US3122104A - Balance ring for pulsating fluid machinery - Google Patents

Description

Feb. 25, 1964 J. O. BYERS, JR

BALANCE RING FOR PULSATING FLUID MACHINERY Original Filed March 5, 1958 3 Sheets-Sheet 1 UTE-l INVENTOR.

Feb. 25, 1964 J. o. BYERS, JR 3,122,104

BALANCE RING FOR PULSATING FLUID MACHINERY Original Filed March 5, 1958 3 Sheets-Sheet 2 INVENTOR.

Jen/5 dfiyms, J4.

United States Patent 3,122,194 eALANcE mun eon rursa'rmc ramp MAcnrNenv James 0. Byers, In, Oalrville, Conn, assignor to The Bendix Corporation, St. Joseph, Mich, a corporation of Delaware Continuation of abandoned application Ser. N 719,285, Mar. 5, 1958. This application Dec. 23, 1.961, Ser. No.

1 (1mm. c1. 103-161) The present invention relates to rotating fluid handling devices having two parts which rotate relative to each other and which have end surfaces through which the axis of rotation passes and which are forced apart by fluid pressure forces; and more particularly to fluid pressure actuated balancing means for holding sliding sealing valving surfaces formed on the end surfaces of two relatively rotating parts into engagement with each other with a generally predetermined force. The present application is a continuation of US. application Serial No. 719,285 filed March 5, 1958, now abandoned.

The present invention has particular advantages in positive displacement hydraulic pumps and motors of the type having a casing member and an internal member which are relatively rotatable about an axis which passes through abutting end valving surfaces on the respective members, and which valving surfaces are biased apart by pressure of the fluid passing through the valving surfaces. In the usual design of pumps and motors of this type, the internal member is usually made the rotating member and is provided with a plurality of radially extending cylinders having pistons therein which reciprocate when the interned member is rotated. In such an arrangement fluid is usually added to and exhausted from the rotor, through ports in the end valving face of the rotor, as these rotor ports are rotated past suction and discharge ports in an abutting valving surface of the casing member. It is inherent in such an arrangement that fluid pressure from the discharge ports flows between the valving surface to bias the valving surfaces apart. An opposing force must therefore be applied between the rotor and casing members in order to hold the valving surfaces into sliding sealing engagement with each other.

It is highly impractical to manufacture rotors with an end to end dimension that provides a suitable sealing fit between two opposite end internal casing surfaces. Prior art attempts to produce pumps of this nature, and with which applicant is familiar, have utilized a plurality of slipper pistons which are biased outwardly of the rotor member into abutment with one inside face of the casing member to force the other face of the rotor member into sealing engagement with the opposite internal face of the casing member. Individual ones of these pistons have been communicated with individual ones of the pumping chambers in the rotor, so that the changing pressure in each purnping chamber will automatically change he force biasing the sealing surfaces together in a manner slightly greater but proportional to the force tending to separate the valving surfaces. Prior art structure have been quite complicated and expensive to mmufacture, and what is more important have necessitated the expenditure of considerable power loss in rotating the plurality of individual slipper pistons over their abutting bearing surface.

An object of the present invention is the provision of a new and improved fluid handling device of the character above referred to which is greatly simplified in its construction, which is more eflicient in its operation, and is cheaper to manufacture than prior art structures.

Another object of the present invention is the provision of a new and improved pump and/ or motor of the above described type whose mechanical efliciency is considerably improved over prior art structures.

3,122,104 Patented Feb. 25, 1964 ice The invention resides in certain constructions and combinations and arrangements of parts; 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 this 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;

FIGURE 3 is an end view of the pump shown in FIG- URE 1 having parts broken away and sectioned and is taken approximately upon the line 33 of FIGURE 1;

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

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

While the invention may be otherwise embodied, it is herein shown and described as embodied in a positive dislacement hydraulic pump capable of producing pressures in the neighborhood of approximately 1,500 p.s.i. 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 relaive 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 20 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 13 is suitably journaled and sealed with respect to the end wall 26. The center portion of the shaft is suitable 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 camming surface 12. Relative rotation between the inner and outer members causes the pistons 16 (of which there are 6 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 fiow 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 into 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 3. 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 to suction and discharge passages in the casing member when the cylinders are properly positioned in the hous ng 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 7 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 36 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 3% 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 surface of a porttion 38 of the pump to be distributed through an annular groove 49 in the end wall 26 to the internal chamber 16 p of the pump. The valving surface 39 of the rotor member B is provided 'with a plurality of identically shaped circular ports 42 each of which communicate with a re spective cylinder bore 14; and these ports 42 are uncovered by the port plate 36 to permit fluid from the internal chamber 19 to be drawn into the cylinder bores during their suction strokes. The ports 42 must, therefore, be sealed off from the internal chamber 14 during their discharge strokes; and inasmuch as the camming surface 12 causes these discharge strokes to be produced over two diametrically opposite 90 arcs, the port plate 36 has an hour glass type of configuration, as best seen in FIG- URES 2 and 4, capable of sealing off the ports 42 from the internal chamber over two diametrically opposite 90 arcs. V V

