Information
-
Patent Grant
-
6503064
-
Patent Number
6,503,064
-
Date Filed
Thursday, July 15, 199924 years ago
-
Date Issued
Tuesday, January 7, 200321 years ago
-
Inventors
-
Original Assignees
-
Examiners
Agents
- Salai, Esq.; Stephen B.
- Shaw, Esq.; Brian B.
- Harter, Secrest & Emery LLP
-
CPC
-
US Classifications
Field of Search
US
- 417 4103
- 417 326
- 418 259
- 418 131
- 418 133
- 418 102
-
International Classifications
-
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.
Claims
- 1. A self-lubricating, bi-directional vane pump comprising:a pump housing including a first port and a second port; a reversible motor connected to a drive shaft such that the drive shaft is reversibly rotatable about its axis; a stationary cam ring mounted in the pump housing, the cam ring having inner elliptical camming surface; a rotor concentrically disposed within the cam ring and connected to the drive shaft to rotate therewith, the rotor having an outer cylindrical surface with radial slots therein, wherein the cam ring and the rotor are each concentrically aligned with the drive shaft axis, and the rotor is disposed in the cam ring such that two diametrically opposed, symmetrical gaps are present between the rotor outer surface and the cam ring camming surface; a plurality of vanes slidingly disposed in the radial slots of the rotor, such that during operation of the pump, the vanes slide outwardly in the radial slots and maintain contact with the cam ring camming surface; first and second port plates disposed in the housing on each side of the rotor and cam ring, wherein each of said port plates comprises: two diametrically opposed ports, two diametrically opposed recesses formed in surfaces adjacent the rotor and cam ring, and a central opening on which the drive draft is journaled directly; wherein said first and second port plates are arranged with respect to the cam ring such that the two diametrically opposed recesses of the first port plate are aligned with the two diametrically opposed ports of the second port plate, the two diametrically opposed recesses of the second port plate are aligned with the two diametrically opposed ports of the first port plate, and the two diametrically opposed ports of the first port plate are offset by 90° from the two diametrically opposed ports of the second port plate, the port plates being in fluid connection with the pump housing first and second ports, such that one of the pump housing ports functions as an inlet port when the pump is operated in a first direction, and the other of the pump housing ports functions as an outlet port when the pump is operated in a second direction, the pump operating with comparable efficiency in both directions; and wherein axial position of the drive shaft is controlled to permit a pumping fluid to lubricate component parts on which the drive shaft is journaled; the pump further comprising a first thrust plate adjacent the first port plate, and a second thrust plate adjacent the second port plate, the first and second thrust plates being centrally attached to the drive shaft to rotate therewith and having diameter sized to avoid obstructing the ports in the first and second port plates, respectively.
- 2. The pump of claim 1, wherein the first thrust place is integrally formed with the drive shaft.
- 3. The pump of claim 1, wherein a surface of the first thrust plate adjacent the first port plate includes a plurality of radially disposed recesses connected to a circular recess surrounding its juncture with the drive shaft, and a surface of the second thrust plate adjacent the second port plate includes a plurality of radially disposed recesses connected to a circular recess around its juncture with the drive shaft.
- 4. The pump of claim 1, wherein a shim is interposed between an end of the drive shaft and an inner shoulder of the second thrust plate, said shim controlling axial movement of the drive shaft.
- 5. The pump of claim 4, wherein axial movement of the drive shaft is controlled to permit formation of a lubricating film of pumping fluid between surfaces of the port plates and adjacent surfaces of the port plates.
- 6. The pump of claim 5 wherein the pumping fluid has low viscosity.
US Referenced Citations (26)
Foreign Referenced Citations (2)
Number |
Date |
Country |
59168291 |
Sep 1984 |
JP |
59180088 |
Oct 1984 |
JP |