WO2006108109A2 - Hydromechanical continuously variable transmission - Google Patents

Abstract

A hydromechanical continuously variable power transmission (Figs. 1 and 2) for converting rotating mechanical power at one combination of rotational velocity and torque to another combination of rotational velocity and torque over a continuous range, inclu a hydraulic pump (30), operatively driven by an input shaft (50), and a hydraulic motor (35) operatively driving an output shaft (51). The hydraulic pump (30) and hydraulic motor (35) are coupled together mechanically through a pair of planet sets (40, 45), and are coupled together hydraulically through a manifold (52), such that hydraulic fluid pressurized by said pump (30) drives the motor (35 and spent fluid from the motor (35) is cycled back to the pump (30) where it is re-pressurized. Both planet sets (40, 45) are arranged axially with the input shaft (50) and the output shaft (51), and the hydraulic pump (30) and hydraulic motor (35) are arranged in seri with each other on opposite sides of the manifold (52), and parallel to the input and output shafts (50, 51), thereby optimizing the us space and keeping the overall length of the transmission to a minimum, and minimizing required lengths of said input and output shafts (50, 51).

Description

HYDROMECHANICAL CONTINUOUSLY VARIABLE TRANSMISSION

This is related to U.S. Provisional Application No. 60/668,730 filed on April 5, 2005 and entitled Hydromechanical Continuously Variable Transmission. This invention pertains to hydro-mechanical power transmissions, and more particularly to a continuously variable hydromechanical power transmission for use especially in light truck and automobile applications where an overdrive final ratio is desired, and where the HSU performance and the overall packaging are optimized to achieve a small and lightweight transmission able to accommodate the highest power engines currently available for the light truck/full size automotive application.

BACKGROUND OF THE INVENTION

Continuously variable transmissions have become recognized in recent years as particularly desirable as vehicle power transmissions because of the operational efficiencies and economies that they can potentially afford. However, conventional prior art hydrostatic transmissions are known by experts in the art to be noisy and inefficient, and mechanical CVTs have power and durability limitations. To achieve significant usage throughout the potential vehicle transmission market, significant improvements in cost, durability, performance, and power capacity will be needed. One such improvement would be in the area of leakage from rotating interfaces, particularly those where working fluid is commutated between the differentially rotating pump and motor.

Another improvement would be in the area of dynamic balancing. The difficulty of balancing rotating equipment to preclude vibration induced by rotating eccentric masses becomes worse exponentially with increasing speed of rotation.

Yet another improvement would be in reducing the losses caused by "windage" and fluid drag associated with the rotating elements inside the transmission housing. In applications having a prime mover with a high rotating speed, such as an electric motor, turbine engine or high performance spark ignition gasoline engine, the input elements would rotate at the prime mover output speed unless a gear reduction unit were interposed between the prime mover and the transmission. Gear reduction units add undesirable cost and weight. The windage and fluid drag losses can be greatly reduced by reducing the speed of rotation of those rotating elements.

Still another desirable improvement would be in the area of manufacturability, simplicity, and cost. Prior art continuously variable hydromechanical transmissions have tended to be excessively complicated and costly to build. It would be a welcome development to original equipment manufacturers to have a continuously variable hydromechanical transmission available that is efficient, small and light weight, and is easily and economically manufactured and maintained.

One approach for achieving these improvements is shown in an international patent application No. PCT/US98/24053 filed on November 12, 1998 by Folsom and Tucker entitled "Hydraulic Machine", now issued as U.S. Patent No. 6,358,174. A variation of this approach in a tandem hydromechanical transmission using low cost conventional components is shown in International Patent Application No. PCT/US99/28,083, now issued as U.S. Patent No. 6,530,855 issued March 11 , 2003, and pending in Application No. 10/386,874 entitled "Parallel Hydromechanical Underdrive Transmission" provides improvements that this technology available for smaller vehicles requiring more compactness and lower cost, such as outboard motors for boats, motor scooters, motor cycles, RVs and snowmobiles and small cars. The greatest benefit for use of a high efficiency continuously variable power transmission would be for automobiles and light trucks, as well as other larger vehicles, particularly those in which an overdrive final ratio is desirable. This class of vehicles consumes by far the most gasoline and diesel fuel and where significant improvements in fuel economy could benefit vehicle owners and make a profound contribution to the betterment of the planet.

Summary of the Invention

A hydromechanical continuously variable power transmission for converting rotating mechanical power at one combination of rotational velocity and torque to another combination of rotational velocity and torque over a continuous range, includes a hydraulic pump, operatively driven by an input shaft, and a hydraulic motor operatively driving an output shaft. The hydraulic pump and hydraulic motor are coupled together mechanically through a pair of planet sets, and are coupled together hydraulically through a manifold, such that hydraulic fluid pressurized by said pump drives the motor, and spent fluid from the motor is cycled back to the pump where it is re-pressurized.

Both planet sets are arranged co-axially with the input shaft and the output shaft, and the hydraulic pump and hydraulic motor are arranged in series with each other on opposite sides of the manifold, and parallel to the input and output shafts, thereby optimizing the use of space and keeping the overall length of the transmission to a minimum, and minimizing required lengths of said input and output shafts. Description of the Drawings

The invention, and its many attendant features and benefits, will become better understood upon reading the following description of the preferred embodiment, in conjunction with the following drawings, wherein:

Fig. 1 is a schematic diagram of a continuously variable hydromechanical power transmission in accordance with this invention;

Fig. 2 is a sectional elevation of a continuously variable hydromechanical power transmission embodying elements of the schematic diagram of Fig. 1 ; Fig. 2A is an enlarged sectional elevation of the left hand side of the transmission shown in Fig. 2;

Fig. 2B is an enlarged sectional elevation of the right hand side of the transmission shown in Fig. 2;

Fig. 3A is a perspective view of the transmission case for the transmission shown in Fig. 2, without the attachment flange, to reveal portions of the interior;

Fig. 3B is a perspective view of the transmission case and attachment flange by which the transmission is attached to a vehicle;

Fig. 4 is a perspective view of a planet carrier of an output planet set of the transmission shown in Fig. 2; Fig. 5 is a sectional elevation of the planet carrier shown in Fig. 4;

Fig. 6 is a perspective view of a planet carrier of an pump planet set of the transmission shown in Fig. 2;

Fig. 7 is a sectional elevation of the planet carrier shown in Fig. 6;

Fig. 8 is sectional elevation of the pump shown in Fig. 2; Fig. 9 is a perspective view of the motor shown in Fig. 2;

Fig. 9A is a perspective view of a torque plate for the pump or motor shown in Fig. 2;

Fig. 9B is a sectional elevation of one side of the torque plate shown in Fig. 9A;

Fig. 10 is a sectional plan view along lines 10-10 in Fig. 2;

Fig. 11 is a sectional elevation along lines 11-11 in Fig. 2 Fig. 12 is a sectional plan view along lines 12-12 in Fig. 2

Fig. 13 is an enlarged sectional plan view of one of the displacement control devices shown in Fig. 12;

Fig. 14 is a perspective view of the motor hydrostatic unit displacement control device shown in Fig. 12, shown connected to the motor yoke and showing the motor at zero displacement;

Fig. 15 is a sectional elevation along lines 15-15 in Fig. 2A; Fig. 16 is a schematic diagram of another embodiment of a continuously variable hydro-mechanical power transmission in accordance with this invention;

Fig. 17 is an enlarged schematic diagram of a portion of the second embodiment shown in Fig. 16; Fig. 18 is a schematic diagram of a third embodiment of a continuously variable hydro-mechanical power transmission in accordance with this invention;

Fig. 19 is a partial sectional elevation of a physical embodiment of the design shown in Fig. 18, showing a rear end planet set and brake replacing the clutch arrangement in the rear end of the embodiment shown in Fig. 2. Fig. 20 is a partial sectional elevation of the rear end planet set and brake shown in

Fig. 19.

Description of the Preferred Embodiments

Turning now to the drawings, wherein like reference numerals designate identical or corresponding views, and more particularly to Fig. 1 thereof, a hydromechanical continuously variable transmission is shown schematically. One embodiment of the transmission design shown in Fig. 1 is shown in Fig. 2. It will be understood that the design shown in Fig. 1 and the physical device shown in Fig. 2 are separate illustrative embodiments of the invention and that other embodiments within the scope of the invention can and will occur to those skilled in the art in light of this description.

The transmission has a case 25, shown in Fig. 3A, and shown in Fig. 3B attached to a mounting flange 26 for attachment to a vehicle. An oil pan 27, and rear module housing 28 are also shown attached to the case 25 in Fig. 3B . Inside the case 25, the transmission includes an input hydrostatic unit or pump 30, operatively driven by an input shaft 50, and an output hydrostatic unit or motor 35, operatively driving a tubular output shaft 51. The hydrostatic units are similar to the hydrostatic unit shown in Patent No. 6,874,994, entitled "Hydraulic Pump and Motor". The hydraulic pump and hydraulic motor 30 and 35 are coupled together mechanically through a pair of coupled planet sets, namely, an output planet set 40 and a pump planet set 45, and are coupled together hydraulically through flow passages 42 and 43 directly through a manifold 52, such that hydraulic fluid pressurized by the pump 30 drives the motor 35, and spent fluid from the motor 35 is cycled back to the pump 30 where it is re-pressurized.

