is locked, the first is when the input and output gears rotate in the same direction and the second when they rotate in opposite directions. In Fig. 24.6(b), the principle is the same, but one of the two gears is an integral pair, so we have three planet gears on two wheels, or a compound planet gear. This arrangement offers the possibility of axially offsetting the sun gear from the annulus so that a wider choice of planet ratios is obtainable than would be possible if two separate planets had to be accommodated between the periphery of the sun gear and the tooth ring in the annulus. With this three wheel arrangement and, as before, the annulus fixed and the sun wheel the input, the overall gear ratio is: P1/P2 • S/(S + A)
(6)
In this instance the planet gear P3 behaves similarly to that in a simple sunplanet-annulus train except that its speed of rotation is influenced by the meshing pair of planet gears with which it is associated. Consequently, the comments that followed equation (5), and those in the subsequent paragraph, also apply here.
27.7
Numbers of teeth
Assembly of an epicyclic gear train is practicable only if the numbers of teeth on the gears has been appropriately chosen. For simple planet epicyclic gearing, the sum of the numbers of teeth on the sun and annulus gears divided by the number of planet pinions Npp must be an integer. For compound planet epicyclic gearing, and when the direction of rotation of the input and output are the same, (P2 • A) - (S • P3) divided by Npp • HCF must be an integer, where Hcr is the highest common factor between P2 and S. If the relative direction of rotation is not the same, (P2 • A) + (S • P3) divided by Npp • must be an integer. During assembly of the gear train, a vernier effect may be experienced,
Epicyclic and pre-selector gearboxes
799
allowing the planet pinions to be inserted incorrectly within the backlash. This can be avoided if the numbers of teeth on the sunwheel and annulus are each divisible by the number of planet pinions, and datum teeth on the planets are marked.
27.8
Another way of applying epicyclic gearing
An alternative to stopping one of the elements of an epicyclic gear train completely is to reduce its speed of rotation, by connecting it to a resistance of some sort. An example already mentioned is the differential gear for splitting the torque equally between two halfshafts driving either the front or rear road wheels. A particularly interesting application of this principle was made in 1964 by E Perkins Ltd (now Varity Perkins). An auxiliary drive from an engine was geared up and taken through a differential train, to drive a supercharger. The engine drove the planet carder and the output was taken through the annulus see I.Mech.E Auto. Div. paper presented by Dawson and Hayward on 14 April 1964. In automotive practice, the brakes or clutches, either pneumatically or hydraulically actuated, are applied gradually to the appropriate gears. The object is to vary the relative speeds progressively during the ratio changes and thus to provide smooth gear shifts. In this way, the need for a separate pedal-actuated clutch is obviated. In modem transmissions, because they require much less attention in service, clutches have largely superseded brakes even though the latter are less costly.
27.9
Epicyclic gearboxes
For car gearboxes, forward gear ratios of from 1 9 1 to about 5 : 1 and one reverse are needed. Moreover, as explained in Section 25.10, the forward ratios must be arranged in a geometrical progression. Epicyclic gearboxes used to be commonly installed in Daimler, Lanchester and Armstrong Siddeley cars, mostly to provide preselection of gears and to eliminate the clutch, although the pedal control was retained for changing gear instead of operating the clutch. However, they have now been superseded by automatic transmissions with two pedal control. In heavy commercial vehicles in which optimum fuel economy is an overriding requirement, the inefficiency of a torque converter is unacceptable so epicyclic gearboxes arc still widely used. For this type of application, extra gear trains arc needed for providing large numbers of gear ratios needed for coping with the heavy loads carried. As previously stated, given that top speed is direct drive, seven different drives are obtainable from a simple train comprising a sun, annulus and carrier. However, this calls for complex construction and, moreover, some of the ratios are unsuitable. Consequently, even for cars, an alternative method is better. This entails interconnecting several epicyclic trains in series, so that suitable ratios can be selected. Generally, the first train is the primary one, the others modifying its ratio before transmitting the drive back through the primary planet carrier to the road wheels. An overall ratio of 1 9 1 is obtainable by locking the whole set of gear trains together. With such arrangements, the number of ratios obtainable rises rapidly with the number of trains although, as previously indicated, some are unsuitable for automotive gearbox applications.
800
The Motor Vehicle
The trains can be all of the same type or of different types. Several arrangements with spur gears are illustrated in Fig. 27.7, in which the brakes are labelled B and the clutches C: in practice multi-plate clutches are employed. The input is on the left and the output on the fight. In diagram (a) two simple trains are in tandem while at (b) two trains with double-sun gears are similarly arranged. By making some individual members function as part of two trains, the arrangement can be simplified, but fewer ratios are obtainable. An example is shown at (c), where the sun gears are integral and a single planet carrier serves both trains. With this arrangement, only one clutch is needed, but only three ratios are obtainable, as compared with four in (a). At (d) two of the suns $1 and $2 are integral and there is a third sun S with a carrier common to all three sets of planet gears. The sun S and its meshing planet gear serves both trains. Because $2 is larger than the driven sun S, reverse is obtained when brake B2 is applied and, since S1 is smaller than S, a forward drive ratio is obtained when brake B1 is applied. A third method of designing an epicyclic gearbox is to compound several simple epicyclic gear trains. This was the basis of the Wilson transmission, originally known as the Wilson-Pilcher gearbox, developed at the beginning of the 20th century. Later Vauxhall Motors worked with Major Wilson to develop it further but dropped it in 1927, when General Motors took the company over. Subsequent development was done by the Daimler company for their cars.
