Patent Document

BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present disclosure relates generally to thrust and journal bearings, and in particular, to hydrostatic bearings. 
   2. Description of the Related Art 
   Fluid bearings are bearings that operate with a layer of fluid, such as a gas or a liquid, between moving parts. In comparison to conventional bearings such as roller bearings or ball bearings, for example, fluid bearings provide significant reduction in friction and wear. One common type of fluid bearing is the hydrostatic bearing, in which a rotor element is supported by a fluid to rotate relative to a stator element. Typically, the bearing is provided with a fluid supply under pressure to one or more cavities, sometimes referred to as pads, which are commonly formed in a stator element, between the rotor and stator elements. When the total surface force in the cavities balances the downward force of the rotor element, the element lifts off the surface of the stator element, so that the rotor floats on the fluid. This eliminates mechanical contact between the rotor and stator, allowing the rotor to rotate virtually without friction. This condition is referred to herein as full hydrostatic operation. 
   The surface force is a function of the surface area of the pad and the pressure (psi) of the fluid in the pad. If the surface force drops below the balance force, the rotor will make contact with the stator, possibly resulting in damage to one or both surfaces. If fluid supply pressure is greater than a pressure necessary to establish the balance force, the rotor element is separated further and fluid escapes from the pads, while the surface force remains substantially constant, and equal to the balance force. It is common to maintain a slight overpressure of the fluid supply to ensure that there is no contact between the rotor and stator. However, any excess supply pressure results in loss of fluid. Because of the energy cost associated with pressurizing the fluid in the first place, this loss of fluid represents a loss of energy and a reduction in economy, so such losses are minimized wherever possible. 
   Several designs have been proposed for the deployment of hydrostatic bearings in hydraulic machines such as pump/motors. However, because of the limitations of hydrostatic bearings, there are problems associated with such use. In an application where the load on the bearing varies, such as in a variable-angle pump/motor, it is important that the fluid supply pressure be sufficiently high that at maximum load levels, the surface force is adequate to maintain the balance force, to avoid damage to the bearing. However, this means that when load levels drop, a significant overpressure will exist, resulting in loss of fluid. While many of the proposed designs attempt to address this problem, they are, for the most part, impractical or ineffective. 
     FIGS. 1A-1C  show sectional views of a portion of a bent-axis pump/motor  100  according to known art. The motor  100  includes a valve plate  102  and a cylinder barrel  104 , having a plurality of cylinders  106  within which pistons  108  travel reciprocally. Each of the pistons  108  engages a respective socket formed in a drive plate  110 . The drive plate  110  is coupled to an output shaft  120  that is rotationally driven by the motor  100 . The drive plate  110  bears against a thrust bearing  118  configured to permit free rotation of the drive plate  110  and shaft  120 , while holding the drive plate in position against radial and axial forces acting thereon. A radial bearing  119  is positioned on the shaft  120  to stabilize the shaft while permitting free rotation. The bearing  118  is shown as a combination bearing, configured to bear radial and axial loads. Many motors employ separate axial and radial load bearings. 
   The cylinder barrel  104  is configured to rotate around a first axis A. The drive plate  110  rotates around an axis B, and is coupled to the rotating cylinder barrel  104  by a constant velocity joint  116  (only portions of which are shown in  FIGS. 1A-1C ). Accordingly, the cylinder barrel  104  and the drive plate  110  rotate at a common rate. 
   The valve plate  102 , barrel  104 , and pistons  108 , which define axis A, are configured to rotate with respect to the drive plate  110 , which defines axis B, for the purpose of varying the displacement volume of the pump/motor  100 . The degree of rotation of axis A away from a coaxial relationship with axis B is typically referred to as the stroke-angle of the device. 
   When the motor  100  is operating in a motor mode, high-pressure fluid is valved into each cylinder  106  as it passes top-dead-center (TDC). The high-pressure fluid applies a driving force on the face of the piston  108 , which acts axially on the piston  108  with respect to axis A. This force is transferred by the piston  108  to the drive plate  110 . As each piston  108  passes bottom-dead-center (BDC), the fluid is vented from the piston  106 , which allows the piston to be pushed back into the cylinder as the barrel rotates it back toward TDC. 
