US3028814A - High speed variable displacement pump - Google Patents

Description

APri] 1962 R. D. RUMSEY ETAL 3,028,814

HIGH SPEED VARIABLE DISPLACEMENT PUMP Filed Oct. 17, 1957 3 Sheets-Sheet 1 54- J All: I53: Z5 O i 0 m I] fi A24 49 H n ,M /& 91 3 50 5e "J 72 58 ZTf al P Eo/flzz Doug/a5 Ramsey April '10, 1962 ,R. D. RUMSEY- ETAL 3,028,814

HIGH SPEED VARIABLE DISPLACEMENT PUMP Filed 001;. 17, 1957 3 Sheets-Sheet 2 Fo/[lfl Doug/a5 Ramsey Emma/f 6. Mal/112L119 April 10, 1962 R. D. RUMSEY ETAL 3,028,814

HIGH SPEED VARIABLE DISPLACEMENT PUMP Filed Oct. 17, 1957 3 Sheets-Sheet 3 Lye .EZZZJLE Pol/112 Doug/as Ramsey Emmett 6. Manning b M %a4 3,028,814 HIGH SPEED VARIABLE DISPLACEMENT PUMP Rollin Douglas Rumsey, Buffalo, and Emmett C. Manning, Lockport, N.Y., assignors to Houdaille Industries, Inc, Butialo, N.Y., a corporation of Michigan Filed Oct. 17, 1957, Ser. No. 699,807 4 Claims. (Cl. 1(23-161) The present invention relates to improvements in constant pressure variable displacement pumps, and more particularly to pumps of this character which are automatically adjustable in response to operating conditions and requirements encountered during performance of the pump.

More particularly, the invention relates to improvements in rotary pumps having a rotor carrying a pumping means such as the type carrying radial pistons which are actuated by engagement with the inner surface of a substantially circular enclosing pump chamber eccentrically located with respect to the rotor. The displacement of the pump is varied by varying the eccentricity of the chamber with respect to the rotor. The invention contemplates improvements in the control of the positional displacement of the axis of the pump chamber between substantially zero displacement and full displacement to achieve a variation in pump output for constant pressure delivery. The invention also contemplates improvements which obviate the difiiculties which occur with change in dimensions with temperature variation. The problem of the generation of heat in pumps of the variable delivery type when the delivery is reduced to zero is also accommodated to correct problems heretofore encountered.

An object of the invention is, therefore, to provide an improved variable displacement constant pressure rotary pump.

Another object of the invention is to provide a constant pressure variable displacement pump which is capable of obtaining greater stability during operation, and-freedom from vibration or chatter, and other difficulties encountered with variation in part dimensions with temperature increase, and especially such as encountered due to the diiterence in expansion between the pump housing and the enclosed pump parts.

Another object of the invention is to provide a pump of the type above described wherein improved positioning and control means are provided for eccentrically locating the chamber surrounding the pump rotor for con trolling the output displacement of the pump.

Another object of the invention is to avoid the problems of overheating in a variable displacement pump when the displacement reaches zero and the fluid passing therethrough reaches a minimum by providing a by-pass supply of fluid for cooling the pump.

A further object of the invention is to provide an improved pump of the type wherein a rotor carries radially reciprocating pistons operated within an eccentric pump chamber, and wherein the pump chamber can easily and accurately be positioned with respect to the rotor.

Still another object of the invention is to provide a pressure balanced valve plate for supplying fluids to a rotary pump wherein the surface pressure between the valve plate and pump will remain substantially constant regardless of changes in fluid supply pressure.

Another object of the invention is to provide for increased fluid supply pressure to a rotary pump to decrease cavitation effects without having to supply a separate means for providing pressurized fluid.

Other objects and advantages will become more apparent with the teaching of the principles of the invention in connection with the disclosure of the preferred embodi- Stas atom 3,028,814 Patented Apr. 10, 1962 FIGURE 2ais a sectional view taken along line Ila- Ila of FIGURE 1 and showing an end view of the valve plate;

FIGURE 3 isa sectional view taken along line IIIIII of FIGURE 1;

FIGURE 4 is a sectional View taken along line IVIV of FIGURE 1; and,

FIGURE 5 is a sectional view taken along line V--V of FIGURE 4.

