GB2181195A - Variable-ratio power transmission mechanism - Google Patents

Abstract

A variable-ratio power transmission mechanism comprises, an input shaft (10), an output shaft (12) and an epicyclic gear unit (16), the epicyclic gear unit comprising a sun gear (14), a planet carrier (18) and a ring gear (22), all rotatable about a common axis, the planet carrier being connected to, and rotatably driving, the output shaft, and having a plurality of rotatable planetary pinions (20) in meshing engagement with the sun gear and the ring gear, the sun gear being directly driven by the input shaft and the ring gear being variably driven by other drive means (24) or vice versa, the other drive means being indirectly driven by the input drive means, the arrangement being such that the rotational speed of the output shaft is proportional to the relative speed of the ring gear to the sun gear. The other drive means is preferably a swash plate hydraulic pump/motor unit,the pump being of variable displacement and the motor of fixed displacement. The fixed swash plate (58) of the motor is preferably fixed to the ring gear (22), with the input shaft (10) driving the drums (26,28) of the pump and motor. First and second valve plates (36, 38) having C-shaped distribution grooves are located between the pump and the motor. <IMAGE>

Description

SPECIFICATION Variable-ratio power transmission mechanism This invention relates to a variable-ratio powertransmission mechanism, for example for use as a vehicle drive.

Hydrostatic (positive-displacement) transmissions have been applied to various industrial uses, and as vehicle drives. For example the Janneytransmission comprises a variable-ratio axial-piston hydraulic pump and motorunitin acoaxial arrangementwith aligned input and output shafts. The pump unit includes a variable-angle swash plate, and the motor unit includes a fixed-angle swash plate, and the pump and motor units have respective cylider-carrying rotary drums (cylinder blocks) separated by a fixed valve plate that is provided with fluid ports for establishing a hydraulic drive in the pump and motor unit.

Improved efficiency is obtainable by the use of a hydrostatic transmission in a shunt (split-torque) configuration with a mechanical torque path, advan tageously with recombination oftorque from the hydraulic and mechanical torque paths by the use of an epicyclic gear unit. An example of such a geared shunt transmission is the Sundstrand dual-mode transmission when operating in its shunt mode. The Sundstrand transmission utilises rotationally fixed drumsforthe axial pistons, and an engine-driven variable-angleswash plate in the pump unit.

By the presentinventionthere is provided avari- able-ratio powertransmission mechanism comprising an input drive means, an input shaft rotatably drivableabove its longitudinal axisbytheinputdrive means, an output shaft rotatably drivable about its longitudinal axis, and an epicyclic gear unit, the epicyclic gear unitcomprising a sun gear, a planet carrier and a ring gear all rotatable about a common axis, the planet carrier being connected to, and rotatablydriving,the output shaft, and having a plurality of rotatable pinions in meshing engagement with the sun gear and the ring gear, the sun gear being driven bythe input shaft and the ring gear being variably driven by other drive means or, alternatively, the ring gear being driven by the input shaft and the sun gear being variably driven by the other drive means, the other drive means being indirectly driven by the input drive means, the arrangement being such that the rotational speed of the output shaft is proportionaltothe relative rotational speed ofthe ring geartothesun gear.

The other drive means preferably comprises at least one hydraulic pump, preferably a variable-ratio axial-piston hydraulic pump and motor unit that is connected to be driven by the input shaft, the pump and motor unit comprising axially spaced first and second rotatable piston-carrying drums, the second drum being connected to rotate with the sun gear of the epicyclic gear unit; first and second ported valve plates that are arranged for co-operation with the first and second drums respectively, the first valve plate being rotationally fixed and the second valve plate being rotatable relative to the second drum, the first drum and either the second drum or the second valve plate being connected to be rotatably driven by the input shaft;; hydraulicconduitsfluidlyinterconnecting ports in the first and second valve plates respectively, a rotat ionallyfixed variable-angle swash plate that is engageable by a plurality of positive-displacement pistons carried bythethefirstdrum;; and afixed-angle swash plate that is connected to rotate with the second valve plate and with the ring gear ofthe epi cyclic gear unit, the fixed-angle swash plate being en- gageable by a plurality of positive-displacement pistons carried by the second drum, whereby the first valve plate, the first drum and its positivedisplacement pistons, and the variable-angle swash plate co-operate to form a firstfluid-displacement unit, and the second valve plate, the second drum and its positive-displacement pistons, and the fixed angle swash plate co-operate to form a second fluiddisplacement unit that in operation is arranged to be driven hydraulically by the firstfluid-displacement unit.

