GB1559493A - Change speed gear boxes - Google Patents

Description

(54) IMPROVEMENTS IN AND RELATING TO CHANGE SPEED GEAR BOXES (71) We, S. R. M. HYDROMEKANIK AB, a Swedish Company, of Box 16, Stockholm Vallingbv 1, Sweden, do hereby declare the invention, for which we pray that a patent may be granted to us, and the method by which it is to be performed, to be particularly described in and by the following statement: The irvention relates to a variable power transmission device for motor vehicles fitted with a rlulti-stage mechanical gear change.

Conventionally such power transmission devices include speed-change gear wheels in the form of pinion gears and, in order to change gear, clutches associated with the pinion gears are actu ted under the control of a gear selector lever, via a mechanical linkage. The use of synchror isation devices for the clutches has certainly rendered unnecessary laborious double-ezclutching when changing down through the gears. However, to achieve synchror isation of the clutches, the drive connections need a certain amount of "gear-ch Inge time"and if a driver disregards the"gear-change time"and"crashes through'the gears, the inevitable result will be not only rapid wear of the gears and synchror isation device, but also early failure or breakdown of the gears and the device.

In the case of vehicle transmissions used in conjunction with hydrodynamic torque converte s, compou id gear boxes are known which are constructed as planetary gears, assemble d from sev ral sets of epicyclic gears. In such a transmission, changing from one speed tc another is carried out with the aid of servo-actuated brakes and/or clutches by means o which one) f three drive elements is stopped or separate drive elements connected together. In order to obtain a sufficient number of gear stages, which is particularly important for heavy trucks and buses and even more important for earth-moving vehicles, this type of compound gearing is very expensive and cumbersome due to the number of gear sets required. Further, for satisfactory operation of gear changes compound gearing requires complicated controls which increase the likelihood of breakdown, and also requires considerable"gear-change time"In addition, fluid pumps with high output are necessary for the controls because motors incorporated in such transmissions generally operate with rotary seals which do not allow high pressures to be used and which possess inherent high losses due to leakage.

It is therefore, an object of the present invention to create a power transmission device for vehicles fitted with a multi-stage mechanical gear-change which can be used both in the conventional manner without a torque converter and also in conjunction with a torque converter, with simple and reliable construction and the lease possible loss, and which allows the shortest possible"gear-change time"both when changing up and when changing down through the gears.

According to this invention there is provided a variable power transmission device for motor vehicles fitted with a multi-stage mechanical gear change wherein the gear change includes planetary gearing with at least two planet wheels of different diameters meshing with associated sun and/or ring gears which can be frictionally braked or clutched into a drive line, each brake or clutch having an associated fluid pressure operated servo-motor for actuating the same and each servo-motor having a non-rotatable housing, a multi-position control valve connected via a non-return valve with a low pressure fluid system, the multi-position valve serving to selectively control operation of the individual brakes and clutches, a maximum pressure valve in each connection between the multi-position control valve and the respective servo motors, and a low capacity high pressure pump for engaging the brakes, and clutches, the low capacity high pressure pump being connected to supply fluid to the multi-position control valve at a position between the multi-position control valve and the non-return valve.

The use of a planetary gear according to the invention and including at least two epicyclic gear sections of different diameter and associated drive components which can be frictionally braked and/or clutched (for which all braking and clutching is actuated using servo-motors having non-rotatable housings) enables a transmission to be made having comparatively small overall dimensions. Furthermore, the use of non-rotatable housings considerably reduces sealing problems in the servo motors thereby reducing, if not eliminating, the usual leakage losses. This invention has an advantage in that it allows the servo motors to be connected through a selectively opening multi-position control valve to an existing pressure fluid system (such as, the lubricating medium system of the drive motor or the filling pump of a hydrodynamic torque converter) which supplies only that quantity of pressure fluid which is required to fill the servo motor which is connected at any particular time, whereas a supplementary high pressure pump produces the holding pressure necessary for the satisfactory engagement of the friction-contact brakes and clutches after the filling process. In this respect the non-return valve which is connected is parallel with the h', gh pressure pump, automatically controls the amount of pressure fluid required to fill the servo motors, in that it opens against a counter pressure which remains low during filling and after filling, shuts off the back-flow of the servo motors which are provided with the holding pressure by the high pressure pump. Thus, owing to the use of non-rotatable servo motors and the consequent easily solved sealing problems permitting low capacity servo motors to be used, the high pressure pump need only compensate for small leakage fosses. It will, therefore, be appreciated that in this way a power transmission device is created which is of simple construction and of compact dimensions and which can be used to advantage both in vehicles without automatic transmission and with automatic transmission, in particular with a hydrodynamic torque converter operating without steps.

Further, the pressure fluid control of the gear change device according to the invention permits location thereof in many different positions and at the same time allows a manufacturer considerable freedom for the construction of the control itself, irrespective of whether the selection of the speeds is carried out by hand or depending on different parameters of the power transmission device, such as, the ratio between the number of revolutions of the drive input shaft and that of the output shaft, the turning moment requirement or some other parameter.