27.10
Basic principle of the Wilson gearbox
A Wilson type gearbox is illustrated in Fig. 27.8(a), and the functioning of its various secondary trains is shown in (b) to (f). The primary epicyclic gear train is common to all the ratios. Its sun S 1, Fig. 27.8(b), is driven by shaft D, which is coupled directly to the engine crankshaft, while its planet carrier C1 is coupled directly to the transmission line to the road wheels. In other
Fig. 27.7 Epicyclic gearbox arrangements
Epicyclic and pre-selector gearboxes
801
Fig. 27.8 Wilson gear ratios
words, these two elements are respectively the input and output regardless of ratio selected. The required gear ratios are obtained by driving the annulus at different speeds in relation to engine speed. How this is done is as follows. If engine speed were constant, at 1000 rev/min, and the annulus braked (zero rev/min), the speed of the carrier would be 1000 x S1/(A1 + S1) where, in general, A, S and C throughout what follows are the numbers of teeth on the annulus and sun wheel respectively, as in Section 27.9. If A1 = 100 and $1 = 25, the speed of C1 = 200 rev/min and the reduction ratio is 5 : 1. On the other hand if, with engine speed still at 1000 rev/min, the annulus is driven at 100 rev/min in the same direction as the engine by other elements in the compounded series of epicyclic gears, the speed of rotation of the sun gear relative to the annulus would be only 900 rev/min. So the speed of the planet carrier would be 900 x 25 = 22 500 divided by 125 = 180 rev/min but, as the carrier would rotate faster because the annulus was rotating, the 100 rev/min of the latter (in these circumstances not multiplied by the sun-toplanet ratio) must be added, giving 280 rev/min for the speed of the planet carrier. Therefore, when the annulus is stationary, the equation in the previous paragraphs is in fact 0 + (1000 x S1/A1 + S1) and, when the annulus is rotating at 100 rev/min, it becomes 100 + (1000 x S1/A1 + $1). In more general terms, the equation is (RE- CA) X (S1/A1 + S1) + RA, where RA and RE are respectively the speeds of rotation of the annulus and engine crankshaft. It follows that, if RE had been 200 rev/min, the output speed of the planet carrier would have been 380 rev/min.
802
The Motor Vehicle
If the annulus is driven in the opposite direction, its speed relative to the crankshaft becomes negative. So given an annulus rotating at a negative speed of 400 rev/min we have (1000 + 400) x (25/100 + 25) - 400 = - 120 rev/min, or in other words a reverse gear ratio of 0.12 : 1.
27.11
The auxiliary trains in the Wilson gearbox
As previously indicated, to drive the annulus at different speeds auxiliary epicyclic gear trains are used. Those for second gear are shown in Fig. 27.8(c). The sun gear $2, in common with S1 adjacent to it, is driven by the engine, the planet carder C2 is coupled to the annulus A1 and, to obtain second gear, a brake is applied to the annulus A2. So long as this brake is holding A2 stationary, the coupled carrier C2 and annulus A1 rotate in the same direction as the engine but at a lower speed. Consequently, C1 rotates at a higher speed than it did in first gear when, as described in the second paragraph of Section 24.10, A1 was the stationary element. The actual speed will, of course, depend on the numbers of teeth on the gears involved. To obtain third gear, Fig. 27.8(d), annulus A1 and therefore also the planet cartier C2 must be made to rotate faster than in second gear. This entails causing A2 to rotate in the same direction as the engine, by applying the brake to drum F to stop $3. The other two elements in the second gear train, C3 and A 3, are coupled respectively to A2 and C2. With $3 fixed and A 3 rotating in the same direction as the engine, C3, $2 and A2 will also be rotating in that direction, the last two because C3 is coupled to A2. Consequently, the speed of rotation of C2 must be greater than when A2 was fixed: therefore annulus A2 must be rotating faster than in second gear. Again, the actual speeds depend on the numbers of teeth on the gears. Direct drive, Fig. 27.8(e), is obtained by sliding the male cone G along the splines on the shaft D, and locking it in the female cone in the drum F which is fixed to $3. This locks the whole epicyclic gear assembly together, so that it rotates en bloc. Obtaining reverse gear, Fig. 27.8(f), entails bringing the fourth epicyclic train into operation by applying the brake to A 4. S u n S 4 is fixed to A1, and carrier C4 is fixed to the driven shaft D. Sun $1 is driven forwards by the engine, and the planet pinion, acting as an idler, drives the annulus A1, and with it the sun $4, backwards. Since A4 is braked, $4 drives the planet carrier C4 backwards too. Moreover, since both C1 and C4 are fixed to the output shaft to the road wheels, both have to rotate at the same speed in the same direction. This speed is in fact determined by the numbers of teeth on A1 and $1 relative to those on $4 and A n, the last mentioned pair determining the speed of rotation of the annulus A. This is not easy to visualise, so let us take an example in which S1 has 25, A1 100, $4 40 and A 4 80 teeth. To represent the conditions in reverse gear, apply the brake to A 4, and then rotate C4, and with it the integral carrier C1, backwards one revolution. Now consider the sequence of events as this rotary motion is transmitted through each of the intermeshing gear pairs in turn. With A 4 fixed, the planet carrier C4 will rotate $4, and with it the annulus A1, backwards 80/40 = 2 revolutions. This, in turn would have rotated S1 backwards 100/25 = 4 revolutions which, at first sight, appears to make an overall ratio of 8 : 1. However, the driven shaft has rotated minus 1 turn and the driving shaft plus 8 turns, so the overall ratio is in fact 7 : 1.
