Abstract:
A hydraulic hybrid powertrain system includes a power plant generating a high pressure fluid at an output, at least one drive motor responsive to the high pressure fluid for generating rotary motion at an output, and a mode selection device connected to the power plant output and the drive motor for selecting a mode of operation from a plurality of drive motor modes of operation including at least two of a drive mode, a neutral mode, a reverse mode and a park mode. The system includes a control device connected to the power plant and the drive motor for controlling operation of the drive motor in the plurality of modes of operation, a selectively actuated brake device for interrupting high pressure fluid flow to the drive motor, and a check valve bridge circuit for connecting the drive motor to a low pressure fluid source when the brake device is actuated.

Description:
CROSS-REFERENCE TO RELATED APPLICATION 
   This application is a continuation-in-part of U.S. application Ser. No. 11/101,837 filed on Apr. 8, 2005 now U.S. Pat. No. 7,179,070. 
   This application claims priority from the provisional application Ser. No. 60/655,221 filed on Feb. 22, 2005. 

   BACKGROUND OF THE INVENTION 
   The present invention relates generally to vehicle powertrain systems and, in particular, to a hydraulic hybrid powertrain system. 
   So-called hybrid powertrains, such as for automotive vehicles, generally refer to a powertrain wherein an internal combustion engine is utilized in combination with an auxiliary motor, such as electric motor or a hydraulic motor, to drive the vehicle. Hybrid powertrain systems known as parallel hybrids include a typical mechanical drivetrain (coupled to the internal combustion engine) along with the auxiliary drivetrain (coupled to the auxiliary motor). These systems are disadvantageously high in weight because of the necessary duplication of parts. Hybrid drive systems known as series hybrids do away with the mechanical powertrain and drive the vehicle solely by a hydraulic motor or motors while utilizing an engine to provide the necessary hydraulic pressure for the hydraulic motor. These systems are more attractive because of the potential reduction in weight and resultant efficiency gains. While the attractiveness of such a hydraulic hybrid powertrain has been recognized, there remain many efficiency issues regarding the operation and control of the engine with respect to the hydraulic drive motor. 
   It is desirable, therefore, to provide a hydraulic hybrid powertrain system that provides increased efficiency of the entire hydraulic hybrid powertrain system. 
   SUMMARY OF THE INVENTION 
   The present invention concerns a hydraulic hybrid powertrain system that includes a power plant generating a high pressure fluid at an output, at least one drive motor responsive to the high pressure fluid for generating rotary motion at an output, and a mode selection means connected to the power plant output and the at least one drive motor for selecting a mode of operation from a plurality of modes of operation of the at least one drive motor including at least two of a drive mode, a neutral mode, a reverse mode and a park mode. The system also includes a control means connected to the power plant and the at least one drive motor for controlling operation of the at least one drive motor in the plurality of modes of operation, a selectively actuated brake means for interrupting a flow of the high pressure fluid to the at least one drive motor, and a check valve bridge circuit for connecting the at least one drive motor to a low pressure fluid source when the brake means is actuated. 

   
     DESCRIPTION OF THE DRAWINGS 
     The above, as well as other advantages of the present invention, will become readily apparent to those skilled in the art from the following detailed description of a preferred embodiment when considered in the light of the accompanying drawings in which: 
       FIG. 1   a  is a schematic view of a hydraulic hybrid powertrain system in accordance with the present invention with a mode select valve in a “Drive” position; 
       FIG. 1   b  is a view of the hydraulic hybrid powertrain system of  FIG. 1   a  with the mode select valve in a “Neutral” position; 
       FIG. 1   c  is a view of the hydraulic hybrid powertrain system of  FIG. 1   a  with the mode select valve in a “Reverse” position; 
       FIG. 1   d  is a view of the hydraulic hybrid powertrain system of  FIG. 1   a  with the mode select valve in a “Park” position; 
       FIG. 1   e  is a view of the hydraulic hybrid powertrain system of  FIG. 1   a  with a brake override device in an override position; 
       FIG. 2  is a schematic view in an enlarged scale of the drive motors and displacement control devices shown in  FIGS. 1   a - 1   d;    
       FIG. 3  is a schematic view in an enlarged scale of the brake override device and check valve bridge circuit shown in  FIGS. 1   a - 1   d;    
       FIG. 4  is an exploded perspective view of an internal gear pump/motor in accordance with the present invention; and 
       FIG. 5  is a partial exploded perspective view of an external gear pump/motor in accordance with the present invention. 
   

   DESCRIPTION OF THE PREFERRED EMBODIMENT 
   The following patent applications are incorporated herein by reference: U.S. patent application Ser. No. 60/655,221 filed on Feb. 22, 2005; U.S. patent application Ser. No. 11/101,837 filed on Apr. 8, 2005. 
   Referring now to  FIG. 1   a , a hydraulic hybrid powertrain system in accordance with the present invention is indicated generally at  10 . The powertrain system  10  may be utilized in a variety of installations, such as, but not limited to, an automotive vehicle, a boat, a submarine, a helicopter, or the like as will be appreciated by those skilled in the art, but for clarity will be referred to as if installed in an automotive vehicle in the following description of the present invention. The powertrain system  10  includes a power plant section  11 , a mode selector module  43 , a control section  59 , and a power delivery section  76 . 
