WO2008037946A1

WO2008037946A1 - Plural power path speed changing gearbox

Info

Publication number: WO2008037946A1
Application number: PCT/GB2007/003154
Authority: WO
Inventors: John Madge
Original assignee: Individual
Current assignee: Individual
Priority date: 2006-09-29
Filing date: 2007-08-20
Publication date: 2008-04-03
Anticipated expiration: 2009-03-29
Also published as: GB0713017D0; GB0903880D0; GB2442294A; GB2456076B; GB2456076A; GB0619234D0

Abstract

A plural power path speed changing gearbox comprises a plurality of co-planar gear sets each comprising an input gear (180-188) meshing two diametrically opposed output gears (160A/B-168A/B); the input gears of each set being mounted on an input shaft (12) and the output gears of each set being mounted on different ones of a pair of output shafts (14A/B); individual co-planar gear sets being selectively drivingly engageable between the input shaft and the output shafts; a closed torque path may be formed between the two output shafts by means of matching output gears (20A/B) and an offset combining gear (22). The driving engagement between the selected co-planar and the input and output shafts to provide a selected speed ratio may be by means of each input gear being clutched to the input shaft through a central aperture (52) engageable with dog teeth (56A, 56B, 56C) carried by the shaft (54). Other clutching arrangements and gearbox configurations are disclosed.

Description

PLURAL POWER PATH SPEED CHANGING GEARBOX. Background of the Invention.

This invention broadly concerns a speed changing gearbox which employs plural power paths. This type of gearbox is in common use in the automotive industry, most usually in the form of planetary gears operated by oil immersed clutches for use in automatic transmissions. Such automatic gearboxes are relatively expensive and inefficient. In a separate line of development, the advent of heavy articulated vehicles and their need for a short and high power gearbox led to the introduction of very successful twin layshaft spur gear arrangements in the 1970's. More recently the pressures from noise legislation have resulted in the introduction of helical gear versions. The advantages of this arrangement has been well documented and proven commercially: i.e, lighter weight, smaller size and lower cost for a given torque carrying capacity. The advantages increase as the torque level increases, particularly when a multiplicity of ratios is required. Thus use in trucks with their need for twelve or more speed ratios and up to 400KW engine power was a natural application, using a twin layshaft speed change section coupled in series with a range change gearbox.

Introduction of twin layshaft plural power path gearboxes into smaller vehicles has been very limited. The lack of automaticity has precluded their use in heavier luxury motor cars. When applied to mass-produced, lower power motorcars with four gear ratios and wide synchronisers then the benefit of reduced length becomes marginal. The demand for ever more efficient motor vehicles is leading to an increased number of gear ratios in mass produced cars. However this factor alone does not make the plural power path arrangement competitive.

Concurrently motor vehicle engines are increasingly being controlled by a computer management system. The larger number of gear change steps and the engine may be integrated within a power train computer management system. The combined technology is gradually encroaching upon the conventional planetary automatic gearbox. However, the new sophisticated control systems are basically being applied to old technology, single layshaft gearboxes.

The conventional plural power path twin layshaft gearbox embodies a closed loop torque path consisting of (1 ) an input dividing gear set comprising a central gear in engagement with two identical, co-planar, diametrically opposed gears each mounted on respective layshafts and (2) a number of combining output gear sets each comprising two matching diametrically opposite gears respectively mounted on the different layshafts and each engaging a central gear positioned on an output shaft co-axial with the input dividing gear. The gearbox speed change ratio is usually effected by selective unsynchronised dog clutch engagement of one of the output gear sets (2) to the other output shaft as described for example in US3105395. Summary of the Invention.

