Patent Document

[0001]    This invention pertains to a continuously variable hydromechanical transmission for a vehicle, and more particularly to a transmission having parallel axial piston pump and motor hydraulically linked through a stationary manifold and mechanically linked through a variable ratio gear set to provide an output torque with a constant mechanical portion and a variable hydraulic portion which diminishes to zero at hydraulic lock-up.  
         BACKGROUND OF THE INVENTION  
         [0002]    Interest in continuously variable hydromechanical transmissions has been increasing in recent years because of the potential operational efficiencies and economies that are increasingly becoming possible in vehicles and other powered systems wherein rotary input power is to be converted to output power at the desired output torque and speed. Continuously variable transmissions provide operational efficiencies and economies in the vehicle that are potentially superior to any known transmission, and theoretically can do so in packages that are smaller and lighter weight than other available transmissions.  
           [0003]    However, conventional prior art hydrostatic transmissions are known by experts in the art to be noisy and inefficient. Convincing those experts and vehicle manufacturers that these new generation hydrostatic transmissions have overcome the intractable problems of the prior art is difficult. Therefore, additional improvements would facilitate acceptance of the new generation hydrostatic transmissions.  
           [0004]    One such improvement would be in the area of leakage from rotating interfaces, particularly those where working fluid is commutated between the differentially rotating pump and motor.  
           [0005]    Another improvement would be in the area of dynamic balancing. The difficulty of balancing rotating equipment to preclude vibration induced by rotating eccentric masses becomes worse exponentially with increasing speed of rotation.  
           [0006]    Yet another improvement would be in reducing the losses caused by “windage” and fluid drag associated with the rotating elements inside the transmission housing. In applications having a prime mover with a high rotating speed, such as an electric motor, turbine engine or high performance spark ignition gasoline engine, the input elements would rotate at the prime mover output speed unless a gear reduction unit were interposed between the prime mover and the transmission. Gear reduction units add undesirable cost and weight. The windage and fluid drag losses can be greatly reduced by reducing the speed of rotation of those rotating elements.  
           [0007]    Still another desirable improvement would be in the area of manufacturability, simplicity, and cost. Prior art continuously variable hydromechanical tansmissions have tended to be excessively complicated and costly to build. It would be a welcome development to original equipment manufacturers to have a continuously variable hydromechanical tansmission available that is efficient, small and light weight, and is easily and economically manufactured and maintained.  
           [0008]    One approach for achieving these improvements is shown in an international patent application No. PCT/US98/24053 filed on Nov. 12, 1998 by Folsom and Tucker entitled “Hydraulic Machine”. A variation of this approach in a tandem hydromechanical transmission using low cost conventional components would make this technology available for smaller vehicles requiring more compactness and lower cost, such as outboard motors for boats, motor scooters, motor cycles, RVs and snowmobiles.  
         SUMMARY OF THE INVENTION  
         [0009]    Accordingly, it is an object of this invention to provide an improved hydromechanical continuously variable transmission for vehicles. Another object of this invention is to provide an improved method of transmitting power from a prime mover of a vehicle to the drive members of the vehicle (wheels, tracks, propeller, belt, etc) at output power in a continuously variable combination output torque and speed selected by the driver.  
           [0010]    These and other objects are attained in a parallel hydromechanical continuously variable transmission having a housing holding a make-up pump and internal cavities for holding operating assemblies of the transmission, including an axial piston pump and an axial piston motor. The pump and motor each have a rotating element and a non-rotating element. Each non-rotating pump element is mounted for tilting movement in its own respective pair of mounting journals in the housing. The tilting axes of the non-rotating elements lie transverse to the axes of rotation of the rotating element. The pump and the motor are disposed side-by-side in the housing with the axes of rotation approximately parallel to each other. A variable ratio gear set couples the pump, motor, and output shaft so that the reaction torque from the pump is delivered directly to the output shaft. The pump and motor are coupled hydraulically through fluid passages in a stationary manifold, fixed in the housing. Internal fluid passages in the stationary manifold convey fluid pressurized in the pump directly to the motor, and convey spent fluid displaced from the motor back to the pump. The transmission ratio is controlled by the tilt angle of the non-rotating pump and motor elements. A tilt angle control apparatus attached to the housing and to the non-rotating pump and motor elements governs that tilt angle. 
       
    
    
     DESCRIPTION OF THE DRAWINGS  
       [0011]    The invention and its many attendant objects and advantages will be better understood upon reading the following detailed description of the preferred embodiment in conjunction with the following drawings, Wherein:  
         [0012]    [0012]FIG. 1 is a schematic diagram of one embodiment of the invention, showing the mechanical and hydraulic power train and the controls;  
         [0013]    [0013]FIG. 2 is a perspective view from the input side of one version of the transmission shown in the schematic diagram of FIG. 1;  
         [0014]    [0014]FIG. 3 is a perspective view from the output side of the transmission shown in FIG. 2;  
         [0015]    [0015]FIG. 4 is an end elevation of the transmission shown in FIG. 2 from the input end;  
         [0016]    [0016]FIG. 5 is a perspective view of the internal components of the transmission shown in FIG. 2 viewed from the input end;  
         [0017]    [0017]FIG. 5A is a perspective view of the gearing shown in FIG. 5, viewed from the output end;  
         [0018]    FIGS.  6 - 9  are sectional plan views of the pump and motor along lines  6 - 6  in FIG. 4 showing the transmission in neutral (FIG. 6), in reverse (FIG. 7), in maximum speed ratio (FIG. 8) and in maximum torque ratio (FIG. 9);  
         [0019]    [0019]FIG. 10 is a sectional elevation along lines  10 - 10  in FIG. 4 on a section through the output shaft;  
