Abstract:
A continuously variable power transmission has a fixed mounting plinth, by which the transmission is mounted to a fixed structure, and an output shaft, to which an output device can be connected. A pintle having a flat radial slab and an axial arbor supports a radial piston pump rotor an one end of the arbor and a radial piston motor rotor on the other end of the arbor. A driven pulley surrounds the transmission and constitutes its exterior shell within which working fluid is contained. The pulley is coupled to a ring gear of a planetary gear set having a planet carrier with planet gears engaged between the ring gear and a sun gear. The sun gear is coupled to and drives the pump rotor, and the carrier is coupled to the output shaft. Working fluid pressurized by the pump is conveyed to the motor rotor to generate torque in the motor which is carried back to the planet carrier.

Description:
This relates to U.S. Provisional Application 60/243,956 filed on Oct. 26, 2000. This invention relates to continuously variable power drive transmissions, and more particularly to an economical radial ball piston hydromechanical continuously variable drive transmission. 
    
    
     BACKGROUND OF THE INVENTION 
     Continuously variable hydrostatic power transmissions have been available for many years and are in use in applications in which their noise and inefficiency have not been seriously objectionable, such as lawn and garden tractors. However, such noise and inefficiency have recently become unacceptable in lawn and garden equipment, and there are numerous other applications for economical low and medium power continuously variable transmissions that require efficient and quiet operation. 
     SUMMARY OF THE INVENTION 
     This invention provides an economical continuously variable power transmission that is efficient and quiet. 
     The continuously variable power transmission according to this invention includes a fixed support plinth at one end by which the transmission is mounted to a fixed support structure, and an output shaft at the other end. A pintle having a flat radial slab and an axial arbor is bolted to the pinth and supports a radial piston pump rotor on one end of the arbor and a radial piston motor rotor on the other end of the arbor. Cylindrical cam rings are mounted around the pump and motor rotors to engage piston balls in the pump and motor rotor cylinders for pumping and torque generation. A driven pulley, having a drive surface engaged with a driving element for coupling rotating mechanical power to the transmission, surrounds the transmission and constitutes its exterior shell within which working fluid such as oil is contained. The driven pulley is supported for rotation on bearings at its two ends. At the support end, the pulley is supported on the plinth, and at the other end is supported on a fixed housing attached to the plinth. The output end of the driven pulley is coupled to a ring gear of a planetary gear set having a planet carrier with planet gears engaged between the ring gear and a sun gear. The sun gear is coupled to and drives the radial piston pump rotor, and the planet carrier is coupled to the output shaft, so the reaction torque from the torque applied to the pump rotor is applied directly to the output shaft. Working fluid pressurized by rotation of the pump is conveyed through channels in the pintle arbor to the motor rotor to pressurize the motor cylinders and generated torque in the motor which is carried via a motor drum back to the planet carrier and thence to the output shaft. The transmission ratio is set by adjusting the radial position of the cam rings which are mounted on a pivot pin at the lower end of the pintle slab and are tilted to the desired radial positions by action of a control system having actuator pistons mounted in upper part of the pintle slab. The cam rings may be coupled together at the top, opposite the pivot pin, by a lever rod mounted for swiveling in the top of the pintle slab, so tilting of the motor cam ring in one direction causes the lever rod to swivel about its swivel mounting in the pintle and tilt the pump cam ring in the opposite direction. 