The present invention provides a twin output shaft speed changing gearbox which is suitable for a broad range of automotive applications, including not only heavy and light commercial but also mass passenger vehicle and autosport usage, and which can provide a larger number of ratios in a given space than existing gearboxes of equivalent power. Such a development is complementary to computer controlled engine and drive train management systems. The invention eliminates the input dividing gear set of the conventional twin layshaft gearbox. Accordingly, the present invention provides a plural power path speed changing gearbox comprising a plurality of co-planar gear sets each comprising an input gear meshing two diametrically opposed output gears; the input gears of each set being mounted on an input shaft and the output gears of each set being mounted on different ones of a pair of output shafts; individual co-planar gear sets being selectively drivingly engageable between the input shaft and the output shafts; the torque path being closed using an output gear on each output shaft, each engaging with a combining gear. For example, the torque path between the output shafts may be closed using a matching gear on each output shaft, each engaging simultaneously with a single output gear. Where this single output gear is generally co-axial with the input gear and shaft, it may be moveable to provide even torque sharing between the two power paths, to compensate for inaccuracies in parts manufacture and assembly. Thus, in the case of spur gears, the single output gear may be moveable transverse to the plane of the output shafts. For helical gears, the movement may be axial translation and/or rotation about an axis normal to the plane of the output shafts, similar to the arrangement described in US3885446 or WO03/046409. Other known torque sharing arrangements are of course possible, such as disclosed in GB 1434928, or as in EP0244263 for double helical or "herringbone" gears, or simply by extremely accurate manufacture and assembly. This gearing layout differs significantly from the conventional twin layshaft speed changing gearbox in that torque sharing gears may be provided at the output shaft rather than the input shaft. The layout also facilitates the use of clutching arrangements operative at the input shaft rather than the output shaft, these being advantageous for speed synchronising, as further described below. Alternatively, the axis of the single output gear is not co-axial with the input shaft and is generally offset from the gearbox and integrated with the drive downstream of the change speed gearbox. This arrangement locates the single combining gear set at a more advantageous position which is usually nearer to the point of the highest torque in the drive path. When the multiple drive paths are combined at a gear with relatively high reduction ratio significant savings in weight and cost are made. In the simplest configuration identical gears on the two layshafts engage an off-centre combining gear with a ratio of say 6 to 1 which provides a means to transfer the gearbox input torque to the front and back wheels of a motor vehicle via a parallel offset axis, or from a transverse mounted engine and gearbox to a parallel axis front wheel drive or, in a similar arrangement for a diesel driven rail application, to the rail wheel axis. In road vehicle applications, the single output gear may be the ring gear of a differential unit.

The foregoing arrangements having a non-co-axial single output gear, or with the torque loop closed at a ring gear or by other drive components that do not provide torque sharing compensation, may instead incorporate such compensation at the input gear of each speed changing gear set, by appropriate movement of the input gear relative to the input shaft. Yet alternatively, the necessary torque sharing compensation can be provided using a torsionally resilient element in the torque loop. The torsionally resilient element may be a compound combining gear that in a preferred form is a differential carrier having two ring gears, each driven by a different one of two output gears and connected to the other ring gear in a torsionally resilient manner. Still alternatively, the torque loop can be closed using oppositely handed helical gears respectively driven from each output shaft, each helical gear engaging different halves of a double helical ("herringbone") combining gear mounted to float axially to provide torque sharing compensation. As yet a further alternative, the helical gears can be of the same hand so as to engage a single helical combining gear mounted to tilt about its axis to provide the torque sharing compensation.

In a conventional speed changing gearbox the centre distance between the driving shaft and the layshaft(s) is usually not determined by the gear loading but by the gear ratio requirements in all of the ratio determining gear sets and the consequent bearing and clutch sizing for the "worst" gear set. These constraints are significantly lessened in the embodiments of the invention using a non-co-axial combining gear. This can provide a large reduction ratio, which in turn enables the co-planar gear sets to provide a speed increase between the input and output shafts. The higher speeds generally mean even further reduction in torque over and above that resulting from the use of twin paths, resulting in smaller and lighter bearings, clutches and rotating masses through to the combining gear.

The driving engagement between the selected co-planar gear set and the input and output shafts to provide a selected speed ratio can be by means of otherwise conventional (e.g. synchronising ) dog clutches operating between each output gear and its respective output shaft. The central input gears of each set may then be mounted to the input shaft in a manner providing torque sharing compensation. Where such compensation is provided elsewhere in the torque loop as described above, or is provided by accurate manufacture and assembly, conventional dog clutches can operate between each input gear and the central shaft, with permanent driving connections between the output gears and the respective shafts. This reduces the number of clutches required.

In a conventional single layshaft speed change gearbox the gear ratio steps are engaged by means of synchronising dog clutches mounted on the face of the driving gears. These clutches can double the length of the gearbox. Existing twin layshaft speed changing gearboxes have replicated this face clutching arrangement (albeit often without synchronisers) and the length disadvantages can sometimes outweigh the gains from narrow facewidth gears.

However, a centrally disposed gear in a multi layshaft arrangement does not need any central bearing because of the balanced gear tooth loads. Thus, in preferred embodiments of the invention, such a gear can be clutched through a central aperture. This can eliminate all face clutches, can significantly reduce the length of the gearbox and enables the power train designer to add more gear ratios without space penalty. Such an arrangement can be used with advantage in a plural power path speed changing gearbox of the kind described above. In this preferred arrangement the present invention provides a speed changing gearbox comprising a plurality of co-planar gear sets each comprising a central gear and a pair of diametrically opposed gears each in driving or driven engagement with the central gear, wherein speed ratios are selected by selected clutched engagement between a central aperture in each central gear and a shaft received in the aperture; one or more members being rotationally coupled to the input shaft, axially moveable with respect to the input . shaft and input gears and engageable with the input gear central apertures to provide the selective clutched engagement; the members axial movement being effected by various means, characterised by their elimination of conventional gear selector forks mounted between co-planar gear pairs.

