Abstract:
A rotary fluid pressure device having a housing member, a manifold assembly, an internally generated rotor type gerotor set, an end plate, and a rotatably journaled torque transfer shaft interconnected with the gerotor set and extending within the housing member and manifold assembly. The gerotor set having at least an internally toothed stator member and a rotating rotor member disposed within the stator member. The rotor member having a first and second axial end surface and external teeth which interengage with the internal teeth of the stator to define a plurality of expanding and contracting volume chambers. The rotor member also having a plurality of circumferentially spaced laterally directed fluid paths fluidly connecting the manifold assembly with a plurality of circumferentially spaced radiating fluid paths which directly connect with the volume chambers. The fluid pressure device further including a plurality of coupling members for interconnecting the components.

Description:
CROSS-REFERENCE TO RELATED CASES  
       [0001]    The present application claims the benefit of the filing date of U.S. Provisional Application Serial No. 60/410,738 filed Sep. 13, 2002. 
     
    
     
       FIELD OF THE INVENTION  
         [0002]    The present invention relates to a rotary fluid pressure device, and more particularly to a gerotor motor wherein a gerotor set has an externally toothed rotor member with a plurality of circumferentially spaced radiating fluid paths in the rotor directly connecting axial fluid paths with volume chambers.  
         BACKGROUND OF THE INVENTION  
         [0003]    One type of rotary fluid pressure devices is generally referred to as gerotors, gerotor type motors, and gerotor type pumps, hereinafter referred to as gerotor motors. Gerotor motors are compact in size, low in manufacturing cost, have a high-torque capacity ideally suited for such applications as turf equipment, agriculture and forestry machinery, mining and construction equipment, as well as winches, etc. Gerotor motors have gerotor sets, which utilize a special form of internal gear transmission consisting of two main elements: an inner rotor and an outer stator.  
           [0004]    The inner rotor and the outer stator possess different centers. The inner rotor has a plurality of external teeth, which contact circular arcs on the interior of the outer stator when it revolves. Gerotor sets have volume chambers, which are separated by continuous contact between the rotor teeth and stator arcs. The volume changes as the rotor revolves with each chamber experiencing expansion or contraction. The rotary mechanism of the gerotor set, by virtue of its continuous chamber volume change, can be used as a positive displacement fluid controller. Gerotor motors, with a stationary outer stator and orbiting inner rotor, have a commutation device for valving flow to and from the chambers in time relation to the movement of the rotor. The output shaft is either directly connected to the orbiting inner rotor or is connected thereto by a drive link splined at each end. When pressurized fluid flows into a motor, the resistance of an external torsional load on the motor begins to build differential pressure, which in turn causes the inner rotor to rotate in the desired direction via a timing valve.  
           [0005]    Gerotor motors are typically manufactured in two forms, an internally generated rotor (hereinafter referred to as “IGR”) gerotor set or an externally generated rotor (hereinafter referred to as “EGR”) gerotor set. The outer stator of both IGR and EGR gerotor sets have one more tooth (N+1 teeth) than the inner rotor (N teeth). When the inner rotor rotates, it also orbits in the opposite direction of rotation with the speed of N times its own rotation. The vane pocket of the EGR is located on the outer stator and the vane pocket of the IGR is located on the inner rotor. During the motor operation, roller vanes mesh with external gear teeth of the inner rotor for an EGR rotor set and mesh with internal gear teeth of the outer ring for an IGR rotor set.  
           [0006]    For both EGR and IGR gerotor sets, the inner rotor can be used as a timing device for valving fluid in a timely manner. Prior art, such as U.S. Pat. No. 2,989,952 to Charlson, and U.S. Pat. No. 3,825,376 to Peterson et al., use EGR gerotor sets which do not efficiently use the inner rotor as a timing device due to the number of volume chambers being one larger than the number of external teeth, and thus one larger than the fluid passages in the inner rotor. This extra volume chamber is trapped during operation, creating excessive high pressure or cavitation during operation. To avoid this, the working fluid has to be detoured into each fluid chamber via a side manifold plate and cannot be directly valved within the EGR gerotor set. The present invention is able to use flow passages in the inner rotor for direct valving since the number of fluid chambers and number of flow passages are the same. The previously-noted prior art patents also use the bearing surface of the inner rotor for openings of the passages into the volume chambers which causes stress concentration and significantly reduces the life of the gerotor set. The U.S. Pat. No. 3,825,376 also has the passageway opening at the bottom-most point of the rotor external gear. Typically, the peaks and valleys of the bearing surfaces are used for sealing. Placing an opening at the valley allows for cross-port leakage which in turn causes poor volumetric efficiency.  
           [0007]    Prior art designs use conventional wear plate assemblies and conventional disk valve assemblies, which typically consist of a rotary disk valve driven by a drive link, a stationary manifold, and a pressure compensation device to close off the clearance of the valve interface at high pressure. The present invention eliminates the wear plate, since the manifold serves as a wear plate between the front housing and the gerotor set, and eliminates the disk valve assembly, since the valving function has been integrated into the rotor. The elimination of these components significantly reduces the number of parts for the gerotor motor. Consequently it reduces the number of areas where cross-port leakage can occur.  
           [0008]    In other prior art constructions, such as those set forth in U.S. Pat. Nos. 4,357,133, 4,697,997, 4,717,320 and 4,872,819 all to White, Jr., the motor uses a conventional EGR gerotor set. A circular commutator ring is integrated on the rotor for fast speed valving of the motor. To avoid possible high no-load pressure drops caused by narrow fluid passages and to reduce the length of the motor, these motors use an inner rotor with a very aggressive rotor profile, having a large eccentricity. Therefore, the drive link of the motor has a very large wobble angle. This causes heavy contact stress on the splines of the drive link, which may reduce the torque capacity or life of the drive link. In order to reduce the large wobble angle of the drive link, these motors are extended by making the drive link longer. The present invention has a similar volume displacement capability of these prior art EGR gerotor motors while having half the eccentricity. This 50% reduction of eccentricity significantly reduces the wobble angle of the drive line. Therefore, the splines of each end of the drive link in the present invention need not be heavily crowned. Also, the contact area of the external (drive link) and internal (rotor and drive shaft) splines is larger than those of the prior art. This increase in spline contact area improves the torque capacity of the drive link and makes the motor more reliable when operated under a high torque load.  