It is a feature of this invention that each servo motor is connected to a pressure-free return duct via its own maximum pressure value which is pre-set to operate at its own individual and characteristic opening pressure. This feature allows the establishment of the connecting pressure on the one hand and holding pressure on the other hand of each of the servo motors to be maintained at an optimum value even with the use of a common hydraulic system and a common high-pressure holding pump for all the servo motors. This facility can be further improved by incorporating means for individually adjusting the opening pressure of the maximum pressure valves, Due to the small capacity of the high pressure holding pump, the removal of small amounts of fluid is sufficient for effective pressure limitation so that the maximum pressure valves can be constructed correspondingly small.

In many cases, t ie use of planetary gear with drive components which can be frictionally braked and/or clui-hed permits the use of a traditional separating clutch to be eliminated since the function thereof has been assumed by the individual brakes and clutches. In m my instances, however, it may be desirable for the brakes and clutches of the planetary gears to be connected whilst under no load for which, according to another feature of the invention, a separating clutch is included in a known way between the drive motor and the planetary gears.

As mentioned, the invention can be applied advantageously to conventional power transmission devices with the exclusion of a multi-stage mechanical gear changing device.

However, it is a particular feature of the invention that the planetary gear of the invention is used in conjunction with a hydrodynamic torque converter and in such a case the torque converter is arranged between the drive motor and the planetary gear. In such a case, it may be desirable to connect the brakes and ctutches of the planetary gears when the former are not under load and this may be achieved using a separating clutch disposed outside the planetary gears. However, it is particularly advantageous if the torque converter has a releasable torque transmitting bladed component (e. g. pump or turbine) which transmits the moment and forms the actual separating clutch. This releasable pump or turbine component can be connected or released from an associated shaft in a known way by reversing the flow of the hydraulic fluid through the toroidal-shaped working chamber of the torque converter chamber via a friction clutch. In order to eliminate the converter under normal driving conditions, an addition direct clutch can be provided in a known way for bv-passing the converter and coupled expediently to a release device for the torque-transmitting bladed component in such a way that the torque converter is made inoperative when the direct clutch is engaged, i. e. the circulation of fluid in the toroidal-shaped converter chamber is brought to a standstill. This can also be achieved by releasing the torque-transmitting bladed component of the converter at the desired rotation when the direct clutch is engaged.

The power transmission device according to the invention is suitable for use with vehicles without automatic transmission, in which therefore changing gear is controlled manually. In such cases, and according to another feature of the invention, the multi-position control valve can be manually operated by means of a lever which runs in a slideway, which preferably includes a neutral position track with finger-shaped tracks branching out from it for the individual connecting positions of the mechanical gear changing device. If desired and, in conjunction with an additional separating clutch (which may be in the form of a separate clutch or a releasable converter impeller component), the lever can further be fitted with a connecting device for operating the separating clutch in such a way that this opens when the lever is in the neutral position track. This arrangement can be still further improved when used in conjunction with a hydrodynamic torque converter which can be by-passed by a direct clutch by connecting in hydraulic drive over part of the finger-shaped tracks adjoining the neutral position track, and by connecting in direct drive over at least part of these tracks in the subsequent terminal area. In this arrangement, the positive disconnection of the hydrodynamic torque converter at each gear change results in a smooth speed change and since the hydrodynamic toque converter is not disconnected for any undue length of time, the operational effectiveness of the arrangement is not reduced to any marked degree.

In a power transmission device fitted with manually operated gear change it may also be desirable, for the optimum exploitation of the efficiency of the engine, to carry out automatic changes between two adjacent gears according to the prevailing driving conditions. This can be achieved by connecting a change-over valve to an outlet of the multi-position control valve, the change-over valve serving to automatically change gear between at least the two highest gears.

The invention will now be described by way of example with reference to the accompanving drawings in which: Figure i is an axial section through a drive connection used in a first embodiment of the invention, showing a hydrodynamic torque converter connected to a planetary gearing: Figure 2 is a diagram showing graphs of the engine speed, in revolutions per minute and the traction power P of a vehicle plotted against the vehicle speed in Km per hour for a drive connection as shown in Figure 1 and in respect of the four forward speeds obtained therewith; Figure 3 shows the drive connection of Figure 1 in simplified form with a hydraulic connecting device for selectively connecting individual servo motors for the clutching and braking. both for hydraulic drive and direct drive of the hydrodynamic torque converter, and in this simplification of the invention, the basic pressure for the control fluid is supplied by the lubricating oil system of the interna combustion engine; Figure 4 is similar to the control system of Figure 3 and modified with supply of the control fluid from the filling pump of the hydrodynamic torque converter; Figure 5 shows a further modification of the control system including a valve for the automatic change-over between the third and fourth gears according to different driving conditions, and Figure 6 is an axial section taken through a drive connection suitable for heavy vehicles and including a hydrodynamic torque converter.

The drive connection shown in Figure 1 has a converter part W and a mechanical drive part M, which are enclosed by stationary housings H, and H2 respectively, and which are rigidly connected to each other.

The housing Hl contains a hydrodynamic torque converter with a rotatable converter casing 10 which can be drive via a coupling 12 bv the flywheel 14 of a vehicle engine (not shown), particularly a disel engine, and is mounted at the rear end in a bearing 16 on a hollow shaft 18. The converter casing 10 defines a part of the circumferentially outer boundary of a toroidal-shaped working chamber 20 and supports a ring of pump blades 22.