The transmission is shown in Fig. 2 at neutral, with the pump 30 at zero displacement, and the motor 35 at maximum displacement. Both hydrostatic units can be either controlled together or independently controlled depending upon the application. The pump 30 is operatively driven by the input shaft 50, acting through the output planet set 40 and the pump planet set 45. The input shaft 50 is driven by a prime mover, such as a vehicle engine 55, by way of a coupling 54, shown only in Fig. Y1. The input shaft 50 has a drive gear 53 which drives a makeup pump 56 housed in the front bulkhead of the mounting flange 26. The coupled planet sets 40 and 45 are mounted in a gear housing 46, which is fastened to a step 47 in the housing with bolts 48. The gear housing also supports the manifold block 52 with bolts 49. The output planet set 40 has a planet carrier 60 that is driven by the input shaft 50, by way of a spline on the inner end of the input shaft 50 engaged with a spline 59 in the bore of a planet carrier 60, as shown in Fig. 5.

Input power from the engine 55 drives the planet carrier 60 of the output planet set 40. The planet carrier 60, as shown in Figs. 4 and 5, carries a plurality of planet gears 63, which are engaged with a ring gear 65 of the output planet set 40, as shown in Fig. 2B. The ring gear 65 is connected drivingly to the main output shaft 51 and delivers reaction torque from the output planet set to the main output shaft 51 , as explained in more detail below.

The planet gears 63 of the output planet set planet carrier 60 are also engaged with a sun gear 70, which is connected axially to a sun gear 71 of the pump planet set 45. The pump planet set also has a planet carrier 75 with planet gears 80, as shown in Figs. 6 and 7. The planet gears 80 are engaged between the sun gear 71 and a pump planet set ring gear 85. The planet carrier 75 of the pump planet set 45 is fixed to ground by way of a spline 77 on the planet carrier 75 engaged with a central splined opening in a grounding flange 81. The grounding flange 81 is mounted in the gear housing 46 by a splined connection 83 to the gear housing 46. The sun gear 70 of the output planet set 40 is connected drivingly to the sun gear 71 of the pump planet set 35, and both sun gears 70 and 71 are supported on bearings on the input drive shaft 50. The ring gear 85 of the pump planet set 45 is connected to a drive flange 90 which extends through an axial opening in the front side of the gear housing and connects to an input chain sprocket 95, which is supported for rotation on a bearing 97 mounted in the axial opening in the front side of the gear housing. A silent (Morse) chain 100 driven by the chain sprocket 95 is also engaged with a pump torque plate sprocket 105 fastened to a torque plate 110 on the pump 30, shown in detail in Fig. 8.

The torque plate 110, shown in detail in Figs. 8, 9, 9A and 9B, is supported for rotation about a longitudinal axis 112 through a manifold block 52 on needle bearings 115. The torque plate 110, shown in Fig. 8, serves as a commutating fluid flow interface between spherical heads 120 of pistons 125 in bores 130 of a pump cylinder block 135, and the face of the manifold block 52, as well as the means for transmitting power to and from the hydrostatic unit. The orientation of the hydrostatic unit in Fig. 8 corresponds to the orientation of the pump 30 in Fig. 2, and the orientation of the hydrostatic unit in Fig. 9 corresponds to the orientation of the motor 35 in Fig. 2, but in fact both the pump and motor hydrostatic units 30 and 35 are identical, so for purposes of this description, only one hydrostatic unit will be described, with the understanding that this description applies to both hydrostatic units 30 and 35.

Pressure in the cylinders acting on the annular area of the pistons 125 places an axial force on the pistons, pushing the spherical piston head 120 into torque plate sockets 126, shown in Figs. 9A and 9B. To react the load, an underbalance hydrostatic bearing between the spherical piston head 120 and the torque plate socket 126 is used. The hydrostatic bearing that is in the torque plate socket 126 is comprised of an internal annular spherical area 127 that is subjected to full pressure from the respective cylinder bore. The bottom of this area 127 terminates in a blind hole 128 that communicates with a kidney slot 260 of that socket on the manifold-side face 129 of the torque plate 110. The separating forces from these two annular areas are such that there is enough clamping force to hold the piston head 120 seated in the torque plate socket 126 to seal working fluid from escaping past this interface while keeping the contact force low enough so as to avoid appreciable wear at this interface. The spherical sockets 126 have a parallel section 119 at the opening of the sockets 126 before the spherical section that is close fitting to the outside diameter of the piston head ball to reduce leakage past the piston ball if it were to become unseated from the socket.

To ensure that full pressure from the respective cylinder bore 130 acts over the internal annular area a small annular groove 131 is placed into the torque plate socket 126, as shown in Fig. 9B. The respective kidney slot 260 opening in the bottom of the torque plate socket 126 breaks into this annular groove 131, so that any pressure that exists in the torque plate kidney slot 260 is communicated to this groove. The internal annular area is now subjected to full pressure from both outside by this groove 131 and inside from the blind hole 128 that communicates with the kidney slot 260.

To change the balance of this hydrostatic bearing, the diameter of the piston spherical head 120 can be increased or decreased and/or the position (and hence diameter) of the annular groove 131 can be changed.

As the annular groove 131 is subjected to full cylinder pressure a small hole 124 is used to feed pressure from this groove to the orifice 280 that is used to feed overbalance grooves on the manifold-side face 129 of the torque plate 110. The cylinder block 135 is mounted for rotation on a bearing 140 mounted on a post 145 fixed in the base 148 of a supporting yoke 150. As shown in Fig. 10, the yokes 150 each have arms 155 that are mounted for swiveling about two parallel lateral pivotal axes in bearings 160 mounted in links 165 attached to both lateral sides of the manifold block 52. The cylinder block bores 130 are through bores; the piston heads 120 protrude from inwardly facing open ends of the bores 130 and seat in the torque plate sockets 126. Pucks 170 seal the opposite ends of the bores 130. The pucks 170 each have a back side with a shallow recess surrounded by a peripheral land. A central restricted fluid orifice 175 communicates through the pucks 170 between the cylinders 130 and the recess to allow a low volume flow of fluid pressurized in the cylinders 130 into the region on the back side of the pucks to create a fluid cushion, acting as a hydrostatic bearing, to lubricate and support the cylinder block 135 as it rotates against the inner face 180 of the yoke base 148, as explained in more detail below.

The working pressure of the hydraulic fluid inside the cylinders 130 acting on the area at the bottom of the bore creates an axial load. This axial load acts in the opposing direction to that of the axial load created by the torque plate 110. This load is then reacted by the yokes 150. To support the axial load between the cylinder blocks and the yokes, whilst allowing for the rotational speed between the cylinder blocks 135 and the yokes 150, the hydrostatic bearing under the outside face of the puck 170 is preferred. It is of course possible to use rolling element bearings, but their size and life ratings make them less desirable in this application. The shallow recess and peripheral land on the outside face of the pucks 170 produce an active area and a sealing land. The active area is designed such that, when oil from the cylinder bore flows to this area via the restricted orifice 175, the pressure of this oil acting over the active area within the land will place the puck in balance with the axial load placed upon it. This balance can be less than, equal to, or more than 100% depending on the geometry of the features used and the size of the passage that allows oil to flow from the piston bore. If the balance is less than 100% (i.e. underbalanced) then there will be a resultant axial load that will force the puck in direct mechanical contact with the yoke. If the balance is more than 100% (i.e. overbalance) there will be a force that will try to separate the puck from the yoke, in this case the lubrication hole will be sized such that oil leaking past the separated puck will cause a pressure drop as it flows through the lubrication hole, therefore reducing the separating force until the puck comes to a equilibrium state. In this state the puck will be floating on a thin film of oil, whose thickness is determined by the leakage rate of the oil, this leakage rate being determined by the pressure drop of the leaking oil flowing thru the small lubrication hole (orifice). Therefore, by changing the diameter of the orifice 175, it is possible to vary the film thickness and the leakage rate. One benefit of using the overbalance design is that the puck will "float" on a film of oil and will have little or no metal-to-metal contact, this will reduce the wear at this interface and result in higher allowable rotational speeds. The orifice 175 will need to be sized such that there will be no failure of this bearing under the harshest of operating conditions whilst keeping the leakage rate to a minimum.

The pucks 170 have springs 185 placed between them and the cylinder blocks that place an axial load separating the puck 170 from the cylinder block 135, keeping the pucks held firmly against the yoke face 180 until hydraulic pressure can properly balance the forces placed upon them. This axial spring force also has the effect of pushing the cylinder block 135 away from the yoke face 180 towards the torque plate 110. As the cylinder block 135 is axially located relative to the torque plate 110 by a center piston 190, this axial spring force also keeps the torque plate held firmly against the manifold. This makes for an efficient use of the spring as it preloads both the puck and the torque plates.

By using individual pucks 170 as opposed to one large hydrostatic bearing plate the hydrostatic bearing is more compliant to deflections and out-of-flat running surfaces. This is because the individual puck can pivot slightly so that it can follow the form of its running surface. Any deviations in flatness acts over the circumference of the relatively small diameter of the puck. If the hydrostatic bearing were formed as one large component (such as if it were formed directly on the back of the cylinder blocks) even if it were allowed to pivot so that it could follow the form of its running surface, any deviations in flatness would be acting over the circumference of a much larger diameter and hence would have a much greater effect on the bearing. This larger hydrostatic bearing would then require much stiffer (and hence larger and heavier) running surfaces so as to keep the leakage and performance of the bearing at the same level as that of the individual puck type hydrostatic bearings.

