GB2451246A

GB2451246A - Powershift gearbox with synchronizing and energy recovery power units

Info

Publication number: GB2451246A
Application number: GB0714400A
Authority: GB
Inventors: John Gordon Madge
Original assignee: Individual
Current assignee: Individual
Priority date: 2007-07-21
Filing date: 2007-07-21
Publication date: 2009-01-28
Also published as: GB0714400D0

Abstract

A powershift gearbox 10 comprises two input shafts 14A, 14B which are driven, via two main clutches 62A, 62B, by a prime mover 90. The input shafts 14A, 14B have input gears 22-34 which are fixed to the input shafts 14A, 14B by dog clutches 60A-60D and are meshed with output gears 25-37 permanently fixed to output shaft 16. Power units, e.g. electric motors/generators, hydraulic motors 80A, 80B coupled to input shafts 14A, 14B may be used to provide addition torque when starting/accelerating or they may be used to synchronize dog clutches 60A-60D with input gears 22-34. Alternatively the power units 80A, 80B may be used to provide regenerative braking. In another independent claim the invention relates to magnetically or hydraulically controlling engagement of the dog clutches. For example, magnetic clutching rings (74A, 74B, fig 5) are caused to move axially by a magnetic field created by an electrical current in windings (112A, 112B). Clutching rings (74A, 74B) haves external splines which engage with internal splines in gears 24-32 and the clutching rings (74A, 74B) are moved in either direction by reversal of current in the winding (112A, 112B).

Description

245d1246 A Dual Path Speed Changing Gearbox

Background

Owing to legislative threat concerning environmental pollution and carbon footprints', vehicle design is increasingly focussed on fuel emissions and power train efficiency.

In one line of development vehicle designers are turning to wholly electrically powered vehicles. Such vehicles are restricted to town-based short trips because of the massive battery systems associated with even this kind of limited use.

Thus the attention paid to the so called hybrids' -combinations of a conventionally fuelled engine with various electrical systems providing driving and energy storage functions. The most successful embodiment relies upon a purpose designed high efficiency expansion ratio cycle engine in split power paths with an electrical generator and an electrical motor. A vital element of the overall efficiency is the braking energy recovery achieved by running electrical motors as generators and storing the energy in advanced technology batteries. The system foregoes the benefits to power train efficiency provided by a multi speed gearbox, and is relatively high cost.

Other approaches have focussed primarily on the aspect of braking energy recovery, using electrical systems as described above or hydraulics based arrangements. The advantages derived from multi speed gearboxes are retained.

However, the regenerative braking systems are most readily suited, almost by definition, to vehicles that are constantly starting and stopping, e.g. town based utility and transport vehicles. Drivers of these vehicles prefer a fully automatic gearbox of the powershift type. Unfortunately, from an efficiency perspective, an inefficient conventional powershift automatic gearbox is not an attractive fit with a regenerative braking system.

This invention addresses this incongruity. It provides an efficient powershift or power-break automatic speed changing gearbox with design capacity for many more speed ratios than a conventional planetary automatic, thereby enabling more efficient engine operation. And braking energy recovery and release are integrally featured.

Summary of the invention

The present invention provides a speed changing gearbox which is suitable for a broad range of automotive applications, including all types of commercial, public transport and passenger car usage, and which can provide a larger number of gear ratios than existing gearboxes of equivalent power and space. The development is complementary to current rapid developments in vehicle and power train management computer systems. The efficiency benefits arising from the increased number of gear ratios are combined with an integral braking energy recovery system to further improve overall power train efficiency. The invention eliminates the oil immersed clutches and planetary gearing of conventional power shift gearboxes.

Accordingly, in a first aspect, the present invention provides a powershift speed changing gearbox comprising a plurality of co-planar gear sets each comprising co-operating gears in engagement; individual co-planar gears being selectively drivingly engageable by a members selected clutched engagement between one of the co-operating gears and a shaft received at the centre; the co-planar gear sets being arranged to provide two primary drive paths; each drive path being separately clutched to the prime mover; during vehicle normal forward driving each primary drive path being alternately in a power path mode and in a pre-engaged mode; each primary drive path being serviced by a power unit that during pre-engagement mode intermittently provides speed synchronising means to enable selective clutched engagement and at other times intermittently provides braking energy recovery means.

Layshaft type gearboxes with selective gear ratio by means of synchronised dog clutch engagement are the most widely prevalent, for good reasons. They are easy to operate, efficient and relatively cheap. Advances have been made in two directions. The first is automation'. By linking computer control to a set of actuators that basically replicate what a human driver does with clutch and gearshift operation, the manual' layshaft gearbox has become an automatic' albeit with a power-break. Various means, including engine speed control and input shaft braking, can be employed to assist with synchronisation of gear and shaft speeds during a gearshift. However, current designs generally retain conventional synchronising means in such developments, to the detriment of gearbox length. In the second line of development two layshaft type gearboxes have been integrated into one, a dual layshaft arrangement with two independent main clutches. Effectively this is two independent gearboxes operating alternately in pre-engaged mode and power path mode. The advantage of this type is the scope to change gear ratio under full power i.e. to powershift. The designer has again continued to rely upon conventional synchronising means for the pre-engagement of the gear ratio prior to switching of the main clutch.

in both of the developments the retention of conventional clutching arrangements restricts the number of gear ratios that can be used. Current conventional speed change gearbox clutches are mostly shift fork arrangements mounted between co-planar gear sets that actuate a dog clutch by axial movement.

