Abstract:
A hydromechanical variable speed transmission ( 1 ) including outer and inner differential gear trains each having a carrier carrying planet pinions meshing with two coaxial side gears, the side gears of the outer differential gear train being ring gears one of which is an input ring gear ( 11 ) and the other ( 52 ) of which rotatably drives the planet pinion carrier of the inner differential gear train, the side gears of the inner differential gear train respectively rotatably driving two coaxial output shafts which extend axially outwards through the ring gears ( 11, 52 ) of the outer differential gear train, the planet pinion carrier ( 14 ) of the outer differential gear train rotatably driving a hydraulic motor ( 2 ) which is fluidly coupled by a hydraulic circuit to rotatably drive a hydraulic pump ( 3 ) having a pump control for changing the displacement of the hydraulic pump ( 3 ) in response to rotation thereof, wherein the pump control is fixed from rotation to thereby isolate the input ring gear ( 11 ) from torque recirculation.

Description:
FIELD OF THE INVENTION  
       [0001]     The present invention relates to a hydromechanical variable speed transmission.  
       BACKGROUND OF THE INVENTION  
       [0002]     The present applicant&#39;s WO 98/17927 discloses a mechanical variable speed transmission having interconnected outer and inner differential gear trains for differentially driving two axles. The present applicant&#39;s WO 99/61820 discloses a hydraulic transmission connected in parallel to the compound differential of WO 98/17927 for controlling the torque/speed ratio of the two axles. The disclosures of WO 98/17927 and WO 99/61820 are incorporated herein by reference.  
         [0003]     It is desirable to provide a hydromechanical variable speed transmission which is compact, clutchless, efficient, reliable and easy to manufacture.  
       SUMMARY OF THE INVENTION  
       [0004]     According to the present invention, there is provided a hydromechanical variable speed transmission including outer and inner differential gear trains each having a carrier carrying planet pinions meshing with two coaxial side gears, the side gears of the outer differential gear train being ring gears one of which is an input ring gear and the other of which rotatably drives the planet pinion carrier of the inner differential gear train, the side gears of the inner differential gear train respectively rotatably driving two coaxial output shafts which extend axially outwards through the ring gears of the outer differential gear train, the planet pinion carrier of the outer differential gear train rotatably driving a hydraulic motor which is fluidly coupled by a hydraulic circuit to rotatably drive a hydraulic pump having a pump control for changing the displacement of the hydraulic pump in response to rotation thereof, wherein the pump control is fixed from rotation to thereby isolate the input ring gear from torque recirculation. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS Embodiments of the present invention will now be described solely by way of non-limiting example and with reference to the accompanying drawings in which:  
       [0005]      FIGS. 1-4  are schematic representations of a first embodiment of a hydromechanical variable speed transmission of the present invention;  
         [0006]      FIGS. 5 and 6  are schematic representations of a second embodiment of a hydromechanical variable speed transmission of the present invention;  
         [0007]      FIGS. 7-10  are schematic representations of a variable displacement radial piston hydraulic pump of the first embodiment;  
         [0008]      FIGS. 11-16  are schematic representations of a stationary commutator of the first embodiment;  
         [0009]      FIGS. 17-20  are schematic representations of a radial piston hydraulic motor of the first embodiment;  
         [0010]      FIGS. 21 and 22  are schematic representations of a rotatable commutator of the first embodiment;  
         [0011]      FIG. 23  is a hydraulic circuit diagram of the first embodiment;  
         [0012]      FIG. 24  is a data sheet for calculating input, restraint and output torques for a conventional heavy duty front end four wheel drive loader;  
         [0013]      FIG. 25  is a data table of calculated input, restraint and output torques for a conventional heavy duty front end four wheel drive loader; and  
         [0014]      FIG. 26  is a graph comparing the performance of a conventional front end loader respectively fitted with an embodiment of a hydromechanical variable speed transmission of the present invention and a standard torque converter.  
     
    
     DETAILED DESCRIPTION OF THE EMBODIMENTS  
       [0015]     The differential transmission of the present applicant&#39;s WO 99/61820 requires a restraint for the annular pinion carrier and displacement of hydraulic fluid allows the restrained torque forces to rotate the pinion carrier at a different speed than the first bevel gear. This differential action allows variation of the speed of rotation and the direction of the second bevel gear output. When input torque is applied to the input first bevel gear and the output second bevel gear is restrained by a work load, then the annular pinion carrier will rotate around the first bevel gear and rotate forward in the same direction as the input direction of rotation and must be restrained to provide rotation and torque to the second bevel gear output to be able to effectively do work.  
