Abstract:
A long-life, low maintenance, bi-directional vane-type water pump has a high degree of symmetry and operates with equal efficiency in either direction. The axial position of the drive shaft is controlled to permit improved lubrication by the pumping fluid of component parts on which the drive shaft is journaled.

Description:
FIELD OF THE INVENTION 
     The present invention relates generally to a fluid pressure energy translating device of the vane type that is suitable for applications such as pumping water in space applications and employs water as the lubricating fluid. 
     BACKGROUND OF THE INVENTION 
     The design of a vane pump for pumping fresh water in space applications presents a serious challenge to the designer because of requirements of light weight and infrequent maintenance. Also, when pumping water it is desirable for the pump to be self-lubricating, i.e., to use the pumped fluid itself as a lubricant. The poor lubricity and low viscosity of water compared with lubricating oils contributes to the challenge. The low viscosity dictates that all design clearances must be an order of magnitude less than for oil lubricated devices. In addition, the potential contamination of scarce water in a space vehicle requires that no oils or greases be used. A high pumping efficiency is clearly advantageous, since a given pumping rate is achievable with the minimum expenditure of power. 
     Generally, vane devices comprise a circular rotor disposed within a non circular cam ring, so that the gap between the rotor and the cam ring varies according to the angular position within the ring. Vanes are disposed in openings around the periphery of the rotor, and when in motion, make sliding contact with the inside of the cam ring. The vanes are free to move back and forth in the openings, being urged into continuous contact with the cam ring by centrifugal force, springs or hydraulic pressure. As the vanes move around the cam ring, they displace fluid into zones of increasing volume, causing more fluid to enter from an inlet port, or into zones of decreasing volume, from which fluid is discharged through an outlet port. 
     Various examples of vane pumps have been disclosed previously. While various examples of pumps perform satisfactorily for their intended purposes, certain limitations prevent them from performing satisfactorily as water pumps in space environments. In particular, space applications demand that pump weight be minimized and that the pump provide efficient trouble-free operation for extremely long periods with minimal maintenance. 
     SUMMARY OF THE INVENTION 
     The invention disclosed herein describes a bi-directional, self-lubricating vane-type water pump. The pump comprises a rotor with a plurality of radial slots, each of which accommodates a vane. The rotor and vanes are driven by a drive shaft to revolve within a non-circular cam ring, displacing fluid and causing it to enter through an inlet port, or to be discharged through an outlet port, the ports being present in port plates. In this invention, the port plates and the cam ring are disposed in a highly symmetrical fashion, which promotes efficiency and furthermore provides equally efficient operation of the pump in either direction. Within narrow prescribed limits, the drive shaft of the pump is free to float back and forth along its axis. This axial movement may be controlled through a shim washer placed at the end the drive shaft. This provides optimum efficiency, permitting sufficient clearance between components to avoid binding and allow the pumping fluid, for example water, to lubricate where required, but nevertheless preventing excessive play. The fluid flows in the pump are subject to minimal constriction, which also contributes to efficient operation. Additionally, wear resistant and friction resistant materials may be employed for specific component parts, so as to obviate the need for conventional bearings. The pump requires very little maintenance, and is suitable for installation in remote locations such as space. 
     Accordingly, it is an object of this invention to provide an improved pump for fluids of low viscosity which has an extremely long operating lifetime with minimal maintenance and is suitable for space applications. 
     It is further an object of this invention to provide an improved pump for fluids of low viscosity which has a simple design, such that fluid flows are minimally constricted, providing optimal efficiency. 
     It is further an object of this invention to provide an improved bi-directional pump for fluids of low viscosity which has a high internal symmetry, allowing effectively equal efficiency in either direction. 
     It is further an object of this invention to provide a pump requiring minimal maintenance via the elimination of dynamic seals. 
     Finally, it is an object of this invention to provide an improved pump for fluids of low viscosity which is self lubricating. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is an exploded, perspective view of a pump according to various preferred embodiments of this invention. 
     FIG. 2 is a partial cross-sectional view of the pump of FIG.  1 . 
     FIG. 3 is a cross-section of a coupling between the pump and a motor. 
     FIG. 4 is a partial perspective, exploded view of an impeller assembly comprising a cam ring, a rotor and vanes. 
     FIG. 5 is a schematic view of an impeller assembly of the pump. 
