US3514232A - Variable displacement turbine-speed hydrostatic pump - Google Patents

Description

May 26, 1970 R. K. MITCHELL ETAL 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet l a 2 a Q m mm mm fw E .9 m 8 AWV/f9/ ROBERT K. MITCHELL JAMES C. SWAIN DAVID L. THOMAS JOHN P WILCOX INVENTORS BYj/L6%,%A1L mu! ATTORNEYS May 26, 1970 VARIABLE Filed Oct. 28, 1968 ZIK 10 Sheets-Sheet 2 Fig. 2

ROBERT K MQTCHELL JAMES C. SWAIN DAVID L. THOMAS JOHN P. WILCOX INVENTORS B%%OZM M ATTDRNEYS May 26, 1970 R. K. MITCHELL ETAL 3,514,232

VARIABLE DISPLACEMENT TURBINE 'SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 5 ZERO DISPLACEMENT STRAIGHT LINE MOVABLE MEMBER I I I2? FULL DISPLACEMENT '3' J ACTUATOR MOTION FD; FULL DISPLACEMENT ROBERT K. MITCHELL JAMES C. SWAIN DAVID L. THOMAS JOHN F? WILCOX INVENTORS svz/myjim m 5 ATTORNEYS y 6, 1970 R. K. MITCHELL ETAL 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 4 Fig. 4

JOHN F. WLCOX u m m M K m a m JAMES C. SWAIN DAVID L THOMAS IN VE NTORS Afibnnevs May 26, 1970 R. K. MITCHELL A 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIG PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 5 ROBERT K MITC JAMES C. A

DAVID L. MAS

JOHN R WILCOX INVENTORS ATTORNEYS May 26, 1970 R. K. MITCHELL L 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 6 2692 &

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! ROBERT KMITCHELL JAMES C. SWAIN E DAVID L. THOMAS JOHN R WILCOX INVENTORS and/1%, Wm m ,{jA mmu ATTORNEYS Fig. 9a Fm @b Mfiy 26, 1970 R. K. MITCHELL 3,

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 7 ROBERT K. MITC L JAMES C. SWAI DAVID L THOMAS JOHN P WILCOX INVENTORS BY)mf,Ww0/nd ATTORNEYS May 26, 1970 M T H L ETAL 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed Oct. 28, 1968 10 Sheets-Sheet 8 ATTORNEYS y 1970 R. K. MITCHELL ETAL 3,514,232

VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP ROBERT K. MITCHELL JAMES C. SWAIN DAVID L. THOMAS JOHN P. WILCOX INVENTORS j m lax 0%? ATTORNEYS May 26, 1970 R. K. MITCHELL L VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed 001;. 28, 1968 1O Sheets-$heet 10 ATTORNEYS United States Patent 3,514,232 VARIABLE DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Robert K. Mitchell, Hilliard, and James C. Swain, David L. Thomas, and John P. Wilcox, Columbus, Ohio, assignors to The Battelle Development Corporation, Columbus, Ohio, a corporation of Delaware Continuation-impart of application Ser. No. 549,679,

May 12, 1966. This application Oct. 28, 1968, Ser.

Int. Cl. F04c 1/02, 5/00, /02

U.S. Cl. 418-27 7 Claims ABSTRACT OF THE DISCLOSURE A sliding-vane rotary pump operating at high rotational speeds and constructed for variable displacement. The pump housing includes two lobes that are enlargeable or reducible to provide variable displacement while maintaining all portions of the cam surface (that surface along which the vane tips track very closely) in tangency. In one embodiment movable lap spaces are connected by a plurality of bridges across the pump ports to nonmovable lap spaces. In a second embodiment the cam surface is constructed of a deformable ring or cylinder that is controllably deformed into circular or elliptical shapes to provide the variable displacement feature. The ports preferably are positioned along the cam surface rather than in the end plates and the ratio of rotor length to rotor diameter is considerably greater than in conventional sliding-vane rotary pumps. The vane tips are pivotally mounted and supported by a hydrodynamic film of oil between the vane tip and vane-tip track which is desirable for the high rotational speeds at normal operating conditions. The tip, vane, and vane slot in the rotor are constructed to react to and control various forces acting on the rotating vanes as centrifugal force, pressures in front of and behind the vane, socket pressures, bending moments, tipping forces, hydrostatic pressures, etc.

CROSS REFERENCE TO RELATED APPLICATION This application is a continuation-impart of our copending application entitled Variable Displacement Turbine-Speed Hydrostatic Pump filed May 12, 1966, Ser. No. 549,679, new Pat. No. 3,407,742.

BACKGROUND OF THE INVENTION Field of the invention This invention relates to a variable-displacement sliding-vane pump for operation at high rotational speeds. More particularly, it concerns a sliding-vane rotary pump operable at high turbine speeds and the invention is especially useful as an element of a transmission for a turbine-powered vehicle.

One of the difficult problems associated with adopting a high-powered turbine to drive a wheeled vehicle is the efficient utilization and control of the high rotational turbine-shaft speeds that are usually in excess of 20,000 r.p.m. The speed of a high-power turbine is relatively unresponsive to load and fuel changes because of the large inertia of the rotor. This speed is much higher than can be used directly at the wheels; therefore, a speed reducing or power conversion mechanism is necessary. For a responsive vehicle, which requires a wide range of speeds at constant power, the speed reducing mechanism must be an efficient variable speed transmission. A hydrostatic transmission, having variable displacement and infinite variability across its range, is a desirable type of transmission for a constant-speed turbine. This invention provides a pump that supplies these desired features. Al-

3,514,232 Patented May 26, 1970 ice though the pump described herein has other applications, it will be discussed mainly with respect to its use with a turbine and the problems associated therewith.

Description of the prior art Numerous sliding-vane rotary pumps have been suggested in the past. The sliding-vane rotary pump includes a rotor surrounded by a housing that may be circularshaped and eccentric to the rotor, or the housing may be elliptical in shape having the two ends of the ellipse serving as lobular pumping chambers. The rotor includes a plurality of slots or sockets with a vane (sometimes called a piston) slidably mounted in each slot. The tip of the vane follows the inner contour of the housing primarily due to the acceleration forces generated by rotation of the rotor. Pumping action occurs within two vanes (or a vane and a point of close proximity between the rotor and the housing), the housing, rotor and the end plates on the housing. The inlet of the pump is generally positioned to supply fluid at a point where the vanes move radially outward and the outlet is positioned at a point where the vanes move radially inward.