This compact clutching arrangement can be used with advantage in conjunction with a computer controlled power train management system to provide automatic or semiautomatic gear changing. The system monitors the engine and vehicle load, speed and acceleration and programmes the change of gearbox ratio. With every increase in the number of available gear ratios, the system gets closer to operating like the desirable continuously variable transmission, albeit without the usual inefficiency inherent in these devices. The combination of the control system and the capacity for more speed ratios with any current gearbox space envelope as provided by this invention enables the vehicle designer to provide automaticity with improved overall fuel efficiency.

The arrangement has application in all existing multiple layshaft gearboxes with a centrally disposed gear, for example as widely adopted in the truck industry. With the capacity for more gear ratios, and in combination with the various forms of closed loop torque sharing as described above, applications are readily and economically extended into all types of front wheel drive and rear wheel drive and multi wheel drive vehicles including but not limited to motor transport, motor sport and mobile industrial equipment. Off road vehicle applications include rail, wheeled and tracked vehicles. The arrangement embodied in a plural power path change speed gearbox of the present invention can be applied to many types of industrial and marine drive. The selected clutched engagement may be achieved by a splined or other connection moveable axially of the shaft for torque transmitting engagement in the central aperture of the selected gear set. For example, the shaft may carry a torque transmitting annular member which is splined or otherwise provided with non-circular sections on its external and internal surfaces respectively to engage co- operating splines/surfaces in the central aperture of the selected gear set and on the central shaft. The annular member may be provided with a single set of external splines/surfaces, or a plurality of such sets, spaced for engagement with the selected one of a larger plurality of the central gears.

The external and/or internal splines may be shaped and dimensioned for co-operation with splines in the central aperture and/or central shaft splines to provide torque sharing compensating movement of the central gears, these forming dividing or combining gears of a plural power path gearbox.

The invention and its preferred features and advantages are described below with reference to illustrative embodiments shown in the accompanying diagrammatic drawings. Brief Description of the Drawings.

Figure 1 is a diagram of a speed changing gearbox embodying the invention;

Figure 2 is a diagram of an arrangement of gears for closing the torque loop of the gearbox of Figure 1;

Figure 3 is a diagrammatic view on line III-III in Figure 2; Figure 4 corresponds to Figure 3, modified to provide non-co-planar input and output shafts;

Figure 5 shows another gearbox layout embodying the invention, with power input and the power output combining gear approximately midway along its central axis;

Figure 6 is a schematic scrap section of dog clutch and synchroniser; Figure 7 is an alternative clutching schematic arrangement forming part of the speed changing gearbox in Figure 1, and

Figure 8 is a diagrammatic representation of a further such clutching arrangement;

Figure 9 represents a yet further clutching arrangement embodied in a diagram of a speed changing gearbox embodying the present invention; Figure 10 is an end view diagrammatic layout of a short length, multiple ratio gearbox embodying the invention;

Figure 11 is a side view corresponding to figure 10;

Figure 12 illustrates a flywheel synchroniser, and

Figures 13 and 14 illustrate a gearbox embodying the invention, used in conjunction with a continuously variable transmission (CVT).

Description of the Preferred Embodiment.

Referring to Figure 1, a speed changing gearbox 10 comprises an input shaft 12 on a central axis in driving engagement with two diametrically opposed output driving shafts 14A and 14B via a central gear 180 meshing two identical co-planar driving gears 160A and 160B mounted on each output shaft. Any number of similar sets of three co-planar gears 182, 162A, 162B; 184, 164A, 164B; 186, 166A, 166B; 188, 168A, 168B of varying diameter and identical centres can be mounted in parallel to the gears 180, 160A, 160B to provide the required number of gear ratios and speeds. The central gear in each gear set may be provided with sufficient degrees of movement to effect equal torque sharing between the output shafts. For example, the central gears may be mounted on or engageable with the shaft 12 by splined or other connections allowing the required degree of movement, whilst transmitting torque and power between the central shaft and the respective gear when under load. The change of gear ratio is effected by clutching the required gear set to the input shaft by means of dog clutches, further described below. The other(s) of the input or output shafts(s) are in permanent driving engagement with their respective gears.

As shown in Figure 2, the torque loop between the shafts 14 A, 14B may be closed by a single gear 22 mounted on a single output shaft 24. The gear 22 meshes simultaneously with gears 2OA, 2OB fixed respectively to the output shafts 14 A, 14B. As further shown in Figures 2 and 3, the shaft 24 may be co-planar with the shafts 14A and 14B (and hence also with input shaft 12, Figure 1) so that the gears 2OA, 2OB mesh with the gear 22 at diametrically opposed zones. (For simplicity, in the drawings, gear teeth are in general not shown; the gears merely being represented by their pitch circles or diameters.) Gear 22 and/or its shaft 24 may therefore be mounted for slight movement normal to the plane of the shafts 14A, 14B as indicated by arrows a in Figure 3, to provide torque sharing compensation. Such movement is appropriate for spur gears. Other forms of movement will be appropriate for other kinds of gear, as discussed in the patent specifications referred to above and known to those skilled in the art. With such an arrangement, there is no need to mount the central gears 180-186 to the input shaft to provide torque sharing. Instead, the gears 180-186 can be individually clutchable to the shaft 12 to select the gear ratios, the gears 160A/B-168A/B being fixed to their respective output shafts 14A and 14B.