     FIG. 6 is an end view of the impeller assembly of FIGS.  1  and  2 . 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring to FIGS. 1 and 2, pump  10  comprises a generally cylindrical housing  12  and an electric motor  40 . A drive shaft  14  includes a first thrust plate  16  integral to its structure, and has a first end  42  and a second end  44 . The first end  42  is in connection with the electric motor  40 . The assembly also includes a spacer  36  and a wave spring  38 , a first port plate  18 , a rotor  20 , a cam ring  22 , vanes  24  including drive pins  26 , a second port plate  28 , a second thrust plate  30 , a screw  32  and a shim washer  34 . In the following description, any references pertaining to an axis will be understood to refer to an axis of rotation  46  of the drive shaft shown in FIG. 2, which axis is shared with the electric motor  40  and the housing  12 . 
     The housing  12  has two ports, a first port  48  axially positioned at the distal end of the housing  12 , and a second port  50  disposed orthogonally to the axis  46  of the housing  12 . A feature of the pump of this invention is that it functions with comparable, and preferably equal, efficiency when pumping in either direction. Thus, when the pump is operating in one direction, port  48  serves as an inlet port and port  50  serves as an outlet port. When the pump is operating in the opposite direction, port  50  serves as an inlet port and port  48  serves as an outlet port. In the following description, for purposes of convenience, port  48  may be referred to as an inlet port and port  50  may be referred to as an outlet port, but it is understood that the inlet and outlet functions of the two ports are interchanged when the pump operating direction is reversed. 
     The end of the housing opposite the inlet port  48  has a circular opening  51 , and this end of the housing is adapted for connection to the electric motor  40 . For example, in the illustrated embodiment, flange  52  of motor  40  includes holes  53 , and housing  12  includes corresponding threaded holes  55  in the surface surrounding opening  51 , whereby the flange  52  is attached to the housing  12  by bolts  54 . As seen in FIG. 2, a portion of the electric motor  40  that extends from flange  52  is received into the interior of housing  12  through opening  51 , and this extending portion may be provided with a circumferential groove  56  for insertion of a seal, such as an O-ring, to provide an effective seal between the motor  40  and the housing  12 . 
     A detailed view of a magnetic coupling between motor  40  and the drive shaft of pump  10  is shown in FIG. 3. A cylindrical, axially aligned permanent drive magnet  58  in the motor  40  is magnetically coupled with a cylindrical mating driven magnet  60 , which is mounted concentrically on end  42  of the drive shaft of pump  10 , for example, magnet  60  may be affixed to end  42  with an adhesive. Thus, driven magnet  60  is disposed radially inward from, and axially aligned to, the drive magnet  58 . Interposed between the drive magnet  58  and the driven magnet  60  is a cup-shaped fluid barrier  61  formed from a thin sheet of nonmagnetic corrosion resistant steel which permits magnetic forces to be transmitted between the magnets. This fluid barrier being integral to the motor housing, it completely seals the motor from the pump to prevent any liquid from passing into the motor. Magnet  60  is free to rotate when driven by magnet  58 , neither magnet having contact with fluid barrier  61 . 
     The first port plate  18  includes an axial opening  63  to rotatably accept the drive shaft  14 , the second end  44  of which is inserted therein such that the first port plate  18  and the first thrust plate  16  are in close proximity. In the assembled pump, the spacer  36  is located between the motor  40  and the first port plate  18 , the spacer having a large enough internal diameter to accommodate the first thrust plate  16  without interference. The wave spring  38  is interposed between the spacer  36  and the motor  40  in order to accommodate any slack in the assembly. 
     Referring to FIGS. 4 and 5, an impeller assembly  64  comprises the rotor  20 , the cam ring  22  and the vanes  24 . The cam ring  22  has an outer cylindrical surface  66  in stationary contact with the inside surface of the housing  12 , and an inner noncylindrical camming surface  69  that defines central opening  68 . Specifically, the opening has an elliptical shape defined by a major diameter  70  and a different minor diameter  72 , the two diameters offset from each other by 90°. The opening  68  is symmetrically disposed about, or concentric with, the axis  46 . As seen in FIG. 2, the outer perimeter of the cam ring  22  may be provided with a circumferential groove  57  for insertion of a seal, such as an O-ring, to provide an effective seal between the cam ring  22  and the housing  12 . 