' Referring now to FIGURE 2 of the drawings, the port plate 36is shown therein in its position providing maximum displacement for the pump. Assuming counter clockwise rotor rotation as seen in FIGURE 2, the pistons 16 start their discharge strokes when the center line of their cylinder bores 14 become coincident with the major axis 44 of the camming surface. When the center line of the ports 42 become coincident with the major axis 44, the

trailing edge of the ports 42 become coincident vith the;

' after the cylinder bores 14 pass the major 'axis 44-, the

leading edge of the circular ports 42 become coincident withthe leading edge 50 of an arcuately shaped discharge port 52 that is centrally positioned within each half of operation of the discharge ports 52 which continues for j approximately 32 of rotor rotation. The leading edge 5t) and the trailing edge 5 of the discharge ports 52 are formed to the same .radiusas the circular ports 42, and about centers'whichare spaced 23 of rotation; apart so that the circular ports 42 remain full open with respect to the discharge port 52 for the. next approximately23? the port plate 36. Thereafter subsequent rotation of the 1 rotor member causes the circular ports to start the opening 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 off 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 old 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 apprordmately 3 past the minor. axis 56 of the camming surface 12; and the individual circular port 42 will remain in 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 7 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 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 35 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 opposeddischarge ports 64 with which the arcuately shaped discharge ports 52 of the porting plate always communicate. The ports at are formed by longitudinally drilled passageways 66 which are intersected by a transverse drilling 63 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 dis lacement,

or quantity of fluid which will be delivered 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 valved off by the leading portion 43 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 or" each cylinder bores maximum displacement is filled with fluid from the inlet of the pump. The individual circular ports 42 will thereafter he valved 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.

placement 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 olf from both the suction and the discharge of the pump .for approximately 3; and'wiil then be communicated with the suction of the punp during the remaining portion of the discharge stroke of the individual pistons 15. It

1 will be passed to the suction of the pump. By this expediency 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 p.s.i. 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 transverse bore 74 in the cover plate 26. The slide member 72 is notched out as at 76 to receive a pin 73 that extends through an arcuately shaped opening 86 within the cover plate 26 that communicates the bore 74 with the back of the port plate 36. Pin 78 is rigidly connected to the port plate 36; and the port plate as 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 74 is held in this position by a coil spring 84 which is biased against an abutment plate 85 positioned against the outer end of the slide member 72 and a closure member 88 which is suitably held in place in the outer end of the transverse bore 74.

The slide member 72 is adapted to be rotated 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 96 in turn controlled by a slide valve structure 9d which is adapted to communicate the bottom end of the piston 9% to the suction pressure of the pump until such time as the discharge pressure of the pump reaches a predetermined level of approximately 1,508 p.s.i. Thereafter, the slide valve structure 94 is moved to modulate discharge pressure of the pump to the cylindrical piston 9% causing the slide member 72 to be moved outwardly 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 @5 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 169 and 102 which when properly positioned will just straddle the bore 8 and close ed the portions of the bore 95 which lie on opposite sides of the bore 3 from communication with the cylindrical piston 90. Pressure from the discharge passage 66 is fed through two intersecting bores 104 and 1% 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 188; so that either suction or discharge pump pressures can be communicated to the cylindrical piston 98 depending upon tire 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 93 by a coil spring 11% 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 1 3 is situated. The spring lid 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 11%) 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 1&2 to begin to throttle flow between the exhaust drilling 1 53 and the inner end of the cylindrical piston and inasmuch as some leakage always occurs past the lands 1% and 162 discharge pressure from the drilling 1% will flow past land 1150 to the drilling 93. Inasmuch as outlet flow from the drilling 98 to the exhaust drilling 1&8 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. 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 1430 to be moved out of overlap with respect to the inner end of the bore 96, and suiiicient pressure is delivered against the cylindrical piston 99 to move the port plate 36 into its no flow position. At pump discharge pressures between 1,500 p.s.i. and 1,650 p.s.i., a proportionate pressure is delivered against the cylindrical piston 90 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 39 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 19 thereby greatly decreasing the pumps hydraulic efiiciency. 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 3% 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 dis charge 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 annular balancing piston 124 which extends around the shaft 18. The annular piston 12 is preferably confined to an area that is as close as possible to the shaft 13; 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 139 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 29 of the casing member A surrounding the shaft is. Fluid under discharge pres sure 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 39 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 12 iis preferably of such a size so as to at all times bias the valving surfaces 39 and 32 together by an amount sufficient to prevent excessive flow between the valving surfaces.