As illustrated in FIGURES 1, 3 and 4, the pump is enclosed by a housing 8 which may be provided with an outwardly extending flange 10 for purposes of mounting the pump housing, such as by bolts inserted through holes 12, 14 and 16 in the housing. The housing is preferably formed of a casting which is suitably cored and tapped for attachment of the various elements to be described.

At one end of the housing is a circular opening 18 through which a pump driving shaft, not shown, will be extended for driving a pump rotor 20. A shaft seal 21 will surround the shaft and the seal is pushed against the rotor 29 by Belleville springs 19 held in place by an annular ring 23 which is held against the housing 3 by screws 25. For enclosing the rotor and other parts within a housing chamber 22 defined by the housing, a housing cap 24 is attached to the end of the housing opposite the opening 18. The housing cap is suitably clamped to the end of the housing and secured thereto such as by bolts 26. A gasket 28 is provided to prevent leakage of fluid from within the housing. A shaftgasket 30 is also provided which will surround the pump driving shaft in a fluid-tight manner. This will permit the flow of cooling fluid through the housing in the manner which will later be described.

The pump rotor 20 is mounted Within the housing chamber 22 for rotation therein, and is substantially cylindrical in shape. The rotor 20 has a shaft receiving socket 32 to which the shaft is connected for driving the rotor in rotation. Rotation will be maintained at a constant speed, and the variable displacement constant pressure output will be obtained in the manner to be described.

A plurality of circumferentially spaced radially extending cylindrical holes are bored into the rotor as shown at 3-4. Fitted into each of the cylindrical bores are pistons 36 which are mounted for reciprocation in a generally radial direction to move into and out of the cylindrical openings or cylinders 34 to cause a pumping action. The cylindrical bores are formed to angle forwardly. As may be observed in FIGURE 3, the pistons are thus canted forward in such a way as to produce as low an overhanging couple as possible. 7

Intake fluid for the cylinders 34 is received through a fluid passageway 38 in the rotor, as illustrated in FIGURES 1 and 2 with a separate passageway provided for each cylinder bore 34. Passageway 38 (FIGURES 1 and 2), is a circular passageway canted in the direction of rotation of the rotor 20 so that as the rotor turns the fluid is rammed into the cylinder chamber, thereby reducing the inlet pressure required to prevent cavitation. The canted passageway 38, as it breaks through the face 39 of the rotor, forms an elliptical opening. This feature is particularly desirable on high speed pumps because it has been found necessary to run inlet pressure as high as psi. in order to force fluid into the pumping chambers with axial passageways. The fact that the canted channel is in the wrong direction for discharge flow is relatively immaterial because at that time no cavitation danger exists.

The pistons are forced inwardly in their reciprocating movement by a first inner modulating ring 40 which surrounds the rotor and is eccentric thereto in normal operating position. The modulator ring 40 may be shifted in a lateral direction to vary its eccentricity with respect to the rotor to thereby vary the displacement or effective output of the pump. For operating the pistons, rollers 42 are positioned between each of the pistons and the modulating ring 40 with the rollers 42 positioned with their axis extending parallel to the axis of the rotor 20 and engaging the outer ends 44 of the pistons and the inner surface 46 of the modulating ring 40. The rollers 42 remain substantially stationary during operation of the pump, inasmuch as the modulating ring is mounted for rotation, the rotary driving force is applied to the piston and the ring is carried therewith by at least one of the rollers 42 frictionally engaging the modulating ring 40 and carrying it in rotation with the rotor.

For carrying the rollers 42 in place and rotating them with the rotor 20, the rollers 42 are carried at one end by an annular flange 48 at the end of the rotor 20. The annular flange is provided with radial slots 50 at the locations of the rollers and in assembly, the rollers are dropped in the slots 50 to be held between the pistons 36 and the inner surface 46 of the modulating ring. The other ends of the rollers are carried in slots 52 in an annular ring 54 which is mounted on the rotor to rotate therewith.

The rotor is supported for rotation within the housing 8 on a ball bearing assembly 56 having bearing balls 58 held between an inner race 60 and an outer race 62. The ball bearing is supported in an annular socket 64 at one end of the housing chamber 22.