           [0009]    In another prior art reference, U.S. Pat. No. 4,741,681 to Bernstrom, the rotary fluid pressure device utilizes a valve-in-star (rotor) type valving. This prior art structure is different from the present invention in several areas. First, the valve-in-star uses an EGR gerotor set rather than the IGR gerotor set, as is the case in the present invention. It is also limited to closed-loop applications due to its intrinsic imbalance, having three pressures at the front side of the rotor and two pressures at the rear side of the rotor. This prior art structure also uses a side plate/manifold to reach the gerotor set volume chambers. Specifically, pressurized fluid flows through the manifold, to the rotor, back to the manifold after timely valving, and then reaches the volume chambers. As noted above, the present invention uses its rotor for direct fluid valving.  
         SUMMARY OF THE PRESENT INVENTION  
         [0010]    A feature of the present invention is to provide a rotary fluid pressure device comprised of a housing member, a manifold assembly, an internally generated rotor type gerotor set, an end plate, a rotatably journalled torque transfer shaft, and a plurality of coupling members for conducting fluid radially through the gerotor set. The housing member has a fluid inlet port, a fluid outlet port, a first flow passage, a second flow passage and an internal bore. The manifold assembly has a first fluid passage, a second fluid passage, an internal bore and one side adjoining the housing member. The internally generated rotor type gerotor set has at least an internally toothed stator member, and an externally toothed rotor member disposed within the stator member having an internal bore and a first and second axial end surface. One of the at least one stator and the rotor members having orbital movement relative to the other member and the rotor member has a rotational movement relative to the stator. The internal teeth of the stator member and the external teeth of the rotor member interengage to define a plurality of expanding and contracting volume chambers. A plurality of circumferentially spaced laterally directed fluid paths in the rotor fluidly connects with the manifold assembly first and second fluid passages. A plurality of circumferentially spaced radiating fluid paths in the rotor directly connect respective ones of the plurality of laterally directed fluid paths in the rotor to the volume chambers. The gerotor set is located between the manifold assembly and the end plate. The rotatably journalled torque transfer shaft is operatively interconnected to the rotor and extends from the housing member. The plurality of coupling members interconnect the endplate, gerotor set, manifold assembly and the housing member.  
           [0011]    Another feature of the noted rotary pressure device includes having the plurality of laterally directed fluid paths extend through the rotor. An added feature includes having the plurality of laterally directed fluid paths being substantially axially directed. Further the plurality of radiating fluid paths can be substantially radially directed.  
           [0012]    A further feature in the noted rotary pressure device includes having the plurality of radiating fluid paths in the rotor being located with the rotor between externally toothed members. Additionally the plurality of radiating fluid paths in the rotor can be substantially laterally centered between the rotor first and second axial ends. Also, the plurality of radiating fluid paths in the rotor can be substantially circumferentially centered between adjacent ones of the externally toothed members thereof. Further the plurality of radiating fluid paths in the rotor can be substantially laterally centered between the rotor first and second axial ends and are substantially circumferentially centered between adjacent ones of the externally toothed members thereof.  
           [0013]    Another feature of the noted rotary pressure device includes having the plurality of radiating fluid paths in the rotor being located in the rotor between externally toothed members thereof at at least one of the first and second axial ends. Further the plurality of radiating fluid paths can be located in the rotor between externally toothed members thereof at both of the first and second axial ends.  
           [0014]    A further feature of the noted rotary pressure device includes having it function as one of a hydraulic pump and motor. Another feature includes having the housing member&#39;s first and second flow passage, and the manifold assemblies&#39; first and second fluid passage being utilized for bi-directional fluid passage.  
           [0015]    An additional feature of the noted rotary pressure device includes having an internal drive link interposed between and operatively interconnected with the rotor and the torque transfer shaft. Additional the torque transfer shaft can be comprised of a straight shaft. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0016]    [0016]FIG. 1 is a perspective view of a hydraulic motor according to the present invention.  
         [0017]    [0017]FIG. 2 is a sectional view of the hydraulic motor.  
         [0018]    [0018]FIG. 3 a  is a cross-sectional view of a gerotor, a component of the hydraulic motor, shown from a first axial end.  
         [0019]    [0019]FIG. 3 b  is a cross-sectional view of the gerotor, similar to FIG. 3 a , but shown from the opposite axial end.  
         [0020]    [0020]FIG. 4 a  is an elevational view of the rotor, as viewed from a first axial end.  
         [0021]    [0021]FIG. 4 b  is an elevational view of the rotor, similar to FIG. 4 a , but shown from the opposite axial end as that in FIG. 4 a.    
         [0022]    [0022]FIG. 5 a  is a frontal view of a manifold plate adjacent the shaft housing of the hydraulic motor.  
         [0023]    [0023]FIG. 5 b  is a frontal view of the middle manifold plate.  
         [0024]    [0024]FIG. 5 c  is a frontal view of a manifold plate adjacent the gerotor.  
         [0025]    [0025]FIG. 6 a  is an end view showing the rotor relative to the stator at 0°.  
         [0026]    [0026]FIG. 6 a ′ shows FIG. 6 together with the manifold plate.  
         [0027]    [0027]FIG. 6 b  is an end view showing the rotor relative to the stator at 18° counterclockwise.  
         [0028]    [0028]FIG. 6 b ′ shows the rotor relative to the adjacent manifold plate at 18° counterclockwise.  
         [0029]    [0029]FIG. 6 c  is an end view showing the rotor relative to the stator at 36° counterclockwise.  
         [0030]    [0030]FIG. 6 c ′ shows the rotor relative to the adjacent manifold plate at 36° counterclockwise.  
         [0031]    [0031]FIG. 7 a  is a frontal view of a channeling plate of the present invention taken along line  7   a - 7   a  in FIG. 2.  
         [0032]    [0032]FIG. 7 b  is a sectional view of the flexible balancing plate taken along line G-G of FIG. 7 a.    
         [0033]    [0033]FIG. 7 c  is a rear view of the channeling plate taken along line  7   c - 7   c  in FIG. 2.  
         [0034]    [0034]FIG. 8 a  is a rear view of an end cover of the present invention.  
         [0035]    [0035]FIG. 8 b  is a cross-sectional side view of an alternate embodiment of end cover taken along line  8   b - 8   b  of FIG. 8 c.    
         [0036]    [0036]FIG. 8 c  is a frontal view of the alternate embodiment of the end cover.  
