Advanced Vehicle Technology 2 Episode 3 doc

Fig. 3.9 (a±d) Split baulk ring synchromesh unit 72 The axial thrust applied by the gear stick to the sliding sleeve will now be sufficient to compress the split synchronizing ring and subsequently permits the sleeve to slide over the gear wheel dog teeth for full engagement (Fig. 3.9(d)). 3.4 Remote controlled gear selection and engagement mechanisms Gear selection and engagement is achieved by two distinct movements: 1 The selection of the required gear shift gate and the positioning of the engagement gate lever. 2 The shifting of the chosen selector gate rod into the engagement gear position. These two operations are generally performed through the media of the gear shift lever and the remote control tube/rod. Any transverse move- ment of the gear shift lever by the driver selects the gear shift gate and the engagement of the gate is obtained by longitudinal movement of the gear shift lever. Movement of the gear shift lever is conveyed to the selection mechanism via the remote control tube. Initially the tube is twisted to select the gate shift gate, followed by either a push or pull movement of the tube to engage the appropriate gear. For the gear shift control to be effective it must have some sort of flexible linkage between the gear shift lever supported on the floor of the driver's compartment and the engine and transmission inte- gral unit which is suspended on rubber mountings. This is essential to prevent engine and transmission vibrations being transmitted back to the body and floor pan and subsequently causing discomfort to the driver and passengers. 3.4.1 Remote controlled double rod and bell cranked lever gear shift mechanism, suitable for both four and five speed transverse mounted gearbox (Talbot) (Fig. 3.10) Twisting the remote control tube transfers move- ment to the first selector link rod. This motion is then redirected at right angles to the transverse gate selector/engagement shaft via the selector relay lever (bell crank) to position the required gear gate (Fig. 3.10). A forward or backward movement of the remote control tube now conveys motion via the first engagement relay lever (bell crank), engagement link rod and second relay lever to rotate the transverse gate selector/engage- ment shaft. Consequently, this shifts the transverse selector/engagement shaft so that it pushes the synchronizing sliding sleeve into engagement with the selected gear dog teeth. 3.4.2 Remote controlled bell cranked lever gear shift mechanism for a four speed transverse mounted gearbox (Ford) (Fig. 3.11) Gear selection and engagement movement is conveyed from the gear shift lever pivot action to the remote control rod universal joint and to the control shift and relay lever guide (Fig. 3.11). Rocking the gear shift lever transversely rotates the control shaft and relay guide. This tilts the selector relay lever and subsequently the selec- tion relay lever guide and shaft until the striker finger aligns with the chosen selector gate. A fur- ther push or pull movement to the gear shift lever by the driver then transfers a forward or backward motion via the remote control rod, con- trol shaft and relay lever guide to the engagement relay lever. Movement is then redirected at right angles to the selector relay guide and shaft. Engagement of the gear required is finally obtained by the selector/engagement shaft forcing the strik- ing finger to shift the gate and selector fork along the single selector rod so that the synchron- izing sleeve meshes with the appropriate gear wheel dog clutch. Fig. 3.10 Remote controlled double rod and bell crank lever gearshift mechanism suitable for both four and five speed transversely mounted gearbox 73 3.4.3 Remote controlled sliding ball joint gear shift mechanism suitable for both four and five speed longitudinal or transverse mounted gearbox (VW) (Fig. 3.12) Selection and engagement of the different gear ratios is achieved with a swivel ball end pivot gear shift lever actuating through a sliding ball relay lever a single