US3078808A

US3078808A - Hydraulic balancing system for rotary positive displacement fluid handling devices

Info

Publication number: US3078808A
Application number: US749133A
Authority: US
Inventors: Jr James O Byers
Original assignee: Bendix Corp
Current assignee: Bendix Corp
Priority date: 1958-07-17
Filing date: 1958-07-17
Publication date: 1963-02-26
Anticipated expiration: 1980-02-26

Description

1963 J. o. BYERS, JR 3,078,808

HYDRAULIC BALANCING SYSTEM FOR ROT Y POSITIVE DISPLACEMENT FLUID HANDLING VICES Filed July 17, 1958 4 Sheets-Sheet 1 TIE. 1

INVENTOR.

Jag/5 0. Made A 7'7'ORA/EY Feb. 26, 1963 J. 0. BY s, JR 3,078,808

HYDRA L C BALANCING SYS. F ROTARY POSITIVE SPLACEMENT FLUID HAN ING DEVICES Filed July 17, 1958 4 Sheets-Sheet 2 INVENTOR.

M245 0. Emma/A. WM Q 4 TTOR/Vf BYERS, JR 3,078,808

Feb. 26, 1963 J, 0

HYDRAULIC BALANCING SYSTEM FOR ROTARY POSITIVE DISPLACEMENT FLUID HANDLING DEVICES 4 Sheets-Sheet 5 Filed July 17, 1958 O O Illa-'5 INVENTOR. dn/vzs a flyzagde 52 BY ATTOAA/E) 3,078,808 IVE Feb. 26, 1963 J. o BYERS, HYDRAULIC BALANCING SYSTEM F 8 DISPLACEMENT FLUID HAN JR OR ROTARY POSIT DLING DEVICES heats-Sheet 4 Filed July 1'7, 195

JA/VZJ 0. 53 2-3 23.

Patented Feb. 26, 1963 3,073,868 HYDRAULIC BALANCENG SYSTEM FOR RUTARY POSITIVE DISPLACEMENT FLUKE) HANDLING DEVICES James 0. Byers, Ilia, St. Joseph, Mich, assignor to The Bendix Corporation, a corporation of Delaware Filed July 17, 1958, Ser. No. 749,133 Claims. (Cl. 103-161) The present invention relates to a balancing piston arrangement for rotating positive displacement fluid handling devices having a plurality of chambers, individual ones of which are alternately pressurized and depressurized; and more particularly to a pressurizing arrangement for a hydraulic balancing piston in radial piston pumps and motors.

An object of the present invention is the provision of a new and improved system for supplying pressure to a balancing piston from the individual fluid pressurizing chambers of a rotary positive displacement device, individual chambers of which are periodically pressurized and then depressurized, and in which pressurizing system for the balancing piston, no moving parts are utilized.

Another object of the present invention is the provision of a new and improved pump and/or motor having a plurality of radial piston chambers therein which are alternately pressurized and depressiurized, and each of which chambers communicate with the balancing piston through substantially identically constructed flow restrictions or orifices.

A further object of the invention is the provision of a new and improved pump and/or motor of the above described type wherein the flow restrictions produce laminar flow.

A still further object of the present invention is the provision of a new and improved pump and/or motor of the above described type having an odd number of fluid pressurizing chambers greater than five.

The invention resides in certain constructions and combinations and arrangements of part-s; and further objects and advantages will become apparent to those skilled in the art to which the invention relates from the following description of the preferred embodiment described with reference to the accompanying drawings forming a part of the specification, and in which:

FIGURE 1 is a cross sectional view taken approximately on the line 11 of FIGURE 3 of a hydraulic pump embodying principles of the present invention;

FIGURE 2 is a cross sectional view taken approximately on the line 22 of FIGURE 1, with the rotor in a different angular position from that shown in FIGURE 1 to better illustrate the Valving cycle of the pump;

FIGURE 3 is an end view of the pump shown in FIG- URE 1 having parts broken away and sectioned approximately upon the line 3--3 of FIGURE 1;

FIGURE 4 is an end view of a porting plate shown in FIGURES 1 and 2;

FIGURE 5 is a cross sectional view taken approximately on the line 55 of FIGURE 4;

FIGURE 6 is a cross sectional view taken approximately on the line 6-6 of FIGURE 1; and

FIGURE 7 is an isometric view of one of the orifice forming plugs shown in FIGURE 6.

While the invention may be otherwise embodied, it is herein shown and described as embodied in a positive displacement hydraulic pump capable of producing pressures in the neighborhood of approximately 1,500 psi. The pump is intended for use in the central hydraulic systems of farm tractors and the like.

The pump shown in the drawing generally comprises an outer casing member A having an internal chamber 10 therein containing a radially inwardly facing annular camming surface 12; and an inner member B having a plurality of radially outwardly extending cylinder bores 14 in which individual pistons 16 are positioned in a manner to be reciprocated by the camming surface 12 during relative rotation of the inner and outer members. The casing member in the present instance is a stationary one, and the internal member B is adapted to be rotated relative to the camming surface 12. by an axially positioned shaft 18 which extends through one end wall 249 of the casing member A. The inner end of the shaft 18 is journaled in a sleeve bearing 22 that is supported in an axial bore 24 in the opposite end wall 26 of the casing member A; and the projecting end of the shaft 18 is suitably journaled and sealed with respect to the end wall 20. The center portion of the shaft is suitably splined to the inner rotor member B substantially on the radial plane passing through the cylinder bores 14.