         [0020]    [0020]FIG. 11 is a sectional elevation along lines  11 - 11  in FIG. 4 on a section line through the axis of the pump;  
         [0021]    [0021]FIG. 12 is a sectional elevation on a section line normal to the parallel axes of the pump and motor and through the middle of the swashplate trunnions and the displacement control bell-cranks;  
         [0022]    FIGS.  13 - 19  are various views of the main housing shown in FIG. 2;  
         [0023]    FIGS.  20 - 23  are various views of the input end housing shown in FIG. 2;  
         [0024]    FIGS.  24 - 27  are various views of the control housing shown in FIG. 2;  
         [0025]    FIGS.  28 - 33  are various views of the make-up pump housing shown in FIG. 3;  
         [0026]    FIGS.  34 - 36  are various views of the input element, including the sun gear of the epicyclic gear set shown in FIG. 6;  
         [0027]    FIGS.  37 - 39  are various views of the ring gear of the epicyclic gear set shown in FIG. 6;  
         [0028]    FIGS.  40 - 42  are various views of the pump drive shaft shown in FIG. 6;  
         [0029]    FIGS.  43 - 45  are various views of the output spur gear shown in FIG. 10;  
         [0030]    [0030]FIGS. 46 and 47 are a perspective and side elevation of the output shaft shown in FIG. 10;  
         [0031]    FIGS.  48 - 50  are various views of the output element shown in FIGS. 3 and 10;  
         [0032]    FIGS.  51 - 53  are various views of the motor drive shaft shown in FIGS. 5A and 6;  
         [0033]    [0033]FIG. 54 is an exploded perspective view of the pump and the motor shown in FIG. 6, both pump and motor being identical;  
         [0034]    FIGS.  55 - 58  are plan views of the pump and motor displacement control unit shown in FIGS. 1, 5,  10  and  11 ;  
         [0035]    [0035]FIG. 59 is an exploded perspective view of the control unit shown in FIGS.  55 - 58 ;  
         [0036]    [0036]FIG. 60 is a perspective view of the internal components of a bent-axis embodiment of a transmission according to this invention viewed from the input end, corresponding to FIG. 5 of the first embodiment;  
         [0037]    [0037]FIG. 61 is a sectional plan view through the axis of the pump and motor of the transmission shown in FIG. 60;  
         [0038]    [0038]FIG. 62 is a sectional elevation on a vertical plane in FIG. 60 along the pump axis and the yoke hinge axis;  
         [0039]    [0039]FIG. 63 is an exploded perspective view of the pump shown in FIG. 61, which is identical to the motor;  
         [0040]    [0040]FIG. 64 is a perspective view of the front housing for the transmission shown in FIG. 60, showing the interior end wall containing the manifold;  
         [0041]    [0041]FIG. 65 is an end elevation of the front housing shown in FIG. 64, showing the slots for fluid flow to underlying pressure and suction passages for fluid flow between the pump and motor;  
         [0042]    [0042]FIG. 66 is an end elevation of the front housing shown in FIG. 64 from the front end;  
         [0043]    [0043]FIG. 67 is a sectional plan view along lines  67 - 67  in FIG. 66;  
         [0044]    [0044]FIG. 68 is a sectional elevation along lines  68 - 68  in FIG. 66;  
         [0045]    [0045]FIG. 69 is a perspective view of the middle housing for the transmission shown in FIG. 60;  
         [0046]    [0046]FIG. 70 is an end elevation looking into the middle housing shown in FIG. 70;  
         [0047]    [0047]FIG. 71 is a plan view of the control valve and control levers for the transmission shown in FIG. 60;  
         [0048]    [0048]FIG. 72 is a schematic diagram of a front wheel drive transaxle vehicle transmission;  
         [0049]    [0049]FIG. 73 is a perspective view of a transmission incorporating the elements of the schematic of FIG. 72;  
         [0050]    [0050]FIG. 74 is a perspective view of the transmission shown in FIG. 73 from the same angle, showing the interior structure;  
         [0051]    [0051]FIG. 75 is a perspective view of the transmission shown in FIG. 73 from the front;  
         [0052]    [0052]FIG. 76 is a perspective view of the transmission shown in FIG. 75 from the same angle, showing the interior structure;  
         [0053]    [0053]FIG. 77 is an elevation of the front end of the transmission shown in FIG. 75;  
         [0054]    [0054]FIG. 78 is sectional plan view along lines  78 - 78  in FIG. 77;  
         [0055]    [0055]FIG. 79 is a sectional elevation along lines  79 - 79  in FIG. 77;  
         [0056]    [0056]FIG. 80 is a sectional elevation along lines  80 - 80  in FIG. 77;  
         [0057]    [0057]FIG. 81 is a sectional elevation along lines  81 - 81  in FIG. 77;  
         [0058]    [0058]FIG. 82A is a perspective view of the drive shown in FIG. 79;  
         [0059]    [0059]FIG. 82B is a sectional elevation of the drive tube shown in FIG. 82A;  
         [0060]    [0060]FIG. 83 is a perspective view of the input shaft shown in FIG. 79;  
         [0061]    [0061]FIG. 84 is a perspective view of the pump shaft shown in FIG. 78;  
         [0062]    [0062]FIG. 85 is a perspective view of the motor shaft shown in FIG. 78;  
         [0063]    FIGS.  86 - 91  are various views of the rear housing shown in FIG. 73 and the integral cradle bearing for the swashplates;  
         [0064]    FIGS.  92 - 96  are various views of the middle housing shown in FIG. 73;  
         [0065]    FIGS.  97 - 103  are various views of the manifold shown in FIG. 78;  
         [0066]    [0066]FIG. 104 is an exploded view of the swashplate and the control crank shown in FIG. 74;  
         [0067]    FIGS.  105 - 107  are plan, elevation, and plan views respectively of the control mechanism and swashplates of the transmission shown in FIG. 74 in the neutral position;  
         [0068]    FIGS.  108 - 110  are plan, elevation, and plan views respectively of the control mechanism and swashplates of the transmission shown in FIG. 74 in the maximum torque position;  
         [0069]    FIGS.  111 - 113  are plan, elevation, and plan views respectively of the control mechanism and swashplates of the transmission shown in FIG. 74 in the maximum forward speed position; and  
         [0070]    FIGS.  114 - 116  are plan, elevation, and plan views respectively of the control mechanism and swashplates of the transmission shown in FIG. 74 in the maximum reverse speed position. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0071]    Turning now to the drawings, and more particularly to FIG. 1 thereof, a parallel hydromechanical underdrive transmission, shown in schematic form, is designed to be used in vehicles where an underdrive final ratio is desired. The transmission is shown having a variable displacement pump  50  and a make-up pump  52  driven by a prime mover  55 , and a variable displacement motor  60  hydraulically coupled through a manifold  70  and mechanically coupled through a gear train  80  which includes a variable ratio gearset such as a planetary gearset  85 . The transmission ratio is controlled by displacement control system  90  under control of a master controller  100  for the vehicle.  