remote control rod (Fig. 3.12). The remote control rod transfers both rotary and push- pull movement to the gate selector and engagement shaft. This rod is also restrained in bushes between the gear shift lever mounting and the bulkhead. It thus permits the remote control rod to transfer both rotary (gate selection) and push-pull (select rod engagement shift) movement to the gate selector and engagement shaft. Relative movement between the suspended engine and transmission unit and the car body is compensated by the second sliding ball relay lever. As a result the gate engagement striking finger is able to select and shift into engagement the appropriate selector rod fork. This single rod sliding ball remote control linkage can be used with either longitudinally or transversely mounted gearboxes, but with the latter an additional relay lever mechanism (not shown) is needed to convey the two distinct movements of selection and engagement through a right angle. Fig. 3.11 Remote controlled bell crank level gear shift mechanism for a four speed transversely mounted gearbox Fig. 3.12 Remote controlled sliding ball joint gear shift mechanism suitable for both four and five speed longitudinally or transversely mounted gearbox 74 3.4.4 Remote controlled double rod and hinged relay joint gear shift mechanism suitable for both four and five speed longitudinal mounted gearbox (VW) (Fig. 3.13) With this layout the remote control is provided by a pair of remote control rods, one twists and selects the gear gate when the gear shift lever is given a transverse movement, while the other pushes or pulls when the gear shift lever is moved longitudin- ally (Fig. 3.13). Twisting movement is thus con- veyed to the engagement relay lever which makes the engagement striking finger push the aligned selector gate and rod. Subsequently, the synchro- nizing sleeve splines mesh with the corresponding dog clutch teeth of the selected gear wheel. Relative movement between the gear shift lever swivel sup- port and rubber mounted gearbox is absorbed by the hinged relay joint and the ball joints at either end of the remote control engagement rod. 3.4.5 Remote controlled single rod with self aligning bearing gear shift mechanism suitable for both five and six speed longitudinal mounted gearbox (Ford) (Fig. 3.14) A simple and effective method of selecting and engaging the various gear ratios suitable for commercial vehicles where the driver cab is for- ward of the gearbox is shown in Fig. 3.14. Movement of the gear shift lever in the usual transverse and longitudinal directions provides both rotation and a push-pull action to the remote control tube. Twisting the remote control tube transversely tilts the relay gear shift lever about its ball joint so that the striking finger at its lower end matches up with the selected gear gate. Gear engagement is then completed by the driver tilting the gear shift lever away or towards himself. This permits the remote control tube to move axially through the mounted self-aligning bearing. As a result, a similar motion will be experienced by the relay gear shift lever, which then pushes the striking finger, selector gate and selector fork into the gear engaged position. It should be observed that the self-aligning bearing allows the remote control tube to slide to and fro. At the same time it permits the inner race member to tilt if any relative movement between the gearbox and chassis takes place. 3.4.6 Remote controlled single rod with swing arm support gear shift mechanism suitable for five and six speed longitudinally mounted gearbox (ZF) (Fig. 3.15) This arrangement which is used extensively on commercial vehicles employs a universal cross- pin joint to transfer both the gear selection and Fig. 3.13 Remote controlled double rod and hinged relay joint gear shift mechanism suitable for both four and five speed longitudinally mounted gearbox Fig. 3.14 Remote controlled single rod with self-aligning bearing gear shift mechanism suitable for both five and six speed longitudinally mounted gearbox 75 