Yet alternatively, as shown in Figure 4, the single output shaft axis 24 can be offset parallel to the plane of the output shafts 14A, 14B. The gears 2OA and 2OB mesh simultaneously with the combining gear 22 having an axis 24 which is parallel to, and offset from the plane of, the shafts 14A and 14B and the speed changing part of the gearbox. Axis 24 may be for example the front wheel axis of a rally car with gear 22 being the ring gear of a limited slip differential unit (not shown) on the axle. Alternatively, in four wheel drive applications, the combining gear 22 may form part of a planetary differential drive (not shown) which divides the torque between a front wheel axis and, via a bevel gear and driveshaft positioned along the central axis of the car, a rear wheel axis.

A further preferred advantage of the twin output shaft arrangement embodying the invention, in comparison with a conventional twin layshaft arrangement, is the freedom to position the plane containing the output gears and combining gear at any point along the axis of the speed changing co-planar gear sets. Power input may likewise be via a transfer gear set positioned anywhere along the central drive shaft. Thus in the arrangement shown in Figure 5, the input is a bevel gear 6 to effect a right angle drive to a complementary bevel gear 8 on the central shaft 12. The plane of the output gears 2OA, 2OB and combining gear 22 lies in approximately the same plane as the bevel pinion axis 4. Arranged with a clutch 56A, 56B at each end, for selecting gears 188, 186, and 184, 182 respectively, this provides a four speed close ratio gearbox for a mid-engined sports car. By providing further clutches and co-planar gear sets at either end of the shaft 12, further gear ratios can be obtained, e.g. three clutches will provide six ratios and four clutches will provide eight ratios for a sports racing car or similar.

Alternatively the planes of the output gears may be offset, with each output gear arranged to drive a divided combining gear (or respective combining gears mechanically coupled to each other) e.g. mounted on each side of a differential unit. The combining gears or divided combining gear parts may be connected for relative torsional movement to provide torque sharing compensation between the two power paths. Alternatively, the combining gear parts may have helical gear teeth of opposite hand and with axial movement to provide torque sharing compensation between the two power paths.

Figure 6 shows a dog clutch and synchroniser ring of otherwise generally conventional form that could be used in the gearbox 10, mounted between the input shaft 12 and each central gear 180-188 or between each output shaft 14A/B and its respective gears 160A/B- 168A/B. The gear wheel is jouraalled to the shaft 12 on bearings 36 and has a dog tooth ring 38 mounted adjacent one face by a splined collar 40. An internally and externally splined ring 42 mounts a clutching ring 46 to the shaft. The clutching ring 46 is axially slideable on the outer splines of the ring 42 by gear selector forks (not shown) engaged in a circumferential groove 44 so as to move its inner splined surface into and out of engagement with the dog teeth 38. The gear wheel is thereby clutched to/disconnected from the shaft. A synchroniser ring 48 is resiliently mounted to the clutch ring 46 by a plate spring 50 so as to frictionally engage a conical centre boss on the dog tooth ring 38 and match the speed of the shaft and gear wheel as the clutch ring 46 is moved towards the dog tooth ring 38. This allows smooth engagement of the clutch ring splines with the dog teeth. A further gear wheel, dog tooth ring and synchroniser ring (not shown) may be mounted on the shaft to the right of the splined ring 42 for co-operation with the same clutch ring 46. The dog clutch arrangement shown in Figure 6 takes up substantial axial space, particularly in a gearbox having a large number of ratios and hence a correspondingly large axial series of dog clutches. Figure 7 shows an alternative arrangement for use at the input shaft 12 or in similar situations where central gears 188, 186, 184, 182 are each simultaneously engaged by a pair of diametrically opposed gears (not shown) whereby circumferentially directed tooth loads on each central gear balance out, and where the central gears are to be selectively clutched to a central shaft 12. Because tooth loads balance out, there is no need to mount the central gears on bearings. The central gears may each therefore be formed with a central, internally splined aperture 52. Tubes 54A and 54B are telescopically slidable relative to each other and on the shaft 12. Tube 54A is internally and externally splined; tube 54B is internally splined and shaft 12 is externally splined so that these elements, whilst being freely relatively axially slidable, are locked together for coaxial rotation. The tubes 54A and 54B each have a freely projecting end provided with a respective coupling ring or flange 56A, 56B. Each coupling ring or flange has external circumferential splines so that flange 56A is moveable from the illustrated central, disengaged, neutral position, for selective engagement in the splined central aperture of either gear 182 or gear 184, as the tube 54A slides axially on the shaft 12. Similarly, flange 56B is selectively engageable in the central apertures of gears 186 or 188, upon axial sliding of the tube 54B. Instead of splines, the tubes, flanges and shaft 12 may have other non-circular external profiles (e.g. with flats to form a triangular, square, pentagonal, hexagonal, etc. section) for engagement with correspondingly shaped internal profiles of other tubes and/or the gear central apertures, as required. Splines are however preferred, as they have the necessary torque transmitting strength. They can also be formed with sufficient clearances for ready engagement and to provide torque compensation movement of the central gears relative to the shaft 12, and/or to allow torque compensating movement of the whole shaft 12 and gears 182, 184 etc. assembly, if necessary. Any number of co- planar sets may be mounted in parallel: the four central gears 182,184, 186, and 188 as shown being for illustrative purposes only. The co-planar gear sets are each mounted on common centres so that the required speed ratios of the gearbox are obtained by means of the differing gear sizes of the gear sets. An alternative mechanism for selectively coupling the required central balanced gears

