US2885960A - High pressure variable delivery rotary vane pump - Google Patents

Description

May 12, 1959 D. J. DESCHAMPS HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP 8 Sheets-Sheet 1 I Filed Nov. 29, 1955 INVENTOR.

DESIRE J. DESCHAMPS ATTORNEY May 12, 1959 D'. JiDESCHAMPS 6 HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP Filed Nov. 29, 1955 83heets-Sheet 2 m MM TA NH E C V S E E mo E m S E D wm E \il EN;

ATTORNE Filed Nov. 29, 1955 y 1959 D. J. DESCHAMPS 2,885,960

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INVENTOR. DESIRE J DESCHAMPS ATTORNEY D. J. DESCHAMPS May 12, 1959 2,885,960 HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP Filed NOV. 29, 1955 8 Sheets-Sqeet 4 INVENTOR. DESIRE J. DESCHAMPS .ATTORNEZ UMP - D. J. DESCHAMPS May 12, 1959 HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE P s Sheets-Sheet 5 Filed Ndv. 29, 1955 7 E 0 w w w. 30 4 M w i 7 a. w

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7 HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP Filed Nov. 29, 1955 8 Sheets-Sheet 6 D. J. DESCHAMPS 2,885,960

May 12, 1959 HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP 8 Sheets-Sheet 7 Filed Nov. 29, l955 W m F INVENTOR DESIRE J. DESCHAMPS 7A ATTORNEY D. JJDESCHAMPS HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP Filed Nov. 29; 1955 May 12, 1959 8 Sheets-Sheet 8 United States Patent HIGH PRESSURE VARIABLE DELIVERY ROTARY VANE PUMP Desire J. Deschamps, San Fernando, Calif assignor to -Iydroaire, Inc., Burbank, Calif a corporation of Caliorma Application November 29, 1955, Serial No. 549,606 17 Claims. (Cl. 103-4) This invention relates to high pressure pumps and, more particularly, to a compact, rugged, reliable and efficient positive displacement, variable capacity, high pressure and self-clearing vane pump particularly adapted for aeronautical uses. In so far as certain common features of novelty are concerned, this application is a continuation-in-part of my copending application Ser. No. 508,955, filed May 17, 1955 and which is, in turn, a division of my application Ser. No. 132,492, filed December 12, 1949, now Patent No. 2,708,884, issued May 24, 1955, and of my copending application Ser. No. 528,920, filed August 17, 1955.

An important object of the present invention is to provide a high pressure, variable displacement fuel pump for aircraft applications, which pump is compact in design, light in weight, and capable of long trouble-free service. Small size and low or light weight are particularly important requirements in pumps for aeronautical applications as, in modern aircraft, space and weight, particularly for engine accessories, are at a premium.

The demand for such a reliable variable displacement fuel pump is not new, but up to the present nearly all turbine jet engines have been equipped with fixed displacement pumps because of the lack of a suitable high pressure variable displacement pump. With a fixed displacement pump, regardless of the type, the regulation of the amount of fuel delivered to the engine, also called engine demand fuel, has been elfected by bypassing the excess fuel back to the pump inlet. This bypassing of part of the pump output back to the inlet means a considerable waste of power in pumping and increasing the pressure of a volume of fuel much larger than required by the engine. Compressing the fuel to a high pressure raises the temperature of the fuel, and the expansion of the hot fuel at the pump inlet generates vapor which, entering the fresh fuel arriving at the pump inlet, not only aifects the operation of the pump but also upsets the accurate metering of the fuel fed to the engine. The globules of vapor in the fuel, and the uneven flow of liquid fuel to the engine, may seriously affect the combustion process and the operation of the engine.

To understand the importance of the fuel by-passing problem, it should be noted that the fuel requirements of a jet engine at high altitude are only to of the requirements at take-off at sea level. This means that, with a pump rated to supply the engine requirements at take-off, the major portion of the pumped fuel is bypassed back to the pump inlet at high altitudes. The resultant vapor entering the fuel is more harmful at such altitudes where, due to the much smaller quantity of fuel used by the engine, the presence of substantial excess vapor upsets the combustion process in the engine.

Not many types of pumps are suitable for the design of a variable displacement high pressure fuel pump. The gear pump, which has been most widely used as a fixed displacement pump, does not lend itself at all to this purpose. As a practical solution, this leaves only the plunger or piston pump and the vane pump.

Despite their very good adaptability to high pressure applications, plunger pumps have disadvantages due to their reciprocating components, and these disadvantages increase with the speed of operation of the pump. As a plunger pump is not adapted, from either the operation or maintenance standpoints, to high speed operation, increase in capacity can be obtained only by substantially corresponding increase in the size and Weight of the pump.

For instance a plunger pump with an output of only 25 g.p.m. has already a size and weight greater than that of another type of pump having double this output. Another drawback of the plunger type pump is the necessity of an extremely close lapped fit of the plungers in their bushings, making the production cost very high, and unless the fuel pumped is very clean, or a suitable fine filter can be installed at the pump inlet, this type of pump is bound to give trouble with sticking plungers.

In passing, the centrifugal pump should be mentioned. Although a centrifugal pump is not a variable displacement pump in the true sense of the Word, it permits the adjustment of its output through throttling of the outlet without by-passing fuel to the inlet. For high fuel pressures as presently required, an aircraft centrifugal pump which, by necessity, has to be of small size and Weight, has to be operated at high speeds, for example, not less than 12000 r.p.m., and must be built with two stages.

In a centrifugal pump, the required high discharge pressure is obtained only at the design or rated pump speed, and the discharge pressure in a centrifugal pump, being a function of the square of the pump operating speed, makes this type of pump entirely unsuitable as an engine driven pump, particularly if the pump has to deliver fuel at the cranking speed of the engine for starting the latter. For instance, With the cranking speed of a representative turbine engine being about of its rated speed, the centrifugal pump would only deliver about of the rated discharge pressure (20 p.s.i. for a rated discharge pressure of 1000 p.s.i.) whereas a minimum fuel pressure of 250 psi. is required at engine cranking speed.

The rotary vane pump, on the other hand, has no difficulty meeting these low speed operation requirements and, when well built, has a much higher lift (a very desirable feature in modern aircraft applications) than any other type of pump excepting the plunger pump. The vane pump also has the ability of remaining primed and to keep operating with high V/ L fuel without stalling out; V/L being the ratio between the vapor and liquid content of the fuel, this ratio attaining up to 25% (vapor) under certain operating conditions of the aircraft.

The rotary vane pump also lends itself to the most simple mechanical solution of a variable displacement pump, but if designed in the manner that has been general practice, this type of pump has just as high bearing loads as the gear pump.

Vane pumps as presently constructed are not Well suited for high pressure operation as a fuel pump, because of their too high bearing loadings. This factor is accentuated by the fact that, for practical operating reasons, the pump bearings must be lubricated by the aircraft fuel being pumped, this fuel having a low viscosity and very poor lubricating qualities.

Even under such poor lubricating conditions, aircraft pumps are in successful service, operating at pressures of 500 psi, but when the discharge pressure of such pumps was increased over this figure, the load on the bearings became greater than advisable for normal operation, particularly if reliable long service life is expected.

While the materials for the shafts and bearings of vane pumps have been constantly improved, and the design of these parts greatly refined, these improvements and refinedischarge periods per revolution.

ments have been such as to permit output pressures and bearing loads of the order of 700 p.s.i., but without the assurance of reliability in operation and long service life characteristic of lower pressure pumps. There are not presently available shaft and bearing materials capable of prolonged trouble free operation at higher pressures with the poor lubrication characteristics of the fuels being pumped.