engagement motion to the remote control tube (Fig. 3.15). Twisting this remote control tube by giving the gear shift lever a transverse movement pivots the suspended selector gate relay lever so that the transverse gate selector/engagement shift moves across the selector gates until it aligns with the selected gate. The gear shift lever is then given a to or fro movement. This causes the transverse selector/engagement shaft to rotate, thereby for- cing the striking finger to move the selector rod and fork. The synchronizing sleeve will now be able to engage the dog clutch of the appropriate gear wheel. Any misalignment between the gear shift lever support mounting and the gear shift mechanism forming part of the gearbox (caused by engine shake or rock) is thus compensated by the swing rod which provides a degree of float for the selector gate relay lever pivot point. 3.5 Splitter and range change gearboxes Ideally the tractive effect produced by an engine and transmission system developing a constant power output from rest to its maximum road speed would vary inversely with its speed. This characteristic can be shown as a smooth declining tractive effect curve with rising road speed (Fig. 3.16). In practice, the transmission has a fixed number of gear ratios so that the ideal smooth tractive effect curve would be interrupted to allow for loss Fig. 3.15 Remote controlled single rod with swing arm support gear shift mechanism suitable for five and six speed longitudinally mounted gearbox Fig. 3.16 Ideal and actual tractive effort-speed characteristics of a vehicle 76 of engine speed and power between each gear change (see the thick lines of Fig. 3.16). For a vehicle such as a saloon car or light van which only weighs about one tonne and has a large power to weight ratio, a four or five speed gearbox is adequate to maintain tractive effect without too much loss in engine speed and vehicle performance between gear changes. Unfortunately, this is not the situation for heavy goods vehicles where large loads are being hauled so that the power to weight ratio is usually very low. Under such operating conditions if the gear ratio steps are too large the engine speed will drop to such an extent during gear changes that the engine torque recovery will be very sluggish (Fig. 3.17). Therefore, to minimize engine speed fall-off whilst changing gears, smaller gear ratio steps are required, that is, more gear ratios are necessary to respond to the slightest change in vehicle load, road conditions and the driver's requirements. Figs 3.2 and 3.18 show that by dou- bling the number of gear ratios, the fall in engine speed between gear shifts is much smaller. To cope with moderate payloads, conventional double stage compound gearboxes with up to six forward speeds are manufactured, but these boxes tend to be large and heavy. Therefore, if more gear ratios are essential for very heavy payloads, a far better way of extending the number of gear ratios is to utilize a two speed auxiliary gearbox in series with a three, four, five or six speed conventional compound gearbox. The function of this auxiliary box is to double the number of gear ratios of the conventional gearbox. With a three, four, five or six speed gearbox, the gear ratios are increased to six, eight, ten or twelve respectively (Figs 3.2 and 3.18). For very special applications, a three speed auxiliary gearbox can be incorporated so that the gear ratios are trebled. Usually one of these auxiliary gear ratios provides a range of very low gear ratios known as crawlers or deep gears. The auxiliary gearbox may be situated either in front or to the rear of the conventional compound gearbox. The combination of the auxiliary gearbox and the main gearbox can be designed to be used either as a splitter gear change or as a range gear change in the following way. 3.5.1 Splitter gear change (Figs 3.19 and 3.20) With the splitter arrangement, the main gearbox gear ratios are spread out wide between adjacent gears whilst the two speed auxiliary gearbox has one direct gear ratio