Such a transmission mechanism is potentially re latively simple, light-weight and low-cost. The pump and motor unit may if desired be based on a conventional positive-displacement unit. There is no need for clutches, belts orfluid coupling, and the transmission mechanism is potentially energysaving because maximum engine power can be utilized throughout the range of transmission ratios.

Essentially, in such a transmission mechanism, con tinuousvariation of the output speed is available by controlling the speed ofthe ring gear or the sun gear of the epicyclic gear unit by means ofthe hydrostatic drive.

Aco-axial arrangement utilising aligned input and output shafts is readily possible.

The first and second valve plates are conveniently disposed adjacent each other, preferably with the interposition of a thrust bearing. Such an arrangement simplifies the externai connections, for example for the supply of make-up fluid to maintain the requisite amount of hydraulic fluid (oil) in the requisite amount of hydraulic fluid (oil) in the positivedisplacement system. The valve plates may each include an opposed pair of circumferentially extending C-shaped grooves for co-operating with the cylinders ofthe respective piston-carrying drum, with the grooves of the respective valve plates communicating with ports that are hydraulically interconnected to provide, in operation, the hydraulic drive in the pump and motor unit.

The swash plates conveniently surround the rotary input shaftwith the interposition of respective bearings; the variable-angle swash plate may be mounted on a swivel pin that is carried by its respective bearing or by the housing within which the pump and motor unit is housed.

Avariable-ratio power transmission mechanism in accordance with the present invention may have the input shaftthereof driven by, for example, an engine that normally operates at constant speed, with the angular position of the variable-angle swash plate controllable by a single regulating device responsive to transmission output speed, for example by means of a hydraulic piston and cylinder arrangement or by an electronic arrangement (eg. a servomotor), in response to the power developed by the engine.

Thereby, it is readily possible to provide at the output shaft, progressively, a range of reverse ratios, a neutral condition (stall condition), and a range offorward ratios including a direct-drive ratio and a range of overdrive ratios.

Power transmission mechanisms in accordance with the present invention are applicable to various industrial uses, and in particular as motor vehicle drives, for example in passenger cars, publictrans- portvehicles, heavy-dutygoodsvehicles,specialised vehicles and waterborne vehicles.

In the drawings: Figure 1 schematically illustrates in fragmentary longitudinal section, with parts in elevation, a preferred embodiment of a variable-ratio powertransmission mechanism in accordance with the present invention; Figure2illustratesC-shaped ports in a pairofvalve platesforming partofthetransmission mechanism shown in Figure 1; Figure3illustrates the flow conditions prevailing at five exemplary settings of a variable-angle swash plate forming partofthe transmission mechanism shown in Figure 1;; Figure4is a graph illustrating relative ring gear speeds for an epicyclic gear unit forming part of the transmission mechanism shown in Figure 1, plotted against a range of transmission output speeds that includes output speeds corresponding to the ex emplaryswash plate settings; Figure 5schematically illustrates in fragmentary longitudinal section, with parts in elevation, an alternative embodiment of a variable-ratio power transmission mechanism in accordance with the invention; and Figure 6schematicaliy illustrates a further alternative embodiment.

As shown in Figure 1 ofthe drawings, a variableratio powertransmission mechanism in accordance with the present invention, applicable for use as a motorvehicle drive, includes a rotary input shaft 10 that is coaxially aligned with a rotary output shaft 12 and is connected to drive a sun gear 14 of an epicyclic gear unit 16.The output shaft 12 is connected to be driven by a planet carrier 18 of the epicyclic gear unit, this planet carrier carrying a plurality of rotatably mounted planetary pinion gears 20thatare in meshing engagement with the sun gear 14 and also with a ring gear22 ofthe epicyclic gearunit. The ring gear 22 constitutes a reaction element ofthe epicyclic gear unit; with control of the rotation ofthe ring gear, in a manner two be described, a mechanical torque path is available from the rotary input shaft 10 to the sun gear 14 of the epicyclic gear unit 16 and thence to the rotary output shaft 12.

