Patent Publication Number: US-6699024-B2

Title: Hydraulic motor

Description:
RELATED APPLICATION 
     This application claims priority under 35 U.S.C. §119(e) to U.S. Provisional Patent Application No. 60/302,257 filed on Jun. 29, 2001. The entire disclosure of this provisional application is hereby incorporated by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates generally as indicated to a hydraulic motor and, more particularly, to a hydraulic motor with a gerotor drive assembly which provides rotational motion to a desired piece of machinery. 
     BACKGROUND OF THE INVENTION 
     A hydraulic motor is a converter of pressurized oil flow into torque and speed for transferring rotational motion to a desired piece of machinery. Of particular relevance to the present invention is a hydraulic motor, wherein this conversion is accomplished by a drive assembly having a gerotor set. A gerotor motor can provide a combination of compact size, low manufacturing cost, and high torque capacity, thereby making it a very popular choice for heavy duty applications requiring low speeds (e.g., 1000 rpm or less) and high torques (e.g., 15,000 In-Lb or more). 
     A gerotor set comprises an outer stator and an inner rotor having different centers with a fixed eccentricity. The stator has internal teeth or “vanes” which form circular arcs, and the inner rotor has one less external “teeth” or lobes. The rotor lobes remain in contact with the circular arcs as the rotor moves relative to the stator, and these continuous multi-location contacts create fluid pockets which sequentially expand and contract. As fluid is supplied and exhausted from the fluid pockets in a timed relationship, the rotor moves hypocycloidally (i.e., orbits and rotates) relative to the stator. 
     A drive link is interconnected to the rotor for movement therewith, and this interconnection usually constitutes crowned external splines on the drive link which engage with internal splines on the rotor. Such a splined mating arrangement allows the drive link to “wobble” during operation of the motor. To prevent the drive link from slipping axially backward out of the splined engagement, an axial stop can be provided adjacent the rear end (or nose portion) of the drive link. 
     The drive assembly of a gerotor motor will typically include a valving system to supply and exhaust the fluid from the gerotor pockets in the desired timed relationship. One common type of valving system includes a disk-type commutator and a stationary valve member (e.g., a manifold). A slow-speed commutator rotates at the speed of rotation of the rotor, and manifold channels are opened/closed in the angular circumferential direction using edges of the valve openings. A fast-speed commutator orbits with the rotor and the commutator&#39;s inner diameter and outer diameter control fluid metering. Generally, a fast-speed commutator is preferred because it allows valving to be synchronized with the volume changes of the gerotor fluid pockets (rather than rotation of the shaft), thereby significantly reducing timing errors. 
     The use of a commutator creates the potential for cross-port leakage (e.g., flow bypasses the drive assembly) at the interface between the commutator and an end cover. To prevent such cross-port leakage, a groove can be formed in the back axial face of the commutator and a triangular or trapezoidal (in cross-section) sealing ring positioned therein. The sealing ring is usually oversized (e.g., the height of the ring is greater than the depth of the groove) so that, when the motor is at rest, the ring projects outwardly from the groove. Upon start-up of the motor, the hydraulic imbalance pushes the sealing ring out of the groove to perform the sealing at the interface between the commutator and end cover. 
     The drive link is interconnected to a shaft to transfer rotational movement thereto. For example, the motor can include a coupling shaft which is connected to the drive link (e.g., by a splined interconnection) and which can be coupled to the input shaft of the desired piece of machinery. In this case, the drive assembly (e.g., the commutator, the manifold and the gerotor set) is commonly positioned between the motor&#39;s end cover and a housing which rotatably supports the coupling shaft. Alternatively, the shaft can be part of the gearbox of the desired machinery and the drive link is directly coupled thereto. In this case, the drive assembly is commonly positioned between the motor&#39;s end cover and a mountable housing for attachment to the gearbox. In either case, a plurality of bolts extend through registered openings in the end cover, the drive assembly and the housing to clamp these components together. A wear plate can be positioned between the drive assembly and the housing, and the clamping bolts can also extend therethrough. Face seals are provided between the various components to prevent leakage at the interfaces. 
