GB2451681A - An integrated plural power speed changing gearbox - Google Patents

Abstract

A plural power speed changing gearbox, e.g. for a winch, comprises one or more input sections each comprising two co-planar sets of gears comprising input gears 22 and 24 each meshing with two diametrically opposed intermediary gears 30A, 30B and 32A, 32B. The input gears 22, 24 of each set being mounted on an input shaft 20 and the intermediary gears 30A, 32A and 30B, 32B being rotationally fixed to shafts 21A and 21B. The input gears 22, 24 being selectively drivingly engageable, via a dog clutch 50 operated by a shift fork, to the input shaft. Intermediary gears 30A and 30B being in engagement with an output gear 40 with an axis offset and substantially parallel to that of the input shaft; the output gear forming a closed torque path. The output gear 40 may form a ring gear integrated in a differential unit of a vehicle driveline. Rotational speed synchronizing may be assisted by external means and double helical gearing may provide axial compensating movement. Compensating movement may be radial, axial, torsional or a combination of any of these.

Description

1 2451681 An inte2rated. niural nower, sneed chan2in2 Learbox This invention concerns a plural power path speed changing gearbox wherein the axes of the input and output shafts are bftetnd substantially parallel. It has widespread automotive and industrial application especially in a dual drive arrangement and in a two speed ratio gearbox being integral with a driven mechanism and providing a corner HP' solution i.e. a drive delivering either high torque at low speed or low torque at high speed.

Known arrangements to provide corner HP include automotive range change' gearboxes, and the various same configurations of gearing are used in industrial machinery e.g. winches. Prior art forms include planetary gearing, conventional single layshaft and plural power layshafi gearing. Speed ratios are changed either manually e.g. ma multi-speed heavy truck gearbox, or automatically with a power break in the drive. There are clear advantages in using plural power paths, the size and the weight being reducei The prior art plural power gearbox arrangements are characterised by the input shaft being co-axial with the output shaft, and by limitations in gear reduction ratios of practically proportioned gear pairs in engagement. An inventive step in this invention is the arrangement of plural power paths in substantially two planes of gears.

In the automotive field the characteristics of prior art range change gearboxes places severe constraints upon the power train design and consequently upon the vehicle design. In a bus for instance, the engine and transmission are preferably positioned at the rear, and lack of axial space in a transversely mounted engine arrangement has encouraged the use e.g. of a spatially effective but poor efficiency torque converter and planetary gear combination.

The present invention seeks to eliminate the above design constraints by the use of non co-axial plural power path gearing, and accordingly provides a speed changing gearbox comprising one or more sets of gear input sections; each section comprising twoS co-planar sets of gears; each set of gears comprising an input gear co-operating with a pair of intermediary gears; each gear of an intermediary pair being diametrically oppositely engaged with the input gear; the intermediary gear pairs in each co-planar set being arranged such that one gear of an intermediary gear pair is rotationally connected on a -common central axis to one gear of the other intermediary gear pair, and the other one gear of an intermediary gear pair is rotationally connected on a common central axis to the other one gear of the other intermediaiy gear pair; each one of the two co-planar input gears being selectively drivingly engageable with an input shaft at the centre; a drivingly engaged input gear transmitting substantially equal force to each gear of the intermediary gear pair in engagement; one of the two co-planar intermediary gear pairs being in engagement with a common output gear forming a closed torque path; each input section being radially disposed around the output gear; the axis of the output gear being substantially parallel and offset with that of the input shaft(s).

In any transmission the use of plural power paths is particularly beneficial at the highest torque position in the drive path. Thus the use of planetazy wheel hubs in e.g. heavy mining vehicles. Using the present invention as a range change gearbox the output gear can be a ring gear of a differential unit, therebypositioning the gearbox in the driveline far more advantageously than a conventional unit positioned upstream of a speed reducing differential. To function properly at this position the range change gearbox has provide a high range' gear reduction equivalent to e.g. a bevel gear pair embedded in a conventional differential, and a useful range shift' ratio. As explained in El' 1 296 847 Bi the range change gearbox as in the present invention can embody co-planar reduction ratios of e.g. 6:1 and 2:1, thereby meeting the practicability requirement. In this preferred embodiment the present invention can replace a conventional range change unit and the conventional bevel gears.