There are also a number of sliding-vane rotary pumps that have been constructed to have variable displacement. Two methods of varying the displacement appear most frequently. In one method, the eccentricity of a circular housing and the rotor are varied with respect to one another. The other method involves varying the size or shape of the pump housing by various means. None of these prior pumps are practical for high-speed operation.

SUMMARY OF THE INVENTION The apparatus of this invention provides a two-lobed sliding-vane rotary pump wherein the lobe spaces are enlargeable and reducible or change their shape to provide variable displacement. 'In one embodiment the construction of the cam surface or vane-tip track includes two fixed surfaces (lap spaces), two movable surfaces (lap spaces) and a plurality of connecting links or bridges. An important feature that distinguishes the construction of the pump of this embodiment of the invention from prior pumps that appear to be similar in construction is that the cam surfaces of the bridges are maintained in tangency (at the intersection points), throughout all displacement changes, to the surfaces of the fixed and movable lap space where the bridges interconnect them. The passage of vanes from the fixed and movable surfaces to the bridge surfaces, and vice versa, is always accomplished smoothly regardless of the displacement settings. In a second embodiment, the cam surface or vane-tip track is a deformable cylinder. Preferably it is constructed from a plurality of sleeves nested together. The sleeves or laminations are in intimate contact in the bridge areas but are free to slide relative to one another. The sleeves are bonded together at the lap spaces and the lap-space blocks (preferably four) are bonded to the deformable cylinder at the places where the sleeves are bonded together. The cam ring is deformed by pulling or pushing on opposite lap spaces, the lap spaces in the bridge area deflect but remain in intimate contact. The resultant cam ring profile provides a two-lobed pumping action with a smooth continuous cam surface. The feature of keeping the vane-tip track elements tangent at all pump displacements is very important as the high speed exaggerates the slightest bump or discontinuity in the vane-tip track causing skips, bounces, leakage, high wear and early failure of the pump.

Most conventional two-lobed pumps have the ports located in the end plates of the pump housing. In this invention, the ports are preferably positioned so that the flow enters and leaves the pump through spaces in the cylinder that forms the cam surface along the complete length of the rotor. This port construction (through the bridges in one embodiment) has the effect of minimizing flow losses through the pump. Preferably, the ratio of porting space to vane support (bridge surface) is about sixty percent open area and about forty percent bridge area. The selection of this ratio is a compromise between the need to provide a sufiicient vane-tip hydrodynamic bearing area and the desire to minimize flow loss.

This invention provides a sliding-vane rotary pump that is two-lobed, hence pressure balanced. The flow rate through the pump is relatively free from pulsations and the pump has a very small clearance volume at the highpressure, low-flow condition which is important in minimizing compressibility effects at high pressure.

This invention further provides a sliding-vane rotary pump wherein the ratio of rotor length to rotor diameter is considerably greater than in conventional sliding-vane rotary pumps. The small-diameter, long-rotor construction fits in very well with the variable-displacement pressure-balanced construction since the bearing loads and rotor distortions are minimized.

One advantage of the small-diameter, long-axis rotor construction for the pump of this invention is the potentially high overall efficiency. The most significant loss in the pump is the flow pressure drop loss which is essentially a square function of vane tip velocity. At any selected angular speed, the vane tip velocity, and therefore pump flow, is a direct function of diameter. Therefore, because flow through the pump is in the turbulent area, flow losses are essentially a square function of rotor diameter. Vane tip friction loss is a function of vane tip load. Both vane tip speed and vane tip load are directly related to the diameter at a given angular speed so that vane tip friction loss is also a square function of the diameter. Other losses, such as bearing loss and viscous drag on the ends of the rotor, are also reduced by small rotor diameter.

Another advantage of the invention is that the long small-diameter rotor of the pump minimizes charging pressure. A charging pressure is required to accelerate the fluid being pumped to vane tip velocity without cavitation occurring. This charging pressure is a square function of vane tip velocity so that a rotor or twice the diameter of another rotor rotating at the same angular speed requires four times the charging pressure. Thus, the smaller diameter rotor has a distinct advantage.

In most conventional sliding-vane rotary pumps the vane tip contacts the cam surface or vane tracks in a substantially sliding relationship. The lubrication is in the nature of boundry lubrication. A number of pivoting vane tips have been used in prior vane pumps essentailly to aid in sealing, to provide a larger load carrying surface, or to distribute the wear on the tip more evenly. Usually the conventional vane tip is constructed with an arc radius substantially equal to the radius of the vane track surface.

In this invention, the vane tip does not touch the vane track but rides on a wedge of oil with friction being about ten times less than that of boundry lubrication. This is possible because the pump of this invention has a vane tip configured to form a thin but very stiff hydrodynamic film of oil between the vane tip and the vanetip track. The vane tip is pivotally mounted in the end of the vane and the contact radius of curvature of the vane tip is selected so as to be smaller than the smallest radius existing in the vane-tip track. There are other unique construction features of the vane and pivoting vane-tip that react to and control various other forces acting on the rotating vanes; for example, such forces as centrifugal force, pressures in front of and behind the vane, socket pressures, bending moments, tipping forces, hydrostatic pressures, etc.

An object of this invention is to provide a sliding-vane rotary pump that is capable of variable displacement and that is relatively simple in construction.

Another object of this invention is to provide a variable-displacement pump that has a smooth, continuous displacement variation.

Another object of this invention is to provide a pump with both hydraulic pressure and dynamic balance to reduce bearing loads and distortion of rotating parts.

Another object of this invention is to provide a variabledisplacement, sliding-vane rotary pump with reduced motion of internal parts and reduced leakage area at reduced displacement.

Still another object of this invention is to provide a variable-displacement, sliding-vane rotary pump especially capable of operating at high rotational speeds.

Still another object of this invention is to provide variable-displacement, sliding-vane rotary pump with smooth flow paths to minimize losses.

Still another object of this invention is to provide a high-capacity, variable-displacement, sliding-vane rotary pump with a high length-to-diameter ratio that minimizes losses and improves efficiency.

Still another object of this invention is to provide a vane-tip construction that adapts to being supported by a hydrodynamic oil film.

Still another object of this invention is to provide a vane construction that minimizes the various loads imposed on the vane and the vane tips.