In single layshaft transmissions the speed synchronisation between the gear and the shaft at the centre is effected by friction between a synchronising ring and the gear.

The clutch adds considerable length to a gearbox, the clutch actuation and shift fork assembly being usually two or three times the facewidth of the engaging gear.

The arrangement severely restricts the number of gear ratios, and as a consequence the application scope and the scope for engine efficiency optimisation. This is because the design of a layshaft gearbox is severely constrained by the span of the gearbox shafts between the bearings (because of bending loads arising from gear loads, and their negative effect upon gear teeth engagement geometry).

In an inventive step the present invention takes the best features of these current developments and combines them with other advances that remove the constraints on the number of gear ratios. Thus the efficiency advantages of layshaft type gearbox layout are retained. And the powershift advantage of the dual layshaft arrangement is gained. The scope to increase the number of co-planar gear sets and therefore gear ratio steps stems from two advances. The first is the provision of external means for precise synchronisation of gear and shaft speeds prior to their selective pre-engagement. A variable speed motor is used to drive the idling shaft in the pre-engaged mode to the required synchronous speed. Each of the drive paths in the dual layshaft arrangement is provided with such a power unit. The power unit feature eliminates the need for conventional friction synchronisers mounted between co-planar gear sets, thereby reducing the bearing span and enabling introduction of further co-planar gear set(s). The second advance arises from the dual path arrangement wherein it provides scope to readily accommodate a speed increasing gear at the engine output. Higher rotating shaft and gear speeds for the same power output further reduces the size of the components, and the shaft loading in the gearbox. An eight speed power shift gearbox can be readily constructed. A transversely disposed arrangement, with co-axial arrangement of the two layshaft gearboxes is ideally suited for a high power, mid engined sports car with rear wheel drive.

The increased number of gear ratios enables the engine to be operated at an advantageous speed-torque regime to optimise fuel efficiency. Twelve gear ratio steps can be envisaged. In this configuration, unlike existing dual path arrangements, the two main clutches may be a dry' type (because of the much reduced differential speeds across the clutch), resulting in substantial improved efficiency compared with the wet' clutches used in current dual path gearboxes.

Efficiency is vitally important in certain applications e.g. luxury passenger cars.

This type of vehicle is coming under particular scrutiny for environmental reasons.

In a further inventive step, the invention provides for even further efficiency improvement by using the previously described synchronising power units as components in a regenerative braking system. For example, when an electrical motor is used and the vehicle is braked the motor is switched to function as a generator and provides retardation to the vehicle as well as delivering energy to the vehicle electrical battery storage system. An electrical system is suited to a luxury passenger car. Alternatively, a hydraulic arrangement uses the synchronising hydraulic motor as a hydraulic pump in retardation mode. The pump can charge e.g. a nitrogen filled hydraulic accumulator for energy storage. A hydraulic system is suited to many types of utility vehicles that are commonly equipped with hydraulic systems e.g. environmental cleaning vehicles. Because the vehicle is constantly stopping and starting, regenerative braking is particularly beneficial.

Release of the stored energy is effected through the power units, either singly or in combination (once a pre-engaged power path has a gear selectively pre-engaged). While this operating condition means that less engine power for a given vehicle load is needed, thus bringing a further efficiency benefit, the condition also extends the potential applications for the invention. A multi speed layshaft gearbox, irrespective of the number of ratios, has limits to the overall gear ratio spread that can be formed between gear pairs on a common shaft. Thus the reason for fitment of e.g. range-change' units in heavy truck vehicles, to increase the ratio spread. Similarly fluid torque converters are applied to ratio limited planetary arrangements. In the invention the power units can be used to provide starting and accelerating torque in parallel to that delivered by the engine. With improved torque spread, application in e.g. transportation buses becomes viable.

In the invention each power unit as described above can function as a precision variable speed setting device and a power absorption device in a regenerative braking system and a power delivery device for stored energy recovery. The power units, either singly or in combination, can be also be configured for reverse rotation, wherein they provide the means for vehicle reversing. This eliminates the clutches and gears required for the purpose in a conventional speed changing gearbox, enhancing the capacity for forward speed changing gear ratios.

A speed increasing gear at the input to the speed changing gearbox has positive benefits for the design of the power units. Being mounted on the layshaft or input shaft they rotate at high speed, meaning a smaller size power unit for given power. Further gear reduction can readily be provided to operate the power units at further increased rotational speed and consequently very high power density. Hydraulic arrangements are particularly space efficacious.

The use of the power units to set precise input or layshaft speed prior to a members selected clutch engagement between one of the co-operating gears and a shaft received at the centre means that the force needed to effect engagement or disengagement of the member is considerably lower than in conventional fork shift clutching mechanisms. Nevertheless, as described above, the conventional arrangement remains prevalent in every type of advanced layshaut gearbox including twin layshafts, automated layshaft and dual layshaft gearbox constructions. Although the fork shift means can be used in the present invention, in preferred embodiments of the invention and in a further inventive step the members are actuated by other means. Relatively light loads mean that conventional mechanical forks can be eliminated, and force can be provided by various electro-magnetic or fluid dynamic means. Although such an arrangement can be used with advantage in a powershift speed changing gearbox of the kind forming the first aspect of the invention as described above, it is of more general applicability, particularly in a power-break automatic gearbox.