A further advantage of plural power paths in this application is the ensuing reduction in pitch circle diameter of the gears, especially the input gears. Noise is closely associated with pitch line speeds. In an overrunning range change unit of the kind in the present invention, small pitch circles are important.

The size of the output ring gear is determined by the axle torque and the space required for the differential gear elements. By providing two equal power paths the tooth loads are substantially halved when compared with e.g. the ring gear in a conventional transversely mounted automotive engine and gearbox. Thus in the present invention the ring gear can be smaller. In situations where the differential gear elements dictate the ring gear diameter the ring gear facewidth can be reduced. Alternatively the benefits of a smaller diameter can be gained by matching the present invention with a compact differential e.g. of the spur gear type.

The two co-planar gear sets each have two torque paths that are closed at the output gear.

While the substantially equal torque in the paths can be achieved by extremely accurate manufacture and assembly, in preferred embodiments some form of compensatory movement is provided in one or more of the rotating members. A variety of means may be employed, as known to those skilled in the art. They include compensation at the input gears e.g. for input gears of the spur type by providing compensatory movement in a radial direction; e.g. for input gears of a helical type the movement may be axial translation and/or rotation about an axis normal to the plane of the input shaft, similar to the arrangement described in US3 885446. Alternatively the torque sharing may be effected using double helical or "herringbone" gears. Yet alternately compensatory movement may be provided by means of a torsionally flexible input shaft.

The above compensatory movements can be applied to the input gears individually or to the input shaft upon which both gears are mounted, or in some cases to the output gear.

Some of the aforementioned mechanisms may be constructed in combination.

The torque sharing means dictates the design of the input shaft In one arrangement the shaft is a splined extension of a gearbox output shaft journalled in bearings. The shaft extension being subject only to torsional loads and suitably hardened can be diametrically sized to e.g. accommodate a coupling sleeve (for torque sharing purposes) without detriment to the overall proportions of the input gears. Alternatively the extension input shaft is of very small diameter and connected to an outer splined tube, the assembly providing torsional flexibility; and yet alternatively a splined shaft is constructed with coupling teeth and is mounted to provide the requisite degrees of freedom. These mechanisms may be constructed in combination. The preferred embodiments will be dictated by the application, as further described below.

A change in speed ratio is effected by selective clutched engagement between one of the input gears and a shaft at the centre. In preferred arrangements the clutching mechanism is a dog clutch rotationally splined to the input shaft and movable axially to engage co-operating dogs or coupling teeth on the face or at the centre of each input gear.

In a simple configuration the clutch is a conventional shift fork arrangement operated manually without speed synchronisation. This arrangement is suitable for stationary operation, e.g. a winch or the selection of a deep range in a stationary vehicle. More often the clutching arrangement will involve some speed synchronisation during a gear change.

The preferred embodiments, including the innovative use of a multi-function motor used in conjunction with a computer controlled power train management system to provide automatic gearchanging, are further described below. Syncronising rotational speeds by external means prior to the clutch engagement of an input gear not joumalled at the centre enables the introduction of various novel clutching arrangements. These allow the construction of substantially adjacent co-planar gear sets, the space saving being particularly important in a vehicle driveline.

The present invention provides an extremely compact two speed gearbox. In a range change automotive application it is particularly suited to transverse mounting e.g. in a front wheel drive passenger car wherein the conventional output transfer gear can be smaller, and a conventional e.g. five speed gearbox can be co-axially connected thereby enabling unconventional and highly efficient power train modes of operation to be considered.