As stated in U.S. Patent No. 6,530,855, entitled "Hydraulic Pump and Motor", the pistons 125 are used to drive the cylinder block in synchronous rotation with the torque plate 110. This is done by means of the tapered outside diameter of the piston 125 running against the cylinder bore 130. The angle of this taper is made large enough to allow for the piston to articulate freely as the cylinder block articulates about the pivot axis, as well as to allow for positional mis-alignment of the cylinder block rotating and pivotal axis relative to the rotational axis of the torque plate 110. However the taper on the piston also allows the cylinder block to 'lag' the torque plate in rotation by a few degrees, and this places an opposing torque on the cylinder block from the torque plate. This opposing torque will add to the bearing drag torque on the cylinder block, increasing the side load on the piston taper, which will result in increased wear at this interface as well as reduced torque efficiency. It is therefore desirable to reduce the angle of the taper to reduce the lag angle as much as possible, which necessitates that the centerline and pivotal axis of the cylinder block be accurately located to intersect the torque plate rotational axis. In order to accurately locate the rotational and pivotal axis of the cylinder block relative to the torque plate rotational axis, the cylinder block is provided with a central bore 195 in which the center piston 190 is located with a precision fit. The center piston 190 has a head ball 200 that seats in a socket 210. The center of the head ball 200 is located on the rotational and pivotal axis of the cylinder block and intersects the rotational axis of the torque plate. The head ball 200 is part of the center piston, and the socket 210 is formed into a ring that is supported and located inside the bore of a protruding end of a support shaft 215 fixed in the manifold 52. The torque plate 110 is supported for rotation against the face of the manifold, and against radial forces acting on it, by a radial bearing 220 mounted in a bearing recess 221 in a central bore 222 through the torque plate 110. The bearing 220 supports the torque plate 110 on the outside of the protruding end of the support shaft 215. This ensures that, when the yoke 155 pivots in its bearings 160, the centerline of the cylinder block 135 will articulate about the ball and socket centerline and therefore accurately locate the cylinder block to the torque plate 110 regardless of in which direction the cylinder block is articulated.

The motor chain sprocket 105 attached to the motor torque plate 110 drives a motor silent chain 100, which is trained around and drives a motor chain sprocket 225. The motor chain sprocket 225 is splined to a tubular output shaft 230, which is connected to the main output shaft 51 by way of a releasable clutch 240. The clutch 240 is actuated by makeup fluid pressure controlled by a solenoid operated valve 242 and boosted, as required by a system controller 245 via a signal to a boost valve 246. The boost valve 246 effectively resets the set point of a makeup pump regulator valve 247 which controls the makeup pump 56 through a makeup pump control piston 248.

A controlled hydrostatic bearing 250 is provided on the manifold-side face 129 of the torque plates 110 shown in Fig. 9, that is, the face of the torque plate 110 that is in fluid engagement with the face of the manifold block 52. This hydrostatic bearing provides a fluid interface between the rotating torque plate 110 with the stationary manifold face, allowing the torque plate to run freely against the face of the manifold block while minimizing fluid leakage out of the interface and transferring fluid at high pressure from the pump through the manifold to the motor, and spent fluid back from the motor to the pump. The hydrostatic bearing 250 has an overbalance hydrostatic bearing in the form of shallow individual wedge recesses 255 radially inside an underbalance hydrostatic bearing in the form of kidney-shaped ports 260 which communicate fluid pressure through the torque plate 110 from the piston head sockets on the other side. The wedge recesses 255 are defined by surrounding land frames 265 which in turn are delineated by a shallow annular groove 270 having holes 275 that communicate with the piston-side face of the torque plate 110. An orifice 280 extends from the center of each wedge recess 255 through to the rear side of the torque plate communicating with the spherical sockets in which the piston heads are seated to supply fluid under system pressure to the wedge recesses 255 to provide the fluid pressure to support the torque plate 110 on a fluid cushion on the manifold face. The excess load carrying capacity of the controlled hydrostatic bearing separates the torque plate 110 from the manifold face to the extent that leakage flow around the land frames 265 into the groove 270 exceeds the flow capacity through the orifices 280 and creates a fluid pressure drop across the orifices between piston head spherical sockets and the wedge recesses 255. This pressure drop reduces the axial force exerted by the controlled hydrostatic bearing until the axial spacing between the torque plate 110 and the manifold face reaches an equilibrium where the axial force exerted by the two hydrostatic bearings just balances the axial force exerted by the pistons 125. The leakage from this hydrostatic bearing can be limited to an acceptable rate by correct choice of the orifice diameter so that the desired balance of leakage through the bearing and reduced torque loss is achieved. An annular groove 281 radially outside the kidney-shaped ports 260 collects any leakage flowing radially outward from kidney-shaped ports 260, and radial spoke groves 283 direct this flow radially to lubricate the interface between the manifold face 129 and the pads 261 formed between the spoke grooves 283.

Control Operation:

The operation of the displacement controls for the hydrostatic units will now be described in connection with Figs. 11-14. These controls are in accordance with a control regime, as follows, that is envisioned to govern operation of transmissions in accordance with the invention:

To achieve the maximum output torque from the transmission, the motor is set at maximum displacement under maximum pressure to generate maximum hydraulic torque, whilst having maximum input torque reacted to the output via the planet set arrangement. To achieve this it is beneficial to have a control regime that will hold the motor at its maximum displacement as the pump is stroked from zero displacement until the pump reaches a displacement where it can generate maximum pressure whilst reacting maximum input torque. Once this position has been reached, the pump and motor can be stroked simultaneously (the motor at a slightly faster rate) so that the pump and motor reach their final displacements (pump at max disp, motor at zero disp) at the same time. The advantage of this control regime is that this will minimize the maximum flow rate in the transmission and hence reduce flow losses and noise generation. Transmission ratio, however, is determined by the ratio of the pump to motor displacements and as long as this ratio is the same, the transmission ratio will be the same regardless of the actual value of the pump and motor displacements. As the pump and motor are controlled individually it is therefore possible to achieve the same transmission ratio with a combination of actual pump and motor displacements, and it may be beneficial, for reasons of efficiency, noise etc under certain driving conditions to have the pump and motor at a smaller or larger displacement to achieve any given ratio. During development and testing it is possible to characterize the transmission to find its optimum value of pump and motor displacements for any given ratio under various driving conditions (in terms of input speed and torque). The controller 245 can then choose the optimum value for the pump and motor displacements based upon the various signals the controller receives to give the best performance for any given required transmission ratio.

One embodiment of displacement controls in accordance with the invention uses system pressure to energize the actuator. The system pressure is tapped off the manifold block 52 through two check valves 282 in a bore 284 that extends through the manifold block 52 and intersects the main flow channels 42 and 43 through which the pump 30 and motor 35 communicate, as shown in Fig. 11. This same bore 284 also holds check valves 290 and 292 through which makeup fluid will flow into whichever of the two main flow channels 42 or 43 is at low pressure. This makeup fluid flow supplies fluid to the pump/motor circuit to make up for fluid lost in leakage, and also to provide fluid to the lubrication and cooling circuit, as described in more detail below.

System pressure, captured from whichever circuit is at the higher pressure, is used actuate a control actuator 300, one for each hydrostatic unit 30 and 35. The two control actuators are mounted, one on each lateral side of the manifold block 52, to the links 165. The pump control actuator is shown in cross-sectional detain in Fig. 13, and the motor control actuator is shown in perspective in Fig. 14. The control actuators 300 control the displacement of the hydrostatic units by controlling the angle that the cylinder blocks 135 and pistons 125 make relative to the fixed (upright) orientation of the torque plate 110 and manifold face. This hydrostatic unit angle is controlled by controlling the tilt angle of the control yokes 150 about the laterally extending pivot axes through the bearings 160. Due to planet set configuration, the pump 30 rotates in the opposite direction to the motor 35, both hydrostatic units are stroked in the same direction, that is, when the transmission is viewed from the side, both the yokes 150 rotate about their respective axes in either a clockwise or counter clockwise direction simultaneously. The pump yoke 150 is connected to a pump control piston 305 via a control arm 304 coupled to a slider block 306, shown in Fig. 14. As shown in Figs. 12 and 13, the control pistons 305 have a small annular area 307 and a large annular area 308. System pressure is tapped off from the manifold via the two check valves 282 (noted above in connection with Fig. 11) or possibly a shuttle valve, and is fed continually to the small annular side 307 of both the pump and motor control pistons. The pressure acting on the small annular areas 307 exerts a control force on both the pump and the motor tending to stroke continually to their maximum displacement. System pressure is tapped off from the manifold 52 via the same two check valves 282, and is fed thru modulating valves, each having spool heads 309 on opposite ends of a valve spool 310, positioned inside of the control pistons 305. These modulating valves will either block flow to and from the large annular area 308 of the control pistons 305 or connect the large annular area 308 of the control pistons 305 to system pressure fed from the manifold, or vent pressure in the large annular area 308 to tank, depending on the position of the valve spool 310 relative to a spool sleeve 312. The spool sleeve 312 is fixed to the control pistons and moves axially with the control pistons, and the valve spools 312 are moved axially by stepper or servo motors 315. When the motor 315 moves the valve spool, the spool head 309 will move from its blocking position to either its venting or pressure feeding position depending on whether the spool is moved inward or outward to the control pistons. If the valve spool 310 is moved outward from the control piston then the valve spool is moved to its position in which it vents pressure from the large annular area 308 of the control piston 305, venting to tank. As pressure is continually fed to the small annular area 307 of the control piston, a force imbalance will be created such that the control piston 305 will move into its bore and pull the yoke 150 toward its maximum displacement position shown in Fig. 8. As the control piston 305 moves, the spool sleeve 312 moves with it, this motion being in the same direction that the spool 310 was moved by the stepper motor 315, and this motion will continue until the spool sleeve 312 reaches the position in which the spool head 309 blocks the hole 313 in the spool sleeve 312, where flow from the large annular area 308 will be blocked, thereby stopping the motion of the control piston 305. The control piston 305 will now be stationary with the pressure in the large annular area 308 of the control piston being at a ratio of the small annular area/large annular area multiplied by the system pressure. If the spool valve 310 is moved inward to the control piston then the spool valve is moved so that system pressure is fed to the large annular area of the control piston. System pressure will now be acting on both sides of the control piston but as there is an area difference there will be a force in-balance that will move the control piston out of its bore and cause the yoke 150 to move the hydrostatic unit toward its minimum displacement position. Again as the control piston 305 moves, the spool sleeve 312 moves with it and this motion being in the same direction that the spool 310 was moved, and this motion will again continue until the spool sleeve 312 reaches the blocking position with the spool 310, where flow from the large annular area will be blocked stopping the motion of the control piston. The control piston will now be stationary with the pressure in the large annular area of the control piston being at a ratio of the small annular area/large annular area multiplied by the system pressure.