Epicyclic and pre-selector gearboxes
27.12
803
The clutches and brakes in the Wilson gearbox
To simplify in Fig. 27.8(e) a cone clutch was shown, but this, of course, would be too harsh for the purpose. In practice, a pneumatically or hydraulically actuated multi-plate clutch such as that illustrated in Fig. 27.9 is employed. The clutch illustrated is air actuated. A control valve lets fluid under pressure enter cylinder A beneath the piston, to actuate the lever B. This moves the ring C to the right, about its pivot D. Mounted on gimbal pivots F (one on each side) within ring C is the housing E for the ball thrust bearing. Consequently, the axial displacement of the centre of C causes the ball thrust bearing to compress the clutch plates between the presser plate G and the clutch hub H, to which are splined the driving plates of the clutch. The drum K, to which the driven plates are splined, is in turn splined to the hub of the sun gear $1. When the fluid pressure is released by the control valve, the clutch is released by a set of coil springs, equally spaced around the hub (only one can be seen in the illustration). A typical brake, Fig. 27.10, for acting upon the periphery of the annulus in this type of gearbox has an outer and an inner band, A and B respectively, having friction linings L. The outer band is anchored by the hooked link D, and the inner one by the lug F projecting through a slot in the outer band. The reactions to the brake torque, at these two diametrically opposite anchorage points, are equal and opposite and therefore do not add to the load on the bearings on which the annulus rotates. A tie rod G pulls the outer band on to the inner band, and thus the inner band on to the drum, which is the periphery of the annulus. Note that the pull of the tie rod on the lower end of the brake band is reacted by an equal and opposite pull applied by the hook on its upper end, and the direction of rotation relative to the fixed anchorage point is such that the brake band is automatically wrapped around the drum. Stop C actuates an automatic adjustment device for compensating for wear. The actuation mechanism is shown in more detail in Fig. 27.11 and the automatic adjustment mechanism in Fig. 27.12. Fluid pressure, in this case
Fig. 27.9
Fig. 27.10
804
Fig. 27.11
The Motor Vehicle
Fig. 27.12
air, on the piston A lifts the piston rod Q in Fig. 27.11. This, in turn, rotates the lever B about its pivot O. Roller C, on the other end of the lever, moves along the cam-shaped lower surface of the lever K, termed the bus bar. As it does so, it lifts the tie rod, G in Fig. 27.10, to apply the brake. The shape of the cam is such that the initial movement of the lever is rapid, to take up the clearance between the drum and brake bands. Subsequently, the slope of the cam is less steep so that the tie lifts more slowly, and the ratio of force on the piston to the leverage on the tie is therefore greater. This effect is enhanced by the fact that, as the roller approaches the end of bus bar K, the lever B on which it is mounted comes up to its top dead centre position, relative to the axis P of the pivot.
27.13
Automatic compensation for wear
In the upper diagram of Fig. 27.12, the adjustment device is in the 'brake off' position and in the lower one it is in the 'brake on position' with zero wear on the lining. Surrounding the round nut screwed on the top of the tie rod G is a coil spring B. One end of this spring is fixed to a pin projecting upwards from plate A. It is then coiled several times round the nut and secured to a second pin, which is fixed to the knife-edge pivot plate J and projects upwards through a slot in plate A. Plate A is free to rotate relative to the nut H. Each time the bus bar K, Fig. 27.11, is pulled upwards to apply the brake, the upper end of the tie rod, and with it the plate-and-spring assembly, moves over to the left until the lug on plate A just makes contact with the adjacent stop C, which is C in Fig. 27.10. For all gears except top, where it is fixed to the gearbox casing, this stop is mounted on the brake band. When wear has occurred, the stop goes further than just contacting the lug: it strikes it and, deflecting it, rotates the plate A anti-clockwise. This uncoils the spring around the nut H, which therefore is then free to rotate, except that the friction between it and its conical seating in the pivot plate prevents it from doing so. When the brake is released, however, the coils tighten around the nut. Then, as the tie rod is lowered and the whole assembly retracts to the fight in the illustration, the plate rotates back to its original 'brake off' position, and the lug on the other side of the plate comes up against the stop D fixed to the casing. At this point, the coil spring grips the
Epicyclic and pre-selector gearboxes
805
nut and, overcoming the friction between it and the plate A, rotates it to reduce the effective length of the tie rod, and thus takes up the clearance due to wear between the brake lining and drum. Wilson type gearboxes are still widely installed in public service vehicles although, in most instances, gear shifting is effected either semi- or fully automatically by means of an electronic control system. This ensures that all gear changes and moving away from rest are effected smoothly. In the semiautomatic systems, the driver uses a small lever to select the gears, but the actual shifting and pressures applied to the actuation mechanisms are effected automatically in relation to the gear selected and factors such as vehicle speed, engine speed and torque being transmitted. Once a gear is fully engaged, the pressure actuating the brake and clutch is increased to the maximum, for prevention of slippage.