This problem of excessive bearing loadings has thus to be solved before a really reliable high pressure fuel pump can be built. The immediate and best solution is in the balancing of the hydraulic loads on the pump shaft and bearings. This cannot be effected conveniently in a gear pump, but can be effected in a vane pump.

In a vane pump, one solution for these ditficulties, successful from the standpoint of providing high outlet pressure, has been to make the pump housing with two eccentric pumping segments rather than one, with two With the discharge periods in 180 opposition, the loads on the rotor shaft are in substantially hydraulic balance and high outlet pressures without excessive bearing loads are possible. However this type of pump is not too suitable for high speed operation, thus necessitating undue size and weight for required outputs. Additionally, this particular design is limited to constant output or capacity pumps with the attendant disadvantages resulting from bypassing large quantities of pumped fuel back to the pump inlet. Constant capacity rotary vane pumps are not self-clearing as respects metallic contaminations in the pumped fuel.

In accordance with the present invention, a variable capacity vane pump is provided in which the bearing loads are substantially hydraulically balanced in a novel manner, fluctuations in the out-put pressure are reduced to a negligible amplitude, and the output capacity is varied substantially instantaneously in response to variations in the fuel demand of the engine.

Hydraulic balance of the bearing loads is effected by mounting three vane pumps side by side with their rotors assembled upon a single pump shaft. The center or intermediate rotor has an axial extent, or width, twice that of the two end or outer rotors, and its porting is inverted relative to that of the two outer rotors. This inversion, which results in the discharge port of the center rotor being substantially 180 from those of the two end rotors, causes the projected area of the center rotor and its extended vanes at its discharge port to be equal to the projected areas of the end rotors and their extended vanes at their discharge ports. As the combined area of the two end rotors subjected to discharge pressure is equal to the area of the center rotor subjected to discharge pressure, the hydraulic loadings on the pump shaft are completely balanced and there is substantially no hydraulic loading transmitted to the shaft bearings.

As the hydraulic loads on the three rotors balance each other on the pump shaft, the latter is subjected to a bending moment, which is a stress the shaft can easily be designed to stand without difficulty. As the shaft rotates, the points of application thereto of the hydraulic loads move around the shaft but are always 180 apart as respects the loading due to the center rotor and the combined loading due to the end rotors, thus maintaining a perfect balance.

While, at first glance, the use of three rotors might appear to introduce mechanical complications in the pump construction, these complications are more apparent than real, and any possible design complications are more than overbalanced by the advantages of removing all of the hydraulic loading from the shaft bearings while still providing for simple, instantaneous and effective variation of the pump capacity without bypassing of fuel back to the pump inlets. Removing the hydraulic loading, amounting to thousands of pounds, from the shaft bearings makes it possible to provide a suc 4 cessful high pressure vane pump with assurance of the shaft bearings having a long time, service free life.

Due to certain design and operational conditions, it is desirable to limit the number of vanes per rotor to a relatively low figure, such as four. Concornitantly this results in variations in the delivery rate four times per revolution if the vanes of the three rotors are exactly in angular alignment or phase with each other, the same as if there were but a single rotor. Such variations or fluctuations in pressure during each revolution are objectionable, particularly in the main fuel pump of a gas turbine engine.

In the invention pump, these undesirable fluctuations in output pressure are reduced to a negligible amplitude by offsetting the vanes of the center rotor relative to those of the end rotors by an angular distance substantially one-half the angular distance between adjacent vanes-as measured at the lines of contact of the vanes with their respective pumping cylinders. For example, with four vanes per rotor, such angular distance between adjacent vanes is and the vanes of the center rotor are offset 45 relative to the vanes of the end rotors. This doubles the number of fluctuations per revolution, with corresponding reduction in the amplitude of the fluctuations.

For example, in a typical embodiment of the invention pump, the peak of the delivery curve, during each cycle or fluctuations, is 2.8% above the mean delivery point of the curve, and the low point of the curve is 3.6% below the mean point, at rated pump output. The flow becomes even smoother at lower outputs prevailing during normal operation, an important consideration as the odds are that the full rated output will never be used, even at take-off, due to the pump rated output being designed to be greater than the maximum fuel demand rate of the engine.

Such relative offsetting of the rotor vanes results in the points of application of the hydraulic loads to the shaft bearings being not in exact opposition to each other and thus not exactly balanced out. In the case of four vanes per rotor, the points of application of the hydraulic loading are apart, creating a resultant force imposing a hydraulic loading on the pump shaft and bearings. However, this loading is still extremely low and, in practice, not more than one-fifth the value considered as safe for long time service-free operation of the bearings.

Variations in the pump output are provided by making the common axis of the pumping cylinders laterally adjustable relative to the common axis of the three rotors. For maximum output, the eccentricity of the two axes is a maximum, and the relative eccentricity is a minimum for minimum output, the limit being alignment or coincidence of the two axes for zero output.

The amount of such relative eccentricity of the cylinder and rotor axes is automatically controlled by a hydraulic servo mechanism operating responsive to the engine master fuel control to adjust, accurately and automatically, the pump output to the fuel demand without the necessity of by-passing pumped fuel back to the pump inlet.

The servo mechanism is of the follow-up type, and includes a valve having a lapped fit for free rotation in a follow-up sleeve. The only frictional resistance to valve movement is a shaft seal where the valve projects from the pump body to connect with the master fuel control. This seal is subjected only to the pump inlet pressure existing within the pump body.

Because of the large size of the fuel passages bringing fuel under pump discharge pressure to and from the hydraulic servo-mechanism, the rather small volume of the servo cylinder and the adopted lever ratio, the operation of the servo is extremely rapid, and the response practically instantaneous when the engine master control calls for a change in fuel flow from the pump. The feed back feature of the servo valve compensates ,automati-' cally for any leakage there might be in the hydraulic servo-system.

The servo mechanism is so designed that, after the engine has stopped, fuel leaking slowly past the valve will cause the pump to return to the maximum output position. This assures immediate fuel delivery and pressure upon cracking of the engine and, as soon as there is fuel flow and pressure, the servo mechanism automatically adjusts the pump output to correspond to the setting of the engine master fuel control, or to that of any manual control provided to override the master control during starting of the engine.

A further object of the invention is to insure reliable service operation of the pump even when pumping so called contaminated fuel containing fine sand and dust of a particle size and in amounts defined in applicable Air Corps Specifications. Any grit, dust or sand present in the fuel is, of course, bound .to accelerate wear in any type of .pmp it operated for long periods under such conditions, but in the invention pump such wear is minimized by the provision of a built-in very fine filter to remove such foreign matter from the small amount of fuel needed for bearing lubrication and for the operation of the hydraulic servo mechanism controlling the pump output.

An important advantage of the capacity variation of the pump is the self-cleaning action on metal contaminants in the fuel. Should a small piece of metal not clear through the outlet port, the vane behind this piece and engaging therewith will momentarily force the pump rotors away from the cylinder walls, due to the reaction on the vane from the metal piece jammed between the vane outer edge and the cylinder wall. Thus, the contaminant is cleared between the rotor and cylinder walls at the price of a momentary decrease in pump output. The vane edge may be nicked slightly, resulting in a minute decrease in output, but vane damage is minimized due to the ability of the vane to retract into its slot in the rotor and thus slide over the obstruction.