Still other objects and advantages of this invention will be apparent from the description that follows, the drawings and the appended claims.

' BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a sectional elevational view of one embodiment of the pump taken along the line 1-1 of FIG. 2;

FIG. 2 is a sectional elevational view of the pump taken along the line 2-2 of FIG. 1;

FIGS. 3a and 3b are diagrams illustrating the relationship of the various cam track elements in the embodiment constructed with bridges;

FIG. 4 is a sectional view of the bridges and a port;

FIG. 5 is a perspective view of two of the bridges and a portion of an associated fixed cam track member;

FIG. 6 is a sectional view of the apparatus for ensuring equalization of the displacement of the two lobes of the pump chamber;

FIG. 7 is an enlarged sectional view of a portion of the rotor, a vane, and a portion of the cam surface;

FIGS. 8a and 8b are diagrams (enlarged) of a vane tip showing the various forces that act thereon;

FIGS. 9a and 9b are diagrams (enlarged) of a vane showing the various forces that act thereon;

FIG. 10 is a perspective view of the deformable cam ring and the lap space blocks, separated from the pump;

I FIG. 11 is a sectional elevational view of a second embodiment of the pump taken along the lines 1111 of FIG. 12;

FIG. 12 is a sectional elevational view of the pump taken along the line 1212 of FIG. 11; and

FIG. 13 is a sectional View showing hydraulic activation of the lap space blocks, using a deformable cam rrng.

DESCRIPTION OF THE PREFERRED EMBODIMENTS In the drawings, the same reference numerals are applied to identically parts in all embodiments and such identically numbered parts are substantially identical in structure, function, and operation, Therefore, to eliminate confusing duplication, these parts, their interrelationship and their function will be described only in conjunction with a single embodiment, such description applying to all embodiments where these parts appear.

Referring to FIGS. 1 and 2, the pump 21 has a pump housing 23 and a manifold section 25 attached at one end. A housing flange 27 and a manifold flange 29 are held together by suitable means such as bolts 31-31.

The contacting surfaces of pump housing 23 and a manifold section 25 are provided with sealing means; preferably O-rings 33-33, that seal around the fixed end plate 37 and also around the outlet passages 39-39 and inlet passages 41-41 that communicate between housing 23 and manifold 25. The fixed end plate is an integral part of the manifold section 2.3 and closes off one end of the variable displacement pump chamber 42.

The opposite end of the pump housing 23 is provided with a retainer 43 held onto the housing 23 by a plurality of bolts 45-45. The retainer 43 has an outwardly extending portion or pump mounting flange 47. A hearing housing 49 fits inside a central opening 50 of the retainer 43 and is attached to the retainer 43 by bolts 51-51. An O-ring 53 seals between the retainer 43 and pump housing 23 and an O-ring 55 seals between the retainer 43 and bearing housing 49.

Pumping energy is transmitted through an input shaft 57, to the rotor 59 and then to each individual vane 61. The input shaft 59 has a spline 63 to receive energy from a power source (such as a turbine, not shown) and a spline 65 engages the rotor 59 at about the midpoint section along the longitudinal axis of rotor 59. The preferred attachment arrangement aids in reducing torsional stresses and deflection in the rotor 59 that would occur if the shaft rotational force were applied to either end of the rotor 59. Such a condition could also exist if the input shaft 57 were arranged for engagement along the length of rotor 59 and only one point or section of the engaging surfaces actually transferred the rotative energy. The rotor 59 is radially positioned and supported by a bearing 67 positioned in the fixed end plate 37 and a sleeve and thrust bearing 69 positioned in a pressureloaded end plate 71. The pressure-loaded end plate 71 and fixed end plate 37 axially position the rotor 59. A collar 73 is threadedly engaged with the input shaft 57 and serves to position and clamp the rotating member 75 of a mechanical face seal 77. The rotating member 75 is clamped against an annular shoulder 79 on the shaft 57. An annular projection 81 on the collar 73 also positions the input shaft 57 axially against sleeve and thrust bearing 83 so that the face seal 77 functions. The collar 73 receives the thrust due to the pump charging pressure which tends to force the input shaft 57 outward. The input shaft 57 is positioned radially by the sleeve and thrust bearing 83, for which the collar 73 is the journal, and by the rotor 59. The fixed member 85 of the face seal 77 is attached to the bearing housing 48. An O-ring seal 87 is provided between rotating face seal member 75 and the input shaft 57., and suitable retaining and seal means 89 is provided for sealing between fixed member 85 and the retainer 49. The mechanical face seal 77 seals the pump charging pressure.

As the rotor 59 turns, the vanes 61 are guided and caused to stroke by the cam surface 90. The cam surface 90 is comprised of lap spaces 91 and 93 of fixed members 95 and 97, the surfaces 99-99 of a plurality of independent bridges 101-101 and lap spaces 103 and 105 of movable members 107 and 109 (FIGS. 2, 3a, 3b, and 4). As seen in FIG. 4 the bridges 101-101 are not closely stacked, but are preferably separated leaving about fifty or sixty percent of the area between the lap spaces 91, 93, 103, and 105 open. It is through these openings that the fluid enters and leaves the pump chamber 42. A plurality of flow passage relief spaces 111-111 are provided on each of the members 95 and 97 (FIGS. 2 and 4). These provide the last exists and entrances for pressurized fluid to exit from in front of a vane 61 on the outlet side and for fluid to first enter behind a vane 61 on the inlet side.

The movable member 107 and 109 are movable in a radial direction, with respect to the rotor 59 to vary the displacement or size of pump chamber 42. The bridges 1111-1111 are positioned by tangs 113-113 on members 95, 97, 107, and 109 fitting closely into slots 115-115 in the bridges 101-101. The bridges 101-101 are positioned axially by their ends 20-120 fitting closely into slots 117-117 on the members 95, 97, 107, and 109. Since each bridge 101 is subjected to the operation of two nonparallel tangs 113-113, its position is uniquely determined for each displacement setting. Portions of the lap space parts 91, 93, 103, and 105 overlap the sides of the bridges 101-101 and serve to carry the vanes 61-61 across the gaps in the cam surface 90 which are formed as the bridges 101-101 slide away from one lap space or another. The cam surfaces on these extended portions 119-119 of the lap space (95, 97, 105, and 107) are planar so that they appear as straight lines (FIGS. 3a and 312). Since these lines (119-119) are parallel to the tanks 113-113, the corners 121-121 of the bridges 101-101 are constrained to be on the lines (119-119). And since in addition, the cam surfaces 99-99 of the bridges 101-101 are constructed to be tangent to the hnes (119-119) at the corner points 121-121, the passage of the vanes 61-61 from lap space to bridge and from bridge to lap space is always made smoothly, irrespective of the displacement settings.