         [0072]    The transmission is shown in FIG. 1 in neutral, with the pump  50  at zero displacement and the motor  60  at maximum displacement. The displacement of both hydrostatic units  50  and  60  are simultaneously controlled by the control system  90  in this embodiment, although they could instead be independently controlled.  
         [0073]    A specific embodiment of the invention using a swashplate version of the pump  50  and motor  60 , shown in FIGS.  2 - 12 , includes a four-piece housing  105 , including a main housing  106  shown in detail in FIGS.  13 - 19 , an input end housing  107  shown in detail in FIGS.  20 - 23 , a control housing  108  shown in FIGS.  24 - 27  and a make-up pump housing  109  shown in FIGS.  28 - 33 . The main housing  106  has three parallel cylindrical lobes  110 ,  112  and  114 . The two top lobes  110  and  112  hold the pump assembly  50  and the motor assembly  60 , respectively, and the bottom lobe  114  holds an output assembly  115  shown in FIG. 10. The input end housing  107  has a locating lip  116  concentric with the pump lobe  110  which is accurately machined to fit a corresponding annular recess in the prime mover  55  so the transmission housing  105  can be rigidly mated to the prime mover  55  with the axis of the output drive of the prime mover aligned with the input element of the transmission.  
         [0074]    It will be noted that torque is input from the prime mover  55  to the transmission through the input end housing  107  and is output to the vehicle drive members through both ends of the housing  105 . For convenience, the end through which torque is input to the transmission from the prime mover will be denominated the “input end” and the opposite end will be denominated the “output end” even though torque is output from the transmission at both the “input end” and the “output end”.  
         [0075]    Power input to the transmission from the vehicle prime mover  55  is via a smooth tapered bore  117  through an input sleeve  118  in an input element  120 , shown in FIGS.  34 - 36 . The output from the prime mover  55  ends in a smooth tapered shaft (not shown) that matches the tapered bore  117  of the element  120 . The tapered shaft is drawn forcefully into the tapered bore  117  by an Allen bolt  122  threaded into a tapped hole in the end of the tapered shaft of the prime mover. The Allen bolt bears against a shouldered washer  123  seated on a shoulder  124  in the bore of the element  120 . The engagement of the tapered shaft in the tapered bore  117  provides a secure torque transmission from the prime mover  55  to the transmission.  
         [0076]    The input element  120  includes a sun gear  125  of the epicyclic gearset  85 , engaged with four planet gears  130  in a planet carrier  132 , as shown in FIGS.  6 - 9 ,  11  and  5 A. The planet gears  130  are engaged with a ring gear  135 , shown in detail in FIGS.  37 - 39  which is engaged with and drives a splined flange  137  on a pump drive shaft  140 , shown in detail in FIGS.  40 - 42 . The input end  142  of the pump drive shaft is supported on a bearing  144  lying between bearing flanges  146  and  148  on the input element  120  and the input end of the pump drive shaft, respectively. A splined section  145  of the pump drive shaft is engaged with a splined bore of the pump cylinder block to drive the pump cylinder block about its axis of rotation. The support for the epicyclic gearset  85  is by way of the bearings of the output shaft of the prime mover  55 . The distal end  142  if the pump drive shaft  140  is supported in bearings  149  mounted in a bore  147  in the output end of the main housing  106 .  
         [0077]    Torque from the prime mover  55  driving the input element  120  is transmitted from the sun gear  125  through the ring gear  135  to the splined flange  137  of the pump drive shaft  140  to drive the cylinder block of the pump  50 . The reaction torque from the pump  50  is reacted back through the pump drive shaft and ring gear  135  to the planet gears  130  and thence to the planet carrier  132 . As shown in FIGS. 4 and 5, the planet carrier is fastened to a carrier spur gear  150  by machine screws  152 . The spur gear  150  is supported on a set of needle bearings  154  on the input sleeve  118  of the input element  120 .  
         [0078]    The carrier spur gear  150  is engaged with an output spur gear  160 , shown in FIGS.  5 ,  6 - 9  and  10 , and shown in detail in FIGS.  43 - 45 , which is journaled on a bearing  162  mounted in a bearing seat  163  on an inwardly projecting tubular nipple  164  on the input end housing  107 . The output spur gear  160  has an outer flange  166  on which the gear teeth  168  are cut, and a concentric inner stub tube  170  having a radial outside surface  172  concentric with the flange  166  for engaging the bearings  162 . The bore  174  of the inner stub tube  170  is splined to receive a splined end  175  of an output shaft  180 , as shown in FIGS. 10, 46 and  47 . The entire bore  174  is splined so that output torque may be taken from both ends of the transmission, as indicated in FIG. 1 and in FIGS. 2 and 3, for convenience in driving a 4-wheel drive vehicle.  
         [0079]    The output end of the output shaft  180  is journaled in a bearing  183 , shown in FIGS. 10 and 12 which is mounted in a stepped axial bore in the output end of the main housing  106  and held in place by a circlip (not shown) in an annular groove  187  in the bore  185 . Torque is output from the output end of the output shaft  180  through an output fitting  190 , shown in FIG. 10 and shown in detail in FIGS.  48 - 50 , having a splined bore  192  engaged with splines  194  on the output end of the output shaft  180 . The output fitting  190  is held on the end of the output shaft  180  by a nut (not shown) that is threaded onto a projecting threaded end  196  of the output shaft  180  and torqued against the output fitting  190 .  