         [0016]      FIG. 1  illustrates a first embodiment of a power transmission  1  which includes a hydraulic restraint motor  2 , a variable displacement control pump  3 , a tail shaft power input  4 , a power input pinion  5  and primary power input crown wheel  6 . This configuration is suited for use in trucks, tractors and all vehicle applications where the power is delivered to the axle and differential via a tail shaft, including four wheel drive vehicles where the power is normally transferred from a transfer case to a front and rear axle. For example, the power transmission  1  may replace a standard differential and eliminate a clutch and a gearbox.  FIG. 2  illustrates an alternative configuration in. which the primary power input is delivered via a gear sprocket or belt  7  via the crown wheel  6 . In both power input configurations, the crown wheel  6  also provides the power transmission for a scavenger pump  8   a  which returns case drain and lubrication oil through a filter back to the oil tank which can be conveniently located any where on the vehicle. The crown wheel  6  also supplies power to the charge pump  8   b  which makes up for any loss of oil via supplying the pump motor closed loop on the low pressure side of the circuit via a check valve  76  as illustrated in  FIG. 23 .  
         [0017]     All of the components of the power transmission  1 , including the motor  2  and the pump  3 , rotate about a common axis A-A and are compactly located within a housing  9 . The power input is transferred via the hollow shaft  10  to the first bevel gear  11  which also is connected to the motor piston group  12 . The hydraulic restraint motor  2  includes a waveform cam track  13  which is attached to the annular pinion carrier  14 . In use, the relative speeds of rotation of the first bevel gear  11  and the annular pinion carrier  14  are controlled via the displacement of hydraulic fluid, for example oil, through the pump  3 , allowing the motor piston rotating group  12  to rotate within the wave form cam track  13  as illustrated in  FIGS. 17, 18  and  23 .  
         [0018]     To allow the oil to be displaced from the motor cylinder cavity  15 , the pump circular cam track  16  must be moved of centre to achieve a displacement variation between pump cylinders  17  and  18  as illustrated in  FIG. 2 . Referring to  FIGS. 1, 2 ,  4  and  23 , when the cam track  16  is moved off centre the oil flows from the pump into galleries  19  through the motor  2  and back through galleries  20  in a closed loop.  FIG. 2  is sectioned horizontally at the midpoint to illustrate the function of components and the oil flow path. If the pump  3  cam track  16  is moved off centre, then the displacement of cylinderl 8  becomes greater and the volume of cylinder  17  becomes less. Referring to  FIGS. 11-13  and  23 , the cam track  16  is fixed at a pivot point  21  so that the pump displacement is via two separate semicircular kidney shaped ports  22  and  23 , and the pump commutator plate  24  is fixed from rotation. Referring to FIGS.  2  and  FIGS. 11-13 , oil always enters the two central galleries of the valve plate  24  and from there to the 360 degree high and low pressure full circle galleries  19  and  20 .  
         [0019]      FIG. 13  illustrates oil flow from the pump semicircular galleries to the full circle high and low pressure ports via galleries  25  and  26  linking with galleries and ports  19  and  20 . As there are four zones of pressure within the valve plate face at  19 ,  20 ,  22 , and  23 , and as hydraulic pressure balancing is used to hold the contact faces together, it is necessary to release any leakage from each zone. This is done via the galleries  27  illustrated in  FIGS. 11-13 . Referring to  FIGS. 14 and 16 , the galleries transfer the pump kidney shaped high and low pressure ports  22  and  23  to the full circle high and low pressure ports  19  and  20 .  
         [0020]     Referring to  FIGS. 15 and 16 , the appropriate balance of hydraulic commutator face clamping forces automatically and accurately adjust to the varying operational pressures by means of annular (or doughnut-shaped) piston rings  28  and  29  for the full circle high and low pressure areas and semispherical banks of individual pistons  30  and  31  to provide balance for the two separate areas of high and low pressure which can reverse subject to acceleration or deceleration. The galleries  37  provide the connection for oil flow from the different pressure zones to the respective pistons. As illustrated in  FIGS. 16 and 14 , the galleries  32  and  33  of the valve plate  24  receive the oil flow in two separate semicircular zones and return it back to full circle high and low pressure ports  19  and  20  to and through the galleries  35  and  36  in the rotating pump body as illustrated in  FIG. 4 . The oil flow then enters the valve plate  38  which has the galleries joining the pump face and sealed via “O” rings as illustrated in  FIG. 3 .  