   Referring to  FIG. 1A , it may be seen that the driving force on the pistons  108  is axial, relative to axis A, but includes both axial and radial force components, relative to axis B. The distribution of the driving force between the axial and radial components depends on the stroke angle of the motor  100 . The axial component tends to drive the drive plate  110  away from the barrel  104  along axis B, which is prevented by the thrust bearing  118 . The radial component of the driving force tends to drive the socket of the drive plate  110 , into which the second end of the piston  108  is seated, to move downward, causing the drive plate  110  to rotate so that the socket moves further away from the barrel, with the barrel  104  rotating in unison with the drive plate  110 . 
   It will be recognized that the lower the stroke angle, the more of the driving force will be distributed to the drive plate  110  as an axial force, until, at a zero stroke angle such as that shown in  FIG. 1C , all of the drive force is distributed to the drive plate  110  as an axial force. On the other hand, when the motor  100  is at a high stroke angle such as that shown in  FIG. 1A , more of the drive force will be distributed radially and will be experienced by the bearing  118  as a radial force. Moreover, because the drive force is in a downward direction, as viewed in the figures, all of that radial force will be experienced by the lower part of the bearing  118 . At the same time, the drive plate  110  and shaft  120  act as a lever, against the bearing  118  as a fulcrum, such that an upward radial force is exerted on the axial bearing  119 . 
   When the motor is at zero stroke angle, as shown in  FIG. 1C , cylinders  106  on one side of the barrel  104 , divided down the line defined by TDC and BDC, are at high-pressure, while those on the opposite side are at low pressure. Thus, the thrust bearing experiences a very high axial load on one side, and a much lower axial load on the other. These high and low sides are separated by 90° from the high and low sides with respect to radial distribution. Furthermore, if the pressure of the fluid circuit that drives the motor is reversed while the motor is rotating forward, the motor switches to pump mode, and the distribution of the axial load is reversed, so that the bearing  118  experiences the high axial load on the opposite side. 
   The motor  100  shown in  FIGS. 1A-1C  is depicted as having cylinders directly opposite one another such that when one cylinder  106  is at TDC, another will be at BDC. This arrangement is pictured to provide a view of cylinders  106  at both TDC and BDC in the same figure. However, in practice, most hydraulic motors employ an odd number of cylinders, typically seven or nine. As a result, in a nine-cylinder motor the number of cylinders that are pressurized at high-pressure will cycle back and forth between four and five cylinders, nine times for each revolution of the cylinder. This means that the axial and radial loads on the motor bearings will also drop by 20% each time there are four pressurized cylinders, then back up by the same amount when there are five pressurized cylinders. 
   In typical applications, pump/motors of the type described here experience frequent changes in direction and speed. While it has been thought desirable to employ fluid bearings with pump/motors of this kind in order to improve efficiency and reduce wear, it has been found problematic, due to the complex nature of the changes in force and vector at play in these systems. 
   It can be seen that the bearings of the motor  100  are subjected to widely ranging forces. Changes from high to low stroke angle, then back to high, can occur very fast and very frequently. Rotation speed and direction varies, and the motor may stop frequently. Finally, because of the odd-number arrangement of the cylinder barrel, there is a constant 20% fluctuation of force as the barrel rotates. Because of these extreme conditions, little success has been shown using fluid bearings. 
   A more detailed discussion regarding the operation and structure of hydraulic pump/motors may be found in U.S. Pat. No. 7,014,429, issued Mar. 21, 2006, entitled HIGH-EFFICIENCY, LARGE ANGLE, VARIABLE DISPLACEMENT HYDRAULIC PUMP/MOTOR; and U.S. Patent Publication No. 2005/0193888 A1, published Sep. 8, 2005, entitled EFFICIENT PUMP/MOTOR WITH REDUCED ENERGY LOSS, which patent and published patent application are incorporated herein by reference, in their entirety. 