     A hydraulic motor will have a flow circuit which determines the path of fluid flow and can be viewed as defining a cylindrical pressure vessel. The diameter of the pressure vessel is determined by the outermost radial reach of the fluid circuit, and the length of the pressure vessel is determined by the longest axial reach of the fluid circuit. 
     The flow circuit of a hydraulic motor includes a working path which extends between the inlet port and the outlet port and through which the fluid passes to cause the drive assembly to rotate the output shaft in the appropriate direction. When the motor is operating in a first direction, the first port is the inlet port and the second port is the outlet port and the output shaft rotates in a first direction (e.g., clockwise). When the motor is operating in a second direction, the second port is the inlet port and the first port is the outlet port and the output shaft rotates in a second direction (e.g., counterclockwise). In either case, the inlet port can be connected to a pump discharge and the outlet port can be connected to a return line to a reservoir which feeds the pump suction. 
     In most hydraulic motor designs, the working path extends through non-working portions of the motor (e.g., the housing and/or an axial passageway in the drive link), whereby the length of the working path extends for a substantial distance of the pressure vessel. Also, most hydraulic motors have a “wet bolt” design, wherein the clamping-bolt openings double as fluid passageways and face seals are located radially outside the diameter of the circular array of clamping bolts. This arrangement results in the diameter of the pressure vessel occupying a substantial portion of the motor&#39;s radial dimension, and requires the clamping bolts to directly absorb corresponding forces. 
     The flow circuit of a hydraulic motor will usually also include a non-working path, including chambers surrounding the drive train components (i.e., the drive link and the coupling shaft) and through which fluid passes for cooling and lubrication of these components. In a two-pressure-zone motor design, fluid traveling through the non-working path rejoins fluid traveling through the working path somewhere upstream of the outlet port. In a three-pressure-zone motor design, fluid traveling through the non-working path does not rejoin the working path and exits the motor through a separate case drain in the housing. 
     A three-pressure-zone motor design is used in applications where contamination flushing must be performed. Additionally or alternatively, a three-pressure-zone design is used for applications in which the drive link is coupled directly to the input shaft of a gearbox. Otherwise, a two-pressure-zone motor design usually is employed because it simplifies plumbing criteria, reduces reservoir size requirements, decreases pump capacity demands, and minimizes the risk of “dead zones” within the motor. 
     Some of the most significant considerations when selecting a hydraulic motor, especially for heavy-duty applications, include the motor&#39;s no-load pressure drop (or mechanical efficiency), its life expectancy, its start-up (or breakaway) efficiency, and/or its torque capacity. Accordingly, motor manufacturers are constantly trying to improve upon these performance parameters. 
     SUMMARY OF THE INVENTION 
     The present invention provides a hydraulic motor which, when compared to conventional hydraulic motors, can be constructed to have an improved no-load pressure drop, a longer life expectancy, a better start-up efficiency and/or a higher torque capacity. The motor can be especially well suited for heavy-duty applications requiring low speeds and high torques. 
     More particularly, the present invention provides a hydraulic motor comprising an end cover, a drive link, a drive assembly, and a flow circuit extending between a first port and a second port. The flow circuit comprises a working path through which fluid flows to cause the drive assembly to hypocycloidally move the drive link in a first direction when the first port is the inlet port and in a second direction when the second port is the inlet port. When the motor is operating in a first direction, the fluid flows in a first direction through the working path of the fluid circuit and, when the motor is operating in a second direction, the fluid flows in a second direction through the working path of the fluid circuit. The motor can be designed to operate in only one direction (either the first or the second) or can be designed to operate in both directions. The flow circuit can also comprise a non-working path passing through chambers surrounding the drive link to cool and lubricate the drive train components. 
     According to one aspect of the invention, the first port and the second port are part of the end cover, and the working path is axially confined to a length between the end cover and the drive assembly. As such, the working fluid is not subjected to no-load pressure drops from unnecessary travel through non-working portions of the motor. This confinement of the working path results in a significantly reduced pressure drop (e.g., 50% less) when compared to conventional hydraulic motors of similar size and/or capacity and this translates into a dramatic improvement in motor efficiency. 