         [0021]     Referring to  FIGS. 2, 4  and  17 - 18 , the valve plate is bolted to the pump rotating group this is also connected by fasteners  39  to the annular pinion carrier  14 , which both serves the purpose of driving the pump piston rotating group  3  and also ensuring that the radial piston hydraulic restraint motor face ports  40 , stay correctly aligned with the waveform motor cam track  13  as illustrated in  FIGS. 2, 3 ,  4 ,  19  and  21 . Referring to  FIGS. 3, 4  and  19 - 21 , the commutator/valve plate  42  is fixed to the pump face  3  and drive assembly  38  by location dowels and notches  41  to ensure proper timing with the waveform motor cam plate  13 .  FIG. 19  illustrates that the alternating high and low pressure ports  43  and  44  remain timed with the wave form motor cam track at all times.  FIG. 20  illustrates the pump face side of the motor commutator/valve plate showing “O” rings  45 ,  46 ,  47  and  48  to seal the full circle high and low pressure port zones against the pump face  3  in  FIG. 2 .  
         [0022]     Referring to  FIGS. 20 and 21 , the separate full circle high and low pressure zones  50  and  51  correspond with the pump pressure zones  19  and  20  illustrated in  FIGS. 11-14 . FIGS.  22 ( a ) and  22 ( b ) illustrate the high and low pressure galleries  40  and  50  which connect to the motor commutator/valve face ports  40  so that there is alternating ports of high and low pressure  43  and  44  in  FIG. 19 . As illustrated in  FIG. 3 , it is necessary to have the proper hydraulic balanced force to maintain the sealing of the hydraulic motor commutator/valve face  40 . This is achieved by having an annular piston  69  connected to the cylinder chamber via gallery  70 . By this means the pressure in the cylinder is applied to the annular (doughnut-shaped) piston  69  applying a force in the direction of the arrow  71 . This in turn applies the force to the rotating piston group causing it to slide along the hollow shaft  10  and the spline  72 , to provide the correct balanced force proportionate to the cylinder chamber and commutator internal hydraulic pressure to seal the rotating commutator/valve plate face connection. Referring to  FIG. 2 , the valve plate attachment  38  restrains the clamping forces applied via the connection  39  to the annular pinion carrier  14 . By this means, the varying forces applied to the pump commutator plate  24  and the motor commutator plate  73  illustrated in  FIG. 1  are isolated.  
         [0023]      FIGS. 5 and 6 , illustrate a second embodiment of a power transmission  1  having a first bevel gear input  11 , pinions  53 , a pinion carrier  14  and a second bevel gear output. The hydraulic power unit  54  is comprised of a fixed displacement motor  55  and variable displacement axial piston pump  56 . The pump housing  57  and variable swash plate  58  are fixed via the pump flange  59 . The power input to the hydraulic power unit  54  is via gear chain or belt  62  at A  60 , this is attached to the motor housing  61  and also provides the power input at A  62  to the first bevel gear  11 . The fixed motor swash plate  63  rotates in unison with the motor housing  61 , and gear sprocket or pulley  62  and is connected to and rotates the first bevel gear  11  at a constant ratio. The motor  55  is free to rotate being controlled by the displacement of the pump in a closed loop circuit. The motor shaft  64  is connected to the gear  65  which in turn is connected via power input B  66  to the annular pinion carrier  14 . The motor shaft  64  is connected via a splined coupling with the pump shaft to provide the input drive to the pump piston rotating group  56  and continue through the stationary pump base  59  for driving a charge pump to supply oil to the low pressure side of the closed loop circuit via a check values. The pump piston rotating group is driven via the annular pinion carrier at B  66 . The flow control and displacement control-pump is restraining the annular pinion carrier  14  which is restrained by the variable swash plate  67  which is fixed. The pump and motor cases have a relatively low pressure seal in  FIG. 5  to allow the motor case to rotate while the pump remains stationary without leakage.  