   The power plant section  11  of the powertrain system  10  includes an engine  12  in communication with a fuel source  14 . The engine  12  may be a conventional internal combustion engine, a turbine engine, an electric motor powered by a battery, a fuel cell, or the like. The engine  12  selectively provides torque to a preferably variable displacement hydraulic pump/motor  16 , which is supplied with a low pressure source  18  of hydraulic fluid on an inlet side thereof and a high pressure conduit  20  on an outlet side thereof. The hydraulic fluid may be a liquid, such as but not limited to water, hydraulic fluid, transmission fluid or the like, or any compressible gas while remaining within the scope of the present invention. The pump/motor  16  is described as such because, depending on the mode of the system  10 , the device functions alternately as a pump or a motor, discussed in more detail below. 
   The power plant section  11  of the system  10  includes a plurality of accessory drives including, but not limited to, a motor generator  22 , an air conditioning compressor  24 , and a heat pump  26 . The motor generator  22  is connected to a power maintenance module  28 , which is in turn connected to a battery pack  30 . The heat pump  26  is in communication with a heater core  32  and both the heat pump  26  and the heater core  32  are in fluid communication with a cooling water source  34  for the engine  12 . The air conditioning compressor  24  is in communication with a heat exchanger  36 . The accessory drives  22 ,  24 , and  26  are preferably run by respective electric or hydraulic motors. Alternatively, the accessory drives  22 ,  24 , and  26  are selectively mechanically clutched to the engine  12 . An accumulator  38  is in fluid communication with the high pressure conduit  20  on the outlet of the pump/motor  16 . The accumulator  38  serves as a reservoir for high pressure hydraulic fluid and maintains high pressure in the system  10 , such as by being charged with a high pressure gas or the like (not shown), as will be appreciated by those skilled in the art. 
   A throttle control module  40  receives an input signal from the air conditioning compressor  24  via a signal on a line  24   a , the power maintenance module  28  via a signal on a line  28   a , and the accumulator  38  via a signal on a line  38   a . Based on the input signals on the lines  24   a ,  28   a , and  38   a , the throttle control module  40  provides an output signal on a line  42  to control either or both of the engine  12  and the pump/motor  16 , discussed in more detail below. The signals on the lines  24   a ,  28   a ,  38   a , and  42  may be electronic signals or mechanical feedback between the various components and the throttle control module  40 . The throttle control module  40  can be any suitable mechanical or electrical device operable to control the operation of the engine  12  and the pump/motor  16  based on one or more inputs. 
   The mode selector module  43  includes a mode select valve  44  that is in fluid communication with the high pressure conduit  20  by a high pressure inlet conduit  46 . The mode select valve  44  is preferably connected to a transmission-like shift lever (not shown) or the like for selectively moving the valve  44  into a one of a “D” or drive position (best seen in  FIG. 1   a ), a “N” or neutral position (best seen in  FIG. 1   b ), a “R” or reverse position (best seen in  FIG. 1   c ), and a “P” or park position (best seen in  FIG. 1   d ). The mode select valve  44  includes a low pressure inlet conduit  48  connected thereto adjacent the high pressure inlet conduit  46 . The mode select valve  44  also includes a high pressure outlet conduit  50  and a low pressure outlet conduit  52  connected thereto and on an opposing side of the mode select valve  44 . Each position P, R, N, D of the mode select valve  44  selectively aligns the internal portion of the position with the conduits  46 ,  48 ,  50 , and  52  and controls the direction of hydraulic fluid flow in the system  10 , discussed in more detail below. While described as “inlet” and “outlet” above during operation each of the conduits  46 ,  48 ,  50 , and  52  may function as an inlet or an outlet depending on the operating condition of the system  10 , discussed in more detail below. 
   The conduits  50  and  52 , in turn, are connected to a brake override device  54 . The brake override device  54  also includes a high pressure outlet conduit  56  and a low pressure outlet conduit  58  connected thereto on an opposing side of the brake override device  54 . The brake override device  54  has a first or normal position  54   a  and a second or override position  54   b , discussed in more detail below. 
   The control section  59  includes a displacement control valve  60  that is in fluid communication with the high pressure conduit  20  by a high pressure inlet conduit  62 . The displacement control valve  60  includes a low pressure inlet conduit  64  connected thereto adjacent the high pressure inlet conduit  62 . The displacement control valve  60  also includes a high pressure outlet conduit  66  and a low pressure outlet conduit  68  connected thereto on an opposing side of the displacement control valve  60 . The displacement control valve  60  is a floating positional valve and includes an accelerator  70  and a brake  72  connected thereto for directing flow from the displacement control valve  60  to a plurality of cylinders  74   a ,  74   b ,  74   c , and  74   d . The accelerator  70  and brake  72  are preferably mechanically connected to a respective accelerator pedal and a brake pedal (not shown). The brake  72  is connected to the brake override device  54  via a connector  73 . The displacement control valve  60  has a first or acceleration position  60   a , a second or hold position  60   b , and a third or deceleration position  66   c . Each position  60   a ,  60   b , and  60   c  of the displacement control valve  60  selectively aligns the internal portion of each position  60   a ,  60   b , and  60   c  with the conduits  62 ,  64 ,  66 , and  68  and controls the direction of hydraulic fluid flow to the cylinders  74   a ,  74   b ,  74   c , and  74   d , best seen in  FIG. 2 . 
