WO2009019430A1 - Boîte de changement de vitesse, intégrée, à plusieurs puissances - Google Patents

Boîte de changement de vitesse, intégrée, à plusieurs puissances Download PDF

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Publication number
WO2009019430A1
WO2009019430A1 PCT/GB2008/002514 GB2008002514W WO2009019430A1 WO 2009019430 A1 WO2009019430 A1 WO 2009019430A1 GB 2008002514 W GB2008002514 W GB 2008002514W WO 2009019430 A1 WO2009019430 A1 WO 2009019430A1
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WO
WIPO (PCT)
Prior art keywords
gear
gearbox
input
gears
shaft
Prior art date
Application number
PCT/GB2008/002514
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
Publication of WO2009019430A1 publication Critical patent/WO2009019430A1/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
    • F16H3/087—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
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B66—HOISTING; LIFTING; HAULING
    • B66D—CAPSTANS; WINCHES; TACKLES, e.g. PULLEY BLOCKS; HOISTS
    • B66D1/00—Rope, cable, or chain winding mechanisms; Capstans
    • B66D1/02—Driving gear
    • B66D1/14—Power transmissions between power sources and drums or barrels
    • B66D1/24—Power transmissions between power sources and drums or barrels for varying speed or reversing direction of rotation of drums or barrels, i.e. variable ratio or reversing gearing
    • 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
    • F16H1/00—Toothed gearings for conveying rotary motion
    • F16H1/02—Toothed gearings for conveying rotary motion without gears having orbital motion
    • F16H1/20—Toothed gearings for conveying rotary motion without gears having orbital motion involving more than two intermeshing members
    • F16H1/206—Toothed gearings for conveying rotary motion without gears having orbital motion involving more than two intermeshing members characterised by the driving or driven member being composed of two or more gear wheels
    • 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
    • F16H63/00—Control outputs from the control unit to change-speed- or reversing-gearings for conveying rotary motion or to other devices than the final output mechanism
    • F16H63/02—Final output mechanisms therefor; Actuating means for the final output mechanisms
    • F16H63/30—Constructional features of the final output mechanisms
    • F16H2063/3093—Final output elements, i.e. the final elements to establish gear ratio, e.g. dog clutches or other means establishing coupling to shaft