Epicyclic and pre -selector gearboxes In an epicyclic gear train, Fig. 27.1, two gears rotate about a common axis XX: that in the centre is termed the sun gear S, while the internally toothed outer one is the annulus gear A. Interposed between and meshing with them is what is termed the planet gear P, so called because it not only rotates about its own axis YY but, at the same time, orbits around the axis XX of the sun gear. In the illustration, the planet gear spindle is carried in a bearing on the end of a lever C, termed the planet carrier. This planet carrier, which is rigidly secured to the output shaft, may be a simple lever, as shown in Fig. 27.1, or a spider or a disc. In this instance, the annulus is fixed, the sun wheel is the driving gear, and the planet carrier the driven component. The larger the number of planet gears the lower is the loading on the meshing teeth. In most such gear trains, there are two or more planet gears equally spaced around the sun gear, so that the output shaft is not subjected to radial loading, and the radial thrusts due to their meshing loads are in balance. Among the major advantages of planetary gearing is that it is axially compact. Furthermore, all the components rotate around a common axis, instead of two shafts, so the gear casing can be of a relatively simple cylindrical shape. Another is that the gear ratio can be altered by using either a brake or
Fig. 27.1 792
Epicyclic and pre-selector gearboxes
793
a clutch, to lock one relative to the other two elements (gears or planet carders). Moreover, either to obtain higher overall gear ratios or more ratios within a casing of any given diameter, planetary gear trains can be compounded, or stacked coaxially in line, as in Figs 27.7 and 27.8. For road vehicle gearboxes, a major advantage of the epicyclic arrangement is that the load is spread over several gears instead of just one pair, as in conventional two-shaft gear boxes. Moreover, all the gears are constantly in mesh, so neither physical movement of gears nor the use of sliding clutches and synchronisers for shifting is necessary. On the other hand, the bandbrakes and clutches needed for changing the ratios in an epicyclic gear train are subject to wear, and obtaining smooth changes demands a high level of skill and experience on the part of the designer. A disadvantage, too, is a tendency to be noisy because of the number of gears meshing together, especially if the loading and overall gear ratio are high. In any case, an epicyclic transmission, because of its larger number of components, tends to be more costly than a simple two-shaft gearbox.
27.1
A simple epicyclic gear train
From the simplified diagram in Fig. 27.2(a) and (b) can be seen how a planetary gear train functions. For the purpose of this explanation, a lever is drawn in black on the planetary gear in both (a) and (b), to define the portion of the gear that is effective when any one set of teeth is in mesh at P and Q. The point R in this illustration is the axis YY of Fig. 27.1. In Fig. 27.2(b) the annulus is stationary, and the planet pinion has been displaced orbitally around axis XX relative to its position in Fig. 27.2(a). Because R is nearer than Q to P which, in this context, is the fulcrum about which the lever is actuated, it will not move as far as point Q. Consequently, the rotational displacement of the planet carrier C is smaller than that of the sun wheel S, but both are displaced in the same direction. To complete the picture, we now
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794
The Motor Vehicle
have to look upon the planet gear as comprising a succession of such levers, one for each meshing point, all of which are integral.
27.2
An alternative epicyclic gear train
Many alternative epicyclic gear layouts are possible. One is illustrated in Fig. 27.3. Instead of meshing with an annulus, the planet in this gear set meshes with sun wheel S1. The relatively large diameter sun wheel S1 is integral with its shaft and free to rotate about its axis XX carded in the frame F. It meshes with the teeth of P1, which is the smaller of the two wheels of a compound, or integral, planet gear pair P1 and P2. This pair rotates about the axis YY of a spindle carried in a bearing on the arm A of the planet carder the shaft of which, rotating about the axis XX, is also carded in the frame F. The larger diameter planet pinion P2, integral with and to the fight of the smaller P1, meshes with a smaller sun wheel $2. This sun wheel, integral with the frame F, is fixed. If the planet carrier arm A is rotated anti-clockwise, the planet wheel cluster Pl+2 orbits in the same direction, as demonstrated in Fig. 27.4(a) and (b), but rotates at a lower speed. If, on the other hand, the fixed wheel $2 were to be larger than S1, as shown at (c), the planet cluster would, for a reason to be explained in the next two paragraphs, rotate wheel S1 in the direction opposite to that of the orbiting planet carder. The lever representing the smaller of the two planet wheels, which meshes with the large sun wheel S1, is shown black in Fig. 27.4, while that representing the larger planet wheel, meshing with fixed, smaller, sun wheel $2, is hatched. At (a) and (b), therefore, P is the meshing point of the longer lever with the fixed smaller sun wheel S1, and point Q is the meshing point of the shorter lever with sun wheel $2. If the planet carrier is rotated anti-clockwise, the short lever pivots about point P and therefore tends to displace points Q and R to the left relative to P. Since point P is fixed, both the planet carrier and the sun wheel S1 have to rotate anti-clockwise. However, because of the differences of their radii, the angular displacement of the planet carrier is larger than that of the sun wheel. On the other hand, as can be seen at (c), if the fixed sun wheel $2 is larger
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Epicyclic and pre-selector gearboxes
(a)
~
795
(c)
Fig. 27.4
than the rotating sun wheel, the situation is similar, but P becomes the fulcrum so, although the planet carrier still rotates anti-clockwise, the sun wheel S1 rotates in the opposite direction. Although there are many other forms of planetary gearing, the two described here are those most commonly used in automotive applications. The broad principles already explained apply to them all. Heavy duty epicyclic gearing is to be found mostly in pre-selector and automatic gearboxes and back axle differentials. In the latter, bevel type planetary gears are usually used for turning the drive through 90~ they have also the advantage of compactness. Examples of bevel gear types are illustrated in Chapters 31, 33, and 34, the clearest being Fig. 34.7. As is explained in those chapters, the terminology for differentials is different: the planets are housed in a carrier called the differential cage bolted to the crown wheel. The latter, in effect, is substituted for the annulus. There are two sun wheels, termed differential gears, one on the inner end of each halfshaft. These mesh with the planet, or differential, pinions.