Each of the individual pistons 16 are provided with a ball 28 for engagement with the camming surface 12; and upon rotation of the shaft 18, centrifugal force causes the individual pistons 16 to be biased radially outwardly into firm engagement with the cumming surface 12. Relative rotation between the inner and outer members causes the pistons 16 (of which there are 7 in the present pump) to be reciprocated in their cylinder bores 14. By properly communicating each cylinder bore 14 to a supply of fluid at suction pressure when its piston 16 is moving radially outwardly, and by properly communicating each cylinder bore to the discharge passages of the pump when its piston 16 is moving radially inwardly, a pumping action is established. By successively Valving each of the cylinder bores to the suction and discharge connections of the pump during their respective suction and discharge strokes, a continuous flow of fluid is achieved.

Inasmuch as the position and duration of the suction and discharge strokes for each cylinder are fixed by the configuration of the camming surface 12; and inasmuch as the camming surface is held stationary with respect to the casing member A, the start of the suction stroke for each cylinder will take place when each cylinder moves precisely the same position relative to the casing member A, and will continue over precisely the same circular arc of the casing member. Likewise, the discharge stroke for each cylinder starts when each moves into precisely the same position and continues over the same circular arc of the casing member. Fluid can be added to and taken from each of the individual cylinders, therefore, by successively Valving each of the cylinders to suction and discharge passages in the casing member when the cylinders are properly positioned in the housing member relative to the camming surface 12.

The pump shown in the drawing has a generally elliptically shaped camming surface 12, so that each individual cylinder has two pumping cycles during each revolution of the rotor member B. Valving of each individual pumping cylinder to the suction and discharge connections of the pump is accomplished by a rotary disc valve arrangement formed between one end of the rotor member B.

and the end wall 26 of the casing member A. The rotary. disc valve arrangement shown generally comprises a pair of matching Valving surfaces which slidingly sealingly engage each other, and one of which surfaces 30 is formed in and rotated by the rotor member B while the other Valving surface 32 is supported on the casing member A. In order that sealing angular alignment of these surfaces can be accomplished easily when the pumps are made on a mass production basis, the matching

Valving surfaces

30 and 32 are spherically shaped; and in order that the Valving operation can be adjusted relative to the camming surface to 'vary the discharge of the pump (as will later be described), the Valving surface 32 is formed as a ura face of a port plate 36 which can be angularly positioned relative to the casing member A.

The pump shown in the drawing is intended to handle an oil having considerable lubricating qualities, and has been designed to permit fluid from the suctioned con- Iiection 38 of the pump to be distributed through an annular groove 46 in the end wall 26 to the internal chamber of the pump. The valving surface 30 of the rotor member B is provided with a plurality of identically shaped circular ports 42 each of which communicate with a res ective cylinder bore 14; and these ports 42 are uncovered by the port plate 36 to permit fluid from the internal chamber 10 to be drawn into the cylinder bores during their suction strokes. The ports 42 must, therefore, be sealed off from the internal chamber 10 during their discharge strokes; and inasmuch as the camrhing surface 12 causes these discharge strokes to be pioduc'ed over two diametrically opposite 90 arcs, the ort plate 36 has an hour glass type of configuration, as best seen in FIGURES 2 and 4, capable of sealing off the ports 42 from the internal chamber 10 over two diametrically opposite 90 arcs.

Referring now to FIGURE 2 of the drawings, the port plate 36 is shown therein in its position providing maxidisplace'nient for the pump. Assuming counter clockwise rotor rotation as seen in FIGURE 2, the pistons 16 stait their discharge s'tro'kes' when the center line of their cylinder bores 14 become coincident with the major axis 44 of the cammi'ng siirfacc. When the center line of the ports 42 become coincident with the major 44, the trailing edge of the ports 42 become coincident with the leading edge 46 of the leading sealing surface 48 of each half of the port plat'es valving surface 32, and the ports 42 become sealed on from both the suctioii and discharge connections of the pump for the following approximately 6 of rotor rotation. Approximately 3 of rotor rotation after the cylinder bores 14 pass the major axis 44, the leading edge of the circular ports 42 become coincident with the leading edge 50 of an arcuately shaped discharge port 52 that is centrally positioned within each half of the port plate 36. Thereafter subsequent rotation of the rotor member causes the circular ports to start the opening operation of the discharge ports 52 which continues for approximately 32 of rotor rotation. The leading edge 50 and the trailing edge 54 of the discharge ports 52 are formed to the same radius as the circular ports 42, and about centers which are spaced 23 of rotation. Thereafter the leading edge of the circular ports 42 begin to move past the trailing edge 54 of the discharge port 52; and after approximately 32 of further rotation, the trailing edge of the circular ports 42 become coincident with the trailing edge of the arcuately shaped discharge port 52 to close on the ports 42 from both the suction and discharge of the pump. This occurs as the center lines of the individual cylinder bores 14 become coincident with the minor axis 56 of the camming surface 12; and the circular ports 42 are valved off from both the suction and discharge of the pump thereafter for approximately 3 of rotor rotation.

The leading edge of the ports 42 become coincident with the trailing edge 58 of the trailing portion 60 of the valving surface 32 after the cylinder bores 14 have moved approximately 3 past the minor axis 56 of the camrning surface 12; and the individual circular ports 42 will remain iii communication with the internal chamber 10 or suction passages of the pump until the trailing edge of the ports 42 become coincident with the leading edge 46 of the other half of the valving surface 32-which occurs when the center line of the cylinder bores 14 become coincident with the major axis 44. This completes one suction and discharge cycle as occurs over 180 of rotor rotation; and thereafter the cycle is repeated with respect to the other half of the porting plate 36 during the second 180 of rotor rotation.

The port plate 36 slidably sealingly engages a planar surface 62 in the end Wall 26 of the casing member A in which are located a pair of diametrically opposed discharge ports 64 with which the arcuately shaped discharge ports 52 of the porting plate always communicate. The ports 64 are formed by longitudinally drilled passageways 66 which are intersected by a transverse drilling 66 which in turn is intersected by the discharge port 70 of the pump.

