Abstract:
A bearing assembly is provided for a fuel delivery system that includes a pump ( 10 ) having a housing that rotatably receives a rotor ( 20 ) carrying vanes ( 26 ) thereon, a cam ring ( 70 ) received between the housing and rotor ( 20 ), and a support member of yoke ( 50 ) encompassing the cam ring ( 70 ) to selectively vary fuel flow. The bearing assembly ( 80 ) is a journal bearing between the yoke ( 50 ) and the cam ring ( 70 ) and includes an annular surface having a central opening therethrough. The annular surface includes a first, high pressure pad ( 102 ) and a second, low pressure pad ( 104 ) substantially diametrically opposite the first pad and separated by first and second lands ( 106, 108 ). The circumferential extent of the first pad ( 102 ) is at least as great as an inner diameter of the cam ring ( 70 ). Circumferential ends of the second pad ( 104 ) are wider than circumferential ends of the first pad. The first and second pads ( 102, 104 ) are formed by circumferentially extending grooves that extend an entire width of the bearing so that the cam ring moves between the first and second pads, and thereby varies a clearance between the lands ( 106, 108 ) and the cam ring ( 70 ).

Description:
This application claims the benefit of International Application No. PCT/US03/022680, International Filing Date 21 Jul. 2003 and U.S. Provisional Application Ser. No. 60/397,212, filed 19 Jul. 2002. 

   BACKGROUND OF THE INVENTION 
   The present invention relates to a bearing arrangement, and more particularly to a bearing arrangement used to support a cam ring within a support member or yoke in a hydrostatic and hydrodynamic configuration for use in fuel pumps, metering, and control for jet engines. 
   PCT/US02/09298, filed Mar. 27, 2002, the details of which are incorporated herein by reference, relates to a fuel delivery system having increased efficiency and reliability over known fuel pump arrangements. Particularly, a pump of a fuel delivery system includes a housing having a chamber with an inlet and outlet in fluid communication with the pump chamber. A rotor is received in the pump chamber, and a cam member surrounds the rotor and is freely rotatable relative to the housing and the rotor. A journal bearing is formed between the cam ring and a support sleeve or yoke that is precluded from rotation within the housing. 
   The bearing arrangement must be responsive to hydrostatic and hydrodynamic forces imposed thereon by the internal components of the pumping mechanism. Known bearing arrangements require improvement to properly support the cam ring in a combined hydrostatic and hydrodynamic arrangement. Accordingly, a need exists for a new bearing assembly. 
   SUMMARY OF THE INVENTION 
   An improved bearing assembly is provided for a fuel delivery system that includes a housing receiving a rotor within a rotatable cam ring, where the cam ring is freely rotatable relative to the housing and the rotor. The bearing assembly includes an annular surface having a central opening dimensioned to receive the associated cam ring. 
   The annular surface includes a first, high pressure pad and a second low pressure pad spaced by first and second lands. 
   The circumferential extension of the first pad is at least as great as an inner diameter of the cam ring. 
   Circumferential ends of the second pad are preferably wider than the circumferential ends of the first pad. 
   A differential pressure is established across the pump chamber and the cam ring is capable of movement between the high and low pressure pads in response to pressure variations. Clearance between the land and the cam ring selectively alters the flow of fluid through the bearing to maintain a pressure. This creates a relatively stiff bearing mount without deflection concerns. 
   A primary advantage of the invention resides in an improved bearing interface between a rotating cam ring and stationary (non-rotatable), but moveable yoke. 
   Another advantage of the invention resides in the structure being capable of providing hydrostatic bearing capabilities, as well as hydrodynamic bearing capabilities. 
   Still other benefits and advantages of the invention will become apparent to those skilled in the art upon a reading and understanding of the following detailed description. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is an exploded perspective view of a preferred embodiment of the fluid pump. 
       FIG. 2  is a cross-sectional view through the assembled pump of  FIG. 1 . 
       FIG. 3  is a longitudinal cross-sectional view through the assembled pump. 
       FIG. 4  is a cross-sectional view similar to  FIG. 2  illustrating a variable displacement pump with the support ring located in a second position. 
       FIG. 5  is an enlarged cross-sectional view of the pump. 