27.3
Epicyclic gear ratios
For ease of recognition of the various gears in all that follows, the annulus gear is called A, the planet carrier C, the sun gear S, and the planet gear P, and it is assumed that the gears have respectively A, S and P teeth. If either end of the planet carrier arm is integral with a tubular shaft and coaxial with the gear shaft at that end, the arm can be used as either the input or output. First, take the case of the planet being fixed and the cartier arm, serving as the input, turned through one revolution, clockwise about the axis of S, Fig. 27.4(a). If the wheel S were not in mesh with P and were fixed to the planet arm, it too would have orbited through 360 ~. However, since it is in mesh with P, it will also have rotated P/S times about its own axis, so its total rotation will be 1 + P/S revolutions. Therefore, the planet wheel has 20 and the sun wheel 40 teeth, the planet wheel will have turned 11/2 revolutions for
796
The Motor Vehicle
1 revolution of the carrier arm, so the gear ratio of this arrangement is 1.5:1. Now consider the situation if the sun wheel S is fixed, with the cartier, still the input, rotating around it, Fig. 27.4(b). In this case, the planet wheel P will have turned 1 + S/P, giving a gear ratio of 1 + 40/20 = 3 : 1. Finally, if the carrier arm is fixed and P is rotated, the ratio is P/S, giving a gear ratio of 0.5 9 1, but if S is rotated, the ratio is A/P giving a ratio of 2 9 1. Note that, in both instances, the driven and driving gears rotate in opposite directions. With such an epicyclic arrangement, therefore, it is possible to obtain four different ratios, simply by applying a brake or clutch to stop one of the elements in the gear train. This, however, is not of much significance for automotive transmissions because the space it takes up is similar to that obtained with a two shaft gearbox, some of the ratios are unsuitable and, in any case, it becomes more complex if a reverse gear is to be provided.
27.4 Simple planetary epicyclic gearing For automotive applications, a coaxial layout comprising an annulus gear with planet gear, or gears, and a sun gear is widely used. From Fig. 27.5(a), it can be seen that, if the annulus A is locked, and the input is clockwise through the sun gear S, the latter will rotate the planet gear anti-clockwise, so it will roll around the stationary annulus and therefore drive the planet cartier clockwise. The speed of the planet gear carrier will be: S/(S + A)
(1)
so, if tile sun gear has 40 and the annulus 80 teeth the gear ratio will be 40/(40 + 80) = 1/3. In other words, this arrangement will give a reduction of 3.0 : 1. However, if the drive is from A to S instead of S to A, the result will be an overdrive ratio of 0.33 9 1. Note that the planet gear, rotating in the direction opposite to that of the sun gear, is simply an idler and therefore has no influence on the gear ratio. If, on the other hand, the planet carrier C is locked, and the input is still the sun gear, Fig. 27.5(b), the output will, of course, be the annulus which is driven, through the idling planet gear, anti-clockwise. The gear ratio is" S/A
Brake on
Drive
Brake off
Locked
(a) Fig. 27.5
(2)
Brake off
Drive
(b)
Brake on
Locked
(c)
(d)
Epicyclic and pre-selector gearboxes
797
so S will drive P anti-clockwise and this, in turn, will drive the annulus anticlockwise at a speed of S/A times that of S, giving a ratio of 40/80 = 1/2 9 1. In other words, for one revolution of the sun gear, the annulus will rotate half a revolution (a reduction ratio of 2 : 1). With the sun gear locked, and the input from the planet carrier, Fig. 27.5(c), the annulus gear A is driven faster in the same direction. The gear ratio is:
(S + A)/A
(3)
the planet and annulus gears rotate in the same direction and the ratio is 120/80 = 1.5 : 1, so the annulus is rotated one and a half turns for every turn of the planet carrier. If the annulus gear A is locked and the planet carrier C is the input, Fig. 27.5 (d), the annulus is driven in the same direction. Again the gear ratio is: (S + A)/A
(4)
and the ratio is 1.5 : 1.
27.5 Simple planet epicyclic gearing in general From the foregoing, it can be concluded that, with an epicyclic gear train comprising an annulus, sun gear and planet carrier, any one of these can be fixed and the drive inputted through one of the other two, then the third will be driven at a different speed. The characteristics of such a gear train are as follows: 1. The output can be driven at a reduced speed relative to the input and in the same direction. 2. The output can be driven at a higher speed in the same direction. 3. The output can be driven at an alternative higher speed in the same direction. 4. The output can be driven at a lower speed than the input but in the opposite direction, to provide a reverse gear. 5. If any two of the gears locked together, the third cannot rotate relative to the others so the whole system turns as one solid mass, giving direct drive at a 1 : 1 ratio. 6. All ratios are dependent upon only the numbers of teeth on the sun and annulus gears, and are independent of the number of teeth on the planet gears. 7. With this arrangement, it is therefore possible to have direct drive, three forward gears, and one reverse gear. 27.6
Compound planet epicyclic gearing
An alternative to Fig. 27.3, in which two integral idler gears are used, is to have two independent but intermeshing 151anetary idler gears, one meshing with the sun and the other with the annulus, Fig. 27.6(a), in which case the gear ratio between these two comes into play. Consequently, if the annulus is fixed, the sun gear is the input and the planet cartier the output, the gear ratio is: P1/P2 • S/(S + A)
(5)
798
The Motor Vehicle P
P3
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(b)
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Fig. 27.6 Double planet pinnions so, if A had 80 teeth, P1 10, P2 20 and S 40, the overall ratio would be that of the two planet gears times that in equation (1), which is 10/20 x 40/(40 + 80) = 0.5 x 0.33 = 0.166 : 1. Note that equation (5) is equation (1) multiplied by the planet gear ratio. It follows that to find the ratios in the other cases, all that is necessary is to multiply the ratios obtained in equations (2), (3) and (4) by the planet gear ratio. However, because there are two intermeshing planet gears, the output will rotate in the opposite direction to that previously obtained. Indeed, all epicyclic gear trains fall into two categories: given that the planet carrier is locked, the first is when the input and output gears rotate in the same direction and the second when they rotate in opposite directions. In Fig. 24.6(b), the principle is the same, but one of the two gears is an integral pair, so we have three planet gears on two wheels, or a compound planet gear. This arrangement offers the possibility of axially offsetting the sun gear from the annulus so that a wider choice of planet ratios is obtainable than would be possible if two separate planets had to be accommodated between the periphery of the sun gear and the tooth ring in the annulus. With this three wheel arrangement and, as before, the annulus fixed and the sun wheel the input, the overall gear ratio is: P1/P2 • S/(S + A)
(6)
In this instance the planet gear P3 behaves similarly to that in a simple sunplanet-annulus train except that its speed of rotation is influenced by the meshing pair of planet gears with which it is associated. Consequently, the comments that followed equation (5), and those in the subsequent paragraph, also apply here.