Coaxiallysurrounding the rotary input shaft 10 there is a variable-ratio axial-piston hydraulic pump and motor unit 24 (the other drive means) ofthe swash-plate type, forming a hydrostatic (positivedisplacement) transmission that is connected to pro videa hydraulictorque path from the input shaft 1 Oto the ring gear 22 ofthe epicyclic gear unit 16. Specifically, the pump and motor unit 24 comprises axially spaced first and second annular drums 26 and 28 which coaxiallysurround the inputshaft 10and are connected to the input shaft by means of respective keys 30 and 32, such that the annular drums are thereby connected to be driven by, and rotate with, the input shaft.

Axially between thefirstand second annular drums 26 and 28 there are first and second ported annular valve plates 34 and 36 which coaxial Iy sur- round the input shaft 10 and are arranged forco- operation with the first and second drums respectively, a thrust bearing 38 being interposed axially between the valve plates. The first valve plate 34 is held rotationallyfixed by co-operation with a stati onaryhousing 40 ofthe pump and motor unit24. The second valve plate 36 is rotatable relative to the input shaft 10 and drum 28.

As is shown in Figure 2 ofthe drawings, the ports in each of the valve plates 34 and 36 include first and second circumferentially extending C-shaped flu iddistribution grooves 42 and 44 (Figure 2) formed in an end face of the respective valve plate. Each ofthe annular drums 26 and 28 correspondingly has formed in an end face thereof, and extending axially through the drum to the opposite end face ofthe drum, a circumferentially extending series of cylinder bores 46 that slidably accommodate individual positive-displacement pistons 48 and are at the same diameter as, and communicatewith the C-shaped fluid-distribution grooves 42 or 44 in the respective valve plate 34 or 36.

As is aiso shown in Figure 1 ofthe drawings, the first C-shaped grooves 42 ofthe respective valve plates 34 and 36 are hydraulically interconnected by means of a first external hydraulic conduit 50 formed in the housing 40, and correspondingly the second C-shaped grooves 44 of the respective valve plates are hydraulically interconnected by means of a second external hydraulic conduit 52 formed in the housing 40. Acommon supply conduit54connected to a source of make-up pressure fluid (not shown), is connected by way of respective one-wayvalves 56 and 58to the first and second external hydraulicconduits 50 and 52 respectively.

The positive-displacement pistons 48 of the first annular drum 26 project from the opposite end face of the drum for co-operation with a rotationallyfixed variable-angle swash plate 60 (first swash plate). The variable-anlge swash plate 60 coaxially surrounds the rotary input shaft 10 with the interposition of a first bearing 62, the swash plate being pivotally mounted on a swivel pin (pivot pin) 64 carried by the first bearing. The angular position ofthe variableangle swash plate 60 is adjustable by means of a hydraulically actuable piston and cylinder arrange- ment 66 in a stepless fashion throughoutthe working range of the variable-angle swash plate. Forthe pur- pose of description, and for correlation with Figures 2 and 4 of the drawings, five exemplary angular positions (settings) of the variable-angle swash plate 60 are identified in Figure 1 as, successively, V, I, 11,111, andlV.

The positive-displacement pistons 48 ofthe second annular drum 28 project from the opposite end face of the drum for co-operation with a fixed- angle swash plate 68 (second swash plate) that is connected by means of a generally annularconnect- ing portion 70to rotate with the second valve plate 36 and also with the ring gear 22 of the epicyclic gear unit 16.Ring seals 72 are interposed between the housing 40 and the second valve plate 36 and its connecting portion 70,to minimise leakage of hydraulic fluid (oil) contained in the cylinder bores, and in the ports in the valve plates 34 and 36 and the external conduits 50 and 52. Asecond bearing 74 is interposed between the fixed-angle swash plate 68 and the rotary input shaft 10,to permit relative rotary movement of these parts.

