Hydraulic device with balanced rotor

Information

  • Patent Grant
  • 6743003
  • Patent Number
    6,743,003
  • Date Filed
    Tuesday, April 29, 2003
    21 years ago
  • Date Issued
    Tuesday, June 1, 2004
    20 years ago
Abstract
A rotary fluid pressure device having a housing member, a manifold assembly, a 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 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 first plurality of circumferentially spaced laterally directed fluid paths which extend through the rotor for fluid connection between the manifold assembly and the volume chambers, and a second plurality of circumferentially spaced, laterally directed fluid paths interposed between the first plurality of fluid paths for sequentially channeling fluid between both axial end surfaces.
Description




FIELD OF THE INVENTION




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, balanced rotor member with a first plurality of circumferentially spaced laterally directed fluid paths extending through the rotor and a second plurality of fluid paths being circumferentially interposed between the first plurality of fluid paths for sequentially channeling fluid between one of the first and second axial end faces.




BACKGROUND OF THE INVENTION




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.




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. An 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. 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.




Due to the flow of pressurized fluid through the gerotor sets, namely into and out of the volume chambers in the gerotor set, the inner rotor tends to have an imbalance of forces acting upon it. This imbalance of forces causes the rotor to tilt to one side during its rotation, resulting in unwanted wear along the surface of the rotor that comes in contact with an adjacent component, e.g. an end cap. Prior art constructions, such as those set forth in U.S. Pat. No. 5,624,248 to Kassen et al. have used an adjacent component, such as a plate, in order to balance the rotor that is tipping in one direction. The plate has hydraulic forces acting on one side, causing it to flex and come in physical contact with the rotor. This contact offsets the differential of forces which tip the rotor, allowing the rotor to rotate uniformly. The present invention uses hydraulically pressurized fluid to balance the rotor without having an extra component that physically contacts the rotor.




Other prior art constructions, such as those set forth in U.S. Pat. No. 4,264,288 to Wüsthof et al., provide opened/recessed slots on both sides of the hydraulic rotor for balance. This causes the rotor to remain axially aligned within the outer stator during its operation. The present invention differs from this prior art construction by providing two sets of axial through holes for balancing. One set of through holes transfers high pressure fluid, while the other set (which alternates between the first set) of through holes transfers exhaust fluid. This alternation of high pressure fluid and exhaust fluid on each side of the rotor provides the desired balance.




SUMMARY OF THE PRESENT INVENTION




The present invention provides a rotary fluid pressure device comprised of a housing member, a manifold assembly, a gerotor set, an end plate, and a rototably journaled torque transer shaft. This invention overcomes the obstacle of balancing components within the gerotor set during operation of the rotary fluid pressure device.




A feature of the present invention is to provide a rotary fluid pressure device where the housing member defines 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, with one side of the manifold assembly adjoining the housing member. The gerotor set has an internally toothed stator member, an externally toothed rotor member disposed within the stator member having an internal bore and a first and a second axial end surface. One of the stator and the rotor members has 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 first plurality of circumferentially spaced laterally directed fluid paths in the rotor extend through the rotor for fluid connection with the manifold assembly first and second fluid passages. A branch conduit for each of the first plurality of fluid paths adapted for directly connecting respective ones of the first plurality of laterally-directed fluid paths in the rotor to the volume chambers. A second plurality of circumferentially spaced, laterally-directed fluid paths extending through the rotor circumferentially interposed between the first plurality of fluid paths for sequentially channeling fluid between one of the first and second axial end faces. The gerotor set is located between the manifold assembly and the end plate. The rotatably journaled torque transfer shaft is operatively interconnected to the rotor and extends from within the housing member. A plurality of coupling members interconnects the endplate, the gerotor set, the manifold assembly and the housing member.




Another feature of the rotary pressure device is that the first and second plurality of laterally directed fluid paths are substantially axially directed, and that the branch conduits are substantially radially directed. A further feature includes having the first plurality of laterally directed fluid paths in the rotor being located in the rotor between externally toothed members thereof, and having the first plurality of laterally directed fluid paths being substantially laterally directed between the rotor first and second axial ends. Further the first plurality of laterally directed fluid paths could be substantially circumferentially centered between adjacent ones of the rotor externally toothed member thereof.




A further feature of the rotary pressure device is that the second plurality of laterally directed fluid paths is circumferentially centered between the first plurality of laterally directed fluid paths. Also the second plurality of laterally directed fluid paths in the rotor is substantially laterally directed between the rotor first and second axial ends, and wherein the second plurality of laterally directed fluid paths in the rotor is substantially radially aligned with adjacent ones of the rotor externally toothed members. Another feature is wherein the pluralities of the first and second laterally-directed fluid paths are substantially parallel.