       FIG. 6  is an exploded perspective view of the bearing assembly. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   As shown in the Figures, a pump assembly  10  includes a housing  12  having a pump chamber  14  defined therein. Rotatably received in the chamber is a rotor  20  secured to a shaft  22  for rotating the rotor within the chamber. Peripherally or circumferentially spaced about the rotor are a series of radially extending grooves  24  that operatively receive blades or vanes  26  having outer radial tips that extend from the periphery of the rotor. The vanes may vary in number, for example, nine (9) vanes are shown in the embodiment of  FIG. 2 , although a different number of vanes can be used without departing from the scope and intent of the present invention. As is perhaps best illustrated in  FIG. 2 , the rotational axis of the shaft  22  and rotor  20  is referenced by numeral  30 . Selected vanes (right-hand vanes shown in  FIG. 2 ) do not extend outwardly from the periphery of the rotor to as great an extent as the remaining vanes (left-hand vanes in  FIG. 2 ) as the rotor rotates within the housing chamber. Pumping chambers are defined between each of the vanes as the vanes rotate in the pump chamber with the rotor and provide positive displacement of the fluid. 
   With continued reference to  FIG. 2 , a spacer ring  40  is rigidly secured in the housing and received around the rotor at a location spaced adjacent the inner wall of the housing chamber. The spacer ring has a flat or planar cam rolling surface  42  and receives an anti-rotation pin  44 . The pin pivotally receives a cam sleeve  50  that is non-rotatably received around the rotor. First and second lobes or actuating surfaces  52 ,  54  are provided on the sleeve, typically at a location opposite the anti-rotation pin. The lobes cooperate with first and second actuator assemblies  56 ,  58  to define means for altering a position of the cam sleeve  50 . The altering means selectively alter the stroke or displacement of the pump in a manner well known in the art. For example, each actuator assembly includes a piston  60 , biasing means such as spring  62 , and a closure member  64  so that in response to pressure applied to a rear face of the pistons, actuating lobes of the cam sleeve are selectively moved. This selective actuation results in rolling movement of the cam sleeve along a generally planar or flat surface  66  located along an inner surface of the spacer ring adjacent on the pin  44 . It is desirable that the cam sleeve undergo a linear translation of the centerpoint, rather than arcuate movement, to limit pressure pulsations that may otherwise arise in seal zones of the assembly. In this manner, the center of the cam sleeve is selectively offset from the rotational axis  30  of the shaft and rotor when one of the actuator assemblies is actuated and moves the cam sleeve ( FIG. 2 ). Other details of the cam sleeve, actuating surface, and actuating assemblies are generally well known to those skilled in the art so that further discussion herein is deemed unnecessary. 
   Received within the cam sleeve is a rotating cam member or ring  70  having a smooth, inner peripheral wall  72  that is contacted by the outer tips of the individual vanes  26  extending from the rotor. An outer, smooth peripheral wall  74  of the cam ring is configured for free rotation within the cam sleeve  50 . More particularly, a journal bearing  80  supports the rotating cam ring  70  within the sleeve. The journal bearing is filled with the pump fluid, here jet fuel, and defines a hydrostatic or hydrodynamic, or a hybrid hydrostatic/hydrodynamic bearing. The frictional forces developed between the outer tips of the vanes and the rotating cam ring  70  result in a cam ring that rotates at approximately the same speed as the rotor, although the cam ring is free to rotate relative to the rotor since there is no structural component interlocking the cam ring for rotation with the rotor. It will be appreciated that the ring rotates slightly less than the speed of the rotor, or even slightly greater than the speed of the rotor, but due to the support/operation in the fluid film bearing, the cam ring possesses a much lower magnitude viscous drag. The low viscous drag of the cam ring substitutes for the high mechanical losses exhibited by known vane pumps that result from the vane frictional losses contacting the surrounding stationary ring. The drag forces resulting from contact of the vanes with the cam ring are converted directly into mechanical losses that reduce the pumps overall efficiency. The cam ring is supported solely by the journal bearing  80  within the cam sleeve. The journal bearing is a continuous passage. That is, there is no interconnecting structural component such as roller bearings, pins, or the like that would adversely impact on the benefits obtained by the low viscous drag of the cam ring. For example, flooded ball bearings would not exhibit the improved efficiencies offered by the journal bearing, particularly a journal bearing that advantageously uses the pump fluid as the fluid bearing. 