   Each of the cylinders  74   a ,  74   b ,  74   c , and  74   d  is mechanically connected via a connector  75   a ,  75   b ,  75   c , and  75   d , to a respective and drive or traction motor  76   a ,  76   b ,  76   c , and  76   d  (in the power delivery section  76 ), on each of the vehicle wheels. The motors  76   a - 76   d  are preferably variable displacement motors. The position of the connectors  75   a - 75   d  determines the displacement of the motors  76   a - 76   d , as will be appreciated by those skilled in the art such as by a connection to a swash plate or the like. The high pressure outlet conduit  66  is in fluid communication with one side of a piston (not shown) in each of the cylinders  74   a - 74   d  and the low pressure outlet conduit  68  is in fluid communication with an opposite side of the piston in the cylinders  74   a - 74   d . While the system  10  is illustrated with a plurality of traction motors  76   a ,  76   b ,  76   c , and  76   d , those skilled in the art will appreciate that as few as one motor may be utilized while remaining within the scope of the present invention. For example, in a single motor installation in an automotive vehicle, the output of the single motor is connected to a differential gear which is in turn mechanically connected to a pair of drive wheels. Each of the traction motors  76   a ,  76   b ,  76   c , and  76   d  have an upper port  77   a ,  77   b ,  77   c , and  77   d  and a lower port  78   a ,  78   b ,  78   c , and  78   d . The direction of the fluid flow through the upper ports  77   a - 77   d  and the lower ports  78   a - 78   d  determines the direction of the motors  76   a - 76   d . A feedback connector  80  extends between the displacement control valve  60  and the pistons of the cylinders  74   a - 74   d.    
   A check valve bridge circuit  82  includes a plurality of check valves  84 ,  86 ,  88 , and  90  and is arranged in a manner similar to a full-wave bridge rectifier, best seen in  FIG. 3 . A conduit  92  is in fluid communication with an inlet of the check valve  84  and an outlet of the check valve  86 . The conduit  92  is also in fluid communication with the high pressure outlet conduit  56 . A conduit  94  is in fluid communication with an inlet of the check valve  86  and an inlet of the check valve  88 . The conduit  94  is also in fluid communication with the low pressure source of hydraulic fluid  18 . A conduit  96  is in fluid communication with an outlet of the check valve  88  and an inlet of the check valve  90 . The conduit  96  is also in fluid communication with the low pressure outlet conduit  56 . A conduit  98  is in fluid communication with an outlet of the check valve  84  and an outlet of the check valve  90 . The conduit  98  is also in fluid communication with the high pressure conduit  20 . 
   The pump/motor  16  and the motors  76   a - 76   d  are preferably variable displacement pump/motors such as that shown in commonly assigned and co-pending patent application Ser. No. 11/101,837 filed on Apr. 8, 2005, the disclosure of which is hereby incorporated by reference and shown in  FIGS. 4 and 5 . Alternatively, the pump/motor  16  and the motors  76   a - 76   d  are vane-type or piston-type variable displacement pump/motors or are fixed displacement pump/motors. 
   Referring now to  FIG. 4 , an internal gear apparatus in accordance with the present invention is indicated generally at  100 . The apparatus  100  may be configured to operate as a motor or as a pump as will be appreciated by those skilled in the art, but will be referred to as a motor in the following description of the present invention. The internal gear motor  100  includes a hollow housing  102  having a base portion  104  and an end cap  106 . The base portion  104  defines a recess or cavity  108  therein that is sized to receive a first mandrel  110  and a first piston member  112 . The end cap  106  includes at least two ports  107  (only one is shown) that each extend between an internal and an external surface thereof, preferably on opposite sides of the end cap  106 . One of the ports  107  is connected to a high pressure segment of a fluid system such as the high pressure conduit  20  of  FIGS. 1   a - 1   e , and another of the ports  107  is connected to a return line or fluid source such as the fluid source  18  of  FIGS. 1   a - 1   e.    
   The first mandrel  110  defines an aperture  114  extending through a base portion  111  thereof and includes a first outer flange  116  and a plurality of spaced apart second outer flanges  118  extending upwardly from an upper surface  113  of the base portion  111 . An inner flange  120  extends upwardly from the base portion  111  of the first mandrel  110  and is located adjacent the aperture  114 . The first outer flange  116  is located adjacent the aperture  114 . The second outer flanges  118  are spaced apart from both the aperture  114  and the inner flange  120 . A first seal bushing  122  is sized to rotatably fit in the aperture  114  and is preferably substantially equal in height to the base portion  111  of the first mandrel  110  such that when the bushing  122  is placed in the aperture  114 , an upper surface of the bushing  122  is substantially flush with the upper surface  113  of the base portion  111 . 