Therefore, in a second aspect, the present invention provides a speed changing gearbox comprising a plurality of co-planar gear sets each comprising co-operating gears in engagement; individual co-planar gears being selectively drivingly engageable by a members selected clutched engagement between an aperture in one of the co-operating gears and a shaft received at the centre; the plurality of members being individually rotationally coupled to the shaft and axially moveable with respect to the shaft and gear and engageable with the gears aperture to provide the selective clutched engagement; the axial movement being effected by a force applied adjacently to a co-operating gear; the force being non-mechanical or radially applied.

The present invention provides a clutch operating at the central aperture of a gear in driving or driven engagement with one or more radially disposed gears, the co-planar gear set comprising part of a speed changing gearbox. it has application in every type of layshaft gearbox, including twin layshaft arrangements, provided that the relative rotational speed difference between the engaging gear and the shaft at the centre is small. While non synchronised engagement is viable in e.g. a racing car application wherein the gear ratio steps are small, for all normal applications a means external to the clutch is provided to synchronise the speed of the member and the speed of the selected gear prior to actual engagement, the whole process being under computer control. Various synchronising means may be employed, as further described below.

A conventional Jayshafi clutching arrangement engages and disengages a clutch mounted between a gear and a shaft received at the centre to effect a speed change and is reliant upon significant axial force to effect final synchronisation and engagement. In the invention this force is unnecessary and the member may be reduced to a small individual cylindrical shaped component with low mass and inertia, albeit being sufficiently robust to transit the torque between connecting elements. The invention eliminates the conventional fork actuation gearshift mechanisms conventionally mounted between co-planar gears.

For example the member may be a permanent magnet in the form of a ring, internally splined to provide driving engagement and mounted on a splined shaft; externally splined to co-operate with a splined aperture in a central gear. Axial movement from a neutral position may engage one or other of two gears straddling the ring. The force required to move the ring into engagement may be by means of a magnetic field induced by passing electrical current in windings mounted between the two central gears. This arrangement can readily be embodied in a six speed changing automatic gearbox for a small passenger car.

Alternatively, the force required to move the ring may be by means of an axially moveable permanent magnet mounted inside a non-magnetic hollow shafl e.g. Kevlar'Such an arrangement would be suited to a five speed twin layshaft speed changing gearbox wherein the member is engaging a floating' centrally disposed gear and axial movement of members are unhindered by bearings.

The aforementioned electro-mechanical arrangements benefit from current rapid advance in vehicle electrical system technology. Today, vehicle electrical systems can operate at much higher voltages than the conventional 12V system and they can provide, together with advanced magnet technology, increased actuating forces. The combination of higher actuating forces and lower reacting forces derived from external synchronising enables the inventive steps made in the invention.

Yet alternatively, the member may be constructed with dog teeth on the side faces to engage with similar dog teeth on the engaging gear. This arrangement offers more latitude for minor discrepancy in exact synchronisation, and is suitable for sports and racing car application. In this case the actuating force may be provided hydraulically or pneumatically, by enclosing the rotating member within a stationary cylinder, wherein the member acts like a piston.

In a further embodiment of the invention the members axial movement may be effected by dynamic fluid force. The shape of the members outer diametral peripheral surface is shaped to change the direction of fluid impinging upon it, thus causing an axial force. Varying the direction and volume of the fluid flow controls the members axial position. The force may be produced e.g. from radially opposing fluid jet streams. Such an arrangement is suited to a utility vehicle application, which is usually well equipped with hydraulic systems and pump drive facilities. Alternatively, the radially directed force can readily be provided mechanically, by means of a rod engaging with the members shaped surface.

While all of the power unit features described in the first aspect may be applied to the invention in the second aspect, further arrangements for external synchronising are envisaged. Known technologies include the use of a flywheel, wherein the operation is integral with the power train computer management system. And synchronisation may be effected by engine acceleration for downshifting and shaft braking for upshifting. Such arrangements are suited to e.g. a six speed automatic transmission for a passenger car, this type of vehicle benefiting less than many commercial types from the regenerative braking feature of the invention in the first aspect.

The invention is less space intensive than presently known technology. The invention enables faster gearshifts owing to lower rotational inertia than other known clutching means. It is particularly suited to an integrated computer controlled vehicle, engine and gearbox management system. The clutch arrangement enables a gearbox designer to incorporate more gear ratios in a given space, giving both improved vehicle performance and improved fuel efficiency.

The invention has application in every arrangement of parallel shaft speed changing gearbox including single, twin and dual gear path types. The invention can be applied to most types of wheeled vehicles.

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

Brief Description of the Drawings.

Figure 1 is a prior art conventional single layshaft gearbox.

Figure 2 is a schematic section of a powershift eight speed dual layshaft speed changing gearbox according to the first aspect of the invention and Figures 3 A-C show examples of its operating modes Figure 4 is a prior art conventional arrangement for a fork operated synchronising dog clutch and; Figure 5 is a schematic section of an electro-magnetic clutch according to the second aspect of the invention and; S Figure 6 is a further such arrangement using permanent magnets; Figure 7 is a schematic section of a hydraulic piston type clutch; Figure 8 is a schematic arrangement of a fluid dynamic clutch; Figure 9 is a schematic arrangement of a six speed automatic gearbox in accordance with the second aspect of the invention.