         [0037]    [0037]FIG. 9 is a schematic illustration of the fluid circuit of the hydraulic motor of this invention showing the high pressure inlet flow and the exhaust flow.  
         [0038]    [0038]FIG. 10 is a further embodiment of the present invention, showing a sectional view of the hydraulic motor.  
         [0039]    [0039]FIG. 11 shows a cross-sectional view of a gerotor of the further embodiment, shown from a first axial end.  
         [0040]    [0040]FIG. 12 shows a cross-sectional view of the gerotor of the further embodiment, similar to FIG. 11, but shown from the opposite axial end. 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0041]    Referring to the drawings, and initially to FIG. 1, it illustrates a compact rotary fluid pressure device  10  utilizing an IGR (Internally Generated Rotor), such as a hydraulic motor or pump (hereinafter referred to as “hydraulic motor” for ease of description) according to the present invention. Hydraulic motor  10  is designed for various applications, but is especially adapted for high torque, low speed use. As is discussed in detail below, hydraulic motor  10  is fully hydraulically balanced, has a simplified flow distribution through the manifold and gerotor set, and has a reduced number of individual components. In addition, this new design provides high starting torque while retaining high durability.  
         [0042]    As shown in FIGS. 1 and 2, hydraulic motor  10  includes the following main components: Shaft housing  13  is located at one end (front) of rotary fluid pressure device  10  and surrounds a torque-transfer shaft, which could be comprised of a coupling shaft  20  or a straight-shaft  120  (shown in FIG. 10). A first and a second port,  15 ,  16 , are integrated into shaft housing  13  and alternately provide, depending on the direction of rotation of shaft  20 , an inlet and outlet port for hydraulic motor  10 . An end cover  70  is located at the other end (rear) of hydraulic motor  10 . A channeling plate  90  is located inwardly adjacent to end cover  70 . A drive assembly  30  is interposed between shaft housing  13  and channeling plate  90 . A drive link  25  extends through drive assembly  30  and into shaft housing  13 . A plurality of peripherally-spaced bolts  80  extend through holes  81  (shown in FIG. 3) and connect end cover  70 , channeling plate  90 , drive assembly  30  and shaft housing  13 .  
         [0043]    Shaft housing  13  has a stepped internal bore  17  for receiving and rotatably supporting coupling shaft  20 . Within an axial front portion of internal bore  17 , a dirt seal  21  is positioned surrounding shaft  20  and prevents outside contaminants from entering internal bore  17 . Two axially-spaced radial bearings  22  are located within internal bore  17  for rotatably supporting shaft  20 . A high pressure shaft seal  23  is provided in a fluid-tight arrangement around shaft  20  in order to prevent any internal fluid from leaking into the front portion of bore  17 . Two axially-spaced thrust bearings  24  are located within internal bore  17  and prevent coupling shaft  20  from moving axially. Extending axially from an inner end of second port  16  is an axial passageway  36  that connects port  16  with a circumferential fluid chamber  37  abutting one end of drive assembly  30 .  
         [0044]    Coupling shaft  20  has a rear clevis portion  27  having a hollow center with internal splines. Coupling shaft rear portion  27  includes an axial passageway  28  that extends from its hollow center into a radial passageway  29 , which in turn is in fluid communication with a fluid chamber  18  located within shaft housing internal bore  17 . The coupling shaft rear portion  27  also includes radial flow passages  19  connecting fluid chamber  26  and fluid chamber  18 .  
         [0045]    Drive link  25  has a front portion  25   a  and a rear portion  25   b , both having external splines. The external splines on front portion  25   a  mate with complementary internal splines on coupling shaft rear portion  27 . The external splines on rear portion  25   b  mate with complementary internal splines in drive assembly  30 . A fluid chamber  26  surrounds drive link  25  and extends along a major portion of its axial extent.  
         [0046]    Drive assembly  30  includes a manifold  32  and a gerotor set  40 . Manifold  32  is comprised of a series of apertured individual plates  33   a - c  (shown in detail in FIGS. 5 a - c ) which are affixed together (e.g. by brazing or via peripherally-spaced bolts) in order to form two separate flow paths. The flow through all three affixed plates is shown in FIG. 9 and will be discussed in greater detail below. Each individual plate has a different path configuration extending therethrough. Referring cursorily to FIG. 9, these affixed plates provide a first flow path  38  extending between shaft housing  13  and gerotor set  40 , and a second flow path  39  extending between gerotor set  40  and shaft housing  13  respectively.  
         [0047]    Referring now to apertured affixed plates  33   a - c , FIG. 5 a  shows plate  33   a , one side of which is directly adjacent to shaft housing  13 . The darker shaded apertures or areas  39   a  signify fluid from second flow path  39  (FIG. 9) through a central bore and the lighter shaded apertures or areas  38   a  signify fluid from first flow path  38  (FIG. 9) through a set of apertures radially spaced from central bore. The lighter shaded areas  38   a  align with fluid chamber  37  of shaft housing  13  when the components are assembled. FIG. 5 b  shows intermediate plate  33   b , one side of which is adjacent to, and aligned with, the other side plate  33   a , on the side opposite shaft housing  13 . As in FIG. 5 a , the lighter shaded areas  38   a  signify fluid from first flow path  38  and the darker shaded areas  39   a  signify fluid from second flow path  39 . As can be seen, lighter shaded areas  38   a  are in a series of comb-like apertures having inwardly directed radial tooth-like members. Darker shaded areas  39   a  are in a single aperture comprised of a plurality of circumferentially spaced outwardly radially directed finger-like openings in communication with the center. It should be noted that the aperture continues from the center of plate  33   b  to the finger-like extensions. As previously noted, plates  33   a - c  are aligned, and affixed together. FIG. 5 c  shows plate  33   c  that is positioned between the other side of plate  33   b  and one end of gerotor set  40 . Again the lighter shaded areas  38   a  signify fluid from first flow path  38  and the darker shaded areas  39   a  signify fluid from second flow path  39 .  
         [0048]    Referring now to FIG. 3 a , which shows gerotor set front side  40   a , and FIG. 3 b , which shows gerotor set back side  40   b , gerotor set  40  consists of an outer stator  41  and an inner rotor  45 . Outer stator  41  has a plurality, N+1, of internal gear teeth  42 , that provide conjugate interaction with a plurality, N, of gear teeth  46  on the outer periphery of inner rotor  45 . Rotor gear teeth  46  preferably have a circular arc shape and can be replaced with hardened rollers for high efficiency gerotor set motors. The use of hardened rollers for rotor gear teeth  46  reduces wear, friction, and leakage in the hydraulic motor.  