   An external gear  124  that is substantially circular in cross section is adapted to be placed atop the upper surface  113  of the base portion  111  wherein a curved outer surface of the gear  124  is adjacent the respective curved inner surfaces of the outer flanges  116  and  118 . The external gear  124  includes a plurality of teeth  126  formed on an inner surface thereof. When placed on the upper surface  113 , the gear  124  is fixed axially between the outer flanges  118  and the inner flange  120 . 
   An internal gear  128  that is substantially circular in cross section includes a plurality of teeth  130  formed on an outer surface thereof and defines an aperture  132  extending therethrough. The teeth  130  are operable to mesh with the teeth  126  formed on the inner surface of the external gear  124 . A lower surface of the gear  128  extends into and rotates with the bushing  122 , wherein the teeth  130  cooperate with corresponding teeth on the bushing  122  when the motor  100  is assembled and operated, as discussed in more detail below. The respective outer surfaces of the teeth  130  of the internal gear  128  are adjacent the inner surface of the inner flange  120 . The aperture  132  is adapted to receive a free end of a drive or output shaft  134  when the motor  100  is assembled. The internal gear  128  is axially moveable along the shaft  134 . The drive shaft  134  is rotatably supported in the end cap  106  by a bearing  135 , such as a ball bearing, a roller bearing or the like. The free end of the drive shaft  134  extends a predetermined distance beyond the upper surface of the end cap  106  and acts as an output shaft for the motor  100 . 
   A second piston member  136  defines an aperture  138  on an interior portion thereof and is adapted to be mounted on respective upper surfaces of the outer flanges  116  and  118  of the first mandrel  110 . The second piston  136  and the first piston  112 , therefore, are mounted on the upper surface and the lower surface, respectively of the lower mandrel  110 . 
   A second mandrel  140  is adapted to be disposed in the aperture  138  of the second piston member  136  and defines an aperture  142  on an interior portion thereof for receiving the drive shaft  134 . The second mandrel  140  includes a downwardly extending flange  144  that cooperates with the upwardly extending inner flange  120  of the first mandrel  110  when the motor  100  is assembled. The upper mandrel  140  includes a pair of bores  146  extending therethrough for fluid communication with the gears  122  and  124  during operation of the motor  100 . 
   A second seal bushing  148  includes a plurality of teeth  150  formed on an exterior surface thereof and defines an aperture  152  extending therethrough. The second seal bushing  148  is adapted to receive the upper mandrel  140  in the aperture  152  and be received in the external gear  124  and rotates therewith, wherein the teeth  126  cooperate with the teeth  150  on the bushing  148  when the motor  100  is assembled and operated, as discussed in more detail below. 
   When the motor  100  is assembled, the first mandrel  110  and the first piston  112  are placed in the base portion  104  of the housing  102 , the first seal bushing  122  is placed in the mandrel  110 , and the external gear  124  is placed on the mandrel  110 . The internal gear  132  and the second mandrel  138  are mounted on the drive shaft  134  and assembled such that the respective teeth  126  and  130  of the gears  132  and  124  rotatably mesh and the internal gear  132  engages with the first seal bushing  122 . The second piston  136  is attached to the upper surface of the mandrel  110 , and the second seal bushing  148  is placed on the second mandrel  138  and engages with the external gear  124 . The downwardly extending flange  144  cooperates with the upwardly extending inner flange  120  to divide the interior of the external gear into an inlet chamber and discharge chamber of the motor  100  and the upper end cap  106  is attached to the base portion  104  to enclose the housing  102 . The flanges  120  and  144  extend radially between the teeth  126  and the teeth  130  to form the inlet chamber on one side of the flanges and the discharge chamber on the other side of the flanges. 
   In operation, the shaft  134  is connected to a load (not shown), such as a wheel of a vehicle or the like. Pressured fluid is introduced from the fluid system such as from the high pressure conduit  20  of  FIGS. 1   a - 1   e , through one of the ports  107 , is routed to the inlet chamber side of the gears  124  and  128  through the bores  146 , acts against the meshing teeth  126  and  130  to rotate the gears and the shaft, flows between the teeth to the discharge chamber and is discharged through the other the bores  146  to the other of the ports  107 . The first seal bushing  122  provides a rotating seal between the internal gear  128  and the first mandrel  110  and the second seal bushing  148  provides a rotating seal between the external gear  124  and the second mandrel  140  to ensure the integrity of the inlet and discharge chambers. The motor  100  in accordance with the present invention requires only the seals  122  and  148  to maintain a fluid seal and allow for efficient operation of the motor  100 . 
   The normal or default spatial relationship between the teeth  126  and  130  of the gears  124  and  128  is such that the teeth  126  and  130  engage substantially all of the axial area of the teeth. In such a relationship, the motor  100  produces its maximum volume flow or maximum output. The motor  100  in accordance with the present invention may advantageously vary from its maximum displacement because the internal gear  128  is axially movable along the shaft  134 . When the internal gear  128  moves towards the first mandrel  110 , less of the axial area of the teeth  126  and  130  engage, which reduces the volume flow or displacement of the motor  100 . 