Still another feature includes having one of the first and second axial end faces on the rotor having a first plurality of circumferentially spaced recesses located thereon, each of the first plurality of recesses in fluid communication with the first plurality of laterally directed fluid paths. Further, the first plurality of circumferentially spaced recesses can receive fluid for reducing the viscous friction between one of the first and second axial end faces and the end plate. Also, another feature is to minimize the number of circumferentially spaced recesses that do not receive a flowing fluid.




An additional feature of the present invention includes having one of the first and second axial end faces on the rotor having a second plurality of circumferentially spaced recesses located thereon, and each of the second plurality of recesses being in fluid communication with the second plurality of laterally directed fluid paths. Further the second plurality of circumferentially spaced recesses receive fluid for reducing the viscous friction between one of the first and second axial end faces and the end plate. Also, another feature is to minimize the number of circumferentially spaced recesses that do not receive a flowing fluid.




Yet another feature of the present invention includes having the plurality of laterally directed first and second fluid paths in the rotor extend through the rotor from the first axial end surface to the second axial end surface. An added feature of the present invention includes having the rotary fluid pressure device function as one of a hydraulic pump and motor.











Still another feature of the present invention involves having some of the housing member first and second flow passage, the manifold assembly first and second fluid passage as well as the pluralities of the rotor first and second laterally directed fluid paths being utilized for both high pressure and exhaust fluid passage. Also the housing member first flow passage and the manifold assembly first fluid passage could be conduits for high pressure fluid, and the housing member second flow passage and the manifold assembly second fluid passage could be conduits for exhaust fluid. Further features of the present invention will become apparent to those skilled in the art upon reviewing the following specification and attached drawings. Further features of the present invention will become apparent to those skilled in the art upon reviewing the following specification and attached drawings.




BRIEF DESCRIPTION OF THE DRAWINGS





FIG. 1

is a perspective view of a hydraulic motor according to the present invention.





FIG. 2

is a sectional view of the hydraulic motor.





FIG. 3



a


is a cross-sectional view of a gerotor, a component of the hydraulic motor, shown from a first axial end.





FIG. 3



b


is a cross-sectional view of the gerotor, similar to

FIG. 3



a


, but shown from the opposite axial end.





FIG. 4



a


is an elevational view of the rotor, as viewed from a first axial end.





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.







FIG. 5



a


is a frontal view of a manifold plate adjacent the shaft housing of the hydraulic motor.





FIG. 5



b


is a frontal view of the middle manifold plate.





FIG. 5



c


is a frontal view of a manifold plate adjacent the gerotor.





FIG. 6



a


is an end view showing the rotor relative to the stator at 0°.





FIG. 6



a


′ shows

FIG. 6

together with the manifold plate.





FIG. 6



b


is an end view showing the rotor relative to the stator at 18° counterclockwise.





FIG. 6



b


′ shows the rotor relative to the adjacent manifold plate at 18° counterclockwise.





FIG. 6



c


is an end view showing the rotor relative to the stator at 36° counterclockwise.





FIG. 6



c


′ shows the rotor relative to the adjacent manifold plate at 36° counterclockwise.





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


.





FIG. 7



b


is a sectional view of the flexible balancing plate taken along line G—G of

FIG. 7



a.







FIG. 7



c


is a rear view of the channeling plate taken along line


7




c





7




c


in FIG.


2


.





FIG. 8



a


is a rear view of an end cover of the present invention.





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.







FIG. 8



c


is a frontal view of the alternate embodiment of the end cover.





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.





FIG. 10

is a further embodiment of the present invention, showing a sectional view of the hydraulic motor.





FIG. 11

shows a cross-sectional view of a gerotor of the further embodiment, shown from a first axial end.





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




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.




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


.




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


.




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


. Coupling shaft rear portion


27


also includes radial flow passages


19


connecting fluid chamber


26


and fluid chamber


18


.




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.




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.




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


.




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.




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


.





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.




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


.




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


.




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.






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


.




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


.




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


).




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


.




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.




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


.




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.




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.




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.






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


.





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.




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'clock to the 7 o'clock position (or those filled with high pressure fluid


39




a


), are expanding and volume chambers


54


, extending (counter-clockwise) from the 5 o'clock to 12 o'clock position (or those filled with exhaust fluid


38




a


), are contracting. The volume chamber at the 6 o'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'clock position, rotor axial passage


48


is intermediate the high-pressure fluid


39




a


and exhaust fluid


38




a


manifold openings.




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'clock to the 11 o'clock position. Volume chambers


54


which are contracting are located (counter-clockwise) from the 11 o'clock to the 6 o'clock position. Volume chamber


54


located at the 5 o'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.






In

FIGS. 6



c


and


6




c


′ rotor


45


has rotated counter-clockwise 36° within stator


41


. Volume chambers


54


from the 10 o'clock to the 6 o'clock position (counter-clockwise) are expanding and volume chambers


54


from the 4 o'clock to the 11 o'clock position (counter-clockwise) are contracting. Volume chamber


54


located at the 5 o'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.