Definitions

  • This invention concerns a plural power path speed changing gearbox wherein the axes of the input and output shafts are offset and substantially parallel. It has widespread automotive and industrial application especially in a dual drive arrangement and in a two speed ratio gearbox being integral with a driven mechanism and providing a 'corner HP' solution i.e. a drive delivering either high torque at low speed or low torque at high speed.
  • Known arrangements to provide corner HP include automotive 'range change' gearboxes, and the various same configurations of gearing are used in industrial machinery e.g. winches.
  • Prior art forms include planetary gearing, conventional single layshaft and plural power layshaft gearing. Speed ratios are changed either manually e.g. in a multi-speed heavy truck gearbox, or automatically with a power break in the drive. There are clear advantages in using plural power paths, the size and the weight being reduced.
  • the prior art plural power gearbox arrangements are characterised by the input shaft being co-axial with the output shaft, and by limitations in gear reduction ratios of practically proportioned gear pairs in engagement.
  • An inventive step in this invention is the arrangement of plural power paths in substantially two planes of gears.
  • the engine and transmission are preferably positioned at the rear, and lack of axial space in a transversely mounted engine arrangement has encouraged the use e.g. of a spatially effective but poor efficiency torque converter and planetary gear combination.
  • the present invention seeks to eliminate the above design constraints by the use of non coaxial plural power path gearing, and accordingly provides a speed changing gearbox comprising one or more sets of gear input sections; each section comprising two co-planar sets of gears; each set of gears comprising an input gear co-operating with a pair of intermediary gears; each gear of an intermediary pair being diametrically oppositely engaged with the input gear; the intermediary gear pairs in each co-planar set being arranged such that one gear of an intermediary gear pair is rotationally connected on a common central axis to one gear of the other intermediary gear pair, and the other one gear of an intermediary gear pair is rotationally connected on a common central axis to the other one gear of the other intermediary gear pair; each one of the two co-planar input gears being selectively drivingly engageable with an input shaft at the centre; a drivingly engaged input gear transmitting substantially equal force to each gear of the intermediary gear pair in engagement; one of the two co-planar intermediary gear pairs being in engagement with
  • the output gear can be a ring gear of a differential unit, thereby positioning the gearbox in the driveline far more advantageously than a conventional unit positioned upstream of a speed reducing differential.
  • the range change gearbox has provide a 'high range' gear reduction equivalent to e.g. a bevel gear pair embedded in a conventional differential, and a useful 'range shift' ratio.
  • the range change gearbox as in the present invention can embody co-planar reduction ratios of e.g. 6:1 and 2:1, thereby meeting the practicability requirement.
  • the present invention can replace a conventional range change unit and the conventional bevel gears.
  • a further advantage of plural power paths in this application is the ensuing reduction in pitch circle diameter of the gears, especially the input gears. Noise is closely associated with pitch line speeds. In an overrunning range change unit of the kind in the present invention, small pitch circles are important.
  • the size of the output ring gear is determined by the axle torque and the space required for the differential gear elements.
  • the tooth loads are substantially halved when compared with e.g. the ring gear in a conventional transversely mounted automotive engine and gearbox.
  • the ring gear can be smaller, hi situations where the differential gear elements dictate the ring gear diameter the ring gear facewidth can be reduced.
  • the benefits of a smaller diameter can be gained by matching the present invention with a compact differential e.g. of the spur gear type.
  • the two co-planar gear sets each have two torque paths that are closed at the output gear.
  • some form of compensatory movement is provided in one or more of the rotating members.
  • a variety of means may be employed, as known to those skilled in the art. They include compensation at the input gears e.g. for input gears of the spur type by providing compensatory movement in a radial direction; e.g. for input gears of a helical type the movement may be axial translation and/or rotation about an axis normal to the plane of the input shaft, similar to the arrangement described in US3885446. Alternatively the torque sharing may be effected using double helical or "herringbone" gears. Yet alternately compensatory movement may be provided by means of a torsionally flexible input shaft.
  • the above compensatory movements can be applied to the input gears individually or to the input shaft upon which both gears are mounted, or in some cases to the output gear.
  • Some of the aforementioned mechanisms may be constructed in combination.
  • the torque sharing means dictates the design of the input shaft.
  • the shaft is a splined extension of a gearbox output shaft journalled in bearings.
  • the shaft extension being subject only to torsional loads and suitably hardened can be diametrically sized to e.g. accommodate a coupling sleeve (for torque sharing purposes) without detriment to the overall proportions of the input gears.
  • the extension input shaft is of very small diameter and connected to an outer splined tube, the assembly providing torsional flexibility; and yet alternatively a splined shaft is constructed with coupling teeth and is mounted to provide the requisite degrees of freedom.
  • a change in speed ratio is effected by selective clutched engagement between one of the input gears and a shaft at the centre.
  • the clutching mechanism is a dog clutch rotationally splined to the input shaft and movable axially to engage cooperating dogs or coupling teeth on the face or at the centre of each input gear.
  • the clutch is a conventional shift fork arrangement operated manually without speed synchronisation.
  • This arrangement is suitable for stationary operation, e.g. a winch or the selection of a deep range in a stationary vehicle. More often the clutching arrangement will involve some speed synchronisation during a gear change.