Avehicle engine (not shown) (input drive means) is connected to drive the rotary input shaft 10 ofthe transmission mechanism, and the rotary output shaft 1 2 of the transmission mechanism is connected to drive a pair of drive wheels (not shown) of the vehicle. The vehicle engine is controlled so as normally to operate essentially at constant speed, and the piston and cylinder arrangement 66 for adjusting the angular position ofthevariabie-anige swash plate 60 is controllable in response to the power developed by the vehicle engine.

lnthevariable-ratio powertransmission mechanism which has been described, the first valve plate 34, the first annular drum 26 with its positivedisplacement pistons 48, and the variable-angle swash plate 60 co-operate to form a firstfluid- displacement unit 76, and corresponding liy the second valve plate 36, the second annular drum 28 with its positive-displacement pistons 48, and the fixed-angle swash plate 68 co-operate to form a second fluid-displacement unit 78 which in oper ation, overthe greater part oftheworking rangeof thevariable-angle swash plate 60, hydraulically drives the first fluid-displacement unit 76.Thereby, a hydraulic torque path is available from the rotary input shaft 10 by way of the first and second fluid- displacement units 76 and 78to the ring gear 22 of the epicyclic gear unit 16 and thence to the rotary output shaft 12.

Overall, therefore, the variable-ratio powertransmission mechanism which has been described utilises a steplesslyvariable-ratio positivedisplacement hydraulic transmission (constituted by the fluid-displacement units 76 and 78) in a splittorque configuration with a fixed-ratio mechanical torque path (constituted essentially by the rotary input shaft 10 driving the sun gear 14), with recombination oftorque from the hydraulic and mechanical torque paths by the use of the epicyclic gear unit 16, and with the rotational movement ofthe ring gear 22 ofthe epicyclic gear unit, and rotation ofthe fixed- angle swash plate 68, being controlled by means of the second fluid-displacement unit 78.

In operation ofthe variable-ratio powertransmission mechanism, rotation of the annular drum 26 and 28 with the rotary input shaft 10 produces positive displacementofthe hydraulic fluid (oil) astheswash plates 60 and 68 co-operate with the positivedisplacement pistons 48 to produce a pumping action of the positive-displacement pistons in the cylinder bores 46.In conjunction with this pumping action, the ported valve plates 34 and 36 provide appropriate selective fluid connections, inasmuch as each of the cylinder bores 46 draws hydraulicfluid from one of the fluid-distribution grooves 42 or 44 of the respective valve plate 34 or 36 during essentially one-half of a revolution of the input shaft 10, and then delivers hydraulic fluid to the otherfluid-distribution groove ofthe valve plate during the succeeding onehalf of a revolution ofthe input shaft, to provide a hydraulic drive in the pump and motor unit formed by the first and second fluid-displacement units 76 and 78.By varying the angular position ofthevariable-angle swash plate 60, which is of the over-centre type, and without the need for further control members, the transmission mechanism provides atthe output shaft 12, progressively, a range of reverse ratios, a neutral (stall) condition, and a range offorward ratios including a direct-drive (1:1) ratio and a range of overdrive ratios.

Figure 3 ofthe drawings illustrates the hydraulic flow conditions prevailing in the pump and motor unit at five specific angular positions of the variableangle swash plate 60, these five positions having been chosen from the full working rangeofstepless variation of swash plate angle to illustratethetypes oftransmission ratio available, and corresponding to the successive settings V, 1,11,111 and IV shown forthe variable-angle swash plate 60 in Figure 1.

In the following, rotational speeds, pumping volume and oil flow are summarised foreach ofthe five settings I and V of the variable-angie swash plate 60, assuming an engine rotating in the clockwise direction (corresponding to the arrow in Figure 1) at 900 r.p.m., and an epicyclic gearunitwith a ratio of 2:1.

References inthefollowing to relative ring gear speed are to the speed of rotation ofthe ring gear 22 relative to the speed of rotation ofthe input shaft 10.