As previously indicated the port plate 36 is made arcuately movable in order that the pumps displacement, or quantity of fluid which will be delivered by the pump during one revolution of the rotor, might be varied. By rotating the port plate 36 in a clockwise direction from the position shown in FIGURE 2, the individual circular ports 42 will be val'ved off by the leading portion 48 of the valving surface 32 prior to the time that the radially outward stroke of the individual pistons 16 have been completed; so that only a fraction of each cylinder bores maximum displacement is filled with fluid from the inlet of the pump. The individual circular ports 42 will thereafter be v'alved off from both the suction and discharge connections of the pump for approximately 3 of rotor rotation; and thereafter the individual circular ports 42 will be connected with the arcuately shaped discharge port 52 in the port plate 36 so that fluid will be taken into each cylinder from the discharge of the pump for the remainder of each pistons suction stroke. Inasmuch as the port plate 36 is proportioned to connect the individual circular ports 42 to the discharge of the pump for approximately 87 of rotor rotation, movement of the center line of the cylinder bores 14 past the major axis 44 of the camming surface causes fiuid to be forced out through the

discharge ports

52 and 64. Discharge through these ports continues past the major axis 44 through an angular displacement of the remainder of 87 of rotation, which will now terminate before the cylinder bores 14 reach the minor axis 56. Thereafter the ports 42 will be valved off from both the suction and the discharge of the pump for approximately 3; and will then be communicated with the suction of the pump during the remaining portion of the discharge stroke of the individual pistons 16. It will therefore be seen that by angularly displacing the porting plate 36, part of the suction stroke for each cylinder bore 14 will be taken up with fluid obtained from the discharge ports of the pump; and a corresponding fluid displacement of the discharge stroke for each cylinder bore 14 will be passed to the suction of the pump. By this expedient the total quantity of fluid passing the outlet of the pump per rotor revolution can be varied or controlled by adjusting the angular position of the porting plate 36 relative to the camming surface 12.

The pump shown in the drawing is provided with automatic means for angularly positioning the port plate 36 in a direction decreasing the displacement of the pump when the pressure in its discharge passages exceeds a predetermined pressure, which in the present instance is; approximately 1,500 psi. The automatic means C for positioning the port plate is best seen in FIGURE 3 of the drawings; and generally comprises a cylindrically shaped slide member 72 which is positioned in a trans verse bore 74 in the cover plate 26. The slide member 72 is notched out as at 76 to receive a pin 78 that extends through an arcuately shaped opening 80 within the cover plate 26 that communicates the bore 74 with the back of the port plate 36. Pin 73 is rigidly connected to the port plate 36; and the port plate 36 is held in its maximum displacement producing position, shown in FIGURE 2, when the inner end of the slide member 72 is held into engagement with a shoulder 82 formed on the inner end of the transverse bore section 74. The slide 72 is held in this position by a coil spring 84 which is biased against an abutment plate 86 positioned against the outer end of the slide member 72 and a closure memher 88 which is suitably held in place in the outer end of the transverse bore 74.

The slide member 72' is adapted to be moved in a direction reducing the displacement of the pump by a piston 99 that is positioned in a smaller diameter bore section 92 in the bottom end of the bore 74; and which in turn is actuated by pressure supplied to its inner surface. Pressure actuation of the piston 90 is in turn controlled by a slide valve structure 94 which is adapted to communicate the bottom end of the piston 90 to the suction pressure of the pump until such time as the discharge pressure of the pump reaches a predetermined level of approximately 1,500 psi. Thereafter, the slide valve structure 94 is moved to modulate discharge pressure of the pump to the cylindrical piston 96 causing the slide member 72 to be moved outwardy compressing the spring 84 and moving the pin 78 in a direction decreasing the displacement of the pump.

The slide valve structure 94 is positioned in a bore 96 which intersects another small diameter bore 98 that communicates with the inner end of the bore section 92. The slide valve structure is provided with a pair of spaced lands 100 and 162 which when properly positioned will just straddle the bore 98' and close off the portions of the bore 96 which lie on opposite sides of the bore 98 from communication with the cylindrical piston 90. Pressure from the discharge passage 66 is fed through two intersecting bores 104 and 106 to the inner end portion of the bore 96. The outer portion of the bore 96- is communicated with the annular suction groove 40 by a longitudinal drilling 108; so that either suction or discharge pump pressures can be communicated to the cylindrical piston 90 depending upon the positioning of the slide valve structure 94. The slide valve structure 94 is biased inwardly to normally communicate suction pressure to the inner end of the cylindrical piston 90 by a coil spring 110 which normally holds an abutment plate 112 that is positioned against the end of the slide valve member 94 into engagement with the bottom end of the counterbore 116 in which the spring 110 is situated. The spring 110 is compressed a predetermined amount by a plug 118 which is forced into the outer end of the counterbore 116 and suitably held in place. When a predetermined pump discharge pressure, which in this instance is approximately 1,500 p.s.i. is delivered to the inner end of the bore 96 the slide valve structure 94 is biased outwardly against spring 110 to cause the abutment plate 112 to begin to move out of engagement with the bottom end 114 of the counterbore 116. This causes the land 102 to begin to throttle flow between the exhaust drilling 1G8 and the inner end of the cylindrical piston 90; and inasmuch as some leakage al ways occurs past the lands 1430 and 102 discharge pressure from the drilling 106 will flow past land 1130 to the drilling 98. Inasmuch as outlet flow from the drilling 98 to the exhaust drilling 108 is now being throttled, a control pressure is established in the drilling 98 which will be of an intensity which depends upon the relative overlap being maintained with respect to the lands 1% and 102. At a pump discharge pressure of approximately 1,650 p.s.i., pressure on the inner end of the slide valve 94 will cause the inner land 1110 to be moved out of overlap with respect to the inner end of the bore 96, and sufficient pressure will be delivered against the cylindrical piston 90 to move the port plate 36 into its no flow position. At pump discharge pressures between 1,500 psi. and 1,650 psi. a proportionate pressure will be delivered against the cylindrical piston 91) to cause the port plate 36 to assume intermediate positions.