Alternatively when this configuration is applied in rear wheel drive buses etc. or truck artic units, the conventional cross drive differential mounted bevel gear pair can be eliminated or made considerably smaller. In the automotive field the combination of a high/low range and a close ratio gearbox can provide operational advantage in e.g. aircraft haulage, tractors and other agricultural equipment, and military vehicles.

The output gear can be the final gear wheel in a winch drive train, e.g. mounted on the winch drum. Again the plural drive path means that the pitch circle diameter and or the facewidth of the gear is reduced. The high reduction ratio feature of the present invention beneficially reduces the size and gear ratio of an upstream gearbox. The two speed feature means that the winch productivity is improved when operating under conditions of light loading e.g. pay out of rope. The present invention is applicable in many of these types of corner HP' problems arising in the alternating conditions of high and low loading of process machinery, or in facilitating start-up of such machinery e.g. conveyor drives. A compact two speed gearbox with scope for automatic operation can improve the operating performance of any device subject to a variable load-speed regime, e.g. a wind turbine, or a boat propeller. S. 5

The present invention and its further preferred features are described below with reference to illustrative embodiments shown in the drawings, in which: Fig I is a diagram of the gearbox with one input section;

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Fig 2 is a side elevation of Fig I; Fig 3 is a drawing showing an input assembly incorporating prior art torque sharing means; Fig 4 is a further such arrangement; Figs is a scrap section showing a conventional shift fork actuated dog clutch; Fig 6 is the arrangement as in Fig 4 showing alternative clutching means; Fig 7 is a drawing showing spur gear torque sharing and a further clutching means; Fig 8 is a scrap section showing diametrically compact input gears; Fig 9 is a drawing showing an input assembly incorporating torsional flexibility torque sharing means; Fig 10 is a schematic showing the invention embodied in a computer managed automotive power train; and 5) Fig 11 is a drawing showing two input sections arranged around an output gear.

Figure 1 is a front elevation diagram showing the gear layout in the present invention with one input section. Input gears 22 and 24 with input shaft 20 at the centre are in co-operating engagement with two diametrically opposed intermediaiy gears 30A, 30B and 32A, 32B respectively, the A ones and the B ones being mounted on the same intermediary shaft 21A and 2 lB respectively. The intermediary gears 30A and 30B are in co-operating engagement with a gear 40 with an axis offset and substantially parallel to the of the input shaft. The output gear may form part of a vehicle differential assembly (not shown) or e.g. a winch drum assembly (not shown).

Figure 2 is a side elevation of Figure 1. The co-planar input gear sets have the diametrically opposed A and B intermediary gears rotationally fixed to a common intermediary shaft 21 A and 21 B respectively, each shaft being journalled in bearings 60 A/D. The input gears 22 and 24 are selectively drivingly engageable to the input shaft 20 at the centre, to provide two gear ratios in the gearbox. Torque sharing between the input shaft and the diametrically opposing intermediary gears in each of the gear ratios is provided by either extremely accurate manufacture and assembly, or provision of compensatoiy movement at the output gear or at each input gear or at the input shaft, or a combination of these.

Figure 3 shows two prior art arrangements for torque sharing means at the input shaft and input gears. The lowest gear ratio uses input gear 22, and this being both the lowest driving rotational speed and the smallest diameter input gear then in many applications e.g. winch drives or an automotive range change for occasional off-road at low vehicle speed or infrequent use in a maximum loaded vehicle operating on steep inclines, spur gearing with simple radial compensation so as to be self-centering may be adequate in the low range.

Spur gear 22 is self centering between the two intermediary gears 30A and 30B (not shown), and is clutched to the input shaft 2013 by the axial movement leftwards of the shift fork operated member 50 wherein co-operating dog teeth engage in the aperture 52A. The member $0 is splined to the input shaft 20B that is coupled to input shaft 20A and 20C at each end at the apertures 54A1B, therein providing radial compensation movement for input gear 22 in driving mode, and thereby substantially equ.alising the force applied to the intermediary gears 30A and 30B.