     According to another aspect of the invention, the clamping bolts are radially positioned outside of the motor&#39;s pressure vessel and, in any event, they do not communicate with any of the motor&#39;s fluid chambers. This radially outward positioning of the clamping bolts, or “dry bolt” design, results in less axial tensile stress per bolt for a motor design having a given number of clamping bolts. Additionally or alternatively, because fluid flow characteristics do not play a part in bolt placement, more clamping bolts can be used in a given motor design. Less strain-per-bolt and/or more bolts-per-motor result in less bolt-stretching and equal bi-directional motor performance which, in turn, results in a longer motor life. Furthermore, this “dry bolt” design avoids the extra manufacturing cost of countersink machining which is required in a “wet bolt” design. 
     According to another aspect of the invention, a non-interference seal arrangement is used at the valving interface between the end cover and the drive assembly. In this arrangement, a sealing ring is positioned in a groove in the commutator. The height of the sealing ring is less than the depth of the groove, whereby the seal does not project outwardly from the groove when the motor is at rest. Also, the groove and seal can each have a roughly rectangular cross-sectional shape such that the ring resides loosely within the groove when the motor is at rest and then, upon start-up of the motor, is appropriately moved to a position which prevents cross-port leakage. Specifically, the seal is pushed rearward by hydraulic imbalance forces and is pushed in the appropriate radial direction by the port-to-port pressure differential. With an oversized seal, mechanical friction is created between the seal and the end cover during startup or very slow speed operation (e.g., 10 rpm or less). With the sealing arrangement of the present invention, this mechanical friction is eliminated thereby enhancing start-up and low speed efficiency and increasing the life of the sealing ring. 
     According to a further aspect of the invention, an axial stop for the drive link is mounted on a moving part of the drive assembly and, more particularly, is preassembled on an internal diameter of the rotor. When the axial stop is mounted on a stationary component of the motor (e.g., the end cover), the drive link will rotate/orbit relative to the axial stop, thereby creating internal mechanical friction therebetween. However, with the axial stop system of the present invention, this internal friction is eliminated, thereby improving the motor&#39;s startup efficiency. 
     According to a further aspect of the invention, the drive link has an axial passageway which allows a component of the drive train (e.g., a coupling shaft) to centrifugally pump a diverted portion of fluid from the working path through the non-working path. Regardless of whether the motor is operating in the first direction or the second direction, the diverted portion of the fluid is centrifugally pumped through the non-working path in the same direction by the output shaft. When the motor is operating in the first direction, the non-working portion of the fluid is diverted from the high pressure (pre-working) fluid and, when the motor is operating in the second direction, the non-working portion of the fluid is diverted from the low pressure (post-worked) fluid. This non-working path is believed to provide superior lubrication for the splined interconnection between the drive link and the rotor and/or the splined interconnection between the drive link and the output shaft. Since, in general, the torque capacity of a motor is limited by the condition of its drive train components, this superior lubrication arrangement can greatly enhance the performance of a motor. This aspect of the invention finds particular application in two-pressure-zone motor designs but can also be used in three-pressure-zone motor designs as well. 
     These and other features of the invention are fully described and particularly pointed out in the claims. The following description and drawings set forth in detail certain illustrative embodiments of the invention, these embodiments being indicative of but a few of the various ways in which the principles of the invention may be employed. 
    
    
     DRAWINGS 
     FIG. 1 is a perspective view of a hydraulic motor  10  according to the present invention. 
     FIG. 2 is an end view of the hydraulic motor  10 . 
     FIG. 3 is a sectional view of the hydraulic motor  10 . 
     FIGS. 4A-4C are close-up sectional views of a commutator sealing arrangement. 
     FIG. 5 is a close-up sectional view of a portion of the motor  10  showing an axial stop for limiting linear movement of a drive link. 
     FIGS. 6A and 6B are schematic illustrations of the fluid circuit of the motor  10  when it is operating in a first direction and a second direction, respectively. 
     FIG. 7 is a sectional elevational view of another motor  110  according to the present invention. 
     FIG. 8 is a close-up sectional view of a portion of the motor  110  showing a commutator end cap and a passageway formed therein. 
     FIGS. 9A and 9B are schematic illustrations of the fluid circuit of the motor  110  when it is operating in a first direction and a second direction, respectively. 
     FIG. 10 is sectional elevation view of another motor  210  according to the present invention. 
     FIG. 11 is a schematic illustration of the fluid circuit of the motor  210  when it is operating in one direction. 