A further feature of the pump is a reducing valve for reducing the pressure of the fuel supplied to the shaft hearings to lubricate the latter. This greatly reduces the internal leakage from the bearings into the pump body, as compared to the leakage if the bearings were supplied with fuel at the high output pressure of the pump. For example, with an output pressure of 1000 p.s.i., the reducing valve supplies fuel to the bearings at 80 psi.

For an understanding of the invention principles, reference is made to the following description of a typical embodiment thereof as illustrated in the accompanying drawing. In the drawing:

Fig. l is a front elevation view of a variable capacity fuel pump embodying the invention;

Fig. 2 is an axial sectional view of the pump, taken on the line 22 of Fig. 1;

Fig. 3 is a side elevation view of the pump;

Fig. 4 is a transverse sectional view on the iine 44 of Fig. 3;

Fig. 5 is a plan view of the pump;

Fig. 6 is a transverse sectional view on the line 66 of Fig. 5;

Fig. 7 is a rear elevation view of the pump, showing the drive end;

Fig. 8 is a front elevation view corresponding to Fig. l with the front cover of the pump removed;

Fig. 9 is a sectional View on the line 9-9 of Fig. 1;

Fig. 10 is an enlarged view of a portion of Fig. 9;

Fig. 11 is a sectional view on the line 11-11 of Fig 9;

Fig. 12 is a transverse sectional view on the line 1212 of Fig. 3;

Fig. 13 is a sectional view on the line 13-13 of Fig. 12;

Fig. 14 is a sectional view on the line 14-14 of Fig. 1;

Fig. 15 is an enlarged view of a part of Fig. 14;

Fig. 16 is :a somewhat diagrammatic partially sectional View illustrating the balancing of the hydraulic loading on the rotor shaft;

Figs. 17 and 18 are diagrams respectively indicating full balancing of the hydraulic loading and the loading with the vanes of the center rotor angdilarly displaced relative to those of the end rotors;

Fig. 19 is a diametric sectional view of one pumping element showing the points of contact of the pump rotor vanes aligned with the leading and trailing edges of the inlet and outlet ports of the pump cylinder;

Fig. 20 is a sectional view of the servo-mechanism, taken on the line 2020 of Fig 1;

Figs. 21 through 24 are sectional views taken on the correspondingly numbered lines of Fig. 20; and

Fig. 25 is a partial plan view of the control valve of the servo-mechanism, illustrating the shape of the valve ports.

Detailed description of pump Referring to Figs 1 through 14, the working parts of the pump are enclosed within a body 30, somewhat oval shape in end elevation, having one end substantially closed by an integral annular wall 31 formed externally with .a circular mounting flange or flat surface 32 for attachment of the pump to the engine (not shown). The open opposite end of body 30 is closed by a removable cover 35 secured to body 30 by bolts 33 and screws 34. Body 30 is formed with a pump inlet port 36 and cover 35 is formed with a pump outlet port 37, these ports being formed for connection of fuel inlet and discharge lines thereinto, as indicated.

Body 30 and cover 35 are constructed and arranged to support the pumping elements and the pump capacity control mechanism. The pumping elements include three vane pumps, or vane pump units, each including a vaned rotor mounted on a common drive shaft 40. Shaft 40 is mounted in coaxial, flanged, cylindrical bearings 41 and 42 seated, respectively, in facing, coaxial, stepped cylindrical recesses 38 and 39, recess 38 being in body 30 and recess 39 in cover 35.

The flange of bearing 42 seats against the bottom surface of the larger part of recess 39, and a dowel pin (not shown) is engaged in the flange and in the cover to restrain bearing 42 against rotation. The smaller diameter part, or body, of bearing 42 has a pair of axially spaced circumferential grooves receiving packing rings 43, and between these grooves is a third circumferential groove 44 provided for a purpose to be described.

Bearing 41 has a sliding fit for axial movement in its recess 38, with its cylindrical surfaces having engagement with the cylindrical walls of recess 38. Three coil springs 26, seated in recesses 27 in wall 31 of body 30, engage the smaller end of bearing 41 to bias the bearing into firm engagement with the pumping elements, as will be described. The relation of parts is such that, under the biasing action of springs 26, a space 28 is left between the flange of bearing 41 and the bottom surface of the larger diameter part of recess 38. A dowel pin (not shown) engaged in the bearing flange and in body 30 restrains rotation of bearings 41. Bearing 41 has three circumferential grooves, one in its flange and two in its body, receiving packing ring 46 and 46A, and between the two grooves in the body is a third groove 47 for a purpose to be described.

The outer or front end of shaft 40 is tubular and has a female spline 43 slidably intermeshing with a male spline 51 on the smaller diameter part of a drive coupling 50. The outer, larger diameter end of coupling 50 has a spline 52 to engage the pump driving means. Coupling 50 is loosely held in shaft 40 by a snap ring 45 located in an internal circumferential groove in the shaft, snap ring 45 holding the coupling assembled to the drive shaft while providing for the coupling to be pulled out and replaced without the use of tools.

Sealing means are provided to prevent fluid leakage at the drive end of the pump. The outer end of shaft 40 is reduced in diameter to form a shoulder resting against a sleeve 55 mounted with a close sliding fit on shaft 40, and having a circumferential groove in its bore receiving a packing ring 53. Sleeve 55 is restrained against rotation on shaft 40 by a small pin 54 engaged in a radial hole in the shaft and extending into a longitudinal slot in the bore of the sleeve. The sleeve has a lapped outer face having a fluid tight fit with the inner end of a stationary standard, commercial shaft seal 56 of known construction. Seal 56 is installed in a seal housing 57 centered in an axial bore in wall 31 of body 30 and held in position by screws 58. The tension of seal 56 against sleeve 55 is adjustable by shims 59 between the outer end of the seal and housing 57. As illustrated, packing rings are disposed between seal 56 and housing 57 and between the latter and the surface of its bore.

The threee pumping elements include slotted rotors 60, 70 and 80 mounted side by side on drive shaft 40. Right end rotor 60 is integral with shaft 40, this construction increasing the resistance of shaft 40 to bending loads or moments. Rotor 60 is seen more clearly in section in Fig. 4, center rotor 70 in section in Fig. 12, and left end rotor 80 in Fig. 8 in outside elevation. From Fig. 2, it will be noted that center rotor 70 has twice the axial extent, or width, of end rotors 60 and 80, which is a fact of fundamental importance in the reduction of the net hydraulic loading of shaft 40 to a minimum. Otherwise, the construction and operation of the three rotors are identical. Center rotor 70 and left end rotor 80 have a close sliding fit on shaft 40 and are held against rotation thereon by keys 79 and 89, respectively.

Each rotor 60, 70 and 80 is cooperable with a cylinder 65, 75, 85, respectively, in a pump housing 66, 76, and 86, respectively. Housings 66, 76 and 86 are mounted on a hollow cylindrical pivot 90 for oscillation about the pivot axis to adjust the relative eccentricity of the axes of the rotors and the axes of the pump cylinders. Pivot 90 is seated at its ends in coaxial facing bores 91 and 92 in body 30 and cover 35, respectively, packing rings 93 forming seals between the trunnion portions of pivot 90 and the bores 91 and 92. The interior of pivot 90 communicates with outlet 37, and ports 96, 97, and 98 in its wall communicate with the interiors of housings 66, 76, and 86, respectively. Center housing 76 is assembled permanently with pivot 90 by means of a shrink fit thereon reinforced by means of a dowel screw 94. The end housings 66 and 86 have a fit on pivot 90 such as to allow a rocking motion, of a small angular amplitude, thereon for adjusting the pump capacity as described hereinafter.