The double-tang construction for bridge positioning has distinct advantages over other methods wherein the bridges are pivoted with respect to one or both of the lap spaces. In other constructions the rotation of the bridges with respect to the lap spaces leads to unfavorable conditions as the vanes pass from one to the other. If sharp corners (against which the vanes would impact) are to be avoided in such construction, when the displacement setting is at one extreme, then a reverse curvature in the cam surface exists when the displacement setting 1s at the opposite extreme. The vanes come out of contact with the cam surface at this point and must be loaded to prevent such skips by some means other than hte1r own centrifugal force. This loading is not only difficult to achieve, but it also increases the maximum tip pressures imposed on the vanes as they cross adjacent areas of the cam surface that have high positive curvature.

FIGS. 3a and 3b show the relationship of a fixed mem' ber 95, a bridge 101, and a movable member 109 at full and zero displacements. The position of the pump elements in FIGI. 3a is designated as zero displacement because with this position of the cam track sections, there is no output flow from the pump. The radii of the curved cam track sections are different and, as shown in FIG. 3a, there are straight sections 119-119 that are a portion of the cam track 90 at the zero displacement position. Obviously the circular cross section of the curved cylinder that comprises the rotor 59 will not be tightly encompassed by the cam track 90, nor will the minute annular space (at zero displacement) between cam track 90 and outer surface of rotor 59 be of a constant width. Actually then, the vanes 61 stroke in and out of their position in the rotor 59 to some degree even at zero displacement. In the embodiment shown in FIG. 3a, the small stroking movement occurs as the vanes 61 cross the bridges 101-101 and adjacent straight portions 119- 119. In FIGS. 3a and 3b the radii of the rotor 59 and lap pspaces (91, 93, 103, and 105) are substantially equal. Their axes coincide along a central axis 123 (shown at a point in FIGS. 30! and 31)) when the pump is at zero displacement. However, other arrangements are also workable and it is sometimes preferred to select unequal radii as discussed subsequently. The radii of the bridge surfaces 99-99 is preferably smaller than the radii of the rotor 59 and lap spaces 91, 93, 103, and 105. The axes 125-125 of the bridge surfaces 99-99 are positioned at about equidistance points around the central axis 123 when the pump is at zero displacement (one of which is represented as a point 125 in FIGS. 3a and 3b). A line 127, perpendicular to the straight section 119 at the point where lap space 91 changes from curved to straight section 119, passes through the central axis 123. Another line 129, perpendicular to straight section 119 at the point where lap space 105 changes from curved to straight section 119, also passes through the central axis 123. Lines 131 and 133, perpendicular to straight sections 119-119 at the corner points 121-121 of the bridge 101, intersect at the axis 125, The above intersections indicate tangency of all elements of the cam track 90 with the straight sections 119-419 being tangent to the curved elements 91, 99, and 105, and this, of course, is similarly true for the other elements of the cam track 90 that are not shown in FIGS. 3:11 and 3b.

In FIG. 3b, the movable member 109 has moved the lap space 105 outwardly (to full displacement) as indicated by the space between lap space axis 135 and central axis 123. The perpendicular line 127 still intersects central axis 123 and perpendicular line 131 still intersects axis 125. Perpendicular lines 129 and 133 coincide and intersect both axis 125 and axis 135. As lap space 105 moves outwardly, bridge 101 moves away from fixed member 95, but moves toward movable member 109. This is due to the tang and slot (113 and 115) arrangements which are positioned with their sides parallehto the straight sections 119-119. The slot 115 bearing against the side of tang 113 on fixed member 95 forces bridge 101 toward movable member 109 as movable member 109 moves outwardly. Thus, regardless of the POSI- tion of movable members 107 and 109 the surfaces (91, 93, 9999, 103, and 105) of the cam track 90 are mamtained in tangency.

The radii of lap spaces 91, 93, 103, and 105 are shown to be equal (FIGS. 3a and 3b) as one example of the construction. The radii are selected depending on the pump application. It is desirable that lap spaces 91 and 93 have their center of radius at point 123 so that no stroking of the vanes 61-61 relative to the rotor 59 occurs while the vanes are traversing the field or immovable lap spaces 91 and 93. In FIG. 311 it is evident that stroking of the vanes 6161 would occur in the lap space 105 because the center of the lap space arc 135 and the center of the rotor 123 do not coincide. The radius of the lap spaces 105 and 107 should be selected to minimize vane stroking in these lap spaces depending upon the specific application of the pump. In some applications the axes of the movable lap spaces 105 and 107 will not colnclde at the axis 123 at zero displacement (the radii of these lap spaces may be selected to be larger) and the stroking of the vanes 6161 will occur as they cross the movable lap spaces 105-107.

FIG. 2 shows the construction for moving the lap spaces 103 and 105 toward and away from the central axis 123. The apparatus that moves the member 107 is duplicated for moving member 109 so that the apparatus will be described only with respect to its relat1onship to member 107. The movable member 107 is attached to a pressure balance piston 137 by suitable attachment means such as bolts 139139. The balance piston 137 includes a seal slot 141 partially enclosed by a base section 143. The balance piston 137 is reciprocal in a chamber 145 that is closed off by a cylinder section 147. The cylinder section 147 is attached to the pump housing 23 by bolts 149-149. The seal slot 141 is provided with a seal 151 that seals between the balance piston 137 and the outer wall of the chamber 145. Another seal 153 is positioned to seal between the balance piston 137 and the movable element 107. The balance piston 137 is provided to essentially balance the pressure load on movable member 107 resulting from pressure generated in the pump chamber 42 and to reduce the force required to move the member 107 toward zero displacement against the pump pressure. High pressure fluid in the discharge port 155 is communicated to the chamber 145 behind the balance piston 137 through passage 157 and ports '159 159. Thus the force (due to pump pressure) that tends to force the movable member 107 away from the rotor 59 is substantially neutralized. The lap space member 107 must be in intimate contact with the case '23 to minimize leakage from the high pressure in the discharge port 155. Position of the sealing land 160, on movable member 107, is selected so pressure forces always cause intimate contact and some nominal load between the sealing lands and the pump housing 23. The axis of the pressure balance piston 137 must also be positioned so that pressure forces always cause intimate contact and some nominal load between the sealing land 160 and the case 23. In positioning both the seailng land 160 and the axis of pressure balance piston 137 it must be recognized that the area over which pressure acts on lap space surface 103 varies with time because of movement of any vane 61 across the lap space which acts as a sealing member (i.e., the area of pressure application on the lap space surface is varied as the sealing vane 61 moves).