         [0049]    Referring to FIG. 4 a , the front side  58 , or the side adjacent manifold plate  33   c , of rotor  45  is shown. Front side  58  shows two sets of pluralities of passages, axial passages  48  and axial through orifices  51 , both extending through the rotor. Both sets of passages  48  and  51  have openings on both axial sides of rotor  45  (as shown in FIGS. 4 a - b ). As will be discussed in detail below, each axial passage  48  is used as a passageway for high-pressure fluid and exhaust fluid. As will also be discussed below, each axial through orifice  51  is used for improving the rotary movement of rotor  45 . The outer periphery of rotor  45  is defined by a series, nine in the example shown in FIG. 4 a , of equally circumferentially-spaced intermediate portions  52  separated via a series of semi-cylindrical pockets or recesses  53  which serve to receive rotor gear teeth or rollers  46 . Spaced portions  52  have a radial outer surface which preferably is substantially perpendicular (but not limited thereto) to rotor front side  58 , rotor back side  63 , and any radial plane emanating from the axial center line of the rotor internal bore, or apertured center. The apertured center of rotor  45  is provided with internal splines  50  located at its peripheral surface for mating engagement with the external splines of drive line rear portion  25   b . This engagement transfers high torque from rotor  45  to drive link  25  and from same to coupling shaft  20 .  
         [0050]    [0050]FIG. 4 b  shows the rear side surface  63 , or the side adjacent channeling plate  90 , of rotor  45 . Axial passages  48  and axial through orifices  51 , both extending from front side surface  58 , are shown. Surrounding each through orifice  51  and extending slightly axially into rotor rear side  63  is a recess  51   a  which can be trapezoidal in shape and is coaxial with orifice  51 . The radial upper or outer portion of each axial passage  48  is provided with another recess  48   a , which also can be trapezoidal in shape, and extends radially outward into flat portion  52 . During operation, recesses  48   a  and  51   a  are filled with fluid for the purpose of reducing the viscous friction between rotating rotor  45  and non-rotating channeling plate  90 . Viscous friction is also reduced due to the reduction of the outer annular area of rotor rear side surface  63  via recesses  48   a  and  51   a . A flower-shaped or multiple-convoluted recess  64  is positioned radially outward of rotor internal splines  50  in rotor rear side surface  63  and continues along the whole circumference thereof. As will be discussed below, recess  64  always receives high pressure fluid in order to overbalance rotor  45 , thus axially biasing rotor  45  towards manifold  32  in order to reduce fluid leakage between manifold  32  and gerotor set  40 , which interface is referred to as the valve interface.  
         [0051]    Rotor  45  has a plurality, N, of central, individual radial fluid channels  47  within flat portions  52 . Radial fluid channels  47  are preferably at least one of substantially axially centered between rotor front side  58  and rear side  63 , and substantially circumferentially centered relative to their adjacent rotor gear teeth  46  (FIG. 3 a ), and preferably both substantially axially and substantially circumferentially centered. One (inner) end of each radial fluid channel  47  opens into an axial passage  48 , extending through rotor  45 , and the other (outer) end opens radially into a gerotor set volume chamber  54  (as shown in FIGS. 3 a - b ). The end of passage  48  that opens into gerotor set volume chamber  54  is preferably centered within equally circumferentially spaced intermediate portions  52 . Each volume chamber  54  is bounded by two nearby inner rotor gear teeth  46 , circumferentially-spaced portion  52  of the rotor outer peripheral surface, and the undulating internal surface of stator  41 . Gerotor set  40  has N volume chambers, which coincides with the number of fluid channels  47 . Rotor  45  also has a plurality, N, of individual radial fluid channels  55  located at either, or both, rotor front side  58  or rotor rear side  63  of rotor  45 . Radial fluid channels  55  are shown at rotor front side  58 , but can also be placed on rotor rear side  63 . Radial fluid channels  55  are preferably circumferentially centered in the manner preferably described with reference to channels  47 , and preferably parallel with channels  47 .  
         [0052]    Referring to FIGS. 2, 3 a  and  3   b , stator  41  is shown in detail. As mentioned above, stator  41  has internal gear teeth  42 , that interact with gear teeth  46  of inner rotor  45 . Located radially outward of gear teeth  42  are bolt holes  81  for receiving bolts  80 , which affix stator  41  between a channeling plate  90  and manifold  32 . A through hole  43  extends axially through stator  41 . Positioned radially outward of through hole  43  are two circumferential seal cavities  44 , located on both axial end surfaces of stator  41 , for receiving seals  67 .  
         [0053]    Referring to FIGS. 7 a - c , channeling plate  90  is shown with bolt holes  81 , for receiving bolts  80  (not shown), extending therethrough. A first check valve opening  91  extends through channeling plate  90 , with check valve opening  91  being defined by a first portion  91   a  and a second portion  91   b . First portion  91   a  has a diameter larger than second portion  91   b  such that it can receive a check ball (not shown) having a diameter larger than that of second portion  91   b . When assembled, as shown in FIG. 2, second portion  91   b  is aligned with stator through hole  43  and is in fluid communication with first flow path  38  (as shown in FIG. 9). A second check valve opening  92  also extends through channeling plate  90 , and, similar to check valve opening  91 , opening  92  has a first portion  92   a  and a second portion  92   b . First portion  92   a  has a diameter larger than second portion  92   b  such that it can also receive a check ball (not shown) having a diameter larger than that of second portion  92   b . When assembled, as shown in FIG. 2, second portion  92   b  is coaxial with the center of gerotor set  40  and is in fluid communication with second flow path  39  (as shown in FIG. 9). At least one further through hole  93  and preferably a plurality of circularly spaced holes  93  extend through channeling plate  90  and are situated in a location between but not radially aligned with both first and second check valve openings  91  and  92 . When assembled, (not shown), at least one through hole  93  is aligned with multiple-convoluted recess  64  on the rotor back side  63  (as shown in FIG. 4 b ). It should be understood that the convoluted shape of recess  64  is due to the fact that rotor  45  both rotates and orbits at the same time. At least one through hole  93  supplies high pressure fluid to multiple-convoluted recess  64 . FIG. 7 c  shows the inner axial surface  90   b  of channeling plate  90  which is directly adjacent end cover  70 . A coaxial circular recess  96  for receiving high pressure fluid, detailed below, is shown. A recessed coaxial annular seal cavity  97  is positioned, radially outside of bolt holes  81  with seal cavity  97  receiving seal  67  (not shown). Recess  96  has a flow channel  96   a  extending radially outward and terminating into seal cavity  97 . Check valve opening  91 , and more specifically first portion  91   a , is centered within flow channel  96   a.    