The input gear 24 being the higher gear ratio has larger diameter and higher pitch line speeds during driving condition, and a helical gear type is more appropriate in most applications. Figure 3 shows a prior art arrangement for torque sharing a single helical gear 24 between the intermediary gears 32A1B (not shown). The compensating movement is provide by mounting gear gear 24 to shaft 20D via a torque transmitting sleeve 25. Gear 24 is formed as a ring gear having internal splines complementing external splines on sleeve 25, internal splines on sleeve 25 in turn engage complementary splines on the member 26 rotationally fixed to the shaft 20D. The co-operating splines are short and slightly crowned to allow pivoting of the axes of shaft 201), sleeve 25 and gear 24 relative to each other.

The lead and hand of the co-operating splines are equal to the lead and hand of the helical gear teeth, thereby ensuring that the axial forces on the sleeve 25 and gear 24 arising from torque transmission cancel each other out. The gear 24 has the degree of movement freedom to ensure that the forces transmitted to intermediary gears 32A1B are substantially equal and the torque is shared between them. The input shaft 201) isjournalled in bearings 66AiB and is provide with co-operating dogs at 52B for selective engagement with member 50. Input shaft 20A as shown is the upstream drive output shaft.

In a preferred arrangement for a vehicle range change application where noise is critical e.g. in a bus, the input gear 22 is also a single helical gear provided with conipensatoiy movement of the kind described above, the plane of the helical gear teeth is offset from that of the two helical splines, to accommodate a smaller gear teeth pitch circle diameter (not shown).

Figure 4 shows an alternative prior art arrangement for torque sharing using double helical gears. Double helical input gears 22 and 24 are journalled in a common input shaft 20 in needle roller bearings 68 and 69 respectively. Intermediary gears 3OAIB and 32AIB are mounted to co-operate with one of the helix angles on gear 22 and 24 respectively. The intermediary gears 30A and 30B co-operate with the output gear formed as a double helical (hemngbone) gear (not shown), therein providing the closed torque path. The bearings 68 and 69 allow small axial movement of gears 22 and 24 respectively thereby ensuring that the forces transmitted by the driving input gear 22 or 24 to 30A, 30B and 32A, 32B are substantially equal.

The input shaft 20 is an extension to the output shaft of the upstream drive and the input gears 22 and 24 are individually rotationally connected to it by co-operating coupling teeth on the member 50 at apertures 52A and 52B respectively, the member being rotationally coupled to the input shaft. This coupling arrangement is compatible with the need for the axial degree of freedom described above.

In this arrangement each helix of the output gear engages one of the intermediary gear pair 30i\ and 30B i.e. there is no plural use of individual gear teeth and the arrangement does * 8 not gain the compactness advantage of those described above. However in a automotive range change gearbox application the arrangement provides quiet gearing and a balanced axial force at the output gear, enabling simpler differential construction.

Figure 5 shows a dog clutch and synchroniser ring of otherwise generally conventional form. The gear 22 isjournalled to the shaft 20 on bearings 66 and has a dog tooth ring 88 mounted adjacent one face by a splined collar 84. An internally and externally splined ring mounts a clutching ring 86 to the shaft. The clutching ring 86 is axially slideable on the outer splines of the ring 80 by gear selector forks (not shown) engaged in a circumferential groove 92 so as to move its inner splined surface into and out of engagement with the dog teeth 88. The gear is thereby clutched to/disconnected from the shaft. A synchroniser ring is resiliently mounted to the clutch ring 86 by a plate spring 82 so as to frictionally engage a conical centre boss on the dog tooth ring 88 and match the speed of the shaft and gear as the clutch ring 86 is moved towards the dog tooth ring 88. This allows smooth engagement of the clutch ring splines with the dog teeth. A further gear 24, dog tooth ring and synchroniser ring (not shown) may be mounted on the shaft to the right of the splined ring 80 for co-operation with the same clutch ring 86.

In the simplest contiguration the clutch is operated manually without speed synchronisation, as in stationary operation, e.g. a winch or the selection of a deep range in a stationary vehicle.

