WO2006108109A2 - Transmission a variation continue hydromecanique - Google Patents

Transmission a variation continue hydromecanique Download PDF

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Publication number
WO2006108109A2
WO2006108109A2 PCT/US2006/012921 US2006012921W WO2006108109A2 WO 2006108109 A2 WO2006108109 A2 WO 2006108109A2 US 2006012921 W US2006012921 W US 2006012921W WO 2006108109 A2 WO2006108109 A2 WO 2006108109A2
Authority
WO
WIPO (PCT)
Prior art keywords
transmission
torque
hydrostatic
torque plate
pump
Prior art date
Application number
PCT/US2006/012921
Other languages
English (en)
Other versions
WO2006108109A3 (fr
Inventor
Lawrence R. Folsom
Clive Tucker
Original Assignee
Folsom Technologies, Inc.
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Folsom Technologies, Inc. filed Critical Folsom Technologies, Inc.
Publication of WO2006108109A2 publication Critical patent/WO2006108109A2/fr
Publication of WO2006108109A3 publication Critical patent/WO2006108109A3/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H47/00—Combinations of mechanical gearing with fluid clutches or fluid gearing
    • F16H47/02—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the volumetric type
    • F16H47/04—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the volumetric type the mechanical gearing being of the type with members having orbital motion
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H37/00—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00
    • F16H37/02—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings
    • F16H37/06—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts
    • F16H37/08—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing
    • F16H37/0833—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths
    • F16H37/084—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths at least one power path being a continuously variable transmission, i.e. CVT
    • F16H2037/0866—Power split variators with distributing differentials, with the output of the CVT connected or connectable to the output shaft

Definitions

  • 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.
  • 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.
  • 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. 1 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 .
  • 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
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 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.
  • 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.
  • 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.
  • Control actuator 300 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • the spool head 309 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • hydraulic regeneration circuits are accessible to the hydraulic circuit in the transmission through these access fittings 390 and 392.
  • 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.
  • Figs. 1 and 2 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • the hydrostatic units 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 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
  • 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:
  • 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:
  • 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.
  • 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.
  • 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.
  • a clutch 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.
  • 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.
  • 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.
  • the gear housing 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.
  • FIG. 16 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.
  • Input speed x (dsg2 x dsp1) / [(dsg2 x dsp1) - (dsg1 x dsp2)] in the same direction.
  • the sun gear 527 of the pump planet set 520 and hence the pump drive sprocket 530 will rotate at the ratio of:
  • the pump 30 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.
  • the pump drive sprocket 530 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:
  • Figs. 18-20 replaces the clutch 240 with an additional of planet set and brake assembly 550.
  • 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.
  • 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.
  • 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.
  • 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 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.
  • 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.
  • 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.
  • a separate support structure such as the gear housing
  • 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.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Motor Power Transmission Devices (AREA)
  • Reciprocating Pumps (AREA)

Abstract

L'invention concerne une transmission de puissance à variation continue hydromécanique destinée qui transforme une puissance mécanique de rotation selon une combinaison de vitesse et un couple de rotation en une autre combinaison de vitesse et de couple de rotation sur une plage continue, qui comporte une pompe hydraulique, fonctionnellement entraînée par un arbre d'entrée, et un moteur hydraulique entraînant fonctionnellement un arbre de sortie. La pompe et le moteur hydrauliques sont couplés mécaniquement par le biais d'une paire de trains planétaires; ils sont également couplés hydrauliquement par le biais d'une rampe, ainsi, le fluide hydraulique sous pression par ladite pompe entraîne le moteur, et le fluide utilisé par le moteur est recyclé vers la pompe où il est mis une nouvelle fois sous pression. Les deux trains planétaires sont disposés coaxialement avec l'arbre d'entrée et l'arbre de sortie, la pompe et le moteur hydrauliques étant disposés en série, chacun de part et d'autre de la rampe, et parallèles aux arbres d'entrée et de sortie, ce qui permet d'optimiser l'utilisation de l'espace et de maintenir au minimum la longueur globale de la transmission, et de réduire les longueurs requises desdits arbres d'entrée et de sortie.
PCT/US2006/012921 2005-04-05 2006-04-05 Transmission a variation continue hydromecanique WO2006108109A2 (fr)

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US60/668,730 2005-04-05

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Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113251127A (zh) * 2021-04-06 2021-08-13 上海宇航系统工程研究所 一种着陆灯驱动机构
US11098792B2 (en) 2019-09-30 2021-08-24 Caterpillar Inc. Transmission system for machine

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3897697A (en) * 1974-02-01 1975-08-05 Caterpillar Tractor Co Infinitely variable drive ratio hydro-mechanical transmission for vehicles or the like
US5683322A (en) * 1993-04-21 1997-11-04 Meyerle; Michael Continuous hydrostatic-mechanical branch power split transmission particularly for power vehicles
US6001038A (en) * 1995-12-27 1999-12-14 Steyr-Daimler-Puch Aktiengesellschaft Method of controlling a power distribution hydromechanical branched transmission in uncertain gear positions
US6007444A (en) * 1996-03-12 1999-12-28 Daikin Industries, Ltd. Hydromechanical transmission
US7063638B2 (en) * 2000-12-19 2006-06-20 Cnh America Llc Continuously variable hydro-mechanical transmission

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3897697A (en) * 1974-02-01 1975-08-05 Caterpillar Tractor Co Infinitely variable drive ratio hydro-mechanical transmission for vehicles or the like
US5683322A (en) * 1993-04-21 1997-11-04 Meyerle; Michael Continuous hydrostatic-mechanical branch power split transmission particularly for power vehicles
US6001038A (en) * 1995-12-27 1999-12-14 Steyr-Daimler-Puch Aktiengesellschaft Method of controlling a power distribution hydromechanical branched transmission in uncertain gear positions
US6007444A (en) * 1996-03-12 1999-12-28 Daikin Industries, Ltd. Hydromechanical transmission
US7063638B2 (en) * 2000-12-19 2006-06-20 Cnh America Llc Continuously variable hydro-mechanical transmission

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US11098792B2 (en) 2019-09-30 2021-08-24 Caterpillar Inc. Transmission system for machine
CN113251127A (zh) * 2021-04-06 2021-08-13 上海宇航系统工程研究所 一种着陆灯驱动机构
CN113251127B (zh) * 2021-04-06 2022-07-01 上海宇航系统工程研究所 一种着陆灯驱动机构

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