   BRIEF SUMMARY OF THE INVENTION 
   According to an embodiment of the invention, a fluid bearing is provided, comprising an insert configured to be received between first and second elements of a machine in which the second element is adapted to rotate with respect to the first element, hydrostatic pads formed in a surface of the insert and positioned to exert a separating force between the first and second elements, and a bushing between the first and second elements to allow rotation of the second element with respect to the first element while the first and second pads are pressurized at a pressure less than that required to establish a hydrostatic balance force. 
   The surface of the insert may have a cylindrical shape configured to receive a cylindrical second element, or it may be substantially planar to receive an element that is configured to rotate around an axis lying at right angles to the surface of the insert. In the case of the cylindrical insert, the insert may be shaped to encompass less than 360 degrees of the cylindrical shape. 
   Flow of pressurized fluid to the pads of the insert is controlled such that the bearing does not operate in full hydrostatic mode. Instead, a separating force generated by surface force of the fluid is controlled such that the force exerted on the bearing exceeds the separating force, although the fluid pressure is also controlled to selected pads to keep the separating force within a selected margin of the force exerted on the bearing, to control friction and wear of the bearing. 
   According to an embodiment, control of the separating force is achieved by selectively pressurizing individual hydrostatic pads, thereby effectively varying the active hydrostatic area of the bearing. 

   
     BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS 
       FIG. 1A  is a cross-sectional elevational view of components of a hydraulic pump/motor according to known art, at a maximum stroke angle. 
       FIGS. 1B and 1C  show sectional views of the pump/motor of  FIG. 1A  at moderate and zero stroke angles, respectively. 
       FIG. 2  shows a simplified sectional view of a hydraulic pump/motor according to an embodiment of the invention. 
       FIG. 3  shows a plan view of an insert of a fluid bearing according to an embodiment of the invention. 
       FIG. 4  is a graph showing the relationship between the stroke angle of the motor of  FIG. 2  and the axial load on the bearing shown in  FIG. 3 , as a percentage of the maximum axial force of the motor. 
       FIG. 5  shows upper and lower races of a radial bearing of the motor of  FIG. 2 . 
       FIG. 6  is a graph showing the relationship between the stroke angle of the motor of  FIG. 2  and the radial load on the bearing shown in  FIG. 5 , as a percentage of the maximum axial force of the motor. 
       FIG. 7  shows a simplified sectional view of a hydraulic pump/motor according to an alternate embodiment of the invention. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   Various embodiments of the invention will now be described with reference to  FIGS. 2-6 . For the purpose of the disclosure and claims, the term fluid pressure will be used to refer to pressure of a fluid relative to area, such as, for example, psi. The term surface force will be used to refer to hydrostatic force exerted on an opposing surface, which is a function of the fluid pressure multiplied by the total surface area of pressurized fluid in contact with the opposing surface. Separating force refers to the force exerted by the surface force to separate elements. The term balance force will be used to refer to a hydrostatic force at which the surface force exerts a force equal to an opposing force exerted by an opposing surface. Either of the terms “motor” and “pump/motor” may be considered to read on a hydraulic motor, pump, or pump/motor. 
   The term axial force is used herein to refer to force vectors that lie substantially parallel to an axis of rotation of a motor&#39;s drive plate, while radial force is used to refer to force vectors that lie in a plane that is substantially perpendicular to the axis of rotation. Neither term is limited to vectors that intersect the axis. In particular, the radial forces referred to herein may lie in vectors some distance from the defined axis such that a device that is configured to rotate about the axis, and upon which the radial forces act, will tend to rotate in reaction to the forces. 