  • the preferred embodiments including the innovative use of a multi-function motor used in conjunction with a computer controlled power train management system to provide automatic gearchanging, are further described below. Syncronising rotational speeds by external means prior to the clutch engagement of an input gear not journalled at the centre enables the introduction of various novel clutching arrangements. These allow the construction of substantially adjacent co-planar gear sets, the space saving being particularly important in a vehicle driveline.
  • the present invention provides an extremely compact two speed gearbox.
  • a range change automotive application it is particularly suited to transverse mounting e.g. in a front wheel drive passenger car wherein the conventional output transfer gear can be smaller, and a conventional e.g. five speed gearbox can be co-axially connected thereby enabling unconventional and highly efficient power train modes of operation to be considered.
  • the conventional cross drive differential mounted bevel gear pair can be eliminated or made considerably smaller.
  • the combination of a high/low range and a close ratio gearbox can provide operational advantage in e.g. aircraft haulage, tractors and other agricultural equipment, and military vehicles.
  • the output gear can be the final gear wheel in a winch drive train, e.g. mounted on the winch drum.
  • the plural drive path means that the pitch circle diameter and or the facewidth of the gear is reduced.
  • the high reduction ratio feature of the present invention beneficially reduces the size and gear ratio of an upstream gearbox.
  • the two speed feature means that the winch productivity is improved when operating under conditions of light loading e.g. pay out of rope.
  • the present invention is applicable in many of these types of 'corner HP' problems arising in the alternating conditions of high and low loading of process machinery, or in facilitating start-up of such machinery e.g. conveyor drives.
  • a compact two speed gearbox with scope for automatic operation can improve the operating performance of any device subject to a variable load-speed regime, e.g. a wind turbine, or a boat propeller.
  • Fig 1 is a diagram of the gearbox with one input section
  • Fig 2 is a side elevation of Fig 1 ;
  • Fig 3 is a drawing showing an input assembly incorporating prior art torque sharing means
  • Fig 4 is a further such arrangement
  • Fig 5 is a scrap section showing a conventional shift fork actuated dog clutch
  • Fig 6 is the arrangement as in Fig 4 showing alternative clutching means
  • Fig 7 is a drawing showing spur gear torque sharing and a further clutching means
  • Fig 8 is a scrap section showing diametrically compact input gears
  • Fig 9 is a drawing showing an input assembly incorporating torsional flexibility torque sharing means
  • Fig 10 is a schematic showing the invention embodied in a computer managed automotive power train.
  • Fig 11 is a drawing showing two input sections arranged around an output gear.
  • Figure 1 is a front elevation diagram showing the gear layout in the present invention with one input section.
  • Input gears 22 and 24 with input shaft 20 at the centre are in co- operating engagement with two diametrically opposed intermediary gears 30A, 30B and
  • the intermediary gears 30A and 30B are in co-operating engagement with a gear 40 with an axis offset and substantially parallel to the of the input shaft.
  • the output gear may form part of a vehicle differential assembly (not shown) or e.g. a winch drum assembly (not shown).
  • FIG 2 is a side elevation of Figure 1.
  • the co-planar input gear sets have the diametrically opposed A and B intermediary gears rotationally fixed to a common intermediary shaft 21 A and 21 B respectively, each shaft being journalled in bearings 60 A/D.
  • the input gears 22 and 24 are selectively drivingly engageable to the input shaft 20 at the centre, to provide two gear ratios in the gearbox. Torque sharing between the input shaft and the diametrically opposing intermediary gears in each of the gear ratios is provided by either extremely accurate manufacture and assembly, or provision of compensatory movement at the output gear or at each input gear or at the input shaft, or a combination of these.
  • Figure 3 shows two prior art arrangements for torque sharing means at the input shaft and input gears .
  • the lowest gear ratio uses input gear 22, and this being both the lowest driving rotational speed and the smallest diameter input gear then in many applications e.g. winch drives or an automotive range change for occasional off-road at low vehicle speed or infrequent use in a maximum loaded vehicle operating on steep inclines, spur gearing with simple radial compensation so as to be self-centering may be adequate in the low range.
  • Spur gear 22 is self centering between the two intermediary gears 30A and 30B (not shown), and is clutched to the input shaft 2OB by the axial movement leftwards of the shift fork operated member 50 wherein co-operating dog teeth engage in the aperture 52A.
  • the member 50 is splined to the input shaft 2OB that is coupled to input shaft 2OA and 2OC at each end at the apertures 54A/B, therein providing radial compensation movement for input gear 22 in driving mode, and thereby substantially equalising the force applied to the intermediary gears 30A and 30B.
  • FIG. 3 shows a prior art arrangement for torque sharing a single helical gear 24 between the intermediary gears 32A/B (not shown).
  • the compensating movement is provide by mounting gear gear 24 to shaft 2OD via a torque transmitting sleeve 25.
  • Gear 24 is formed as a ring gear having internal splines complementing external splines on sleeve 25, internal splines on sleeve 25 in turn engage complementary splines on the member 26 rotationally fixed to the shaft 2OD.
  • the co-operating splines are short and slightly crowned to allow pivoting of the axes of shaft 2OD, sleeve 25 and gear 24 relative to each other.
  • the lead and hand of the co-operating splines are equal to the lead and hand of the helical gear teeth, thereby ensuring that the axial forces on the sleeve 25 and gear 24 arising from torque transmission cancel each other out.
  • the gear 24 has the degree of movement freedom to ensure that the forces transmitted to intermediary gears 32A/B are substantially equal and the torque is shared between them.
  • the input shaft 2OD is journalled in bearings 66A/B and is provide with co-operating dogs at 52B for selective engagement with member 50.
  • Input shaft 2OA as shown is the upstream drive output shaft.
  • the input gear 22 is also a single helical gear provided with compensator ⁇ ' movement of the kind described above, the plane of the helical gear teeth is offset from that of the two helical splines, to accommodate a smaller gear teeth pitch circle diameter (not shown).
  • Figure 4 shows an alternative prior art arrangement for torque sharing using double helical gears.