         [0024]     By altering the angle of the variable swash plate  67 , the pump displacement is changed causing the speed of rotation between A  62  input to the first bevel gear and input B  66  annular pinion carrier with the proportionate variation of speed and direction of rotation of the second bevel gear  52  output. As the pump valve plate and variable swash plate  67  are fixed, the relationship between the semicircular kidneys shaped high and low pressure ports remain the same. The motor piston rotating group  55  rotates at a variable speed to the motor housing  61  and the fixed swash plate  63  which is fixed to the motor housing  61 . The motor commutator valve plate  68  is therefore fixed to the motor housing  61  to maintain proper alignment with the separate semi circle kidneys shaped high and low pressure ports.  
         [0025]     The transfer of oil between the pump and motor is achieved with the same methods as described with reference to  FIGS. 11-14  so as to provide common high and low pressure commutation between the pump and motor semispherical shaped porting on each of the rotating groups.  
         [0026]      FIG. 23  illustrates the closed loop circuit (but excludes showing the detail of the motor alternating high and low pressure zones) in schematic form. The hydraulic restraint motor  2  and the variable displacement control pump  3  are connected via a high pressure half of a closed loop circuit  74  and a low pressure half of the circuit. The flow direction and the rotation of the motor piston group  13 , the motor wave form cam track  13  and the pump piston group  80  always rotate in the same direction, hence the hydraulic fluid always flows in the same direction in the closed loop circuit  74  and  75 , however the displacement volume, speed and pressure varies constantly to meet operational demands and the high and low pressure areas  74  and  75  alternate with changing output loads from a vehicle accelerating or decelerating.  
         [0027]     The restraint pump rotating piston group  80  applies the restrained load to the cam track  16  which is fixed to the differential housing via a pivotal point  21  the cam track is restrained from rotating about the pivotal point  21  by a hydraulically assisted servo control  78  being attached to the cam track  16  at a pivotal point  77 . The movement of the cam track off centre for variable displacement to meet operational demands is controlled by a microprocessor  79 .  
         [0028]     Both the pump and the motor are of the same configuration in the area of rotating piston groups to meet the rigorous torque restraint forces, to have reliability and a long life and to reduce manufacturing costs.  FIG. 7  illustrates a partial cutaway of the pump  3  exposing the pistons  81  and piston rings  82 —if side load forces are removed from the piston and cylinder entirely then if a ring or rings are used for sealing manufacture tolerances can be reduced for high speed high volume manufacture without a reduction in performance and service life. As illustrated in  FIGS. 7-10 , the piston  81  and piston ring  82  and cylinder  17 ,  18  are free from any side load and only subject to direct linear force against the hydraulic fluid. This is achieved by the extended walls of the pistons  81  having piston pins  85  fitted centrally through the piston walls with the cam rollers  83  fitting into the piston on centre line. Thus, any vectored thrust from the action of the roller  83  against the cam track  16  will be imparted precisely on the pin/shaft  85  centre line which is supported via radius rollers  85  that run in machined tracks in the pump body  80 . Referring to  FIG. 9 , as the cam track  16  is flat in the contact area, there is no axial load applied but a radial tangent load is applied due to the offset cam track ring in one embodiment. As illustrated in  FIG. 8 , the radius rollers  85  running in the machined track  87  takes all the radial loads including side loads imposed on the pin and avoid any of these forces reaching the piston. The pin  85  is stepped so that when the needle roller bearings  84  and the rollers  83  are assembled on the pin and the rollers  87  are lowered into the machined radius slot  87  and the cam track ring  16  is placed around the piston group there is no other assembly necessary in one embodiment. Referring to  FIG. 10 , in one alternative embodiment for low cost manufacturing, flat faced rollers  83  can run on a flat face  87  with side load being prevented by thrust washers retained by a face plate (not shown).  
         [0029]     Embodiments of the present invention provide control and restraint of the free wheeling rotation of the annular pinion carrier without loss of power other than the power required for the normal operation of the charge pump and other hydraulic inefficiencies. Embodiments of the invention also provide a compact rotatable assembly of a variable displacement pump and a fixed displacement motor rotating about the same axis as the transmission differential, with the internal differential and output axles all housed within a differential axle housing. These advantages are provided by hydraulic flow control via multipurpose commutator and valve plates with in one embodiment reversal of flow direction within the pump stationary commutator plate and back through the rotating pump which is controllably varying the flow and volume of the hydraulic fluid being displaced through a commutator which transfers two separate zones of high and low pressure to a 360 degree series of separate ports all at the same diameter to provide discrete areas of high and low pressure to match the rotating motor ports which is achieved by rotating the valve plate in unison with the motor wave form cam track. On the pump oil delivery side, the annular variable off centre cam track is fixed as well as the valve plate so that the semicircular kidney shaped commutator plate orifices stay properly aligned with the pump high and low pressure delivery ports, and then as described above reversed through the commutator and the rotating pump body.  