    
    
     DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic diagram of an engine compartment of a vehicle in which a transmission in accordance with this invention issued as a demand fan drive; 
     FIG. 2 is a perspective view of the transmission shown in FIG. 1, detached from the fan and the support mount in the engine compartment and viewed from the fan connection end; 
     FIG. 3 is a perspective view of the transmission shown in FIG. 2, viewed from the support connection end; 
     FIGS. 4 and 5 are sectional elevations along lines  4 — 4  and  5 — 5  in FIGS. 7 and 6, respectively, through the axial centerline of the transmission; 
     FIGS. 6 and 7 are sectional elevations along lines  6 / 7 — 6 / 7  in FIG. 4, at two different settings of the transmission, respectively; 
     FIGS. 8 and 9 are sectional elevations along lines  8 / 9 — 8 / 9  in FIG. 4, at the two different settings of the transmission, respectively, shown in FIGS. 6 and 7; 
     FIG. 10 is a perspective view of the output shaft shown in FIG. 4; 
     FIG. 11 is an end elevation of the output shaft shown in FIG. 10; 
     FIG. 12 is a sectional elevation along lines  12 — 12  in FIG. 11; 
     FIG. 13 is a perspective view of the support plinth shown in FIG.  4 ;: 
     FIG. 14 is an end elevation of the support plinth shown in FIG. 13; 
     FIG. 15 is a sectional elevation of the support plinth along lines  15 — 15  in FIG. 14; 
     FIG. 16 is a perspective view of the pulley flange shown in FIG. 4; 
     FIG. 17 is an end elevation of the pulley flange shown in FIG. 16; 
     FIG. 18 is a sectional elevation of the pulley flange along lines  18 — 18  in FIG. 17; 
     FIG. 19 is a perspective view of the fixed housing shown in FIG. 4; 
     FIG. 20 is an end elevation of the fixed housing shown in FIG. 19; 
     FIG. 21 is a sectional elevation of the fixed housing along lines  21 — 21  in FIG. 20; 
     FIG. 22 is a perspective view of the planet carrier shown in FIG. 4, viewed from the output end; 
     FIG. 23 is a perspective view of the planet carrier shown in FIG. 22, viewed from the other end; 
     FIG. 24 is an end elevation of the planet carrier in FIG. 23; 
     FIG. 25 is a sectional elevation of the planet carrier along lines  25 — 25  in FIG. 24; 
     FIG. 26 is a perspective view of the ring gear for the planetary gear set shown in FIG. 4; 
     FIG. 27 is an end elevation of the ring gear shown in FIG. 26; 
     FIG. 28 is a sectional elevation of the ring gear along lines  28 — 28  in FIG. 27; 
     FIG. 29 is a perspective view of one of the planet gears which are mounted in the planet carrier shown in FIG. 22 an used in the planetary gear set shown in FIG. 4; 
     FIG. 30 is an end elevation of the planet gear shown in FIG. 29; 
     FIG. 31 is a side elevation of the planet gear shown in FIG. 29; 
     FIG. 32 is a perspective view of the sun gear of the planetary gear set shown in FIG. 4; 
     FIG. 33 is an end elevation of the sun gear shown in FIG. 32; 
     FIG. 34 is a sectional elevation of the sun gear along lines  34 — 34  in FIG. 32; 
     FIG. 35 is a perspective elevation of the pump drive ring shown in FIGS. 4 and 43; 
     FIG. 36 is an end elevation of the pump drive ring shown in FIG. 35; 
     FIG. 37 is a sectional elevation of the pump drive ring along lines  37 — 37  in FIG. 36; 
     FIG. 38 is a perspective view of the rotor used in both the pump and motor shown in FIG. 4; 
     FIG. 39 is an end elevation of the rotor shown in FIG. 38; 
     FIG. 40 is sectional plan view of the rotor along lines  40 — 40  in FIG. 39; 
     FIG. 41 is a side elevation of the rotor shown in FIG. 38; 
     FIG. 42 is a sectional elevation along lines  42 — 42  in FIG. 41; 
     FIG. 43 is a perspective view of the pump and motor assembly shown in FIG. 4; 
     FIG. 44 is a perspective view of the pintle shown in FIG. 4, viewed from the output end; 
     FIG. 45 is a perspective view of the pintle shown in FIG. 44, viewed from the input end; 
     FIG. 46 is an elevation of the pintle shown in FIG. 44; 
     FIG. 47 is a sectional side elevation of the pintle along lines  47 — 47  in FIG. 46; 
     FIG. 48 is a sectional plan view of the pintle along lines  48 — 48  in FIG. 46; 
     FIG. 49 is an end elevation of the pintle as viewed from the right in FIG. 46; 
     FIG. 50 is a sectional end elevation of the pintle along lines  50 — 50  in FIG. 49; 
     FIG. 51 is a sectional view on a diagonal section along lines  51 — 51  in FIG. 46; 
     FIG. 52 is a perspective view of the motor drum shown in FIG. 4, as viewed form the output end; 
     FIG. 53 is an end elevation of the motor drum shown in FIG. 52; 
     FIG. 54 is a sectional side elevation of the motor drum along lines  54 — 54  in FIG. 53; 
     FIG. 55 is an exploded perspective view of the planet carrier, the motor drum, and the motor drive disc shown in FIG. 4, showing the alignment for the cog drive connection between the planet carrier and the motor drum, and showing the alignment for the castellated joint between the motor drum and the motor drive disc; 