Pressure from the arcuately shaped discharge ports 52 in the port plate 36 will, of course, tend to flow through the space between the valving surfaces 30 and 32 and will tend to bias the valving surfaces apart. Should the surfaces become separated, discharge from the cylinder bores 14 will be short circuited directly to the internal chamber thereby greatly decreasing the pumps hydraulic efficiency. The valving surfaces 30 and 32 must therefore be biased together by an amount of force which will prevent excessive leakage between the valving surfaces. In order that the pressure seepage between the valving surfaces might be confined to as small an area as possible, and thereby decrease the amount of force tending to bias the valving surfaces apart, an annular groove 120 is formed in the surface 36 a short distance radially outwardly from the radially outer edge of the arcuately shaped discharge ports 52. Similarly, an axially positioned recess 122 is formed in the rotor member B with its radially outer edge positioned a short distance radially inwardly from the inner edge of the arcuate shaped discharge ports 52. It will therefore be seen that pressure forces upon the port plate are confined to its area bounded by the annular groove 120, the axially positioned recess 122 and its leading and trailing

edges

46 and 58 respectively. A full discharge pressure will be exerted against the rotor member B on areas defined by the arcuately shaped discharge ports 52 and the pressure distribution on the remainder of the area bounded as previously set forth, will vary from substantially full pump discharge pressure adjacent the arcuate opening 52 to substantially suction pressure around the outer edges of the area previously set forth. An approximation of the force biasing the valving surfaces apart can be obtained by adding: the force obtained by multiplying the area which is in sliding sealing engagement by a pressure which is approximately one half of the difference between suction and discharge pressures, and the force obtained by multiplying full discharge pressure to the area of the arcuately shaped discharge ports plus the area of all circular ports 42 which are communicated to pressure.

According to principles of the present invention, the valving surfaces 36 and 32 are forced into sliding sealing engagement with each other with a generally predetermined force by a single balancing piston 124 which in the present embodiment extends around the shaft 18. The annular piston 124 is preferably confined to an area that is as close as possible to the shaft 18; and in the embodiment shown in the drawing, is positioned in a counterbore 126 in the end of the opening in the rotor member B through which the shaft extends. O-ring seals 128' and 130 are provided between the annular piston 124 and the sidewalls of the counterbore and shaft respectively; and another O-ring seal 132 is provided in the shaft opening of the rotor member inwardly from the counterbore 126. The outer surface 134 of the annular piston 124 bears against an annular abutment or slipper plate 136 which is non-rotatably supported on the end wall 20 of the casing member A surrounding the shaft 18. Fluid under a function of discharge pressure is admitted to the inner surface of the annular piston 124 to force the annular piston into abutment with the slipper plate 136 to produce a reaction which holds the valving surface 30 of the rotor member B into sealing engagement with the valving surface 32 of the port plate 36. The cross-sectional area of the annular piston 124 is preferably of such a size so as to at all times bias the valving surfaces 30 and 32 together by an amount suiiicient to prevent excessive flow between the valving surfaces.

According to the principles of the present invention, pressure is supplied to the counterbore 126 from the cylinder bores 14 through interconnecting passageways 13 3, one for each cylinder, and each of which passageways contain a flow restriction or orifice therein, so that flow continually proceeds through those passageways 138 of the cylinders under pressure to the balancing ring, while at the same time How is proceeding from the balance ring through those passageways 138' which are connected to suction pressure to establish an intermediate pressure behind the balance piston. The number of passageways 138 which are communicated to pressure at any instant will depend upon the total number of pistons in the pump, and the instantaneous position of the rotor member B. In any given pump design, there will be a maximum number of its cylinders which can be communicated to pressure at any one time, and a minimum number of cylinders which can be communicated at any one time. Generally speaking the maximum number can be ascertained by determining the maximum number of cylinder centerlines which can overlie the segments of the cam which produce the discharge stroke; and in the pump shown in the drawing wherein 7 pistons are shown, the maximum number will be 4. The minimum number will of course be 7 minus 4 or 3. In a six piston pump the maximum number will be 4, and the minimum number will be 2.

The force tending to separate the port plate 36 from the rotor B, has been found to vary with rotor speed, and it has also been found that the pressure that is developed behind the balancing piston of a given pump will depend upon the type of orifice that is used in the passageways 138. It will be remembered that full pump discharge pressure is subjected to that portion of the sliding sealing valving surfaces of the port plate and rotor to which the discharge port 52 is communicated, and that a pressure intermediate pump suction and discharge pressure acts upon the remainder of the sliding sealing valving surfaces. It will be seen, therefore, that the total force tending to separate the valving surfaces varies throughout the rotor cycle and depends upon the number and position of rotor ports 42 communicating with the discharge port 50 of the port plate. My prior application Serial Number 719,285, filed March 5, 1958, now abondoned, describes an arrangement whereby a substantially constant pressure force is exerted upon the balance piston to oppose the fluctuating pressure force between the valving surfaces. The present invention provides means whereby a changing pressure force can be exerted upon the balance piston which will at all times be more generally proportional to the fluctuating pressure force tending to separate the valving surfaces.