         [0030]     The power unit rpm, torque load and other inputs are input into a microprocessor which provides synergistic analysis instantly and continually making real time outputs to a control device which via hydraulic servo control adjusts the position of the circular cam track off centre to vary the displacement of the pump to thereby act as a variable displacement flow controller. This allows for the power unit to operate at a precise rpm a torque balance for the optimum ratio so that input power matches the output requirements and allows the power source to always operate at the most efficient balance of torque and speed (rpm).  
         [0031]     The input power drive is connected to both the first bevel gear and to the rotating piston group. The outer wave form cam track (in the case of the radial piston motor) is connected to the annular pinion carrier. Variation of the pump displacement causes the same displacement of fluid through the radial piston motor this allows the outer cam track to rotate at a different speed relative to the piston group. As the cam track is attached to the annular pinion carrier this causes the corresponding variation of rotational speed between the first bevel gear and the annular pinion carrier and the effect on the out put speed and direction of rotation of the second bevel gear. Displacement of the fluid from the motor is caused by the adjustment of displacement of the pump and therefore must at all times be the same volume of variation. The torque formula for hydraulics is as follows.  
           Cc   ⁢     /     ⁢   Rev   ×   Pressure   ⁢           ⁢   Bar       62.857142   ⁢               =     Nm   ⁢           ⁢     (     Newton   ⁢           ⁢   metres     )           
 
         [0032]     Torque forces must always remain balanced between the restraint motor and the flow control pump across the speed and volumetric range of operation and at all times are restrained by a fixed pump cam track or a pump swash plate which is fixed with no possibility of recirculation of torque interfering with the power unit. Because the pump rotating piston group is driven by the rotational speed of the annular pinion carrier which the pump controls, then the rotational speed of the pump will always be the same as the annular pinion carrier.  
         [0033]     The very high reactive torque forces restrained by the radial piston motor through the hydraulic circuit reacts directly through the expanding cylinder volume and the movement of the pistons in the pump flow control and is restrained by the cam track which is caused to try to rotate. This rotational force is restrained by the cam track being fixed to the differential housing this then prevents any of the restraint force being applied to the power input source or requiring any power from the power unit to restrain the annular pinion carrier. Thus the power required is directly related to the overall ratio affected between the input and the output.  
         [0034]     The power input into the first bevel gear rotates the pinions and the annular pinion carrier in the same direction as the power input. The second bevel gear will remain stationary if subject to load until the pinion carrier is restrained. In preferred embodiment, the pinion carrier is restrained by the waveform cam track acting against the rollers connected to the rotating piston group. The pressure of restraining the pinion carrier is restrained by the pump piston rollers reacting against the variable cam track. For the displacement to be variable at the pump, the pump cylinder/piston chamber must expand in the direction of the pump rotation which is always in the direction of the power input rotation. This means that the torque generated is in the same rotational direction as the power input, and the torque being equal cancel out and remain balanced. The torque restraint at the pump is directly against the variable cam track which is fixed from rotation in one embodiment directly to the differential housing. The annular pinion carrier restraint forces and the variable volume flow control of the pump are restrained against a variable position cam track which is fixed to prevent rotation and is free from any contact or means of causing the annular pinion carrier restraint forces being transferred to the power input apart from normal charge pump, lubrication and hydraulic efficiency losses. The power input is always only that required to meet the demands of the ratio selected to move the out put load at the second bevel gear. The resultant action of the pinion rotating about the pinion bearing causes the torque load applied to the annular pinion carrier to be always equal to twice the output torque.  
         [0035]     When a vehicle is at operating speed the input torque at the first bevel gear and the restraint torque at the annular pinion carrier will be in a ratio of 1:1, and the output torque of the second bevel gear will be also the same with an overall ratio from the power input to output of 1:1. At this point, the cam track on the radial piston pump will be central with zero displacement and zero oil flow, likewise the axial piston pump will be at 90 degrees to the axis of the rotating piston group with the same result as described above. High torque forces are being restrained at this point but with zero pump displacement and with no hydraulic fluid flow there is no work being done and the power source is applying power directly to the out put load. By way of analogy it is like raising a load with a hydraulic jack and leaving the jack support the full heavy load. The high load is continuously present, but no movement, and thus no work, is being done. There is therefore no requirement for power other than conventional power losses.  