         [0080]    Fluid pressurized in the pump  50  is conveyed through passages in the manifold  70  to the motor  60  where it is converted to output motor torque and conveyed to a motor output shaft  200  by way of a spline  202  on the motor output shaft  200  engaged with a splined bore  204  in the motor cylinder block  206 , shown in FIGS. 6 and 54. The motor output shaft  200  is journaled in a bearing  208  in the manifold block  70  and a front bearing  210  in a bearing recess  212  within a bearing boss  214  in the input end housing  107 , as seen in FIGS. 2, 6 and  20 . The motor output torque is conveyed through the motor output shaft  200  and through an integral gear  216  in meshing engage with the output spur gear  160  to add the torque, conveyed from the motor  60  through the motor output shaft  200  to the output shaft, to the torque conveyed from the carrier spur gear  150 , so that the total output torque to the output shaft is the mechanical portion conveyed from the carrier spur gear and the hydraulic portion conveyed from the motor  60  through the motor output shaft  200 .  
         [0081]    The hydraulic torque from the motor  60  is generated by the action of fluid pressurized in the pump  50 , shown in FIGS.  7 - 9 . Rotation of the pump cylinder block  206 P is by spline engagement of the splined section  145  of the pump drive shaft  140  to rotate the pump cylinder block  206 P against a valve plate  220 P which commutates the fluid displaced from cylinders  203 P in the pump cylinder block  206 P into pressure passages in the manifold  70  opposite the “descending” slope of the swashplate surface, and suction passages opposite the “ascending” slope of the swashplate. Pump pistons  205 P in the pump cylinders  203 P have piston heads  225 P which swivel in slippers  230 P held against the flat surface of a swashplate  235 P by a hold-down plate  240 P. The structure shown in FIG. 54 is conventional and is commercially available, e.g. from Sundstrand Hydrogear.  
         [0082]    The pressurized fluid commutated by the pump valve plate  220 P to a pressure channel in the manifold block  70  is conveyed directly to a pressure port in the manifold  70  where it is distributed by the pressure slot in the motor valve plate  220 M to the cylinders  203 M on the “ascending” side of the motor cylinder block  206 M. The fluid pressure acting against the motor pistons  205 M to drive them axially outward against the motor swashplate  235 M. The action of the axially acting pistons against the tilted surface of the motor swashplate  235 M is resolved into a circumferential force which drives the motor cylinder block “downhill” relative to the tilt angle of the surface of the motor swashplate  235 M. Continued rotation of the motor cylinder block  206 M forces the motor pistons  205 M back into the cylinders  203 M to displace fluid in the cylinders  203 M back through the suction passages in the manifold and thence into the pump cylinders  203 P on their suction stroke.  
         [0083]    The make-up pump  52  is provided to make up any fluid lost in the system by leakage, and also to pressurize the displacement control system, as described below. The make-up pump  52  is a conventional commercially available pump such as a gerotor type available from a number of sources. It is located in a cavity  243  in the manifold block  70  and is driven by a hexagonal section  246  of a quill shaft  245  having a hex head  247  engaged in the hex recess of the bolt  122  shown in FIG. 5A. The make-up pump  52  draws fluid from the housing through a suction passage  249  and the fluid pressurized in the pump is conveyed through an external fluid line through a filter  250  and thence through a one-way valve  252  to the pressure channel in the manifold block  70 . Pressure is limited to a predetermined value, e.g. 100 psi, by a pressure relief valve  254 .  
         [0084]    The displacement control system  90  shown in FIGS. 1, 5,  10 - 12  and  55 - 59  is designed to control the tilt angle of the pump and motor swashplates  235 P and  235 M. The two shashplates  235 P and  235 M each have top and bottom trunnions  258  and  260 , respectively. The top trunnions  258  are mounted in sockets  262 P and  262 M in the lid of the control housing  108 . The drawings of these sockets  262 P and  262 M are erroneous since they do not show the top trunnions  258 P and  258 M supported in the sockets  262 P and  262 M as intended. That error is easily remedied by repositioning the sockets  258 P and  258 M on the control housing  108  to align with the position of the trunnions  258 . Likewise, bottom sockets are to be provided for the bottom trunnions  260 P and  260 M, and FIG. 12 does not reflect the presence of these sockets in the floor of the main housing. This is an omission easily corrected.  
         [0085]    A pump control bell-crank  265  is mounted on the top pump swashplate trunnion  258 P and a motor control bell-crank  270  is mounted on the top motor swashplate trunnion  258 M for controlling the tilt angle of the pump and motor swashplates, and thereby controlling the pump and motor displacements. As shown in FIGS.  55 - 58 , the bell-cranks have ball-ends  272  and  274  engaged in the ends of pump and motor control pistons  280  and  285  in cylinders  290  and  295  projecting from the control housing  108  as shown in FIG. 2. The stepper motor  300  moves a control rod  305 , shown in FIG. 59, attached to a control spool  310  inside a spool valve  320 . The spool valve is driven by fluid pressure to position itself at the same position on the control spool  310  and the pump control cylinder  280  follows the spool valve  320  to position the pump control cylinder  280  at the desired location determined by the position of the control spool  310 . The motor control piston  285  is stopped at the maximum displacement position shown in FIGS. 55 and 56 by an internal stop and is biased to that position by system pressure in the cylinder  295 . The position of the motor control bell crank  270  away from the maximum displacement position is controlled by the pump control piston engaging and pushing the motor control piston  285  against the system pressure in the cylinder  285  by virtue of the greater area of the cylinder  290 .  
         [0086]    In operation, input from the engine is connected to the sun gear (Sp)  125  of the planet set  85  and then on to the make-up pump housed in the manifold. The ring gear (Rp)  135  of the planet set  85  is connected drivingly to the cylinder block of the pump  50 . The planet carrier  132  of the planet set  85  is connected to the spur gear (Sg 3 )  150  which drives the output spur gear  160  connected to the output shaft (Sg 1 )  180 . The cylinder block of the motor is connected to a spur gear (Sg 2 ) which also drives the spur gear connected to the output shaft (Sg 1 ).  
         [0087]    When the transmission is at neutral, the output shaft is stationary, hence the motor and planet carrier are also stationary. The sun gear rotates at input speed and therefore the ring gear (and hence the pump) rotates at input speed multiplied by the ratio of the numbers of teeth in the sun gear and ring rear (Sp/Rp), in the opposite direction to the input. In the preferred embodiment, the ratio is (43/77)=0.558 times input speed. Since the pump is at zero displacement, there is no pumping; therefore, no reaction torque can be generated at the pump. Hence, the pump rotates freely and there is no transmission of output torque to the output shaft.  