It has been found that in any given pump, larger pressure fluctuations will be created behind the balance piston when a sharp edged orificeis used in each passageway 138, than is created when an orifice producing laminar flow is used in each passageway 138. By changing the shape and construction of these orifices, therefore, the amount of pressure change behind the balancing piston can be regulated. In the preferred arrangement, it will usually be desirable to cause the pressure forces on the balancing piston to be made proportional to the force tending to separate the valving surfaces. As previously stated the speed of rotor rotation changes the pressure forces tending to separate the valving surfaces, but it will generally be true that the separating forces will vary as a function of the percentage of valving surface to which full discharge pressure is communicated. The changes in separating force will usually be abrupt as a rotor port 42 is brought into communication with the discharge port 50, and the degree of change depends upon the percentage change in area to which full discharge pressure is communicated.

Thepump shown in the drawing is provided with seven pistons, and will alternately bring 4 and then 3 of its cylinder chambers into communication with the discharge ports 50. In order to qualitatively show the elfect that orifice configuration plays upon the pressure forces exerted against its balancing piston, the following equations are given by way of illustration:

Assume that a sharp edged orifice is used, flow to the counterbore '126 must at all times be equal to the flow out, four equal orifices are feeding the counterbore while three are bleeding flow out, and flow through a sharp edged orifice varies generally as the square root of the pressure drop:

AP =pressure drop in AP =pressure drop out 8 AP=pressure drop (discharge to suction) 16AP1=9AP2 AP1=%GAP2 AP1+APZ=AP 9ia 2+ 2= AP= AP AP2=1%5A.P 1=%5 P When three orifices are connected to pump discharge and four to pump suction,

AP1=1%5AP so that the fluctuation represents AP or 28%.

Now assume that laminar flow orifices are used in the same pump arrangement, and their flow through a laminar flow orifice is proportional to its pressure drop:

AP =%AP and AP =%AP When three orifices are connected to pump discharge and four to pump suction:

so the fluctuation here represents AI" or 14.3% which is approximately one-half that for the sharp edge orifices.

The laminar flow orifices shown as forming the passageways 138 of the pump shown in the drawing are formed by a drilling 140 of comparatively large sizeone for each piston chamber I l communicating its piston chamber 1 with the counterbore 126. Each of the drillings is of the same size, and each of the orifices are conveniently made of the identical size and shape by pressing identically shaped plugs 14?; into each of the drillings 140. When laminar flow orifices are to be used, the plugs 142 may be conveniently made by the method disclosed in my copending application Serial Number 706,059, filed December 30, 1957, now Patent No. 2,952,071, issued December 9, 1959. A plurality of small coil springs (not shown) will preferably also be interpositioned between the annular balancing piston 124 and the rotor member B to hold the valving surfaces together during the starting operation of the pump when no pressure is available for causing the balancing piston to hold the valving surfaces into scaling engagement. One end of these springs may be positioned in respective recesses in the balancing piston, while the other end of each spring may bear against the bottom of the counterbore 126.

In order to lubricate the sliding surface between the annular piston I24 and the slipper plate 136, a pair of concentric annular grooves 144 and 146 are provided in the outer surface 134 of the annular piston. Fluid under pressure from the counterbore 126 flows through a passageway 148 in the annular piston and then through a groove 150 in the surface 134 which extends between the recesses 144 and 146. The same pressure that is delivered against the inner edge of the annular piston 124 is therefore delivered to the sliding surface between the piston 124- and slipper plate 136 to relieve the mechanical hearing forces between these surfaces. The annular area between the grooves 144 and 146 is sized in such a way that the hydraulic forces tending to separate the annular piston 124 from the slipper plate 136 will at all times be slightly less than the force against the end of the annular piston 124 positioned in the counterbore 126; and

Q as previously indicated the annular piston 124 is so sized as to hold the valving surfaces 30 and 32 together. Rotation of the annular piston 124 relative to the shaft 18 is prevented by a pin 152 which extends into aligned opening 154 and 156 in the piston 124 and rotor member B, respectively.

Describing now the operation of the pump with the porting plate 36 in its maximum flow producing position shown in FIGURE 2, fluid enters through the pump suction 38 to the annular groove 40 where it is distributed uniformly to the internal chamber of the pump. With the port plate 36 in the position shown in FIGURE 2, the circular ports 42 in the rotor member B will be in communication with the internal chamber 10 for substantially the full suction stroke of their pistons 16, and until the centerline of the bores 14- are coincident with the major axis 44 of the camrning surface 12. The ports 42 are valved oif from both the internal chamber 10 and the discharge port 52. for the next 3 of rotor rotation. At aproximately 3 of rotation after the centerline of the cylinder bores have passed the major axis as, the circular ports 4-2 become communicated with the arcuately shaped discharge port 52 so that inward movement of the pistons 16 causes fluid to flow out through the ports 42, and the arcuately shaped discharge port 52 to one of the discharge ports 64 in the removable end wall 26 of the casing member. Inasmuch as the camming surface 12 is elliptically shaped to produce two pumping cycles during each revolution of the rotor member, flow simultaneously proceeds through both of the diametrically opposed drilled passage Ways 66 to the transverse drilling 6% and out through the discharge port 70 of the pump. When the centerline of the cylinder bores 14 reach the minor axis 55 of the camming surface 12, the trailing edge of the circular ports 42 move out of engagement with the arcuately shaped discharge port 52 to valve off the cylinder bores 14 from both the suction and discharge connections of the pump. The ports 42 remain sealed off from both of the suction and discharge passages of the pump for the next 3 of rotation, or until their centerlines have moved approximately 3 or rotation past the minor axis 56; and thereafter the leading edge of the circular ports 52 move past the trailing edge 58 of the port plate 36 to establish communication with the suction of the pump. The ports 42 remain in communication with the pump suction for approximately 87" of rotation thereafter; and the entire cycle will thereafter be repeated with respect to the diametrically opposed portion of the port plate 36.