At setting I of the variable-angle swash plate 60, the speed of rotation of the output shaft 12 is zero (stall condition). The ring gear 22 rotates anti-clockwise at 900 . = 450 r.p.m. Relative ring gear speed is 1350 r.p.m.,so pumpingvolume is 1350timesthe pum- ping volume per r.p.m. There is high oil flow, but without pressure, in conduit 50 from unit 78 to unit 76, and in conduit 52 from unit 76 to unit 78.

At setting 11, the speed of rotation of the output shaft 12 is 90 r.p.m. (forward underdrive condition).

The ring gear 22 rotates anti-clockwise. The transmission ratio (outputshaftspeed input shaft speed) is 90 - 900 = 0.1. Relative ring gear speed is 900 x 3/2(1 ) = 1215 r.p.m., so pumping volume is 1215timesthe pumping volume per r.p.m. There is reduced oil flow, at high pressure, in conduit 50 from unit 78 to unit 76, and reduced oil flow, without pressure, in conduit 52 from unit 76 to unit78.

At setting Ill, the speed of rotation of the output shaft 12 is 900 r.p.m. (direct-drive condition), corresponding to a transmission ratio of 900 -: 900 = 1.

The ring gear 22 rotates clockwise. Relative ring gear speed is 900 x 3/2(1 - 1) = 0 r.p.m., so pumping volume is correspondinglyzero. There is onlyannular oil flow, at moderate pressure, in the C-shaped grooves 42 of the valve plates 34 and 36, and only annular oil flow, without pressure, in the C-shaped grooves 44 ofthe valve plates 34 and 36.

At setting IV, the speed of rotation of the output shaft 12 is 1,000 r.p.m. (forward overdrive condition).

The ring gear 22 rotates clockwise. The transmission ratio is 1,000 + 900 = 1.11. Relative ring gear speed is 900 x 3/2(1.11 - 1) = 150 r.p.m., with a pumping volume of 150 times the pumping volume per r.p.m.

In this overdrive condition, the unit 76 acts as the pump. There is low oil flow, at high pressure, in conduit 50 from unit 76 to unit 78, and lowoilflow, without pressure, in conduit 52 from unit 78 to unit 76.

At setting V, the speed of rotation of the output shaft 12 is 100 r.p.m. in the reverse direction (reversedrive condition). The ring gear 22 rotates anticlockwise. The transmission ratio is -100 . 900 = -0.11. Relative ring gearspeed is 900 x 3/2(1 + 0.11) = 1500 r.p.m., such thatthe pumping volume is 1500 timesthe pumping volume perr.p.m.Thereisvery high oil flow,without pressure, in conduit 50 from unit78to unit 76, and very high oil flow, atveryhigh pressure, in conduit 52 from unit76to unit78.

With regard to the direction of rotation of the ring gear 22, is should be noted that the ring gear is stati onary ifthe relative ring gear speed is equal to the speed ofthe input shaft 10 (namely 900 r.p.m.), in the opposite direction. If the transmission ratio forthis condition is denoted by R, one obtains the equation 900x3/2(1 - R) = 900,from which 1 - R = 2/3 = 0.6666, and hence R = 0.3334. Below this ratio of 0.3334 the ring gear 22 rotates anti-clockwise, and above this ratio of 0.3334the ring gear 22 rotates clockwise.

In Figure 1 ofthe drawings the epicyclic gear unit 16 is shown as a simple epicyclic gear unit, having a ratio of2:1 The epicyclicgear unit could alternatively be derived from a Ravigneaux gearset, by elimination of redundant elements. Also, the ratio ofthe epicyclicgearunitcould have a greaterorlesser valuethan the above figure of2:1.

In Figure 1 of the drawings, the swash plates 60 and 68 are shown schematically. To reduce friction, either or both of the swash plates could in practice comprise a drive plate portion and a main swash plate portion,with a bearing interposed to permit relative rotation of the said portions.