More often the clutching operation will involve some speed synchronisation during a gear change. In the present invention, as shown in Figure 2, during the speed synchronising period only the rotational speed of the input shaft 20 is changed i.e. there being no gears connected to the shaft, either in the gearbox or in any upstream gearbox (as shown in Figure 10). However, while the rotating mass is small, the rotational speed variation is high and as the inertia forces are proportional to the square of the latter, substantial mechanical and thermal loadings are imposed on the synchronising elements.

In Figures 3 and 4 the gear shift actuation is shown diagrammatically using these conventional shift fork actuated clutching means. However, the dog clutch arrangement shown in Figure 5 takes up substantial space, particularly in a range change gearbox application incorporating the synchronising elements. While the combination of prior art torque sharing means and prior art gear shifting means can be used in the present invention, improvement upon conventional means is preferred.

Advanced computer control systems are now commonly applied in e.g. vehicle power train control systems to optimise engine and driveline operation for high efficiency and to provide automatic gear changing. In preferred embodiments of the present invention these computer control systems are used in combination with synchronising means applied to a simple clutch operating between a selected input gear and a shaft at the centre, as further described below.

Figure 6 shows one example of apreferreci clutching arrangement when external synchromsing means are used. The arrangement of gears is as in Figure 4, however the conventional shift-fork member 50 is replaced by an electro-magnetic clutch system comprising electrical windings 70 and a clutching ring 72. Application of an electrical current in the windings will induce an axial force in the clutching ring, and reversal of the current will induce a force in the opposite direction. The current is actuated and the movement effected simultaneously at substantial synchronisation of the rotational speed of input gear 22 or 24 with the shaft 20, achieved by external means. At that point the loads required to engage the external coupling teeth on clutching ring 72 with co-operating ones on input gear 22 or 24 are relatively small. The axial space between the co-planar input gear sets is substantially reduced. Other simple means to produce sufficient force to effect axial movement include arrangements using hydraulic or fluid4ynamic or radially applied mechanical forces (not shown).

Figure 7 shows a further clutching arrangement applicable to the present invention, also reliant upon the externally provided means for synchronising speeds prior to engagement as described above. In some arrangements of gears e.g. as in Figure 3 above, the input gear floats' without any permanently connected member at the centre, especially the bearing required in a conventional single power path arrangement. . Such an input gear can be simply clutched using coupling teeth at the centre, and the clutching actuating means can readily be mounted externally to the axial space between the two co-planar gear sets, thereby further reducing the gearbox size. In Figure 7 the input gears 22 and 24 are self- centering spur gears between intermediary gears 30A, 30B and 32A,32B respectively (not shown).

Shaft 20 is splined and carries a clutching actuator 73 with coupling teeth at one end alternately in engagement with co-operating teeth at the centre of the input gears 22 and 24; the clutching ring is moved axially by a force applied at the opposite end. With no actuating mechanisms positioned between the input gears the face of the gears are substantially in the same plane, as shown.

In this arrangement the co-planar gear sets are substantially adjacent to each other, and the bearing span across the intermediaiy gears and the gearbox width is further reduced.

In an automotive range change application it is especially important to keep the pitch line speed of the low ratio input gear as low as possible in the unloaded over-run' condition.

The use of plural power paths can significantly reduce the size of the input gear for a given torque carrying capacity, thereby going some way towards acceptable standards. In some cases the geomet!y of the gearbox in the present invention used as a two speed range change unit may be dictated by the size of the differential unit and therefore the output gear surrounding it. The constraint may have adverse effect on the pitch line speeds of the low ratio input gear, in which case in a preferred arrangement the present invention is used in combination with a differential unit of the spur gear type, diametrically more compact than conventional bevel gears (not shown).