   When the unit  100  is configured as a motor, an external source of pressure, such as hydraulic fluid from an external hydraulic pump, compressed air from an air compressor or the like, provides a volume flow to the ports  107  to spin the gears  124  and  128  and produce an output torque on the shaft  134 . As the pressure is varied, the internal gear  128  will move along the axis of the shaft  134  in order to vary the output horsepower of the motor  100 . The motor  100  may be advantageously utilized to control output rpm under widely changing output loads including, but not limited to automotive vehicles, turrets, large machinery, earth movers, large well drills, ships, farm equipment, or the like. 
   When the unit  100  is configured as a pump and a prime mover, such as the engine  12  of  FIGS. 1   a - 1   e , rotates the shaft  134  at a lower speed or with a lower torque, the pump  100  will react to the reduced input speed or input torque by varying its output based on the internal pressures in the pump housing  102 . In this condition, the output port  107  will create a higher back pressure in the discharge chamber, and the internal gear  128  will move along the axis of the shaft  134  to a point along the axis where the gear  128  is at or near equilibrium to continue operation. The pump  100 , therefore, can vary from a maximum output or displacement where the internal gear  128  is substantially adjacent the upper mandrel  140  to a minimum displacement where the internal gear  128  is substantially adjacent the lower mandrel  110 . 
   Referring now to  FIG. 5 , an external gear apparatus in accordance with the present invention is indicated generally at  200 . The apparatus  200  may be configured to operate as a pump or a motor as will be appreciated by those skilled in the art, but will be referred to as a pump in order to simplify the description of the present invention. The external gear pump  200  includes a hollow housing  202  having a first end cap  204  and a second end cap  206  connected by a body portion  208 . Preferably, the first end cap  204  and the second end cap  206  are attached to the body portion  208  by a plurality of fasteners  210 , such as high strength bolts or the like. The body portion  208  defines a recess  212  therein. 
   A first gear  214  having a plurality of teeth  216  formed on an external surface thereof and a second gear  218  having a plurality of teeth  220  formed on an external surface thereof are adapted to be disposed in the recess  212  of the housing  202 . The teeth  216  and  220  of the respective gears  214  and  218  are operable to rotatably mesh in the recess or pump cavity  212  during operation of the pump  200 . The first gear  214  has a shaft  222  extending therefrom and the second gear  216  has a stepped shaft  224  extending therefrom. The first gear  214  is fixed on the shaft  222  and the second gear  218  is axially moveable along the shaft  224 . The shafts  222  and  224  extend in opposite axial directions and the shaft  224  is greater in length than the shaft  222 . A first seal sleeve  226  having internal teeth receives the first gear  214  and a second seal sleeve  228  having internal teeth receives an end of the second gear  218 . 
   A plate fitting  230  includes a flange  232  extending downwardly therefrom and is attached to a first thrust plate  234  on a planar upper surface thereof. Preferably, the thrust plate  234  is attached to the fitting  230  by a plurality of fasteners  236 , such as high strength bolts or the like. A free end of the shaft  222  extends through an opening formed in the fitting  230  and the thrust plate  234 . The free end of the shaft  222  is rotatably secured in the fitting  230  and the thrust plate  234  by a pair of nuts  238  and is rotatably supported by a bearing  240 , such as a ball bearing, a roller bearing or the like. The second seal sleeve  228  is operable to be received in a recess in the fitting  230  adjacent the flange  232 . When the shaft  222  is mounted in the fitting  230  and the thrust plate  234 , the gear  214  is fixed axially with respect to the housing  202 . 
   A second thrust plate  242  is attached to an upper surface  205  of the first end cap  204  by a plurality of fasteners  244 , such as high strength bolts or the like. The plate  242  includes an aperture for receiving a free end of the shaft  224  and a larger aperture for receiving and locating the first seal sleeve  226  adjacent the upper surface of the first end cap  204 . The free end of the shaft  224  extends through the aperture in the plate  242 , threadably engages a pair of nuts  246  at the step and is rotatably supported by a bearing  248 , such as a ball bearing, a roller bearing or the like. The bearing  248  is preferably disposed in a cavity  250  formed in the upper surface  205  of the first end cap  204  while the nuts  246  attach the shaft  224  to the end cap on a lower surface opposite the upper surface  205 . The free end of the shaft  224  extends a predetermined distance beyond the lower surface of the end cap  204  and acts as a drive shaft or output shaft for the pump  200 . 
   The body portion  208  defines a first port  252  and a second port  254  that each extend between an internal and an external surface thereof. One of the ports  252  and  254  is connected to a low pressure segment of a fluid system such as the hydraulic fluid source  18  of  FIGS. 1   a - 1   e  or the like, and another of the ports  252  and  254  is connected to a high pressure or pressurized segment of a fluid system such as the high pressure conduit  20  of  FIGS. 1   a - 1   e.    