Thus, the rotor 20 rotates about a fixed axis with respect to the housing 8, and the pistons 36 are reciprocated by being held within the modulator ring 40 which is positioned eccentric with respect to the rotor. The eccentricity of the modulator ring is adjustable to vary the amount of reciprocation given the pistons and thus the output of the pump.

To control the eccentric position of the modulator ring 40, it is held within a second outer ring 66 which is located within the housing chamber 22. The outer positioning ring 66 supports the inner ring by a series of bearing rolls 68 located between the rings. The rings are thus concentrically located with respect to each other and the inner ring is free to rotate with respect to the fixed outer ring. The outer ring is moved within the housing chamber 22 by a number of positioning members with the primary positioning members being controlled by the pressure of the delivery fluid.

As the rotor 20 rotates and the pistons 36 reciprocate, fluid is taken in and expelled from the cylinders 34 through the fluid delivery passageways 38. For the intake stroke of the pistons 36, the passageways 38 are in communication with an elongated arcuate slot 70 in a valve plate 72 which is spring pressed against the fiat end face 74 of the rotor. Communicating with the elongated arcuate intake slot 70 is an intake passageway 76 which is formed by a connector 78 and which has an axis parallel to the axis of rotation of the pump. The arcuate slot 70 has substantially the same area as the outside diameter of the connector tubes 78, so that changes in inlet pressure will have no effect on the valve plate rubbing force.

The connector tube 78 has an outer diameter which permits it to fit snugly within an opening 86 within the valve plate and Within an opening 82 in the housing cap. O- ring seals 84 and 86 prevent the leakage of fluid as it flows into the connector tube 78. The cap has an internally threaded passageway 88 through which the intake fluid flows on entering the housing. For the delivery stroke of the piston, the fluid passageway 38 communicates with an arcuate slot 90 which extends around the other half portion of the valve plate 72. This slot 90 leads to a delivery passageway 94 within a projection 96 integral with the valve plate. The arcuate slot 90 has substantially the same area as the outside diameter of the valve plate projection 96, such that change in discharge pressure will have a minimum effect upon the valve plate rubbing pressure.

The valve plate projection 96 extends into an opening 98 in the housing cap, and the projection 96 and the connector sleeve 73 prevent the valve plate 72 from rotating with the rotor 20. The projection 96 contains a coil compression spring 100, which bears against a spring supporting washer 102 within the opening 98 in the housing cap and urges the valve plate 72 to a non-leaking engagement with the face 74 of the rotor. Fluid under pressure is delivered from the pump through the delivery or outlet passageway 104 in the housing cap. An 0- ring 106 seals the projection 96 within the opening 98 of the cap.

In the position illustrated in FIGURE 3, the modulating ring 40 is shifted to the left to reduce the pump displacement and shifted to the right to increase the displacement. For movement of the modulating ring to vary the pump output, lateral forces are applied to the outer ring 66. To guide the outer ring in its lateral movement within the housing chamber 22, opposed guide ascmblies 108 and 110 are provided. The guide assemblies include a ring supporting bearing member 112 for the guide assembly 108, and 114 for the guide assembly 110. An important feature of the invention is the provision of thermally expansible backing elements 116 and 118 for the guide bearings 112 and 114. The parts may be made of different materials having dilferent rates of expansion. For example, the housing 3 may be of a material having a greater coefficient of expansion and as the temperature of the pump increases the guides normally would move away from firm contact with the ring.

The bearing rings 66 and 40 preferably are constructed of alloy steel to carry the high contact stress and the housing is constructed of either aluminum or magnesium in order to minimize weight. Thus the expansion of the aluminum or magnesium case would be nearly double the rate of the alloy steel rings and considerable differential expansion will occur. The thermal expansive backing members are designed to have a rate of thermal expansion so that the difference in expansion with increase in temperature of the pump will be compensated for. The backing elements 116 and 118 are formed of Teflon, which is a plastic material sold commercially under the foregoing trade name, and which has an expansion approximately five times greater than aluminum or steel. The Teflon pads 116 and 118 are held within caps 120 and 122 which are threaded and screwed into sockets in the sides of the housing. O-ring gasket seals 124 and 126 are provided to prevent leakage from the housing chamber 22. Thus, the bearing members 112 and 114 continually hold the ring 66 snugly within the housing chamber preventing vibration and chatter, and permitting the ring to be moved laterally to shift the modulating ring 40.