Figure 4 of the drawings is a graph illustrating the above-described relationships between relative ring gear speed (which determines pumping volume) and the speed ofthe output shaft 12, as well as thetransmission ratios at the five exemplary setting I to V of the variable-angle swash plate 60. The graph is thus plotted for an epicyclic gear unit ratio of 2:1 and a speed of the input shaft 10 of 900 r.p.m.: variation of the input shaft speed (engine speed) would vary the vertical position of the straight-line characteristic.

In the alternative embodiment shown in Figure 5, parts which correspond to those shown in Figure 1 have been give the same reference numbers. In this arrangement, the sun gear 14 and the ring gear 22 of the epicyclic gear unit 16 reverse roles. The sun gear 14 constitutes the reaction element of the epicyclic gear unit 16 and the ring gear is driven by the input shaft 10. As can be seen from Figure 5, the second valve plate 36 is connected to the input shaft 10 by means of key 79, thereby providing the drive to the ring gear 22. The second drum 28 is mounted on a separate rotary shaft 80, and connected to it by key 81. The shaft 80 is axially aligned with input shaft 10 and output shaft 12, and is connected to the sun gear 14 of the epicyclic gear unit 16.In this arrangement, thrust bearings are not required between the first and second valve plates 34,36. In practice, the operation ofthis arrangement is identical to that shown in Figure 1.

The further alternative shown in Figure 6 is identical to the arrangement shown in Figure 1 (again, reference numbers correspond), except that the engine 82 (the input drive means) is positioned between the first fluid displacement unit 76 and the second fluid displacement unit 78. The firstfluid displacement unit 76 can be either directly driven by the engine 82, or it can be belt driven.

In the arrangements shown above, the sun gear 14, planet carrier 18 and ring gear 22 of the epicyclic gear unit 16 all rotate about a common axis which is aligned with the axis of rotation ofthe input shaft 10 and output shaft 12. This invention is not, however, restricted to such an arrangement as the axes ofthe input shaft 10 and outputshaft 12may be offsetfrom the common axis of the epicyclic gear unit 16, with a belt of cogwheel drive interconnecting the input shaft 10 with the sun gear 14 or ring gear 22, and the output shaft 12 with the planet carrier 18.

As still further alternatives, thefirstfluid displacement unit 76 may be replaced by a variable radial piston pump, and/orthe second fluid displacement unit 78 may be replaced byeithera gear pump ora vane pump ora radial piston pump.

Claims (18)

1. Avariable-ratio power transmission mechanism comprising an input drive means, an input shaft rotatably drivable about its longitudinal axis by the input drive means, an output shaft rotatably drivable about its longitudinal axis, and an epicyclic gear unit, the epicyclic gear unit comprising a sun gear, a planet carrier and a ring gear, all rotatable about a common axis, the planet carrier being connected to, and rotatably driving, the output shaft, and having a plurality of rotatable planetary pinions in meshing engagement with the sun gear and the ring gear, the sun gear being driven by the input shaft and the ring gear being variably driven by other means or, alternativelythe ring gear being driven by the input shaft and the sun gear being variably driven by the other drive means, the other drive means being indirectly driven by the input drive means, the arrangement being such that the rotational speed of the output shaft is proportional to the relative rotational speed ofthe ring geartothe sun gear.

2. Avariable-ratio powertransmission mechanism according to claim 1, in which the other drive means comprises at least one hydrau I ic pum p.

3. Avariable-ratio powertransmission mechanism according to claim 1 or claim 2, in which the other drive means comprises a variable-ratio axialpiston hydraulic pump and motorunitthatisconnec- ted to be driven by the input shaft, the pump and motor unit comprising axially spaced first and second rotatable piston-carrying drums, the second drum being connected to rotate with the sun gear of the epicyclic gear unit; first and second ported valve plates that are arranged for co-operation with the first and second drums respectively, the first valve plate being rotationally fixed and the second valve plate being rotatable relative to the second drum, the firstdrum and either the second drum or the second valve plate being connected to be rotatably driven by theinputshaft; hydraulicconduitsfluidlyinterconnecting ports in the first and second valve plates respectively; a rotationally fixed variable-angle swash plate that is engageable by a plurality of posi- tive-displacement pistons carried by the first drum; and a fixed-angle swash plate that is connected to rotate with the second valve plate and with the ring gear of the epicyclic gear unit, the fixed-angle swash plate being engageable by a plurality of positivedisplacement pistons carried by the second drum, whereby the firstvalve plate, the first drum and its positive-displacement pistons, and the variableangle swash plate co-operateto form a firstfluiddisplacement unit, and the second valve plate, the second drum and its positive-displacement pistons, and the fixed-angle swash plate co-operate to form a second fluid-displacement unit that in operation is arranged to be driven hydraulically by thefirstfluiddisplacement unit.