   The term bushing material is used to refer to a material configured to receive a moving surface against its nonmoving surface without suffering damage due to galling, scoring, etc. Bushings made from such material are well known with reference, for example, to cylindrical bushings, such as bronze sleeves that act as journal bearings. Such sleeves are positioned within an aperture or fitting in a machine, and a shaft is positioned within the bushing, where it is supported for rotation. Frequently a thin oil film is provided to further reduce friction. In the present specification and claims, bushing is used more broadly to refer to materials and structures that serve a similar function, whether in a cylindrical shape or some other shape, such as, for example, a component having a planar surface. The term bushing is also used to refer to conventional mechanical bearings such as, for example, roller bearings and ball bearings. In the specification, solid bushing may be used to distinguish an element made from bushing material from an element configured to operate as a conventional mechanical bearing, with rollers, balls, etc. Nevertheless, when used in the claims, bushing is to be construed broadly to include either general type of structure unless otherwise specified. 
   Referring to  FIG. 2 , a simplified sectional view of a portion of a hydraulic pump/motor  200  is illustrated. The motor  200  operates according to the principles outlined with respect to  FIGS. 1A-1C , and includes a drive plate  202  carried in a casing  204  and driven by pistons  206 . The pistons  206  are driven by pressurized fluid in cylinders of a cylinder barrel that is not shown in  FIG. 2 , since the basic operation of a typical bent-axis hydraulic motor is previously described and is well known in the art. Fluid bearings are provided for operation of the motor  200 . More particularly, an axial, or thrust bearing  208  is positioned and configured to receive axial loads from the drive plate  202 , while radial bearing  210 , is configured to receive radial loads exerted thereon by the drive plate  202 . Radial bearing  212  may be a fluid type bearing or a conventional bearing, and is configured to support the output shaft  220 . 
   Fluid supply lines indicated generally and diagrammatically at  214  provide pressurized fluid to the thrust bearing  208 , while fluid supply lines  216  provide pressurized fluid to the radial bearing  210 . The fluid supply lines  214  and  216  illustrated in  FIG. 2  are not intended to represent the actual number or arrangement of supply lines necessary for any particular embodiment, inasmuch as such details will depend on various design factors that will become obvious to one of ordinary skill upon review of the present description. 
   Some bent-axis pump/motors are referred to as over-center machines because they are capable of varying the stroke angle in the negative direction, i.e., downward, as viewed in  FIG. 1 . This allows the machine to reverse rotation by moving to a negative angle rather than by reversing polarity of the fluid pressure driving the motor. The motor  200  of  FIG. 2  is not described below as an over-center machine, but as the more common type, in which the piston angle is always 0° or above. Nevertheless, the principles described herein may be applied to the operation of an over-center machine, as well. 
   In the description that follows, the structure and operation of various embodiments will be described with reference to a nine-cylinder motor. One of ordinary skill will recognize that the principles described may be easily adapted for use with motors having other configurations, including motors having an even-number of cylinders. 
   The force from the pistons  206  is distributed as axial force, which is parallel to the axis C, and radial force, which is perpendicular to the axis C. As the angle of the pistons is varied, the distribution of force varies, between axial and radial, that is applied to the drive plate  202 . If the angle is 0°, in which the pistons are parallel to the axis C, the distribution will be 100% axial and 0% radial. As the stroke angle increases, the axial force decreases as a function of the cosine of the stroke angle, while the radial force increases as a function of the sine of the stroke angle. Depending on the design of the motor  200 , it may be capable of a maximum angle of 45° or greater. At 45° the radial and axial forces will each be about 70% of the maximum axial force. 
   Referring now to  FIG. 3 , the thrust bearing  208  is shown in plan view, showing the surface that contacts the drive plate  202 . The upper surface of the bearing  208  includes a land area  328 , first and second arcuate hydrostatic pads  330 ,  331 , and third and fourth arcuate hydrostatic pads  332 ,  334 , which may include a plurality of radial wetting grooves  336 . According to one embodiment, the pads  330 ,  331 ,  332 , and  334  are formed in an insert  338  that fits within a recess provided in the casing  204 . A pressurized fluid supply is provided to the grooves  330 ,  331 ,  332 , and  334  via supply lines  214 . 