In other words, the difference between the diameter of the outer ring 66 and the diameter of the housing is taken up by the guide assemblies 108 and 110. This difference will change with temperature change. The difference change must be equaled by the expansion and contraction of the pads 116 and 118 plus the expansion and contraction of the guide bearings 112 and 114. Since the expansion and contraction of the guide bearings 112 and 114 is small by comparison with the pads 116 and 118, because they are made of metal which has a low coeificient of expansion as compared with the Teflon pads 116 and 118, their expansion in many constructions can be ignored.

When the ring 66 and the housing are selected, the difference in expansion per degree of temperature change can readily be determined, from either measurement or from known expansion of annular members formed of given metals. The length of the pads 116 and 118 is then chosen to substantially equal said dilference in expansion per degree of temperature change. When said difference is a larger figure, the pads will be made longer, and when said difference is a smaller figure the pads are made shorter. Then, at temperatures between room temperature and operating temperature, the pads will fill the space caused by the difference in expansion between the ring and housing, but they will not overfill this space and the ring will not have play but also will not bind. If the expansion and contraction of the metal bearings 114 and 116 is to be taken into account, with metal bearings of a known coefficient of expansion, it is a simple matter to select a conibination of lengths of pads and bearings which will together yield an expansion equal to said expansion dilference between the ring 66 and the housing 8. As an alternative, a given bearing length may be selected, and when the Teflon pad length is determined, caps 120 and 122 are provided to support the pads so that the bearings are in engagement with the ring 66.

A lateral pressure of a constant force is applied to the ring 66 by a spring biasing element 128. The biasing element includes a hollow cap 130 threaded into an opening in the side of the housing 8, and having a cylindrical inwardly facing opening 132 in which slides a piston 134 biased toward the ring 66 by a spring 136. The spring 136 applies a constant pressure against the ring urging the modulating ring 40 toward a position of maximum displacement of the pump. Thus, when no other forces are applied, such as when the pump is first started, a maximum delivery will be received until pressure is built up.

Another lateral force is applied in the direction to urge the modulating ring 40 toward a position of maximum capacity of the pump by a positioning piston assembly 138. This assembly is held within a hollow boss 140 projecting from the housing 8 and facing inwardly toward the housing chamber 22. A piston 142 is slidably mounted within a lining 144 within the boss. The piston 142 has its inner end received Within a hollow sliding cup 145 also slidably mounted within the lining 144. The piston supporting lining 144 is threaded into the boss 140 from inside the casing chamber 22, and is held in place by threads 146. An O-ring seal 143 prevents leakage of pressurized fluid from a space 150 behind the piston 142. v

The space or chamber 150 behind the piston is filled with fluid communicated thereto from the output of the pump, and is, therefore, at the pump delivery pressure. As shown in FIGURE 4, a fluid pressure line 152 leads through a ridge 154 in the housing cap 24 and communicates at one end with the discharge passageway 104 from the pump'and at the other end with a lateral passageway 156 which leads to the space 150 behind the piston. Thus, the hollow sliding cup 145, is pressed against the ring 66 in accordance with the output of pressure of the pump, and tends to urge the modulating ring 40 to a position of full pump output.

A lateral force is applied against the ring 66 in an opposite direction by a feathering piston assembly 158. A feathering piston 160 is slidably held within a cap 162 threaded into the open end of a boss 164 on the housing 8. Seals 166 and 168 prevent leakage of pressurized fluid from the chamber 170 behind the feathering piston 160 and to the housing chamber 22. The piston is provided with a piston ring 172 which prevents fluid leakage past the piston. The piston 160 bears directly against the ring 66, and its position varies to vary the position of the modulating ring 40 with variance in discharge pressure of the pump. The pump discharge fluid is communicated to the chamber 170 behind the feathering piston 160 through a passageway 174 formed in a rib 176 on the housing cap 24. A lateral passageway 178 leads the fluid to the chamber 170.