The outer boundary of the toroidal-shaped working chamber 20 is completed by a turbine wheel 24 which carries a ring of turbine blades 28 and a guide or reaction wheel 26 which carries a ring of guide blades 30.

The guide wheel 26 is mounted on the hollow shaft 18 via a freewheel 32 in such a way that the guide wheel 26 can only rotate in the same direction as the converter casing 10. The turbine wheel 24 is rotatably mounted on a hollow shaft 34 and is sealed relative to the hollow shaft 18 by a sealing ring 36. The hollow shaft 34 is carried on a main drive shaft 40 by a spline connection 38 with the splines on the shaft 34 machined to include a plurality of axially extending grooves 38A. The drive shaft 40 is mounted at the front end in the converter casing at IOA and at the rear end in the housings H, H2 by means of a ball bearing 42.

The hollow shaft 34 is mounted for axial movement on the central main drive shaft 40 and in the turbine wheel 24 and it is sealed during axial movement by a seal 44 located at the rear end of the hollow shaft 18. At its front end, the hollow shaft 34 carries a friction disc 46 having a conical outer flange 46A, the outer peripheral face serving to co-operate with the inner peripheral face of a conical ring 40 of the casing 10, whilst its inner peripheral face serves to co-operate with a conical ring 50 on the turbine wheel 24. The conical ring 50 and the friction disc 46/46A together forms a release clutch between the turbin wheel 24 and main drive shaft 40, by which the ring of turbine blades 28 can be released from the main drive shaft 40. Further, with the conical ring 48 the friction disc 46/46A forms a direct clutch between the converter casing 10 and the main drive shaft 40, by which the drive motor can be connected directly to the main drive shaft 40 while the turbine 24 of the hydrodynamic torque converter is free. Leaf springs 52,54 carried by the converter casing 10 and on the turbine wheel 24 respectively normally serve to centralise the flange 46A between the conical ring 48 and the conical ring 50 and thus keep both clutches disengaged. The engagement of the clutches 46A/48 and 46A/50 is achieved by control of the flow of fluid into and out of the toroidal-shaped working chamber 20. For this purpose, between the hollow shaft 34 and the central main drive shaft 40 there is a pressure fluid channel 56 through which converter fluid under pressure can be supplied by a filling system (not shown in Figure 1), via a transverse supply bore 58 and through to the front side (i. e. left hand side as viewed of Figure 1) of friction disc 46. In a similar way, between the hollow shaft 34 and the hollow shaft 18 a channel 60 connects a supply bore 62 from which fluid under pressure can be introduced via the axial passages 64 in bearing 66, between the hollow shaft 34 and the turbine wheel 24 to the rear side (i. e. right hand side as viewed in Figure 1) of the friction disc 46. In the turbine wheel 24 there is a first maximum pressure valve 68 which opens into the space formed between the turbine wheel 24 and the friction disc 46 when a predetermined pressure is reached in the converter chamber 20.

A second maximum pressure valve 70 is located in the friction disc 46 and opens when a predetermined pressure is reached in the space between the turbine wheel 24 and the friction disc 46, this space being connected to the channel 56. The hollow shaft 18 has on its outer side a channel 72 through which the toroidal-shaped working chamber 20 can be connected to supply 74.

It will be understood that when fluid under pressure is introduced through the supply bore 58 and the channel 56 to the front side of the friction disc 46, a pressure force will be exerted which presses this against the turbine wheel 24 and thus engages the release clutch 46A/48, since, when this clutch engages, the pressure fluid is prevented from directly flowing away between the converter casing 10 and the friction flange 46A into the space between the flange 46A and the turbine wheel 24. The pressure fluid which is supplied thus flows round the periphery of the turbine wheel'4 into the working chamber 20, increases the pressure in the working chamber and opens the maximum pressure valve 68, so that the converter fluid returns through the channel 60 and the supply bore 62 to a valve of the pressure fluid system which wll be later described. If, on the other hand, the pressure fluid is passed through the supply bore 62 and the channel 60 into the space between the flange 46A and the turbine wheel 24, it exerts a pressure on the rear side of the friction disc 46 and, by moving the friction disc 46 forces the direct drive clutch 46Ai48 into engagement. Since also in this instance the pressure fluid is prevented from flowing away over the flange 46A of the friction disc 46 into the space between the flange 46A and the converter casing 10 and onwards to the channel 60 and the supply bore 62. the maximum pressure valve 70 opens to the supply bore 56,58 when the necessary pressure has been reached. In this way the valve 70 allows the pressure fluid to flow away whilst at the same time maintaining an adequate pressure difference between the two sides of the friction disc.

The housing H, supports in a ball bearing 76 at its rear end, an output shaft 78, the front end of which is mounted in an axial bore formed in the central main drive shaft 40. A two part planet carrier 80 is rotationaiiy fixed to the output shaft 78 at 82, and bears a cluster of planet gears 84, of which only one is visible in the plane of the section shown in Figure 1.