As the two end housings are identical in design, only housing 66 will be described in detail. This housing comprises a steel casting formed with parallel flanges 61, in radial planes, between which is formed pump cylinder 65, the flanges having hub portions engaged with pivot 90. At the upper end of housing 66 is formed a cross web 62 machined to form a thrust face for bearing shoe 63 engaging an eccentric 105 on shaft 100. Inlet port 64 and outlet port 67 are cut through the wall of cylinder or cavity 65. Outlet port 67 opens into a chamber 68 formed between flanges 61 and wall or web 69, chamber 68 communicating with the interior of pivot 90 through ports 96. The hub portions of flanges 61 form an extended bearing surface for the housing on pivot 90, and packing rings 99 in grooves in pivot 90 engage these hub portions to form a seal between the housing and pivot 90.

Center pump housing 76 has the same general construction as end pump housings 66 and 86, except that it is twice as wide in an axial direction. Housing 76 is a steel casting including parallel flanges 71 united at their upper ends by pump cylinder 75, the flanges being joined at their lower ends by a cylindrical hub 101 having, as described, a press fit on pivot 90. In the case of center pumping cylinder 75, inlet port 74 is at the bottom of the cylinder and outlet port 77 is at the top, these ports being thus oppositely located relative to the corresponding ports of end cylinders 65 and 85. Inlet port 74 opens directly into the cavity of pump body 30, whereas outlet port 77 opens into a chamber 78 in housing 76 above cylinder 75. Tubing 102 connects chamber 78 into a chamber 103 extending from hub 101 and communicating with port 97 in pivot 90. A cross web 72 between the upper ends of flanges 71 is machined to form a thrust surface for a bearing shoe 73 engaging an eccentric 105 on shaft 100.

Each of the pumping cavities or cylinders 65, 75, is of larger diameter than its rotor 60, 70, or 80, respectively. With the housings 66, 76, 86 pivoted to oscillate about the axis of pivot 90, the axes of the pumping cylinders can be displaced, relative to the common fixed axis of the rotors, between a position of maximum eccentricity of the cylinders and rotors and a position in which the cylinder and rotor axes are concentric or coaxially aligned. At the position of maximum eccentricity, the pump delivery is a maximum, and at the coaxial position, the pump output is zero. Automatic servo-mechanism means are provided to control the relative eccentricity of the cylinder and rotor axes to vary the pump output, these means rotating shaft to rotate eccentrics bearing against shoes 63, 73, 83. The means holding these shoes pressed against the eccentrics will be described more fully hereinafter.

Except for the fact that center rotor 70 has twice the axial width of end rotors 60 and 80, the rotors are identical in construction. Hence, only end rotor 60, shown in diametric section in Fig. 4, will be described in detail. Referring to Fig. 4, the body of each rotor has cut therein four longitudinal slots 111 parallel to its axis and preferably equispaced circumferentially of the rotor. Slots 111 may be in radial planes or in planes parallel to radii. The latter relation is illustrated. The slots are closed at each end by the flanges of bearings 41 and 42, or by spacer disks 112 between the center rotor and each end rotor, but communicate with the pumping cylinder 65 through the medium of longitudinal bores or passages 113, at the inner end of each slot, each connected by one or more pressure balancing passages 114 to the exterior surface of the rotor.

Each slot 111 slidably mounts a pump vane 115 in the form of a flat plate, these vanes, when the pump is in operation, being projected by centrifugal force into engagement with the inner surface of cylinder or cavity 65. The inner ends of vanes 115 are substantially flat with rounded edges, but the outer ends of the vanes are bevelled, as at 116, to form a chisel shaped edge or nose having a slightly rounded point 117 at the leading edge of the vane. The chisel shaped nose of each vane has the advantage of keeping the line of contact between the vane and the surface of cavity 65 substantially in the same location adjacent the leading edge of the vane, which is not the case when the vane has a rounded nose, as is usually provided. Furthermore, the chisel shaped nose plays an important part in the pressure balancing of the vane.

Referring more particularly to Fig. 19, when rotor 60 is mounted in cavity 65 and the two parts are in eccentric or output producing relation, there is an eccentric space 118 between the rotor and cavity divided by vanes 115 into four parts A, B, C and D of nearly equal circumferential length. In the rotor position illustrated in Fig. 19, the edges 117 of vanes 115-1 and 115-2 contact the wall of cavity 65 exactly at the edges defining cavity inlet port 64, and the edges of vanes 115-3 and 115-4 contact the cavity wall exactly at the edges defining cavity outlet port 67.

In this position of the rotor 60, portion A of space 118 is at inlet pressure, with the inlet pressure acting directly against the outer end of vane 115-2 and, through passages 113 and 114, against the inner end of this vane. The pressures on the outer and inner ends of vane 115-2 are thus balanced, and this balance would ;be 100% efiective if vane edge 117 were actually a sharp edge. However, for practical reasons, edge 117 is rounded to a very small radius, thus slightly reducing the area of the outer end of the vane subjected to inlet pressure as indicated by the dimension lines. This slight reduction, resulting in a very slight pressure diflerential on the ends of the vane, has no detrimental effect on the pump operation. However, if the vanes had the usual fully rounded nose, the pressure differential on the inner and outer ends would be substantial and would have a very substantial elfect on the pump operation. With such a fully rounded nose, the line of contact moves constantly, during a revolution of the rotor, between a minimum and a maximum distance from the leading edge of the vane.

The other three vanes are likewise pressure balanced. In the illustrated position of rotor 60, the ends of vane 115-3 are both at the pressure existing in part B of space 118, and the ends of vane 115-1 are at the pressure existing in part D of space 118; as vanes 115-2 and 115-3 shut off space B and vanes 115-4 and 115-1 shut off space D. Both ends of vane 115-4 are at the outlet pressure at the outlet port 67, through the medium of passages 113 and 114.

if rotor 40 is turned a few degrees clockwise, vanes 1151 and 115-3 will clear the edges of inlet port 64 and outlet port 67, respectively, thus placing part D of space 118 at inlet pressure and part B at outlet pressure. Part A remains at inlet pressure and part C at outlet pressure, as before. The pressures on the inner and outer ends of each of the four vanes remain substantially balanced.

As the pressures on the inner and outer ends of the vanes 115 are balanced for any position of rotor 60, and for any position of the vane edges 117 relative to the ports 64 and 67, there is no pressure differential operating to move the vanes in their slots 111 in either direction, except for the very slight and negligible pressure differential due to the slight rounding of edges 117. Thus, the only practically effective force acting on vanes 115 is the centrifugal force resulting from rotation of rotor 40 during operation of the pump.

In the position of pump cylinder 65 shown in Fig. 19, wherein the rotor 60 and cavity 65 have their maximum concentricity, the surface of rotor 60 nearest the surface of cavity 65 is only a few thousandths of an inch therefrom, and the pump delivery is a maximum. If body 66 is swung counterclockwise, about the axis of pivot 90 until cavity 65 is concentric with rotor 60, the pump delivery will be zero as there will no longer be any reciprocation of vanes 115 and thus no change in the volume of the spaces enclosed by rotor 60, cavity 65, vanes. 115, the

flange of bearing 41 and spacer disk 112.