and a second gear which is either a step up or down ratio (Fig. 3.19). The auxiliary second gear ratio is chosen so that it splits the main gearbox ratio steps in half, hence the name splitter gear change. The splitter indirect gear ratio nor- mally is set between 1.2 and 1.4:1. A typical ratio would be 1.25:1. A normal upchange sequence for an eight speed gearbox (Fig. 3.20), consisting of a main gearbox with four forward gear ratios and one reverse and a two speed auxiliary splitter stage, would be as follows: Auxiliary splitter low ratio and main gearbox first gear selected results in `first gear low' (1L); auxiliary splitter switched to high ratio but with the main gear- box still in first gear results in `first gear high' (1H); Fig. 3.17 Engine speed ratio chart for a vehicle employing a five speed gearbox Fig. 3.18 Engine speed ratio chart for a vehicle employing either a ten speed range change or a splitter change gearbox 77 splitter switched again to low ratio and main gear- box second gear selected results in 2L; splitter switched to high ratio, main gearbox gear remaining in second gives 2H; splitter switched to low ratio main gearbox third gear selected gives 3L; splitter switched to high ratio main gearbox still in third gives 3 H. This procedure continues throughout the upshift from bottom to top gear ratio. Thus the overall upshift gear ratio change pattern would be: Gear ratio 1 2 3 4 5 6 7 8 Reverse Upshift sequence 1L 1H 2L 2H 3L 3H 4L 4H RL RH It can therefore be predicted that alternate changes involve a simultaneous upchange in the Fig. 3.19 Eight speed constant mesh gearbox with two speed front mounted splitter change Fig. 3.20 Splitter change gear ratio±speed chart 78 main gearbox and downchange in the splitter stage, or vice versa. Referring to the thick lines in Figs 3.2, 3.17 and 3.18, these represent the recommended operating speed range for the engine for best fuel economy, but the broken lines in Fig. 3.17 represent the gear shift technique if maximum road speed is the criteria and fuel consumption, engine wear and noise become secondary considerations. 3.5.2 Range gear change (Figs 3.21 and 3.22) In contrast to the splitter gear change, the range gear change arrangement (Fig. 3.21) has the gear ratios between adjacent gear ratio steps set close together. The auxiliary two speed gearbox will have one ratio direct drive and the other one normally equal to just over half the gear ratio spread from bottom to top. This is slightly larger than the main gearbox gear ratio spread. To change from one gear ratio to the next with, for example, an eight speed gearbox comprising four normal forward gears and one reverse and a two speed auxiliary range change, the pattern of gear change would be as shown in Fig. 3.22. Through the gear ratios from bottom to top `low gear range' is initially selected, the main gear- box order of upchanges are first, second, third and fourth gears. At this point the range change is moved to `high gear range' and the sequence of gear upchanges again becomes first, second, third and fourth. Therefore the total number of gear ratios is the sum of both low and high ranges, that is, eight. A tabulated summary of the upshift gear change pattern will be: Gear ratio 12345 6 7 8 Reverse Upshift sequence 1L 2L 3L 4L 1H 2H 3H 4H RL RH 3.5.3 Sixteen speed synchromesh gearbox with range change and integral splitter gears (Fig. 3.23) This heavy duty commercial gearbox utilizes both a two speed range change and a two speed splitter gear change to enable the four speed gearbox to Fig. 3.21 Eight speed constant mesh gearbox with two speed rear mounted range change 79 extend the gear ratio into eight steps and, when required, to sixteen split (narrow) gear ratio intervals. The complete gearbox unit can be considered to be divided into three sections; the middle section (which is basically a conventional double stage four speed gearbox), and the first two pairs of gears at the front end which make up the two speed splitter gearbox. Mounted at the rear is an epicyclic gear train