     A rotor  20  is placed with the cam ring  22 , the rotor having a circular perimeter  76  and an outer cylindrical surface  78 . Rotor  20  is symmetrically disposed about the axis  46 , such that rotor  20  is concentric with respect to the cam ring  22 . The diameter of the rotor outer surface  78  approximates the minor diameter  72  of the cam inner surface. Accordingly, the insertion of the outer cylindrical surface  78  of the rotor within the elliptical camming surface of the cam ring  22  provides two diametrically opposed gaps  80  therebetween, the gaps arranged symmetrically with respect to one another about diameter  72 . 
     As best seen in FIG. 4, outer surface  78  of the rotor  20  has a plurality of spaced radial slots  82  formed therein to accept vanes  24 . The rotor  20  also has a central axial opening  84  and a plurality of smaller openings  86  around the periphery of the axial opening  84 , these recesses being aligned with the axis  46  and sized to accommodate the drive pins  26 . Disposed around a circumferential zone of the drive shaft are recesses  88  which correspond and align with the smaller openings  86  in the rotor  20  so that the drive pins  26  may be inserted into the openings  86  and recesses  88  to engage the drive shaft  14  with the rotor  20 . 
     Inserted into the slots  82  are the vanes  24 . Each of the vanes  24  is generally rectangularly shaped with a base and an arcuate outer end surface  90 . Vanes  24  are free to translate within slots  82 , such that when the rotor  20  revolves during the operation of the pump  10 , centrifugal force maintains surfaces  90  of the vanes in sliding contact with the inner surface  69  of the cam ring  22 . In other words, the cam ring remains stationary, and as the rotor rotates, the vanes are free to translate radially according to their position relative to the cam ring. It is noted that it is unnecessary for the vanes to be spring-biased according to the illustrated embodiment. 
     The impeller assembly  64  is positioned between the first port plate  18  and the second port plate  28 , with the cam ring  22  remaining stationary with respect the port plates, and the rotor  20  rotating with respect to the port plates  18  and  28 . The axial location of second port plate  28  is defined by a step  91  in the interior wall of the housing. As will be described further, the port plates  18  and  28 , which are essentially identical in their geometry, differ in their orientation within the pump assembly. 
     The second thrust plate  30 , which is mounted within the housing  12  at the same end of the housing as the inlet port  48 , has an annular region  92 , an extension  94  and an opening  96  sized to receive the second end  44  of the drive shaft  14 . The opening  96  penetrates the entire thickness of the annular region  92  and into the extension  94 , terminating at a cap  98 . The cap  98  has an axial hole  100  sized to pass the screw  32 , by which the second thrust plate  30  is fixedly bolted into a corresponding threaded hole  102  in the second end  44  of the drive shaft  14 . A shim washer  34  is situated between the distal end  44  of the drive shaft and the inner shoulder of cap  98  of the second thrust plate. The second thrust plate  30  is in close proximity with the second port plate  28 . 
     Referring further to the port plates  18  and  28 , port plate  18  has two diametrically opposed reniform ports  104  through which fluid can pass, and port plate  28  similarly has two diametrically opposed reniform ports  105 . Port plate  18  also has two diametrically opposed reniform recesses  106 , and port plate  28  includes two similar recesses  107 , which act as fluid reservoirs. The recesses  106  are staggered from the ports  104  by 90°, and the recesses  107  are staggered from the ports  105  by 90°. The ports  104 ,  105  and recesses  106 ,  107  are symmetrically positioned about the axis  46  in a circular band so that they straddle the gap  80  between the rotor  20  and the cam ring  22 . Each such port  104 ,  105  and recess  106 ,  107  extends around an arc of about 45°. In addition to the reniform recesses  106 ,  107 , each of the port plates  18  and  28  also has, facing the rotor, a circular recess  108  close to but not abutting the central opening. Besides their role in providing fluid channels and reservoirs, the port plates  18  and  28  also function as journal bearings for the drive shaft  14 ; the drive shaft is inserted directly in, and journaled by, the port plates requiring no anti-friction bearings. The thrust plates  16  and  30  are sufficiently smaller in diameter than the port plates  18  and  28 , so that the thrust plates do not cover the ports  104 ,  105 . 