Figure 8 shows another arrangement of gears designed to minimise the pitch line speed of low range input gear 22. The double helical gear teeth are formed integrally with the shaft, the shaft being journalled in bearings 65, the assembly being mounted at the centre of the high range input gear 24. The high range input gear is journalled in bearings 67 grounded to the housing. The input gears 22 and 24 are alternately drivingly engaged to the input shaft 20 by means of co-operating coupling teeth at the apertures 52A and 52B respectively by means of the moveable clutching ring 72.

Figure 9 shows a further arrangement for single helical gears wherein the torque sharing means includes torsional flexibility in the driving shafts. Input shaft 20A is an output shaft extension of the upstream drive and is hollow with coupling splines on its inner circumference, the splines co-operate with and drivingly engage splines at one end of shaft 20B. The identical splines at the other end of shaft 208 co-operate and drivingly engage with the annular ring 27 upon which are mounted single helical input gears 22 and 24 journalled in bearings 68 and 69 respectively. The shaft 20B is characlerised by the shaft section between the splines at each end being torsionally flexible. The input gears are restrained axially to react the axial thrust arising from the helix angle of the teeth (not shown), and during drive mode the torsional flexibility in shaft 20B ensures that the forces applied to the intermediary gears are substantially equal.

In the arrangement in Figure 9 the clutching ring 72 is splined to the annular ring 27. The clutching ring is connected to a clutching actuator 73 constructed as an inner sleeve by e.g. three pegs operating in slots in the annular ring (one shown). The input gears are drivingly engaged by axial movement of the clutching actuator, the arrangement once more providing an extremely compact gear cluster.

In an alternative construction of Figure 9 the external coupling splines at the ends of shaft 20B are helical splines. The annular ring 27 engages the input gears by means of' helical splines and the annular ring acts in the manner of clutching sleeve 25 of Figure 3 to ensure torque sharing (not shown).

The various arrangements of input gear types and clutching means and torque sharing means described above can be used in the present invention in different combinations than those specifically described; and more than one of the various torque sharing means, particularly precision manufacture and assembly, may be featured together in a power path.

Figure 10 shows a single input section arrangement of the present invention embodied in a vehicle as a range change gearbox wherein the gear shift operations are carried out in combination with a power train control computer system.

The prime mover 100 drives a four speed gearbox through a main clutch 160, first gear being provided by engagement of gear 150 to shaft 20 by means of clutch 156, and fourth gear being provided by the engagement of shaft 20 to the input gear 152 by means of clutch 154. The range change gearbox of the present invention is co-axial with this four speed gearbox and the output gear 40 forms a ring gear driving the differential unit 110, the differential gears driving the wheel axles 120 and 130. During low range the input gear 22 is drivingly engaged with shaft 20 by means of clutch 50 equal torque being transmitted to intermediate gears 30A and 30B (not shown) each engaging the output gear with axis 120, 130 parallel and offset from that of the shaft 20.

The driveline can be constructed using conventional shift fork clutches with integral synchronising means at 154, 156 and 50. Operated manually for a range upshift the driver will in sequence: disengage the main clutch 160, disengage clutch 154, disengage clutch 50, engage clutch 50 to input gear 24 (high range), engage clutch 156 with gear 150 (upstream gearbox first gear) and engage 160. This manual operation can readily be replicated using a computer system to provide an eight speed automatic gearbox with power breaks.

In a preferred arrangement the synchronising function of clutch 50 above is eliminated, the rotational speed synchronisation between an input gear and a shaft at the centre being by means of the external motor 140 drivingly engaged to shaft 20, and the clutch is simplified as described in Figure 6,7,8 and 9.

In one arrangement the motor 150 is a flywheel and pulley system as described in GB 2424682A. Alternatively the motor is a variable speed hydraulic or electrical motor being used to control the speed of shaft 20. Advanced vehicle power trains are increasingly featuring e.g. regenerative braking and/or dual power paths incorporating an electrical or hydraulic motor/generator, and in a preferred arrangement the device can readily be used to provide synchronising means in the present invention, and additionally also provi de vehicle reversing means. Following disengagement of clutch 50 the device is used to accelerate or decelerate the shaft 20, the input gear and shaft speeds are continuously measured and a computer control system activates the axial movement of the clutch to execute engagement at substantially synchronous rotational speeds of the appropriate input gear and shaft. The arrangement is ideal for a bus application wherein a regenerative braking system is particularly advantageous.