182 -188 to the driveshaft 12 is shown in Figure 8. The shaft 12 is stepped and hollow. Sets of axially extending, circumferentially distributed slots 60A₃ 6OB are provided in each stepped section. A central selector rod 54B and telescoped selector tube 54A are slideably received within the shaft 12. Externally splined coupling rings 56A, 56B are pinned to the tube/rod 54A, 54B respectively, through the slot sets 6OA, 6OB. The shaft 12, coupling . rings 56A, 56B, selector rod 54B and selector tube 54A are thus pinned together for rotation as a unit; the external splines of the coupling rings 56A, 56B being selectively engageable with the splined apertures of the gears 182, 184; 186, 188 respectively, upon appropriate axial sliding movement of the selector rod 54B and tube 54A.

Access for movement of the coupling rings or flanges 56 A and 56B of Figures 7 and 8 along the gearbox central axis may be provided by mounting one or more offset power input/output bevel gears or transfer gears on the shaft 12 for transferring power from or to one or more sources. Alternatively, the power input/output may be directly to the shaft 12. The rod/tubes 54 A and 54B may be moved axially for engagement of the flanges or rings 56A, 56B with the required gear set by suitable linear actuators or mechanical linkages, e.g. hydraulic or pneumatic cylinders, electromagnetic actuators including screw driven electomechanical actuators or Bowden cables. Connection of the selector rod and/or each selector tube to its actuator may for example be by means of a ring attached to the distal end of the tube, circumferentially slotted for reception of a selector fork.

Alternatives to the dog clutches formed by the splined balanced gear central apertures and coupling rings/flanges discussed in relation to Figures 7-9 are readily possible. For example, the ends of the selector tubes/rod, or the coupling rings/flanges carried by these, may be provided with radially disposed dog teeth selectively engageable with dog teeth provided on (e.g. cut in) the face of the gears 182, 184 etc. The engagement will lock the gear concerned to the shaft, precluding any gear movement for torque balance. Consequently this may be provided mounting the whole telescopic selector assembly and drive shaft 12 or rotating member 54 for radial floating movement. Torque sharing compensation can alternatively be provided elsewhere in the torque loop, e.g. at a combining gear or compound combining gear assembly.

As a further alternative the mechanism to selectively couple the required central drive gear to the driveshaft 12 (not shown) may be a series of axially or radially moveable pegs or latches extendable/retractable from the driveshaft 12 under mechanical, electromechanical, hydraulic or pneumatic actuation, for selective engagement with the corresponding formations in the apertures 52 of the central gears 182, 184. Yet alternatively, the selective coupling mechanism may be a friction drive obtained by application of hydraulic pressure to thin walled chambers formed in the tube 54 or driveshaft 12 where it passes through the apertures 52. Application of hydraulic pressure can be used to expand a localised section of the shaft or tube into frictional driving engagement with the inner surface of the selected aperture in the gear wheel 182 or 184 etc.

Figure 9 is a diagrammatic representation of another clutching and ratio selection mechanism operable at the central balanced gears 180-188. The central shaft 12 is coupled to rotate a further shaft 54 whilst allowing axial telescoping motion between these parts, e.g. by a sliding splined connection. Shaft 54 passes through the balanced gear apertures 52 and is linearly movable as represented by arrow b. Five sets of three co-planar ratio determining gears 180-188; 160A/B-168A/B are disposed along the shafts 12, 14A/B at even pitch spacing P. The shaft 54 carries three rings 56A, 56B, 56C of external dog teeth of width approximately P/6 engageable in corresponding grooves formed in the central apertures 52 of the gears 180-188 of width approximately P/6. The dog tooth rings are evenly spaced along the shaft 12 at a pitch 4/3 that of the gear sets, i.e. 4/3P. As shown, dog tooth ring 56B is engaged in the central aperture of gear 184 for driving engagement with the shafts 14A, 14B via the gears 164 A, 164B. All the other central gears 180, 182, 186, 188 are idling around the input shaft 12 and shaft 54. Linear movement of the shaft by 1/3P to the right from the position shown disengages gear 184 and engages gear 186 via dog tooth ring 56A. Similarly, movement of the shaft 54 to the left by 1/3P engages gear 182 via dog tooth ring 56C. For a total linear movement of the shaft 54 over 4/3P it is possible to select any of the five available gear ratios. Other numbers and configurations of dog tooth rings will be readily apparent, to suit other numbers of available gear ratios. This arrangement is simple, with low parts count and straightforward single linear actuation. This embodiment is preferred when sequential up/down selection may be acceptable, for example in racing cars applications.