     Considering their spatial relationship with the impeller assembly  64 , the port plates  18  and  28  are disposed so that the recesses  106 ,  107  are on the faces of the port plates that abut the rotor. Further, the port plates are radially displaced from each other by 90° with respect to their ports  104 ,  105 , and the ports  104  and  105  are radially equidistant from the major and minor diameters  70  and  72  of the cam ring  22  by 45°. 
     The cam ring  22  and port plates  18  and  28  have corresponding alignment holes  110  and are secured in place with an alignment pin  112  which is inserted in the alignment holes  110  and bolted into a threaded hole in the step  91  of the housing. 
     The second thrust plate  30  has a plurality of radial recesses  114  extending from its outer edge to meet with a circular recess  116  around the opening  96 , the recesses being in the surface which abuts the second port plate  28 . The first thrust plate  16  has like radial recesses meeting with a circular recess  118  where the first thrust plate meets the drive shaft  14 , the recess  118  being shown in FIG.  2 . The recesses of the first thrust plate  16  abut the first port plate  18 . 
     A primary function of shim washer  34  is to control the amount of axial play in the entire assembly of components about the drive shaft  14 . In effect, shim washer  34  determines the distance by which the thrust plates  16  and  30  are separated; the drive shaft  14  is allowed to float axially back and forth by a small but fixed distance, which allows for a film of fluid to be interposed between proximate faces of the port plates  18  and  28  and the thrust plates  16  and  30 . The fluid film acts as a lubricant, which avoids the need to introduce a separate lubricating liquid which potentially may be a source of contamination. Generally, for a given lubricating action, generally, a fluid of low viscosity must be present as a thinner film than a fluid of higher viscosity. In other words, the lubricity of a fluid film tends to degrade more rapidly with increasing film thickness if the fluid has a lower viscosity. Therefore, by controlling axial play, the thickness of the fluid film may be controlled to provide a desired range of lubricity, thereby contributing to the efficiency of the pump. 
     The operation of the pump is dependent on the relationship of the port plates  18  and  28  to the impeller assembly  64 . In the context of this invention, the term fluid will normally but not exclusively refer to a liquid, since a liquid would better fulfill the potential efficiency of the invention. Referring to FIG. 5, there is shown schematically the cam ring  22 , the rotor  20  positioned within the cam ring, and the vanes  24 . It will be seen that gaps  80  are present between inner surface  69  of the cam ring  22  and the outer surface  78  of the rotor, these gaps varying in width about the circumference of the rotor. As the rotor  20  rotates, each of the vanes  24  tends to be displaced outwardly from its respective slots  82  by centrifugal force, so that the outer surfaces  90  of the vanes slidingly contact the inner surface  74  of the cam ring  22 . 
     FIG. 5 shows in outline the position of the ports  104 ,  105  in the first and second port plates  18  and  28 , respectively. Although the rotor  20  may equally well be driven in either direction, the explanation which follows will assume that the rotation is counter-clockwise as viewed in FIG.  5 . It will be seen that the ports  104 ,  105  of the first port plate  18  and the second port plate  28  are staggered by 90° when viewed along the axis  46 . 
     Considering first in FIG. 5 the vane  24  in position  120 , as the rotor rotates counter-clockwise, fluid is pushed ahead of this vane. Because of the widening gap between the rotor  20  and the cam ring  22 , each given quantity of fluid is impelled into a larger volume than it previously occupied. Since the fluid does not expand to fill such additional volume, the additional volume is filled with incoming liquid, which enters through port  105  in the second port plate  28  from an inlet chamber  121 . Considering now the vane in position  122 , the volume is still increasing ahead of this vane as the rotor rotates counterclockwise, and the rotation of the vane in this position continues to cause the admission of fluid into gap  80 . Position  124  is essentially a dwell point, where the available volume is at a maximum and therefore there is neither an increase nor decrease of fluid. Thus, the portion between positions  120  and  124  is a fluid inlet region. By contrast, from position  124  through  126  and up to position  128 , there is a region of decreasing volume, from which an incompressible fluid is necessarily expelled through port  105  in the second port plate  28  into an outlet chamber  129 . The position  128  has minimum available volume; just as with the region of maximum volume, the available volume neither increases nor decreases, whereby position  128  is essentially another dwell point. Thus, the portion between positions  124  and  128  is a fluid discharge region. 