27.7
Numbers of teeth
Assembly of an epicyclic gear train is practicable only if the numbers of teeth on the gears has been appropriately chosen. For simple planet epicyclic gearing, the sum of the numbers of teeth on the sun and annulus gears divided by the number of planet pinions Npp must be an integer. For compound planet epicyclic gearing, and when the direction of rotation of the input and output are the same, (P2 • A) - (S • P3) divided by Npp • HCF must be an integer, where Hcr is the highest common factor between P2 and S. If the relative direction of rotation is not the same, (P2 • A) + (S • P3) divided by Npp • must be an integer. During assembly of the gear train, a vernier effect may be experienced,
Epicyclic and pre-selector gearboxes
799
allowing the planet pinions to be inserted incorrectly within the backlash. This can be avoided if the numbers of teeth on the sunwheel and annulus are each divisible by the number of planet pinions, and datum teeth on the planets are marked.
27.8
Another way of applying epicyclic gearing
An alternative to stopping one of the elements of an epicyclic gear train completely is to reduce its speed of rotation, by connecting it to a resistance of some sort. An example already mentioned is the differential gear for splitting the torque equally between two halfshafts driving either the front or rear road wheels. A particularly interesting application of this principle was made in 1964 by E Perkins Ltd (now Varity Perkins). An auxiliary drive from an engine was geared up and taken through a differential train, to drive a supercharger. The engine drove the planet carder and the output was taken through the annulus see I.Mech.E Auto. Div. paper presented by Dawson and Hayward on 14 April 1964. In automotive practice, the brakes or clutches, either pneumatically or hydraulically actuated, are applied gradually to the appropriate gears. The object is to vary the relative speeds progressively during the ratio changes and thus to provide smooth gear shifts. In this way, the need for a separate pedal-actuated clutch is obviated. In modem transmissions, because they require much less attention in service, clutches have largely superseded brakes even though the latter are less costly.
27.9
Epicyclic gearboxes
For car gearboxes, forward gear ratios of from 1 9 1 to about 5 : 1 and one reverse are needed. Moreover, as explained in Section 25.10, the forward ratios must be arranged in a geometrical progression. Epicyclic gearboxes used to be commonly installed in Daimler, Lanchester and Armstrong Siddeley cars, mostly to provide preselection of gears and to eliminate the clutch, although the pedal control was retained for changing gear instead of operating the clutch. However, they have now been superseded by automatic transmissions with two pedal control. In heavy commercial vehicles in which optimum fuel economy is an overriding requirement, the inefficiency of a torque converter is unacceptable so epicyclic gearboxes arc still widely used. For this type of application, extra gear trains arc needed for providing large numbers of gear ratios needed for coping with the heavy loads carried. As previously stated, given that top speed is direct drive, seven different drives are obtainable from a simple train comprising a sun, annulus and carrier. However, this calls for complex construction and, moreover, some of the ratios are unsuitable. Consequently, even for cars, an alternative method is better. This entails interconnecting several epicyclic trains in series, so that suitable ratios can be selected. Generally, the first train is the primary one, the others modifying its ratio before transmitting the drive back through the primary planet carrier to the road wheels. An overall ratio of 1 9 1 is obtainable by locking the whole set of gear trains together. With such arrangements, the number of ratios obtainable rises rapidly with the number of trains although, as previously indicated, some are unsuitable for automotive gearbox applications.
800
The Motor Vehicle
The trains can be all of the same type or of different types. Several arrangements with spur gears are illustrated in Fig. 27.7, in which the brakes are labelled B and the clutches C: in practice multi-plate clutches are employed. The input is on the left and the output on the fight. In diagram (a) two simple trains are in tandem while at (b) two trains with double-sun gears are similarly arranged. By making some individual members function as part of two trains, the arrangement can be simplified, but fewer ratios are obtainable. An example is shown at (c), where the sun gears are integral and a single planet carrier serves both trains. With this arrangement, only one clutch is needed, but only three ratios are obtainable, as compared with four in (a). At (d) two of the suns $1 and $2 are integral and there is a third sun S with a carrier common to all three sets of planet gears. The sun S and its meshing planet gear serves both trains. Because $2 is larger than the driven sun S, reverse is obtained when brake B2 is applied and, since S1 is smaller than S, a forward drive ratio is obtained when brake B1 is applied. A third method of designing an epicyclic gearbox is to compound several simple epicyclic gear trains. This was the basis of the Wilson transmission, originally known as the Wilson-Pilcher gearbox, developed at the beginning of the 20th century. Later Vauxhall Motors worked with Major Wilson to develop it further but dropped it in 1927, when General Motors took the company over. Subsequent development was done by the Daimler company for their cars.