     FIG. 56 is a perspective view of the motor drive disc shown in FIG. 55; 
     FIG. 57 is an eld elevation of the motor drive disc shown in FIG. 56; 
     FIG. 58 is a sectional side.elevation of the motor drive disc along lines  58 — 58  in FIG. 57; 
     FIG. 59 is a perspective view of the pulley end cap shown in FIG. 4; 
     FIG. 60 is an end elevation of the pulley end cap shown in FIG. 59; 
     FIG. 61 is a sectional side elevation of the pulley end cap along lines  61 — 61  in FIG. 60; 
     FIG. 62 is a perspective view of the support flange shown in FIG. 4; 
     FIG. 63 is an end elevation of the support flange shown in FIG. 62; 
     FIG. 64 is a sectional side elevation of the support flange along lines  64 — 64  in FIG. 63; 
     FIG. 65 is a perspective view of the one of the pistons shown in FIG. 4; 
     FIG. 66 is an end view of the piston shown in FIG. 65; 
     FIG. 67 is a sectional side view of the piston along lines  67 — 67  in FIG. 66; 
     FIG. 68 is a perspective view of one of the two cam races shown in FIG. 4; 
     FIG. 69 is a end elevation of the cam race shown in FIG. 68; 
     FIG. 70 is a sectional side elevation of the cam race along lines  70 — 70  in FIG. 69; 
     FIG. 71 is a perspective view of the pump cam ring shown in FIGS. 4 and 43; 
     FIG. 72 is an end elevation of the pump cam ring shown in FIG. 72; 
     FIG. 73 is a sectional side elevation of the pump cam ring along lines  73 — 73  in FIG. 72; 
     FIG. 74 is a perspective view of the motor cam ring shown in FIGS. 4 and 43; 
     FIG. 75 is an end elevation of the motor cam ring shown in FIG. 74; 
     FIG. 76 is a sectional side elevation of the motor cam ring along lines  76 — 76  in FIG. 75; 
     FIG. 77 is a perspective view of one of the two cylindrical joint plugs shown in the top of the am rings in FIGS. 4 and 43; 
     FIG. 78 is an end elevation of the cylindrical joint plug shown in FIG. 77; 
     FIG. 79 is a sectional side elevation of the joint plug along lines  79 — 79  in FIG. 78; 
     FIG. 80 is a perspective view of the control piston shown in FIGS. 8 and 9; 
     FIG. 81 is an end view of the control piston shown in FIG. 80; 
     FIG. 82 is a sectional view of the control piston along lines  82 — 82  in FIG. 81; 
     FIG. 83 is a perspective view of the load piston shown in FIGS. 8 and 9; 
     FIG. 84 is an end view of the open end of the load piston shown in FIG. 83; 
     FIG. 85 is a sectional view of the load piston along lines  85 — 85  in FIG. 84; 
     FIG. 86 is a perspective view of the control spool shown in FIGS. 8 and 9; 
     FIG. 87 is a sectional view on a section plane intersecting and parallel to the axis of the control spool shown in FIG. 86; 
     FIG. 88 is a perspective view of spool piston shown in FIGS. 8 and 9; 
     FIG. 89 is an end view of the open end of the spool piston shown in FIG. 88; 
     FIG. 90 is a sectional view of the spool piston along lines  90 — 90  in FIG. 89; and 
     FIG. 91 is a schematic diagram of the transmission shown in FIG.  4  and the drive arrangement for the transmission shown in FIG.  1 . 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Turning now to the drawings, and more particularly to FIG. 1 thereof, a continuously variable transmission  50  is shown in a schematic layout of a vehicle engine compartment. The transmission  50  is used in this application as a demand fan drive unit for driving a cooling fan  52  behind the vehicle engine coolant radiator  51  for drawing cooling air through the radiator. The transmission  50  includes a support plinth  53  by which the transmission  50  is mounted at a support end of the transmission on a fixed support  54  in the engine compartment in a position that aligns an input pulley  56 , driven by the vehicle prime mover  58 , with a driven pulley  60  surrounding the transmission  50  concentrically. A drive belt  62 , trained around the input pulley  56  and the driven pulley  60 , transmits mechanical power in the form of torque at the rotation speed of the input pulley  56  to the driven pulley  56 . The transmission  50  converts that mechanical power to rotation of the fan  52  at the desired speed, as set by the adjustable controls, described in detail below. The drive belt  62  could be replaced with a drive chain or a gear train for this or other applications of this transmission, such as a drive transmission for a small automobile such as a city car, or a lawn and garden tractor, between the prime mover and the drive wheels. 
     Turning now to FIGS. 2 and 3, the output end of the transmission  50  is shown having a fan connection flange  64  by which the fan  52 , or other driven load as noted above, is connected to and driven by the transmission  50 . The fan connection flange  64  is integral with an output shaft  66 , shown in FIG.  4  and shown in detail in FIGS. 10-12 and described in detail below. The support plinth  53 , shown in FIGS. 3 and 4, and shown in detail in FIGS. 13-15 and described in detail below, has a square base plate  68  with four holes  70  by which the transmission is attached to the fixed support  54  in the engine compartment. 