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'clock will now be discussed. Referring to

FIG. 6



a


, volume chamber


54


(at 11 o'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'clock, in axial alignment with opening


34


filled with pressurized fluid


39




a


, and 10 o'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'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.






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.




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'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).




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.




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.




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


.




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.




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.




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.




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 straight, or through, 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


.




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.




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


.




The principles, preferred embodiments and modes of operation of the present invention have been described in the foregoing specification. The invention which is intended to be protected herein should not, however, be construed as limited to the particular form described as it is to be regarded as illustrative rather than restrictive. Variations and changes may be made by those skilled in the art without departing from the scope and spirit of the invention as set forth in the appended claims.



Claims
  • 1. A rotary fluid pressure device comprising:a housing member defining a fluid inlet port, a fluid outlet port, a first flow passage, a second flow passage and an internal bore; a manifold assembly having a first fluid passage, a second fluid passage, and an internal bore, one side of said manifold assembly adjoining said housing member; a gerotor set having an internally toothed stator member; and an externally toothed rotor member, disposed within said stator member, said rotor member having an internal bore and a first and a second axial end surfaces, one of said stator and said rotor members having orbital movement relative to the other said member, said rotor member having a rotational movement relative to said stator, with the internal teeth of said stator member and the external teeth of said rotor member interengaging to define a plurality of expanding and contracting volume chambers; a first plurality of circumferentially spaced laterally directed fluid paths in said rotor extending through said rotor for fluid connection with said manifold assembly first and second fluid passages, a branch conduit for each of said first plurality of fluid paths adapted for directly connecting respective ones of said first plurality of laterally-directed fluid paths in said rotor to said volume chambers, with one side of said gerotor set adjoining another side of said manifold assembly; a second plurality of circumferentially spaced, laterally-directed fluid paths extending through said rotor, said second plurality of fluid paths being circumferentially interposed between said first plurality of fluid paths for sequentially channeling fluid between one of said first and second axial end faces; an end plate, adjoining another side of said gerotor set; a rotatably journaled torque transfer shaft operatively interconnected to said rotor and extending from within said housing member; and a plurality of coupling members for interconnecting said endplate, said gerotor set, said manifold assembly, and said housing member.
  • 2. The rotary pressure device as in claim 1 wherein said first and second plurality of laterally directed fluid paths are substantially axially directed.
  • 3. The rotary pressure device as in claim 1 wherein said branch conduits are substantially radially directed.
  • 4. The rotary pressure device as in claim 1 wherein said first plurality of laterally directed fluid paths in said rotor is located in said rotor between externally toothed members thereof.
  • 5. The rotary pressure device as in claim 4 wherein said first plurality of laterally directed fluid paths in said rotor is substantially laterally directed between said rotor first and second axial ends.
  • 6. The rotary pressure device as in claim 4 wherein said first plurality of laterally directed fluid paths in said rotor is substantially circumferentially centered between adjacent ones of said rotor externally toothed members thereof.
  • 7. The rotary pressure device as in claim 4 wherein said first plurality of laterally-directed fluid paths in said rotor is substantially laterally centered between said rotor first and second axial ends and are substantially circumferentially centered between adjacent ones of said rotor externally toothed members thereof.
  • 8. The rotary pressure device as in claim 4 wherein said plurality of laterally-directed fluid paths in said rotor is at least one of substantially laterally directed between said rotor first and second axial ends, and are substantially circumferentially centered between adjacent ones of said rotor externally toothed members thereof.
  • 9. The rotary pressure device as in claim 1 wherein said plurality of laterally directed pluralities of first and second fluid paths are circumferentially centered between adjacent ones of externally toothed members thereof.
  • 10. The rotary pressure device as in claim 1 wherein said second plurality of laterally directed fluid paths is circumferentially centered between said first plurality of laterally directed fluid paths.
  • 11. The rotary pressure device as in claim 1 wherein said second plurality of laterally directed fluid paths in said rotor is substantially laterally directed between said rotor first and second axial ends.
  • 12. The rotary pressure device as in claim 1 wherein said second plurality of laterally directed fluid paths in said rotor is substantially radially aligned with adjacent ones of said rotor externally toothed members thereof.
  • 13. The rotary pressure device as in claim 1 wherein said pluralities of first and second laterally-directed fluid paths are substantially parallel.