         [0054]    Referring to FIG. 8 a , the substantially flat outer axial surface of end cover  70  is shown. In the present invention, the inner axial surface of end cover  70  is substantially similar to that of the axial outer surface shown in FIG. 8 a . Bolt holes  81  extend through end cover  70  and receive bolts  80 , not shown, which align end cover  70  with channeling plate  90 . As part of another embodiment of the invention, FIGS. 8 b - c  show how recess  96  and seal cavity  97  of channeling plate  90  can alternately be incorporated into the inner axial surface of end cover  70  rather than being incorporated in channel plate  90 . Similar to the design of FIGS. 7 b  and  7   c , a coaxial circular recess  72  is incorporated into the inner axial surface of end cover  70  for receiving high-pressure fluid. A recessed coaxial annular seal cavity  71  is positioned, radially outside of bolt holes  81 , in end cover  70 , with seal cavity  71  receiving a seal, similar to seal  67 . FIG. 8 c  shows the inner axial surface of end cover  70 , as part of the alternate embodiment, which is directly adjacent channeling plate  90 . Recess  72  has a flow channel  73  extending radially outward, with flow channel  73  having its radial outer portion  74  terminating into end cover seal cavity  71 . When assembled, flow channel radial outer portion  74  is radially and axially aligned with first portion  91   a  of first check valve opening  91 .  
         [0055]    The hydraulic circuit and operation of hydraulic motor  10  will now be discussed. Referring first to FIG. 9, the fluid path for hydraulic motor  10  is shown when it operates in a first direction. High pressure fluid  38  enters second port  16  and follows the path indicated by darker shading with triangular shapes. It should be noted that although fluid  38  is shown entering port  16  in FIG. 9, this path could be reversed with exhaust fluid emanating therefrom. Ports  15  and  16  can be either inlet or outlet ports, depending on the desired direction of rotation of hydraulic motor  10 . For sake of description, the triangular shaded path was chosen to represent high pressure inlet fluid  38 , with fluid  38 , entering port  16 , traveling axially through passageway  36  and entering fluid chamber  37 . Fluid  38  then travels into manifold  32  through the axially aligned passages in manifold plate  33   a  (as seen and indicated by  38   a  in FIG. 5 a ). Fluid  38  further flows axially from plate  33   a  into plate  33   b  (as shown and indicated by  38   a  in FIG. 5 b ) and travels radially inwardly while passing through this plate. Fluid  38  continues its flow into and axially through a plurality, N+1, of aligned openings  34  in plate  33   c  (as shown and indicated by  38   a  in FIG. 5 c ), with openings  34  being aligned with rotor axial passages  48  and fluid  38  passing into these passages. Finally, fluid  38  then flows radially outwardly through fluid channels  47  (FIG. 4 b ) within rotor  45  into gerotor set volume chambers  54 . Fluid  38  also flows radially outward through fluid channel  55  (FIGS. 4 a  and  9 ) into volume chambers  54 . The pressurized fluid  38  causes volume chambers  54  to expand. As well known to those skilled in the art, this fluid communication causes rotor  45  to rotate and orbit within fixed stator  41 . The expanding volume chambers, coupled with the rotation and orbiting of rotor  45 , i.e., hypocloidal movement, will cause other volume chambers  54  to contract. Contraction of volume chambers  54  provides the exhausting, or return fluid flow indicated by second flow path  39 .  
         [0056]    Exhausting fluid  39  is indicated with dotted shading, and begins its flow with the contraction of gerotor set volume chambers  54  forcing exhaust fluid  39  radially inwardly through rotor fluid channels  47 . Fluid  39  enters axial fluid passages  48  (FIG. 4 c ), flows towards plate  33   c  and enters the aligned openings  34  therein (as shown and indicated by  39   a  in FIG. 5 c ). Fluid  39  then travels into manifold plate  33   b  and flows radially inwardly while passing therethrough (as shown and indicated by  39   a  in FIG. 5 b ). Fluid  39  continues its flow axially through the center of plate  33   a  (as shown and indicated by  39   a  in FIG. 5 a ).  
         [0057]    Drive link  25  (FIG. 9) extends freely through the center of manifold plates  33   a - c  and its rear end  25   b  is linked to rotor  45 , via the previously-described cooperating spline arrangement, and rotates and orbits with rotor  45 . Therefore, the portion of drive link  25  that extends through the center of manifold plates  33   a - c  is not sealed against the inside surface of plates  33   a - c . Thus fluid  39 , upon reaching the center of plate  33   b  is free to travel along the outside surface of drive link  25 . This provides a lubricant for drive link  25 , as well as being an exhaust path for the fluid flow. Exhaust fluid  39  will travel axially along drive link  25  towards coupling shaft  20  then radially outward through passageway  19  within shaft housing  13 . Exhaust fluid  39  then reaches fluid chamber  18  where it continues radially outward and exits through first port  15 , which in this example functions as an outlet port. Exhaust fluid  39  will occupy all gap areas between drive link front portion  25   a  and coupling shaft  20 , and all areas between coupling shaft  20  and shafting housing  13 . Radial passageway  29  provides a path between the areas surrounding coupling shaft  20  and the areas within coupling shaft  20 . Fluid  39  passing through these areas provides lubrication for these moving parts and removes heat. Due to the rotation of coupling shaft  20 , the centrifugal flow of fluid through radial passageway  29  takes the heat away from seal  23  and thrust bearings  24 , while traveling towards and out of first port  15 .  