The areas of the control pistons are selected so that the force in-balance created when the spool 310 is moved is large enough to overcome the control forces generated on the yokes 150 by the hydrostatic units 30, 35, as well to accelerate the pivoting masses so that adequate control times are achieved. When the spool 310 is in the blocking position and the control piston 305 is stationary, the system pressure acting on the small annular area 307 of the control piston 305 and the resultant pressure acting on the large annular area 308 of the control piston 305 generates enough holding force so that the control forces generated on the yokes 150 by the hydrostatic units can not stroke the hydrostatic units.

The transmission is designed to readily adapt to supplemental hydraulic circuits through access fittings 390 and 392. For example, hydraulic regeneration circuits are accessible to the hydraulic circuit in the transmission through these access fittings 390 and 392. Accordingly, separate and identical control piston, spool, stepper motor and associated components are provided for both the pump and motor hydrostatic units 30 and 35 to allow for individual control of displacements of the pump and motor so as to fully exploit the benefits of hydraulic brake energy recovery.

The configuration shown in Figs. 1 and 2 has been designed to optimize both the hydrostatic unit performance and the packaging requirements, to achieve a small lightweight transmission able to accommodate the highest power engines currently available for the light truck/full size sedan applications.

In order to achieve maximum hydrostatic unit efficiencies (and hence overall transmission efficiency) it is desirable to minimize the size of the rotating interfaces of the hydrostatic units as this will reduce both the torque loss and the leakage loss from the hydrostatic units as well as give size and weight advantages. To achieve this, it is helpful to reduce the size of, or eliminate, any shafts that pass through the center of the hydrostatic units as this will help to minimize the piston bore circle and hence the torque plate land diameters. For the same reason, it also helps to reduce the number of pistons used for a given displacement. Obviously as the number of pistons used decreases, so the smoothness of operation of the hydrostatic units decreases, therefore a compromise between acceptable operation and efficiency is sought. In this design seven pistons are used, as this has proven to offer a good compromise.

To reduce the rotating diameter of the hydrostatic units, the input and output shafts 50, 51 that transmit power to and from the hydrostatic units 30, 35 have been located at a position away from the center of the torque plate 110, unlike the conventional bent axis design, and power to and from the hydrostatic units is transmitted via the outside diameter of the torque plates 110 by means of a sprocket or gear. In the disclosed designs, a silent chain sprocket has been used, although a geared transfer could be used instead. The radial bearing 220 is placed in the center of the torque plate 110 for location as well as to support the radial load placed upon the torque plate 110. This radial bearing 220 is supported by the shaft 215 that is secured in the manifold 52. The axial center of the radial bearing 220 and the chain sprocket 105 is located coincident with the axial position of the center of the spherical piston heads 120 in torque plate so that there is no moment produced on the radial bearing 220 and torque plate 110 from any radial loads placed upon it from the either the hydrostatic unit pistons or the chain sprocket 105. Taking power from the outside as opposed to the inside of the torque plate not only gives the advantage of being able to keep the hydrostatic unit torque plate size to a minimum, but by careful angular orientation of the line of force from the chain (or gear) it is possible to use the radial force induced by the chain (or gear) to reduce the radial force induced by the hydrostatic unit pistons. This will reduce the resultant radial forces that are transmitted to ground which will enable the use of a smaller radial bearing to support this load, as well as reduce noise transmitted to ground from the hydrostatic units. A combination hydrostatic bearing supports the axial load on the torque plate. This combination hydrostatic bearing, shown in Figs. 9, 9A and 9B, is similar to that described in detail in U.S. Patent Application No. 10/311,983, entitled "Hydraulic Pump and Motor", now U.S. Patent No. 6,874,994 issued on April 5, 2006.

The device that is used to transmit power to and from the hydrostatic units 30, 35 (i.e. a chain sprocket 105 or gear) via the torque plate 110 is shown as a separate component from the torque plate 110. This enables the torque plate 110 to be made from a different material from that of the torque transmission device 105, therefore making it possible to manufacture these components from their ideal material, taking into consideration performance, durability, manufacturability and cost etc. In the design shown in Fig. 2, the chain sprocket 105 is connected to or integral with a retainer plate, which holds the piston heads in the spherical sockets in the torque plate. The retainer plate is pinned to the torque plate so that it can transfer torque to and from the torque plate to the retainer plate and hence the chain sprocket 105. The chain sprocket 105 or gear form can of course be a separate component from the torque plate and retainer plate, being splined or connected to the torque plate in a manner so that torque can be transmitted between it and the torque plate. It is also possible to have the chain sprocket 105 (or gear) be directly formed to the outside of the torque plate if material selection allows. It order to keep the transmission as small and light as possible, it is best to reduce the loading that all of the hydraulic components impart on their supporting structures, thereby reducing the required size and weight of these structures. By placing the hydrostatic units so that the torque plates 110 face each other across the manifold block 52, in a series configuration, the large axial force from the torque plates 110 cancel each other out and place the manifold block 52 mainly in compression. As the manifold block 52 is mainly under a compressive load, the manifold structure is inherently strong and stiff thereby reducing the size required to keep the manifold faces flat and deflection free, which affords the best performance of the combination hydrostatic bearing. The axial load placed upon the yokes 150 that support the hydrostatic units 30, 35 can be reacted from the pump yoke to the motor yoke by connecting the two yokes 150 together through the links 165. These links 165 are placed mainly in tension where they are inherently strong and stiff, thereby reducing the size of the structure taking this load. These links 165 are rigidly connected to the manifold block 52, but the only loads that are placed upon the manifold block 52 from the links are due to the imbalance of axial forces when the pump and motor hydrostatic units are at different displacements, and the radial loads that are induced from the yokes when the hydrostatic units are at angle other than zero degrees.

To further decrease the size of the hydrostatic units and reduce the size weight of the transmission, the hydrostatic unit displacement is reduced, so that under maximum transmission output torque conditions a maximum operating pressure of 5000 psi is reached. Using a maximum operating pressure of 5000 psi also has the added benefit of increasing the power density of any hydraulic storage devices used for supplemental hydraulic circuits such as regenerative brake energy recovery, if so incorporated. The flow to and from the hydrostatic units is passed through the hollow pistons 125 and the torque plate 110 to the manifold 52. An added benefit of placing the hydrostatic units in a series configuration is that the passages that carry the fluid in the manifold to and from the hydrostatic units, can now be relatively short and straight, thereby minimizing the flow losses through the manifold and increasing transmission efficiency.

When locating the power shaft apart from the center of the hydrostatic units, the best location for the axis of the hydrostatic units is in parallel to the input/output axis as opposed to concentric to the input/output axis. To achieve the optimum packaging shape and size for the intended application, the series hydrostatic units have been placed beneath the input/output centerline. To keep the depth of the transmission within acceptable limits, this means the assembly of hydrostatic units and manifold will now be below the normal sump height. To have the hydrostatic units rotating in the sump oil reservoir would cause undesirable windage losses and oil foaming. This can be prevented by several methods including:

Utilize a dry sump with an oil reservoir separated from the hydrostatic units and geartrain and use scavenge pumps to replenish the oil reservoir and keep the sump and hydrostatic units dry. Use of a baffle 400 that closely follows the contour of the hydrostatic units assembly to limit the amount of reservoir oil that comes into contact with the rotating components of the hydrostatic units. When the hydrostatic unit elements start to rotate, the oil that is in contact with them will be flung clear, evacuating the area between the baffle and the hydrostatic unit assembly. The baffle 400 is designed to allow this oil to return to the reservoir oil on the outside of the baffle and be de-aerated on the way. This method is the one utilized in the preferred embodiment of Fig. 2.

Make up pressure oil is fed to the manifold from make up pumps driven from the input shaft 50. Make up pressure is used to replenish system oil that leaks from the pump and the motor to the transmission sump via the various hydraulic interfaces, as well as to keep a positive pressure on the low pressure side of the flow passages to prevent cavitation. The makeup pressure is fed to the main flow passages in the manifold block 52 via the check valves 290 and 292 so that this oil will flow to the flow passage that is at the lower pressure.