Each planet gear 84 has three gear rings 86,88,90 with different diameters and is mounted with roller bearings 92,94 outside the largest gear ring 86 and between the gear rings 88,90 on the two-part planet carrier 80.

A flange 96 on the end of the main drive shaft 40 is connected by a spline connection to a drive flange 98 which is rotationally fixedly connected to an internally toothed cylindrical ring gear wheel 100 which meshes with the largest gear 86 of the planet gears. A cylindrical sun gear wheel 102 rotatably and axially movably mounted on the drive shaft 78 has gear rings 102A, 102B at its ends as shown, The gear rings 102A and 102B mesh at the front end with the largest planet 86 of the planet cluster 84 and at the rear end with a friction disc 104.

The friction disc 104 can be craked by a servo motor piston 106, which is movable in the housng H2, against the force of a plate spring 108 carried by a fixed housing insert 109. A further cylindrical sun wheel 112 is mounted to mesh at its front end with the middle gear 88 of the planet cluster and, at its rear end, is rotationally fixed to a friction disc 14, which can be firmly braked by a servo motor piston 116 against the effect of a plate spring 228 carried by the above-mentioned housing insert 110. A third sun wheel 120 meshes over the greater part of its length with the smallest planet 90 of the planet cluster 84 and is connected at its rear end to a friction disc 122 which can be braked by a servo motor piston 124 against the effect of a plate spring 126 supported on an extension 127 of the housing H2. Finally, a further internally toothed ring gear wheel 130 meshes radially outwardly with the smallest planet 90 of the planet cluster 84. The ring gear 130 is formed integrally with a friction disc 132 which can be firmly braked by a servo motor piston 134 against the effect of a plate spring 136 on a housing insert 128.

The planet carrier 80 has an annular member 138 on which a pressure ring 140 of a disc clutch 142 can be moved between the ring gear 100 and the planet carrier 80. For engagement of the clutch 142 a servo motor piston 144 which is located in the rear area of the housng part H, and which acts via a first needle bearing 146 on the axially moveable sun wheel 102 and a second needle bearing 148 together with push rod 150 acting on a movable pressure ring 152 on the hub of the planet carrier 80. When moved to the left the pressure ring 152 engages a number of radially extending levers 154 which are pivotally mounted on the annular member 138 of the planet carrier 80. The outer ends of the levers 154 engage the pressure ring 1.-0 for the operation of the disc clutch 142. A plate spring 156 located in the region of the cuter ends of the levers 154 is stressed to bias the levers 154 into the disengaged directi n of the clutch 142 so that, when the servo motor piston 144, is not actuated, the clutch 142 is disengaged and, at the same time, the needle bearings 146, 148 remain axially loaded.

It is apparent that upon engagement of the disc clutch 142 by setting the servo motor piston 144 under pressure, a direct connection of the man drive shaft 40 and the output shaft 78 is achieved and this corresponds to the fourth gear of the mechanical drive part M formed by the epicyclic gears. Apart from this direct drive, a first gear, with the greatest step-down in the gearing, can be connected by braking the sun wheel 120 with the aid of the servo motor 124, a second gear can be connected bv braking the sun wheel 112 with the aid of the servo motor 116 and a third gear can be connected by braking the sun wheel 102 with the aid of the servo motor 106. If, on the other hand, the annular wheel 130 is braked with the aid of the servo motor 143, the planet carrier rotates backwards in relation to the driving ring wheel 100 and reverse drive is obtained on the drive shaft 78.

Figure 2 is a diagram showing the development of the engine revolutions and the traction of the vehicle at the wheels, against the vehicle speed for the four gears designated as I-IV.

The reduction ratio of the epicyclic gears given by example, are Gear I (braking 122/124 engaged) 3.'-'1 Gear II (braking 114/116 engaged) 1. 95 : 1 Gear III (braking 104/106 engaged) 1. 32 : 1 Gear IV (clutch 142 engaged) 1-1 (Direct drive) When reverse gear is engaged with the aid of the brake 132,134 the reduction in this example is 1.8'1.

Figure 3 shows a first embodiment for the control of the different clutches and brakes with the aid of a manually operated gear lever 160 which can be moved frm a neutral position into four forward gear positions for the gears 1, 2, 3, 4 (in Figure 3) (and corresponding to 1, II, III, IV in Figure 2) and into the reverse gear position R. The gear lever 160 acting via a link 162 operates a multi-position control valve 164, which is fed with pressure fluid through a pipe 166.

Pressure fluid in the embodiment of Figure 3 is obtained from a sump 172 by means of the lubricating oil pump 168 of a motor vehicle engine which is schematically shown and designated 170. The oil sucked out of a sump 172 and is fed via a filter 174 and a non-return valve 176. Between the filter 174 and the non-return valve 176 a pipe feeds fluid to a high pressure pump 180 of relatively small capacity and driven by the rotating converter casing 10 of the hydrodynamic torque converter. The outlet of the high pressure pump 180 is connected to the inlet pipe 166 which is connected to the multi-position control valve 164.

Between the inlet pipe 178 and the outlet pipe 166 of the high pressure pump a non-return valve 182 opening to the outlet pipe 166 and a maximum pressure valve 184 opening to the inlet pipe 178 are connected in parallel to one another.