In a further arrangement two or more of the above speed changing gear assemblies for example similar to the arrangement shown in Figure 1 are mounted radially around a combining gear, with each pair of output gears engaging the combining gear. As shown in Figure 10 each shaft 12A, 12B of the speed changing gear assembly carries a gear 13A, 13B respectively, both engaged with a centrally disposed input gear 11 on the engine driveline. A gearbox thus configured with two speed changing sections each with two clutches 62, 64, Figure 11 and four gear sets 146A, 146B, 146C, 146D provides an in-line drive with eight gearchange ratios in a very short length. As shown in Figure 11 such an arrangement can be provided with two co-planar sets of output gears and combining gears , the first set comprising gears 2OC, 2OD, 2OG, and 2OH engaged with the first combining gear 22A; the second set comprising gears 2OE, 2OF, 201 and 2OJ engaged with the second combining gear 22B. Either the combining gear 22 A, or alternatively the combining gear 22B is connected to a common output shaft by a clutch 66. This provides a divided torque path range change section in series with the speed change sections. Speed (ratio) selection is by a further pair of clutches 62, 64 supplying power from a shaft 12A via a selected one of the co-planar gear sets 146A, 146B, 146C, 146D to the output gears 2OC, 2OD, 2OE, 20F₃ with similar clutches and similar gear sets (not shown) supplying power from the shaft 12B to the output gears 2OG, 2OH, 201, 2OJ. The number of gear ratios (sixteen as shown) and the overall speed ratios thus provided are suitable for heavy vehicle applications.

The central balanced gears in any of the embodiments described above may be restrained from axial movement arising from vibration or other forces by means of a bearing surface on adjacent gears, or by means of side plates mounted on the gear casing or any other suitable support, or by use of double helical gears.

As a further alternative, the mechanism to selectively couple the required central, balanced gear to the driveshaft 12 is by gear selector forks directly engaging the gears, or engaging in grooved bosses rigidly attached to or integrally formed with the gears, to slide them axially in/out of engagement with dog clutches at the central aperture, for example like the ratio selection mechanism disclosed in US 3105395. This arrangement is preferred when there is restricted input and output co-axial access, e.g. for an in-line engine and gearbox.

Any of the aforementioned clutching arrangements including those of Figures 7-9 can be applied to the multiple speed change section arrangements as described including those as shown in any of the other Figures. Particularly advantageous is the use of the clutching arrangements of Figures 7-9 in the speed (and range) change section of Figure 5, 10 and 11. Here access to the central balanced gear axes for axial sliding of selector rods/tubes is easily achieved because both ends of the balanced gear axes are offset from the axes of the gearbox input and output shafts. This further improves the already advantageous short length of these gearboxes. It also provides the basis for a full power shift gearbox.

In all of the speed changing gearboxes described thus far the clutching mechanisms, whether by "crash" dogs in racing cars or otherwise by controlled synchronised clutching, involve a time period, albeit of very short duration, where the drive from the engine to the ground wheels is disconnected. In a modification of the arrangement shown in Figures 10 and 11, each input shaft 12A, 12B is driven through a separate clutch (not shown). This enables the control system to pre-engage the next up or down gear ratio during the period that power is transmitted through another change speed section. The gearchange is then simply effected by engaging the clutch on the pre-engaged unit under full power. While this kind of dual clutch power shift gearbox is not new, when matched to the compactness of the twin output shaft gearbox arrangement it presents many advantages over conventional planetary automatic gearboxes. An in-line dual path arrangement is ideally suited to luxury passenger vehicles because the conventional torque converter is eliminated and efficiency is much improved.

Ratio selection in autosport applications is usually of the "crash" kind, using a combination of engine torque "back off', including "sparking" the engine rather than clutch disengagement, allied with robust application of force on dog clutches. These arrangements can be replicated in embodiments of the present invention. In more conventional vehicles, near synchronisation of the rotational speed between the connecting shaft and the connecting gear is necessary prior to full clutch engagement. Ratio selection may be automatic or semi-automatic, for example using a micro processor controlled power train management system, which senses and controls one or more of engine speed, acceleration, braking and engine output torque and gearbox shaft speeds and executes synchronous clutch engagement.