The clutch 240 used to connect the motor torque through the tubular output shaft 230 to the output shaft 51 is energized by the makeup pressure generated by the make up pump. To keep the makeup pressure at a minimum, and thereby reduce parasitic losses from the make up pump and leakage from the hydrostatic units, the makeup pressure may not be high enough to prevent the clutch from slipping under high output torques. But in this kind of vehicle application, high output torques are used very infrequently, so it would waste too much energy to continually produce a higher makeup pressure purely to stop clutch slippage. For this reason a make up pressure boost valve 405 has been incorporated. This valve will increase the make up pressure when required by the higher output torques so as to supply enough force to the clutch to prevent the clutch from slipping. A PWM solenoid valve will take a makeup pressure that has been regulated down to a constant 50 psi and send this regulated pressure to act upon a piston in the pump pressure control valve. The PWM solenoid valve can send anything from 0 - 50 psi to this piston. The pump pressure control valve uses a constant mechanical spring force to control normal make up pressure, and this mechanical spring force can be augmented by force from the piston with in it, so when the 50psi regulated pressure is fed to act upon the piston in the pump pressure control the mechanical spring force will increase therefore increasing makeup pressure. By controlling the signal to the PWM valve it is possible to control the make up pressure infinitely between zero boost (normal make up pressure) and maximum boost pressure. The transmission controller will control the PWM solenoid valve so that the makeup pressure will be just high enough to stop the clutch from slipping at all output torques. The transmission controller can receive a signal from the engine controller indicating the current engine output torque, and as the transmission controller will know the transmission ratio it will be able to calculate the current transmission output torque. Once this is known the transmission controller will use a look up table to find what the signal to the PWM solenoid valve should be in order to prevent the clutch from slipping. As seen in Fig. 15, a shuttle valve 350, located in the manifold, connects the two main flow passages to the lubrication circuit. This shuttle valve 350 (also known as a flushing valve) is designed such that the flow passage at the higher pressure is blocked off from the lubrication circuit, and the flow passage that is under the lower pressure (i.e. make up pressure) is opened to the lubrication circuit. The lubrication circuit receives all of its flow from the flushing valve 350. This ensures that the flow passage that is under make up pressure has a continual flow of filtered cool oil from the makeup pump 56 that is at least equal to the lubrication flow rate. This will avoid heat build up that is possible in the manifold when the transmission is used for extended periods of time at a relatively low load and the leakage from the various hydrostatic unit interfaces is also correspondingly low, and the working fluid is not renewed often enough.

There are two ports in the manifold that connect to the two main flow passages, these ports are fed to the outside of the transmission case by connecting tubes. The ports can then be connected to an external circuit to gain direct access to the main flow passages, for use in a hydraulic energy recovery circuit, as well as other devices if desired. Operation

The operation of the transmission shown in Fig. 1 will be described in its several drive modes, using a shorthand notation of the planet gearsets 40 and 45 to indicate the number of teeth on meshing gears. This notation is explained on the following list:

Rop - output ring gear 65 Sop - output sun gear 70 Rp - pump ring gear 85 Sp - pump sun gear 71 SPp - pump chain sprocket 95

SPtp - pump torque plate sprocket 105 SPm - motor torque plate sprocket 105' SPop - output chain sprocket 225

Neutral:

When the transmission is at neutral and the vehicle is stationary, the main output shaft 51 , and hence the output ring gear 65, are stationary. The clutch 240 is disengaged. As the input shaft 50 and output planet carrier 60 are rotating at input speed, the connected sun gears 70, 71 and the pump planet set ring gear 85 will rotate and hence cause the pump to rotate at the ratio of:

Input Speed x (1 +(Rop/Sop)) x (Sp/Rp) x (SPp/SPtp) (eq1)

If the pump and motor hydrostatic units 30, 35 are at some displacement, there will be some pumping from the pump 30, which will cause the motor 35 to rotate. If this is the case, this will cause the motor shaft 230 to rotate, but as the clutch 240 is released, this will allow the motor hydrostatic unit 35 to rotate freely without transmitting any torque to the main output shaft 51.

Whilst the transmission is at neutral and before the clutch 240 is engaged, the transmission controller will ensure that the pump hydrostatic unit is at zero displacement and the motor hydrostatic unit is at maximum displacement. This will ensure that there will be no flow from the pump and hence no rotation from the motor. This can be achieved by several ways including (but not limited to): A speed sensor on the motor or the motor sprocket that will detect speed and rotation direction, and hence determining the pump HSU displacement; or, an angular position sensor on the pump and the motor HSU - the design shown incorporates both of these methods. Low Ratio (high torque multiplication):

When the transmission controller is signaled to go to Drive mode the clutch C1 will be engaged (once the controller has determined the pump HSU is at zero displacement).

Due to the planet set configurations, as described above, the input torque is split into two parallel paths, these being a direct mechanical path fed continually to the output shaft at the ratio of:

Input torque x (1/(1+(Sop/Rop))) eq2 and a 'hydraulic' torque path fed continually to the pump at the ratio of:

Input torque x (1/(1 +(Rop/Sop))) x (Rp/Sp) x (SPtp/SPp) eq3 When instructed by the transmission controller, the pump is stroked to give a small displacement. As the pump is rotating at the speed as described by eq1, pumping and hence fluid flow will take place. This fluid flow passes directly thru' the manifold to drive the motor. As the clutch C1 is now engaged, this motor rotation will cause the motor shaft and hence the main output shaft to rotate to give an output speed. Due to the fact that the pump is at a small displacement, a small amount of torque to the pump results in a high pressure and small flow rate, and as the motor is at a large displacement this high pressure and small flow rate results in a high output torque and low output speed at the main output shaft in the ratio of:

Motor Torque x (SPop/SPm) eq4 Due to the fact that the pump is at a small displacement it will react a torque from the input shaft as described in eq3. This torque on the input shaft will then be reacted directly to the main output shaft as described in eq2.

The high 'hydraulic' output torque as described by eq4 is added directly to the mechanical output torque as described by eq2. Therefore the total output torque can be expressed as:

[Input torque x 1/(1+(Sop/Rop))]+ [Input torque x 1/(1+(Rop/Sop)) x (Rp/Sp) x (SPtp/SPp) x (SPop/SPm) x (motor disp/pump disp)] eqδ

It can therefore be seen that there is a total output torque comprising of a fixed mechanical torque plus a variable hydraulic torque, and as the motor displacement to pump displacement ratio decreases, the amount of hydraulic torque decreases, and if the pump or motor displacement equals zero then there is no hydraulic torque, just mechanical torque.

As pump displacement increases, flow rate from the pump increases, and this increased flow causes the motor and hence the output planet set ring gear and main output shaft to increase in speed. As the main output shaft increases in speed relative to the input shaft speed, the output ring gear speed increases relative to the output planet carrier speed; this causes the output and pump sun gears speed to decrease. This has the effect of causing the pump speed to decrease as the transmission ratio increases and the main output shaft increases relative to input speed. This has the effect of reducing the total system flow rate, (when compared to a conventional hydrostatic transmission of the same capacity), to approximately V₃ to % depending on planet set ratios used. This reduces the flow losses and noise levels normally associated with pure hydrostatic machines.

Under the normal control regime, as the transmission approaches final ratio, the motor approaches zero displacement whilst the pump approaches maximum displacement. Therefore the pump speed will approach zero as motor speed approaches its maximum.

Overdrive or final ratio:

When the motor reaches zero displacement it can no longer accept fluid flow, so the pump can no longer displace fluid and therefore stops rotating. This causes the pump and output sun gears to stop rotating. The pump is now acting as a reaction unit for output sun gear and now all of the input torque is transferred thru' the output planet set to the output shaft. Due to the ratio of the output sun gear to the output ring gear" the output speed is increased and the output torque decreased give a final overdrive ratio of: Input speed x (1 + (Sop/Rop)) eqrβ As the pump has been stroked to its full displacement, hydraulic pressure required to react the input torque has been reduced to a minimum, thus reducing hydraulic leakage losses and hydraulic loading of bearings to a minimum.

As the motor is now at zero displacement it is no longer adding any torque to the output, and as stated above, the motor is now rotating at its maximum speed. This will create an efficiency loss, as the motor is now being 'driven' by the main output shaft and will subtract torque from the output shaft. In order to reduce this torque loss the clutch C1 can be dis-engaged. This will allow the motor to stop rotating as the torque loss from the motor spinning will be greater than the drag torque thru the open clutch. As all of the input power is now transferred thru' the output planet set, and the hydraulics are acting only as a reaction unit to hold the sun gears, the efficiency of the transmission will be very high (97+%), as the only losses are the normal gear set losses (approx. 1%), slippage on the pump due to leakage and windage losses thru' the open clutch and makeup pump losses. To further increase the efficiency at this point, a brake could be applied to the pump. This will stop the pump HSU from rotating due to hydraulic leakage and hence reduce speed loss at the output shaft. Reverse:

Starting from the same conditions as in neutral, with the motor at its maximum displacement and the pump at zero displacement: the controller will stroke the pump in the opposite direction (i.e. to a negative angle) causing fluid flow to go in the opposite direction. This will cause the motor and hence the output shaft to rotate in the reverse direction. Due to the planet set gear configuration, the mechanical torque (as described in eq2) is still acting in the forward direction. Therefore the total output torque, in reverse, can be expressed as:

[Input torque x 1/(1+(Sop/Rop))] - [Input torque x 1/(1+(Rop/Sop)) x (Rp/Sp) x (SPtp/SPp) x (SPop/SPm) x (motor disp/pump disp)] eq7

Maximum Torque Limitation:

Due to the fact the motor to pump displacement ratio can be infinitely large, at or around the neutral zone in forward and reverse, it is possible to generate infinitely high pressures and output torques. Obviously these have to be limited to reasonable values, as determined by the structural limitations of the transmission. This can be achieved in several ways, including but not limited to:

The use of a pressure relief valve that will limit the maximum pressure that the pump can generate, and hence the maximum output torque. Since the pump will be at relatively small displacements, the flow rate thru the relief valve will be at reasonable levels.

The system can be designed to be self regulating; by designing the pump and motor to have a leakage rate (which is necessary for hydrostatic bearing interface cooling and lubrication) which at a specified pressure is equal to the pump discharge. This will prevent the pump from generating a higher pressure than this. The transmission will then reach a 'stall' torque.