Description of the Preferred Embodiment.

Figure 1 shows a prior art schematic drawing of a conventional six speed, speed changing gearbox as used in many front wheel drive passenger car applications. The gear ratios are obtained by selective clutching of six pairs of gears 22-23, 24-25, 26-27, 28-29, 30-31, and 32-33 to either the input shaft 12 or the output shaft 16 by means of the three synchronising dog clutches 60. (For simplicity the gear teeth are shown schematically throughout the description, showing engagement at the pitch circle diameter). The input shaft 12 is usually rotating at the speed of the engine. The output shaft 16 is drivingly engaged with the differential unit through a reduction gear. The schematic illustrates the design constraints caused by the size of the clutches 60 in the conventional two shaft gearbox. They engage with the larger gears which means providing shift fork facilities to both shafts. And because minimal shaft deflection between bearings I O2AIB and 1 OOAIB is critical for effective gear tooth loading, the width of the clutches 60 limit the number of gear sets that can be accommodated between bearings.

The present invention concerns various combinations of advancement that eliminate the aforementioned design constraint thereby enabling new arrangements of speed changing gearbox, especially of the automatic type. Figure 2 shows a schematic diagram of a preferred arrangement for an eight speed, powershift speed changing gearbox 10 according to the first aspect of the invention, illustrating how these combined advances facilitate new constructions of gearing.

The axis of the gearbox is parallel with that of the vehicle axle and is at right angle to prime mover 90 that drives two input shafts 14A/B via a speed increasing bevel gear set 38 and two main clutches 62A/B. The speed increasing gear is preferably of not less than 1.25:1 ratio, meaning that all of the rotating elements in the gearbox are smaller than in a convention arrangement as Figure 1. In a geometrically similar arrangement all of the speed change gear elements, including the input gears 22,24,26,28,30,32 and 34 are sized at least 7.7% smaller.

Additionally, each of the two input shafts 14AIB is engaged with a multi-function power unit 80A/B. One of the functions is to intermittently rotate the shafts I 4AIB at a speed equal to that to one of two clutches 6OAIB and 6OCID respectively. This synchronisation process eliminates the conventional internal mechanical means and reduces the length of the clutches 60A/D.

The gearbox 10 provides two primary power paths independently co-axially disposed about the prime mover 90 output shaft The first path comprising the main clutch 62A driving input shaft 14A; four input gears 24,28,32 and 36 being journaled to the input shaft; one of the four input gears being selectively engageable to the input shaft in an aperture at the centre by means of one of a pair of dog clutches 60A18; the four input gears individually engage an output gear 25,29,33 and 37 respectively. The second independent power path being 62B and 14B; 22,26,30 and 34 and 60C/D; and 23,27,3 land 35 respectively. All of the output gears being rigidly connected with the output shaft 16, the output torque being transferred to the vehicle differential mounted on shaft 18 by means of a high reduction ratio gear pinion 40 and gear wheel 42. In a further arrangement (not shown) the two sets of output gears are mounted on different output shafts radially disposed with identical output gear pinions engaging gear wheel 42.

in Figure 2 all of the clutches 6OAID are shown in the neutral position. In normal operation the dual primary power paths are in either a power path mode or a pre-engaged mode. For instance, in the first path in power path mode the main clutch 62A is drivingly engaged and one of the dog clutches 60A18 is drivingly engaged with the input shaft 14A and one of the input gears 24,28,32,36. At the same time the second path is in pre-engage mode; clutch 62B is disengaged, power unit 80B sets the speed of input shaft I 4B to the requisite level to pre-engage an appropriate clutch 6OCID with an appropriate input gear 22,26,30,34. A powershift gear change is effected by disengaging main clutch 62A and simultaneously engaging main clutch 62B. While the principle of a dual path powershift gearbox is not new, the invention introduces new configurations facilitated by the compact dog clutches 60A/D used in combination with multi-function power units 8OAIB.

The synchronising function of power units 80A/B is described above. The further functions that each power unit intermittently provides, either in single or simultaneous operation, are; providing reversal means; providing additional starting and accelerating torque and; providing means to recover the vehicle braking energy. These functions are practicable owing to the prevalence of advanced vehicle and power train computer control systems. The modus operandi of the power units are described below by reference to diagrams Figures 3A/C showing various operating conditions.

Figure 3A shows gearbox 10 in a starting/accelerating mode. The second path is in power path mode; main clutch 62B is engaged as is dog clutch 60C thus providing first gear driving power path from the prime mover output shaft to the gearbox output shaft 16 through the input gear 22 and output gear 23. Additional torque is provided by each power unit 8OAIB.

Power unit 80B drives input shaft 14B, and continues to do so while accelerating torque is required and until a powershift is effected to the first power path; whereupon the computer system can in sequence: disengage the main clutch 62B to place the second path in pre-engaged mode; disengage clutch 60C; use power unit 80B as a variable speed motor to reduce the input shaft 1 4B rotational speed to that of gear 26; engage clutch 60C with input gear 26 (to be readied for a further power upshift) and; re-establish power unit 80B function to that of supplementary torque provision (if desired).