As previously indicated, the amount of fluid discharged from the pump can be varied or regulated by rotation of the port plate 36 with respect to the camming surface 12 of the casing member. Angular displacement of the port plate 36 with respect to the casing member A is accomplished by the structure best shown in FEGURE 3, and which comprises a slide member 72 that is normally biased into its maximum flow producing position by the coil spring 84. The slide 72 is caused to angularly displace the port plate 36 in a direction reducing the output of the pump when a pressure exceeding approximately 1,500 psi. is supplied to the piston 90 which abuts the inner end of the slide member 72. When the discharge pressure of the pump exceeds approximately 1,500 p.s.i., the spool valve structure 94 moves outwardly to compress spring 110 sufliciently to cause land 102 to lap with respect to bore 96. Thereafter variable leakage rates pass the lands 100 and 102 causes increasing control pressure to be delivered against the piston 90 which in turn causes the slide 72 to compress spring 84 and rotate the port plate 36. The port plate 36 will be rotated by increasing amounts as the discharge pressure exceeds 1,500 p.s.i.; and when approximately 1,650 p.s.i. discharge pressure is reached, the slide member 72 will abut plug 88 and the port plate 36 will be rotated to its no flow producing position for the pump. As the discharge pressure of the pump falls below 1,650 p.s.i. the reverse operation is produced; and it will therefore be seen that the pump is capable of adjusting its rate output to correspond with the consumption of the hydraulic system to which it is connected. The precise manner in which angular displacement of the port plate 36 reduces the displacement of the pump has previously been set forth in detail and will not further be described.

During the operation of the pump, pressure is supplied from individual ones of the cylinder bores 14 which are producing a pumping action through the passageways 138 to the recess 122 to bias the annular piston 124 into engagement with the slipper plate 136, and thereby hold the valving surfaces 3%? and 32 of the rotor and port plate, respectively, into sliding sealing engagement. As has previously been explained some of the passageways 138 will be communicated with the pump discharge while others of the passageways 133 will be communicated to the pump suction so that the pressure that is established in the recess 122, will be a balance as determined by the flow rates in and out of the passageways 1.33. As has been previously explained there will be times when four passageways 138 will be supplying fluid to the recess 122 and three passageways 138 are conducting flow out of the recess 122; while at other times three passageways will be conducting flow in and four passageways will be conducting flow out. The pressure behind the annular piston 124 will therefore vary, and as previously explained, this fluctuation when using laminar flow orifices of the type shown in the drawing will be approximately one-half that which would be produced it the flow restriction in the passageways 138 were of the sharp-edged orifice type. In the preferred embodiment of pump shown in the drawing, the variation in force tending to separate the valving surfaces 39 and 32, more nearly corresponds with the pres sure fluctuation as produced by the laminar flow orifices shown in the drawing; so that the force produced by the balancing piston will slightly exceed and be generally protional to the pressure forces tending to separate the valving surfaces.

As previously indicated, some of the fluid supplied to the recess 122 passes through passageway 148 to the annular grooves 144- and M6 to pressurize the abutting faces of the annular piston 124 and slipper plate 136 to thereby reduce the direct bearing force between these sliding surfaces. As previously indicated the hydraulic force tending to separate these surfaces is less than the pressure force on the inner end of the annular piston biasing it into engagement with the slipper plate 136, which force in turn is greater than the hydraulic pressure forces tending to separate the valving surfaces 30 and 32 by an amount preventing excessive flow therebetween. A continuing amount of leakage occurs out of grooves 144 and 146, as well as past the O-rings 123, 13% and 132. inasmuch as these flows are derived from the fluid flowing through the passageways 138 into the recess 122, these flows may be utilized to further tailor t .e balancing pressure behind the annular piston 12.4 to more nearly approach the pressure forces tending to separate the valving surfaces 30 and 32.

I have further found that pump designs having odd number of pistons greater than five, as for example, seven, nine, eleven, thirteen, etc., can be designed to have valving surfaces in which the number of cylinder ports which are valved to discharge at any one instant during its rotor cycle will not vary by more than one. Laminar flow orifices of the type shown in the drawing can therefore be used to good advantage in pumps and/or motors of these configurations, and the pulsations in the pressure discharge of pumps having these configurations decreases as the number of cylinder chambers increases. Applicant has further found that pumps having six fluid pressure chambers alternately connects four and then two of its chambers to the pump discharge to produce large pressure fluctuations in the forces tending to separate its valving surfaces.

It will be apparent that the objects heretofore enumerated as well as others have been achieved; and that an improved pressure balancing arrangement for a single balancing piston has been provided for fluid devices having end valving surfaces which are biased apart by pres sure fluid between the valving surfaces.

While the invention has been described in considerable detail, I do not wish to be limited to the particular constmctions shown and described; and it is my intention to cover hereby all novel adaptations, modifications and arrangements thereof which come within the practice of those skilled in the art to which the invention relates.

I claim:

1. In a positive displacement fiuid handling device: an outer casing member having first and second opposite end walls forming an internal chamber therein and through which end walls an axis of rotation extends; an internal member in said chamber; means for providing relative rotation between said internal member and said casing member; said internal member having a plurality of cylinder bores therein; a piston in each cylinder bore constructed and arranged to be reciprocated in said bore during relative rotation of said internal and outer casing members; said first end wall of said casing member and the adjacent end of said internal member having respective valving surfaces which are in sliding sealing engagement with each other and which define areas of high and low pressure that are exerted against said end of said internal member; said internal member having a flow communicating port for each cylinder bore extending between each bore and the valving surface of said internal member; said second end wall of said casing member having a coaxially positioned abutment surface facing said internal member; said internal member having a coaxially positioned recess therein facing said abutment surface; a balance ring sealingly engaging the sides of said recess and having a surface with annulm grooves slidingly engaging said abutment surface, the area, of said balance ring disposed between said grooves being communicated by through passages with the opposite face of said balance ring and proportioned of slightly less area than the area of said balance ring exposed to the pressure at its opposite face to effect slight engagement force between said balance ring and abutment surface; a plurality of individual flow passages communicating respective cylinder bores with said recess; said individual flow passages each having a flow restriction defined by a sector shaped cross-section and an elongated length located at a radial distance inwardly of the outer periphery of the wall structure defining said recess for limiting flow therethrough from said recess to its respective cylinder bore when its cylinder bore is communicated to low pressure to provide laminar flow at all times therethrough.