   The bearing insert  338 , or at least the upper surface comprising the land  328 , is formed of a bushing material that is configured to tolerate contact with the drive plate  202  within selected limits. Such material is known in the art. For example, various types of polymerized metals have been developed that are effective in controlling friction. In other cases, a lubricant-impregnated metal may be employed to facilitate a low friction contact. 
   According to an embodiment, the surface area of the land and formulation of the bearing insert  338  are selected such that the bearing  208  can tolerate a direct load during operation of the motor  200  of up to around 35% of the maximum axial load of the motor  200 . Because the land  328  of the bearing  208  is configured to operate as a solid bushing in contact with the drive plate  202 , it is beneficial to maintain a thin film of lubricating fluid between the bearing  208  and drive plate  202 . Accordingly, in the embodiment illustrated, radial wetting grooves  336  are provided to place hydraulic fluid in contact with a large portion of the surface area in contact between the drive plate  202  and bearing  208  as the drive plate  202  rotates, in order to distribute the fluid as a lubricant. Alternative embodiments may omit such grooves, or provide other means for wetting the contact surfaces of the bearing and drive plate. 
   In operation, before the stroke angle of the motor  200  is rotated from 0° to begin rotation from a stopped condition, either the pads  330  and  334  or  331  and  332  are provided with hydraulic fluid at a fluid pressure sufficient to offset at least 65% of the maximum axial load so that the remaining axial load falls within the 35% that the bearing  208  can tolerate. The determination of which grooves are pressurized is determined by the polarity of the motor  200 : if the cylinders on the left, as viewed from the orientation of the bearing  208  pictured in  FIG. 3 , are pressurized at high-pressure, grooves  330  and  334  will be pressurized to offset the force exerted by the pistons of those cylinders. Conversely, if the polarity of the motor is reversed so that the high-pressure force is exerted on the right, pads  331  and  332  will be pressurized. This may be accomplished by the same valves that control polarity of the motor, or separate fluid valves may be employed for this purpose. For the purpose of this description, it will be assumed that the polarity of the motor is such that the high-pressure force is exerted on the left, over pads  330  and  334 . It will be understood that operation of the motor in the opposite polarity is substantially identical, except that pads  331 ,  332  will support the high-pressure force. 
   As the stroke angle increases from 0° and the drive plate  202  begins to rotate with respect to the casing  204  and the bearing  208 , the axial load on the bearing  208  will begin to drop.  FIG. 4  is a graph showing the relationship between the stroke angle of a motor having nine cylinders and the axial load on a thrust bearing such as bearing  208  of  FIG. 3 , as a percentage of the maximum axial force of the motor. Line L 1  traces the axial load exerted while five of the nine cylinders are pressurized, and line L 2  traces the axial load exerted while four of the nine cylinders are pressurized. As the cylinder barrel of the motor rotates, the axial load will constantly fluctuate between L 1  and L 2  along a vertical line corresponding to the particular stroke angle of the motor. Line H 1  traces the offsetting surface force applied by the pressurized fluid in the pads  330  and  334 , and the area between the line H 1  and the lines L 1  and L 2  shows the residual force exerted by the drive plate  202  on the land  328  of the bearing  208  under the pressure of the five and four pistons, respectively at any given stroke angle. This force may be referred to as a clamping force, holding the drive plate  202  against the bearing  208  and preventing fluid loss from the bearing  208 . 