         [0058]    It should again be noted that the directions of fluid travel are chosen for example purposes only and can be reversed by switching the fluid streams communicating with ports  15  and  16 . If the fluid streams were reversed, high-pressure fluid would then enter port  15  and would travel in the direction indicated by the dotted shading. After entering port  15 , high pressure fluid would flow into shaft housing  13 , axially along drive link  25  through the central aperture of plate  33   a  and radially upwardly into manifold plate  33   b . Unlike the above discussed example, in which high pressure fluid enters manifold  32  axially, high pressure fluid would now enter manifold  32  radially. As mentioned above, the aperture in manifold plate  33   b  extends from the center radially outwardly so high-pressure fluid can travel from directly from the central internal bore radially outward before flowing in the axial direction.  
         [0059]    Referring again to FIG. 9 and the example where high pressure fluid  38  enters port  16 , when high pressure fluid  38  reaches manifold plate  33   c , a certain amount of fluid travels through an axial passageway  35  (which is comprised of portions  35   a - c ) in manifold plates  33   a - c  respectively into aligned stator through hole  43 . If the pressure of this fluid  38  is greater than a predetermined value it will crack a first check valve  94  and fill channeling plate recess area  96 . Fluid  38  will then travel via at least one through-hole  93  in channeling plate  90  and fill flower-shaped recess  64  (as shown in FIG. 4 b ) in rotor back side  63 . In a similar fashion, when high pressure fluid enters port  15  and travels in a direction indicated by the dotted shading in FIG. 9, fluid  39  will travel along the outer surface of drive link rear portion  25   b  and will crack, if the pressure is sufficient, a second check valve  95  in channeling plate  90 . Fluid  39  will fill channeling plate recess area  96 , flow via at least one through-hole  93  in channeling plate  90  and fill flower-shaped recess  64  in rotor back side  63 . In either of these flow examples, high pressure fluid in flower-shaped recess  64  would act on rotor back side  63  and axially bias rotor  45  toward manifold  32 . This biasing action will substantially reduce leakage between gerotor set  40  and manifold  32 .  
         [0060]    Although channeling plate  90  has high-pressure fluid passing (in both axial directions) therethrough, it remains substantially rigid due to its thickness. As an example, a 5″ diameter channeling plate  90  can have a thickness of approximately 0.5″, so that it will only negligibly deform and not physically contact rotor  45 . This lack of deformation is unlike prior art designs which provide thinner, flexible balancing plates which come in physical contact with the rotor to provide stability to an unbalanced rotor. Channeling plate  90  acts as a passageway for directing high-pressure fluid, either 38 or 39, towards rotor  45 . Unlike prior art designs, where the channeling plate will flex and contact the rotor in order to minimize the gap between the rotor and the manifold set, the present invention uses only high-pressure fluid to bias rotor  45  toward manifold  32  in order to minimize the gap. Therefore channeling plate  90  does not physically contact rotor  45  as a result of the negligible elastic deformation of channeling plate  90 , but merely provides a passageway for the high-pressure fluid. A thin layer of high-pressure fluid separates channeling plate  90  and rotor  45 . Since only high-pressure fluid is received within flower-shaped recess  64 , the pressure on rotor backside  63  is greater than the pressure on rotor front side  58 . Without the hydraulic biasing force provided by the high-pressure fluid acting on rotor  45  via recess  64 , the pressure forces on opposite rotor sides,  58  and  63 , is substantially equal.  
         [0061]    Referring to FIGS. 6 a - c  and  6   a ′- c ′, gerotor set  40  has an inherently balanced rotor  45  due to axial passages  48  and through orifices  51 . Manifold  32 , and specifically manifold plate  33   c , has twenty aligned openings  34  which are adjacent to gerotor set  40 . Aligned openings  34  have alternating pressures, exhaust fluid  38   a  and high pressure fluid  39   a , which are valved with rotor axial passages  48  and through orifices  51 . Referring to FIG. 6 a , during operation axial passages  48  on the left side are filled with high pressure fluid  39   a  while axial passages on the right side are filled with exhaust fluid  38   a . Through orifices  51  on the left side are filled with exhaust fluid  38   a  while through orifices on the right side are filled with high pressure fluid  39   a . Without through orifices  51 , rotor  45  would have an imbalance of hydraulic force (half seeing forces from high-pressure fluid  39   a  and the other half seeing forces from exhaust fluid  38   a ). With through orifices  51 , these forces are equally distributed throughout the circumference of rotor  45 . Forces on rotor backside  63  are similarly distributed throughout the rotor circumference since axial passages  48  and through orifices  51  extend through rotor  45 . If axial passages  48  and through orifices  51  did not extend through to rotor back side  63 , the center of hydraulic force at rotor back side  63  would move away from the center of rotor  45  since half of rotor back side  63  would have high pressure fluid  39   a  acting upon it (from volume chambers  54  which axial extend from gerotor set front side  40   a  to gerotor set back side  40   b ) and the other half would have exhaust fluid  38   a  acting upon it. This significant offset of hydraulic force would tip rotor  45  and cause excessive mechanical loading on rotor gear teeth  46 , thus creating excessive frictional loss. Once rotor  45  is tipped, it is no longer balanced. Adding high pressure filled flower shaped recess  64  to rotor back side  63  does not change the balance of rotor  45  since this high pressure force has a center that matches rotor  45  center.  
         [0062]    Referring to FIGS. 4 b  and  9 , when fluid  38  enters axial passage  48  and through orifice  51  in rotor  45 , it continues to flow to rotor back side  63  and fills axial passage recess  48   a  and through-orifice recess  51   a . As previously discussed, filling of recesses  48   a  and  51   a  with fluid reduces the viscous friction between rotating rotor  45  and channeling plate  90 . Fluid that flows through axial passage  48  and through-orifice  51  during the routine valving process will fill recesses  48   a  and  51   a  thus reducing the friction therebetween. Friction is also reduced due to the reduction of the outer surface area of rotor backside surface  63  via recesses  48   a  and  51   a . Reduction of friction not only improves the overall efficiency of rotary fluid pressure device  10  but also improves its longevity. The inclusion of recesses  48   a  and  51   a  on rotor back side  63  also reduces the area of transition pressure. Recesses  48   a  and  51   a  will be filled with either pressurized fluid or exhaust fluid. By maximizing, with the recesses, the area that is receiving a flowing, working fluid (the pressurized or exhaust fluid), the area that is not seeing the flowing, working fluid is minimized. The area not seeing working fluid is the transition area between recesses  48   a  and  51   a.    