Figure 3A shows the first path in pre-engage mode. Main clutch 62A is disengaged and dog clutch 60A is drivingly pre-engaged with input shaft 1 4A and input gear 24. The power unit 80A is delivering supplementary torque to the input shaft 1 4A which is drivingly engaged through output gear 25 to the output shaft 16.

The use of power units 80A!B to provide supplementary torque extends the effective spread' of the speed ratios in the gearbox in a similar way to that provided by a torque converter in a conventional planetary gearbox -albeit in a much more efficient manner. The arrangement is particularly suited to a rear mounted transverse speed changing powershift gearbox fitted to a bus.

Figure 3B shows another novel feature of power units 80A/B. Either singly or together the power units are used as retarders, providing a reaction torque for braking, while being a component of a braking energy recovery system. Figure 38 shows the position with the vehicle in top' i.e. eighth gear, and being braked. The vehicle is driving gear 40 mounted on the output shaft 16; which is driving respectively output gears 37 and 35, and input gears 36 and 34 and input shafts 14A and 14B in the first and the second paths. Input shafts 14A and 14B are driving power units 80A and 80B respectively. Both of the main clutches 60A and 60B are disengaged.

The regenerative braking system, of which power units 8OAIB are a component, can be either conventional electric or hydraulic arrangements. In electrical configuration, the power units 80A/B are operating as electrical generators, supplying electrical energy to a battery to store the vehicle braking energy making it available for motive purpose through re-delivery to the power units 8OAIB operating as motors e.g. as in Figure 3A. In a hydraulic system the power units 8OAIB are used as hydraulic motors, the energy is stored in a hydraulic accumulator. Gearbox 10 fitted with an electrical system is well suited to a bus transport application, and with a hydraulic system to a utility vehicle.

In Figure 3C the speed changing gearbox 10 is shown in normal accelerating mode, with third gear engaged and the fourth gear pre-engaged. The first path is in power path mode. Main clutch 62A is engaged as is clutch 60A with input shaft I 4A and input gear 28; input gear 28 is driving output gear 29 and output shaft 16.

The power unit 80A is idling'. The second path is in pre-engage mode; main clutch 62B is disengaged; power unit 80B has synchronised the input shaft and input gear speeds enabling dog clutch 60D pre-engagement thus making ready for power upshift. The power unit 80B is idling'.

An arrangement of the eight speed, speed changing gearbox 10 fitted with close ratio steps is suited to a sports car. Here the main clutches 62A and 62B are subject to the lightest load, deriving from low inertia high speed engine and small inertia steps. The use of dry clutches, replacing conventional oil immersed wet types, provides enhanced vehicle performance.

Alternatively, the transverse gearbox layout in Figure 2 can be arranged with a speed increasing gear set with axis parallel to that of the prime mover, two identical output gears each separately driving a main clutch and a bevel gear pair.

The arrangement provides a narrow construction in transverse layout (not shown).

Yet alternately the transverse gearbox layout in Figure 2 can be arranged to incorporate a range change' gear unit (not shown) e.g. a unit providing either a 1:1 drive or a 3.5:1 speed reduction mounted at the position of output pinion 40 provides a 16 speed gearbox. Constructed e.g. as a powershift planetary unit then full powershift functionality is retained. Alternatively the range change can be constructed as an automatic gearchange with a power break. For example an up range shift can be carried out in the sequence: clutch 62A engaged, gear pair 36 and 37 in drive, clutch 62B disengaged, gear pair 22 and 23 in pre-engagement; disengage low range, disengage clutch 62A; power units 80A and 808 synchronise output shaft 16 rotational speed to high range; engage high range; engage clutch 62B to restore prime mover power path. The enhanced ratio spread is suitable for e.g. an articulated truck tractor unit. Alternatively a splitter' unit can be mounted at the input, positioned between the input bevel gear 38 and prime mover 90..

All of the transverse dual power path arrangements described above can readily be configured with in-line aspect. An eight speed gearbox comprises four co-planar gear sets with four output gears mounted on the central axis and two parallel axis input shafts (not shown).

Figure 4 shows a fork-actuated dog clutch and synchroniser ring of otherwise generally conventional form used e.g. as clutches 60 in Figure 1.

Adjacent gear wheels 30,32 are journaled to the shaft 12 on bearings 104 and each have a dog tooth ring 64A1B mounted adjacent one face by splined collars 65A1B.

An internally and externally spimed ring 68 mounts a clutching ring 66 to the shaft. The clutching ring 66 is axially slideable on the outer splines of the ring 68 by gear selector forks (not shown) engaged in a circumferential groove 67 so as to move its inner splined surface into and out of engagement with the dog teeth 64A/B. The gear wheels are thereby clutched to/disconnected from the shaft.

Synchroniser rings 70A/B are resiliently mounted to the clutch ring 68 by plate springs 1 O6AIB so as to frictionally engage a conical centre boss on the dog tooth rings 64A/B and match the speed of the shaft and gear wheel as the clutch ring 66 is moved towards the dog tooth rings 64A1B respectively. This allows smooth engagement of the clutch ring splines with the dog teeth. The fork actuated clutching arrangement could be used in gearbox 10, at 60A/D operating at the input shafts I 4AIB (albeit without the synchronising elements, this function being provided by the power units 8OAIB).