   As long as the offsetting surface force of the pads  330 ,  334  remains below the balance force necessary to operate in full hydrostatic operation, the bearing remains clamped and there is no appreciable leakage of fluid from the pads  330 ,  334 . The bearing  208  of the present embodiment is configured to operate in this manner to minimize fluid loss, and thereby improve operational economy. It may be seen that as the axial load decreases in response to the increase of the stroke angle, the line H 1  converges with L 1 , and L 2 . If at any point during the operation of the motor the forces represented on the graph by the lines H 1  and L 2  intersect, the surface force will exceed the balance force, and at that stroke angle the bearing will begin to operate in full hydrostatic mode and fluid will be forced from the pads  330 ,  334  each time the axial force drops from L 1 , to L 2 . As indicated above, this condition results in a loss of pressurized fluid, and is to be avoided. Accordingly, when the stroke angle of the motor increases above around 31°, according to the embodiment described herein, fluid pressure to pad  330  is removed, i.e., shut off by a valve in the appropriate supply line  214 . This reduces the total effective area of the hydrostatic pads of the bearing  208 , and thus reduces the surface force, as shown in  FIG. 4 , thereby maintaining clamping force on the bearing  208 . This ability to adjust the effective area of the hydrostatic pads enables the bearing to withstand the varying forces without resulting in significant overpressure or leakage as would otherwise occur. 
   For the purpose of this description, the range in which both pads  330 ,  334  are pressurized will be referred to as the first zone of operation, and the range in which only pad  334  is pressurized will be referred to as the second zone of operation. The respective areas of pads  330  and  334  are selected such that throughout the second zone of operation, the clamping force is still within the 35% limit of the bearing land  328 , yet will not intersect L 2  at the maximum stroke angle of 45°. To avoid chatter when the motor is operating at a stroke angle very near the transition point between the first and second zones, a hysteresis path may be provided such that the system transitions from the first zone to the second zone at a higher angle, shown as path H 1A  in  FIG. 4 , than the return transition from the second zone to the first, shown as path H 1B . 
   Additional zones of operation may be provided for by incorporating additional fluid pads on the face of the bearing. This will permit the formulation of bearings that are not required to withstand such a high load, but would also require more frequent switching as the stroke angle changes and requires changes to the additional zones of operation. In some alternative embodiments, a single fluid pad is provided, thereby reducing the complexity of the bearing, as compared to a bearing having two or more pads. In such embodiments, the single fluid pad may be configured to provide a constant separating force at about the level shown in the second zone of  FIG. 4 , for example, or the fluid pressure may be reduced or switched off at a selected stroke angle. It will be recognized that in some of these embodiments, it will be necessary for the bushing to be configured to tolerate a higher load than that of the embodiments pictured. 
   Particular features of the bearing, such as, for example, the number of pads, the area of each of the pads and lands of the bearing, the fluid switching scheme, and the arrangement of wetting grooves, are all matters of design that will be influenced by factors such as the maximum axial load, duty cycle, machine size, number of cylinders, etc., and are within the abilities of one of ordinary skill in the art. 
   According to an embodiment, the areas of the pads are selected such that the system operates as described above when the fluid supply pressure used to pressurize the pads is equal to the high-pressure fluid used to drive the motor. This reduces the complexity and increases reliability of the system as compared to systems that require regulated pressure for hydrostatic operation. Alternatively, fluid pressure to the hydrostatic pads may be regulated to a pressure that is different from the fluid pressure employed to provide power to the motor. 
   An issue that arises in many hydraulic motors of the type described here is the problem of stiction. When the motor is at a zero stroke angle and there is no rotation, the maximum axial force is applied to the bearing, but there is no opportunity to maintain fluid lubrication such as occurs when the motor is rotating and the wetting grooves  336  continually wipe the surface of the drive plate  202 . As a result, the thin film of lubricating fluid may be squeezed out from between the drive plate  202  and the bearing  208 . This creates a sticking effect between the land  328  and the drive plate, which resists initial rotation of the motor. To prevent stiction, one or both of the opposite pads  331 ,  332  may be pulsed with fluid pressure when the motor is first rotated from a zero stroke angle. The clamping force can easily be overcome by such a fluid pulse, which will momentarily lift drive plate  202  and force fluid between the land  328  and the drive plate  202 . 