     As shown in FIGS. 4 and 5, the driven pulley  60  is connected to a bell-shaped pulley flange  75 , shown in detail in FIGS. 16-18, by machine screws  77  extending through peripheral holes  79  in the outside peripheral edge of the pulley flange  75  and threaded into tapped holes in the edge of the driven pulley  60 . The inside peripheral edge of the pulley flange  75  ends in an integral inner axial collar  80  having a radially outside cylindrical surface  82  that may be machined or ground to receive with a snug fit an inner bearing  85 . The bearing  85  supports the pulley flange  75  on a fixed housing  88 , shown in detail in FIGS. 19-21, for rotation about the central axis  90  of the transmission  50 . The collar  80  of the pulley flange  75  also has a radially inside cylindrical surface  92 , that likewise may be machined or ground to receive with a snug fit an outer bearing  93 . The outer bearing  93  supports a planet carrier  95 , shown in detail in FIGS. 22-25 and described in detail below. The planet carrier is connected to a splined section  156  of the output shaft  66  and the output shaft is held in place against axial translation by a snap ring (not shown) in a snap ring groove  96  on the splined section  156  inside the planet carrier  95 . Another way to hole the output shaft against axial translation is to attach an oil seal cup  98  to the collar  80  of the pulley flange. 
     The collar  80  of the pulley flange  75  has a radially inside cylindrical surface  97  that is axially adjacent and inside the surface  92 . The inside cylindrical surface  97  of the collar is splined and receives a ring gear  100 , shown in detail in FIGS. 26-28, that is splined on its radially exterior surface  102  for torsional coupling to the pulley flange  75 , and is provided with gear teeth on its radially inside surface  104  for engaging a series of planet gears  106 , shown in FIGS. 29-31, mounted in the planet carrier  95 . The planet gears  106  are engaged between the ring gear  100  and a sun gear  110 , shown in FIGS. 32-34. 
     The sun gear  110  has an exterior surface provided with the usual gear teeth  112  and an inner ring  113  with exterior splines that are engaged with matching splines  114  around the inner periphery of a pump drive ring  115 , shown in detail in FIGS. 35-37. The pump drive ring  115  has radially extending teeth  118  on its outside periphery that define slots  120  between the teeth  118 . The slots  120  receive cogs  124  on the side of a pump rotor  125 , shown in detail in FIGS. 38-42, and shown assembled with the pump drive ring  115  in FIG. 43, by which the pump rotor is driven in rotation about the axis  90  by the sun gear  110  by way of the pump drive ring  115 . 
     The pump rotor  125  has an axial bore  127 , which receives a pump rotor arbor  128  of a pintle  130 , shown in detail in FIGS. 44-51. The opposite side of the pintle  130  also has an axial arbor  132  which is received into the axial bore of a motor rotor  135 , also shown in detail in FIGS. 38-42 since the pump rotor  125  and the motor rotor  135  are identical. The pinde  130  has a stepped axial bore  138  by which the pintle is fixedly mounted concentric with the axis  90  on a stub shaft  140  projecting axially from the support plinth  53 , as shown in FIGS. 4,  5 ,  13  and  15 . The pintle is attached to the plinth  68  by bolts (not shown) in aligned bolt holes  141 , as shown in FIG.  5 . The other end  139  of the stepped bore  138  receives the inner end of the output shaft  66  and supports and stabilizes the inner end of the output shaft on needle bearings  142 . 
     A cup-shaped motor drum  145 , shown in FIGS. 4 and 5 and shown in detail in FIGS. 52-54, has a cylindrical wall  147  disposed concentrically inside the fixed housing  88 , and a radially extending end wall  148 . The end wall  148  has anaxial opening  149  having radially extending teeth  150  which fit between and drive cogs  152  on the outer periphery of the inner face of the planet carrier  95 , as shown in FIGS. 4 and 23. The engagement of the teeth  150  between the cogs  152  transmits torque generated in the motor  130  to the planet carrier  95  and thence to the output shaft, by way of a spline coupling between splines  154  in the bore of the planet carrier  95  and splines  156  on the output shaft, as shown in FIG.  10 . The cylindrical wall  147  of motor drum  145  ends in a castellated free circular edge  158  which is coupled with and driven by a corresponding castellated free edge  159  of a motor drive disc  160 , shown in detail in FIGS. 56-58. The motor drive disc  160  is fastened to the outer face of the motor rotor  130  by screws  162 , as shown in FIG. 4, so torque generated in the motor rotor  130  is transmitted directly to the motor drum  145  through the castellated joint  158 / 159 , and then to the planet carrier  95  through the cogs  152 , as shown in FIG.  55 . 