According to further principles of the present invention, pressure is supplied to the counterbore 126 from individual ones of the cylinder bores 14. Inasmuch as the pressure in these cylinder bores varies between suction and discharge pressures, a check valve is provided in these passages to prevent return flow from the counterbore to the cylinder bores during their suction strokes. In the embodiment shown, the interconnecting passageway 138 is counterbored as at 14% adjacent the bottom of the counterbore 126, and a ball 142 is positioned therein to prevent the return flow. The depth of the counterbore 140 is preferably just sufficient to provide from .005 to i010 of an inch clearance between the bottom of the counterbore, such that very little travel is required of the ball before it is brought into engagement with its seat.

In order to lubricate the sliding surface between the annular piston 124 and the slipper plate 136, a pair of concentric annular grooves 144 and 146 are provided in the outer surface i34- of the annular piston. Fluid under pressure from the counterbore 125 flows through a pas sageway 148 in the annular piston and then through a groove in the surface 134 which extends between the recesses 144 and 14a 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 plat 135 to relieve the mechanical bearing 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 as pre' viously 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 13 is procircular ports 42 in the rotor member B will bein communication with the internal chamber 155 for substantial- .ly the full suction stroke of their pistons 15, and until the center line of the bores 14 are coincident with the major.

axisedof the camrning surface 12. The ports 42 are valved o& from both the internal chamber iii and the dis-v charge port 52 for the next 3 of rotor rotation. At appromately 3 of rotation after the center line of the cylinder bores have passed the major axis 44, the circular ports iZ become communicated with the arcuately'shaped discharge port 52 so that inward movement of the pistons 3.6 causesiluid to flow out through the ports 42, and the 1 oer.

arcu'ately shaped discharge port 52 to one of the discharge ports 64 in the removable end wall 26 of the casing morn- Inasmuch as the carnmin-g surface 12 is eliiptically shaped to produce'two pumping cycles during each revolution of the rotor member, flow simultaneously proceeds through both of the diametrically opposed drilled passageways 66 to the transverse drilling '63 and out through the discharge port 79 of the pump. When the centerline of the cylinder bores 14 reach the minor 56 of the cam-i ming surface 12, the trailing edge of the circular ports 42 nove 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 oil from both of the suction and discharge passages of the pump for the next 3 of rotation, or until their center lines have moved approximately 3 of rotation past the minor axis 56; and thereafter the leading edge of the circular ports 52 moves past the trailing edge 53 of the port plate as to establish con rnunioationwith the suction of the pump. The ports 4-2 einain in communication with the pump suction for a proximately 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 as widi respect to the carrrning surface 12 of the casing member. Angular displacement of the port plate 36 with respect to the casing member A is accon plished by the structure best shown in FEGURE 3, and ch comprises a slide member 72 that is normally biased into its maximum flow producing position by the coil spring The slide '72 is caused to angularly displace the port plate 3% in a direction reducing the output of the pump when a pressure exceeding approximately 1,500 psi. is supplied to the piston 9%) which abuts the inner end of the slide member '72. When the discharge pres: sure of the pump exceeds a proximately 1,500 psi, the spool valve structure 94 moves outwardly to compress spring 114 sufficiently to cause land M2 to lap with respect to bore 95. the lands 1% and N2 causes increasing control pressure to be delivered against the piston which in turn causes the slide 72 to compress spring, 34' androtate the port plate 36. The port plate 36 will be rotated by increasing amounts as the discharge pressure exceeds 1,500 psi; and when approximatel 1,650 psi. discharge pressure is reached, the'slide member 72 will abut plug 88 and the port plate 35 will be rotated to its no flow producing position for the pump. As the discharge pressure of the pump falls below 1,650 psi. 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 discharge stroke of the cylinder bores 14, pressure flu d flows from individual ones of the cylinder bores 14, passes through the passageway 133, past the ball check 142 into the recess 149 to bias the annular piston.