     Once a given vane  24  passes position  128  it begins to repeat the pumping cycle in a fashion equivalent to position  120 ; similarly, positions  130 ,  132  and  134  are equivalent to positions  122 ,  124  and  126 , respectively. In other words, for every revolution of the rotor, a given vane  24  goes through two pumping cycles. Therefore, there are two diametrically opposed inlet regions and two diametrically opposed discharge regions, the inlet and outlet regions being radially positioned at 90° from one another. The profile of the cam ring opening  68  is defined as a high power polynomial curve, which is selected to reduce both the acceleration and change in acceleration to zero at dwell points. This greatly reduces impact forces and therefore minimizes wear on the cam ring and vanes. 
     For the described counterclockwise rotation of the rotor, FIG. 5 shows the ports  104  of the first port plate  18  are lined up with the inlet regions, and the ports  105  of the second port plate  28  lined up with the discharge regions. The ports are sized and shaped to be most compatible with the flow rates at the regions of optimum inlet and discharge, providing the minimum possible constriction to flow and minimizing frictional energy losses. The use of radially opposed port plates results in a balance of forces on the rotor and thus promotes efficiency in operating the pump. 
     The housing, the drive shaft and the thrust plates are preferably made from stainless steel. Preferably, the drive shaft and the thrust plate are coated with a wear- and corrosion resistant coating, such as tungsten carbide. The vanes and cam ring are preferably made from tungsten carbide or other ceramic material, with tungsten carbide most preferred for the vanes because its high density provides greater centrifugal force than other ceramic materials, thus maintaining better contact with the cam ring. The rotor and the port plates are preferably made from a ceramic material exhibiting good wear resistance and corrosion resistance. The hardness and dimensional stability of an alumina ceramic renders it ideal for hydrodynamic journal bearings. The rotating drive shaft runs directly in the port plate journals; the inclusion of a wear resistant coating such as tungsten carbide on the drive shaft precludes the need for antifriction bearings. Additionally, the drive shaft and its thrust plate bear on the outboard faces of the port plates; such a coating serves to provide a hydrodynamic thrust bearing. Accordingly, the need to include antifriction bearings is eliminated, especially for applications of a water pump of relatively low pressure (i.e., no greater than 100 psi). Overall, the stability of the preferred materials provides resistance to the degradation of pump efficiency over long periods of time, thus reducing maintenance of the pump which is important for applications where the pump is installed in a remote location, such as in space. 
     It is clear that the pump  10  of this invention has a high degree of symmetry. In particular, if the revolution of the rotor  20  is reversed, the fluid flow patterns in the vicinity of the rotor  20  and port plates  18  and  28  are identical except in their direction. Such a reversal merely converts an inlet region to an outlet region and an outlet region to an inlet region, thus reversing the roles of the ports  104 ,  150  in the port plates  18  and  28 , the inlet and outlet chambers  121  and  129 , and the inlet and outlet ports  48  and  50  in the housing  12 . The aforementioned symmetry mandates that the efficiency of the pump is independent of the direction in which it is operated. An exception to this symmetry is in the positioning of the inlet port  48  and outlet port  50  of the housing  12 . Since the openings at these ports are much larger than the fluid clearances at other points in the system, they provide little resistance to flow by comparison, and will therefore have only a negligible effect on pump efficiency. 
     The arrangement of the various reniform recesses  106  and circular recesses in port plates  18  and  28 , and of the radial recesses  114  in the thrust plates  16  and  30 , is such that a film of the fluid being pumped is formed at the interfaces between the stationary port plates  18  and  28 , and the rotating rotor  20  or thrust plates  16  and  30 . This film acts as a lubricant which avoids the need to introduce a separate lubricating liquid which could be a source of contamination. 
     In summary, the combination of high internal symmetry, minimal constriction of fluid flow, control of play and inter-surface clearances, and low- corrosion, low-wear materials provides a long-life self-lubricating pump of high efficiency which operates equally well in either direction. Further, the ceramic material used for some components allows them to have a reduced weight by comparison with metal, which is important in space applications. 
     While the invention has been described with reference to preferred embodiments, it will be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation of material to the teachings of the invention without departing from the scope of the invention. Therefore, it is intended that the invention not be limited to the particular embodiments disclosed as the best mode contemplated for carrying out this invention, but that the invention will include all embodiments falling within the scope and spirit of the appended claims.