2. In a positive displacement fluid handling device: an outer casing member having first and second opposite end walls forming an internal chamber therein and through which end walls an axis of rotation extends; an internal member in said chahber; means for providing relative rotation between said internal member and said casing member; said internal member having a plurality of cylinder bores therein; a piston in each cylinder bore constructed and arranged to be reciprocated in said bore during relative rotation of said internal and outer casing members; said first end wall of said casing member and the adjacent end of said internal member having respective valving surfaces which are in sliding sealing engagement with each other and which define areas of high and low pressure that are exerted against said end of said internal member; said internal member having a flow communicating port for each cylinder bore extending between each bore and the valving surface of said internal member; said second end wall of said casing member having a coaxially positioned abutment surface facing said internal member; said internal member having a coaxially positioned recess therein facing said abutment surface; a balance ring sealingly engaging the sides of said recess and having a surface which slidingly engages said abutment surface and having two annular grooves therein defining a surface area therebetween proportioned to be of less area than its opposite face which is exposed to pump pressure; a plurality of individual flow passages located at a radial distance inwardly of the outer periphery of the wall structure defining said recess and communicating respective cylinder bores with said recess; said individual flow passages each having a flow restriction of sector shaped cross-section and of substantial length to extend between the opposite faces of said balance ring for effecting laminar flow therethrough from said recess to its respective cylinder bore when its cylinder bore is communicated to low pressure; and a flow passage in said balance ring communicating the pressure of said recess to the area between said grooves to relieve pressure between the sliding engaging surfaces of said balance ring and abutment surface.

3. In a positive displacement fluid handling device: an outer casing member having first and second opposite end walls forming an internal chamber therein and through which end walls an axis of rotation extends; an internal member in said chamber; an axially extending shaft extending through said second end wall of said casing member into said internal chamber for supporting said internal member relative to said casing member; means for providing relative rotation between said internal member and said casing member about said shaft; said internal member having a plurality of cylinder bores therein; a piston in each cylinder bore constructed and arranged to be reciprocated in said bore during relative rotation of said internal and outer casing members; said first end wall of said casing member and the adjacent end of said internal member having respective valving surfaces which are in sliding sealing engagement with each other and which define areas of high and low pressure that are exerted against said end of said internal member; said internal member having a flow communicating port for each cylinder bore extending between the radially inner end of each bore and the valving surface of said internal member; said second end wall of said casing member having an annular abutment surface surrounding said shaft and facing said internal member; said internal member having an annular recess therein surrounding said shaft and facing said abutment surface; an annular balance ring sealingly engaging the sides of said recess and having a surface with annular grooves which slidingly sealingly engages said abutment surface and having passages extending therethrough to communicate the area between the annular grooves with the opposite face of said balance ring whereby pressure from said recess will be communicated to said area between the annular grooves; and a plurality of individual flow passages in said internal member communicating respective cylinder bores with said annular recess; said individual flow passages each having a full restriction having a cross section area and length of flow path relationship for effecting laminar flow at all times therein and for limiting flow therethrough from said recess to its respective cylinder bore when its cylinder bore is communicated to low pressure.

4. In a positive displacement fluid handling device: an outer casing member having first and second opposite end walls forming an internal chamber therein and through which end walls an axis of rotation extends; an internal member in said chamber; an axially extending shaft extending through said second end wall of said casing member into said internal chamber for supporting said internal member relative to said casing member; means for providing relative rotation between said internal member and said casing member about said shaft; said internal member having a plurality of cylinder bores therein; a piston in each cylinder bore constructed and arranged to be reciprocated in said bore during relative rotation of said internal and outer casing members; said first end wall of said casing member and the adjacent end of said internal member having respective valving surfaces which are in sliding sealing engagement with each other and which define areas of high and low pressure that are exerted against said end of said internal member; said internal member having a flow communicating port for each cylinder bore extending between the radially inner end of each bore and the valving surface of said internal member; said second end wall of said casing member having an annular abutment surface surrounding said shaft and facing said internal member; said internal member having an annular recess therein surrounding said shaft and facing said abutment surface; an annular balance ring sealingly engaging the sides of said recess and having a surface with annular grooves which slidingly engages said abutment surface and having passages extending therethrough to communicate the area between the annular grooves with the opposite face of said balance ring whereby pressure from said recess will be communicated to said area between the annular grooves; and a plurality of individual flow passages in said internal member communicating respective cylinder bores with said annular recess; said how passages being located at a radial distance inwardly of the outer periphery of the wall structure defining said recess; said individual flow passages each having a flow restriction having a cross section area and length of flow path relationship for efiecting laminar flow at all times therein and for limiting flow therethrough from said recess to its respective cylinder bore when its cylinder bore is communicated to low pressure.