         [0063]    When rotor  45  rotates, valving is accomplished at the flat, transverse interface of rotor front side  58  and the adjacent side of manifold plate  33   c . This valving action communicates pressurized fluid  38  to volume chambers  54 , causing the chambers to expand, and communicates exhaust fluid from the contracting volume chambers via radial fluid channels  47  and axial passages  48  in rotor  45 . FIGS. 6 a - c  and  6   a ′- c ′demonstrate the correctness of timely valving when rotor  45  is located at three different angular positions, 0°, 18° (counter-clockwise), and 36° (counter-clockwise). Since the valving is integrated into rotor  45 , there is no timing error resulting from extra drivetrain components which have been eliminated here. In prior art designs, separate componentry, e.g. conventional disk valve assemblies, is needed for valving and the possibilities for cogging, or clocking, are much greater. A conventional disc assembly usually consists of a rotary disk valve driven by a drive link, a stationary manifold, and a pressure compensation device to close off the clearance of the valve interface at high pressure. By eliminating the separate disk valve assembly, the timing error is minimized which in turn improves the low speed performance of hydraulic motor  10 .  
         [0064]    [0064]FIGS. 6 a - c  show rotor  45  rotating, and orbiting, within stator  41 . High pressure fluid is shown with a darker, denser, shading. Exhaust fluid is indicated by a lighter, less dense, shading. FIGS. 6 a ′- c ′show gerotor set  40  over (or transposed onto) manifold  32 , and specifically manifold plate  33   c , with only the fluid inside manifold plate  33   c  having the shading. In this fashion, the positions of axial passages  48  and through orifices  51  relative to aligned openings  34  in manifold plate  33   c  are clearly shown.  
         [0065]    Referring to FIGS. 6 a  and  6   a ′, fluid denominated by numeral  39   a  in alternating aligned manifold plate openings  34  (FIG. 5 c ), indicates high pressure fluid and fluid denominated by  38   a , in alternate manifold plate openings  34 , indicates exhaust fluid. With rotor  45  rotating in a counter-clockwise direction within stator  41 , volume chambers  54 , extending (counter-clockwise) from the 12 o&#39;clock to the 7 o&#39;clock position (or those filled with high pressure fluid  39   a ), are expanding and volume chambers  54 , extending (counter-clockwise) from the 5 o&#39;clock to 12 o&#39;clock position (or those filled with exhaust fluid  38   a ), are contracting. The volume chamber at the 6 o&#39;clock position is in transition from expansion to contraction. As can be seen, each rotor axial passage  48  in the expanding region is axially aligned with a high pressure  39   a  manifold plate opening  34 . Each rotor axial passage  48  in the contracting region is axially aligned with an exhaust fluid  38   a  manifold plate opening  34 . At the six o&#39;clock position, rotor axial passage  48  is intermediate the high-pressure fluid  39   a  and exhaust fluid  38   a  manifold openings.  
         [0066]    In FIGS. 6 b  and  6   b ′ rotor  45  has rotated counter-clockwise 18° within stator  41 . Volume chambers  54  which are expanding are located (in a counter-clockwise fashion) from the 4 o&#39;clock to the 11 o&#39;clock position. Volume chambers  54  which are contracting are located (counter-clockwise) from the 11 o&#39;clock to the 6 o&#39;clock position. Volume chamber  54  located at the 5 o&#39;clock position is in transition from contraction to expansion. As can be seen, volume chambers  54  which are contracting have axial passages  48  aligned with exhaust fluid  38   a  and volume chambers  54  which are expanding have axial passages  48  aligned with pressurized fluid  39   a.    
         [0067]    In FIGS. 6 c  and  6   c ′ rotor  45  has rotated counter-clockwise 36° within stator  41 . Volume chambers  54  from the 10 o&#39;clock to the 6 o&#39;clock position (counter-clockwise) are expanding and volume chambers  54  from the 4 o&#39;clock to the 11 o&#39;clock position (counter-clockwise) are contracting. Volume chamber  54  located at the 5 o&#39;clock position is in transition. Volume chambers  54  which are expanding have axial passages  48  aligned with pressurized fluid  39   a  and volume chambers  54  which are contracting have axial passages  48  aligned with exhaust fluid  38   a.    
         [0068]    Illustrating the operation of gerotor set  40  from another perspective, the movement of rotor  45  relative to a stator internal gear tooth  42  situated at 11 o&#39;clock, will now be discussed. Referring to FIG. 6 a , volume chamber  54  (at 11 o&#39;clock) is expanding as it is filled with high-pressure fluid  39   a . As seen in FIG. 6 a ′, axial passage  48  is in partial axial alignment with opening  34  (which is filled with pressurized fluid  39   a ) in manifold plate  33   c . As rotor  45  rotates 18° counter-clockwise to the position shown in FIG. 6 b , rotor gear tooth  46  is in adjacent contact with stator internal gear tooth  42 . As seen in FIG. 6 b ′, axial passages  48  are located at 12 o&#39;clock, in axial alignment with opening  34  filled with pressurized fluid  39   a , and 10 o&#39;clock, in axial alignment with opening  34  for receiving exhaust fluid  38   a . As rotor  45  rotates 36° counter-clockwise to the position shown in FIGS. 6 c  and  6   c ′, the 11 o&#39;clock volume chamber  54  is contracting as fluid flows from volume chamber  54  through fluid channel  47  (as best shown in FIG. 4 b ), through axial passage  48  and into axially aligned opening  34  in manifold plate  33   c . Axial passage  48  is in partial axial alignment with opening  34  for exhaust fluid  38   a  in manifold plate  33   c.    
         [0069]    Referring back to FIG. 2, prior art designs typically have a wear plate located between shaft housing  13  and gerotor set  40  that absorbs any axial stresses caused by moving components. A wear plate can be replaced more readily than other componentry and ensures that the other componentry is not negatively affected by axial stresses. But the wear plate also provides another leak path at its connection with adjacent components. In the present invention, the wear plate has been eliminated. Manifold  32 , in addition to its manifold function, also serves as a wear plate between shaft housing  13  and gerotor set  40 . The elimination of a conventional wear plate reduces the number of parts for hydraulic motor  10  and also eliminates another possible leak path.  
         [0070]    Referring to FIG. 3 a , since rotor  45  has nine gear teeth  46  and stator  41  has ten gear teeth  42 , nine orbits of rotor  45  result in one complete rotation thereof and one complete rotation of coupling shaft  20  (FIG. 2). Thus, a 1:9 ratio of gear reduction is achieved. A 1:9 gear reduction along with gerotor set&#39;s  40  smooth rotor  45  profile significantly improves the low speed performance of hydraulic motor  10 . Similar motors have gear reduction ratios of 1:6 (for 6×7 EGR motors) or 1:8 (for 8×9 EGR motors).  