When the rotors are in rotation, the reaction of the fluid against vanes 115 forces the vanes against the trailing surfaces of slots 111, thus providing good fluid sealing between the vanes and such trailing surfaces. Special sealing means, of the type shown and described in my U. S. Patent No. 2,708,884, are provided between the outer portions of the leading surfaces of slots 111 and the associated vanes to minimize leakage along the vane surfaces.

Looking toward the closed driving or front end of the pump body 30, center pump housing 76 is biased in a counter-clockwise direction, and both end pump housings 66 and 86 are biased in a clockwise direction by springs 121 seated in recessed pockets 122 in the outer surfaces of webs 62, 72 and 82, the springs having their outer ends seated in caps 123 screwed into threaded bosses in the side walls of pump body 30. Such biasing is augmented by the hydraulic pressure existing in the eccentric purmping spaces 118 of the pump cylinders when the pump is in operation. When the pump is stopped, springs 121 bias the pump housings to the position of maximum eccentricity of the axes of the rotors and pumping cylinder.

Lubrication of the bearings Fuel at inlet pressure enters the pump through inlet port 36, filling the entire pump cavity, and enters pump cylinders 65, and through inlet ports 64, 74 and 84, respectively. The three pumping units deliver fuel at outlet pressure to hollow pivot 90, as desciibed more fully hereinafter.

Bore 91, behind pivot 90, is connected by a drilled passage 126 (Figs. 6 and 11) to a filter chamber 125 containing a porous filter element 127, the outer end of pas sage 126 being closed by a plug 128. Filter 127 cleans the small amount of fuel required to lubricate and cool the pump shaft bearings and to operate the servomechanism adjusting the pump output. After passing through element 127, the fuel flows through channel 131 and passages 132, 133 to a pressure reducing valve 130, which reduces the pressure of the fuel from the pump outlet pressure to about 80 p.s.i. for hearing lubrication, and cooling. This pressure reduction reduces considerably the internal leakage and thus improves the volumatic efliciency of the pump.

The fuel at outlet pressure in passage 132 acts on the outer end of a small diameter piston 134, slidable in a guiding extension of a ported sleeve 135, to move piston 134 to the left, as viewed in Figs. 9 and 10, thus forcing large diameter hollow piston 136 to the left. This compresses a spring 137 seated in piston 136 and in cup shaped spring seat 138. As soon as a ported circumferential groove 139 in the outer surface of piston 136 overlaps a circumferential groove 141 in the inner surface of sleeve 135, fuel at outlet pressure in passage 133 flows through passage 142, grooves 141 and 139, and ports 143 into piston 136.

Fuel for lubricating and cooling the shaft bearings exits from piston 136 and sleeve through ports 144 and passage 148 into hollow dowel between body 30 and cover 35. The fuel for lubricating and cooling the cover bearing 42 flows from dowel 140, through passages 146 and 147 to groove 44 in bearing 42, and thence through radial passages to a shallow longitudinal groove 149 in the bearing surf-ace (Figs. 2, 14 and 15). Fuel flows from dowel 140 to body bearing 41 through a crosshole 151 (Figs. 14 and 15), channel 152, and passage 1:53. entering circumferential groove 47 and flowing through radial passages 154 to a shallow longitudinal groove in the bearing surface.

As soon as the resistance to fuel flow to the bearings is high enough to build up a pressure of 80 p.s.i. (or Whatever pressure valve 130 may be set for by selection of spring 137 and selective use of shims with the spring), this reduced pressure within large piston 136 balances the pump discharge pressure on piston 134 so that these pistons move to the right under the influence of spring 137. This movement either closes off groove 141 or else leaves grooves 139 and 141 overlapped an amount just suflicient to maintain the regulated pressure against the leakage through the bearings.

Fuel if leaking past pistons 134 and 136 into space 202, between the closed end of piston 136 and the bottom of sleeve 135, would create an hydraulic lock, keeping the pressure reducing valve from operating as intended. Therefore space 202, through radial holes 203 in the wall of sleeve 135, communicates with annulus 204 which is vented to the pump cavity by means of a suitable passage (not shown).

Fuel leaking past cover bearing 42 fills space 155 at the inner end of this hearing and returns to the pump cavity through drilled passage 156. Similarl fuel leaking past body bearing 41 collects in space 157 between the outer end of this hearing and shaft seal housing 57, from which it flows to the pump cavity through drilled passage 158. It is important that the pressure of fuel in space 157 be kept low, and at substantially the pump 11 inlet pressure, in order to prevent leakage past shaft seal 56.

Balancing of the hydraulic shaft loading As previously described, the three pump housings 66, 76, and 86 are mounted side by side on pivot 90, and the three rotors 60, 70 and 80 are mounted side by side on shaft 40, with rotor 70 being twice the width (along shaft 40) of end rotors 60 and 80 which are equal in width. Also, the outlet port 77 of center pump cylinder 75, which is the high pressure side of this cylinder, is 180 opposite the outlet ports 67 and 87 (high pressure sides) of end pump cylinders or cavities 65 and 85. All three outlet ports, due to their common connection to hollow pivot 90, are at the same high outlet pressure during operation of the pump.

Referring to the diagrams of Figs. 16 and 17, the total outlet pressure loading transmitted to shaft 40 through center rotor 70, as indicated by the loading lines E, is exactly twice the loading transmitted to shaft 40 through each end rotor 60 or 80, as indicated by the loading lines F. The opposing loads are thus equal and opposite, as indicated at E and F in Fig. 17, and cancel each other out as far as bearing loads on shaft 40 are concerned. The loads E and F, of course, impose a bending stress on shaft 40 which thus has to be designed to carry the bending stress without deformation or excessive strain for the material used for the shaft. Such design can be easily effected without undue difiiculty.

The foregoing analysis applies only when the vanes of all three rotors are in phase with each other. However, and as previously described, the vanes 115 of center rotor 70 are offset, by half the angular distance between adjacent vanes, relative to the vanes 115 of rotors 60 and 80, which are in phase with each other. With four-vane rotors having adjacent vanes at 90 to each other, such offset is 45. This means that the loading E" (Fig. 18) due to the center rotor is no longer directly opposite to the combined loading F" due to the end rotors, and whose magnitude is equal to that of loading E. The equal magnitude loadings E" and F" are now at an angle of 135 to each other. This produces a resultant hydraulic loading G on the shaft bearings which, however, is only a fraction of what the bearing loading would be if the discharge outlets of all three pump cylinders were on the same side of shaft 40.

Taking a practical example of a high pressure vane pump embodying the invention, the center rotor 70 applies a load of 4125 lbs. to the shaft at the mid-point of the center rotor. Each of the two end rotors 60 and 80 applies a load of 2062.5 lbs. to shaft 40 at a distance 1.344 from the mid-point of the center rotor, these points being on opposite sides of rotor 70 and the loads being applied at an angular spacing of 135 from the direction of application of the loading from the center rotor. The end loadings cause bending loads at 1.344 each side of the midpoint of the center rotor. If a single rotor having a width equal to the combined widths of the three rotors were used, a load of 8250 lbs. would be applied to shaft 40 at the rotor midpoint, resulting in bearing loads of 4125 lbs. at points spaced 1.344 from such midpoint. Such bearing loads cannot be successfully carried by known bearing materials within the design parameters and under the specified operating conditions of a pump of this type.

The shaft bearings have been the limiting factor in designing vane pumps even at output loads as low as 600 to 700 p.s.i., quite considerably below the 1000 p.s.i., or higher, outlet pressures of the invention pump. Even with the above-mentioned 45 offset of the vanes, the bearing PV factor, at rated pump speed and pressure, is only 18,500.