   Referring now to  FIG. 5 , the radial bearing  210  is illustrated according to an embodiment of the invention. Bearing  210  comprises upper and lower bearing races  502 ,  504  configured to be received in recesses in the casing  204  of the motor  202 , as shown in  FIG. 2 . Lower race  504  includes first, second, and third hydrostatic pads  506 ,  508 , and  510  surrounded by a land  512 . The first pad  506  is centered in the lower race  504 , while the second pad  508  comprises sections  508   a  and  508   b  that are spaced outward from the first pad  506 , on either side, and the third pad  510  comprises sections  510   a  and  510   b  that are spaced outward from pads  506  and  508 . Sections  508   a  and  508   b  are in fluid communication with each other via fluid lines not shown in detail, and sections  510   a  and  510   b  are likewise in fluid communication with each other. The bearing  210  is formed of a suitable bushing material and is configured to support direct contact of the drive plate up to about 25% of the maximum axial load, or about 35% of the maximum radial load. Fluid supply lines  216  provide an individually switchable pressurized fluid supply to each of pads  506 ,  508 , and  510 . The upper and lower races  502 ,  504  are centered over TDC and BDC, respectively, as shown in  FIG. 2 . Because any radial load in this machine is always directed downward (as oriented in  FIG. 2 ), the upper race  502  will receive only nominal loads, and so the upper race is not provided with fluid pads. 
     FIG. 6  is a graph showing the relationship between the stroke angle of the motor  200  and the radial load on the bearing  210 , as a percentage of the maximum axial force of the motor  200 . Line L 3  traces the radial load exerted while five of the nine cylinders are pressurized, and line L 4  traces the radial load exerted while four of the nine cylinders are pressurized. Line H 2  traces the offsetting surface force applied by pressurized fluid in pads  506 ,  508 , and  510 . The clamping force is represented by the vertical distance between the line H 2  and the lines L 3  and L 4 . 
   In contrast to the axial load described with reference to  FIG. 3 , the radial load on the bearing  210  is substantially zero while the motor is at a zero stroke angle, and rises as the stroke angle increases, as a function of the sine of the stroke angle. Accordingly, only sufficient fluid pressure to wet the land  512  of the bearing  210  is provided until the stroke angle reaches about 18°, at which point high-pressure fluid is supplied to the first pad  506 . In like manner, as the stroke angle reaches about 25°, the second pad  508  is also pressurized, and as the stroke angle reaches about 36°, the third pad is also pressurized. Because the second and third pads  508 ,  510  are each separated into two sections, the offsetting force provided is balanced with respect to the drive plate  202 . Though not shown in  FIG. 6 , switching of the fluid pressure in the respective pads  506 ,  508 , and  510  may be provided hysteretically to avoid chatter, as described in more detail with reference to  FIG. 4 . 
   In a like manner as was described with reference to the axial bearing  208  to avoid stiction, any or all of pads  506 ,  508  or  510  may be pulsed with fluid pressure at a lower stroke angle to momentarily overcome the radial clamping force and allow the motor to begin rotation. Additionally, in cases where the motor is configured to operate under conditions requiring extremely high torque such that a maximum stroke angle is required to initiate rotation, an additional fluid pad, or provision for a momentarily boosted fluid supply pressure, may be provided to create a sufficient separating force to overcome stiction. 
   Because the radial forces on the drive plate are substantially unidirectional, the bearing  210  receives those forces only in the region supported by the hydrostatic pads  506 ,  508 , and  510 . The drive plate  202  does not undergo significant lateral radial loads, and so there is no requirement for increased bearing surfaces on the sides of the plate  202 . Accordingly, the bearing  210  can comprise the upper and lower races  502 ,  504 , with substantial area between that is not supported by the bearing. This is particularly advantageous in a bent-axis pump/motor of the type described herein. Though not shown, such motors typically include a yoke that supports the valve plate and cylinder barrel, and which rotates on pins or trunnions that are positioned on either side of the drive plate to control the stroke angle rotation. In motors employing conventional bearings, such as the motor illustrated in  FIGS. 1A-1C , the radial bearings must be positioned below the drive plate, as shown at  118 , to avoid interfering with the trunnions. However, this gives rise to the lever/fulcrum action described with reference to  FIGS. 1A-1C , which transmits radial loads down the length of the output shaft, necessitating additional substantial radial load bearings. In contrast, radial bearings according to embodiments of the invention can be positioned higher on the drive plate so that they are more nearly directly opposite the radial force vectors, thereby substantially eliminating the lever/fulcrum action. Accordingly, bearings that support the output shaft (such as the second radial bearing  212  of  FIG. 2 ) need not be configured to tolerate large radial loads. 