  • Double helical input gears 22 and 24 are journalled in a common input shaft 20 in needle roller bearings 68 and 69 respectively.
  • Intermediary gears 30A/B and 32A/B are mounted to co-operate with one of the helix angles on gear 22 and 24 respectively.
  • the intermediary gears 30A and 30B co-operate with the output gear formed as a double helical (herringbone) gear (not shown), therein providing the closed torque path.
  • the bearings 68 and 69 allow small axial movement of gears 22 and 24 respectively thereby ensuring that the forces transmitted by the driving input gear 22 or 24 to 30A, 30B and 32A, 32B are substantially equal.
  • the input shaft 20 is an extension to the output shaft of the upstream drive and the input gears 22 and 24 are individually rotationally connected to it by co-operating coupling teeth on the member 50 at apertures 52A and 52B respectively, the member being rotationally coupled to the input shaft.
  • This coupling arrangement is compatible with the need for the axial degree of freedom described above.
  • each helix of the output gear engages one of the intermediary gear pair 30A and 30B i.e. there is no plural use of individual gear teeth and the arrangement does not gain the compactness advantage of those described above.
  • the arrangement provides quiet gearing and a balanced axial force at the output gear, enabling simpler differential construction.
  • Figure 5 shows a dog clutch and synchroniser ring of otherwise generally conventional form.
  • the gear 22 is journalled to the shaft 20 on bearings 66 and has a dog tooth ring 88 mounted adjacent one face by a splined collar 84.
  • An internally and externally splined ring 80 mounts a clutching ring 86 to the shaft.
  • the clutching ring 86 is axially slideable on the outer splines of the ring 80 by gear selector forks (not shown) engaged in a circumferential groove 92 so as to move its inner splined surface into and out of engagement with the dog teeth 88.
  • the gear is thereby clutched to/disconnected from the shaft.
  • a synchroniser ring 90 is resiliency mounted to the clutch ring 86 by a plate spring 82 so as to frictionally engage a conical centre boss on the dog tooth ring 88 and match the speed of the shaft and gear as the clutch ring 86 is moved towards the dog tooth ring 88. This allows smooth engagement of the clutch ring splines with the dog teeth.
  • a further gear 24, dog tooth ring and synchroniser ring (not shown) may be mounted on the shaft to the right of the splined ring 80 for co-operation with the same clutch ring 86.
  • the clutch is operated manually without speed synchronisation, as in stationary operation, e.g. a winch or the selection of a deep range in a stationary vehicle.
  • Advanced computer control systems are now commonly applied in e.g. vehicle power train control systems to optimise engine and driveline operation for high efficiency and to provide automatic gear changing.
  • these computer control systems are used in combination with synchronising means applied to a simple clutch operating between a selected input gear and a shaft at the centre, as further described below.
  • FIG 6 shows one example of a preferred clutching arrangement when external synchronising means are used.
  • the arrangement of gears is as in Figure 4, however the conventional shift-fork member 50 is replaced by an electro-magnetic clutch system comprising electrical windings 70 and a clutching ring 72.
  • Application of an electrical current in the windings will induce an axial force in the clutching ring, and reversal of the current will induce a force in the opposite direction.
  • the current is actuated and the movement effected simultaneously at substantial synchronisation of the rotational speed of input gear 22 or 24 with the shaft 20, achieved by external means.
  • the loads required to engage the external coupling teeth on clutching ring 72 with co-operating ones on input gear 22 or 24 are relatively small.
  • the axial space between the co-planar input gear sets is substantially reduced.
  • Other simple means to produce sufficient force to effect axial movement include arrangements using hydraulic or fluid-dynamic or radially applied mechanical forces (not shown).
  • Figure 7 shows a further clutching arrangement applicable to the present invention, also reliant upon the externally provided means for synchronising speeds prior to engagement as described above.
  • the input gear 'floats' without any permanently connected member at the centre, especially the bearing required in a conventional single power path arrangement.
  • Such an input gear can be simply clutched using coupling teeth at the centre, and the clutching actuating means can readily be mounted externally to the axial space between the two co-planar gear sets, thereby further reducing the gearbox size.
  • the input gears 22 and 24 are self- centering spur gears between intermediary gears 30A, 30B and 32A,32B respectively (not shown).
  • Shaft 20 is splined and carries a clutching actuator 73 with coupling teeth at one end alternately in engagement with co-operating teeth at the centre of the input gears 22 and 24; the clutching ring is moved axially by a force applied at the opposite end. With no actuating mechanisms positioned between the input gears the face of the gears are substantially in the same plane, as shown.
  • the co-planar gear sets are substantially adjacent to each other, and the bearing span across the intermediary gears and the gearbox width is further reduced.
  • the pitch line speed of the low ratio input gear is especially important to keep the pitch line speed of the low ratio input gear as low as possible in the unloaded 'over-run' condition.
  • the use of plural power paths can significantly reduce the size of the input gear for a given torque carrying capacity, thereby going some way towards acceptable standards.
  • the geometry of the gearbox in the present invention used as a two speed range change unit may be dictated by the size of the differential unit and therefore the output gear surrounding it. The constraint may have adverse effect on the pitch line speeds of the low ratio input gear, in which case in a preferred arrangement the present invention is used in combination with a differential unit of the spur gear type, diametrically more compact than conventional bevel gears (not shown).
  • Figure 8 shows another arrangement of gears designed to minimise the pitch line speed of low range input gear 22.
  • the double helical gear teeth are formed integrally with the shaft, the shaft being journalled in bearings 65, the assembly being mounted at the centre of the high range input gear 24.
  • the high range input gear is journalled in bearings 67 grounded to the housing.