providing a two speed low and high range change (Fig. 3.23). The epicyclic gear train at the rear doubles the ratios of the four speed gearbox permitting the driver to initially select the low (L) gear range driving through this range 1, 2, 3 and 4 then select- ing the high (H) gear range. The gear change sequence is again repeated but the gear ratios now become 5, 6, 7 and 8. If heavy loads are being carried, or if maximum torque is needed when overtaking on hills, much closer gear ratio intervals are desirable. This is provided by splitting the gear steps in half with the two speed splitter gears; the gear shift pattern of 1st low, 1st high, 2nd low, 2nd high, 3rd low and so on is adopted. Front end splitter two speed gearbox power flow (Fig. 3.23) Input power to the gearbox is supplied to the first motion shaft. When the splitter synchro- nizing sliding sleeve is in neutral, both the splitter low and high input gear wheels revolve on their needle bearings independently of their supporting first motion shaft and mainshaft respectively. When low or high splitter gears are engaged, the first motion shaft drive hub conveys power to the first or second pair of splitter gear wheels and hence to the layshaft gear cluster. Mid-four speed gearbox power flow (Fig. 3.23) Power from the first motion shaft at a reduced speed is transferred to the layshaft cluster of gears and subsequently provides the motion to all the other mainshaft gear wheels which are free to revolve on the mainshaft, but at relatively different speeds when in the neutral gear position. Engagement of one mid-gearbox gear ratio dog clutch locks the corresponding mainshaft drive hub to the chosen gear so that power is now able to pass from the layshaft to the mainshaft through the selected pair of gear wheels. Reverse gear is provided via an idler gear which, when meshed between the layshaft and mainshaft, alters the direction of rotation of the mainshaft in the usual manner. Rear end range two speed gearbox power flow (Fig. 3.23) When the range change is in the neu- tral position, power passes from the mainshaft and sun gear to the planet gears which then revolve on the output shaft's carrier pin axes and in turn spin round the annular gear and synchronizing drive hub. Engaging the low range gear locks the synchron- izing drive hub to the gearbox casing. This forces the planet gears to revolve and walk round the inside of the annular gear. Consequently, the carrier and output shafts which support the planet gear axes will also be made to rotate but at a speed lower than that of the input shaft. Changing to high range locks the annular gear and drive hub to the output shaft so that power flow from the planet gears is then divided between the carrier and annular, but since they need to rotate at differing speeds, the power flow forms a closed loop and jams the gearing. As a result, there is no gear reduction but just a straight through drive to the output shaft. 3.5.4 Twin counter shaft ten speed constant mesh gearbox with synchromesh two speed rear mounted range change (Fig. 3.24) With the quest for larger torque carrying capacity, closer steps between gear ratio changes, reduced gearbox length and weight, a unique approach to fulfil these requirements has been developed Fig. 3.22 Range change gear ratio±speed chart 80 Fig. 3.23 Sixteen speed synchromesh with range change and integral splitter gears 81 [...]... planet gear teeth Overdrive gear ratio ˆ Fig 3 .27 ˆ 75 ˆ 0:78 125 96 b) A ˆ S ‡ 2P Therefore Pˆ AÀS 2 ˆ ˆ Simple epicycle overdrive gear train 87 75 À 21 2 54 ˆ 27 teeth 2 30 00 ˆ 38 40 rev/min 0:78 125   38 40 À 30 00 d) Percentage of overdrive ˆ 100 30 00 c) Output speed ˆ ˆ b) A ˆ PL ‡ PS ‡ S ˆ 21 ‡ 24 ‡ 15 ˆ 60 teeth c) Output speed ˆ 840  100 ˆ 28 % 30 00 Example 2 A compound epicyclic gear train overdrive... constant K can be determined by:  GT Kˆ GB a) 1st gear ratio G1 ˆ 20 :0:1 2nd gear ratio G2 ˆ G1 K ˆ 20  0:7 ˆ 14:0X1 3rd gear ratio G3 ˆ G1 K 2 ˆ 20  0: 72 ˆ 9:8X1 4th gear ratio G4 ˆ G1 K 3 ˆ 20  0: 73 ˆ 6:86X1 5th gear ratio G5 ˆ G1 K 4 ˆ 20  0:74 ˆ 4:8X1  1À1 nG So ; GT ˆ GB K nG b) G1 ˆ 9:8 ˆ 2: 042X1 4:8 G4 ˆ For commercial vehicles, the gear ratios in the gearbox are often arranged in geometric... blockage, ˆ 10Cr W ‡ CD AV 2 where Cr ˆ coefficient of rolling resistance W ˆ gross vehicle weight (kg) CD ˆ coefficient of aerodynamic resistance (drag) A ˆ projected frontal area of vehicle (m2 ) V ˆ speed of vehicle (km/h) 93 Fig 3. 