         [0088]    A ‘dump valve’ may be opened to ‘short circuit’ the high and low pressures of the pump and motor, so if there were to be some small displacement of the pump, there would still be no pressure, and hence, no torque would be generated with the dump valve open. The dump valve is closed electronically only when the operator selects the ‘drive’ or ‘reverse’ mode on the mode selector switch. The controller closes the dump valve only after ensuring, via a sensor, that the pump is at zero displacement.  
         [0089]    Due to the planet set configuration, the input torque is split into two parallel paths. One is a direct mechanical path fed continually to the output shaft at the ratio of input torque multiplied by (1+(Rp/Sp)). The other is a hydraulic path fed continually to the pump at the ratio of input torque multiplied by (Rp/Sp).  
         [0090]    As the pump is stroked to give a small displacement and is rotating at input speed multiplied by (Sp/Rp), it pumps fluid which flows directly through the manifold and drives the motor in the same direction to give output torque. Due to the fact that the pump is at a small displacement, a small amount of torque to the pump results in a high pressure and low flow rate. Since the motor is at a large displacement, the low flow rate from the pump at high pressure results in a high output torque and low output speed. This high ‘hydraulic’ output torque is multiplied by the gear ratio (Sg 1 /Sg 2 ) and is then added directly to the mechanical output torque as described above. Therefore the total output torque can be expressed as:  
         Output Torque=Input Torque×[(1+( Rp/Sp ))+( Rp/Sp )×motor disp/pump disp×( Sg   1 / Sg   2 )] 
         [0091]    It can therefore be seen that there is a total output torque comprising a fixed mechanical torque portion plus a variable hydraulic torque portion. As the ratio of motor displacement to pump displacement decreases, the amount of hydraulic torque decreases. When the motor displacement has been reduced to zero, the hydraulic torque portion reduces to zero and the only output torque is the fixed mechanical torque portion.  
         [0092]    As the pump displacement increases, flow rate from the pump increases, and this increased flow causes the motor and hence the output shaft to increase in speed. As the output shaft increases in speed, the planet carrier increases in speed relative to the input shaft and hence sun gear speed, this causes the ring gear speed to decrease, which causes the pump speed to decrease. This has the effect of reducing the total system hydraulic fluid flow rate, when compared to a conventional hydrostatic transmission of the same capacity, to approximately {fraction (1/3)} to {fraction (1/4)}, depending on planet set ratios used. This reduces the flow losses and noise levels normally associated with hydrostatic machines.  
         [0093]    As the motor displacement approaches zero and the pump displacement approaches its maximum, the pump speed approaches zero and motor speed approaches its maximum. When the motor reaches zero displacement it can no longer accept fluid flow so the pump can no longer displace fluid and therefore stops rotating, causing the ring gear (Rn) to stop rotating. The pump now acts as a reaction unit for the ring gear. In this case all the input torque is now transferred through the planet set, via the planet carrier and spur gears Sg 3  and Sg 3 , to the output shaft. Due to the ratio of the sun gear to ring gear, the output speed is decreased and the output torque increased, by a factor of 2.79:1 in the disclosed preferred embodiment. Naturally, the ratio would be different in designs with different size gears. As the pump has been stroked to its full displacement, hydraulic pressure required to react the input torque has been reduced to a minimum, thus reducing hydraulic leakage losses and hydraulic loading of bearings to a minimum.  
         [0094]    As all the power is now transferred through the planet set and spur gears Sg 3  and Sg 1 , and the hydraulics are acting only as a reaction unit to hold the ring gear, the efficiency is very high (95+%). The only losses are the normal gearset losses (approx. 2%), slippage on the pump due to leakage, and windage losses on the motor due to the fact it is spinning at output speed×(Sg 1 /Sg 2 ) with the unit at some pressure. To further increase the efficiency at this point a brake could be applied to the pump. This will help in two ways: first it will stop the input unit from slipping due to hydraulic leakage and second it will reduce the hydraulic system pressure to makeup pressure therefore reducing the load and hence windage loss of the motor. The brake could be actuated by makeup pressure or by electro-mechanical means.  
         [0095]    To drive the vehicle in reverse, the transmission is first placed in neutral, with the motor at maximum displacement and the pump at zero displacement. The selector switch is moved to “reverse” which causes the controller to stroke the pump displacement control in the opposite direction (i.e. a negative angle) causing fluid flow to go in the opposite direction. This causes 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 above, still acts in the forward direction. Therefore the total output torque, in reverse, can be expressed as:  
         Output Torque=Input Torque×[(1+( Rp/Sp ))−( Rp/Sp )×motor disp/pump disp×( Sg   1 / Sg   2 )] 
         [0096]    Due to the fact that the pump and motor rotate in the same direction, both swashplates are stroked in opposite directions, i.e. when the transmission is viewed from the top the pump swashplate is rotated clockwise as the motor swashplate is rotated counter clockwise, for forward ratios. The pump swashplate is rotated counter clockwise as the motor swashplate is held stationary for reverse ratios.  
         [0097]    The pump swashplate is connected to the pump control arm, which is connected to the pump control piston in such a way as to allow the pump control arm to pivot and slide relative to the pump control piston. As the pump control piston moves axially in its bore, the pump control arm and pump swashplate rotate about the pump swashplate axis. Similarly, the motor swashplate is connected to the motor control arm, which is connected to the motor control piston in such a way as to allow the motor control arm to pivot and slide relative to the pump control piston. As the motor control piston moves axially in its bore the motor control arm and motor swashplate rotate about the pump swashplate axis.  
         [0098]    System pressure is tapped off from the manifold via a shuttle valve and is fed continually to the back of the motor control piston  285 . The area of this piston is equal to  1 A. The pressure acting on this area biases the motor toward maximum displacement. System pressure is tapped off from the manifold via the same shuttle valve and is fed continually to the small annular area of the pump control piston  280 . The area of this annulus is also equal to  1 A. The pressure acting on this annular area biases the pump toward its maximum displacement in reverse. System pressure is tapped off from the manifold and is fed thru a modulating valve to the large annular area of the pump control piston  280  which is three times greater than the back of the motor control piston, or  3 A. When system pressure acts on this large annular area the force generated overcomes the force generated on the small annular side by a factor of 3 due its larger area. This causes the pump to stroke towards its maximum displacement in the forward direction.  