    
    
     DETAILED DESCRIPTION 
     Referring now to the drawings, and initially to FIGS. 1-3, a hydraulic motor  10  according to the present invention is shown. The illustrated hydraulic motor  10  is especially designed for heavy duty applications requiring low speeds and high torques. As is explained in more detail below, the motor  10  can be constructed to have an improved no-load pressure drop, a longer life expectancy, a better start-up efficiency and/or a higher torque capacity. 
     The motor  10  comprises an end cover  12  defining a first port  14  and a second port  16 , a drive assembly  18 , a shaft housing  20 , a drive link  22  and a coupling shaft  24 . (FIGS. 1 and 3.) In the illustrated embodiment, the end cover  12  is a separate component which functions as a rear lid for the motor  10 . However, end covers integral with other components of the motor  10  and/or end covers which do not necessary perform as rear lids are possible with, and contemplated by, the present invention. 
     A plurality of bolts  26  (e.g, nine bolts in a circular array) extend through registered openings in the end cover  12 , the drive assembly  18  and the shaft housing  20  to clamp these components together. (FIGS. 2 and 3.) In the illustrated embodiment, the motor  10  also includes a wear plate  28  positioned between the drive assembly  18  and the shaft housing  20  and the clamping bolts  26  also extend therethrough. (FIGS. 1 and 3.) Face seals  30  are provided between the end cover  12  and the drive assembly  18 , between two components of the drive assembly  18  (namely a manifold  34  and a rotor set  36 , introduced below), between the drive assembly  18  and the wear plate  28 , and between the wear plate  28  and the shaft housing  20 . (FIG. 3.) 
     When the motor  10  is operating in a first direction (e.g., the coupling shaft  24  rotates clockwise), the first port  14  is the inlet port and the second port  16  is the outlet port. When the motor  10  is operating in a second opposite direction (e.g., the coupling shaft  24  rotates counterclockwise), the second port  16  is the inlet port and the first port  14  is the outlet port. In either case, the inlet port can be connected to a pump discharge and the outlet port can be connected to a return line to a reservoir which feeds the pump suction. In response to pressurized fluid passing from the inlet port to the outlet port through a working fluid path, the drive assembly  18  hypocycloidally moves (i.e., orbits and rotates) the drive link  22  and the coupling shaft  24  rotates in a corresponding direction. The motor  10  does not include a case drain whereby it has a two pressure zone design. 
     The drive assembly  18  comprises a commutator  32 , a manifold  34 , and a gerotor set  36 . The commutator  32  is positioned in a space between the end cover  12  and the manifold  34  for movement with the drive link  22  during operation of the motor  10 . Accordingly, the illustrated commutator  32  is a fast-speed commutator which orbits at the orbiting speed of the moving member of the gerotor set  36  (namely its rotor  52 , introduced below). 
     The commutator  32  comprises an inner ring  38 , an outer ring  40 , and spoke-like members extending between the rings so that the commutator&#39;s inner diameter and outer diameter can control fluid metering. The inner ring  38  captures a portion of the drive link  22  (namely its nose portion  66  introduced below). The outer ring  40  divides the space between the end cover  12  and the manifold  34  into a first chamber  42  which communicates with the first port  14  and a second chamber  44  which communicates with the second port  16 . 
     As can best be seen by referring additionally to FIGS. 4A-4C, the axial face of the outer commutator ring  40  adjacent the end cover  12  includes a groove  46  which houses a sealing ring  48 . The sealing ring  48  can be made of a polyimide resin, such as VESPEL® which is a trademark of DuPont for a temperature-resistant thermosetting polyimide resin. In any event, the depth of the groove  46  is greater than the height of the sealing ring  48  whereby there will be no mechanical friction between the seal  48  and the end cover  12  at very low speed operation of the motor  10  as is found, for example, with an oversized commutator seal. This elimination of internal friction enhances the starting efficiency of the motor  10  and increases the life of the sealing ring  48 . 