Leading away from the multi-position control valve 164 and connecting the pipe 166 (according to the corresponding position of the gear lever 160) to the servo motors 124,116, 106,144 and 134, there are five pipes 1,2,3,4, and R, from each of which a branch pipe lads via an individually ajustable maximum pressure valve MV to a pressure-free return pipe 186, which returns to the sump 172. In Figure 3 the five individually ajustable maximum pressure valves are bracketed together and generally designated MV for ease of reading the Figure.

In addition to leading to the high pressure pump 180, the pipe 178, which is under the pressure of the engine lubricating oil pump 168 also leads via fa filter 188 to a centrally located pressure inlet of a multi-position valve 190 and to two electromagnetically actuated primary valves 192,194. When the two primary valves 192,194 are operated, the pressure fluid which has accumulated at the appropriate primary valve inlet, is introduced into the upper or lower control chamber of the multi-position valve, thereby displacing the multi-position valve from a central position where it shuts off the further flow of the pressure fluid into a lower or upper position respectively. In the lower position of the multi-position valve 190, the pressure fluid arriving through the pipe 178 is delivered via a pipe 196 to the connection 58 on the hydrodynamic torque converter whereas the connection 62 is connected via a pipe 198 with the pressure-free pipe 186, so that the release clutch 46A/50 for the turbine blade ring 28 is engaged. In the upper position of the multi-position valve 190, the connection 62 is connected via the pipe 198 with the pressure pipe 178 and on the other hand the connection 58 is connected via the pipe 196 with the pressure-free pipe 186, so that the direct clutch 46A/48 is connected. Correspondingly, in the neutral position both the release clutch 46A/50 and also the direct clutch 46A/48 are disengaged under the effect of the leaf springs 52,54 which centralise the friction disc 46.

As long as neither of the primary valves 192,194 is actuated, the flow of power through the hydrodynamic torque converter is interruped. If the upper primary control valve 192 is actuated, then hydraulic drive is engaged by the clutching of the turbine blade ring to the main drive shaft 40, and by actuating the lower primary control valve 194 direct drive connection occurs when the hydrodynamic torque converter rotates. In Figure 3 designations D and H accompanied by arrows indicates the direction of flow of the pressure fluid for direct and hydraulic drive respectively.

The actuation of the primary control valves 192 and 194 is accomplished electrically under control of the gear lever 160 via a 3-position switch 200, on which the gear lever 160 acts when moved about its pivotal axis. To assist understanding, the part designated X shows the view of the gear lever and it will be seen how the gear lever 160 is guided in a slideway 202 with a generally vertical neutral position track N, from which finger-shaped, relatively long tracks 1, 2,3,4 corresponding to the four forward gears and a shorter track R for the reverse gear extend. In the neutral position, i. e. as long as the gear lever is located in the neutral position track N, the actuating coils of the primary control valves 192, 194 are not energised. When the gear lever 160 is put into one of the tracks 1 to 4, firstly the actuating coil of the upper primary valve 192 for hydraulic drive is energised, and then the actuating coil of the lower control valve 194 for direct drive is energised. When the lever 160 is put into the shorter reverse track R the actuating coil of the upper bore valve 192 for hydraulic drive is connected to voltage and excited.

In addition, the electro-hydraulic control system shown also contains a valve 204 including two valve pistons separated from each other and kept at a predetermined distance by a rod rigidly fixed to one piston. The two valve pistons are pressed towards the left by a compression spring acting on the right-hand piston. In this position a pipe 206 connected into the supply 74 (Figure 1) is, in turn, connected to the pressure fluid sump, thus forming a direct outlet pipe to the sump for the working fluid contained in the toroidal-shaped working chamber 20 of the converter. A pipe 208 which leads from the primary control valve 192 to the upper actuation side of the multi-position valve 190 also leads to the valve 204 at a position between the two pistons thereof. In a similar way the pipe 210 which leads from the primary control valve 194 to the lower actuation side of the multi-position valve 190 also leads to the left side of the left actuating piston of the valve 204. This construction enables the valve 204 to be operated so as to interrupt the connection between the toroidal-shaped working chamber of the converter 20 and the sump, via the channel 72 and between the connection 74 and the pipe 206, as well as interrupting the opening of one of the two valves 192,194 with the aid of the manual lever 160 and the 3-position switch to connect the hydraulic drive or the direct drive.

The electro-hydraulic control shown in Figure 3 operates as follows. Assuming that the vehicle equipped with the drive and the controls shown in Figure 3 is stationary, with its engine idling and with the gear lever 160 located in the neutral position slot N of the slideway 202. Under these conditions the lubricating oil pump 168 is running and supplies not only the drive motor 170 with oil, but also pressurises the pipe 178 without the pressure fluid therein, namely oil, being able to flow away.

If the driver now wishes to move his vehicle in a forwards direction he moves the gear lever 160 out of the slot N downwards to track 1 for gear 1. He thereby connects, via the multi-position control valve 164,. the pipe 166 with servo motor 124 (Figure 1) for the first gear, which means that the planetary gear 120 is firmly braked. This occurs when all parts of the planetary gearing are stationary since the friction disc 46 in the torque converter W is in its central position under the influence of the two leaf springs 52, 54, and, therefore, both the release clutch 46A/50 (given the reference F in Figure 3) and the direct clutch 46A/50 (given the reference D in Figure 3) are disengaged.