Fluid is admitted to the passageway 174 to control the position of the feathering piston by an output pressure control valve assembly 180. The output pres sure control valve 180 has a valve body 182 which is threaded into a threaded opening 184 in the housing cap 24, as shown in FIGURE 4. The flow through the valve passes through an orifice 186 at the inner end which communicates withthe pump discharge passageway 104. The orifice is'formed in an inset fitting 188 in the end of the valve 180. The orifice leads to a cylindrical chamber 190 in which is located the control piston 192. The piston is slidably movable within the chamber 190 to open the lateral passageway 194 and permit a flow of pump delivery fluid through the passageway into an annular groove 196 around the valve body. The groove is in direct communication with the passageway 174 leading to the chamber behind the feathering piston 160. Thus, when the control piaston 192 is permitted to slide rearwardly and uncover the lateral passageway 194, pressurized fluid will be permitted to move the feathering piston 160, FIGURE 3, against the ring 66 to move the modulating ring 40 toward at position of decreased pump output. This will occur when the pump output reaches the predetermined pressure at which the output is to be maintained. When the pressure drops below the predetermined constant pressure, the piston 192, FIGURE 4, will cover the lateral passageway 194. Fluid will then bleed out of the chamber 170 behind the feathering piston 160 through the passageway 174, through lateral passageway 194, and through an axially extending passageway 198 through the valve body. This passageway leads to a spring chamber 200 in the valve body and the fluid will flow into the space 202 behind the piston and out through a lateral passageway 204 in the valve body which communicates with an opening 206 draining into the housing chamber 22. The fluid pressure behind the feathering piston 160 will thus be relieved, .again permitting the ring 66 to move the modulating ring 40 to a position of increased discharge. A balance is maintained at the proper discharge pressure by movement of the piston 192. During starting, it will be noted in FIGURE 4 that the piston 192 covers passageway 194 and opens bleed passageway 206. Further, it will be noted from FIGURE 4 that piston 192 will cover the lateral passage 204 to seal the spring chamber 200 when pressurized fluid is being directed up through the passageway 174 to the feathering piston. 1

Control of the valve piston 192 is accomplished by a plunger 208 controlled by a spring 210 positioned in the spring chamber 200. The plunger 208 is slidably mounted in the spring chamber 200 and has openings 211 to permit the free flow of fluid to either side of the plunger. The plunger has an extension which engages the piston 192 so that the piston is urged toward the end of the valve body 182 by the spring, and urged in the opposite direction by the pressure at the end of the piston in the chamber 190.

The spring pressure against the piston 192 is adjusted by a pressure adjusting cap 212 adjustably threaded onto a threaded end 213 of the valve body. The cap 212 has a spring engaging member or plunger 214, which is held within a cup 216 with the bottom of the cup carrying an expansion pad 218.

As the temperature of the parts of the pump increase, the spring modulus of the spring 210 decreases, thus decreasing the pressure of the spring against the piston 192. To compensate for this and insure a constant spring pressure against the piston 192, the pad 218 is formed of Teflon or a like material which has a coeflicient of expansion suflicient to close the coil compression spring 210 a distance so that its force against the piston 192 will remain constant regardless of the temperature of the pump. Thus, as the temperature of the pumpincreases and modulus of the spring decreases, the spring will be shortened due to the expansion of the thermal expansive pad 218. The function of this pad will remain the same for any setting of the pressure adjusting cap 212. It will be understood, however, that the thiclness of the pad chosen may be varied to suit the pressure setting in order to get the proper thermal compensation.

The pump is adapted to be used for supplying a systern wherein the demand for pressurized fluid is intermittent, so that a constant pressure will be delivered by the pump under varying quantities of delivery. Under some circumstances, the system will require no fluid and the pump will thus operate under zero delivery conditions. The friction of the pump parts naturally creates an amount of heat and this heat is dissipated during normal operation and absorbed by the fluid passing through the pump. Under conditions of no delivery, a special by-pass fluid is directed through the housing chamber 22 to provide the cooling necessary.

The by-pass fluid is provided by a bypass valve assembly 220 supported within a hollow boss 222 at the side of the housing 8. A valve body 224 is threaded into the hollow boss 222 from within the housing 8. The valve body carries a valve plunger 226, which is reciprocable within the body to control flow through a lateral passageway 228.