     The groove  46  and the sealing ring  48  each have substantially rectangular cross-sectional shape and the width of the groove  46  is also greater than the width of the sealing ring  48 . When the motor  10  is at rest (i.e., not operating), the sealing ring  48  resides loosely within the groove  46 . (FIG. 4A.) However, when the motor  10  is operating in the first direction, and high pressure fluid is introduced into the first chamber  42 , the high pressure fluid presses the radially outer side of the sealing ring  48  against the radially outer side of the groove  46 . Also, the imbalance between the hydraulic forces on the rear and the front of the sealing ring  48  cause it to be pushed axially rearward towards the end cover  12 . (FIG. 4B.) Likewise, when the motor  10  is operating in the second direction, and high pressure fluid is introduced into the second chamber  44 , the high pressure fluid presses the radially inward side of the sealing ring  48  against the radially inner side of the groove  46 . Again, the imbalance between the hydraulic forces on the rear and the front of the sealing ring  48  cause it to be pushed axially rearward towards the end cover  12  (FIG. 4C.) 
     The manifold  34  has a first set of channels which extend between the first chamber  42  and the gerotor set  36  and a second set of channels which extend between the second chamber  44  and the gerotor set  36 . The number of channels in each set and their circumferential spacing corresponds to the fluid pockets formed by the gerotor set  36  and these channels are systematically opened and closed by the commutator  32  as it is moved with the drive link  22 . In the illustrated embodiment, the manifold  34  is made from a plurality of layers which are laminated together in a certain stacked arrangement to form the flow channels. 
     The gerotor set  36  comprises a stator  50  and a rotor  52  having different centers with a fixed eccentricity. The stator  50  has internal teeth or “vanes” which form circular arcs and the rotor  52  has one less external “teeth” or lobes. As fluid is supplied and exhausted from the fluid pockets in a timed relationship, the rotor  52  moves hypocycloidally (i.e., orbits and rotates) relative to the stator  50 . 
     The illustrated gerotor set  36  is a 8×9 gerotor set, that is, the stator  50  has nine vanes and the rotor  52  has eight teeth, and these components cooperate to form nine fluid pockets. When compared to, for example, a 6×7 gerotor set, the 8×9 gerotor set  36  allows a larger drive link to be assembled inside the rotor  52  thereby providing a higher torque capacity. Also, the 8×9 gerotor set  36  allows a lower eccentricity (e.g., 3 mm) for a desired displacement capacity thereby providing smoother rotation of the rotor  52  and better spline engagement between the drive link  22  and the rotor  52 . That being said, other gerotor designs (e.g., a 6×7 gerotor set) are possible with, and contemplated by, the present invention. 
     The shaft housing  20  has a central bore  54  in which the coupling shaft  24  is rotatably supported. The central bore  54  has portions of varying diameters to accommodate the stepped profile of the coupling shaft  24  as well as radial bearings  56  and thrust bearings  58 . A fluid chamber  60  surrounds the coupling shaft  24  within the bore  54  and a fluid-tight seal  62  is provided to prevent leakage therefrom. A dirt seal  64  can also be provided at the exposed axial end face of the shaft housing  20 . 
     The drive link  22  includes a nose portion  66  captured within the commutator inner ring  38 , an externally splined intermediate portion  68  which mates with internal splines on the rotor  52 , and an externally splined end portion  70  which mates with an internal splines on the coupling shaft  24 . A fluid chamber  72 , in communication with the first chamber  42 , surrounds the drive link  22  as it extends through the manifold  32 , the rotor  52 , the wear plate  28  and into a portion (namely a sleeve portion  84  introduced below) of the coupling shaft  24 . The drive link  22  also includes a passageway  74  extending between its axial ends. 
     As is best seen by referring additionally to FIG. 5, an axial stop member (e.g., a metal washer) is mounted on the rotor  52  adjacent its splined portion and held in position by a snap ring  78 . The axial stop  76  has an annular shape and its inner diameter is greater than the diameter of the nose portion  66  of the drive link  22  but less than the diameter of its splined portion  68 . In this manner, possible axial movement of the drive link  22  towards the end cover  12  is prevented. By mounting the axial stop  76  on a component which moves with the drive link  22 , internal mechanical friction therebetween is minimized as compared to when the axial stop  76  is mounted on the end cover  12 . Accordingly, the use of the inner rotor  52  as an axial stop translates into an enhancement of the motor&#39;s start-up efficiency. Also, since an axial stop does not have to be positioned in the first chamber  42 , flow area within this chamber is optimized thereby further enhancing the no-load pressure drop characteristics (i.e., mechanical efficiency) of the motor  10 . 