With today's sophisticated electronic engine controls it is possible to limit engine torque when the pump is at small displacements so as to regulate the maximum system pressure.

There is a minimum pump angle at which the pump can react full input torque without exceeding the maximum allowable system pressure, and hence generate maximum output torque. At pump angles less than these, the output torque will not increase as the maximum pressure is limited by methods as described above, but the input to output speed ratio will continue to decrease and will approach infinity as the pump angle becomes infinitely small. The hydromechanical continuously variable transmission disclosed in Fig. 2 has many advantages over existing transmission designs including: Taking power to and from the hydrostatic units via the outside diameter of the torque plate enables the hydrostatic units rotating diameters to be kept as small as possible, as well as allowing for a small number of pistons to be used. This increases the hydrostatic units efficiency as both torque loss and leakage loss across the HSU increase as the rotating diameters increase.

Taking power to and from the hydrostatic units from the outside diameter of the torque plate also allows for the hydrostatic units to be placed in parallel to the input/output axis and facing each other in series. Placing the hydrostatic units in series enables the pump forces to counteract the motor forces, significantly reducing the resultant forces that are exerted to the supporting structure and transmission case. This allows the transmission to be as small and light as possible whilst being able to handle high powers.

Using a chain (or gear) to drive the outside diameter of the torque plates enables the radial reaction load from the chain or gear to counteract the radial load from the pistons acting on the torque plate. This reduces the force on the bearing that radially locates the torque plate, therefore reducing the size of this bearing and its supporting structure.

By making the torque input/output members of the hydrostatic units separate from the torque plates allows for different materials to be used for the chain sprocket/gear and the torque plates. It is therefore possible to manufacture these components from their ideal material, taking in to consideration performance, durability, manufacturability and cost etc.

By connecting the pump and motor hydrostatic units in series as described above, the pump and motor rotation directions are such that both the high and low pressure flows are directly inline with each other between the pump and motor. This ensures that the flow passages are as short and as straight as possible, thereby reducing flow losses and maximizing hydraulic efficiency.

The motor is connected to the output shaft via a clutch, so that power from the motor can be disengaged to the transmission output. This is beneficial for several reasons including : At initial start up in neutral the clutch will be released, so that when the pump rotates at a ratio of input speed, and if it has moved away from zero displacement during rest, any subsequent rotation of the motor will not cause the vehicle to leap forward or backward unexpectedly. The clutch will only be applied when the transmission is placed in Drive mode and the controller receives a signal from a sensor that the pump and motor are at their correct displacements. As the pump is stroked to give some displacement and hence some output speed and torque, the clutch can be modulated to give some slip to allow for a smooth start, in the same manner in which a clutch is slipped in a regular manual transmission during vehicle launch. This will eliminate any jerking 'kangaroo' takeoffs common with previous hydrostatic transmission designs.

To further increase the transmission efficiency, the clutch can be released when the transmission is at final ratio and the motor is no longer adding any power to the transmission output. As the released clutch will have less drag torque than the rotating motor, the motor will come to rest and the parasitic losses will be reduced. The pump can be fixed to ground by a releasable brake so that when the brake is activated when the CVT is at final ratio the pump can not rotate due to reaction torque generating pressure and hence causing leakage thru the various HSU interfaces. This will further increase efficiency at this ratio.

By adjusting the makeup pressure that controls the clutch and or brake so that the minimum pressure is used to prevent slippage as the torque through these devices changes, will limit the parasitic losses of the makeup pump system, and hence increase efficiency through the ratio range of the CVT.

Placing the output and pump planet sets within the same axial space occupied by the manifold, optimizes the use of axial space and keeps the overall length of the transmission to a minimum. It also reduces the lengths of the input and output shafts, which is an important manufacturing benefit.

Separating the gear housing from the manifold allows for the gear housing to be manufactured from a material different from that of the manifold. This allows for optimal material selection for these two components, taking into account their required structural weight, manufacturing processes and cost etc. This also allows the gear housing to be made from a material with good sound dampening coefficients, such as magnesium for example, to help in the reduction of hydraulic noise transmission from the hydraulic assembly to the transmission case.

As the manifold and the hydraulic assembly are separate to the gear housing, all of the high pressure oil passages are contained within the hydraulic assembly and do not need to be passed into the gear housing or transmission case. This reduces structural integrity requirements of the gear housing and transmission case as well as reduces hydraulic noise transmitted to these components.

An alternate geartrain schematic for a transmission in accordance with this invention, shown in Fig. 16, .has the input shaft 50 connected to a large sun gear 502 of a double sun planetary gearset 500 and the output is connected to the small sun gear 505 of the double sun planetary gearset 500. The planet carrier 507 of the double sun planetary gearset contains a compound planet gear arrangement where the large planet gear 510 meshes with the small sun gear 505 and the small planet gear 512 meshes with the large sun gear 502. The planet carrier 507 of the double sun planetary gearset is connected to a ring gear 515 f a simple pump planetset 520. The carrier 522 of the pump planet set 520 is fixed to ground and the sun gear 525 of the pump planet is connected to the pump drive sprocket 530. The motor 35 is connected to the output shaft as described in connection with Fig. 1 , and the pump / motor displacement and speed and all other systems will also act as previously described in connection with Fig. 1. The operation of the transmission shown in Fig. 16 will be described in its several drive modes, using a shorthand notation of the planet gearsets pump and output gearsets to indicate the number of teeth on meshing gears. This notation is explained on the following list: dsg1 - small sun gear 505 dsg2 - large sun gear 502 dsp1 - large planet gear 510 dsp2 - small planet gear 512 rg1 - ring gear 515 of a pump planetset sg1 - sun gear 527 of a pump planetset

In neutral the output shaft 51 is stationary and the carrier 507 of the double sun planetary gearset 500 will rotate at a ratio of:

Input speed x (dsg2 x dsp1) / [(dsg2 x dsp1) - (dsg1 x dsp2)] in the same direction.

As the carrier of the double sun planetary gearset 500 is connected to the ring gear 515 of the pump planetset 520, and the carrier 522 of the pump planetset 520 is fixed to ground, the sun gear 527 of the pump planet set 520 and hence the pump drive sprocket 530 will rotate at the ratio of:

Input speed x(dsg2 x dsp1) / [(dsg2 x dsp1) - (dsg1 x dsp2)] x (rg1 / sg1) in the opposite direction.

When the transmission is at final ratio, the pump 30 will be at max displacement and the motor 35 is at zero displacement, so the pump 30 and the pump drive sprocket 530 will be stationary, as described previously. As the pump drive sprocket 530 is connected to the sun gear 527 of the pump planet set 520, and the carrier 522 of the pump planetset 520 is fixed, this will have the effect of locking the pump planet set ring gear 515 and hence the planet carrier 507 of the double sun planetary gearset 500. The output speed will now rotate at an overdrive speed of:

Input speed x (dsg2 x dsp1) / (dsg1 x dsp2)

The advantage of this geartrain as opposed to the double simple planet set shown in Fig. 1 is that the planet speeds at and around neutral and into reverse are a lot slower and the pump speed at these ratios will be more or less equal to input speed. This will alleviate any issues that may exist with planet gear bearing speeds and with gear noise that may exist at these ratios. It will also alleviate the need to limit engine speed at these ratios. Also the double sun planetary gearset 500 has no ring gear which is typically the most expensive part of a planetset. The disadvantage of this type of arrangement is that the loads on the planet gears of a double sun planetary gearset are not as balanced as they are in a simple planet set requiring special attention to thrust bearing design for these planet gears. Also the cost of producing a compound planet gear may offset the reduction in cost due to the lack of a ring gear. Another embodiment, shown in Figs. 18-20, replaces the clutch 240 with an additional of planet set and brake assembly 550. This puts a gear ratio between the motor 35 and the output shaft 51 to increase the torque coming from the motor. This makes the transmission suitable for high torque applications, such as large suv, light trucks, etc .where a top speed of no more than 100 mph is anticipated, but due to their heavier weight, more output torque is desired. The rest of the transmission is identical, except as noted, for the back end where the clutch 240 is replaced with the brake and planetset assembly 550.

The different rear end of this third embodiment is shown using another planet set and brake assy instead of the clutch assy. This variation makes use of an advantage of the design shown in Fig. 2, namely, that it is possible to change the rear 'module' of the transmission, replacing the clutch assy with this new planetset and brake assembly, so as to make a high torque low speed transmission that is more suitable in vehicles requiring more torque. In this design shown in Figs. 18-20, it is preferable to limit the maximum transmission output speed because the rear end gearset speeds up the motor relative to the output speed (at the time multiplying motor torque relative to the output shaft) it is well to keep the speed of the motor well within its maximum speed limit. The way the gearset is arranged in Fig. 18, the speed and torque multiplication factor is about 1.5 from the motor to output shaft. Everything else in the transmission is similar to the embodiment shown in Fig. 1.

A similar effect could be achieved by just changing sprocket ratio on the motor chain, but this would cause the center distance between the hydrostatic units and the planetsets to increase due to the larger sprocket, which would increase the depth of the transmission from the centerline to the bottom of the pan. This greater depth could be undesirable in some applications. It would also increase the torque requirement of the clutch by a factor of 1.5. The way the gearset is arranged in Fig. 18, the sun gear is held to ground by a releasable brake and the sun gear sees only approx half of the motor torque.

By way of an overview of some of the salient features of the continuously variable transmissions disclosed herein, the following list is presented. Naturally, there will be many applications that do not need to use all of these features, or would use some modification of these features. Therefore, the inclusion of this list should not be interpreted as a limitation to the claims in which this invention is defined.