   In prior applications these mechanical drag losses can far exceed the mechanical power to pump the fluid in many operating regimes of the jet engine fuel pump. As a result, there was a required use of materials having higher durability and wear resistance because of the high velocity and load factors in these vane pumps. The material weight and manufacturing costs were substantially greater, and the materials also suffer from high brittleness. The turning speed of those pumps was also limited due to the high vane sliding velocities relative to the cam ring. Even when using special materials such as tungsten carbide, high speed pump operation, e.g., over 12,000 RPM, was extremely difficult. 
   These mechanical losses resulting from friction between the vane and cam ring are replaced in the present invention with much lower magnitude viscous drag losses. This results from the ability of the cam ring to rotate with the rotor vanes. A relatively low sliding velocity between the cam ring and vanes results, and allows the manufacturer to use less expensive, less brittle materials in the pump. This provides for increased reliability and permits the pump to be operated at much higher speeds without the concern for exceeding tip velocity limits. In turn, higher operating speeds result in smaller displacements required for achieving a given flow. In other words, a smaller, more compact pump can provide similar flow results as a prior larger pump. The pump will also have an extended range of application for various vane pump mechanisms. 
     FIG. 3  more particularly illustrates inlet and outlet porting about the rotor for providing an inlet and outlet to the pump chamber. First and second plates  90 ,  92  have openings  94 ,  96 , respectively. Energy is imparted to the fluid by the rotating vanes. Jet fuel, for example, is pumped to a desired downstream use at an elevated pressure. 
   As shown in  FIG. 4 , neither of the actuating assemblies is pressurized so that the cam sleeve is not pivoted to vary the stroke of the vane pump. That is, this no flow position of  FIG. 4  can be compared to  FIG. 2  where the cam sleeve  50  is pivoted about the pin  44  so that a close clearance is defined between the cam sleeve and the spacer ring  40  along the left-hand quadrants of the pump as illustrated in the Figure. This provides for variable displacement capabilities in a manner achieved by altering the position of the cam sleeve. 
   In the preferred arrangement, the vanes are still manufactured from a durable, hard material such as tungsten carbide. The cam ring and side plates, though, are alternately formed of a low cost, durable material such as steel to reduce the weight and manufacturing costs, and allow greater reliability. Of course, it will be realized that if desired, all of the components can still be formed of more expensive durable materials such as tungsten carbide and still achieve substantial efficiency benefits over prior arrangements. By using the jet fuel as the fluid that forms the journal bearing, the benefits of tungsten carbide for selected components and steel for other components of the pump assembly are used to advantage. This is to be contrasted with using oil or similar hydraulic fluids as the journal bearing fluid where it would be necessary for all of the jet fuel components to be formed from steel, thus eliminating the opportunity to obtain the benefits offered by using tungsten carbide. 
   As illustrated in greater particularity in  FIGS. 5 and 6 , the journal bearing assembly defined by the interface between the cam sleeve or yoke  50  and the cam ring  70  is shown in greater detail. Particularly, the inner surface  100  of the support sleeve or yoke is a non-constant diameter to define discrete portions of the bearing arrangement. Specifically, a first or large diameter portion  102  defines a first, high pressure pad and a diametrically opposite, second or low pressure pad  104 . For ease of description, and as will be appreciated from  FIG. 5 , the high pressure pad portion  102  extends from approximately 4 o&#39;clock to 8 o&#39;clock while the low pressure pad extends from approximately 10 o&#39;clock to 2 o&#39;clock. Separating the high pressure pad from the low pressure pad are first and second seal lands  106 ,  108 . The first seal land  106 , therefore extends from approximately 2 o&#39;clock to 4 o&#39;clock, while the second seal land  108  extends from approximately 8 o&#39;clock to 10 o&#39;clock. 
   The bearing arrangement defines a combination hydrostatic and hydrodynamic configuration. The hydrostatic portion of the bearing is the two pad arrangement defined by the high pressure and low pressure pads  102 ,  104 , respectively. The high pressure pad is a groove cut through the full width or extent of the yoke, i.e., from a front face  50   a  to a rear face  50   b , as will be more clearly appreciated from a review of  FIG. 6 . Likewise, the low pressure pad is also a groove through the full width of the yoke. The high pressure pad is capable of supporting the forces generated by the internal components of the pumping mechanism. Between the two pads, in the yoke, are the seal lands  106 ,  108  that create a hydrodynamic effect that enables smooth start-up and centers the cam ring within the bearing during operation. 