     The coupling shaft  24  has a rear portion  82  which projects outwardly from the shaft housing  20  and a wider front sleeve portion  84  which receives the end portion  70  of the drive link  22 . The shaft  24  includes an axial passageway  86  which extends from the internal end face of the sleeve portion  84  to a radial passageway  88  communicating with the shaft-surrounding chamber  60 . The chamber  72  surrounding the drive link  22  extends into the sleeve portion  84  and the shaft  24  has radial passageways  92  which connects the chamber  60  to the chamber  72 . 
     Referring now to FIGS. 6A and 6B, the fluid circuit for the motor  10  is schematically shown when the motor  10  is respectively operating in a first direction (e.g. the shaft  24  rotates clockwise) and in a second direction (e.g., the shaft  24  rotates counterclockwise). In these schematic illustrations, high pressure regions (pre-working) are represented by dark shading and low pressure regions (post-working) are represented by light shading. Also, the working path of the fluid (e.g., the path fluid follows to cause rotation of the coupling shaft  24 ) is represented by solid arrows and the non-working path of the fluid (e.g., the path fluid follows for cooling, lubrication and/or sealing purposes) is represented by dashed arrows. 
     When the motor  10  is operating in the first direction shown in FIG. 6A, high pressure fluid is introduced through the first port  14  into the first chamber  42  and the commutator  32  sequentially directs a primary portion of the high pressure fluid through the first set of flow channels in manifold  34 . The manifold  34  thereby channels the high pressure fluid to the fluid pockets of the gerotor set  36  and the rotor  52  orbits/rotates in a first direction (e.g, clockwise). The now-low-pressure (post-working) fluid then flows through the second set of flow channels in the manifold  34  to the second chamber  44  and exits the motor  10  through the second port  16 . (See solid arrows in FIG. 6A.) 
     When the motor  10  is operating in the first direction, a secondary portion of the high pressure fluid bypasses the working path and travels through the non-working path. Specifically, the secondary portion of the high pressure fluid travels through the axial passageway  74  in the drive link  22  into the axial passageway  86  in the coupling shaft  24 . The rotation of the coupling shaft  24  produces centrifugal forces causing the high pressure fluid to be flung through the shaft&#39;s radial passageway  88  into the chamber  60 . The fluid flows from the chamber  60 , through the radial passageways  92  into the chamber  72 , and back into the first chamber  42  whereat it mixes with the inlet high pressure fluid being introduced through the first port  14 . (See dashed arrows in FIG. 6A.) 
     When the motor  10  is operating in the second direction shown in FIG. 6B, high pressure fluid is introduced through the second port  16  into the second chamber  44 . The commutator  32  sequentially directs all of the high pressure fluid (i.e., none of the high pressure fluid is diverted from the working path) through the second set of flow channels in the manifold  34 . The manifold  34  thereby channels the high pressure fluid to the fluid pockets of the gerotor set  36  thereby causing the rotor  52  to orbit/rotate in a second opposite direction (e.g., counterclockwise). The now-low-pressure (post-working) fluid then flows through the first set of flow channels in the manifold  34  to the first chamber  42  and a primary portion of the low pressure fluid exits the motor  10  through the first port  14 . (See solid arrows in FIG. 6B.) 
     When the motor  10  is operating in the second direction, a secondary portion of the low pressure fluid does not exit the motor through the first port  14  but instead travels through the non-working path. Specifically, the secondary portion of the low pressure fluid travels through the drive link&#39;s axial passageway  74 , into the shaft&#39;s axial passageway  86 , through the shaft&#39;s radial passageway  92 , into the chamber  60 , through the shaft&#39;s radial passageways  92  into the chamber  72 , and back into the first chamber  42  whereat it mixes with the low pressure fluid being exited through the first port  14 . (See dashed arrows in FIG. 6B.) 