The three pump housings 66, 76 and 86 are held in fluid tight relation with bearings 41, 42, and with spacer disks 112, by the three springs 26 bearing against the outer end of bearing 41, supplemented by hydraulic pressure in the space 28 behind the flange of bearing 41. Channel 159 drilled to connect space 28 with bore 91 of pivot provides for the pump discharge or outlet pressure to load the bearing flange. The area of this flange is calculated to provide the proper loading, which is somewhat higher than necessary to balance the opposing fluid pressures within the pump housings 66, 76 and 86 but low enough so as not to cause too much frictional resistance of the elements to oscillation of the pump housings by the output control mechanism to adjust the pump displacement. The pump rotors have a very small clearance, at their ends, with bearings 41, 42 and spacer disks 112 to permit free rotation while still providing as good a fluid seal as possible therewith.

Output control servo-mechanism The pump output, in normal operation, is adjusted in correspondence with the engine fuel demand by a servomechanism linked to the engine master fuel control. A manual override for the master control may be provided for use in starting the engine. Suitable linkage connects the engine master control to a lever on a shaft extension 161 of a rotary tubular valve rotatable in substantially fluid-sealed relation in a valve sleeve 180 rotatably mounted in fluid sealed relation in a bore 162 in body 30, shaft extension 161 extending through a shaft seal 163 in a boss 164 on cover 35. The outer end of bore 162 is closed by a plug 174 held in place by a snap ring. Valve 160 and its sleeve 180 constitute the control means for a hydraulic servo-mechanism for adjusting the relative eccentricity of the common axis of rotors 60, 70, 80 and their associated pumping cylinders or cavities 65, 75, 85, respectively, to adjust the pump output.

An indicator 165 has a hub keyed to shaft extension 161 and is cooperable with an arcuate scale 166 secured to cover 35 for limited angular adjustment about the axis of valve 160. A collar 167, adjustably clamped around boss 164, carries abutments 168 engageable by indicator 165 at either end of scale 166. A nut 169 threaded on the outer end of shaft extension 161 forces the serrated hub of lever 120 into firm meshing engagement with a serrated spacer 171 meshing with the serrated hub of indicator 165. Handle 120 is thus angularly adjustable relative to valve 160 and its indicator 165.

Fuel at discharge or outlet pressure is supplied to servovalve 160-180 by a channel 172 connected to channel 131 and to a passage 173 aligned with an outer circumferential groove 181 in sleeve 180. From groove 181, fuel flows through sleeve ports 182 into circumferential groove 176 in valve 160 and then through ports 177 into the interior of the valve, which is closed near its outer end by a threaded plug associated with suitable packing means.

In a diametric plane spaced axially from that of groove 176, valve 160 has a pair of diametrically opposite ports having oval peripheries with major axes extending axially of the valve. Ports 175 are adapted to be selectively partially or completely registered circumferentially with holes in sleeve 180. These holes open into an external circumferential groove 183, in sleeve 180, communicating with a channel 191. Beyond plug 17 0, valve 160 has a pair of diametrically opposite ports 178 shaped like ports 175 and displaced 90 relative to ports 175. Ports 178 are adapted to be selectively partially or completely registered with holes 188 in sleeve 180. These holes open into an external circumferential groove 184, in sleeve 180, communicating with passage 192.

The leading edges of ports 175 and 178 must register exactly with the edges of holes 185 and 188, respectively, the holes of each pair being diametrically opposite each other. The oval or oblong shape of ports 175 and 178 enables this circumferential registry with the holes to be obtained Without the necessity .of too high precision in the axial spacings between ports 175 and 178 and between holes 185 and 188. The diametrically opposed relation of the ports provides hydraulic balance, thus preventing any drag restricting free rotation of valve 160.

Between groove 176 and shaft extension 161, valve 160 has a relatively wide circumferential groove 179 communicating with holes 186 in sleeve 180. Holes 186 open into a circumferential groove 187, in sleeve 180, registering with a passage 193.

When lever 120 is turned clockwise rotating valve 160 in the same direction and degree, ports 175 communicate With holes 185, and high pressure fuel flows into channel 191 and thence into channels 194 and 196 (Fig. 13). Channel or passage 196 communicates (not shown) with a circumferential groove 197 ,(Fig. 6),, .surrounding a servocylinder 190. From this groove, the fuel flows through holes 198 into the space beneath a servo-piston 195. Piston 195 operates shaft 100 by means of a connecting rod 106 and a crank 107. The high pressure fluid beneath piston 195 moves the piston outwardly to rotate shaft 100 and its eccentrics 105 in a direction to decrease the relative eccentricity of the pump rotor and pump cylinder or cavity axis to reduce the pump output as called for by the engine control.

A gear 108 is mounted on the end of shaft 100 opposite to the crank end, and this gear, through an idler gear 199, rotates a gear 189 formed on valve sleeve 180. The limit of motion of shaft 100 is 45, and the overall gear ratio is 3:4 so that, for 45 rotation of shaft 180, sleeve 181 is turned 60 in the same direction.

Assuming the engine control, through lever 120, rotates valve 160 9 clockwise allowing fuel under pressure to flow to cylinder 190 beneath piston 195, piston 195, moving outwardly, turns shaft 100 clockwise. After 6 of movement of shaft 100, follow-up valve sleeve 180 will have turned 9 clockwise and be in its original angular relation to valve 160. Thus, communication between ports 175 and holes 185 is interrupted and, as no more high pressure fuel flows to cylinder 190, motion of piston 195 ceases leaving the pumping elements at the position of axis eccentricity necessary for the required output. The fuel beneath piston 195, and in ports 185, groove 183, and channels 191, 194 and 196 now acts as a hydraulic lock.

Should piston 195 leak or ports 178 in valve 160 not seal properly in sleeve 180, fuel will slowly leak out from beneath piston 195. The pressure in the pumping cavities 65, 75, 85, acting on eccentrics 105, would tend to rotate shaft 100 counter-clockwise. If this were allowed to happen, the pump output would be increased beyond that called for by the engine control. However, the servomechanism automatically compensates for these conditrons.

Should shaft 100 rotate the slightest amount counterclockwise, as above, sleeve 180 is also rotated counterclockwise. Valve 160 is held stationary by the fuel control, so that ports 175 again communicate with holes 185, and high pressure fuel flows to servo-cylinder 190. Piston 195 is moved outward to rotate shaft 190 clockwise back to its previously set position, in turn rotating sleeve 180 clockwise to shut off the fuel flow. Actually, instead of hunting back and forth, sleeve 180 takes a position allowing continuous seepage of fuel to cylinder 190 in an amount just sufficient to compensate for the leakage.

If the engine requires more fuel, lever 128 is operated by the fuel control to rotate valve 160 counterclockwise. This motion of valve 160 moves ports 175 further away from holes 185, shutting off fiow of high pressure fuel to servo-cylinder 191). At the same time, ports 178- are brought into communication with holes 188. Fuel now returns from beneath piston 195 through holes 198, groove 197, passages or channels 196, 194 and 192, groove 184, holes 188, ports 178, valve 160, bore 162, and channel 201 into the pump cavity. Piston 190 is moved inwardly, increasing the relative eccentricity .of the pump rotors and their cavities to increase the pump output. Eccentric shaft 100, being thus rotated counterclockwise, correspondingly rotates follow-up sleeve 180 to close off ports 178 at the new position of valve 160. The timing of the servo ports has to be correct to avoid a lag in the control when shifting from a decrease to an increase in pump output. This requires that, in the neutral position of the servo-mechanism 16l180, the .controlling edges of ports 175 and 178 should just meet the controlling edges of holes 185 and 188, respectively, as shown in Figs. 21 and 22.