When the valve plunger 226 is in the position shown, the passageway 228 is closed. When the valve plunger 226 moves laterally, or to the left, as shown in FIGURE 3, the passageway 228 is opened, and a flow of fluid will be permitted from the delivery passageway 104 of the pump, FIGURE 4, to the housing chamber 22. The fluid will flow through a by-pass passageway 230, formed in a rib 232 in the housing cap 24. A lateral passageway 234 leads from the rib passageway 230 into an annular groove 236, FIGURE 3, communicating with the valve passageway 228. Fluid will flow through the hollow core 238 of the valve plunger 226 and through the port 240 in the base of the cup 242 which is slidably mounted within the valve body. The cup 242 supports the flared base of the valve plunger 226 and the flared base receives the force of a coil compression spring 244 bearing against the end of the hollow chamber 246 in the valve body 224.

It will be observed that a position of the valve plunger 226 is controlled by the position of the outer ring 66 which bears against the cup 242 holding the valve plunger against its spring 244. When the ring 66 is in the position wherein the pump is delivering fluid, the plunger 226 closes the orifice 228. However, when the ring 66 moves to the left from the position as shown in FiGURE 3, to a location where the modulating ring 40 is in a location of minimum delivery, the valve plunger 226 will uncover the passageway 228, and a flow of by-pass cooling fluid will be permitted from the delivery passageway of the pump. Since this delivery passageway is connected to a supply line or supply tank, or the like, a constant supply of fluid will be available for cooling the pump.

As shown in FIGURE 1, the by-pass cooling fluid will be permitted to escape from the housing chamber 22 through a passageway 247 in the pump housing 8 leading to an internally threaded drain opening 248. The by-pass fluid will, of course, permit lubrication for the roller bearings 68 between the rings 40 and 66, and the ball bearings supporting the rotor 20.

As a brief summary of operation, the rotor 20 is driven in rotation within the modulator ring 40 which is mounted to be moved eccentrically with respect to the rotor. The modulator ring rotates with the rotor, being carried within bearing rollers 68 held within an outer ring 66. The inner modulator ring transmits radial forces to pistons 36 carried in radially extending pump chambers 34 through axially extending rollers 42.

The outer ring 66 is laterally positoned to vary the output of the pump by a positioning piston 342, FlGURE 3, which receives pressure fluid within the piston chamber from the pump output. A variable positioning pressure is applied by a feathering piston 166 which intermittently receives discharge fluid from the pump by the action of a control valve 180. The feathering piston is larger than the piston 142 and therefore, applies a greater force.

A control valve has a valve piston 192 controlled by a spring 216 and alternately moves to expose a lateral passageway 194 to permit a flow of fluid up to the feathering piston. A thermal expansive pad 218 maintains the pressure of the valve spring 210 constant regardless of temperature change. During periods of minimum pump delivery, a bypass of fluid through the housing chamber 22 is permitted the by-pass valve 220 having the plunger 226 which is positionally controlled by the ring 66.

Thus, it will be seen that we have provided an improved constant pressure variable displacement pump which meets the objectives and advantages hereinbefore set forth. The pump has very important basic advantages in that it can be constructed in a small size for operation at a high speed, and has a potential long life.

It will be understood that while the rotor and pumping means are shown in the form of a member carrying reciprocating pistons that other types of pumping means may be employed utilizing certain features of the invention. It will also be recognized by those skilled in the art, that certain other changes may be made in various operating elements retaining the advantages of certain elements embodying the principles of the invention.

The pump is compact and capable of accurate control with high speed operation and substantial delivery output. It will be recognized, that in certain circumstances, different positioning devices for the modulator ring can be employed. Further, various elements of the pump may be used in different operating circumstances, although the pump is shown in its preferred environment.

We have, in the drawings and specification, presented a detailed disclosure of the preferred embodiments of our invention, and it is to be understood that we do not intend to limit the invention to the specific form disclosed, but intend to cover all modifications, changes and alternative constructions and methods falling within the scope of the principles taught by our invention.