     At the support end of the transmission  50 , the driven pulley  60  is supported by a pulley end cap  165 , shown in detail in FIGS. 59-61, that is held in a shallow groove around the inside surface of the driven pulley  60  by a snap ring  167 , as shown in FIGS. 4 and 5, or it could be fastened to an inwardly extending flange on the driven pulley by machine screws. The outer edge of the pulley end cap  165  has an outwardly opening groove  168  for receiving a seal ring (not shown) such as a conventional elastomeric static seal ring, and the inner edge  169  is supported on a needle bearing  170 . At the support end of the transmission  50 , the fixed housing  88  is supported on a support flange  175 , shown in detail in FIGS. 62-64. The support flange  175  includes an inner ferrule  177  mounted in a fixed position on a cylindrical mount  179  of the support plinth  53  between the square base plate  68  and the stub shaft  140 , against rotation relative to the cylindrical mount  179 . The radially outer edge of the support flange  175  is provided with a number of integral bushings  180  which receive machine screws  182  that are threaded into tapped holes in bosses  184  in the free edge of the cylindrical wall  186  of the housing  88  to connect the support flonge  175  rigidly to the housing  88 . 
     A pitot tube  190  is formed as an elongated radially extending boss on the outside face having a tube bore communicating between the inside of the inner ferrule  177  and the outside edge of the support flange  175 . The pitot tube  190  is used to drain excess lubricating oil and oil leakage from the pump and motor out of the case of the transmission, as discussed in detail below. 
     Referring now to FIGS. 38-42, the pump includes the pump rotor  125  which has a number of radial pump cylinders  195 , each having a radially inner opening  197  communicating with the bore  127  of the pump rotor  125 . In this embodiment, there are 7 pump cylinders. A pump piston  200 , shown in FIGS. 4 and 5 and shown in detail in FIGS. 65-67 is fitted into each pump cylinder  195 . Each piston  200  has a cylindrical outside surface that fits snuggly into its pump rotor cylinder  195 , and has a semi-spherical outer face, forming a pocket  202  that receives a ball  205 . The balls  205  are intended to rotate in the pockets  202 , floated in an oil film that is pressurized by the system pressure created by reciprocation of the pistons in the cylinders. The system pressure is communicated through the piston between the inside face of the piston and the ball/pocket interface by way of a central axial hole  207  and two side holes connected by a circular groove  210 . The pistons are made of bronze, although other conventional materials and even high-strength plastics could be used. The sealing of the pistons in the cylinders could be improved, at a greater cost, by the use of piston rings. 
     Reciprocation of the pump pistons  260  is effected by rotating the pump rotor  125  with the balls  205  of the pistons engaged with a hardened cam race  215 , shown in FIGS. 68-70, mounted in a pump cam ring  220 , shown in FIGS. 71-73. The cam race  215  has a concave inside surface  217  that is shaped to match the surface of the balls  205 , thereby mining the contact pressure of the balls  205  on the cam race  215 . The pump cam ring  220  is mounted on the pintle  130 , as shown in FIGS. 4-7 and  43 , in a position surrounding the pump rotor  125 , with the cam race  215  aligned radially over the balls  205 . 
     The motor, also shown in FIGS. 4 and 5, is structurally similar to the pump. It includes the motor rotor  130 , motor pistons  201  and motor piston balls  206  inside a motor cam race  216 , all identical to the corresponding parts in the pump. The motor cam race  216  is mounted in a motor cam ring  221 , shown in FIGS. 74-76, which is a mirror image of the pump cam ring, and also includes two attachment bosses  223  which receive pivot pins by which push blocks  225  are pivotally attached to the motor cam ring  221 . The push blocks  225  are engaged by control pistons to control the transmission ratio, as described in greater detail below. 
     As shown in FIGS. 4,  6 - 9  and  43 , the cam rings  220  and  221  are pivotally supported on the pintle  130  by way of a pivot pin  228  that is mounted in a bore  230  in the pintle  130  on an axis parallel to and below the stepped bore  138 . Effectively, the cam rings  220  and  221  are mounted on a stationary pivot pin in the transmission  50 . The top end of each of the cam rings  220  and  221  has a cylindrical joint plug  233 , shown in FIGS. 77-79, fitted into a cylindrical recess opening in the top of the cam ring  220  and  221 . The cylindrical joint plugs  233  each have a diametrical hole  234  drilled through the cylindrical join plug, each of which receives one of two opposite ends of a cylindrical lever rod  235 . The lever rod  235  is mounted at its longitudinal center in a center cylindrical joint plug  237  that is mounted for rotation about its vertical axis in a cylindrical recess  238  in the top of the pintle  130 . Each side of the pintle  130  on either side of the cylindrical recess  238  has an elongated opening  239  that is tapered to allow the cylindrical lever rod  235  to swivel about the vertical axis of the center cylindrical joint plug  237  in the opening  239  when the cam rings are rotated in opposite directions about the pivot pin  228  by the control pistons, as described below. Since the pump and motor cam rings are pivotally supported about the pintle by means of the pivot pin  228  and the opposite ends of the cam rings  220  and  221  are connected to each other via the swiveling lever rod  235  thru the sliding and pivoting joints  233  which is pivotally supported on the pintle at the center cylindrical joint plug  237 , so as one cam ring is moved, the other cam ring is forced to move in the opposite direction. 