         [0036]     At the other extreme, when the input first bevel gear and the annular pinion carrier is moved slightly off 2:1, it can give an overall input/output ratio between the first bevel gear and the second bevel gear of one million to one, and thus the power source is seeing a torque reduction of one million to one.  FIG. 24  is a data sheet of example calculations of torque required for a conventional four wheel drive front end loader to skid the wheels.  
         [0037]      FIG. 25  is a data table illustrating that the power input required to skid the loader wheels is only 4.7 kW.  
         [0038]      FIG. 26  is a graph which illustrates the 100% power loss experienced by conventional front end loaders fitted with torque converters. This occurs particularly at the point of loading the front end loader bucket when the torque converter is on stall. The engine will be running at full peak speed and maximum power for the torque converter to generate torque but the wheels will not be turning or only turning very slowly. This means that on stall the torque converter is turning 100% of the engine horsepower to heat and which requires conventional front end loaders to be fitted with heat exchangers capable of dissipating around 30% of the engine power in heat. This requires further power loss in terms of oil pumps heat exchangers and high capacity fans to cool the torque converter oil.  FIG. 26  illustrates that a hydromechanical variable speed transmission of an embodiment of the present invention has the torque restrained by the fixed pump cam track, and that the engine in the front end loader can only see the power required and effected by the very high differential gearing and ratio between the first bevel gear input and the second bevel gear output. As the low torque requirement is input into the microprocessor the output from the microprocessor instantly adjusts the engine speed, and in this case while loading the bucket and just skidding or spinning the wheels while crowding and loading the bucket as illustrated in  FIG. 25  this would only require 4.74 kW of power so the output from the microprocessor would balance the engine torque and speed closing the engine throttle to run at a low speed near idle speed.  
         [0039]     Illustrative component operating speed calculations are provided for the operational speed of trucks. The calculations use the following data. 
        Loaded radius of 10.00−20 truck tyre=20.25 inches (514.35 mm)×2=1028.7 mm diameter×3.1428571=3.233 metre circumference     100 Kilometres per hour=1666 metres/minute     1666 divided×3.233 meters=515.5 rpm     At a cruise speed of 100 kilometres per hour the truck axle and differential are rotating at 515.5 rpm.        
 
         [0044]     The outer differential transmission input ratios of the first bevel gear and the annular pinion carrier will be 1:1 and there will be zero hydraulic oil flow or hydraulic pump/motor rotation relative to each other but they will be rotating in unison at 515.5 rpm.  
         [0045]     With the vehicle stationary at dynamic neutral lock regardless of power input rpm at dynamic neutral lock the differential hydraulic speed variation between the first bevel gear and the annular pinion carrier will be 2:1 requiring the hydraulic speed variation to be 515.5 divided×2=257.75 rpm. This makes the choice of radial piston pumps and motor very suitable for the high torque loads and relatively low rpm speed requirement.  
                                       If 0 to 100 km/hr requires an rpm variation of the pinion    257.5 rpm       carrier of:       Then for 10 km/hr reverse will require a further reduction    23.03 rpm       of:       It can be seen the total hydraulic speed variation will be:   232.03 rpm       515.5 rpm/232.03 rpm = required ratio of 2.22:1.                  
 
         [0046]     An illustrative embodiment of the present invention using a fixed displacement multilobed wave form cam track radial piston motor provides a motor that is compact with high displacement per RPM for high torque load restraint. Illustrative performance data for this embodiment are as follows. 