         [0071]    The fluid displacement capacity of hydraulic motor  10  is proportional to the multiple of N (number of rotor external gear teeth), N+1 (number of stator internal gear teeth), and the volume change of each volume chamber  54  of gerotor set  40 . The change of volume of each volume chamber  54  is approximately proportional to the eccentricity of gerotor set  40  if the value of N is fixed. The present invention, which uses a 9×10 gerotor set  40  (9 rotor gear teeth  46  and 10 stator gear teeth  42 ) has similar displacement capacity and overall size as a conventional 6×7 EGR gerotor set while its eccentricity is only one half of that of the 6×7 gerotor set. This 50% reduction of eccentricity significantly reduces the wobble angle of drive link  25  (which is used for operatively connecting rotor  45  and coupling shaft  20 ). Therefore, the splines of each end of drive link  25  do not need to be heavily crowned. The internal and external spline contact areas between drive link  25 , rotor  45  and coupling shaft  20  have a much larger contact area than that of a conventional 6×7 EGR gerotor set. Usually the life of gerotor set orbit motors is limited by the life of drive link  25 . The increase of spline contact area improves the torque capacity of drive link  25  and makes rotary fluid pressure device  10  more reliable when it is operated under high torque load.  
         [0072]    Referring to FIG. 7 c , when high pressure fluid fills recess  96 , fluid between end cover  70  and channeling plate  90  migrates into bolt holes  81 , classifying this motor as a “wet-bolt” type. It should be noted that regardless of the direction of rotation of compact hydraulic motor  10  (or the direction of fluid flow), high pressure fluid will fill bolt holes  81  since in both flow directions recess  96  will be filled with high pressure fluid. Therefore, it is necessary that seal  67  (FIG. 2) is placed radially outside of bolt holes  81  (into seal cavity  97 ) and that bolt holes  81  avoid first and second ports  15 ,  16  respectively. Since ports  15 ,  16  could either be at high or low pressure and the pressure within bolt holes  81  is only high pressure, it is necessary that the high pressure fluid within bolt holes  81  does not interconnect with a low pressure exhaust port. The use of a “wet-bolt” design in a motor is another way to reduce its size and weight.  
         [0073]    Leakage in hydraulic motors occurs at locations where components are connected or abut and is generally referred to as cross-port leakage. The present invention significantly reduces cross-port leakage by eliminating componentry. Specifically, since the valving operation is integrated into rotor  45 , hydraulic motor  10  has eliminated possible areas, e.g. the disk valve assembly, for cross-port leakage. In the prior art, in order to prevent leakage, designs have used tight fitting gerotor sets that create high friction and wear, thus negatively affecting the mechanical efficiency of the motor. In the present invention, the integration of parts has also eliminated extra mechanical friction between componentry which in turn increases the mechanical efficiency of hydraulic motor  10 .  
         [0074]    Referring to FIGS. 3 a  and  4   b , it should be noted that the present invention has an exceptionally high volumetric efficiency since rotor gear teeth  46  can compensate for any wear between the outer surface of rotor  45  and the inner surface of stator  41 . Over the operating lifespan of hydraulic motor  10 , the conjugation of rotor  45  and stator  41  will cause wearing to each surface. Typically this would create a leak path. Since each rotor gear roller  46  can move radially outwardly, relative to its pocket  53 , it can provide a reliable seal between adjacent volume chambers  54 . Otherwise fluid could leak from one volume chamber, at the roller/stator interface, to an adjacent volume chamber and fluid would not be discharged through radial fluid channel  47  as intended.  
         [0075]    Hydraulic motors can be classified as either having a two-pressure zone or a three-pressure zone. One skilled in the art will appreciate that this invention is applicable to both two and three-pressure zone motors. One skilled in the art will further appreciate that fluid pressure device  10  can be used as either a bi-directional hydraulic pump or motor. When used as a pump, coupling shaft  20  of course acts as an input or driving member in contrast to acting as the output or driven shaft in a motor.  
         [0076]    It should be noted that while the valve in rotor feature of the present invention is specifically applicable to an IGR-Type gerotor set, the features pertaining to the inherently balanced rotor  45 , the reduced sized manifold set  32 , and channeling plate  90  are not limited to an IGR-Type gerotor set, and could be utilized, for example, with an EGR-Type gerotor set.  
         [0077]    Referring to FIGS.  10 - 12 , a further embodiment 10′ of the present invention is shown. In this embodiment the componentry shown in FIG. 2 for hydraulic motor  10  remains the same with the exception of coupling shaft  20 , drive link  25 , and gerotor set  40 . Coupling shaft  20  and drive link  25  (in FIG. 2) have been replaced with a through, or straight, shaft  120 . Two piece gerotor set  40  (comprised of rotor  45  and stator  41 ) has been replaced with a three-piece gerotor set  140 , which now includes a rotor  145 , and inner orbiting stator  186 , and a fixed outer stator  141 . Straight shaft  120  is now directly connected with rotor  145  since rotor  145  only rotates, rather than rotating and orbiting as in prior embodiment  10 . Since rotor  145  only rotates, a circular recess  164  is provided to receive high pressure fluid rather than convoluted recess  64  in prior embodiment  10 . Outer stator  141  functions similarly to stator  41  in prior embodiment  10 . Orbiting inner stator  186  is added to gerotor set  140  and moves in a hypocycloidal fashion, similar to rotor  45  in prior embodiment  10 .  
         [0078]    Straight shaft  120  gerotor sets similar to this embodiment  10 ′ are well known in the art. An example of a commercially available straight shaft hydraulic motor having a three-piece gerotor set similar to embodiment  10 ′ of the present invention is fully shown and described in U.S. Pat. No. 4,563,136 to Gervais et al., as well as also being assigned to the assignee of the present invention.  
         [0079]    As stated above, all other componentry of this embodiment is the same as that shown in embodiment  10 . All inventive features, shown and described with reference to embodiment  10  are also present in embodiment  10 ′. Since embodiment  10 ′ has straight shaft  120 , three-piece gerotor set  140  is used in order for inner stator  186  to compensate for the orbiting movement within gerotor set  140 .