     Accordingly, when the motor  10  is operating in a first direction, the fluid flows in a first direction through the working path of the fluid circuit and, when the motor  10  is operating in a second direction, the fluid flows in a second direction through the working path of the fluid circuit. In either case, a portion of the fluid is centrifugally pumped through the non-working path in the same direction by the coupling shaft  24 . When the motor  10  is operating in the first direction, the non-working portion of the fluid is diverted from the high pressure (pre-working) fluid and, when the motor  10  is operating in the second direction, the non-working portion of the fluid is diverted from the low pressure (post-worked) fluid. 
     As is best shown in FIGS. 6A and 6B, that the motor  10  defines a cylindrical pressure vessel having a diameter D and an axial length L. (The diameter D is defined by the outermost radial reach of the fluid circuit and the axial length is defined by the distance between the outermost axial reach of the fluid circuit.) The working portion of this pressure vessel (i.e., the portion occupied by the working path), has an axial length L working  confined to the end cover  12  and the drive assembly  18 . As such, the working fluid avoids the essentially inevitable pressure-dropping resistance it would be subjected to if the fluid traveled through non-working portions of the motor  10 . This confinement of the working path results in a substantially less no-load pressure drop (e.g., 50% less) of the fluid as it travels through the working path than that found in conventional hydraulic motors which translates into a dramatic improvement in motor efficiency. 
     As is best seen by referring back to FIGS. 2 and 3, the clamping bolts  26  are radially positioned outside the diameter D of the motor&#39;s pressure vessel. The bolt-receiving openings do not communicate with any of the motor&#39;s fluid chambers and the face seals  30  (which define the diameter D of the pressure vessel) are located radially inward from the bolts  26 . 
     The “dry-bolt” design of the hydraulic motor  10  results in less strain-per-bolt for a motor design having a given number of clamping bolts. Also, because fluid flow characteristics do not play a part in bolt placement considerations, more clamping bolts  26  can be used in a given motor design thereby additionally or alternatively reducing the strain-per-bolt. As the life of the clamping bolts directly influences the life of the motor, such a strain-per-bolt reduction can make a major contribution towards increasing motor life. Further, the integrity of the clamping bolts during their working life provides consistent performance regardless of whether the motor  10  is being operated in the first or second direction. Moreover, from a manufacturing point of view, this “dry bolt” design avoids the extra manufacturing cost of countersink machining which is necessary in a “wet bolt” design. 
     Referring now to FIG. 7, another hydraulic motor  110  according to the present invention is shown. The motor  110  is similar in many ways to the motor  10  whereby like reference numerals (plus  100 ) are used to designate corresponding parts. It should be noted, however, that the shaft housing  120  includes a case drain  194  extending from the chamber  60  whereby the motor  110  has a three pressure zone design. Also, the drive link  122  does not include an axial passageway (although one could be provided). Further, as is best seen by referring additionally to FIG. 8, the inner commutator ring is replaced with a cap  196 . The cap  196  covers the nose end  166  of the drive link  122  and separates the first chamber  142  from the chamber  172  surrounding the drive link  122 , except for passageways  198  extending therebetween. 
     The fluid circuit for the motor  110  is schematically shown in FIGS. 9A and 9B when the motor  110  is respectively operating in a first direction (e.g. the shaft  124  rotates clockwise) and in a second direction (e.g., the shaft  124  rotates counterclockwise). As in FIGS. 6A and 6B, the high pressure regions are represented by dark shading, the low pressure regions are represented by light shading, the working path is represented by solid arrows and the non-working path is represented by dashed arrows. 
     The working path for the motor  110  is essentially the same as the working path for the motor  10  in the first direction and the second direction. (See solid arrows in FIGS. 9A and 9B.) Also, the working portion of the pressure vessel of the motor  110  has an axial length L working  confined to the end cover  112  and the drive assembly  118 . As with the motor  10 , this confinement of the working portion of the pressure vessel significantly reduces the no-load pressure drop of the motor  110  which translates directly into an increased mechanical efficiency. 
     When the motor  110  is operating in the first direction (the first port  114  is the inlet port), a secondary portion of the high pressure fluid bypasses the working path and travels through the non-working path. (See dashed arrows in FIG. 9A.) When the motor  110  is operating in the second direction (the second port  116  is the inlet port), a secondary portion of the low pressure fluid bypasses the working path and travels through the non-working path. (See dashed arrows in FIG. 9B.) In either case, the non-working fluid travels from the first chamber  142  through a passageway (passageway  198  in FIG. 8) to the chamber  172  surrounding the drive link  122 . Part of the non-working fluid in the chamber  172  flows through the axial passageway  186  in the coupling shaft  124 , through the radial passageway  188  to the chamber  160 . The rest of the working fluid in the chamber  172  flows through the radial passageway  192  in the coupling shaft  124  to the chamber  160 . The non-working fluid in the chamber  160  exits the motor  110  through the case drain  194 . 