In some vehicle drivelines of the kind in Figure 10 the power breaks described above are unacceptable, and gear changing under full power is preferred. In a full powershift version of the two speed gearbox in the present invention the clutch 50 is a multiple plate clutch operating at each input gear (not shown). The arrangement can be used in series with a torque converter (to provide usable torque spread for a vehicle application) or with a fluid coupling (to provide fully variable speed for e.g. a wind turbine application). Alternatively the arrangement can be matched with a conventional e.g. upstream three speed powershifI gearbox.

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The various gearbox synchronising means and operating modes described above, and the various arrangements of input gear types and clutching means and torque sharing means described above can be used in the present invention in different combinations than those specifically described.

The combination of the present invention and advanced computer control systems and dual power arrangements provides the power train designer with scope for significant improvement in vehicle efficiency over a complete duty cycle. For example, the availability of more gear ratios provides opportunity for extending vehicle operation in higher gears and the engine operation at lower speeds and higher torque (improved efficiency). Alternatively, loaded and unloaded vehicles can be operated in efficient low or high range respectively, with ratio overlap between them.

The above description and embodiment of the present invention is predominately concerned with applications using a single input section, and especially its use in a vehicle driveline as a range change gearbox. Figure 11 shows a speed changing gearbox as in Figure 1 with another input section, to provide four gear ratios. The second input section comprising input gears 180, 182 are in engagement with intermediary gears I 84A, I 84B and I 86A, 1 86B respectively. The intermediary gears 1 84A and I 84B are in engagement with the output gear 40, in common with gears 30A and 30B. As drawn the gear ratios of 3.5 and 4.5:1 are added to the original 2 and 6:1, providing a dual, plural power path four speed gearbox. The gearbox can incorporate in any combination any of the input gear configuration, torque sharing means, clutching arrangements, synchronising methods including the multi-function motor, and upstream driver configurations described above.

In a preferred dual plural power path arrangement one input section or plural power path transfer gearbox or single gear in engagement with the output gear is drivingly engaged with an electrical drive for lower vehicle speeds; and an input section of the same output gear is drivingly engaged with an engine and torque converter for high speed cruise and burst' acceleration (not shown). While this gearbox can be installed in a conventional in-line layout driving a bevel differential, the preferred arrangement is a transverse mounting with the output gear being integral with the differential.

Another application for the gearbox as described in the present invention is in a direct wheel or sprocket drive wherein the gearbox is mounted upon the driving shaft of one or more wheels of a vehicle; in such a drive of e.g two opposing wheels the arrangement also provides a steering function (not shown). In combination with the use of high power density hydraulic motors the arrangement is applicable for an agricultural tractor, or off-road vehicles e.g. a quad' bike.

Claims (28)