     The angular tilt of the cam rings  220  and  221  about the pivot pin  228  between the two extreme positions shown in FIGS. 6 and 7 is controlled by the control system shown in FIGS. 8 and 9. Because one cam ring is forced to follow the other cam ring, only one cam ring need be controlled, in this case, the motor cam ring  221 . Tilt control of the motor cam ring  221  is achieved by means of two separate pistons housed in cylinders the pintle  130 , a control piston  245 , shown in detail in FIGS. 80-82, in a stepped cylinder  247  shown in FIG. 50, and a load piston  250 , shown in detail in FIGS. 83-85 in a cylinder  252 . These pistons act upon the push blocks  225  that are pivotatly connected to opposite sides of the motor cam ring  221  on the attachment bosses  223 . The control piston  245  is continually fed with system pressure through a fluid channel  254  via two check valves  255  and  256  and strokes the motor cam ring  221  toward its maximum displacement. The load piston  250  is fed with a modulated pressure, via a control spool  260 , shown in detail in FIGS. 86 and 87, disposed in an axial bore  263  in the control piston  245 , and strokes the motor cam ring  221  toward zero displacement. The load piston  250  has a larger area than that of the control piston  245 , so that it can overcome both the control forces on the cam ring and the control force from the control piston  245 . 
     System pressure is tapped off from the pintle  130  via the two check valves  255  and  256  and is fed continually to the control cylinder  247  to act against the annular area of the control piston  245 . System pressure is fed from the fluid channel  254  through openings  265  and  266  in the control piston  245  into the bore  263  of the control piston to act on the .control spool  260  that sits inside of the control piston  245 . The control spool  260  modulates the system pressure that is fed to the load piston  250 . The porting to the control spool  260  is such that when the control spool  260  is moved to the right, relative to the control piston  245 , system pressure is fed directly into the load piston chamber. When system pressure acts upon the load piston  250 , the load piston generates enough force to overcome both the control forces on the cam ring and the control force from the control piston and hence strokes the cam ring toward zero displacement. This has the effect of moving the control piston  245  to the right relative to the control spool  260  and thereby closing of the port feeding system pressure to the load piston  250 . When the control spool is moved to the left relative to the control piston, pressure acting upon the load piston is vented. Therefore the force that the load piston generates fall to zero, and as system pressure is continually fed to the control piston, the control piston strokes the motor cam ring toward maximum displacement. This causes the control piston to move to the left relative to the control spool, and thereby closes the port venting the load piston chamber. 
     Looking at FIGS. 8 and 9, a bias spring  267  is compressed between the inside face  268  of an end disc  269  on the control spool  260  and a shoulder in the control cylinder  247 . The spring force of the spring  267  acts to move the control spool toward the left (i.e. stroke the transmission toward final ratio). Control pressure is admitted to the control cylinder through passages  271  and acts upon a spool piston  270  engaged with the control spool  260  to move the control spool to the right when there is sufficient control pressure acting upon the spool piston to overcome the spring force. As the control spool is moved further to the left the spring is further compressed and hence the spring force increases, thereby requiring a higher control pressure to overcome this force. Therefore by modulating the control pressure, the control spool can be accurately positioned relative to the control piston and hence control the position of the motor and pump cam ring, thus controlling the ratio of the transmission. In the design presented, control pressure is supplied and modulated by an external source and fed into the pintle, via a fitting (not shown) in a port  272  in the base of the support plinth  68 , and on into the spool piston chamber. Alternatively, the position of the control spool could be controlled by a miniature stepper motor or servo motor in the pintle, controlled by wires extending through the axis of the stub shaft  140  or by wireless telemetry. 
     Rotation of the pump rotor  125  around the inside of the pump cam ring  235  in contact with the cam race  215  when the pump cam ring  220  is tilted to a non-concentric position with respect to the pump rotor  125 , as shown in FIG. 6, causes the balls  205  to drive the pistons  200  radially into the cylinders and displace fluid which is pumped out radially into fluid channels that run axially between the motor rotor and the pump rotor. As the pump rotor is rotated against the half of the cam race that drives the pump cylinders into the pump rotor cylinders, fluid in those cylinders is pressurized and pumped out through the high pressure passages in the pintle arbor to the motor rotor cylinders, causing the motor pistons to move forcefully under fluid pressure radially outward against the cam race of the motor cam ring and exert a torque on the motor rotor that is transmitted via the motor drive disc and the motor drum to the planet carrier and thence to the output shaft. 