         [0099]    At a predetermined angle of pump displacement, the pump control piston  280  contacts the motor control piston  285  (which is being forced to the motor maximum displacement position as described above). When the pump control piston  280  contacts the motor control piston  285 , the force acting on the front of the pump control piston  280  overcomes the force acting on the annular side of the pump control piston  280  plus the force acting on the motor control piston  285 , by a factor of 1.5, and forces the motor  60  to stroke toward zero displacement whilst stroking the pump  50  toward its maximum displacement. The built-in lag in stroke angle between pump control piston contacting the motor control piston, allows the motor to stay at its maximum displacement whilst some displacement is achieved by the pump. Therefore, the pressure generated by the pump is allowed to act on the largest possible displacement of the motor, and hence generate the maximum amount of output torque possible. The amount of lag in stroke angle between pump control piston contacting the motor control piston, is equal to the minimum pump angle at which the pump can react full input torque whilst not exceeding the maximum allowable system pressure.  
         [0100]    As the modulating valve releases pressure from the large annular area of the pump control piston, the force acting on the motor control piston and the force acting on the small annular area of the pump control piston causes the motor to stroke toward its maximum displacement and the pump toward zero displacement. This will continue to happen until the motor eventually reaches its maximum displacement, when it can stroke no further. The pump will then continue stroking toward zero displacement until it reaches neutral. If the modulating valve further releases pressure from the large annular area of the pump control piston, the pump will continue to stroke into a reverse angle. By keeping the motor at its maximum displacement and stroking only the pump in reverse, the maximum possible torque is obtained from the motor but a limited speed capability, which is desirable.  
         [0101]    As stated above system pressure is tapped off from the manifold via a shuttle valve to control the pump and motor, but similarly, make-up pressure could be used to the same effect. This would however require larger piston diameters to generate enough force to smoothly and accurately control the pump and motor, and may therefore require a larger package.  
         [0102]    The modulating valve as mentioned above can be of several types, including a classic ‘leader-follower’ type spool valve actuated by a stepper motor or servomotor, or a solenoid operated spool valve etc.  
         [0103]    An advantage of this type of control regime is that it enables just one modulating valve (and associated control hardware, such as computer controls etc.), to control both the pump and motor from neutral thru final drive and into reverse. A simple, reliable and low cost control system is the result.  
         [0104]    Due to the fact the motor to pump displacement ratio can be infinitely large, at or around the neutral zone in forward and reverse, it is therefore theoretically possible to generate infinitely high pressures and output torque, and practically possible to generate pressure and output torque which exceeds the capability of the materials to contain them. Obviously these have to be limited to reasonable values, as determined by the structural limitations of the transmission. Torque limitation is achieved by use of a pressure relief valve mounted in the manifold, limiting the maximum pressure the pump can generate, and hence the maximum output torque. Since the pump will be at relatively small displacements when the pressure is at such high levels, the flow rate thru the relief valve will be at acceptable levels.  
         [0105]    Alternatively, the system can be inherently torque limited by designing the pump and motor to have a leakage rate that, at a specified pressure, is equal to the pump discharge. The leakage functions as a pressure relief and prevents the pump from generating any more pressure than that specified pressure. The transmission will then reach a ‘stall’ torque. A certain leakage rate is necessary for hydrostatic bearing interface cooling and lubrication anyway, so designing a leakage rate which also provides a torque limiting function, would have the advantage of doing both functions without need for a separate relief valve.  
         [0106]    There is a minimum pump angle at which the pump can react full input torque without exceeding the maximum allowable system pressure, and hence maximum output torque. At pump angles less than these, the output torque will not increase as the maximum pressure is limited as described above, but the input to output speed ratio will continue to decrease and will approach infinity as the pump angle becomes infinitely small.  
         [0107]    The stated and other benefits of the invention are also achieved in a bent axis design shown in FIGS.  60 - 71 . The gearing  85  and input/output arrangement of this embodiment is similar all significant respects to the embodiment of FIGS.  1 - 59 . The only significant difference is that pump and motor cylinder blocks  330 P and  330 M in this bent axis embodiment, as best shown in FIGS.  60 - 63 , are turned around with pistons  332  facing a manifold  335  and engaged in pump and motor torque rings  337 P and  337 M running against the manifold  335 , as in Applicant&#39;s International Patent Application PCT/US98/24053 entitled “Hydraulic Machine”, the disclosure of which is incorporated herein by reference. In this embodiment, as shown in FIGS. 61 and 62, the manifold  335  is in an interior end wall of a front housing  340 , shown in FIGS.  64 - 68 , which also supports bearings for the pump and motor shafts  140  and  200 . The pump and motor cylinder blocks  330 P and  330 M in this design rotate against non-rotating tilting yoke seats  342 P and  342 M. Torque is input and output to this bent axis unit through splined engagement of the pump and motor shafts  140  and  200  with the torque rings  337 P and  337 M.  
         [0108]    As shown in FIGS. 60, 62 and  63 , the back face of the pump and motor cylinders  330 P and  330 M each bear against a flat face of the yoke seat  342 . Two arms  344  are attached to the yoke seat  342 , one on each side, and extend forward to gudgeons  345  which are fixed on trunnions  346  pivotally supported in a rear housing  350 , shown in FIGS. 69 and 70. The outer ends of the trunnions  346  are supported in bosses in the exterior of the rear housing  350 , and the inner ends of the trunnions  346  are supported in bosses in internal webs  352  in the housing  350 . The trunnions at the top of the housing  350  protrude beyond the housing and are fixed to the proximal ends of two control crank arms  265  and  270 , of distal ends of which extend inwardly toward each other and are engaged in the ends of control pistons  280  and  285 .  
         [0109]    This bent axis embodiment is advantageous because it has greater efficiency and power density, can result in a reduction in size, weight complexity and cost, and has the ability to run faster than a same size swashplate unit. It is thus possible to use gear ratios that make the bent axis unit spin faster, thereby increasing its torque and power output. The greater power throughput makes it possible to design the unit with smaller hydrostatic units (to achieve the same torque at the same pressure) or run it at a lower pressure and hence use smaller and lighter supporting structures since the loads will be less, or the unit can be made available at the same size with higher torque capacity.  