  • the input gears 22 and 24 are alternately drivingly engaged to the input shaft 20 by means of co-operating coupling teeth at the apertures 52 A and 52B respectively by means of the moveable clutching ring 72.
  • Figure 9 shows a further arrangement for single helical gears wherein the torque sharing means includes torsional flexibility in the driving shafts.
  • Input shaft 2OA is an output shaft extension of the upstream drive and is hollow with coupling splines on its inner circumference, the splines co-operate with and drivingly engage splines at one end of shaft 2OB.
  • the identical splines at the other end of shaft 2OB co-operate and drivingly engage with the annular ring 27 upon which are mounted single helical input gears 22 and 24 journalled in bearings 68 and 69 respectively.
  • the shaft 2OB is characterised by the shaft section between the splines at each end being torsionally flexible.
  • the input gears are restrained axially to react the axial thrust arising from the helix angle of the teeth (not shown), and during drive mode the torsional flexibility in shaft 2OB ensures that the forces applied to the intermediary gears are substantially equal.
  • the clutching ring 72 is splined to the annular ring 27.
  • the clutching ring is connected to a clutching actuator 73 constructed as an inner sleeve by e.g. three pegs operating in slots in the annular ring (one shown).
  • the input gears are drivingly engaged by axial movement of the clutching actuator, the arrangement once more providing an extremely compact gear cluster.
  • the external coupling splines at the ends of shaft 2OB are helical splines.
  • the annular ring 27 engages the input gears by means of helical splines and the annular ring acts in the manner of clutching sleeve 25 of Figure 3 to ensure torque sharing (not shown).
  • Figure 10 shows a single input section arrangement of the present invention embodied in a vehicle as a range change gearbox wherein the gear shift operations are carried out in combination with a power train control computer system.
  • the prime mover 100 drives a four speed gearbox through a main clutch 160, first gear being provided by engagement of gear 150 to shaft 20 by means of clutch 156, and fourth gear being provided by the engagement of shaft 20 to the input gear 152 by means of clutch 154.
  • the range change gearbox of the present invention is co-axial with this four speed gearbox and the output gear 40 forms a ring gear driving the differential unit 1 10, the differential gears driving the wheel axles 120 and 130.
  • the input gear 22 is drivingly engaged with shaft 20 by means of clutch 50 equal torque being transmitted to intermediate gears 3OA and 30B (not shown) each engaging the output gear with axis 120, 130 parallel and offset from that of the shaft 20.
  • the driveline can be constructed using conventional shift fork clutches with integral synchronising means at 154, 156 and 50. Operated manually for a range upshift the driver will in sequence: disengage the main clutch 160, disengage clutch 154, disengage clutch 50, engage clutch 50 to input gear 24 (high range), engage clutch 156 with gear 150 (upstream gearbox first gear) and engage 160.
  • This manual operation can readily be replicated using a computer system to provide an eight speed automatic gearbox with power breaks.
  • the motor 150 is a flywheel and pulley system as described in GB 2424682A.
  • the motor is a variable speed hydraulic or electrical motor being used to control the speed of shaft 20.
  • Advanced vehicle power trains are increasingly featuring e.g. regenerative braking and/or dual power paths incorporating an electrical or hydraulic motor/generator, and in a preferred arrangement the device can readily be used to provide synchronising means in the present invention, and additionally also provide vehicle reversing means.
  • the device is used to accelerate or decelerate the shaft 20, the input gear and shaft speeds are continuously measured and a computer control system activates the axial movement of the clutch to execute engagement at substantially synchronous rotational speeds of the appropriate input gear and shaft.
  • the arrangement is ideal for a bus application wherein a regenerative braking system is particularly advantageous .
  • the clutch 50 is a multiple plate clutch operating at each input gear (not shown).
  • the arrangement can be used in series with a torque converter (to provide usable torque spread for a vehicle application) or with a fluid coupling (to provide fully variable speed for e.g. a wind turbine application).
  • the arrangement can be matched with a conventional e.g. upstream three speed powershift gearbox.
  • the combination of the present invention and advanced computer control systems and dual power arrangements provides the power train designer with scope for significant improvement in vehicle efficiency over a complete duty cycle. For example, the availability of more gear ratios provides opportunity for extending vehicle operation in higher gears and the engine operation at lower speeds and higher torque (improved efficiency). Alternatively, loaded and unloaded vehicles can be operated in efficient low or high range respectively, with ratio overlap between them.
  • Figure 1 1 shows a speed changing gearbox as in Figure 1 with another input section, to provide four gear ratios.
  • the second input section comprising input gears 180, 182 are in engagement with intermediary gears 184 A, 184B and 186A,186B respectively.
  • the intermediary gears 184A and 184B are in engagement with the output gear 40, in common with gears 30A and 30B.
  • the gear ratios of 3.5 and 4.5:1 are added to the original 2 and 6:1, providing a dual, plural power path four speed gearbox.
  • the gearbox can incorporate in any combination any of the input gear configuration, torque sharing means, clutching arrangements, synchronising methods including the multi-function motor, and upstream driver configurations described above.
  • one input section or plural power path transfer gearbox or single gear in engagement with the output gear is drivingly engaged with an electrical drive for lower vehicle speeds; and an input section of the same output gear is drivingly engaged with an engine and torque converter for high speed cruise and 'burst' acceleration (not shown).
  • this gearbox can be installed in a conventional inline layout driving a bevel differential, the preferred arrangement is a transverse mounting with the output gear being integral with the differential.
  • gearbox as described in the present invention is in a direct wheel or sprocket drive wherein the gearbox is mounted upon the driving shaft of one or more wheels of a vehicle; in such a drive of e.g. two opposing wheels the arrangement also provides a steering function (not shown).
  • the arrangement is applicable for an agricultural tractor, or off- road vehicles e.g. a 'quad' bike.