Hydrostatic Unit Related Features:

Torque plate has overbalance grooves inside of the main kidney slots Torque plate has overbalance grooves outside of the main kidney slots

Hydrostatic units placed facing each other so that the axial force of the pump is counteracted by the axial force of the motor, placing the manifold in compression.

Individual pucks are used to support the axial load from the cylinder block to the cylinder block support structure Individual pucks that have an overbalance hydrostatic bearing using an throttling device so that the leakage flow rate across the bearing causes a pressure drop in the recess area so that the separating force in the recess area is equal to the clamping force from the pressure acting over the piston area, placing the puck in equilibrium or near- equilibrium. Individual pucks that have an underbalance hydrostatic bearing where the separating area and force is less than that of the clamping area and force.

Power is transferred to and from the each hydrostatic unit by means of a gear or chain sprocket that is positioned around the outside of the torque plate.

The axial center of the sprocket is positioned such that the centerline of the chain is co-incident (or near co-incident) with the center of the piston spheres and the hydrostatic unit articulation center in the torque plate. A chain is connected to the chain sprocket on the torque plate and is orientated such that the radial force of the chain is used to counteract the radial force that is generated by the pistons acting on the torque plate.

The torque plate is radially supported by a bearing that is positioned at the radial center of the torque plate.

The torque plate is radially supported by a bearing that is positioned such that the center of the bearing is co-incident (or near co-incident) with the center of the piston spheres and the hydrostatic unit articulation center on the torque plate.

Geartrain Related Features

A simple planetset is used to obtain a power split from the input power path so as to generate a parallel power path from the input to the output and to the pump hsu.

The power transferred to the pump is transferred to hydraulic power which is used to drive the motor which then transfers this hydraulic power to the output. The motor is connected to the output by a releasable clutch.

In another embodiment the motor is connected to a planetset that is connected to the output in such a manner that torque from the motor is multiplied as it is transferred to the output.

One member of the above planet set is connected to ground via a releasable brake. The hydrostatic sub-assembly is connected to the main transmission case via a separate support structure (such as the gear housing) to isolate noise from the hydraulic sub assembly from being transmitted to the main transmission case. ,

The planetary geartrain sub-assembly is supported by a separate support structure (such as the gear housing) that is connected to the main transmission case to isolate noise from the planetary geartrain sub assembly from being transmitted to the main transmission case.

Controls Related Features

A valve is used so that makeup supply will flow thru whichever of the two main flow channels is at low pressure to the lubrication and or cooling circuit.

System pressure is captured from whichever circuit is at the higher pressure and used to actuate a control actuator for hydrostatic unit displacement control.

An individual control actuator is used to control the displacement of each pump and motor hydrostatic unit. There are ports that connect the two main flow passages to connections that are external to the transmission to allow for an external hydraulic source to flow in and out of the main flow circuits for supplemental hydraulic circuits such as hydraulic regeneration connection etc.

Obviously, numerous other modifications, combinations and variations of the preferred embodiments described above are possible and will become apparent to those skilled in the art in light of this specification. For example, many functions and advantages are described for the three preferred embodiments, but in some uses of the invention, not all of these functions and advantages would be needed. Therefore, we contemplate the use of the invention using fewer than the complete set of noted functions and advantages. Moreover, several species and embodiments of the invention are disclosed herein, but not all are specifically claimed, although all are covered by generic claims. Nevertheless, it is our intention that each and every one of these species and embodiments, and the equivalents thereof, be encompassed and protected within the scope of the following claims, and no dedication to the public is intended by virtue of the lack of claims specific to any individual species. Accordingly, it is expressly intended that all these embodiments, species, modifications and variations, and the equivalents thereof, are to be considered within the spirit and scope of the invention as defined in the following claims, wherein we claim:

Claims

1. A transmission for converting rotating mechanical power at one combination of rotational velocity and torque at an input shaft to another combination of rotational velocity and torque at an output shaft over a continuous range, comprising: an input hydrostatic unit and an output hydrostatic unit hydraulically coupled together with fluid passages so that pressurized working fluid from said input hydrostatic unit drives said output hydrostatic unit, and spent fluid from said output hydrostatic unit recharges said input hydrostatic unit for being pressurized therein; each of said hydrostatic units having cylinders in a cylinder block mounted for rotation about an axis for rotation and for pivoting about a pivotal axis relative to a stationary manifold block through which said fluid passages extend; said cylinders each having a piston therein, and each piston having a piston head thereon protruding from an open axial end of said cylinder on a first axial end of said cylinder block; each of said hydrostatic units includes a torque plate having sockets which receive said piston heads, said torque plate being supported between said manifold and said cylinder block to rotate with said cylinder block about a torque plate center of rotation against a face of said manifold, said torque plate having kidney slots communicating with said sockets in said torque plate that convey fluid through said torque plate between said pistons and said manifold; an epicyclic gear train through which said hydrostatic units are mechanically coupled, said epicyclic train including an epicyclic gear set that includes first and second planet gear sets coupled together such that said first planet gear set is coupled to said input hydrostatic unit, and said output hydrostatic unit is coupled to said second planet gear set and said output shaft, and said input shaft is drivingly coupled to said second planet gear set.

2. A transmission as defined in claim 1 , further comprising: a third planet set is coupled between said output hydrostatic unit and said output shaft; a ground member for grounding an element in said third planet set; and a releasable brake for grounding said ground member, thus enabling said ground member to rotate freely when said brake is released such that said output hydrostatic unit is mechanically isolated from output torque reaction when said brake is released.

3. A transmission as defined in claim 1 , wherein: said first planet set is coupled to a double planet gear engaged with a double sun gear of said second planet set; wherein said input shaft drives one of said double sun gears to drive a first of said double planet gears, and the other of said double planet gears drives said output shaft by way of said second sun gear of said second planet set, and the planet carrier of said second planet set is coupled to said first planet set.

4. A transmission as defined in claim 1 , further comprising: a sprocket around an outside diameter of each of said torque plates for transmitting torque to and from said torque plates to and from said planet sets to provide said mechanical coupling between said hydrostatic units.

5. A transmission as defined in claim 4, further comprising: a mechanical drive train between said torque plate sprockets and sprockets on each of said planet sets, by way of which mechanical power is transmitted between said hydrostatic units and said planet sets.

6. A transmission as defined in claim 4, wherein: said sprocket is a separate component from said torque plate and is drivingly connected to said torque plate.

7. A transmission as defined in claim 4, wherein: said sockets in said torque plate of each hydrostatic unit are in a circular array around said torque plate, said circle defining an axial centerline of said array of sockets; said sprocket has an axial centerline that is located at the axial centerline of said array of spherical sockets.

8. A transmission as defined in claim 1 , further comprising: a torque plate bearing that is located at the center of said torque plate for supporting said torque plate for rotation about said torque plate center of rotation.

9. A transmission as defined in claim 8, wherein: said torque plate bearing has an axial centerline that is located on said axial centerline of said spherical sockets.

10. A transmission as defined in claim 5, further comprising: a bearing located at the center of said torque plate for supporting said torque plate for rotation about said torque plate center of rotation; said cylinder blocks each have a pivotal axis that is so orientated that radial forces from said pistons acting upon said torque plate are counteracted by separating forces from said mechanical drive train on said sprocket thereby reducing radial force reacted by said radial bearing.

11. A transmission as defined in claim 1 , further comprising: an underbalance hydrostatic bearing between each said spherical piston head and a respective torque plate socket, said hydrostatic bearing including an internal annular spherical area that is subjected to full pressure from the respective cylinder bore receiving said piston, said annular spherical area having a bottom that terminates in a blind hole communicating with a kidney slot of said socket on the face of said torque plate facing said manifold; a spherical sealing land between the said annular spherical area and said face of said torque plate; a small annular balance groove in said torque plate socket between the said annular spherical area and the said spherical sealing land area to convey full pressure from the respective cylinder bore acts over the said internal annular area, said respective kidney slot opening in the bottom of the torque plate socket communicating with the said annular groove so that any pressure that exists in the said torque plate kidney slot is communicated to the said annular groove; whereby said internal annular area is subjected to full pressure from both outside by said groove and inside from said blind hole that communicates with said kidney slot causing full pressure to migrate over said annular spherical area such that said annular spherical sealing land will drop the full system pressure that exists in the said annular area to atmospheric pressure at the external face of the said torque plate.

12. A transmission as defined in claim 11 , further comprising: a hydrostatic bearing for supporting said torque plate on said manifold face, said hydrostatic bearing having orifices in said torque plate that feed hydraulic pressure to overbalance grooves on said face of said torque plate, said orifices are fed hydraulic pressure via a hole that is connected to said annular balance groove.

13. A transmission as defined in claim 1 , further comprising: a hydrostatic bearing for axially supporting said torque plate on said manifold face, said hydrostatic bearing having a combination underbalance and overbalanced hydrostatic bearing, including overbalance grooves in said manifold-side face of said torque plate positioned inside of said kidney slots.

14. A transmission as defined in claim 1 , wherein: said torque plate is axially supported by a combination underbaiance and overbalanced hydrostatic bearing, and including balance pads on said manifold-side face of said torque plate located outside of said kidney slots.

15. A transmission as defined in claim 1 , further comprising: a cylinder support yoke for each hydrostatic unit, each yoke being articulated around said pivotal axis; said cylinders in said cylinder blocks are thru formed bores, and fluid is contained in said cylinders at the opposite side from said piston heads by individual pucks; said pucks having a hydrostatic overbalance bearing that runs against said cylinder support yoke, such that axial force that is generated on said puck by hydraulic pressure that exists in its cylinder bore is supported by said cylinder support yoke.