4. Avariable-ratio power transmission mechanism according to claim 3, in which the variableratio axial-piston hydraulic pump and motor unit coaxially surrounds the input shaft, with the first drum and the second drum orthe second valve plate connected to rotate with the input shaft.

5. Avariable-ratio powertransmission mechanism according to claim 3 or claim 4, in which the first and second valve plates are disposed adjacent each other, with or without an interposed thrust bearing.

6. Avariable-ratio power transmission mechanism according to any one of claims 3 to 5, in which the ports in each valve plate include a pair circum- ferentially extending C-shaped fluid-distribution grooves formed in an end face of the respective valve plate, each drum correspondingly has formed in an end face thereof a circumferentially extending series of cyii nders that accom modate the individual positive-displacement pistons and communicate with the C-shaped grooves of the respective valve plate, and the hydraulic conduits each fluidly connect one ofthe C-shaped fluid-distribution grooves in one valve plate to the corresponding fluid-distribution groove in the other valve plate, to provide an arrangement in which in operation each cylinder draws hydraulic fluid from one of the fluid-distribution grooves ofthe respective valve plate during one-half of a revolution of the input shaft and then delivers hydraulic fluid to the other fluid-distribution groove ofthe valve plate during the succeeding one-half of a revolution of the input shaft, to provide the hydraulic drive in the pump and motor unit.

7. Avariable-ratio power transmission mechanism according to claim 6, including a source of pressure fluid connected by means of a one-way valve to the fluid-distribution grooves, for the supply of make-upfluid forthe hydraulic conduits.

8. Avariable-ratio power transmission mechanism according to any one of claims 4to 7, in which a first bearing is interposed between the input shaft and the variable-angleswash plate.

9. Avariable-ratio powertransmission mechanism according to claim 8, in which the variableangle swash plate is pivotally mounted on a swivel pin that is carried by the first bearing.

10. Avariable-ratio power transmission mechanism according to claim 8, in which the variableangle swash plate is pivotally mounted on a swivel pin that is carried by the housing within which the pump and motor unit is housed.

11. Avariable-ratio power transmission mechanism according to any one of claims 3 to 10, in which the angular position ofthe variable-angle swash plate is adjustable by means of a hydraulically actuabie piston and cylinder arrangement or by an electronic arrangement to provide at the output shaft, progressively, a range of reverse ratios, a neutral condition, and a range of forward ratios including a direct-drive ratio and a range of overdrive ratios.

12. Avariable-ratio power transmission mechanism according to claim 11, in which the piston and cylinder arrangement for adjusting the angular position of the variable-angle swash plate is controllable in response to the power developed bythe input drive means that normally operates at constant speed.

13. Avariable-ratio power transmission mechanism according to any one of claims 1 to 12, in which the output shaft is connected to drive a pair of drive wheels of a vehicle.

14. Avariable-ratio power transmission mechanism according to any one of claims 1 to 13, in which the input drive means comprises a vehicle engine.

15. Avariable-ratio power transmission mechanism according to claim 2, in which the at least one hydraulic pump comprises a gear pump, or a vane pump, or a radial piston pump.

16. Avariable-ratio power transmission mechanism substantially as hereinbefore described with reference to, and as shown in figures 1 to 4 ofthe accompanying drawings.

17. Avariable-ratio power transmission mechanism substantially as hereinbefore described with reference to, and as shown in, figure5 of the accompanying drawings.

18. Avariable-ratio power transmission mech anism substantially as hereinbefore described with reference to and as shown in Figure 6 ofthe accompanying drawings.