The balance piston 137 is attached to a rod 161 that passes through the wall 163' of cylinder section 147 into cylinder 165. Appropriate seals 167167 are provided around rod 161 to seal cylinder from balance chamber 145. A double acting piston 169 is positioned in cylinder 165 and is attached to rod 161 by a nut 171 that force piston 165 against a shoulder 173 on the rod 161. An -ring 175 is provided around piotsn 169 to seal against the wall of cylinder 165. The cylinder 165- is closed 0 by a cylinder head 177 also with an O-ring seal 179. Ports 181 and 183 are provided at each end of the cylinder 165 to admit fluid under pressure thus actuating the piston 169 and moving member 107. Fluid admitted through port 18 1 moves member 107 toward rotor 59 and zero displacement and fluid admitted through port 183 increases the pump displacement.

In FIG. 1, the pressure loaded end plate 71 is radially positioned by the housing 23. The end plate 71 has an annular shaped extension or annular piston 185 that fits into an annular cavity 187 positioned between the retainer 43 and bearing housing 49, with O-ring seals 189-189 provided on the piston to seal-off the open side of the cavity 187. At least one port 191 passes through the annular piston 18 5 and delivers the outlet pressure of the pump to chamber 187 thus pressure-loading the plate 71 and urging it toward the rotor 59' and against the ends of fixed lap spaces 95 and 97 and of movable lap spaces 10 7 and 109, and against shoulders 192 and 194 on the case 23. The end plate 71 is pressure loaded to maintain a small nominal rotor end clearance and to act as a seal to minimize leakage from the discharge pressure passages 39'-39. Though there must be intimate contact between the end plate 71 and the fixed lap spaces 95 and 97, the movable lap spaces 107 and 109 and the case at shoulders 192 and 194, the loading must be such as to allow stroking of the movable lap spaces members with relative case.

To keep pressure forces acting on the rotor 59 balanced, it is desirable to ensure that movable members 107 and 109 will always be substantially equidistant from the central axis 123 (or rotor 59). By maintaining the movable members 107 and 109 equidistant from central axis 123, the displacement of the two lobes of pump chamber 42 has equal displacement and equal flow losses resulting in pressure balance of the rotor. The construction for maintaining the movable members 107 and 109' equidistant from the central axis 123 at all displacements is partially shown in FIG. 1 and more completely in FIG. 6. The equalizer plate 193 is positioned between the pressure-loaded end plate 71 and the retainer 43 for longitudinal positioning. The equalizer plate 193 is provided with two slots 197-197 with a slider 199 fitted into each slot 197. The slider is provided with a pin 201, and each pin passes through the end plate 71 and engages one of the movable members 107 or 109. The equalizer plate 193 is free to rotate on the bearing surface on end plate 71 so that if one of the movable members, for example 107, moves toward the central axis 123, that movement is transferred through one of the pins 201 to equalizer plate 193 which causes the plate 193 to rotate and of necessity, through engagement of the other movable member 109 with the other pin 201, causes member 10 9 to move toward the central axis 123 an equal amount (the actual engagement of pins 201-201 with movable members 107 and 109 is not shown).

The cam surface 90 includes eight circular arcs (9'1, 93, 103, 105, and four 99-99) connected by straight tangent surfaces (119119) of variable length. Some of the operating problems of an example pump are: (1) rotational speeds of about 22,000 r.p.m.; (2) a pressure range of about 2000 to 8000 psi; (3) absorption of constant power over the range of pressures; (4) some stroking of the vane while the vane is sealing; (5) the vanes must cross over the ports in the bridges; (6) disappearance of centrifugal force and reduction of acceleration forces as the vanes cross the straight sections of the cam surface; and (7) full pump pressure differential across the vane while the vane crosses a lap space. The foregoing conditions require that the vanes 61-61 be of a more specialized construction than the vanes of conventional pumps. The relatively large hydrostatic pressures tend to force the vanes 6161 away from the cam surface 90 and must be balanced. The vane tip is hydrodynamically lubricated, i.e., it is separated from the cam surface 90 by a thin film of oil.

The vane 61 and vane tip 203- mounted in the rotor 59 are shown in the enlarged cross-sectional view of FIG. 7. The vane tip 203 is a type of pivoted slider hearing mounted in a socket 205 of the vane 61. The slider surface 207 is relatively small and this accomplishes two things: (1) the hydrostatic force on the tip 203 is relatively small compared to the hydrodynamic force, thereby allowing the hydrodynamic force to controlthe angle of tilt; and (2) hydrostatic radial vane forces can be 209 is considered to be on the front of the vane, 61.

The pressure in chamber 42a, in front of the vane 61, communicates with the chamber 213 by means of aport 221 and passage 223. The pressure in chamber 42b,in back of the vane 61, communicates with chamber 215 by means of a port 225 and passage 227. The undervane area of step 209 is larger than the undervane area of the vane step 217 because the hydrostatic force on the slider surface 207 (represented by force 249'in FIG. 8a) is greater when outlet pressure occurs in pumping chamber 42a than when outlet pressure occurs in pumping chamber 42b due to the forward tilt of the tip 203. The vane 61 is just wide enough to provide'sufiicient strength in the tip socket 205 to hold the tip 203 against hydrostatic forces. The moment due to tip hydrodynamic and hydrostatic drag is small, compared to the potential hydrodynamic moment. This is due primarily to low drag, but

'also because the distance from the slider surface 207 to the pivot center 229 is selected to be relatively small, at least smaller than the radius of the tip socket 205. The pivot center 22-9 is also positioned toward the rear of the slider surface 207, which is the preferred location, according to hydrodynamic theory. This tends to tilt the tip 203 forward. The tip 203 is also provided with a straight section 231 on the trailing edge and a'slightly larger straight section 233 on the leading edge.