16. A transmission as defined in claim 15, wherein: axial force that is generated on said puck by hydraulic pressure that exists in said cylinder bore is further supported by a hydrostatic underbalance bearing on said puck that runs against said cylinder support yoke.

17. A transmission as defined in claim 15, further comprising: a spring in contact with said puck on one end and in contact with said cylinder block on the other, thereby forcing the puck away from the cylinder block and into firm engagement with said supporting yoke with an initial preload.

18. A transmission as defined in claim 1, further comprising: a cylinder support yoke for each hydrostatic unit, each yoke being articulated around said pivotal axis; said cylinders in said cylinder blocks are thru formed bores, and fluid is contained in said cylinders at the opposite side from said piston heads by individual pucks; a spring compressed between said puck on one end and said cylinder block on the other, said spring thereby applying a force to push said puck away from the cylinder block and into firm engagement with said supporting yoke with an initial preload.

19. A transmission as defined in claim 18, further comprising: a center piston having a piston head held in a spherical bearing located at said pivotal axis so as to axially locate said cylinder block relative to said torque plate.

20. A transmission as defined in claim 18, further comprising: said sockets in said torque plate of each hydrostatic unit are in a circular array around said torque plate, said circular array having an axial centerline of said array of sockets; a center piston that is located in a central bore of the cylinder block and having a spherical head that is held in a spherical cup that is supported by said torque plate such that the spherical center of said cup is located on the rotational and pivotal axis of said cylinder block and is also coplanar with the centers of the spherical sockets in the\torque plate and on said axial centerline of said array of sockets; said center piston holding said cylinder block is centered on said rotational and pivotal axis and reacting said force that is applied to said cylinder block from said spring that preloads said puck while applying a preload to said torque plate via said center piston.

21. A transmission as defined in claim 1 , further comprising: a cylinder support yoke for each hydrostatic unit, each yoke being articulated around said pivotal axis; said cylinder blocks are each supported on said yoke for articulating therewith about said pivotal axis to give displacement, and for rotation about said rotational axis of said cylinders; support of rotation of said cylinder blocks about the rotational axis thereof is provided by radial bearing that is positioned between a shaft that is connected to said yoke and a center bore of said cylinder block such that said radial bearing accommodates rotational speed between said cylinder block and yoke, and said radial bearing allows for relative axial motion between said cylinder block and said yoke.

22. A transmission as defined in claim 21 , wherein: said yokes each include two arms that straddle said cylinder block and are pivotally connected by way of two radial bearings to two links fastened to lateral sides of said manifold block, such that articulation of said yoke about said pivotal axis affords control of fluid displacement of said cylinder block as said cylinder block rotates, and said links carry substantially all longitudinal separating forces exerted by said cylinder blocks on said yokes.

23. A transmission as defined in claim 1 , further comprising: tubes protruding into high and low pressure flow passages in said manifold to allow external access to high and low pressure flows from said hydrostatic units for supplemental hydraulic circuits.

24. A transmission as defined in claim 1 , wherein: said spherical sockets that are formed in said torque plates have a cylindrical section before the spherical section that is close fitting to the outside diameter of the piston head ball to reduce leakage past the piston ball if it were to become unseated from the socket.

25. A transmission as defined in claim 1 , further comprising: a main housing containing said hydrostatic units and said epicyclic gear train, and a separate gear housing enclosing said an epicyclic train set for noise and vibration suppression, said separate gear housing being connected to said main housing and to said manifold block.

26. A transmission as defined in claim 1 , further comprising: a valve located between high and low pressure flow paths of said hydrostatic units, said valve communicating said low pressure flow path to a lubrication and cooling circuit such that flow for the cooling and lubrication circuit is supplied by a make-up pump thru a low pressure side of the hydrostatic unit flow path.

27. A transmission as defined in claim 1 , further comprising: two separate control actuators for individually controlling said hydrostatic units, said actuators including hydraulic actuators that use high system pressure fluid supplied by the hydrostatic units via a valve configured to supply flow from whichever flow passage of said hydrostatic units is at a higher pressure.

28. A transmission as defined in claim 27, wherein: said actuator includes a control piston with a first annular area and a second annular area, wherein said first annular area is larger than said second annular area and, during operation, said second annular area is continually supplied with high system pressure from said valve and said first large annular area is selectively supplied with high system pressure or vented to a sump tank via a control valve to control the axial position of said control piston.

29. A transmission as defined in claim 28, wherein: said control valve is a leader/follower spool type valve that is actuated by an electro mechanical actuator such that, as a spool of said spool type valve is moved, said spool uncovers a port to said first annular area of said control piston and selectively connect said first annular area to high system pressure or vent said first annular area to said tank, depending on which direction said spool is moved relative to said control piston; whereby, when said spool valve is moved in the direction that opens said first annular area to said high system pressure, axial force generated on said first annular area overcomes force on said second annular area causing said control piston to move in the same direction that said spool valve is moved and relative to said spool valve until said port is closed off, and when said spool valve is moved in the direction that vents said first annular area to tank, relieving force generated on said first annular area, force on said second annular area causes said control piston to move in the same direction that said spool valve is moved and relative to said spool valve until the said port is again closed off.

30. A transmission as defined in claim 29, wherein: said leader/follower spool valve is positioned co-axially within said control piston

31. A transmission as defined in claim 1 , further comprising: a cylinder support yoke for each hydrostatic unit, each yoke being articulated around said pivotal axis; said cylinder blocks are each supported on said yoke for articulating therewith about said pivotal axis to give displacement, and for rotation about said rotational axis of said cylinders; two separate control actuators for individually controlling said hydrostatic units, said actuators each including a hydraulic actuator having a control piston in a cylinder that is acted upon by high system pressure fluid supplied by said hydrostatic units via a valve configured to supply flow from whichever flow passage of said hydrostatic units is at a higher pressure; said actuators each include a slider block located in a groove in said control piston such that, as said control piston moves axially, said slider block moves axially correspondingly while being free to move perpendicular to the axis of said control piston; said slider block is pivotally coupled to said support yoke so that, as said control piston and hence said slider block moves axially, said slider block moves radially around said pivot axis causing said support yoke to articulate around said pivot axis, and wherein differential motion between said slider block and said control piston caused by the said control piston moving axially and said slider block moving radially is accommodated by said slider block moving transverse to said control piston axis in said groove.

32. A transmission as defined in claim 31 , wherein: said two hydrostatic units are positioned in a series configuration on opposite sides of said manifold block and said yokes are coupled together across said manifold block and said torque plates face each other across said manifold block, whereby large axial forces from said torque plates and said yokes cancel each other out and place the manifold block mainly in compression.

33. A transmission as defined in claim 1 , wherein: fluid flow to and from said hydrostatic units passes through hollow bores in said pistons and through said torque plate to said manifold.

34. A transmission as defined in claim 1 , further comprising: a baffle around said hydrostatic units separates reservoir oil from coming into contact with rotating components of said hydrostatic units, such that oil that is in contact with said hydrostatic units when said hydrostatic unit elements start to rotate will be flung clear, evacuating regions between said baffle and said hydrostatic units, and said oil flung clear of said hydrostatic units is free to drain to a oil reservoir outside of said baffle.

35. A transmission as defined in claim 1 , further comprising: a releasable brake connected to said input hydrostatic unit so that, when said brake is activated, said input hydrostatic unit is fixed to ground and cannot rotate.

36. A transmission as defined in claim 35, wherein: said releasable brake is actuated by hydraulic pressure from a makeup pump via an electro-hydraulic valve, and whereby makeup pressure from said makeup pump is elevated when required to provide enough clamping force on said clutch to prevent slippage as the torque thru the brake demands.

37. A process for converting rotating mechanical power in a vehicle engine output shaft at one combination of rotational velocity and torque to another combination of rotational velocity and torque in a transmission output shaft over a continuous range, comprising: coupling a pump and motor through an epicyclic gear set; starting said vehicle in neutral, and with a clutch connected to said motor so that, when said clutch is released, said motor is isolated a transmission output shaft, such that at initial start up in neutral with said clutch released, when said pump rotates, and it has moved away from zero displacement during rest, the vehicle does not leap forward or backward unexpectedly; engaging said clutch only when said transmission is placed in Drive mode and a controller receives a signal from a sensor that said pump and motor are at their correct displacements; stroking said pump to give some displacement and hence some output speed and torque; modulating said clutch to give some slip to allow for a smooth start and eliminate any jerking 'kangaroo' takeoff; and releasing said clutch when said transmission is at final ratio, thereby allowing said motor to come to rest when it is not adding any power to the output, and thereby providing increased transmission efficiency.

38. A transmission for converting rotating mechanical power in an input shaft at one combination of rotational velocity and torque, to another combination of rotational velocity and torque in an output shaft over a continuous range, comprising: a hydraulic pump, operatively driven by an input shaft, and a hydraulic motor operatively driving an output shaft, said hydraulic pump and said hydraulic motor coupled mechanically through a pair of planet sets and coupled hydraulically through a manifold, such that hydraulic fluid pressurized by said pump drives said motor and spent fluid from said motor is cycled back to said pump to be pressurized therein; said planet sets both arranged co-axially with said input shaft and said output shaft, and said a hydraulic pump and hydraulic motor arranged in series with each other on opposite sides of said manifold, and parallel to said input and output shafts, thereby optimizing the use of space and keeping the overall length of said transmission to a minimum, and minimizing required lengths of said input and output shafts.

39. A transmission as defined in claim 38, wherein: said hydraulic couping of said pump and motor through said manifold is direct straight flow channels, and rotation directions of said pump and motor are such that both the high and low pressure flows are directly inline with each other between the pump and motor.