     As the motor pistons pass the top-dead-center position, they are driven back into their cylinders radially and displace spent fluid through, the low pressure fluid passages in the pintle arbor to the pump pistons. The low pressure fluid flow fills the pump cylinders in preparation for their next pressure stroke. 
     The transmission shown in FIGS. 2-9 and described above is designed to be used in applications where an underdrive final ratio is required, and where the input centerline to the output centerline is offset from the transmission centerline, as when power is transmitted by means of a belt, chain, gear train, etc. The described example is a demand fan drive for a large vehicle radiator cooling fan. This design is shown at Neutral. The input hydrostatic unit (HSU) or pump, is at zero displacement and the output HSU, or motor, is at maximum displacement. Both HSUs are simultaneously controlled in this case, although they can be independently controlled. 
     In operation, the input from the engine  58  drives the input pulley  56 . This pulley  56  drives the driven pulley  60  which is the input to the transmission. The driven pulley  60  is connected to the ring gear  100  of the planet set (Rp). The sun gear  110  of the planet set (Sp). drives the pump rotor  125  about the axis  90 . The planet carrier  95  of the planet set is connected to the output shaft  66  (Sg 1 ), and is connected to the motor rotor  130  via the motor drum  145 . 
     Make-up pressure is fed externally from a separate source into the support plinth  53  via a fluid passage  274  and two check valves which connect the make-up supply to the high and low pressure lines of the pump and motor rotors  125  and  130 . This make-up supply also acts as the lubrication supply for the bearings and gears. Control pressure is also fed into the support plinth through a fluid passage  272 , externally from a separate source, and then on into the pintle through the fluid passages  271  where it acts upon the control spool. 
     When the fan drive is at neutral, the output shaft  66  (and hence the motor and planet carrier  95 ) is stationary, the ring gear  100  is rotating at input speed and therefore the sun gear  110  (and hence the pump) is rotating at (Sp/rp) multiplied by the driven pulley speed, in the opposite direction to the input, (in the disclosed embodiment, this is [ 64 / 23 ] 2.78 times input speed). If the pump is at zero displacement, there would be no pumping and therefore no reaction torque could be generated at the pump. Hence the pump rotor  125  would rotate freely and allow no output speed. This true neutral would be desirable for some applications of this transmission, such as the drive transmission for small vehicles. However, in one intended use of the disclosed embodiment, wherein the transmission is used as a vehicle radiator cooling fan drive, it is desirable always to keep the fan spinning at some speed for safety reasons. Therefore, in this application, the pump will not be allowed to go fully to zero displacement, thereby keeping the fan drive at some ratio above neutral. 
     The planet set configuration splits the input torque into two parallel paths: 1) a direct mechanical path fed continually to the output shaft  66  at the ratio of input torque multiplied by (1+(Sp/Rp)) in the same direction, and 2) a mechanical path fed continually to the pump at the ratio of input torque multiplied by (Sp/Rp) in the opposite direction. 
     As the pump cam ring  220  is stroked to give the pump a small displacement and it is rotating at input speed multiplied by (Rp/Sp), pumping takes place. This fluid flow passes directly through the pintle  130  to the motor rotor  135  and drives the motor (in the opposite direction to the pump) to give output speed. Due to the fact that the pump is at a small displacement, a small amount of torque to the pump results in a high pressure and small flow rate. The motor is at a large displacement, so the high pressure and small flow rate of the pump results in a high output torque and low output speed. This high ‘hydraulic’ output torque is added directly to the mechanical output torque as described above. Therefore the total output torque can be expressed as: 
     
       
         Output Torque=Input Torque×[(1+(Sp/Rp))+((Rn/Sn)×motor disp/pump disp)] 
       
     
     It can therefore be seen that there is a total output torque comprising of a fixed mechanical torque plus a variable hydraulic torque. As the motor displacement to pump displacement ratio decreases, the amount of hydraulic torque decreases, and if the motor displacement equals zero then there is no hydraulic torque, just mechanical torque. 
     An advantage of using this kind of hydromechanical transmission in a fan drive application is that when the transmission is at low ratios (where most of the output power is generated hydraulically), the fan speed is slow and hence does not take much power to drive it. This means that very little power will ever get passed through the hydraulic path of the transmission and hence hydraulic losses will be low. 