In case of fuel seepage at valve 160 or piston 195, allowing the piston to move inwardly, the compensating action of sleeve 180 occurs, in a manner analogous to that previously described, to maintain the pump elements in the adjusted position.

A similar automatic corrective action takes place should the servo valve 160 allow fuel under pump discharge pressure to leak past closed ports 175 and, flowing into the servo cylinder, unduly cause piston 195 to move outwardly decreasing the pump output beyond the requirements of the engine master control, or when not called for. This outward motion of the servo piston causes eccentric shaft and follow-up sleeve 180 to turn in a clockwise direction, opening ports 178 and allowing fuel from under piston 195 to escape until the pump capacity adjustment has returned to the position of the servo valve 160, which has not been moved.

No hunting occurs as sleeve 180 takes position allowing continuous leakage back to the pump cavity, in the amount just suflicient to compensate for the unwanted excess of high pressure fuel leaking past valve ports 175.

High pressure fuel leaking to the right along valve 160 enters groove 179 and returns to the pump cavity through ports 186, sleeve groove 187 and passage 193. Leakage of high pressure fuel to the left along the valve flows into bore 162 and returns to the pump cavity through passage 201.

Pump operation during engine stopping and starting In stopping the engine, valve 160 is moved to the zero position of indicator 165 on scale 166. Piston 195 is thus moved outwardly its limit of movement, where the pump rotors and their cavities are axially concentric. The fuel pressure thus drops to zero, and springs 121 acting against shoes 122 bearing on eccentrics tend to rotate shaft 100 counter-clockwise, which action can be augmented, if desired, by a torsion spring on shaft 100. Although valve 169 has a very close lapped fit in sleeve 180, fuel from cylinder 191) will leak past the vave and return to the pump cavity through bore 162 and passage 201. After a while, piston 195 will come to rest at the end of its inward stroke, leaving the pump in maximum output position.

At the next starting of the engine, the master fuel control, or an overriding manual throttle operates lever to move valve to the required reduced fuel demand position, but the pumping elements are in the maximum output position. The engine is cranked at about 15% of its rated operating speed, the pump being then rotated for example, at 600 r.p.m. As soon as the pump delivers fuel and builds up pressure, piston 195 is moved outwardly to adjust the pumping elements to the required output position as set by lever 12%). The follow-up sleeve moves to shut off ports at this position, so that the pump immediately attains and maintains the prerequisite output position.

The invention has been described, by way of giving a typical practical illustration, as applied to a pump having four blades per rotor. However, it will be readily perceived that the invention principles are not limited to a four blade per rotor pump, and are equally applicable where a greater or less number of blades are used per rotor. A practical example is a pump having six blades per rotor.

Furthermore, while a specific embodiment of the invention has been shown and described in detail to illustrate the application of the invention principles, it will be understood that the invention may be embodied otherwise without departing from such principles.

What is claimed is:

1. A high pressure vane pump comprising, in combination, housing means forming a fluid inlet cavity; a plurality of equal diameter cylindrical pumping cavities in axially adjacent relation in said housing means, each cavity having a fluid inlet in communication with said inlet cavity and a fluid outlet diametrically opposite its fluid inlet; a common pump outlet in communication with all of said fluid outlets; a drive shaft extending through said cavities parallel to the axes thereof and mounted in bearings in said housing means; a plurality of equal diameter pump rotors secured in axially adjacent'relation on said drive shaft for rotation therewith, each rotor being rotatable in a different one of said pumping cavities and the diameter of therotors being less than that of said cavities; and vanes slidably mounted in circumferentially equi-spaced longitudinal slots in each rotor for engagement of their outer ends with the surface of the associated pumping cavity; the fluid outlet of at least one of said pumping cavities being diametrically opposite the fluid outlets of the other pumping cavities, and the axial width of the rotor associated with said one pumping cavity being equal to the combined axial width of the other rotors; whereby the hydraulic loads on said drive shaft, during operation of the pump, are substantially in balance; said other rotors having substantially equal axial widths and the vanes of said one rotor being offset angularly relative to the vanes of said other rotors by one-half the circumferential spacing of the vane outer edges.

2. A high pressure vane pump comprising, in combination, housing means forming a fluid inlet cavity; three equal diameter cylindrical pumping cavities in axially adjacent relation in said housing means, each cavity having a fluid inlet in communication with said inlet cavity and a fluid outlet diametrically opposite its fluid inlet; a common pump outlet in communication with all of said fluid outlets; a drive shaft extending through said cavities parallel to the axes thereof and mounted in bearings in said housing means; three equal diameter pump rotors secured in axially adjacent relation on said drive shaft for rotation therewith, each rotor being rotatable in a different one of said pumping cavities and the diameter of the rotors being less than that of said cavities; vanes slidably mounted in circumferentially equi-spaced longitudinal slots in each rotor for engagement of their outer ends with the surface of the associated pumping cavity; the fluid outlet of the intermediate pumping cavity being diametrically opposite the fluid outlets of the two end pumping cavities; means mounting said cavities and rtors for relative bodily displacement of the cavity and rotor axes to vary the pump displacement; biasing means continuously operable to urge relative movement of the cavities and rotors toward the position of maximum pump displacement; and pump output adjusting mechanism selectively operable to effect relative movement of the cavities and rotors toward the position of minimum pump displacement in opposition to said biasing means; the axial width of the intermediate rotor being equal to the combined axial width of the two outer rotors, and the outer rotors being equal in width; whereby the hydraulic loads on said drive shaft, during operation of the pump, are substantially in balance, thus removing the hydraulic loads from the drive shaft bearings.

3. A high pressure vane pump as claimed in claim 2 in which said end rotors have substantially equal axial widths.

4. A high pressure vane pump as claimed in claim 2 in which the vanes of said intermediate rotor are offset angularly relative to the vanes of the end rotors by onehalf the circumferential spacing of the vane outer edges.

5. A high pressure vane pump comprising, in combination, housing means forming a fluid inlet cavity; a plurality of equal diameter cylindrical pumping cavities in axially adjacent relation in said housing means, each cavity having a fluid inlet in communication with said inlet cavity and a fluid outlet diametrically opposite its fluid inlet; a common pump outlet in communication with all of said fluid outlets; a drive shaft extending through said cavities parallel to the axes thereof and mounted in bearings in said housing means; a plurality of equal diameter pump rotors secured in axially adjacent relation on said drive shaft for rotation therewith, each rotor being rotatable in a different one of said pumping cavities and the diameter of the rotors being less than that of said cavities; vanes slidably mounted in circumferentially equi-spaced longitudinal slots in each rotor for engagement of their outer ends with the surface of the associated pumping cavity; the fluid outlet of at least one of said pumping cavities being diametrically opposite the fluid outlets of the other pumping cavities, and the axial width of the rotor associated with said one pumping cavity being equal to the combined axial width of the other rotors; whereby the hydraulic loads on said drive shaft, during operation of the pump, are substantially in balance; means mounting said cavities and rotors for relative bodily displacement of the cavity and rotor axes to vary the pump displacement; biasing means continuously operable to urge relative movement of the cavities and rotors toward the position of maximum pump displacement; fluid outlet pressure operated servo-mechanism operable to effect relative movement of the cavities and rotors toward the position of minimum pump displacement, in opposition to said biasing means; and control means selectively operable to control flow of fluid at pump outlet pressure to said servo-mechanism to effect relative movement of the cavities and rotors toward such position of minimum pump displacement, and to exhaust fluid from said servo-mechanism to effect relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means; whereby seepage of fluid from said servomechanism through said control means, when the pump is not operating, will provide relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means for maximum fluid delivery when the pump is re-started.