When the driver moves the gear lever 160 to a position half-way along the track 1 of the slideway 202, the 3-position switch 200 closes the contact switch leading to the actuating coil of the primary control valve 192 to open this valve. Since at the inlet side of the valve 192 pressure builds up from the pipe 178, this is released to the upper actuation side of the multi-position valve 190 and changes this into the lower position for hydraulic drive. In this condition oil under pressure from the pipe 178 flows through the multi-position valve 190 to pipe 196 and from this through connection 58 and channel 56 to the front side of the friction disc 46, thus engaging the release clutch F. Engagement of the release clutch F and consequent engagement of parts 46A/50 results in the closure of a direct connection around the periphery of the friction disc ! flange 46A to the space between the friction disc 46 and the turbine wheel 24. In addition, the valve 204 also seals off the pressure release of the toroidal-shaped working chamber 20 to the sump, and pressure builds up in the working chamber 20 until the maximum pressure valve 68 opens. Thus, the maximum pressure valve 68, whilst maintaining a predetermined pressure in the working chamber 20 of the converter, allows the pressure fluid to flow back to the multi-position valve 190 through the channel 60, the connection 62 and the pipe 198 and on from there through the pressure-free pipe 186 to the sump 172. With the engagement of the release clutch F the turbine blade ring 28, which is carried along by the pump blade ring 22, initially Aithout circulation of fluid in the working chamber 20, is also braked. This the circulation of the tluid in the working chamber 20, and produces a gent) y increasing turning moment at the turbine wheel 24, which is transmitted via the main drive shaft 40, the drive flange 98 and the ring wheel 100 to the large gear ring 86 of the epicyclic gears 84. Since the small planet gear 90 of the epicyclic gears 84 is held stationary on the inner circumference of the planetary gearing by the firmly braked sun wheel 120, the turning moment which arises at the ring wheel 100 is further increased at the output shaft 78 which carries the planet gears 84 and leads to a rapid and yet smooth moving off of the vehicle in the selected gear 1.

It is apparent that here the release clutch F acts as a separating clutch, whilst the brake 122 has the sole function of holding the sun wheel 120. Despite this, the load on the release clutch F is minimal because at the moment when coupling takes place, the fluid in the toroidal-shaped working chamber 20 is still stationary and, therefore, there is still no turning moment produced at the turbine wheel 24.

As soon as the vehicle has started to move, the gear can be changed to the next highest gear without first having to move the gear lever 160 right up to the end of the track 1 and thereby, by releasing the release clutch F and engaging the direct clutch D, connecting the direct drive. In practice, connection of the direct drive in gear 1 is much more likely to be quite unnecessary, so that the track 1 can be made as short a track as reverse track R to make exclusively hydraulic drive possible.

The change to the next highest gear is carried out by moving the gear lever 160 back to the neutral position track N, thus opening the release clutch F once more, and subsequently switching the multi-position control valve still further by moving the gear lever 160 into the next position in which the sun wheel 12 is rigidly braked instead of the sun wheel 120. As with the opening of the release clutch F, the circulation of the fluid in the toroidal-shaped working chamber again ceases, engagement of the brake 114 for the sun wheel 112 is also achieved without a load, and with the entry of the gear lever 160 into the track 2 of the slideway 202 the release clutch F is also engageci almost unloacled. In this condition, the circulating flow once again builds up in the converter chamber 20 and causes the creation of a turning moment at the turbine wheel 24 which is carried via the release clutch F, the main drive shaft 40, the drive flange 98 and the annular wheel 110 to the planetary gearing and there transmitted, further increased, to the output shaft 78.

If, for example, due to the vehicle encountering an incline or prhaps a speed limitation in local traffic, the drive speed of the vehicle produced at the change-over point of the converter is not to be further increased, the gear lever 160 can be put into the terminal position in the track 2 of the slideway 202 as required, so that the release clutch F is disengaged by closing of the valve 192, and the direct clutch is engaged in its place by actuation and opening the valve 194. When the valve 194 is opened, the multi-position valve 1-90 is moved into the upper terminal position and the pressure fluid, which builds up at the inlet of the multi-position valve, flows through the pipe 198, the connection 62 and the channel 60 to the rear side of the friction disc 46 so that the flange 46A engages the conical insert 48 of the converter casing 10. The seal thus created between the friction disc flange 46A and the conical insert 48 prevents flow round the periphery of the friction disc flange 46A and leads to the building-up on the rear side of the friction disc 46 of the appropriate holding force for the direct clutch until the maximum pressure valve 70 opens and, whilst maintaining a sufficiently great pressure difference, allows the pressure fluid to flow away through the channel 56 and the connection 58 as well as the pipe 196 to the multi-position valve 190 and further on through the pressure-free pipe 186 to the sump 172.