     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 put shaft and hence ring gear speed, this causes the sun gear speed to decrease, which causes the pump speed to decrease. This has the effect of reducing the total system flow rate, when compared to a conventional hydrostatic fan drive of the same capacity, to approximately ⅓ to ¼ depending on planet set ratios used. This reduces the flow losses and noise levels normally associated with hydrostatic machines. 
     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. This causes the sun gear (Sp) to stop rotating. The pump is now acting as a reaction unit for the sun gear. In this case all the input torque is now transferred thru&#39; the planet set, via planet carrier to the output shaft, and due to the ratio of the ring gear to sun gear, the output speed is decreased and the output torque increased (in the disclosed embodiment, this is by a factor of 2.78:1). 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. 
     With the transmission at final ratio, the fan speed is, at its maximum speed and will hence require full input power to drive it. But as all the power is now transferred through the planet set, and the hydraulics are acting only as a reaction unit to hold the sun gear, the efficiency will be very high (95+%). The only losses being 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 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. 
     Due to the fact that the pump and motor rotate in the opposite directions, the control system is designed to tilt the cam rings such that they are stroked in opposite directions when making adjustments to the transmission ratio, so as to keep the flow passages and the pressure fields inline. When the fan drive is viewed from the front, the pump cam ring is rotated to the left, (from zero displacement to maximum), as the motor cam ring is rotated to the right, (from maximum to minimum displacement), as the CVT is stroked from neutral to final ratio. 
     Make-up fluid is fed externally from a separate source via a port  273  in lower edge of the base of the support plinth and a passage  274  through the stub shaft  140 , then through two check valves in the pintle arbor to the high and low pressure fluid passages in the arbor. Make-up pressure is also fed into the output shaft whereby it is directed to the various gears and bearings for lubrication. 
     Lubrication oil and oil that leaks from the pump and motor rotors will fall to the sump of the transmission, which in this design is the inside diameter of the driven pulley  60 . As the driven pulley is rotating at some ratio of engine speed, the oil that collects there will be centrifuged out to the inside diameter of the driven pulley. This oil is collected and returned to its external source through the pitot tube  190  and a passage  275  to a port  278  in the top of the plinth base  68 . A small elbow may be attached to the outer end of the pitot tube  190  with an angled opening facing into direction of the rotating oil layer on the inside of the driven pulley  60 , and close to inside diameter of the driven pulley  66 , to collect the oil before it reaches the level of the house  88 . As oil rotates at driven pulley speed and comes into contact with the pitot tube, the dynamic head of the oil gets converted into pressure head, and pumps itself down the tube into the outlet port in the fan drive support, or it merely flows down the pitot tube  190  by gravity. This ensures that this sump oil will not be churned by the rotating pulley  60 . 
     This invention provides numerous advantages and benefits as well as a plethora of additional possibilities, including the following: 
     Low power is throughput at low transmission ratio when hydraulics do most of the work. This reduces the hydraulic losses normally incurred in hydrostatic devices. 
     When clipping fan speed, no power is wasted. 
     With the gear train, a range of final ratio speed ranges is possible, as contrasted with 1:1 with a clutch. 
     No particulates are generated in the lube oil due to clutch wear because the clutch is eliminated. 
     Improved life and efficiency. 
     Lower fan noise due to the ability to keep the fan at an optimum speed at all operating conditions. 
     Due to the fact that the hydraulics do little work, low cost hydrostatic units can be used. 
     The transmission uses the same inputs (i.e lube and control supply) as current clutch type fan drives, making integration and retrofitting easy. 
     Balls in spherical pockets enhance piston sealing in the cylinders. 
     The piston/ball interface is hydrostatically balanced by means of a pressure fed balance groove to reduce the contact loading between the ball and the piston ensuring ball rotation in the pockets and reducing friction between the ball and the piston. 
     By stroking the cam rings in the opposite directions and rotating them in opposite directions the rotating high pressure fields are kept in line with each other between the pump and motor. This places the control shaft in shear and the pintle in tension, not in bending, thereby reducing the stress and deflections in those parts. 
     All of the hydraulic forces and control forces are contained within the pintle and hydrostatic unit assembly, and are not transferred to the support or housing, thereby reducing noise transmitted to the outside. Only reaction torque is transmitted to the support. 
     Obviously, numerous modifications and variations of the preferred embodiment described above are possible and will become apparent to those skilled in the art in light of this specification. For example, vane-type hydrostatic units could be used instead of the radial piston type hydrostatic units in situations where greater power density is required and the disadvantages of shorter life and greater leakage rate of vant-type hydrostatic units would be acceptable. Moreover, many functions and advantages are described for the preferred embodiment, but in many 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 features, process steps, benefits, 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, in all their combinations, are to be considered within the spirit and scope of the invention as defined in the following claims, wherein.