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WO2008037946A1 - Boîte de vitesses de changement de vitesse à schéma de puissance multiple - Google Patents

Boîte de vitesses de changement de vitesse à schéma de puissance multiple Download PDF

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Publication number
WO2008037946A1
WO2008037946A1 PCT/GB2007/003154 GB2007003154W WO2008037946A1 WO 2008037946 A1 WO2008037946 A1 WO 2008037946A1 GB 2007003154 W GB2007003154 W GB 2007003154W WO 2008037946 A1 WO2008037946 A1 WO 2008037946A1
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WO
WIPO (PCT)
Prior art keywords
gear
gearbox
gears
input
output
Prior art date
Application number
PCT/GB2007/003154
Other languages
English (en)
Inventor
John Madge
Original Assignee
John Madge
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by John Madge filed Critical John Madge
Priority to GB0903880A priority Critical patent/GB2456076B/en
Publication of WO2008037946A1 publication Critical patent/WO2008037946A1/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/02—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion
    • F16H3/08—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts
    • F16H3/083—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts with radially acting and axially controlled clutching members, e.g. sliding keys
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/02—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion
    • F16H3/08—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts
    • F16H3/087—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears
    • F16H3/093—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears with two or more countershafts
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/02—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion
    • F16H3/08—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts
    • F16H3/087—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears
    • F16H3/093—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears with two or more countershafts
    • F16H3/095—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears with two or more countershafts with means for ensuring an even distribution of torque between the countershafts
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/003—Transmissions for multiple ratios characterised by the number of forward speeds
    • F16H2200/0047—Transmissions for multiple ratios characterised by the number of forward speeds the gear ratios comprising five forward speeds
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/02—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion
    • F16H3/08—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts
    • F16H3/087—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears
    • F16H3/093—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears with two or more countershafts
    • F16H3/097—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts characterised by the disposition of the gears with two or more countershafts the input and output shafts being aligned on the same axis

Definitions

  • 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.
  • 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 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.
  • 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.
  • 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.
  • 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.
  • the single output gear may be moveable transverse to the plane of the output shafts.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • a centrally disposed gear in a multi layshaft arrangement does not need any central bearing because of the balanced gear tooth loads.
  • 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.
  • 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.
  • 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.
  • FIG. 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
  • FIGS 13 and 14 illustrate a gearbox embodying the invention, used in conjunction with a continuously variable transmission (CVT).
  • CVT continuously variable transmission
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 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.
  • 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.
  • 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.
  • 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 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.
  • 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.
  • flange 56B is selectively engageable in the central apertures of gears 186 or 188, upon axial sliding of the tube 54B.
  • 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
  • the shaft 12 is stepped and hollow. Sets of axially extending, circumferentially distributed slots 60A 3 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.
  • 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.
  • 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.
  • the mechanism to selectively couple the required central drive gear to the driveshaft 12 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.
  • 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.
  • FIG. 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.
  • 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.
  • 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.
  • 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, 20F 3 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • the central driving shaft 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.
  • 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.
  • 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.
  • 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.
  • pulley set 82B, 84B is re-engaged in place of the other pulley sets.
  • 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.
  • the driving shaft rotational speed is increased by a controlled engine speed increase while a brake is used to reduce it.
  • 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.
  • CVT speed variator
  • 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.
  • 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.
  • the gear 22 is a split ring gear comprising torsionally resiliently interconnected parts 22 A and 22B providing torque compensation.
  • 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.
  • 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.
  • a single helical combining gear 22 can be used, which tilts about its axis to provide torque sharing compensation.
  • 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.
  • 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.
  • Gearboxes according to the present invention can be of simpler construction and more robust than existing units, particularly in the motorsport field.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Structure Of Transmissions (AREA)