FIGS. 8a and 8b are enlarged diagrams of the tip 203 indicating a typical static-pressure distribution 235 (FIG. 8a) and a typical hydrodynamic pressure distribution 237 (FIG. 8b). FIG. 8a also includes a plurality of arrows that indicate the imposed forces on the tip 203 and FIG. 8]) includes a plurality of arrows that indicate the. reaction forces on the tip 203. In FIG. 8a, the arrow 219 indicates the rotational direction. The profile of the hydrostatic-pressure distribution indicates that the pressure ahead of the tip 203 is greater than the pressure behind the tip 203. Thus, the tip 203 is moving along a section of the cam track where the inlet port is behind and the outlet port ahead of the tip 203. The arrow 239 shows a force, due to pump pressure, acting mainly on the vanetip surface 233 that tends to rotate the tip 203 counterclockwise. The pressure behind the tip 203 results in a force acting mainly on the surface 231 represented by the arrow 241 and tends to rotate the tip 203 clockwise. The arrow 243 represents a hydrostatic pressure from the space in front of the tip 203 that invades the space between the tip 203 and vane 61 and provides a hearing force for the tip 203. The arrow 245 represents the moment due to friction between the vane 61 and tip 203. A frictional drag caused by hydrostatic pressure flow that tends to tile the tip 203 counterclockwise is represented by the arrow 247. The force due to the hydrostatic pressure distribution 235 that tends to turn the tip 203 clockwise is shown by the arrow 249. FIG. 8b shows an arrow 251 representing the force of the vane 61 tending to hold the vane tip 203 in the socket (205). The arrow 253 represents the reaction moment of the vane due to friction. A drag force due to shear rate in the hydrodynamic film is indicated by the arrow 255. The force due to the hydrodynamic pressures distribution 237 represented by the arrow 257, tends to turn the tip 203 counterclockwise (at this particular set of conditions). However, the position and magnitude of the hydrodynamic pressure force (represented by the arrow 257), varies to balance out the other forces and keep the tip 203 stabilized to travel on, and be supported by, a thin wedge of oil. The friction of the hydrodynamically supported vanetip 203 is about ten times less than that of boundary (conventional) lubricated vanes. The radius of curvature of the slider surface 207 is selected to be less than the smallest radius encountered in the cam surface 90. The slider surface 207 is selected to be sufficiently large that a hydrodynamic pressure occurring between the tip surface 207 and cam surface 90 is sufficient to counteract other forces acting on the tip from pump pressures and from the vane 61 that supports the tip 203. The straight sections 231 and 233 On the tip aid in balancing the hydrostatic-pressure forces that affect the tip 203. The hydrodynamic action provides a very stiff bearing since the thickness of the hydrodynamic film varies inversely as the square root of the net vane load thus allowing a wide range of forces to act on the tip 203.

FIGS. 9a and 9b show some of the various forces acting on'the vane 61. FIG. 9a shows the imposed forces acting on the vane. The hydrostatic force, imposed on the slider surface 207, is represented by the arrow 249 acting on the tip 203. Arrows 238438 indicate the hydrostatic pressure ahead of the vane 61 acting on the top and side of the vane 61 and also introduced beneath the sure flow is indicated by the arrow 247. Acting through the vane center of gravity 259 are a Coriolis force (arrow 261), a centrifugal force (arrow 263) and an acceleration force (arrow 265); the Coriolis and centrifugal force are actually due to rotor rotation while the acceleration force is an imposed force that depends on the cam track configuration.

The reaction forces are shown in FIG. 9b. The hydrodynamic force (arrow 257) acts through the slider surface 207 on tip 203. Vane socket reaction forces shown by arrows 267 and 269 are imposed on the vane 61 by the rotor 59. A hydrodynamic drag force is represented .by the arrow 255.

The vanes 6161 are responsible for pumping the fluid in the pump 21 and the vane tips 203 provide the support and sealing necessary for the vanes 61-61 to operate. The hydrodynamic film beneath the vane surface 207 provides lubrication and support to each vane 61 through the tip 203. The steps 209 and 217 and chambers 213 and 215 beneath each vane 61 are means for applying forces to the vane 61 to balance the various other forces that arise due to rotation and the various pressures existing within the pump 21. The construction of the vanes 61-61 and their associated tips 203-203 are necessary to function properly at the high pressures and high rotational speeds required for the pump 21.

FIG. shows another embodiment of a cam ring for a turbine-speed-type vane pump. The cam ring is deformable and is removed from the pump in order to show the structure better. The cam ring 301 includes a deformable cylinder 303 preferably constructed from a plurality of nested thin flexible cylinders 305-305. The cylinders 305-305 are bonded together at four places forming the lap spaces 307-307 which are inflexible while the portion of the cylinder 303 between the lap spaces 307-307 remain flexible. At the same time that the cylinder 305- 305 are bonded together at the lap spaces 307-307, the lap space blocks 309-309 and 311-311 are attached to the cylinder 301. The lap space blocks 309-309 and 311-311 are tapered (for reasons explained subsequently) with the blocks 309-309 each having a small end 313 and a large end 315, while the blocks 311-311 each have a small end 317 (at the opopsite end of the cylinder from small end 313 of each block 309) and a large end 319 (at the opposite end of the cylinder from large end 315 of each block 309). A plurality of slots 321- 321 are cut through the cylinder 303 in each of the flexible portions between the lap spaces 307-307 and provide the openings for the inlet and outlet ports. The interior surface 323 of the cylinder 303 forms the cam ring for the vane tips 203-203.

FIGS. 11 and 12 show a simplified version of an embodiment of the pump using the deformable cam ring 301. The shaft 57, rotor 59, and vanes 61 are similar to the pump shown in FIGS. 1 and 2. The pump housing is divided into a front housing 325, a center housing 327, and a rear housing 329. Flanges 331 and 333 are provided on the central housing for attachment to flanges 335 and 337 on the front and rear housings, respectively, by suitable means such as fasteners 339. The front housing 325 includes a mounting flange 341 with holes 343 for receiving fasteners for mounting purposes. The shaft 57 is supported by a front bearing 344 and also bearingend plate combinations 345-345 at each end of the rotor 59. Appropriate seals 348-348 are positioned the housing parts.