27.10
Basic principle of the Wilson gearbox
A Wilson type gearbox is illustrated in Fig. 27.8(a), and the functioning of its various secondary trains is shown in (b) to (f). The primary epicyclic gear train is common to all the ratios. Its sun S 1, Fig. 27.8(b), is driven by shaft D, which is coupled directly to the engine crankshaft, while its planet carrier C1 is coupled directly to the transmission line to the road wheels. In other
Fig. 27.7 Epicyclic gearbox arrangements
Epicyclic and pre-selector gearboxes
801
Fig. 27.8 Wilson gear ratios
words, these two elements are respectively the input and output regardless of ratio selected. The required gear ratios are obtained by driving the annulus at different speeds in relation to engine speed. How this is done is as follows. If engine speed were constant, at 1000 rev/min, and the annulus braked (zero rev/min), the speed of the carrier would be 1000 x S1/(A1 + S1) where, in general, A, S and C throughout what follows are the numbers of teeth on the annulus and sun wheel respectively, as in Section 27.9. If A1 = 100 and $1 = 25, the speed of C1 = 200 rev/min and the reduction ratio is 5 : 1. On the other hand if, with engine speed still at 1000 rev/min, the annulus is driven at 100 rev/min in the same direction as the engine by other elements in the compounded series of epicyclic gears, the speed of rotation of the sun gear relative to the annulus would be only 900 rev/min. So the speed of the planet carrier would be 900 x 25 = 22 500 divided by 125 = 180 rev/min but, as the carrier would rotate faster because the annulus was rotating, the 100 rev/min of the latter (in these circumstances not multiplied by the sun-toplanet ratio) must be added, giving 280 rev/min for the speed of the planet carrier. Therefore, when the annulus is stationary, the equation in the previous paragraphs is in fact 0 + (1000 x S1/A1 + S1) and, when the annulus is rotating at 100 rev/min, it becomes 100 + (1000 x S1/A1 + $1). In more general terms, the equation is (RE- CA) X (S1/A1 + S1) + RA, where RA and RE are respectively the speeds of rotation of the annulus and engine crankshaft. It follows that, if RE had been 200 rev/min, the output speed of the planet carrier would have been 380 rev/min.
802
The Motor Vehicle
If the annulus is driven in the opposite direction, its speed relative to the crankshaft becomes negative. So given an annulus rotating at a negative speed of 400 rev/min we have (1000 + 400) x (25/100 + 25) - 400 = - 120 rev/min, or in other words a reverse gear ratio of 0.12 : 1.
27.11
The auxiliary trains in the Wilson gearbox
As previously indicated, to drive the annulus at different speeds auxiliary epicyclic gear trains are used. Those for second gear are shown in Fig. 27.8(c). The sun gear $2, in common with S1 adjacent to it, is driven by the engine, the planet carder C2 is coupled to the annulus A1 and, to obtain second gear, a brake is applied to the annulus A2. So long as this brake is holding A2 stationary, the coupled carrier C2 and annulus A1 rotate in the same direction as the engine but at a lower speed. Consequently, C1 rotates at a higher speed than it did in first gear when, as described in the second paragraph of Section 24.10, A1 was the stationary element. The actual speed will, of course, depend on the numbers of teeth on the gears involved. To obtain third gear, Fig. 27.8(d), annulus A1 and therefore also the planet cartier C2 must be made to rotate faster than in second gear. This entails causing A2 to rotate in the same direction as the engine, by applying the brake to drum F to stop $3. The other two elements in the second gear train, C3 and A 3, are coupled respectively to A2 and C2. With $3 fixed and A 3 rotating in the same direction as the engine, C3, $2 and A2 will also be rotating in that direction, the last two because C3 is coupled to A2. Consequently, the speed of rotation of C2 must be greater than when A2 was fixed: therefore annulus A2 must be rotating faster than in second gear. Again, the actual speeds depend on the numbers of teeth on the gears. Direct drive, Fig. 27.8(e), is obtained by sliding the male cone G along the splines on the shaft D, and locking it in the female cone in the drum F which is fixed to $3. This locks the whole epicyclic gear assembly together, so that it rotates en bloc. Obtaining reverse gear, Fig. 27.8(f), entails bringing the fourth epicyclic train into operation by applying the brake to A 4. S u n S 4 is fixed to A1, and carrier C4 is fixed to the driven shaft D. Sun $1 is driven forwards by the engine, and the planet pinion, acting as an idler, drives the annulus A1, and with it the sun $4, backwards. Since A4 is braked, $4 drives the planet carrier C4 backwards too. Moreover, since both C1 and C4 are fixed to the output shaft to the road wheels, both have to rotate at the same speed in the same direction. This speed is in fact determined by the numbers of teeth on A1 and $1 relative to those on $4 and A n, the last mentioned pair determining the speed of rotation of the annulus A. This is not easy to visualise, so let us take an example in which S1 has 25, A1 100, $4 40 and A 4 80 teeth. To represent the conditions in reverse gear, apply the brake to A 4, and then rotate C4, and with it the integral carrier C1, backwards one revolution. Now consider the sequence of events as this rotary motion is transmitted through each of the intermeshing gear pairs in turn. With A 4 fixed, the planet carrier C4 will rotate $4, and with it the annulus A1, backwards 80/40 = 2 revolutions. This, in turn would have rotated S1 backwards 100/25 = 4 revolutions which, at first sight, appears to make an overall ratio of 8 : 1. However, the driven shaft has rotated minus 1 turn and the driving shaft plus 8 turns, so the overall ratio is in fact 7 : 1.