  • 14. The rotary pressure device as in claim 1 wherein the laterally-directed center lines of said plurality of first and second fluid paths are substantially located on a common circumferentially-directed circle.
  • 15. The rotary pressure device as in claim 1 wherein one of said first and second axial end faces on said rotor has a first plurality of circumferentially spaced recesses located therein, each of said first plurality of recesses in fluid communication with said first plurality of laterally directed fluid paths.
  • 16. The rotary pressure device as in claim 15 wherein said first plurality of circumferentially spaced recesses functions to reduce the surface area of one of said end faces thereby reducing the viscous friction between said one of said end faces and said end plate.
  • 17. The rotary pressure device as in claim 15 wherein the efficiency of said rotor is increased by increasing the surface area of said recesses thus increasing the area of said one of said end faces that receives a flowing fluid.
  • 18. The rotary pressure device as in claim 1 wherein one of said first and second axial end faces on said rotor has a second plurality of circumferentially spaced recesses located thereon, each of said second plurality of recesses in fluid communication with said second plurality of laterally-directed fluid paths.
  • 19. The rotary pressure device as in claim 18 wherein said second plurality of circumferentially spaced recesses receive fluid for reducing the viscous friction between said one of said first and second axial end faces and said end plate.
  • 20. The rotary pressure device as in claim 18 wherein the majority of said recesses receive one of high pressure fluid and exhaust fluid, thus reducing the recesses not receiving one of high pressure fluid and exhaust fluid.
  • 21. The rotary pressure device as in claim 1 wherein said plurality of laterally-directed first and second fluid paths in said rotor extend through said rotor from said first axial end surface to said second axial end surface.
  • 22. The rotary pressure device as in claim 1 wherein said device functions as one of a hydraulic pump and motor.
  • 23. The rotary pressure device as in claim 1 wherein some of said housing member first flow passage, said housing member second flow passage, said manifold assembly first fluid passage, said manifold assembly second fluid passage as well as said pluralities of rotor first and second laterally-directed fluid paths are utilized for both high pressure and exhaust fluid passage.
  • 24. The rotary pressure device as in claim 1 wherein said housing member first flow passage and said manifold assembly first fluid passage are conduits for high pressure fluid, and said housing member second flow passage and said manifold assembly second fluid passage are conduits for exhaust fluid.
  • 25. The rotary pressure device as in claim 1 wherein said housing member second flow passage and said manifold assembly second fluid passage are conduits for high pressure fluid, and said housing member first flow passage and said manifold assembly first fluid passage are conduits for exhaust fluid.
  • 26. A gerotor hydraulic pressure device for use in one of a hydraulic motor and pump having an internally toothed stator member; an externally toothed rotor member, eccentrically disposed within said stator member, and having an internal bore and first and second axial end surfaces; one of said stator and said rotor members having an orbital movement relative to the other said member, said rotor member having a rotational movement relative to said stator, the internal teeth of said stator member and the external teeth of said rotor member interengaging to define a plurality of expanding and contracting volume chambers, a first plurality of laterally-directed fluid paths extending through said rotor; and a plurality of radiating fluid paths in said rotor, each radiating fluid path being connected to both one of said plurality of laterally-directed fluid paths and one of said plurality of volume chambers; a second plurality of laterally-directed fluid paths extending through said rotor, said second plurality of fluid paths being circumferentially interposed between said first plurality of fluid paths for sequentially channeling fluid between one of said first and second axial end faces.
  • 27. In a gerotor hydraulic pressure device for use in one of a hydraulic pump and motor application including:a. an internally toothed stator member; b. an externally toothed rotor member, eccentrically disposed within said stator member, having an internal bore and first and second axial end surfaces, with the external teeth thereof being separated by equally circumferentially spaced connecting portions; and c. one of said stator and rotor members having an orbital movement relative to the other said member and said rotor member having at least a rotational movement relative to said stator, with the internal teeth of said stator member and the corresponding external teeth of said rotor member interengaging to define a plurality of repeating expanding and contracting volume chambers, wherein the improvement comprises: i. a first plurality of substantially laterally-directed fluid paths in said rotor; a plurality of radiating fluid branches in said rotor, each said radiating fluid branches being connected to both respective ones of said first plurality of laterally-directed fluid paths and one of said plurality of volume chambers; ii. a second plurality of substantially laterally-directed fluid paths extending through said rotor, said second plurality of fluid paths being circumferentially interposed between said first plurality of fluid paths for sequentially channeling fluid between one of said first and second axial end surfaces.
CROSS-REFERENCE TO RELATED CASES

The present application claims the benefit of the filing date of U.S. Provisional Application Serial No. 60/410,680 filed Sep. 13, 2002.

US Referenced Citations (6)
Number Name Date Kind
2989951 Charlson Jun 1961 A
3233524 Charlson Feb 1966 A
4264288 Wiisthof et al. Apr 1981 A
4411606 Miller Oct 1983 A
4717320 White, Jr. Jan 1988 A
5624248 Kassen et al. Apr 1997 A
Provisional Applications (1)
Number Date Country
60/410680 Sep 2002 US