     If the diameter of the pressure vessel for the motor  110  is defined by the outermost radial reach of the flow circuit, this would include the case drain  194 . However, the clamping bolts  126  are positioned outside a pressure vessel defined by the working portion of the motor  110  (i.e., D working  and L working ). Moreover, the flow circuit of the motor  110  does not intersect with the registered openings for the clamping members  126  and thus the motor  110  also has a “dry bolt” design with the same associated advantages as found in motor  10 . 
     Referring now to FIG. 10, another hydraulic motor  210  according to the present invention is shown. The motor  210  is similar in many ways to the motor  110  whereby like reference numerals (plus  100 ) are used to designate corresponding parts. It should be noted, however, that in the motor  210 , the drive link  222  is inserted into the gearbox of the mechanism and directly coupled to its input shaft whereby the motor  210  does not have a coupling shaft and/or a shaft housing. Accordingly, the motor  210  does not include the bearings  56 / 156  and  58 / 158  found in motors  10 / 110  whereby the motor  210  can be considered to be “bearingless.” A mounting face housing  220  is provided for attachment to the gearbox and this housing  220  includes a case drain  294  extending from the chamber  272 . Thus, the motor  210  has a three-pressure-zone design. 
     The fluid circuit for the motor  210  is schematically shown in FIG. 11 with the high pressure regions being represented by dark shading, the low pressure regions being represented by light shading, the working path being represented by solid arrows and the non-working path being represented by dashed arrows. Since most gearboxes are not designed to accommodate high pressure lubricating/cooling fluid, the motor  210  is appropriate for unidirectional applications wherein high pressure fluid is introduced through the second port  216 . Specifically, the high pressure fluid is introduced through the second port  216  and travels through the drive assembly  218  and back to the first chamber  242  as low pressure fluid and a primary portion of the low pressure fluid exits the motor through the first port  214 . (See solid arrows.) A secondary portion of the low pressure fluid bypasses the working path and travels through the non-working path, that is it travels from the first chamber  242  through a passageway (see passageway  198  in FIG. 8) to the chamber  272  to the case drain  294 . (See dashed arrows.) 
     The working portion of the pressure vessel of the motor  210  has an axial length L working  confined to the end cover  212  and the drive assembly  218  and, as with the motors  10  and  110 , this confinement significantly reduces no-load pressure drops. Also, the clamping bolts  226  are positioned outside a pressure vessel defined by the working portion of the motor  110  (i.e., D working  and L working ) and the motor&#39;s flow circuit does not intersect with the registered openings for the clamping members  226 . Thus, the motor  210  also has a “dry bolt” design with the same associated advantages as found in motors  10  and  110 . 
     One can now appreciate that a hydraulic motor  10 / 110 / 210  according to the present invention can provide decreased no-load pressure losses, an extended life expectancy, an enhanced start-up efficiency, and/or an increased torque capacity. It should be noted that while the illustrated motor  10  was designed for heavy duty applications requiring low speed and high torque, the principals of the invention can be employed in motors designed for other applications. It should also be noted that while the various aspects of the invention have been described as being incorporated into the same motor design, these aspects could be used separately and/or in different combination in a plurality of motor designs. By way of an example, the valve interface sealing arrangement can be used on a fast-speed commutator (as shown), a slow-speed commutator or, for that matter, in a variety of valve interface settings to prevent friction during start-up and/or very low speed operation. By way of another example, the rotor-mounted axial stop system could be utilized in many other motor designs to limit internal mechanical friction upon engagement of the drive link with the axial stop. By way of a further example, a drive link with an axial passageway could be used in certain three-pressure-zone motor designs. Accordingly, although the invention has been shown and described with respect to certain preferred embodiments, it is obvious that equivalent and obvious alterations and modifications will occur to others skilled in the art upon the reading and understanding of this specification.