The layout of the gearing and clutch arrangements in the twin output shaft speed changing gearbox are conducive to movement of clutches under computer control. When a central gear is disengaged from its drive, the central driving shaft carries no rotating gears, which means that the forces required to speed up or down this assembly to synchronising speeds are minimised because of its low inertia. In one arrangement shown in Figure 12 the central driving shaft downstream of the main clutch drives a flywheel 80 by means of a belt and pulleys 82B, 84B. Pulley 82B (or pulley 84B) is connected to its shaft via a computer controlled (e.g. electromechanical) friction clutch (not shown) the other pulley being directly coupled to its shaft which normally remains engaged. Two further belt and pulley sets 82A, 84A and 82C, 84C are provided in parallel to the set 82B, 84B. These incorporate similar clutches, but which remain disengaged, save during gear changing. Alternatively, jockey wheels or any other suitable means can be used to engage/disengage the various belt drives.

When the control system determines that a gear change down is necessary, drive from the central shaft to the flywheel 80 via the pulley set 82B, 84B is automatically disconnected and drive from the flywheel to the central shaft is instead established via pulley set 82C, 84C₅ which have a higher speed ratio than the pulleys 82B, 84B. Angular momentum is therefore transferred from the flywheel to the central shaft, causing it to increase its rotational speed by a predetermined proportion. The moment of inertia of the flywheel and the ratios of the pulley sets 82B, 84B; 82C, 84C are chosen in relation to the moment of inertia of the central shaft, such that the speed increase matches that necessary for smooth engagement of the dog clutches on the central shaft to effect the change down. Similarly, when the control system determines that a gear change up is necessary, pulley set 82A, 84A is engaged in place of pulley set 82B, 84B. Angular momentum is thus transferred from the central shaft to the flywheel 80, to ensure the necessary proportional central shaft speed decrease for smooth dog engagement of the next higher gear. On completion of the gear change down or up, pulley set 82B, 84B is re-engaged in place of the other pulley sets. Alternatively, some or all of the illustrated pulley sets can be replaced by one in which the effective diameter of one or both of the pulleys can be varied. In an alternative arrangement, not shown, the driving shaft rotational speed is increased by a controlled engine speed increase while a brake is used to reduce it.

In a further embodiment shown in Figures 13 and 14, the drive between the engine and the twin output shaft gearbox is through a speed variator (CVT) 90, for example a friction toroidal ring type or a variable diameter belt type. In this case the variator is used in series with the gearbox 10 to run the engine at a constant and optimum speed for efficiency and performance. The series combination of variator and speed changing gearbox allows the use of a smaller and lighter construction variator, and it is the variator which readily increases or decreases the driving shaft rotational speed toward synchronisation. While this type of "shunt" CVT is not new as such, the new use in combination with the twin output shaft gearbox with its capability to package more gear ratio steps and very low moment of inertia of the input shaft during gear changing is a particularly advantageous arrangement. This is especially so in a transverse engine front wheel drive application as shown in Figures 13 and 14 with a belt variator 90 driving an under axle centrally disposed gearbox 10 with a combining gear 22 which forms part of the axle driveline differential. As shown, the gear 22 is a split ring gear comprising torsionally resiliently interconnected parts 22 A and 22B providing torque compensation.

Alternatively, one of the output gears 20 can be torsionally resiliently mounted to its output shaft or such torsional resilience can be provided elsewhere in one of the output shafts, with both gears 20 simultaneously engaging a unitary combining gear 22. As a further alternative, the combining gear 22 may be a double helical or "herringbone" gear, free to float axially relative to the output gears 20 to provide torque sharing compensation. Similarly, a single helical combining gear 22 can be used, which tilts about its axis to provide torque sharing compensation. Still further alternatively, torque sharing compensation may be provided by movement of the central balanced gears of the speed, ratio selection sets. The split ring gear, axially floating double helical combining gear, single helical tilting combining gear or torsionally resiliently mounted output gear torque compensating arrangements are of course applicable in other embodiments, such as modifications of Figures 4, 5, 9, 11 and 12.

In the preferred arrangement for an in-line racing car gearbox the co-planar gear sets comprise a spur central gear with engagement to an axial clutching arrangement of the single linear actuation type. The preferred torque compensation means is at the wheel axis combining gear so that the input drive shaft and clutching arrangement can be mounted robustly. The output shaft gears are mounted solidly upon the two output shafts mounted in bearings at each end. This eliminates all of the pressure fed needle roller bearings on the output shaft of a conventional racing car gearbox, lowering the parts count and simplifying manufacture by eliminating the associated oil ways. Because of the smaller gears in the twin output arrangement, the gearbox inertia is lower and in idle the only rotating mass is the input shaft. This enables faster gearchanging compared to known gearboxes, a benefit that is enhanced by the relatively stiffer gear change selector mechanisms of the preferred embodiments. Gearboxes according to the present invention can be of simpler construction and more robust than existing units, particularly in the motorsport field.