If, on the other hand, the vehicle is to be accelerated further after reaching the change-over or shift point in the torque converter, the gear lever 160 is brought back into the neutral position track N of the slideway 202 and the multi position control valve 164 is brought into the next position in which the sun wheel 102 is made stationary instead of the sun wheel 112. In this condition the gear lever 160 is moved irto the position for hydraulic drive (i. e. into the track 3) the release clutch F is first opened and then re-engaged so that the smallest reduction in the planetary gearing is introduced into the drive line and the converter again increases the turning moment.

A change into gear IV is carried out in a similar way but, instead of the braking of the sun wheel 102, the direct clutch 142 in the planetary gearing is actuated. When the direct clutch is so actuated and engaged, and the gear lever 160 is in the position H, the torque converter again first undertakes multiplication of the turning moment until the shift point is reached, whereupon the gear lever 160 is moved into the terminal position D in the track 4 and instead of the release clutch, the direct drive clutch is engaged into the drive line. In this position the entire drive connection from the converter casing 10, which forms the input component, up to the drive shaft 78 is then directly and mechanically connected right through the transmission and, in this condition the hydrodynarnic torque converter and the planetary gearing is disconnected.

The above description of the connection of all stages of the planetary gearing in conjunction with hydraulic drive in the converter part W is normally only required when particularly high traction power is required either for the acceleration of the vehicle with a heavy load due to the actual load carried, or on inclines. With more lightly loaded goods vehicles and pesonnel transporters, on the other hand, one to two of the intermediate gears, such as for example gears I and III, can be omitted. Conseyuently, direct drive is only engaged in the lower gears if the vehicle is to be driven for a long time in these gears at correspondingly lower speeds in the most favourable working range of the drive motor.

It will be self-evident from the foregoing that the establishment of the reverse gear is carried out in a similar manner. Since the reverse gear is engaged only for short periods and over short distances, it is never necessary to engage direct drive (with consequent by-passing of the hydrodynamic torque converter) so that the track R in the slideway 202 can always be a relatively short track.

The control system shown in Figure 4 is different from that in Figure 3 only in that, instead of the lubricating oil pump of the drive motor, the hydrodynamic torque converter filling pump (reference 210 in Figure 4). produces the basic pressure for the control system.

The filling pump'10 sucks fluid out of the sump 212 of the converter. The pressure pipe of the filling pump 210 corresponds to the pipe 178 in Figure 3 and is therefore also given the reference 178. In other respects the control system shown in Figure 4 is identical to that in Figure 3, and therefore functions in the same way.

It will be understood that, as Ion (, as the multi-position control valve 164 remains in the neutral position in which all outlets 1 to 4 and R are closed, that a pressure build-up will occur in the inlet pipe 166 to this valve due to operation of the filling pump 210 or operation of the lubricating oil pump 168 (Figure 3) and the subsequently connected high pressure pump 180 and that the build-up is limited to a specific maximum value by the maximum pressure valve 184. As soon as the multi-position control valve 164 is switched to connect the pipe 166 with one of the control servos for the brakes or couplings of the ptanetarv gearing, the pressure will momentarily drop and with the rotation of the high pressure pump 180, the filling pump 210 or the lubricating oil pump 168, pressure fluid is supplice under comparatively low pressure through the non-return valve 182 directly to the servo motor to which it is connected. The so connected servo motor is, therefore, rapidly fillcd by virtue of the high capacity of the filling or lubricating oil pump, and the relevant brake or clutch is rapidly engaged with comparatively low power. As soon as this filling is completed and when, therefore, no further fluid replenishment is required by the filling or lubricating oil pump through the non-return valve 182 and the pipe 166, the pressure in the pipe) 66 rises again and closes the non-return valve 182. This allows the high pressure pump 180 to still further increase the pressure in the pipe 166 and provides the ultimate pressure required to hold the brakes and clutches in engagement. The level of this pressure is finally individually limited by the maximum pressure valves MV associated with each individual servo motor and the individual valves are adjusted so that the associated servo motor operates at the pressure required to hold the relevant brake or clutch.

The control system as shown in Figure 5 differs from that shown in Figure 4 only in that the multi-position control valve 164 has four outlets, and the slideway has only three forward tracks 1,2 and 3. In the position 3 the pipe 166 is connected by a pipe 213 to a change-over valve 214 which is constructed as a 4/2 way valve. The two outlets of the change-over valve 214 lead to the servo motors 106 and 124 for gears III and IV (Figure 1), and the other inlet of the valve 214 is connected to the torque converter sump.

With the switching over of the valve, which can be carried out electrically, hydraulically or pneumatically depending on certain parameters such as, for example, the number of revolutions of the engine or perhaps the position of the accelerator pedal, in position 3 of the multi-position control valve 164 either the servo motor 106 is connected to the pressure pipe 166 and the servo motor 124 released (gear III) or, the servo motor 124 is connected to the pressure pipe 166 and the servo motor 106 released (gear IV). With the aid of this change-over valve the driver can, for example, sometimes be spared from changing the position of the gear lever 160 when a reduction in the speed makes travel in gear III instead of gear IV desirable with regard to the efficiency of the engine. Further, at higher travelling speeds, a higher turning moment can momentarily be made available by pressing the accelerator pedal right down to produce so-called"kick down".