The dog clutch arrangement shown in Figure 4 takes up substantial length, particularly in a gearbox having a large number of ratios and hence a correspondingly large number of dog clutches. Figure 5 shows an alternative clutching arrangement for use in the first power path of gearbox 10, operating at the input shaft l4A. The input gears 24,28,32 and 36 are journaled to shaft 14A by bearings I 08A, II OA, 11 OB and 1 08B respectively. At each input gear centre is an aperture 72AfD that embodies face dogs or internal splines. Each input gear can be selectively engaged to the input shaft by the axial movement of one of two clutching rings 74AiB rotationally coupled to input shaft 14A e.g. by co-operating splined surfaces as shown. Each clutching ring 74A1B is shaped on the outer surface with either dogs or splines to engage and co-operate with the input gears.

Clutching rings 74A and 74B service input gears 24,28 and 32,36 respectively.

Whereas the description of Figure 5 thus far is of a conventional dog clutch, the means of actuating the clutching rings 74A1B is substantially different and eliminates the conventional fork actuators mounted between adjacent gears.

In Figure 5 the clutching rings 74A/B function as magnets, and the force required to move the clutching rings into engagement with the input gears is provided by passing an electric current through the windings 11 2A/B held in position by retaining plates 11 4A/B. For example, a magnetic field created by an electrical current in winding II 2A causes axial movement of clutching ring 74A.

A reversal of current in winding 11 2A causes axial movement in the opposite direction. Thereby under the control of a computer system the means to rotationally engage and disengage the input gears 24, 28 with and from respectively the input shaft 1 4A is provided. While the force available from such an arrangement is enhanced by the generally higher voltages used in modern vehicle electrical systems. the principle of making use of electro-magnetic force in the application is enabled by the external synchronisation means and a computer control system. Precise speed synchronisation of the rotating parts to be engaged and precise torque "back off' reduces the axial clutching force required to engage and disengage respectively. The construction of a clutch 74A1B as a ring engaging with an input gear at a central aperture provides minimal inertia. In Figure 5 the neutral position of ring clutches 74A/B can be assisted by mounting lightly loaded springs between the clutch ring end faces and the adjacent bearing face; 74A with 108A and 1 IOA and; 74B with 1 lOB and 108B (not shown). Alternately, the clutch 74A1B is constructed with integral winding and serviced via slip rings (not shown) Figure 6 shows another arrangement of clutch using permanent magnets.

The five gears 43-47 are the centrally disposed gears in a twin layshaft or twin output shaft gearbox, wherein the e.g. central spur gears are subject to purely rotational forces and thus no central bearing is required to support the central gear.

In Figure 6 the central gears 43-47 are selectively engaged with the shaft 15 by means of three clutching rings 76A/C rotationally engaged with the input shaft 15 and having external splines that engage selectively with co-operating splines mounted in the central apertures 72A/E of the central gears. The clutching rings 76A/C are ermanent magnets. The shaft 15 is of non-magnetic material e.g. Kevlar"nc1 is tubular. In the centre of the tube is mounted an actuating rod 17 upon which are mounted three permanent magnets 75 A/C wherein the magnetic fields cause the permanent magnets 76A1C respectively to move into co-planar alignment. The central gears and the rod/clutching rings assembly are arranged such that axial movement of the rod by one linear unit can e.g. disengage a central gear and by a further unit e.g. engage a central gear. In Figure 6 the clutching ring 76A is shown engaged with the central aperture of gear 43, the lowest, first, gear ratio. Axial movement of the rod leftwards by one unit causes the clutching rings 76A/C to move leftwards by one unit, thus disengaging clutching ring 72A from gear 43. A further movement of one unit to the left engages clutching ring 76C with central gear 44, the second gear. Further one unit moves to the left cause disengagement of 44; engagement of 45; disengagement of 45; engagement of 46; disengagement of 46 and engagement of 47 thereby moving into the highest, fifth gear.

Arrangements using the transient electro-magnetic forces described in Figure 5 can readily be configured for use in the gearbox type described in Figure 6 (not shown).

An alternative to the electro-magnetic clutch arrangement shown in Figure 5 is shown in the scrap section Figure 7. Magnetic force is replaced by hydraulic or pneumatic force. Clutching Pistons 78AfB are internally splined to co-operate with the splined input shaft 14A and externally splined to co-operate selectively with the input gears 24, 28 respectively central apertures. The outer diametral surface of pistons 78A1B rotate inside a cylinder 116 divided into two chambers each housing a clutching piston. The chamber formed between the stationary cylinder and clutching piston 78A,78B is served by hydraulic or pneumatic feed tube 118, 120 respectively. Selective application of pressure to the chambers causes the clutching pistons to move axially into engagement with the input gear, 78A with 24 and 78B with 28. Because the forces required to effect engagement are low, the leak path between the input shaft and the clutching pistons can be accommodated. Disengagement is effected by creating a vacuum in the chamber, alternately with light springs 1 22A1B in combination with precise torque back off, and yet alternately by connecting the members 78A and 78B and applying pressure to the opposite side.