124 intoengagement with the slipper plate 136. When the individual cylinders 14 become communicated with the internal chamber it? of the pump, a slight reversal of flow causes the ball check 142 to engage its seatand thereby maintain pressure within the recess 122. As pre- 7 viously indicated, some of the fluid supplied to the recess 12?. passes through passage 14-8 to the annular grooves 144 and 146 to pressurize the abutting surfaces of the annular piston 12 i and slipper plate'lfs to thereby reduce the direct bearing force between their sliding surfaces. As previously indicated the hydraulic 'force tending to separate the surfaces of the m nular piston and slipper plate is less than the pressure force on the inner end of the annular piston biasing it into engagement with the slipper plate 136 and which force in turn is greater than the hydraulic pressure forces tending to separate the .valving surfaces 3 0 and 32 byan amount prevent ng excessive flow therebetween. A continuing amount of leakage occurs out of the grooves 144 and 145, as well as past the O-ringslild, 13%, and 132 which must be replaced by flow, through passage 13% and check valve 1A2.- In order that at least one fiow passage 13% will at all times be communicated Thereafter variable leakage rates pass.

to pressure, three of the cylinder bores 14 in tne present instance are provided with these passage and check valve structures.

The balancing piston structure of the present invention is not only simpler than that of prior art structures employing individual balancing pistons one for each or -re cylinder bores but has proven to have a much greater mechanical efiiciency, requiring considerably less horspower to produce rotation of the rotor r ember B with respect to the casing member A. The single annular piston of the present invention greatly reduces 1e frictional forces between itself and its slipper clate; inasmuch as considerably less surface is swept by the singular annular piston of the present invention than is re uired to be swept by a plurality of individual circular pr ons. What is more, the eifective moment are through whi these frictional resis nces are exerted upon the shaft is reduced to a prac ice-.1 minimum by the present invention. A further advantage is believed obtained by the s' annular piston construction in that the fluid surface on the slipper elate 136 is maintained by the single annular piston; While that in the p structures is constantly changing in a manner can pressure fluid to be swe,;t from between t. e balanc'ug pistons and the slipper plate which they abut.

It will be apparent t at thc objects heretofore enumerated as as others nave been achieved, and that an improved balancing piston arrangement been provided for fluid devices having end valving surfaces which are biased apart by pressure fluid between the valving surfaces. While the inven n has been described in considerable detail, I do not h to be limited to the particular constructions shown and described, and it is my 11 ention to cover hereby all novel adaptations, modifications and arrangements thereof which come within the practice of those sl; :vd in the art to which the invention relates and which come within the scope of the following claim.

I claim:

In a hydro-mechanical device: a casing member having an internal chamber with opposite end walls through which an axis of rotation extends, a rotor mounted for rotation in said chamber about said axis, one of said end Walls of said casing havin a rotary val-ling surface wi h inlet and outlet ports therein, and the other of said walls of said casing having a smooth abutment surface thereon, one end of said rotor having a cooperating rotary valving surface in sliding sealing engagement with said rotary valving surface of said casing, and the opposite end of said rotor having an axially positioned recess therein opposite said smooth abutment surface on said casing, an annular balancing piston operatively connected to said rotor for rotary movement therewith and having an inner surface and an outer surface, said balancing piston being received in said recess and slidably and sealingly engaging the side of said recess for relative axial movement therebetween, the inner surface of said balancing piston being disposed within said recess and the outer surface thereof slidably engaging the smooth abutment surface of said casing, passage means communicating a portion of said outer surface with said inner surface whereby pressure in said recess will be communicated to said portion of said outer surface, a shaft rotatably mounted to the end walls of said casing for rotation about the axis, said shaft being operatively connected to said rotor for rotating the same and extending axially through said rotor, said recess and said balancing piston, said rotor having at least one expansible chamber which communicates with said rotary valving surfaces and which expands and contracts during rotation of said rotor, said rotor having a passageway communicating said annular recess with said expansible chamber, a valve seat in said last named passageway, 21 ball check valve member adapted to be seated said seat, said ball valve being so aranged in said last named passageway to be unseated by pressure flow to said recess from said expansible chamber and to be seated by pressure in said recess to prevent reverse flow to said expansible chamber, whereby said pressure in said recess acts on said piston inner face to bias the outer face of said piston into sliding engagement with said smooth abutment surface and said rotor valving surface is biased against said casing valving surface by substantially constant pressure in said recess.

References Cited in the file of this patent UNiTED STATES PATENTS 2,712,794 Humphreys July 12, 1955 2,871,798 Thoma Feb. 3, 1959 2,895,426 Orshansky July 21, 1959 FOREIGN PATENTS 27,721 Great Britain Dec. 9, 1911 of 1911

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