   The high pressure pad geometry is determined so that the force generated by the fluid pressure is slightly greater than the forces generated by the internal pumping elements. The circumferential extent of the pad  102 , i.e., from 4 o&#39;clock to 8 o&#39;clock, is determined by the radial thickness of the cam ring. It is preferred that the edges  102   a ,  102   b  of the high pressure pad are located outside the inside diameter  72  of the cam ring (see  FIG. 5 ). The seals and the sides of the high pressure groove, that is along the faces  50   a ,  50   b  of the yoke, are created by the port plates  90 ,  92  ( FIG. 3 ) which clamp across the pumping element. High pressure fluid (jet fuel) is fed into the pad through openings  120  shown in  FIG. 6  and the flow to the interface between the yoke and the cam ring is restricted through orifices  122  (only one of which is seen in the view of  FIG. 6 ). As will be appreciated, the high pressure orifices  122  communicate with respective openings or holes  120  in this region of the bearing assembly. 
   The geometry of the low pressure pad  104  is determined by setting circumferential edges  104   a ,  104   b  slightly wider than the circumferential edges of the high pressure pad, i.e., slightly wider than  102   a ,  102   b , respectively. Venting from the high pressure pad to the low pressure pad must be provided in this pad such that high pressure does not build. This is provided through openings  124 , one of which is illustrated in  FIG. 6 . As will be apparent, openings  124  have a substantially larger diameter than openings  122 . Therefore, a differential pressure is established across the yoke to react the forces within the pumping element. 
   The high and low pressure pads  102 ,  104  are cut completely through the bearing, i.e., they extend completely from face  50   a  to  50   b , to allow the cam ring to move in the vertical direction as depicted in  FIG. 5 . The movement in the vertical direction allows for radial deflection of the yoke in the horizontal direction, thus increasing the clearance between the lands and the cam ring. When the clearance increases, the flow through the bearing must increase to maintain the pressure in the high pressure pad, or the clearance must be reduced. The orifices  122  on the high pressure pad side restrict the flow and thus the cam ring moves vertically forward decreasing the clearance to re-establish an equilibrium force condition. This creates a relatively stiff bearing without the concerns of deflection. 
   The entire bearing, yoke  50  and cam ring  70  is free to roll within the pumping mechanism as described above. As shown in  FIG. 5 , the bearing rolls leftwardly or rightwardly along the generally planar surface  42  provided in the spacer ring  40 . This rolling on the surface  42  acts to provide a linear translation of the cam ring. Linear cam ring translation is critical to minimizing fluid pump pressure pulsation during operation. Sliding and rotation of the yoke are prevented by the anti-rotation disks  44  inserted on each side of the yoke. As will be apparent from  FIG. 6 , these anti-rotation disks  44  are dimensioned for receipt in arcuate recesses or cutouts  130 , only one of which is illustrated in  FIG. 6 , although it will be appreciated that a similar cutout recess is provided on the rear surface  50   b  of the yoke. Thus, these anti-rotation disks  44  do not pass completely through the yoke, or corresponding recesses provided in the spacer ring, and thereby allow the forces in yoke to be transmitted to the housing structure through the spacer ring. 
   It will also be appreciated that in the preferred embodiment of the yoke  50 , an undercut  140  is provided on the first and second surfaces  50   a ,  50   b . The undercut  140  is provided at the outer radial perimeter of these faces. Moreover, the undercut extends circumferentially around substantially the entire yoke, i.e., from approximately 6:30 in a clockwise direction to approximately 5:30. The undercut facilitates control of pressure on the face of the yoke and accurately predicts or controls the pressure of the overall pump arrangement. 
   The invention has been described with reference to the preferred embodiments. Obviously, modifications and alterations will occur to others upon reading and understanding the preceding detailed description. It is intended that the invention be construed as including all such modifications and alterations in so far as they come within the scope of the appended claims or the equivalents thereof.