In yet a further alternative, the force to move the selectively engageable clutch is derived from fluid dynamics. In scrap section Figure 8 the clutching ring 79 is internally splined to co-operate with the splined input shaft 14A and externally splined to co-operate selectively with the input gears 24 and central apertures. The outer diamet.ral surface of the clutching ring 79 is shaped to change the direction of a fluid path impinging upon it. The change of direction produces a force that is used to move the clutching ring into and out of engagement with the input gears 24 and 28. The force is derived from a fluid jet delivered from tubes 124 and 126. A jet from tube 124, 126 impinging a sloping surface on clutching ring 79 is vectored in forces both along the angle of the slope and at right angles to it, and the latter in turn produces a vectored axial force to move the clutching ring into engagement with input gear 28, 24 respectively and similarly disengaging.

Alternately, the fluid dynamic forces may be derived from the construction of the clutching ring acting as a pump or as a thrust nozzle (not shown).

The radial arrangement described above can be replicated using mechanical forces, the tubes 124 and 126 being replaced by radially disposed rods engaging the shaped surface of the clutching ring, the rods being moved radially to effect axial movement of the clutching ring (not shown).

The clutching arrangements described are most favourably used in conjunction with external synchronising means to substantially equalise the rotating speeds of the shaft and gear to be engaged prior to their engagement. This can be achieved by means of a power unit as described in gearbox 10, Figure 2.

Alternative external synchronising means are well known to those knowledgeable in the art and include, either singly or in various combinations, flywheel based arrangements, engine speed control and input shaft brakes. In some specialist applications the clutching arrangements may be used without synchronising. For instance, in a twelve speed gearbox for a racing car the speed steps are very small, and the engine inertia is very low. In this application a clutching arrangement as e.g. described in Figure 7 is effective without synchronisation.

In the context of gearbox 10 and the alternative regenerative braking systems described, then the various clutching arrangements described above have natural partners. For instance electrical regenerative braking is compatible with electro-magnetic clutching because of the high voltage availability.

Any of the aforementioned clutching arrangements including those in Figures 5-8 contain features that can be applied to any or all of them, and any combination can be applied to the various speed change section arrangements described, including Figures 1, 2 and 6.

Particularly advantageous is the use of regenerative electrical braking allied with eiectro-mechanical clutch engagement applied to a single layshaft, six speed automatic gearbox of the power break type and fitted to e.g. a town based commercial delivery van. Figure 9 shows the schematic arrangement, gearbox 11.

Prime mover 90 drives the input shaft 12 via speed increasing gears 39 and the main clutch 62. The six input gears are joumaled to the input shaft by bearings 104 and are selectively clutched to the input shaft 12 by means three clutching rings 74 internally splined to provide driving engagement with the input shaft 12; the clutching rings are externally splmed to co-engage with splines at the central apertures of the input gears; a clutching ring axial movement being effected by means of an electrical current applied to an appropriate one of three windings 112 supported by plates 114.

Under computer control a change of gear involves the following steps either simultaneous or in sequence; disengagement of an input gear from the input shaft; disengagement of the main clutch 62; activation of the variable speed electrical motor power unit 80 to set the rotational speed of the input shaft 12 to that of the input gear to be selectively engaged; throttle control of the unloaded engine to synchronise the output shaft of gear set 39 to that of the input shaft 14; application of an electric current in the appropriate winding 112 to effect the axial movement and engagement of the appropriate clutching ring to the input gear to be selectively engaged to the input shaftl4; re-engagement of the main clutch 62 and return of engine management responsibility to the driver.

Figure 9 shows the input gear 26 engaged with input shaft 12, delivering torque to the output transfer gear 42 via output gear 27, output shaft 16, and output pinion 40. The six input gears are in engagement with the six output gears; the six gear pairs providing six forward speed ratios in the speed changing gearbox. The output gears are all rigidly mounted to the output shaft 16; the output shaft 16 incorporates a pinion gear 40 being engaged with transfer gear wheel 42 in a high gear reduction; the whole output assembly being mounted between the bearings 102. Owing to the benefits of the high rotating speeds and the compact clutches the bearings set 102 span is smaller than any other known arrangement.

The power unit 80 functioning as a generator provides regenerative braking means as described for gearbox 10; and additionally provides the vehicle reversal means. The combination of features provides an extremely efficient power train incorporating a high efficiency speed changing automatic gearbox.