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  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
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Abstract

L'invention concerne une boîte de changement de vitesse à plusieurs puissances, qui comprend une ou plusieurs sections d'entrée comprenant chacune deux ensembles coplanaires d'engrenages comprenant des engrenages (22 et 24) engrenant chacun respectivement deux engrenages intermédiaires diamétralement opposés (30A, 30B et 32A, 32B) ; les engrenages d'entrée de chaque ensemble étant montés sur un arbre d'entrée (20) et les engrenages intermédiaires (30A, 32A et 30B, 32B) étant respectivement fixés à rotation à des arbres (21A et 21B) ; les engrenages d'entrée étant aptes à s'engager à entraînement de façon sélective à l'arbre d'entrée ; les engrenages intermédiaires (30A et 30B) étant en engagement avec un engrenage de sortie (40) avec un axe décalé et sensiblement parallèle à celui de l'arbre d'entrée ; l'engrenage de sortie formant un trajet de couple fermé. L'engrenage de sortie peut former une couronne intégrée dans une unité de différentiel d'une chaîne cinématique de véhicule. L'engagement sélectif des engrenages d'entrée peut se faire au moyen d'un embrayage à griffes (50) classique, actionné par une fourche de changement, avec une synchronisation de vitesse de rotation aidée par des moyens externes.
PCT/GB2008/002514 2007-08-09 2008-07-24 Boîte de changement de vitesse, intégrée, à plusieurs puissances WO2009019430A1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
GB0715566.6 2007-08-09
GB0715566.6A GB2451681B (en) 2007-08-09 2007-08-09 An Integrated plural power speed changing gear box