Abstract

La présente invention concerne une boîte de vitesses de changement de vitesse à schéma de puissance multiple comportant une pluralité de trains de pignons coplanaires chacun comprenant un pignon d'entrée (180-188) s'engrenant avec deux pignons de sortie diamétralement opposés(160A/B-168A/B); les pignons d'entrée de chaque train étant montés sur un arbre d'entrée (12) et les pignons de sortie de chaque train étant montés sur des arbres différents d'une paire d'arbres de sortie (14A/B); des trains de pignons individuels coplanaires étant aptes à un engagement d'entraînement sélectif entre l'arbre d'entrée et les arbres de sortie; un chemin de couple fermé peut être formé entre deux arbres de sortie au moyen de pignons de sortie correspondants (20A/B)et un pignon combiné décalé (22). L'engagement d'entraînement entre des trains de pignons coplanaires et les arbres d'entrée et de sortie pour fournir un rapport de vitesses sélectionné peut s'effectuer au moyen de l'embrayage de chaque pignon d'entrée à l'arbre d'entrée via une ouverture centrale (52) apte à un engagement avec des dents de crabot (56A, 56B, 56C) portées par l'arbre (54). L'invention concerne également d'autres agencements d'embrayage et configurations de boîte de vitesses.
PCT/GB2007/003154 2006-09-29 2007-08-20 Boîte de vitesses de changement de vitesse à schéma de puissance multiple WO2008037946A1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
GB0903880A GB2456076B (en) 2006-09-29 2007-08-20 Plural power path speed changing gearbox

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
GB0619234.8 2006-09-29
GBGB0619234.8A GB0619234D0 (en) 2006-09-29 2006-09-29 Plural power path speed changing gearbox

Publications (1)

Publication Number Publication Date
WO2008037946A1 true WO2008037946A1 (fr) 2008-04-03

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GB (3) GB0619234D0 (fr)
WO (1) WO2008037946A1 (fr)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN110345245A (zh) * 2019-08-23 2019-10-18 河南科技大学 液力机械式多档多离合变速器

Families Citing this family (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2024218750A1 (fr) * 2023-04-19 2024-10-24 Eaton Intelligent Power Limited Transmission à double arbre concentrique à plusieurs vitesses

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3105395A (en) 1962-12-26 1963-10-01 Eaton Mfg Co Automotive device
GB1094811A (en) * 1965-01-06 1967-12-13 A E C Ltd Improvements in or relating to gear-boxes
US3885446A (en) 1972-04-29 1975-05-27 Eric Alexander Pengilly Gearing-torque division
GB1434928A (en) 1972-04-29 1976-05-12 Pengilly E A Power transmissions
GB2112883A (en) * 1982-01-07 1983-07-27 Brown Gear Ind Multi-ratio gearbox
EP0244263A2 (fr) 1986-05-02 1987-11-04 Eric Alexander Pengilly Transmission avec division de couple multiple
WO2003046409A1 (fr) 2001-11-26 2003-06-05 Worrall, Trevor, James Transmissions d'automobile
GB2424682A (en) * 2005-04-01 2006-10-04 John Gordon Madge Plural power path gearbox.

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB1509880A (en) * 1976-09-21 1978-05-04 Kubota Ltd Transmission for tractors
FR2862366B1 (fr) * 2003-11-18 2006-12-15 Peugeot Citroen Automobiles Sa Dispositif de commande d'un rapport d'une boite de vitesses.

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3105395A (en) 1962-12-26 1963-10-01 Eaton Mfg Co Automotive device
GB1094811A (en) * 1965-01-06 1967-12-13 A E C Ltd Improvements in or relating to gear-boxes
US3885446A (en) 1972-04-29 1975-05-27 Eric Alexander Pengilly Gearing-torque division
GB1434928A (en) 1972-04-29 1976-05-12 Pengilly E A Power transmissions
GB2112883A (en) * 1982-01-07 1983-07-27 Brown Gear Ind Multi-ratio gearbox
EP0244263A2 (fr) 1986-05-02 1987-11-04 Eric Alexander Pengilly Transmission avec division de couple multiple
WO2003046409A1 (fr) 2001-11-26 2003-06-05 Worrall, Trevor, James Transmissions d'automobile
GB2424682A (en) * 2005-04-01 2006-10-04 John Gordon Madge Plural power path gearbox.

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN110345245A (zh) * 2019-08-23 2019-10-18 河南科技大学 液力机械式多档多离合变速器

Also Published As

Publication number Publication date
GB0713017D0 (en) 2007-08-15
GB2442294A (en) 2008-04-02
GB0903880D0 (en) 2009-04-22
GB2456076A (en) 2009-07-08
GB2456076B (en) 2011-03-30
GB0619234D0 (en) 2006-11-08

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