When one force is applied to one set of opposing lap space blocks and an opposite force applied to the other set of lap space blocks ninety degrees away, the cam ring 301 is changed from a circular shape (shown in FIG. 12) to elliptical shape (shown in FIG. 13) to form a variablesized two-lobed pumping chamber. Pumping begins as soon as the cross section of the cam ring 301 changes from circular to elliptical and continues only so long as the shape is elliptical. Deformation of the cam ring 301 to a pumping condition is also possible by just applying a force to one set of opposing lap space blocks and allowing the other set to move to a position resulting from the applied force. FIG. 10 shows the lap space blocks 309309 tapered in one direction while lap space blocks 311-311 are tapered in the opposite direction. A wedge 343 is positioned between each lap space block 309 and a tapered or cam surface 345 of the center housing 327. A wedge 347 (tapered opposite to wedge 343) is positioned between each lap space block 311 and a tapered or cam surface 349 of the center housing 327. Each lap space block 309-309 and 311-311 is attached to one end 351 of a rod 353. The other end 355 of rod 353 has a spring retainer 357 held in place by a nut 359. A spring 361 bears against retainer 357 and a recess 363 in central housing 327. The spring 361 urges the lap space blocks 309- 309 and 311-311 outward against wedges 343-343 and 347-347. A housing 365 is provided to cover each spring 361 and a slot 367 is provided in each wedge (343-347) to allow rod 353 to be connected to the lap space blocks (309-311). A seal 369 is provided on each side of the lap space blocks 309-309 and 311-311.

The wedges 343-343 and 347-347 are connected to control rods 371-371 (one shown only in FIG. 11). The central rod 371 passes through the rear housing 329, where a seal 372 is provided, and through a slot 373 in a plate 375. The rod 371 is restrained from movement (horizontal in the drawing) with respect to the plate by washers 377-377 and nuts 379-379 threaded into the rod 371 and positioned on each side of the plate 375.

The plate 375 has a central opening 378 and a collar 370 that is slideably mounted on a spindle 381. The spindle 381 is attached to the rear housing 329 by suitable means such as fasteners 383 (one shown). The spindle 381 also has a smaller threaded end portion 385 threadedly engaged by a sleeve 387. The sleeve 387 is threaded on its outer surface and has an eternal lip 389 that engages an internal lip 391 on collar 380. A control nut 393 threaded onto the external threads of the sleeve 387 and is moved until it engages end of collar 380. A lock nut 395 holds control nut 393 against collar 380 so that rotation of control nut 393 rotates sleeve 387 on threaded spindle 385.

Rotation of control nut 393 in one direction (for example, clockwise) moves the plate 375 toward rear housing 329 by the nut 393 exerting a force against collar lip 391, and rotation in the opposite direction moves the plate 375 away from the housing by the sleeve lip 389 pulling on the collar lip 391. The cam ring 301 is shown in the neutral position in FIGS. 11 and 12. When the plate 375 moves away from the rear housing 329, rod 371 pulls the wedge 343 to the right (as shown in FIG. 11) and spring 361 through rod 353 pulls block 309 outwardly. The large end of wedge 343 moves into the relief space 397 provided in rear housing 329. The other wedges (343 and 347-347) also move except that wedges 347-347 have their small and large ends positioned end for end with respect to wedges 343-343 and the lap space blocks 311-311 move inwardly due to the wedge movement. Thus, cam ring 301 assumes an elliptical shape with its long axis vertical (as the pump is shown in FIG. 12). When the cam ring 301 is in the shape of a circle, the fluid between vanes 61-61 is merely pushed around the inside of the cam ring 301 with very litle or no pumping action. With the cam ring in the elliptical shape (as it is in FIG. 13), pumping occurs as a function of the difference between the major and minor diameters of the ellipse. The vanes seal solid cam ring portions between the rows of ports 321-321 pumping" is accomplished by successive pairs of the most outwardly extending vanes as they move across the pumping lap spaces.

The undervance areas participate in the pumping to make up for the vane volume. The amount of pumping is determined by the deformation of cam ring 301 which is, as shown, infinitely variable.

When the pump is used in a vehicle transmission, it is advantageous to be able to reverse without the need of gears or other auxiliary devices. When plate 375 is moved toward rear housing 329, wedge 343 moves to the left (as shown in FIG. 11) with the small end moving into relief space 399. This forces the lap space blocks 309-309 inwardly while blocks 3 11-311 are moved outwardly by rods 353-353. The cam ring 301 again assumes an elliptical shape, but this time the long axis of the ellipse is horizontal (as the pump is shown in FIG. 12). The inlet 41 now becomes the outlet and the outlet 39 now becomes the inlet as the flow is reversed.

FIG. 13 shows another construction for moving the lap spaces 307 to deform the cam ring 301 to a selected position. The lap space blocks 309'-309 are each urged outwardly by spring 361 acting through rod 353 as are blocks 309-309 in the embodiment shown in FIGS. 11 and 12 except the blocks 309'--309 are not tapered. The force for moving each block 309 inwardly against the force of spring 361 is provided by control fluid passing through an annular opening 401 around rod 353. The control fluid is supplied through control line 403 into housing 365. The pressure of the control fluid is varied (by means not shown) to move the lap space blocks 309-309 to the desired position and maintain position after the position has been selected.

Lap space blocks 311'-311 are modified to include a stepped end with a short end 405 and a long end 407 fitting into and dividing a chamber 409 into a high pressure compartment 411 and a low pressure compartment 413. A duct 415 from the outlet 39 communicates with a small orifice 417 that connects duct 415 and high pressure compartment 411. A duct 419 connects low pressure compartment 413 and inlet 41. Seals 421421 on the long end 407 separate compartments 411 and 413.

When displacement of cam ring 301 from circular shape to elliptical is desired, the control pressure is reduced and lap space blocks 309'-309, actuating as a piston-in-a-cylinder-type actuator, are pulled outward and pushed outward by hydrostatic pressure in the pumping chamber acting on cam ring surface 323. This exerts a force within the cam ring 301 so that lap space blocks 311'311' move inwardly. There is also a pressure force acting through duct 415, orifice 417 to chamber 411 from the outlet 39. The pressure in chamber 411 is used substantially as a balancing force against the pressure within the cam ring 301. For this reason, it is preferred that only a portion of the available back surface of the lap space block 311 be subjected to the outlet pressure as, for example, the step 405.