6. A high pressure vane pump comprising, in combination, housing means forming a fluid inlet cavity; a plurality of equal diameter cylindrical pumping cavities in axially adjacent relation in said housing means, each cavity having a fluid inlet in communication with said inlet cavity and a fluid outlet diametrically opposite its fluid inlet; a common pump outlet in communication with all of said fluid outlets; a drive shaft extending through said cavities parallel to the axes thereof and mounted in bearings in said housing means; a plurality of equal diameter pump rotors secured in axially adjacent relation on said drive shaft for rotation therewith, each rotor being rotatable in a different one of said pumping cavities and the diameter of the rotors being less than that of said cavities; vanes slidably mounted in circumferentially equi-spaced longitudinal slots in each rotor for engagement of their outer ends with the surface of the associated pumping cavity; the fluid outlet of at least one of said pumping cavities being diametrically opposite the fluid outlets of the other pumping cavities, and the axial width of the rotor associated with said one pumping cavity being equal to the combined axial width of the other rotors; whereby the hydraulic loads on said drive shaft, during operation of the pump, are substantially in balance; means mounting said cavities and rotors for relative oscillatory displacement of the cavity and rotor axes about an axis spaced from and parallel to said drive shaft to vary the pump displacement; biasing means continuously operable to urge relative movement of the cavities and rotors toward the position of maximum pump displacement; fluid outlet pressure operated servo-mechanism operable to effect relative movement of the cavities and rotors toward the position of minimum pump displacement, in opposition to said biasing means; and control means electively operable to control flow of fluid at pump outlet pressure to said servo-mechanism to effect relative movement of the cavities and rotors toward such position of minimum pump displacement, and to exhaust fluid from said servo-mechanism to effect relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means; whereby seepage of fluid from said servomechanism through said control means, when the pump is not operating, will provide relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means for maximum fluid delivery when the pump is re-started.

7. A high pressure vane pump as claimed in claim 6 in which said common pump outlet is a hollow pivot forming the axis for such relative oscillation.

8. A high pressure vane pump comprising, in combination, housing means forming a fluid inlet cavity; three equal diameter cylindrical pumping cavities in axially adjacent relation in said housing means, each cavity having a fluid inlet in communication with said inlet cavity and a fluid outlet diametrically opposite its fluid inlet; a common pump outlet in communication with all of said fluid outlets; a drive shaft extending through said cavities parallel to the axes thereof and mounted in bearings in said housing means; three equal diameter pump rotors secured in axially adjacent relation on said drive shaft for rotation therewith, each rotor being rotatable in a different one of said pumping cavities and the diameter of the rotors being less than that of said cavities; vanes slidably mounted in circumferentially equi-spaced longitudinal slots in each rotor for engagement of their outer ends with the surface of the associated pumping cavity; the fluid outlet of the intermediate pumping cavity being diametrically opposite the fluid outlets of the two end pumping cavities, and biasing means continuously operable to urge relative movement of the cavities and rotors toward the position of maximum pump displacement; the axial width of the intermediate rotor being equal to the combined axial width of the two outer rotors, and the outer rotors being equal in width; whereby the hydraulic loads on said drive shaft, during operation of the pump, are substantially in balance; means mounting said cavities and rotors for relative oscillatory displacement of the cavity and rotor axes about an axis spaced from and parallel to said drive shaft to vary the pump displacement; biasing means continuously operable to urge relative movement of the cavities and rotors toward the position of maximum pump displacement; fluid outlet pressure operated servo-mechanism operable to effect relative movement of the cavities and rotors toward the position of minimum pump displacement, in opposition to said biasing means; and control means selectively operable to control flow of fluid at pump outlet pressure to said servomechanism to effect relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means; whereby seepage of fluid from said servo-mechanism through said control means, when the pump is not operating, will provide relative movement of the cavities and rotors toward the position of maximum pump displacement under the influence of said biasing means for maximum fluid delivery when the pump is re-started.

9. A high pressure vane pump as claimed in claim 8 in which said common pump outlet is a hollow pivot forming the axis for such relative oscillation; said pumping cavities are formed in individual housings each oscillatable about the axis of said hollow pivot and in communication with the interior thereof, each pumping cavity being in communication with the interior of its housing through its fluid outlet; the intermediate housing, during such relative axial displacement of its cavity and rotor, being moved by said biasing means and said servo-mechanism in a direction opposite to the direction of corresponding displacement varying movement of the two end housings.

10. A high pressure vane pump as claimed in claim 8 including follow-up means operable by said servo-mechanism responsive to attainment by the cavities and rotors of the relative axial displacement selected by said control means, to effect interruption of fluid flow between said control means and said servo-mechanism.

11. A high pressure vane pump as claimed in claim 9 in which said servo-mechanism comprises a piston movable in a cylinder; a cam shaft rotatably mounted in said housing means parallel to said drive shaft and hollow pivot; cams on said cam shaft each engaged with one of said housings; and means connecting said cam shaft to said piston for oscillation by the latter.

12. A high pressure vane pump as claimed in claim 11 including follow-up means operable by said cam shaft, responsive to attainment by the cavities and rotors of the relative axial displacement selected by said control means, to effect interruption of fluid flow between said control means and said servo-mechanism.

13. A high pressure vane pump as claimed in claim 12 in which said control means comprises a sleeve oscillatably mounted in said housing means and formed with port means in communication with said cylinder beneath said piston; a tubular valve having a substantially fluidtight fit in said sleeve; a control element operable to oscillate said valve relative to said sleeve; and means connecting separated sections of the interior of said valve to said hollow pivot and to said fluid inlet cavity; said valve having port means registerable with the port means of its sleeve, upon oscillation of said valve by said element, to selectively connect the interior of said cylinder to either of the sections of the valve interior; said follow-up means rotating said sleeve in the same direction as the movement of said valve by said control element to move the sleeve port means out of registry with the valve port means.

14. A high pressure vane pump as claimed in claim 13 in which said follow-up means comprises a gear train operatively interconnecting said cam shaft and said sleeve.

15. A high pressure vane pump as claimed in claim 13 in which the registerable port means of said sleeve and valve each include a pair of diametrically opposite ports for hydraulic balance.

16. A high pressure vane pump as claimed in claim 14 in which the valve ports are axially elongated to facilitate registry with the sleeve ports.

17. A high pressure vane pump as claimed in claim 1 including passage means interconnecting said common pump outlet and said bearings for lubrication of the latter by the pumped fluid; and outlet pressure responsive pressure reducing means interposed in said passage means.

References Cited in the file of this patent UNITED STATES PATENTS 1,132,413 Whipple Mar. 16, 1915 1,779,757 Streckert Oct. 28, 1930 2,142,275 Lane Ian. 3, 1939 2,322,568 De Lancey June 22, 1943 2,600,632 French June 17, 1952 2,612,114 Ernst Sept. 30, 1952 2,691,482 Ungar Oct. 12, 1954 2,708,884 Deschamps May 24, 1955 FOREIGN PATENTS 162,12& Australia Mar. 23, 1955

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