Claims

Claims:- 1. A powershift speed changing gearbox comprising a plurality of co- planar gear sets each comprising co-operating gears in engagement; individual co-planar gears being selectively drivmgly engageable by a members selected clutched engagement between one of the co-operating gears and a shaft received at the centre; the co-planar gear sets being arranged to provide two primary drive paths; each primary drive path being separately clutched to the prime mover; during vehicle normal forward driving each primary drive path being alternately in a power path mode and in a pre-engaged mode, each primary drive path being serviced by a power unit that during pre-engagement mode intermittently provides speed synchronising means to enable selective clutched engagement and at other times intermittently provides braking energy recovery means.
2. A speed changing gearbox as in claim 1 wherein the power units are intermittently used simultaneously to provide braking energy recovery means.
3. A speed changing gearbox as in claim I wherein a primary drive path in power path mode is intermittently provided with additional power provided by its power unit.
4. A speed changing gearbox as in claim 3 wherein simultaneously the primary drive path in pre-engaged mode is provided with power provided by its power unit.
5. A speed changing gearbox as in claim I wherein the power units may rotate in either direction thus providing a reversing means when either one or both are in drive engagement and the prime mover is disconnected.
6. A speed changing gearbox as in claims 1-5 wherein each primary drive path gear cluster comprises a plurality of input gears mounted on one or more bearings; each bearing being mounted on an input shaft; each input gear being selectively drivingly engageable with the input shaft; each input gear co-operating with an output gear rotationally fixed to the output shaft; the power units being engaged with the input shafts.
7. A speed changing gearbox as in claim 6 wherein the two input shafts are arranged co-axially.
8. A speed changing gearbox as in claim 6 wherein the two input shafts are arranged in parallel.
9. A speed changing gearbox as in claim 8 wherein the input shafts are radially disposed with differing centres around a single output shaft; each co-planar gear set comprising of two input gears engaging a common output gear.
10. A speed changing gearbox as in claims 7-9 wherein in power path mode the input shaft rotates at a higher speed than that of the prime mover.
11. A speed changing gearbox as in claims 7-9 incorporating a range change or splitter gearbox
12. A speed changing gearbox as in claim 11 wherein an automatic range gear change is effected with a power break; one or both of the power units being used to synchronise the range change.
13. A speed changing gearbox as in claims 7-12 wherein the output shaft(s) engage with the driven or driving axle via high reduction ratio gears.
14. A speed changing gearbox as in claim 6 wherein the power units function intermittently as a variable speed motor capable of precise speed variation.
iS. A speed changing gearbox as in claim 14 in which a power unit is intermittently used to set the input shaft speed while in pre-engagement path mode to enable pre-engagement of a members clutched engagement between an appropriate gear and the shaft received at the centre.
16. A speed changing gearbox as in claim 6 wherein the power units function intermittently as a power absorber and a component of a brake energy recovery system.
17, A speed changing gearbox as in claims 15-16 in which a power unit is intermittently an electrical motor and is intermittently an electrical generator.
18. A speed changing gearbox as in claim 16-17 wherein the braking energy is recovered and stored in an electrical battery.
19. A speed changing gearbox as in claims 15-16 in which a power unit is intermittently a hydraulic motor and is intermittently a hydraulic pump.
20. A speed changing gearbox as in claims 16 and 19 wherein the braking energy is recovered and stored in a nitrogen filled hydraulic accumulator.
21. A speed changing gearbox as in claim 6 wherein the power units are engaged via a intermediary speed reducer. ( 27
22. A speed changing gearbox as in any of the preceding claims wherein the clutching and power unit operation to provide full powershift and automatic gear changes and braking energy recovery are managed by a computer system.
23. A speed changing gearbox comprising a plurality of co-planar gear sets each comprising co-operating gears in engagement; individual co-planar gears being selectively drivingly engageable by a members selected clutched engagement between an aperture in one of the co-operating gears and a shaft received at the centre; the plurality of members being individually rotationally coupled to the shaft and axially moveable with respect to the shaft and gear and engageable with the gears aperture to provide the selective clutched engagement; the axial movement being effected by a force applied adjacently to a co-operating gear; the force being non-mechanical or radially directed. Is
24. A speed changing gearbox as defmed in claim 23 in which the members are coupled rotationally to the shaft by means of splines.
25. A speed changing gearbox as defmed in claim 24 in which the members are annular members provided with splines or dog teeth sections on their external surfaces to engage co-operating surfaces in the aperture of the selected gear.
26. A speed changing gearbox as defined in claim 23 in which the members axial movement is effected by direct magnetic or electro-magnetic force.
27. A speed changing gearbox as defmed in the claims 25-26 in which each member is a permanent magnet and the member is moved axially by means of magnetic force.
28. A speed changing gearbox as defmed in the claims 25-26 in which the members axial movement is effected by means of initiating an electrical current and the member is moved axially by means of an induced electromagnetic force.
29. A speed changing gearbox as defined in claim 23 in which the members axial movement is effected by means of a directly applied fluid hydro-dynamic or hydraulic or pneumatic force.
30. A speed changing gearbox as defined in claims 25 and 29 in which the member is a piston subject to pressure or vacuum at one or both ends.
31. A speed changing gearbox as defined in claims 25 and 29 in which the member is configured as a nozzle reacting to the transient fluid flow.
32. A speed changing gearbox as defined in claims 25 and 29 in which the members outer surface is shaped to exert axial force upon the initiation of fluid flow.
33. A speed changing gearbox as defmed in any of the claim 23 in which the member is moved axialiy by means of a radially directed mechanical force.
34. A speed changing gearbox as defmed in any of the claims 23-33 wherein a variable speed power unit is drivingly engaged with the shaft and intermittently provides shaft rotational speed near synchronisation simultaneously to a members selective clutched engagement with a co-operating gear.
35. A speed changing gearbox as defined in any of claims 23-3 3 in which the selective clutched engagement is synchronised by means of a flywheel.
36. A speed changing gearbox as in claim 34 wherein the power unit intermittently functions as a supplementary power to the prime mover; a vehicle reversal means and; a power absorption device being part of a braking regenerative system.
37. A speed changing gearbox as defmed in any of the claims 23-36 used in conjunction with a computer controlled power train management system to provide automatic or semi-automatic gearchanging.
38. A speed changing gearbox according to claim I constructed, arranged and adapted to operate substantially as hereinbefore described with reference to as illustrated by the accompanying drawings.
39. A speed changing gearbox according to claim 23 constructed, arranged and adapted to operate substantially as hereinbefore described with reference to as illustrated by the accompanying drawings.