The small orifices (401 and 417) that supply pressure to the lap space block chamber are preferred to minimize lap space block (309'-311') radial vibration amplitude because of the variation in the pressure force acting outwardly due to the motion of vanes across the lap spaces 307-307. The orifices are selected to be large enough to allow rapid change in the cam ring shape but as small as possible to minimize lap space block vibration amplitude.

Although the lap spaces 307307 are shown as sharply defined in the drawings, they are not so sharp in the actual pump construction. This is due to the preferred methods of constructing the cam :ring. One of the simplest methods is to coil is flat sheet of metal to produce the laminations and then attach the lap space blocks 309 309 and 311-311 by diffusion bonding, brazing, or electron beam welding. This produces the substantially solid lap spaces 307-307 with the thin flexible cylinders 305305 nested together.

It will be understood of course that while the forms of the invention herein shown and described constitute the preferred embodiments of the invention, it is not intended herein to illustrate all of the possible and equivalent forms or ramifications of the invention. It will also be understood that the words used are words of descrip tion rather than of limitation and that various changes such as change in shape, relative size, and arrangement of parts may be substituted without departing from the spirit or scope of the invention herein disclosed.

We claim:

1. A pump chamber for a variable-displacement, vane,

(a) a plurality of nested flexible tubes forming a multiwalled tubular chamber;

(b) four rows of ports through said tubular chamber, each said row centered along a line parallel to the central axis of said tubular chamber and each row spaced ninety degrees from the adjacent rows; and

() means for applying a variable force on opposite sides of said tubular chamber each along a line substantially midway between a pair of port rows to deform said tubular chamber to various crosssectional shapes between circular and elliptical thereby providing a variable-size, two-lobed pumping chamber.

2. A pump chamber for a variable-displacement, vane,

rotary pump, comprising:

(a) a plurality of nested flexible cylinders arranged to form a multi-walled cylindrical chamber, said nested flexible cylinders being fused in an area along four lines each line parallel to the central axis and spaced ninety degrees from one another;

(b) a row of ports extending along a line substantially midway between each of the fused areas; and

(c) means for applying a variable force on each pair of opposing fused areas to vary said cylindrical chamber between a circular and elliptical cross-section thereby providing a variable-capacity two-lobed pumping chamber.

3. A pumping chamber and cam track for the pump vanes of a variable-displacement, vane, rotary pump comprising:

(a) a first pair of opposing arcuate cam track elements and a second pair of opposing arcuate cam track elements arranged ninety degrees from each other substantially on the loci of a circle;

(b) a plurality of nested flexible arcuate laminations connecting each of said first pair of opposing arcuate cam track elements to each of said second pair of opposing arcuate cam track elements to form a substantially cylindrical pumping chamber with four flexible laminated sections; and

(c) means for moving said first and second pair of arcuate cam track elements inwardly and outwardly with respect to the central axis of said pumping chamber to vary the cross-section of said pumping chamber between cylindrical and elliptical thereby providing a variable-capacity two-lobed pumping chamber.

4. A pumping chamber for a variable-displacement,

vane, rotary pump comprising:

(a) a plurality of nested flexible tubes arranged to form a multi-walled tubular chamber said nested flexible tubes being fused together along the length of the tubes at four places spaced substantially equidistant from each other to form foor relatively unfiexible lap spaces;

(b) a plurality of ports arranged in a row between each of said lap spaces;

(c) means for moving one opposing pair of said lap spaces in one direction relative to the centered axis of said tube while moving the other pair of opposing lap spaces in the opposite direction to vary the crosssection of said tube between circular and elliptical thereby providing a variable-capacity two-lobed pumping chamber.

5. A pumping chamber for a variable-displacement, vane, rotary pump in accordance with claim 4 wherein said means for moving said lap spaces includes a block attached to each said lap space, resilient means attached to the blocks for urging said lap spaces away from said central axis and wedges bearing on said blocks and slideable along a line substantially parallel to said central axis for overcoming said resilient means to move said blocks towards said central axis.

6. A pumping chamber for a variable-displacement, vane, rotary pump in accordance with claim 4 wherein said means for moving said lap spaces includes piston and cylinder means attached to each lap space for movement hydraulically.

7. A variable-displacement, vane, rotary pump, comprising, in combination:

(a) a central pump housing having a central bore;

(b) inlet and outlet manifolds in said pump housing and communicating with said central bore;

(0) a pump chamber mounted in said bore, the inner surface of said pump chamber being a cam track for the pump vanes, said pump chamber being a deformable laminated cylinder and variable from circular to substantially elliptical cross-section to provide a variable-displacement two-lobed pump chamber, said laminations being fused substantially along the entire length of the cylinder in four places equidistantly spaced from each other around the circumference to form four lap spaces;

(d) a rotor mounted in said pump chamber having a plurality of vane sockets;

(e) a plurality of vanes each slideably mounted in a vane socket of said rotor, each said vane having a pivotally mounted tip constructed to produce a hydro- 16 pump chamber said ports being arranged in four rows with one row between each lap space; and (h) means attached to said lap spaces for moving said lap spaces inwardly and outwardly to vary the pump displacement.

References Cited UNITED STATES PATENTS 1,190,139 7/1916 Ford.

2,149,337 3/1939 Deming 103136 3,099,964 8/1963 Eickmann 1()3136 3,187,677 6 1965 Stieber.

3,407,742 10/1968 Mitchell et a1.

dynamic film between the vane tipand the cam track 15 WILLIAM FREEH: Primary Examiner when said rotor is rotating;

(f) an end plate closing each end of said pump chamher; 7

(g) inlet and outlet ports through the wall of said W. I. GOODLIN, Assistant Examiner US. Cl. X.R.

- 2 3 UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION Patent No. 232 Dated May 26, 1970 Inventor(s) Robert K. Mitchell et al.

It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

In the Specification:

Column 3 line 44, after "rotor", "or" should read Column 5, line 48, after "housing", "48 should read Column 6, line 2, after "ends", "'20" should read 120 line 36, "hteir" should read their line 44, after "in", "FIGI. should read FIG.

In the Claims:

Claim 1, line 1, after "vane, insert rotary pump,

comprising:

ollmfii) AM Q-EALED OCTG E l Attest:

EdwardMFletchmIr.

' mm B. W, .18. LAnestmg 0mm Oomissioner of Patents

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