1. Claims:- 1. A speed changing gearbox comprising one or more sets of gear input sections; each section comprising two co-planar sets of gears; each set of gears comprising an input gear co-operating with a pair of intermediay gears; each gear of an intermediaiy pair being diametrically oppositely engaged with the input gear; the intermediary gear pairs in each co-planar set being arranged such that one gear of an intermediary gear pair is rotationally connected on a common central axis to one gear of the other intermediary gear pair, and the other one gear of an intermediary gear pair is rotationally connected on a common central axis to the other one gear of the other intermediary gear pair; each one of the two co-planar input gears being selectively drivingly engageable with an input shaft at the centre; a drivingly engaged input gear transmitting substantially equal force to each gear of the intermediary gear pair in engagement; one of the two co-planar intermediary gear pairs being in engagement with a common output gear forming a closed torque path, each input section being radially disposed around the output gear; the axis of the output gear being substantially parallel and offset with that of the input shaft(s).
2. 2. A gearbox as in claim I wherein the output gear is characterised by its indirect mounting to one or more shafts at the central axis.
3. 3. A gearbox as in claim 2 wherein the output gear is in the form of an annular ring and forms part of a vehicle differential.
4. 4. A gearbox as in claim 2 wherein the output gear forms part of a winch drum.
5. 5. A gearbox as in claim I wherein a fi.irther one or more gears are drivingly engaged with the output gear.
6. 6. A gearbox as in claim 3 wherein the gearbox is a range change gearbox section in a vehicle driveline.
7. 7. A gearbox as in claim 1 wherein the force equalisation and torque sharing means is either by precision manufacture, or by providing compensatory movement in one or more of the gearbox drive path members, or is a combination of these.
8. 8. A gearbox as in claim 7 wherein the compensatory movement is provided at or by the selected driving input gear of each input section.
9. 9. A gearbox as in claim 8 wherein the compensatory movement is either radial or pivotal or axial or torsional or is a combination of any of these.
10. 10. A gearbox as in claim 9 wherein the gear is a self centering spur gear with coupling teeth at the central aperture selectively drivingly engaged with coupling teeth allowing compensatory movement or is connected to a shaft at its centre providing the requisite radial degrees of freedom.
11. 11. A gearbox as in claim 9 wherein the gear is a single helical gear.
12. 12. A gearbox as in claim 11 wherein the compensatory movement is provided by two sets of co-operating helical splines mounted in offset and substantially parallel planes to that of the helical gear with appropriate lead and hand relative to the helical gear teeth to balance the axial forces.
13. 13. A gearbox as in claim 11 wherein the compensatory movement is provided by torsional flexibility in a shaft.
14. 14. A gearbox as in claim 9 wherein the gear is a double helical gear configured to provide axial compensating movement.
15. 15. A gearbox as in any of the previous claims wherein at each input section the rotational speed of an input gear and a shaft at the centre prior to selective engagement is effected by means positioned externally to the annular space between the co-planar gear sets.
16. 16. A gearbox as in claim 15 wherein rotational speeds are substantially synchronised by means is a pulley and flywheel system.
17. 17. A gearbox as in claim 15 wherein the rotational speeds are controlled by means of a variable speed motor operating on the shaft.
18. 18. A gearbox as in claim 17 wherein the motor is a hydraulic or electric motor.
19. 19. A gearbox as in claim 18 wherein the motor also operates as a part of a braking energy recovery system or as a drive power supply device or as a vehicle reversing means device or a combination of any of these.
20. 20. A gearbox as in any of the previous claims wherein in each input section the selective driving engagement means is the axial movement of a member rotationally coupled to the input shaft.
21. 21. A gearbox as in claim 20 characterised by the actuation means for the axial movement being other than forks mounted between the co-planar gear sets operating on a circumferential groove connected to the member.
22. 22. A gearbox as in claim 21 wherein an electro-magnetic or hydraulic or fluid dynamic or radially applied mechanical force is applied to the member between the planes of the co-planar gear sets to produce the axial movement
23. 23. A gearbox as in claim 21 wherein a force external to the annular space between the co-planar gear sets is applied to the member to produce the axial movement
24. 24. a gearbox as defined in any of the claims 1-5 and 7-23 wherein the input sections comprise differing input gear type or means of torque sharing or clutching or synchronising or any combination of these.
25. 25. An arrangement wherein two or more of the gearboxes as defined in any of the previous claims are individually mounted to drive an individual vehicle wheel or sprocket.
26. 26. An arrangement as in claim 25 wherein one or more of the gearboxes transiently operate at different rotational speeds at the output shaft thereby providing a steering means for the vehicle.
27. 27. A gearbox as defined in any of the previous claims used in conjunction with a computer controlled power train management system to provide automatic or semi-automatic gearchanging.
28. 28. A gearbox according to claim 1 constructed, arranged and adapted to operate substantially as hereinbefore described with reference to as illustrated by the accompanying drawings.