        8 pistons×6 lobes==48 strokes/rev say 12 mm stroke=576 mm piston travel per rev=57.6 cm     If the piston diameter was 30 mm diameter=3 cm     Area=1.5×1.5 cm×3.1428571=7.071 square cm     7.071 square cm×57.6 cm piston travel/revolution=407 cc/revolution     The total variation of the input rpm to the annular pinion carrier of 232.03 rpm to give 10 km/hr reverse will be 515.5 rpm minus 232.03 rpm=283.47 rpm     283.47 rpm at 407 cc/revolution divided by 1000=115.37 litres/min     As the annular pinion carrier is directly rotating the pump piston group at 232.03 rpm 115.37 litres divided×232.03=0.49722 litres per rev or 497 cc/revolution     Using a cam ring pump that can be offset of centre for displacement variation allows the entire control in this one simple function.     An adjustable off centre cam ring provides one piston stroke per rotation, using 8 pistons in the rotating group 497 cc/revolution divided×8=62 cc capacity per piston stroke per rev at full cam track at full offset full piston travel for reverse.     Using a 5 cm piston diameter 2.5 cm×2.5 cm×3.1428571=19.6428 square cc     62 cc divided×19.6428=3.156 cm piston/cylinder stroke length required for maximum displacement in reverse        
 
         [0058]     It will be understood that the above performance data are example calculations only, and that actual performance data for embodiments of the invention may be relatively more advantageous.  
         [0059]     Embodiments of the present invention provide a hydromechanical mechanical vehicle transaxle or powertrain in which an oil flow control device, for example an hydraulic pump and/or a hydraulic motor, is driven by a rotating annular pinion carrier which the pump restrains and controls the volumetric displacement and hydraulic fluid flow, which in turn controls the speed of rotation of the rotating pinion carrier and pinions, including the speed at which the pump rotates. The restraint forces being retained by the pump variable cam track being fixed to the differential housing in the case of the radial piston pump embodiment. The restraint forces are retained by the pump swash plate, which is fixed and not free to rotate in the case of the axial piston pump embodiment. In both embodiments, the power source is isolated and free from torque from the differential, the annular pinion carrier and pinions and can not be transferred back to the input power source. The piston groups and the cam track of the radial piston motor embodiment (or the fixed displacement swash plate of axial piston motor embodiment) rotate as well as rotating at a variable speed to each other subject to the pump displacement. In both embodiments, the means of varying the pump displacement is fixed and not free to rotate with the pump with stationary semicircular commutator valve port plates which convert the high pressure and low pressure semispherical sides of high and low pressure to two 360 degree separate high and low pressure areas which can communicate with valve commutator port plates that rotate in a fixed position relative to the cam track lobes and the fixed displacement swash plate of the respective hydraulic motors to allow proper communication of hydraulic fluid at the correct position to allow the filling and discharge of the cylinders regardless of the rotation of the rotating piston groups and cam tracks and plates.  
         [0060]     Embodiments of the present invention provide the following advantages. 
        A compact hydromechanical mechanism that will fit within a differential or transaxle housing for all components to rotate about a common axis.     Control reactive and recirculation torque to react directly between a fixed mounting and the rotating differential pinions, free from direct connection to the power source.     A combination of fluid displacement control with interconnected components rotating about the axle axis of the differential.     A means of fluid flow commutation and valve control between the rotating components by using the pump commutator valve plates to convert flow and high pressure semispherical sides of the pump to a high and low pressure full circle separate high and low pressure areas and to reverse flow direction from the pump back to and through the pump to connect to the motor which rotates about the same axis. The motor commutator valve plates to convert flow and high pressure from two separate spherical areas of high and low pressure to one radial location but in intermittent high and low pressure areas to match the multi lobed radial piston motor cam track.     Use pump and motor common valves and commutator plates and piston groups as a high pressure high volume rotary hydraulic fluid connector.     The elimination of hosing, and high speed and high volume rotating connectors external to pump and motor commutator valve plates by means of direct interfacing rotatable components.     Mechanical control of divergent thrust on hydraulic components thereby enabling simpler lower cost manufacture of hydraulic components and to eliminate side thrust.     The elimination of complicated costly relief valves by using real time instantaneous transmission speed and torque inputs to and outputs from a microprocessor to stay within pre-programmed pressure limits.     Multistack overlapping pancakes of axial piston cylinder assemblies for ease of selection and provision of torque control requirements.     The elimination of oil bath splash lubrication in the differential and by so doing eliminate hydrodynamic drag and resultant power loss.     Allow filtration of common hydraulic and differential lubricating oil.     The provision of spray jet lubrication delivered precisely to the high load areas.     Provision of oil cooling of oil for the hydraulic and differential mechanical components.        
 
         [0074]     The present invention is not limited to the embodiments that have been described and depicted, but variations and modifications may be made without departing from the scope of the present invention. For example, the invention is not limited to radial or axial piston hydraulic pumps but may also be implemented using other conventional variable displacement hydraulic pumps, for example a bent axis variable speed hydraulic pump.