         [0110]    Turning now to FIG. 72, a schematic diagram of a third embodiment of the invention is shown particularly for use in a front wheel drive transaxle arrangement having a substantial offset between the prime mover output shaft and the output differential by which the front axle is driven. This particular design was made light weight and inexpensive for a European microcar, but could also be adapted for small automobile applications as noted below.  
         [0111]    The transmission is shown in FIG. 72 in neutral, with the pump  50  set at zero displacement and the motor  60  at maximum displacement. Both the pump  50  and motor  60  are simultaneously controlled in this case, although they could be independently controlled.  
         [0112]    As shown in FIG. 72, and also in FIGS. 74, 76 and  79 - 81 , the input from the prime mover  55  is connected through an input spline coupling  354  to an input shaft  355 . The input shaft  355  extends through a drive tube  357 , shown in detail in FIGS. 82A and 82B, and has an intermediate spline  356  adjacent its inner end that engages and drives an interior spline  358  at the inner end of the drive tube  357 . This input shaft  355  is used to accommodate mis-alignments and eccentricities between the engine and transmission whilst being a torsionally rigid coupling. Since the two splines are relatively far apart, a small amount of clearance in the splines will accommodate these mis-alignments and eccentricities.  
         [0113]    The input shaft  355 , shown in detail in FIG. 83 has an end spline  359  that engages and drives a make-up pump  366 , as shown in FIGS.  79 - 81 . The drive tube  357  has an integral sun gear  360  of a planet set  365  driving a series of planet gears  362  engaged with an encircling ring gear  367  of the planet set  365 . As shown in FIG. 78, the ring gear  367  has an integral spur gear  368  which drives a spur gear  370  connected to the pump cylinder block shaft  371 , shown in detail in FIG. 84. The planet gears  362  are mounted in a planet carrier  372  of the planet set  365  which is machined in its outer periphery as a spur gear  374 . The spur gear  374  is driven by a spur gear  376  splined to the motor output shaft  380 , shown in detail in FIG. 85, which is driven by a spline connection with the motor cylinder block  206 M. The planet carrier  372  is also splined to a transmission chain sprocket  384 , as shown in FIGS.  79 - 81 , which is coupled via a drive chain  386  to a differential chain sprocket  388  connected to the output differential  390 , as shown in FIG. 76.  
         [0114]    One advantage of driving the pump by way of spur gears  368  and  370  is that the ratio between these spur gears can be selected to spin the pump faster than the ring gear speed. In the first embodiment shown in FIGS. 1 and 11, the input is connected to the sun gear and the pump is driven directly from the ring gear, so the pump will spin at a slower speed and with a higher torque than the input shaft. This can be disadvantageous for the pump as it will generate a higher pressure to react the input torque, thus giving greater leakage and higher bearing loads. It also means that the maximum pump speed will be lower than its design maximum speed, so the full potential horse power of the unit will not be produced. By using a ratio between the spur gears  368  and  370 , it is now possible to spin the pump at it&#39;s maximum design speed. Therefore the maximum potential horse power can be extracted from the unit and the system pressure will be lower at any given input torque.  
         [0115]    The chain  386  is used to drive the output differential  390  to facilitate spacing the front wheel drive shafts on a centerline  393  far from the engine centerline  395  to accommodate an existing installation, without using a series of gears to achieve the same center distance. Naturally, a series of gears could be used and a different centerline spacing could be used to provide closer coupling between the transmission/engine drive centerline and differential  390 .  
         [0116]    The pump and motor cylinder blocks  206 P and  206 M lie on parallel axes coincident with the axes of their shafts  371  and  380 , as shown in FIG. 78. Pistons  400  in the cylinders of each cylinder block engage a thrust ring which rotates with the cylinder block and is mounted by way of a thrust bearing  404  on a non-rotating, tilting swash plate  408 . The displacement of the pump  50  and motor  60  can be varied by adjusting the tilt angle of the swashplate  408  by a crank arrangement. The swashplate  408  is supported in a cradle bearing  410  on the rear housing  415  of the transmission, shown in FIGS.  86 - 91 . The cradle bearing is preferably provided with a low friction polymer surface such as PTFE or the like. The rear housing is connected to a middle housing  417 , shown in FIGS.  92 - 96  by multiple machine screws  419  to provide a reaction path for the axial forces exerted by the pump  50  and motor  60  through the housing and back to a manifold  420  supported by an internal transverse bulkhead  422  inside the middle housing, as shown in FIG. 95.  
         [0117]    The manifold  420 , shown in FIGS.  97 - 103 , is held against the bulkhead  422  by compression coil springs (not shown) inside the hollow pistons in the pump and motor cylinder blocks  206  which also maintains sealing contact of the cylinder blocks  206  with the manifold to enable system pressure to develop when the transmission is started. During operation, the axial forces exerted by the pump  50  and motor  60  maintain the manifold forcefully engaged with the bulkhead  422 .  
         [0118]    The manifold  420  has two flat round faces  425 P and  425 M in contact with the flat faces of the pump and motor cylinder blocks  206 P and  206 M. Each face  425 P and  425 M has a pair of opposed curved slots  428  and  430  for conveying high pressure fluid on the pressure stroke from the pump cylinder block  206 P to the motor cylinder block  206 M, and for conveying spend low pressure fluid displaced from the motor cylinder block  206 M back for recharging the pump cylinder block  206 P on suction stroke. Four bosses  435  on the manifold  420  hold check valves for passing make-up fluid from the make-up pump  366  through passages  437  in the bulkhead  422 , and for passing high pressure fluid to the control unit  450  through a passage  438  in the bulkhead  422 . Four valves are needed instead of just two because the high and low pressure sides switch when the transmission is back driven through the vehicle wheels during downhill or decelerating travel when engine braking is used. The hydraulic operation of the pump and motor  60  in this transmission is the same as that described in the first embodiment.  