We claim as our invention:

1. In a variable displacement pump the combination comprising, a pump rotor carrying pumping means, an annular modulator ring positioned with its axis parallel and eccentric with respect to the rotor and operating the pumping means with rotation of the rotor within the ring, a housing defining a chamber for enclosing the ring and rotor for movement of the ring in the housing in a radial direction for changing the distance between the ring axis and rotor axis, pump output control means for shifting the ring radially within the housing, bearing guides supported on the housing and engaging the peripheral outer surface of the ring and located on opposite sides of the ring laterally of the path of radial shifting movement of the ring, and a backing member between the housing and at least one of the bearing guides non-yieldably holding the guides in engaging supporting non-binding contact with the ring and having a thermal expansion rate equal to the difference between the expansion rate of the housing and the ring so that the ring will be sup ported between the guides at varying temperatures without the formation of spaces between the ring and guides and without binding the ring.

2. In a variable displacement pump the combination comprising a pump carrying a radially reciprocating pump element, an annular modulator ring positioned with its axis parallel and eccentric with respect to the rotor axis and operative to reciprocate the pump element with rotation of the rotor within the ring, a housing defining a chamber for enclosing the ring and rotor for movement of the ring in the housing in a radial direction for changing the distance between the ring axis and rotor axis, said ring and said housing formed of diiterent materials with different coefficients of thermal expansion, pump output control means for shifting the ring radially within the housing, bearing guides supported on the housing and engaging the peripheral outer surface of the ring and located on opposite sides of the ring laterally of the path of radial shifting movement of the ring, and backing members between the housing and each of the guides nonyieldably holding the guides in engaging supporting nonbinding contact with the ring and having a thermal expansion rate equal to the diiference between the expansion rate of the housing and the ring so that the ring will be supported between said guides at varying temperatures without the formation of spaces between the ring and guides and without binding the ring.

3. In a variable displacement pump the combination comprising, a pump rotor carrying radially reciprocating pump elements, an annular bearing ring having an outer race and an inner race positioned with its axis parallel and eccentric with respect to the rotor axis and operative to reciprocate the pump elements with rotation of the rotor within the ring, a housing defining a chamber for enclosing the ring and rotor for movement of the ring in the housing in a radial direction for changing the distance between the ring axis and the rotor axis, pump output control means engaging the ring for shifting the ring radially within the housing, bearing guides supported in the housing and engaging the peripheral outer surface of a the ring and located on opposite sides of the ring laterally of the path of radial shifting movement of the ring, recesses in the housing supporting the bearing guides, and backing members in each of the recesses behind the hearing guides non-yieldably holding the guides in engaging supporting non-binding contact with the ring and having a thermal expansion rate equal to the difference between the expansion rate of the housing and the ring so that the ring will be supported between said guides at varying 10 temperatures without the formation of spaces between the ring and guides and without binding the ring.

4. A combination of elements in a variable displacement pump in accordance with claim 3 in which the backing members are formed of Teflon.

References Cited in the file of this patent UNITED STATES PATENTS 1,325,434 Todd Dec.-16, 1919 1,988,213 Ott Jan. 15, 1935 2,006,112 Heid June 25, 1935 2,143,937 Chandler Jan. 17, 1939 2,271,336 Goldsmith Jan. 26, 1942 2,273,468 Ferris Feb. 17, 1942 2,292,181 Tucker Aug. 4, 1942 2,309,833 Elze Feb. 2, 1943 2,345,952 I Smith Apr. 4, 1944 2,374,592 Ernst Apr. 24, 1945 2,422,864 Taylor June 24, 1947 2,506,974 Sorensen May S, 1950 2,525,498 Naylor et a1 Oct. 10, 1950 2,547,645 Horton Apr. 3, 1951 2,612,418 Krotz Sept. 30, 1952 2,646,755 Joy July 28, 1953 2,680,412 Entwistle June 8, 1954 2,729,165 Kremer Jan. 3, 1956 2,823,619 May Feb. 18, 1958 2,845,941 Wagner Aug. 5, 1958 FOREIGN PATENTS 161,911 Great Britain Apr. 21, 1921 443,041 Great Britain Feb. 20, 1936 953,223 Germany Nov. 29, 1956 OTHER REFERENCES Technical Service Bulletin No. 13, Teflon, DuPont Teflon Components and Coatings, Product Engineer- 7 ing, September 1952, pages 149-153,

Images