Claims

Claims:-

1. A plural power path speed changing gearbox comprising a plurality of co-planar gear sets each comprising an input gear meshing two diametrically opposed output gears; the input gears of each set being mounted on an input shaft and the output gears of each set being mounted on different ones of a pair of output shafts; individual co-planar gear sets being selectively drivingly engageable between the input shaft and the output shafts; a closed torque path being formed between the two output shafts; the torque path being closed using an output gear on each output shaft, each engaging with a combining gear; the axis of the combining gear is not co-axial with the input shaft.

2. A gearbox as defined in claim 1 in which each output gear is mounted part way along the length of its output shaft.

3. A gearbox as defined in either of the claims 1-2 in which the combining gear is a ring gear of a differential unit.

4. A gearbox as defined in any of the claims 1-3 in which the combining gear is a split ring gear comprising torsionally resilient interconnected parts.

5. A gearbox as defined in any of claims 1-3 in which the torque path is closed using oppositely handed helical gears respectively driven from each output shaft, each helical gear engaging different halves of a double helical combining gear mounted to float axially to provide torque sharing compensation.

6, A gearbox as defined in any of claims 1-3 in which the torque path is closed using helical gears respectively driven from each output shaft to engage a single helical combining gear mounted to tilt about its axis to provide the torque sharing compensation.

7. A gearbox as defined in any of the claims 1-3 in which torque sharing compensation is provided at the input gear of each speed changing gear set, by movement of the input gear relative to the input shaft, or movement of the input gear and input shaft relative to the remaining gears in each speed changing set.

8. A gearbox as defined in any of the claims 1-3 in which torque sharing compensation is provided by one or more torsionally resilient components.

9. A gearbox as defined in claim 1-2 in which the planes of the output gears are offset, with each output gear arranged to drive a divided combining gear.

10. A gearbox as defined in claim 9 in which the divided combining gear comprises respective combining gears mechanically coupled to each other.

11. A gearbox as defined in claim 10 in which the respective combining gears are connected for relative torsional movement to provide torque sharing compensation between the two power paths.

12. A gearbox as defined in claim 10 in which the respective combining gears have helical gear teeth of opposite hand and can move axially to effect torque sharing compensation between the two power paths.

13. A gearbox comprising a further plurality of co-planar gear sets and output gears as defined in claim 1 wherein the plurality of co-planar gear sets and output gears mesh with a common combining gear.

14. A gearbox as defined in claim 13 comprising a further combining gear which, together with the combining gear, forms a range change section of the gearbox.

15. A gearbox as defined in any of claims 1-14, in which the driving engagement between the selected co-planar gear set and the input and output shafts to provide a selected speed ratio is by means of a dog clutch operating between each input gear and the central shaft.

16. A gearbox as defined in any of the claims 1-14, in which the driving engagement between the selected co-planar gear set and the input and output shafts to provide a selected speed ratio is by means of each input gear being clutched to the input shaft through a central aperture.

17. A gearbox as defined in claims 15-16 wherein the selected clutched engagement is by means of one or more members being rotationally coupled to the input shaft and axially moveable with respect to the input shaft and input gears and engageable with the input gear to provide the selective clutched engagement; the axial movement being effected by a force applied to the member externally to a possessive or connected shift fork groove positioned between co-planar gear sets selectively drivingly engageable by the member.

18. A gearbox as defined in claim 17 in which the member or members comprise a selector rod or tube telescopically coupled to the input shaft via a non-circular section.

19. A gearbox as defined in claim 18 in which the non-circular section comprises splines.

20. A gearbox as defined in claim 18 or 19 in which the selector rod or tube carries a torque transmitting annular member which is provided with non-circular sections on its external surfaces to engage co-operating surfaces of the selected gear.

21. A gearbox as defined in claim 20 in which the non-circular section comprises splines or dog teeth.

22. A gearbox as defined in claim 20 or 21 in which the annular member is provided with a plurality of such sections, axially spaced for engagement with the selected one of a larger plurality of the central gears.

23. A gearbox as defined in any of claims 20-22 in which the annular member is shaped and dimensioned for co-operation with the surfaces in the central aperture to provide torque sharing compensating movement of the central gears.

24. A gearbox as defined in claim 16 in which the clutching means at the central aperture comprises a part comprising or rotably attached to the input shaft and radially moveable into rotational driving engagement with a selected input gear central aperture.

25. A gearbox as defined in any of claims 15-24 in which clutching engagement is synchronised by means of a flywheel.

26. A gearbox as defined in any of the claims 1-25 in which power is transmitted to the input shaft by a set of transfer gears.

27. A gearbox as defined in claim 26 in which one of the transfer gears is mounted part way along the length of the input shaft.

28. A gearbox as defined in any of the claims 1-27 used in combination with a CVT in an automotive driveline.

29. A speed changing gearbox as defined in any of the claims 1-28 used in conjunction with a computer controlled power train management system to provide automatic or semi-automatic gearchanging.

30. A speed changing gearbox according to claim 1 constructed, arranged and adapted to operate substantially as hereinbefore described with reference to as illustrated by the accompanying drawings.