   In operation, the shaft  224  is connected to a prime mover, such as the engine  12  of  FIGS. 1   a - 1   e  or the like. When the prime mover rotates the shaft  224 , the gear  218  rotates and causes the gear  214  to rotate. Fluid is introduced from the fluid system through one of the ports  252  or  254 , is trapped between the meshing teeth  216  and  220  as is well known in the art and is discharged through the other of the ports  252  or  254 . Suitable passages are formed in the housing  202  to ensure that the fluid is routed correctly during operation of the pump  200 . The first seal sleeve  226  provides a rotating seal between the first gear  214  and the upper surface  205  and the second seal sleeve  228  provides a rotating seal between the second gear  218  and the fitting  230  to ensure the integrity of the pump cavity  212 . The pump  200  in accordance with the present invention requires only the seal sleeves  226  and  228  to maintain a seal and allow for efficient operation of the pump  200 . 
   The normal or default spatial relationship between the teeth  216  and  220  of the gears  214  and  218  is such that the teeth  216  and  220  engage substantially all of the axial area of the teeth. In such a relationship, the pump  200  produces its maximum volume flow or maximum displacement. The pump  200  in accordance with the present invention may advantageously vary from its maximum displacement because the second gear  218  is axially movable along the shaft  224 . When the second gear  218  moves towards the lower thrust plate  242 , less of the axial area of the teeth  216  and  220  engage, which reduces the volume flow or displacement of the pump  200 . Typically, this will occur when the prime mover rotates the shaft  224  at a lower speed or with a lower torque and the pump  200  will react to the reduced input speed or input torque by varying its output based on the internal pressures in the pump housing  202 . In this condition, the output port  252  or  254  will create a higher back pressure in the recess  212 , and the second gear  218  will move along the axis of the shaft  224  to a point along the axis where the gear  218  is at or near equilibrium to continue operation. The pump  200 , therefore, can vary from a maximum output or displacement where the gear  218  is substantially adjacent the fitting  230  to a minimum displacement where the gear  218  is substantially adjacent the lower thrust plate  242 . 
   When the apparatus  200  is configured as a motor, an external source of pressure, such as hydraulic fluid from an external hydraulic pump, compressed air from an air compressor or the like, provides a volume flow to the ports  252  and  254  to spin the gears  214  and  218  and produce an output torque on the shaft  224 . As the pressure is varied, the second gear  218  will move along the axis of the shaft  224  in order to vary the output horsepower of the motor  200 . The motor  200  may be advantageously utilized to control output rpm under widely changing output loads including, but not limited to automotive vehicles, turrets, large machinery, earth movers, large well drills, ships, farm equipment, or the like. 
   In operation of the system  10 , the engine  12  is started and supplies torque to the pump/motor  16 , which in turn supplies pressurized hydraulic fluid to the high pressure conduit  20 . The accumulator  38  ensures that the hydraulic pressure within the conduit  20  remains relatively stable and provides energy storage in a manner well known to those skilled in the art. The pressure in the conduit  20  is transmitted to the conduits  46 ,  62 , and  98 . 
   Referring to  FIG. 1   a , when the mode select valve  44  is in the D or drive position and the brake override device  54  is in the  54   a  position, hydraulic fluid will flow through the conduit  46 , through the mode select valve  44  and out the conduit  50  in the direction shown by the arrow in the D position, through the brake override device  54  and out the conduit  56  in the direction shown by the arrow in the  54   a  position, and to the respective upper ports  77   a - 77   d  of the motors  76   a - 76   d , through the motors  76   a - 76   d  and to the respective lower ports  78   a - 78   d , dropping in pressure and providing an output torque in a forward direction for each of the motors  76   a - 76   d  in a manner known to those skilled in the art. The lower pressure hydraulic fluid in the lower ports  78   a - 78   d  travels through the conduit  58 , through the brake override device and out the conduit  52  in the direction shown by the arrow in the  54   a  position, and through the mode select valve  44  and out the conduit  48  in the direction shown by the arrow in the D position to the hydraulic fluid source  18 . 
   Referring to  FIG. 1   b , when the mode select valve  44  is in the N or neutral position, and the brake override device  54  is in the  54   a  position, hydraulic fluid will flow through the conduit  46  but will be prevented from flowing through the mode select valve  44  by the cap adjacent the conduit  46  in the N position. The outlet conduits  50  and  52  are in fluid communication with the lower pressure hydraulic fluid in the conduit  48  and, therefore, there is no fluid flow through the brake override device  54  or to the motors  76   a - 76   d , as the pressure in the conduits  50  and  56  will balance with the pressure in the conduits  52  and  58 . When the in N position, oil from the reservoir  18  is available to flow through to the motors  76   a - 76   d  should any of the motors  76   a - 76   d  require oil flow. 