Publications (1)

Publication Number Publication Date
WO2009019430A1 true WO2009019430A1 (fr) 2009-02-12

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CN107117547A (zh) * 2017-05-25 2017-09-01 杭州航验环境技术有限公司 一种可换挡的卷扬机构
CN108953497A (zh) * 2018-09-20 2018-12-07 江苏泰隆减速机股份有限公司 一种用于lt型重型水箱拉丝机上的专用减速机
CN117145939A (zh) * 2023-11-01 2023-12-01 无锡车联天下信息技术有限公司 一种气动离合转向桥轮边减速器

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DE2819293A1 (de) * 1978-05-02 1979-11-08 Zahnradfabrik Friedrichshafen Zahnraederwechselgetriebe mit mehreren vorgelegewellen
SU1316858A1 (ru) * 1985-04-03 1987-06-15 Предприятие П/Я Г-4695 Коробка перемены передач транспортного средства
EP0636813A1 (fr) * 1993-07-30 1995-02-01 Maag Getriebe Ag Transmission à engrenages avec des arbres parallèles
WO2005108822A1 (fr) * 2004-04-28 2005-11-17 Zf Friedrichshafen Ag Boite de vitesses a arbre de renvoi
GB2424682A (en) * 2005-04-01 2006-10-04 John Gordon Madge Plural power path gearbox.

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FR2787162B1 (fr) * 1998-12-15 2001-01-19 Renault Boite de vitesses compacte a deux arbres secondaires
DE102005012226B3 (de) * 2005-03-15 2006-06-08 Linsinger-Maschinenbau Ges.M.B.H. Zahnradgetriebe

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1358615A (en) * 1919-05-10 1920-11-09 Bbc Brown Boveri & Cie Toothed-wheel gearing
DE2819293A1 (de) * 1978-05-02 1979-11-08 Zahnradfabrik Friedrichshafen Zahnraederwechselgetriebe mit mehreren vorgelegewellen
SU1316858A1 (ru) * 1985-04-03 1987-06-15 Предприятие П/Я Г-4695 Коробка перемены передач транспортного средства
EP0636813A1 (fr) * 1993-07-30 1995-02-01 Maag Getriebe Ag Transmission à engrenages avec des arbres parallèles
WO2005108822A1 (fr) * 2004-04-28 2005-11-17 Zf Friedrichshafen Ag Boite de vitesses a arbre de renvoi
GB2424682A (en) * 2005-04-01 2006-10-04 John Gordon Madge Plural power path gearbox.

Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN107117547A (zh) * 2017-05-25 2017-09-01 杭州航验环境技术有限公司 一种可换挡的卷扬机构
CN107117547B (zh) * 2017-05-25 2024-06-07 杭州航验环境技术有限公司 一种可换挡的卷扬机构
CN108953497A (zh) * 2018-09-20 2018-12-07 江苏泰隆减速机股份有限公司 一种用于lt型重型水箱拉丝机上的专用减速机
CN117145939A (zh) * 2023-11-01 2023-12-01 无锡车联天下信息技术有限公司 一种气动离合转向桥轮边减速器
CN117145939B (zh) * 2023-11-01 2024-01-26 无锡车联天下信息技术有限公司 一种气动离合转向桥轮边减速器

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GB0715566D0 (en) 2007-09-19
GB2451681B (en) 2012-06-20

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