Epicyclic and pre-selector gearboxes
27.12
803
The clutches and brakes in the Wilson gearbox
To simplify in Fig. 27.8(e) a cone clutch was shown, but this, of course, would be too harsh for the purpose. In practice, a pneumatically or hydraulically actuated multi-plate clutch such as that illustrated in Fig. 27.9 is employed. The clutch illustrated is air actuated. A control valve lets fluid under pressure enter cylinder A beneath the piston, to actuate the lever B. This moves the ring C to the right, about its pivot D. Mounted on gimbal pivots F (one on each side) within ring C is the housing E for the ball thrust bearing. Consequently, the axial displacement of the centre of C causes the ball thrust bearing to compress the clutch plates between the presser plate G and the clutch hub H, to which are splined the driving plates of the clutch. The drum K, to which the driven plates are splined, is in turn splined to the hub of the sun gear $1. When the fluid pressure is released by the control valve, the clutch is released by a set of coil springs, equally spaced around the hub (only one can be seen in the illustration). A typical brake, Fig. 27.10, for acting upon the periphery of the annulus in this type of gearbox has an outer and an inner band, A and B respectively, having friction linings L. The outer band is anchored by the hooked link D, and the inner one by the lug F projecting through a slot in the outer band. The reactions to the brake torque, at these two diametrically opposite anchorage points, are equal and opposite and therefore do not add to the load on the bearings on which the annulus rotates. A tie rod G pulls the outer band on to the inner band, and thus the inner band on to the drum, which is the periphery of the annulus. Note that the pull of the tie rod on the lower end of the brake band is reacted by an equal and opposite pull applied by the hook on its upper end, and the direction of rotation relative to the fixed anchorage point is such that the brake band is automatically wrapped around the drum. Stop C actuates an automatic adjustment device for compensating for wear. The actuation mechanism is shown in more detail in Fig. 27.11 and the automatic adjustment mechanism in Fig. 27.12. Fluid pressure, in this case
Fig. 27.9
Fig. 27.10
804
Fig. 27.11
The Motor Vehicle
Fig. 27.12
air, on the piston A lifts the piston rod Q in Fig. 27.11. This, in turn, rotates the lever B about its pivot O. Roller C, on the other end of the lever, moves along the cam-shaped lower surface of the lever K, termed the bus bar. As it does so, it lifts the tie rod, G in Fig. 27.10, to apply the brake. The shape of the cam is such that the initial movement of the lever is rapid, to take up the clearance between the drum and brake bands. Subsequently, the slope of the cam is less steep so that the tie lifts more slowly, and the ratio of force on the piston to the leverage on the tie is therefore greater. This effect is enhanced by the fact that, as the roller approaches the end of bus bar K, the lever B on which it is mounted comes up to its top dead centre position, relative to the axis P of the pivot.
27.13
Automatic compensation for wear
In the upper diagram of Fig. 27.12, the adjustment device is in the 'brake off' position and in the lower one it is in the 'brake on position' with zero wear on the lining. Surrounding the round nut screwed on the top of the tie rod G is a coil spring B. One end of this spring is fixed to a pin projecting upwards from plate A. It is then coiled several times round the nut and secured to a second pin, which is fixed to the knife-edge pivot plate J and projects upwards through a slot in plate A. Plate A is free to rotate relative to the nut H. Each time the bus bar K, Fig. 27.11, is pulled upwards to apply the brake, the upper end of the tie rod, and with it the plate-and-spring assembly, moves over to the left until the lug on plate A just makes contact with the adjacent stop C, which is C in Fig. 27.10. For all gears except top, where it is fixed to the gearbox casing, this stop is mounted on the brake band. When wear has occurred, the stop goes further than just contacting the lug: it strikes it and, deflecting it, rotates the plate A anti-clockwise. This uncoils the spring around the nut H, which therefore is then free to rotate, except that the friction between it and its conical seating in the pivot plate prevents it from doing so. When the brake is released, however, the coils tighten around the nut. Then, as the tie rod is lowered and the whole assembly retracts to the fight in the illustration, the plate rotates back to its original 'brake off' position, and the lug on the other side of the plate comes up against the stop D fixed to the casing. At this point, the coil spring grips the
Epicyclic and pre-selector gearboxes
805
nut and, overcoming the friction between it and the plate A, rotates it to reduce the effective length of the tie rod, and thus takes up the clearance due to wear between the brake lining and drum. Wilson type gearboxes are still widely installed in public service vehicles although, in most instances, gear shifting is effected either semi- or fully automatically by means of an electronic control system. This ensures that all gear changes and moving away from rest are effected smoothly. In the semiautomatic systems, the driver uses a small lever to select the gears, but the actual shifting and pressures applied to the actuation mechanisms are effected automatically in relation to the gear selected and factors such as vehicle speed, engine speed and torque being transmitted. Once a gear is fully engaged, the pressure actuating the brake and clutch is increased to the maximum, for prevention of slippage.