   Referring to  FIG. 1   c , when the mode select valve  44  is in the R or reverse position, and the brake override device  54  is in the  54   a  position, hydraulic fluid will flow through the conduit  46 , through the mode select valve  44  and out the conduit  52  in the direction shown by the arrow in the R position, through the brake override device  54  and out the conduit  58  in the direction shown by the arrow in the  54   a  position, and to the respective lower ports  78   a - 78   d  of the motors  76   a - 76   d , through the motors  76   a - 76   d  and to the respective upper ports  77   a - 77   d , dropping in pressure and providing an output torque in a reverse direction for each of the motors  76   a - 76   d  in a manner known to those skilled in the art. The lower pressure hydraulic fluid in the lower ports  77   a - 77   d  travels through the conduit  56 , through the brake override device and out the conduit  50  in the direction shown by the arrow in the  54   a  position, and through the mode select valve  44  and out the conduit  48  in the direction shown by the arrow in the D position to the hydraulic fluid source  18 . 
   Referring to  FIG. 1   d , when the mode select valve  44  is in the P or park position, and the brake override device  54  is in the  54   a  position, hydraulic fluid will not flow through any of the conduits  46 ,  48 ,  50 , and  52  as the caps adjacent each of the conduits  46 ,  48 ,  50 , and  52  in the P position prevent any flow through to the motors  76   a - 76   d.    
   As outlined above, in the first position  54   a , the brake override device  54  allows hydraulic fluid to flow (depending on the position of the mode select valve  44 ) between the conduits  50  and  56 , and between the conduits  52  and  58 . In the second position  54   b , however, best seen in  FIG. 1   e , hydraulic fluid will not flow through any of the conduits  50 ,  52 ,  56 , and  58  as the caps adjacent each of the conduits  50 ,  52 ,  56 , and  58  in the second position  54   b  prevent any flow through the brake override device  54 . The brake override device  54  is moved from its normal first position  54   a  to the second position  54   b  by actuation of the brake  72  and the transmission of a signal along the connector  73  and prevents hydraulic fluid flow from the displacement control valve  44  to the motors  76   a - 76   d.    
   In operation, if the brake  72  is engaged when the mode select valve  44  is in the D or drive position, and the override device  54  is moved to the second position  54   b , the only source of hydraulic fluid for the motors  76   a - 76   d  is through the check valve bridge circuit  82  and, therefore, all fluid flow is routed through the check valve bridge circuit  82 . During braking, the motors  76   a - 76   d  will begin to function as pumps, advantageously recapturing energy from the rotation of the vehicle wheels during braking. When braking in the D position, hydraulic fluid will flow from the hydraulic fluid source  18 , through the conduit  94 , through the check valve  86 , through the conduit  92 , to the upper ports  77 - 77   d  and to the motors  76   a - 76   d , where the hydraulic fluid pressure is raised. High pressure hydraulic fluid will then flow from the motors  76   a - 76   d , through the lower ports  78   a - 78   d , through the conduit  96 , and, if the pressure in the conduit  96  is greater than the conduit  98 , through the check valve  90  and into the conduit  98 , where the high pressure hydraulic fluid flows to the conduit  20  and recharges the accumulator  38 . 
   When braking while the mode select valve  44  is in the R position, hydraulic fluid will flow from the hydraulic fluid source  18 , through the conduit  94 , through the check valve  88 , through the conduit  96 , to the lower ports  78   a - 78   d  and to the motors  76   a - 76   d , where the hydraulic fluid pressure is raised. High pressure hydraulic fluid will then flow from the motors  76   a - 76   d , through the upper ports  77   a - 77   d , through the conduit  92 , and, if the pressure in the conduit  92  is greater than the conduit  98 , through the check valve  84  and into the conduit  98 , where the high pressure hydraulic fluid flows to the conduit  20  and recharges the accumulator  38 . 
   The check valve bridge circuit  82  functions to prevent flow of hydraulic fluid to the motors  76   a - 76   d  in a reverse direction once the vehicle has come to a complete stop. When braking and when the mode select valve  44  is in the D position, the brake override device  54  moves to the position  54   b  and prevents flow from the mode select valve  44  to the motors  76   a - 76   d . Flow from the high pressure conduit  20  will attempt to reach the motors  76   a - 76   d  via the conduit  98  but is prevented from flowing to the motors via the check valves  84  and  90 . The check valve bridge circuit  82  will allow flow to the conduit  98  only from the conduit  92  through the check valve  84  or from the conduit  96  via the check valve  90 , which will only occur when the pressure in the conduits  56  and  92  or the conduits  58  and  96  are greater than the pressure in the conduit  98 . If the pressure in the conduit  92  is less than the pressure in the conduit  98  and the conduit  94 , the check valve  86  will open but since the conduit  94  is at a low pressure, no flow can occur from the reservoir  18  to the conduit  92 . Similarly if the pressure in the conduit  96  is less than the pressure in the conduit  98  and the conduit  94 , the check valve  88  will open but since the conduit  94  is at a low pressure, no flow can occur from the reservoir  18  to the conduit  96 , and advantageously preventing high pressure hydraulic fluid from causing the motors  76   a - 76   d  to engage in a reverse direction after the vehicle has come to a complete stop. 
   In operation, the flow of the hydraulic fluid through the system  10  is controlled by the operator via the accelerator  70  and the brake  72  connected to the displacement control valve  60 . The connector  80  and the connections  75   a - 75   d  are connected together via suitable linkage or the like, which allows the motors  76   a - 76   d  to provide feedback to the displacement control valve  60  via the connections  75   a - 75   d  in a similar manner as the connector  80  provides control to the motors  76   a - 76   d  through the connections  75   a - 75   d.    