The description of the control system according to Figures 3 to 5 is carried out in conjunction with the drive connection shown in Figure 1, the torque converter part W of which contains a hydrodynamic torque converter with a release clutch for the turbine and with a direct clutch, which can be engaged alternately. It will be apparent, however, that the control systems shown and described can be used in conjunction with other drive connections which, for example, have only one mechanical drive in conjunction with a separating clutch of conventional design, which is open when the gear lever 160 is in the neutral position track. In this case, naturally the whole valve complex which is encircled in dashes in Figures 3,4 and 5, is replaced by a simple actuation valve for the separating clutch, which can if desired, also be changed electro-magnetically by a switch operated by the gear lever, similar to the 3-position switch 200.

The control systems shown and described can, however, also easily be adapted to drive connections with other hydrodynamic torque converters, which do not contain any releasable vaned component. Such a drive connection is shown in Figure 6 and, although the mechanical gear M is the same as that of the drive connection of Figure 1, the torque converter of the drive shown in Figure 6 includes a hydrodynamic torque converter with a direct clutch D between the converter housing 216, which forms the input component and a main dri e shaft 218 which corresponds to the main drive shaft in Figure 1. A turbine wheel 220 havi ig two turbine blade rings T1 and T2 is non-rotatably carried by the drive shaft 218 and a rii g of guide blades L is disposed between the turbine rings T1 and T2. The guide ring L is sup sorted on an impeller wheel 222 which, in turn, is supported on an impeller wheel shaft 2* which is rotatably mounted in the converter housing and encloses the main drive shaft 21s as a hollow shaft. A multiple disc brake 226 enables the impeller wheel 222/shaft 224 to be locked with the non-rotatabie housing.

Due to the lack of a release clutch as well as a separating clutch, in the drive connection shown in Figure 6 the braking or clutching of the individual parts of the planetary gearing is carried out under load. Thus the brake and the direct clutch operate not only as holding couplings, but also as separating couplings or clutches and, consequently, are generally constructed with more braking and clutch discs. A certain amount of relief for these brakes and this clutch during a change can, however, be achevez by releasing the impeller wheel brake'16 during the change-over process so that the hydrodynamic torque converter does not bear the full turning moment.

The adaptation of the control systems shown in Figures 3 to 5 can then be achieved in that the sear) lever 160 is equipped with a 3-position switch which in the position H opens a valve to connect the servo motor of the impeller wheel brake 226 and in the position D opens a further valve through which the front face of an actuating piston 2'8 for the direct clutch D formed as a disc clutch is pressurised.

Reference is also made to the description and claims of copending Application 3572/76 (Serial No. 1559491) and 28964/76 (Serial No. 1559492).

Claims (13)

WHAT WE CLAIM IS:

1. A variable power transmission device for motor vehicles fitted with a multi-stage mechanical gear change, wherein the gear change includes planetary gearing with at least two planet wheels of different diameters meshing with associated sun andlor ring gears which can be frictionally braked or clutched into a drive line, each brake or clutch having an associated fluid pressure operated servo-motor for actuating the same and each servo motor having a non-rotatable housing, a multi-position control valve connected via a non-return valve with a low pressure fluid system, the multi-position valve serving to selectively control operation of the individual brakes and clutches, a maximum pressure valve in each connection between the multi-position control vale and the respective servo-motors and a low capacity high pressure pump for engaging the brakes, and clutches, the low capacity high pressure pump being connected to supply fluid to the multi-position control valve at a position between the multi-position control valve and the non-return valve.

2. A transmission device as in Claim 1, wherein each servo motor is connected via its associated maximum pressure valve to a press-free return pipe and wherein the pressure at which each maximum pressure valve operates can be individually adjusted for a predetermined maximum pressure.

3. A transmission according to Claim 1 or Claim 2 wherein the low pressure fluid system is included in the transmission and is supplied by a high capacity low pressure pump.

4. A transmission claimed in any preceding claim including a separating clutch between the drive motor and the planetary gearing.

5. A transmission according to any one of Claims 1 to 4 including a hydrodynamic torque converter between the drive motor and the planetary gearing.

6. A transmission according to Claim 3 or Claim 4 wherein the torque converter includes a releasable torque transmitting component (pump or turbine) and wherein the said component constitutes a part of a transmission clutch.

7. A transmission according to any one of Claims 1 to 5 wherein the torque converter can be by-passed by a direct drive clutch.

8. A transmission according to any preceding claim wherein the multi-position control valve is manually operable by means of a lever which is movable in a slideway.

9. A transmission according to Claim 8 wherein the slideway has a neutral position track with finger-shaped tracks leading from it for [he individual gear steps of the gear change.

10. A transmission according to Claim 9 and any one of Claims 4 to 6, including a change-over device for the operation of the separaiing clutch wherein the device opens when multi-position control valve is in a position which corresponds to the neutral position.

11. A transmission according to Claims 9 and 10 wherein the slideway includes an area designated for hydraulic drive and disposed in a position adjacent the neutral position track.

12. A transmission according to Claim 11 wherein the slideway includes an area designated for direct drive.

13. A transmission device according to any one of Claims 7 to 11 including a change-over valve at one outlet of the multi-position control valve for automatic change-over between at least the two highest gears.