31 range Forces opposing vehicle motion over its speed Fig 3. 32 Relationship of power developed and road power required over the vehicle' s speed range 3. 8 .2 Setting bottom gear... ‡ S) a) Overdrive ˆ gear ratio PL (PL ‡ PS ‡ S) ‡ PS S 24 (24 ‡ 15 ‡ 21 ) 24 (24 ‡ 15 ‡ 21 ) ‡ (15  21 ) ˆ 4878 À 4000  100 4000 ˆ 21 :95% The overdrive gear ratio, the number of annulus ring gear teeth, the annulus ring and output shaft speed, the percentage of overdrive ˆ 4000 ˆ 4878 rev/min 0: 82 24  60 1400 ˆ (24  60) ‡ 31 5 1755 ˆ 0: 82 Fig 3 .28 Compound epicycle gear train 88 the annulus and therefore... economy Consider the engine to vehicle speed characteristics for each gear ratio as shown (Fig 3. 35) When changing gear the engine speed will drop from the highest NH to the lowest NL without any change in road speed, i.e V1 , V2 , V3 etc Let G1 G2 G3 G4 G5 where Also ; G3 ˆ G2 and NH NL ˆ G3 G4 ; G4 ˆ G3 NH NL ˆ G4 G5 ; G5 ˆ G4 ˆ 1st overall gear ratio ˆ 2nd overall gear ratio ˆ 3rd overall gear ratio ˆ... the ratio of the rolling resistance to the normal load on the tyre i:e: GF GB M E T R R Cr ˆ Wr Rr ˆ 10Cr W ˆ 10  0:015  150 ˆ 22 5N Rg ˆ ˆ 10W G 10  1500 ˆ 37 50N 4 E ˆ Rr ‡ Rg ˆ 37 50 ‡ 22 5 ˆ 39 75N GB ˆ Fig 3. 33 Gradient resistance to motion ˆ eR TGF M 39 75  0 :27 ˆ 3: 1X1 100  4:07  0:85 Tractive effort ˆ Resisting forces opposing motion EˆR ˆ Rr ‡ Rg (N) where E ˆ tractive effort (N) R ˆ resisting...  3: 1 42  0:54  6000 ˆ 150 GF ˆ 10W or 10W sin  G where Rg ˆ gradient resistance (N) Rg ˆ W ˆ gross vehicle weight (10W kg ˆ WN) ˆ 4:07 X 1 G ˆ gradient (1 in x) ˆ sin  94 Table 3 .2 resistance Average values of coefficient of rolling Vehicle type where Coefficient of rolling resistance (Cr) Concrete Passenger Car 0.015 Trucks 0.0 12 Tractors 0. 02 Medium hard soil Sand 0.08 0.06 0.04 0 .30 0 .25 0 .20 ... (rev/min) second gear G2 ˆ G2 These wheel speeds must be equal for true rolling Hence NH NL ˆ G2 G3 gear G3 ˆ G2 NL NH Hence the ratios form a geometric progression Fig 3. 35 Gear ratios selected on geometric progression 96 The following relationship will also apply for a five speed gearbox: Example A transmission system for a vehicle is to have an overall bottom and top gear ratio of 20 :1 and 4.8 respectively... gear ratio ; G2 ˆ G1 NL NH NL NH NL is known as the minimum to maxNH imum speed range ratio K for a given engine Now, gear G2 ˆ G1 Overall ˆ Engine speed (rev/min) gear ratio Road wheel speed (rev/min) NH NL ˆ G1 G2 NL NH The ratio since NL ˆ G1 K, NH NL ˆ k (a constant) NH NL ˆ G2 K ˆ (G1 K)K NH ˆ G 1 K2 NL gear G4 ˆ G3 ˆ G3 K ˆ (G1 K 2 )K NH ˆ G 1 K3 NL gear G5 ˆ G4 ˆ G4 K ˆ (G1 K 3 )K NH ˆ G 1 K4... (undergearing) used for cars, vans, coaches and commercial vehicles varies from as little as 15% to as much as 45% This corresponds to undergearing ratios of between 0.87:1 and 0.69:1 respectively Typical overdrive ratios which have been frequently used are 0. 82: 1 (22 %), 0.78:1 (28 %) and 0.75:1 (37 %) Simple gear train (Fig 3 .27 ) A S‡S A ˆ S ‡ 2P A ˆ number of annulus ring gear teeth S ˆ number of sun . P S  S)  P S S  24 (24  15  21 ) 24 (24  15  21 ) (15  21 )  24 Â60 (24 Â60)  31 5  1400 1755  0: 82 b) A  P L  P S  S  21  24  15  60 teeth c) Output speed  4000 0: 82  4878 rev/min d) Percentage. S  75 75 21  75 96  0:78 125 b) A  S  2P Therefore P  A ÀS 2  75 21 2  54 2  27 teeth Fig. 3 .27 Simple epicycle overdrive gear train 87 c) Output speed  30 00 0:78 125  38 40 rev/min d). be: Gear ratio 1 23 45 6 7 8 Reverse Upshift sequence 1L 2L 3L 4L 1H 2H 3H 4H RL RH 3. 5 .3 Sixteen speed synchromesh gearbox with range change and integral splitter gears (Fig. 3 . 23 ) This heavy duty