   For example, if a user (not shown) of the vehicle presses the accelerator  70 , this causes the feedback connector  80  to move in an acceleration direction and causes the displacement control valve  60  to move toward the position  60   a . High pressure fluid from the conduit  62  will flow through the ports on the displacement control valve  60 , increasing the pressure in the conduit  66  and flowing to the cylinders  74   a - 74   d . Since the pressure in the conduit  66  will be greater than the pressure in the conduit  68 , the connectors  75   a - 75   d  will be moved in an acceleration direction, increasing the displacement and, therefore, the output torque of the motors  76   a - 76   d.    
   Once a desired output torque of the motors  76   a - 76   d  has been reached, the motors  76   a - 76   d  will throttle back, moving the connectors  75   a - 75   d  in a deceleration direction, decreasing the pressure in the conduit  66  and increasing the pressure in the conduit  68 . This movement is translated back to the displacement control valve  60  by the feedback connector  80 , which moves the displacement control valve towards the position  60   b . In the position  60   b , there is no flow through the displacement control valve  60  and thus the connectors  75   a - 75   b  remain stationary and the displacement and, therefore, the output torque of the motors  76   a - 76   d  remains constant. 
   If the user removes his or her foot from the accelerator  70 , this causes the feedback connector  80  to move in a deceleration direction and causes the displacement control valve  60  to move toward the position  60   c . High pressure fluid from the conduit  62  will flow through the ports on the displacement control valve  60 , increasing the pressure in the conduit  68  and flowing to the cylinders  74   a - 74   d . Since the pressure in the conduit  68  will be greater than the pressure in the conduit  66 , the connectors  75   a - 75   d  will be moved in a deceleration direction, decreasing the displacement and, therefore, the output torque of the motors  76   a - 76   d.    
   Advantageously, there is no direct connection between the accelerator  70  and the engine  12 . Rather, the engine  12  is operated and controlled based on a combination of engine speed (based on the signal on the line  42 ), torque (based on the position of the displacement control valve  60 , which is affected by the position of the accelerator  70 ), and system pressure (based on the signal on the line  38   a ). This combination of inputs allows the throttle control module  40  of the system  10  to always run the engine  12  at its peak efficiency, based on known engine efficiency parameters and, therefore, provide proportional control of the engine  12  and system  10 . At times when the system  10  is fully charged, the engine  12  can be advantageously turned off, reducing the instant fuel consumption to zero. When the system pressure drops, the engine  12  is restarted to again provide pressure to the conduit  20 . 
   Based on the condition or operating state of the air conditioning compressor  24 , the power maintenance module  28 , and the accumulator  38  (as determined by their respective signals on the lines  24   a ,  28   a , and  38   a ), the throttle control module  40  sends a signal on the line  42  to start or stop the engine  12  and/or vary the displacement of the pump/motor  16 . 
   As the system pressure in the conduit  20  increases, the accumulator  38  fills and the rate of flow from the pump/motor  16  is reduced. The flow of the pump/motor  16  continues to be reduced until the system pressure drops due to an output to the motors  76   a - 76   d . If at any time the flow of the pump/motor  16  reaches zero flow, the engine  12  may be turned off until flow is again needed. The flow of the pump/motor  16  may also be reduced if an accessory requires power to prevent the engine  12  from stalling (assuming the accessory is clutched to the engine  12 ). The powertrain system  10  obtains its efficiency by averaging the rate of power consumption. Energy needed for intermittent bursts is supplied by the stored energy in the accumulator  38 . The pump/motor  16  provides flow greater than the average flow needed to propel the vehicle. The extra flow created by the pump  16  is then stored in the accumulator  38 . 
   The hydraulic hybrid powertrain system  10  in accordance with the present invention advantageously providing an uncomplicated and straightforward control methodology and a very responsive control means for the system  10  by virtue of the fact that output torque response from the motors  76   a - 76   d , once their displacement is increased, is very quick. 
   Those skilled in the art will appreciate that the system  10  in accordance with the present invention may be utilized to supply hydraulic power to any number of systems including, but not limited to, a propulsion system for a floating or submersible vessel such as a ship, a boat, or a submarine, a propulsion system for a helicopter, among others. In short, the output of the pump/motor  16  could be utilized with the powertrain system  10  to run any number of hydraulic motors, such as the motors  76   a - 76   d  for any number of purposes while remaining with the scope of the present invention. 
   The connectors  73 ,  75   a - 75   d , and  80 , and the signals on the lines  24   a ,  28   a ,  38   a , and  42  may be any type of mechanical connector, such as a hydraulic line, a cable, a metal bar or the like, or an electrical signal communicating with solenoid valves or the like, while remaining within the scope of the present invention. 
   In accordance with the provisions of the patent statutes, the present invention has been described in what is considered to represent its preferred embodiment. However, it should be noted that the invention can be practiced otherwise than as specifically illustrated and described without departing from its spirit or scope.