         [0119]    The control unit  450  operates basically like the control units in the first and second embodiments. Due to the fact that the pump  50  and motor  60  rotate in opposite directions, both swashplates  408  are stroked in the same direction for forward ratios. When the transmission is viewed from the top, as in FIGS. 105 and 107, the pump swashplate  408 P is rotated counter-clockwise as the motor swashplate  408 M is rotated counter-clockwise. The pump swashplate  408 P is rotated clockwise as the motor swashplate  408 M is held stationary for reverse ratios.  
         [0120]    The pump swashplate  408 P is connected to a pump control arm  454  which is connected to a pump control piston  458  in such a way as to allow the pump control arm  454  to pivot and slide relative to the pump control piston  458 . As shown in FIG. 104, the pivot axis  460  of the pump control arm  454  coincides with the axis of rotation of the pump swashplate. As the pump control piston  458  moves axially in its bore  464 , the pump control arm  454  and pump swashplate  408 P rotate about the pump swashplate axis. The motor swashplate  408 P is connected to a motor control arm  466  which is connected to a motor control  468  piston in such a way as to allow the motor control arm  466  to pivot and slide relative to the pump control piston  468 . As the motor control piston  468  moves axially in its bore  470  the motor control arm  466  and motor swashplate  408 M rotate about their common axis.  
         [0121]    System pressure is tapped off from the manifold through one of the check valves in the manifold and is fed continually to the motor control cylinder  470  behind the motor control piston  468 . The area of the face of the motor control piston  468  is about one third of the area of the face of the piston control piston  458 . The pressure acting on this area biases the motor continually toward its maximum displacement. System pressure is tapped off from the manifold via the same check valve and is fed continually to the small annular area  472  of the pump control piston. The area of this annulus is equal to the area of the motor control piston  468 , and the pressure acting on this area biases the pump continually toward its maximum displacement in reverse (i.e. to rotate the pump swashplate  408 P clockwise) as shown in FIGS.  105 - 107 .  
         [0122]    System pressure is tapped off from the manifold and is fed thru the modulating valve  474  to the large annular area  476  of the pump control piston  458 . The area of this large annular face  476  of the pump control piston is equal to three times the area of the face of the motor control piston  468 , so when system pressure acts on this annulus  476 , the force generated overcomes the force generated on the small annular side by a factor of 3 due its larger area. This strokes the pump towards its maximum displacement in the forward direction.  
         [0123]    As shown in FIGS.  108 - 110 , at a predetermined angle of pump displacement, the pump control piston  458  contacts the motor control piston  468  (which is being forced to the motor maximum displacement position as described above). When the pump control piston  458  contacts the motor control piston  468 , the force acting on the front of the pump control piston overcomes the force acting on the annular side of the pump control piston and the force acting on the motor control piston, by a factor of 1.5, and strokes the motor toward zero displacement whilst stroking the pump toward its maximum displacement. The built in lag in stroke angle between pump control piston  458  contacting the motor control piston  468  allows the motor  60  to stay at its maximum displacement whilst some displacement is achieved by the pump  50 , thereby allowing the pressure generated by the pump to act on the largest possible displacement of the motor, and hence generating the maximum amount of output torque possible. The amount of lag in stroke angle between pump control piston contacting the motor control piston is equal to the minimum pump angle at which the pump can react full input torque whilst not exceeding the maximum allowable system pressure. Continued movement of the pump control piston  458  to the fully extended position shown in FIGS.  111 - 113  shifts the pump swashplate  408 P to maximum displacement position and the motor swashplate to its zero displacement position, resulting in hydraulic lock-up and full mechanical drive through the transmission.  
         [0124]    As pressure is released from the large annular area of the pump control piston, by the modulating valve, the force acting on the motor control piston and the force acting on the small annular area of the pump control piston causes the motor to stroke toward its maximum displacement and the pump toward zero displacement. This will continue to happen until the motor eventually reaches its maximum displacement, when it can no longer stroke. The pump will then continue stroking toward zero displacement until it reaches neutral, shown in FIGS.  105 - 107 . If the modulating valve further releases pressure from the large annular area of the pump control piston, the pump will continue to stroke into a reverse angle, as shown in Figs. By keeping the motor at its maximum displacement and stroking only the pump in reverse, the maximum possible torque from the motor is attained but with a limited speed capability, which is desirable.  
         [0125]    As stated above system pressure is tapped off from the manifold via a shuttle valve to control the pump and motor, but similarly make-up pressure could be used to the same effect. This would however require larger piston diameters to generate enough force to smoothly and accurately control the pump and motor, and may therefore pose some packaging problems.  
         [0126]    The modulating valve as mentioned above can be of several types, including a classic leader-follower type spool valve actuated by a stepper motor, or a solenoid operated spool valve etc.  
         [0127]    The advantage of this type of control regime is that it enables just one modulating valve (and associated control hardware, such as computer controls etc.), to control both the pump and motor from neutral thru final drive and into reverse. Thus reducing cost and complexity of the control system. It also has the advantage of mechanically linking the pump and motor swashplate displacements together eliminating possible control errors that may occur if each swashplate is individually controlled.  
         [0128]    The use of the front wheel drive transaxle shown in this third embodiment could be readily be modified to incorporate the yoke support for the swashplate as shown in the first embodiment of FIG. 5, or the yoke supported bent axis arrangement of the pump and motor as shown in the second embodiment shown in FIG. 60. Obviously, numerous other modifications, combinations and variations of the preferred embodiments described above are possible and will become apparent to those skilled in the art in light of this specification. For example, many functions and advantages are described for the three preferred embodiments, but in some uses of the invention, not all of these functions and advantages would be needed. Therefore, we contemplate the use of the invention using fewer than the complete set of noted functions and advantages. Moreover, several species and embodiments of the invention are disclosed herein, but not all are specifically claimed, although all are covered by generic claims. Nevertheless, it is our intention that each and every one of these species and embodiments, and the equivalents thereof, be encompassed and protected within the scope of the following claims, and no dedication to the public is intended by virtue of the lack of claims specific to any individual species. Accordingly, it is expressly intended that all these embodiments, species, modifications and variations, and the equivalents thereof, are to be considered within the spirit and scope of the invention as defined in the following claims, wherein we claim:

Technology Category: f