   According to an embodiment of the invention, the upper race  502  of the radial bearing  210  is provided with hydrostatic pads similar to those described with reference to the lower race  504 , for operation with an over-center motor, such that when the motor is stroked to a negative angle, the pads of the upper race are pressurized as described above with reference to the lower race  504 . 
   Control of hydrostatic fluid pressure in the fluid bearings, according to the various embodiments of the invention, may be provided by valving that is integral with the device associated with the bearing. For example, in the case of a bent-axis pump/motor of the kind described with reference to  FIG. 2 , the same valves that are configured to provide the high- and low-pressure fluid supply to the motor  200  may also be adapted to provide pressurized fluid to the bearings. The valve may also incorporate pressure regulators or the like. Alternatively, separate control valves may be employed for these purposes. Furthermore, control devices such as mechanical linkages, electronic devices and circuits, and computer modules may be employed to regulate fluid switching and pressure. All of these control systems are within the abilities of one of ordinary skill in the art. 
   Embodiments of the invention have been described with respect to an insert received in a recess of a component, usually the stationary component, such as the casing of the motor of  FIG. 2 . According to another embodiment of the invention, the hydrostatic pads are formed directly in a surface of one of the components. Furthermore, according to an embodiment, the pads may be formed in the rotating component. In such a case, fluid pressure may be provided via supply lines in the stationary component and opening onto a face of the stationary component such that as the pads rotate over the end of the supply lines, fluid is provided to the pads. 
   Embodiments of the invention have been described in which solid bushings are employed as lands to withstand the entire clamping force on the bearing. It will be recognized that, especially in motors having an odd number of bearings, the clamping force may at some stroke angles be significant, necessitating a land having a substantial surface area. According to alternate embodiments, some or all of the clamping force may be received by bushings configured as conventional mechanical bearings. For example,  FIG. 7  shows a motor  700  that is substantially identical to the motor  200  illustrated in  FIG. 2  except that the motor  700  includes a small tapered roller bearing  718 . Bearing  718  is configured to act as a supplemental bushing, receiving a portion of the clamping force that would otherwise be applied to the lands of bearings  208  and  210 . This permits the surface area of the respective lands to be significantly reduced, thereby reducing the overall dimensions of the motor  700  as compared to those of the motor  200  of  FIG. 2 . Nevertheless, because most of the load exerted by the drive plate  202  is supported by the fluid pads of the bearings  208  and  210 , the collective size of the bearings  208 ,  210 , and  718  is at least substantially comparable to the size of the bearing or bearings of an otherwise equivalent conventional motor, while the comparable efficiency is significantly superior. 
   According to embodiments of the invention, conventional bearings may be employed to supplement or replace solid bushings in both radial and axial load bearings, or in only one or the other. Furthermore, fluid bearings may be used in combination with non-fluid bearings. For example, In some embodiments it may be more practical to use a fluid axial bearing and a mechanical radial bearing, or vice-versa. Finally, while embodiments of the invention have been described with reference to their operation in a hydraulic motor, the scope of the invention is not limited to that application. The principles of the invention may be practiced in a wide range of applications to support axial and radial loads. 
   All of the above U.S. patents, U.S. patent application publications, U.S. patent applications, foreign patents, foreign patent applications and non-patent publications referred to in this specification and/or listed in the Application Data Sheet, are incorporated herein by reference, in their entirety. 
   From the foregoing it will be appreciated that, although specific embodiments of the invention have been described herein for purposes of illustration, various modifications may be made without deviating from the spirit and scope of the invention. Accordingly, the invention is not limited except as by the appended claims.

Technology Category: 2