5. in a positive displacement fluid handling device: an outer casing member having first and second opposite end walls forming an internal chamber therein and through which end walls an axis of rotation extends; an internal member in said chamber; an axially extending shaft extending through said second end wall of said casing member into said internal chamber for supporting said internal member relative to said casing member; means for providing relative rotation between said internal member and said casing member about said shaft; said internal member having a plurality of cylinder bores therein; a piston in each cylinder bore constructed and arranged to be reciprocated in said bore during relative rotation of said internal and outer casing members; said first end wall of said casing member and the adjacent end of said internal member having respective valving surfaces which are in sliding sealing engagement with each other and which define areas of high and low pressure that are exerted against said end of said internal member; said internal member having a flow communicating port for each cylinder bore extending between the radially inner end of each bore and the valving surface of said internal member; said second end wall of said casing member having an annular abutment surface surrounding said shaft and facing said internal member; said internal member having an annular recess therein surrounding said shaft and facing said abutment surface; an annular balance ring sealingly engaging the sides of said recess and having a surface with annular grooves which slidingly engages said abutment surface and having passages extending therethrough to communicate the area between the annular grooves with the opposite face of said balance ring whereby the pressure from said recess will be communicated to said area be tween the annular grooves; and a plurality of individual openings in said internal member communicating respective cylinder bores with said annular recess; each said individual opening having a plug firmly fit therein, said plug being of the same length as said opening and of less cross section area than said opening thereby forming a flow passage between said plug and opening, said fiow passage having a cross section area and length of flow path relationship for effecting laminar flow at all times therein and for limiting flow therethrough from said recess to its respective cylinder bore when its cylinder bore is communicated to low pressure.

References Cited in the tile of this patent UNITED STATES PATENTS Re. 20,026 Thoma June 30, 1936 669,193 Alexander Mar. 5, 1901 2,111,657 Benedek Mar. 22, 1938 2,155,455 Thoma Apr. 25, 1939 2,257,724 Bennetch Oct. 7, 1941 2,393,773 Hoffer Ian. 29, 1946 2,698,585 Cotner et al. Jan. 4, 1955 2,741,993 Orshansky Apr. 17, 1956 2,895,426 Orshansky July 21, 1959 2,972,961 Clark Feb. 28, 1961 FOREIGN PATENTS 1,122,271 France May 22, 1956 1,132,654 France Nov. 5, 1956 2,784 Great Britain Feb. 3, 1913 678,917 Great Britain Sept. 10, 1952 801,678 Great Britain Sept. 17, 1958

Claims

1. IN A POSITIVE DISPLACEMENT FLUID HANDLING DEVICE: AN OUTER CASING MEMBER HAVING FIRST AND SECOND OPPOSITE END WALLS FORMING AN INTERNAL CHAMBER THEREIN AND THROUGH WHICH END WALLS AN AXIS OF ROTATION EXTENDS; AN INTERNAL MEMBER IN SAID CHAMBER; MEANS FOR PROVIDING RELATIVE ROTATION BETWEEN SAID INTERNAL MEMBER AND SAID CASING MEMBER; SAID INTERNAL MEMBER HAVING A PLURALITY OF CYLINDER BORES THEREIN; A PISTON IN EACH CYLINDER BORE CONSTRUCTED AND ARRANGED TO BE RECIPROCATED IN SAID BORE DURING RELATIVE ROTATION OF SAID INTERNAL AND OUTER CASING MEMBERS; SAID FIRST END WALL OF SAID CASING MEMBER AND THE ADJACENT END OF SAID INTERNAL MEMBER HAVING RESPECTIVE VALVING SURFACES WHICH ARE IN SLIDING SEALING ENGAGEMENT WITH EACH OTHER AND WHICH DEFINE AREAS OF HIGH AND LOW PRESSURE THAT ARE EXERTED AGAINST SAID END OF SAID INTERNAL MEMBER; SAID INTERNAL MEMBER HAVING A FLOW COMMUNICATING PORT FOR EACH CYLINDER BORE EXTENDING BETWEEN EACH BORE AND THE VALVING SURFACE OF SAID INTERNAL MEMBER; SAID SECOND END WALL OF SAID CASING MEMBER HAVING A COAXIALLY POSITIONED ABUTMENT SURFACE FACING SAID INTERNAL MEMBER; SAID INTERNAL MEMBER HAVING A COAXIALLY POSITIONED RECESS THEREIN FACING SAID ABUTMENT SURFACE; A BALANCE RING SEALINGLY ENGAGING THE SIDES OF SAID RECESS AND HAVING A SURFACE WITH ANNULAR GROOVES SLIDINGLY ENGAGING SAID ABUTMENT SURFACE, THE AREA OF SAID BALANCE RING DISPOSED BETWEEN SAID GROOVES BEING COMMUNICATED BY THROUGH PASSAGES WITH THE OPPOSITE FACE OF SAID BALANCE RING AND PROPORTIONED OF SLIGHTLY LESS AREA THAN THE AREA OF SAID BALANCE RING EXPOSED TO THE PRESSURE AT ITS OPPOSITE FACE TO EFFECT SLIGHT ENGAGEMENT FORCE BETWEEN SAID BALANCE RING AND ABUTMENT SURFACE; A PLURALITY OF INDIVIDUAL FLOW PASSAGES COMMUNICATING RESPECTIVE CYLINDER BORES WITH SAID RECESS; SAID INDIVIDUAL FLOW PASSAGES EACH HAVING A FLOW RESTRICTION DEFINED BY A SECTOR SHAPED CROSS-SECTION AND AN ELONGATED LENGTH LOCATED AT A RADIAL DISTANCE INWARDLY OF THE OUTER PERIPHERY OF THE WALL STRUCTURE DEFINING SAID RECESS FOR LIMITING FLOW THERETHROUGH FROM SAID RECESS TO ITS RESPECTIVE CYLINDER BORE WHEN ITS CYLINDER BORE IS COMMUNICATED TO LOW PRESSURE TO PROVIDE LAMINAR FLOW AT ALL TIMES THERETHROUGH.