US3779674A

US3779674A - High-pressure gear pump

Info

Publication number: US3779674A
Application number: US00163750A
Authority: US
Inventors: O Eckerle; R Jung
Original assignee: Individual
Current assignee: Individual
Priority date: 1970-07-17
Filing date: 1971-07-19
Publication date: 1973-12-18
Anticipated expiration: 1990-12-18
Also published as: DE2035575A1; FR2101659A5; GB1346474A; DE2035575C2; JPS5522633B1

Abstract

A high-pressure gear pump with a drive pinion, internal gear ring, and filler wedge, where both the axial thrust and radial thrust are substantially compensated; a deflecting axial thrust plate cooperates with an O-ring to define an axial compensation pressure field; a pair of transverse sealing members, either elastic or radially sliding in grooves, cooperates with the adjacent faces of the periphery of the internal gear assembly and of the housing bore to define a radial compensation pressure field. A bearing ring around the internal gear, together with means for directionally restricting the displacements of the internal gear assembly, allow maintenance of a fixed gear center distance and backlash.

Description

United States Patent 1191 Eckerle et al.

[ HIGH-PRESSURE GEAR PUMP [75] Inventors: Otto Eckerle, Am Bergwald 3,

Malsch; Robert Jung, Malsch, both of Germany [73] Assignee: said Eckerle, by said Jung [22] Filed: July 19, 1971 [21] Appl. No.: 163,750

[30] Foreign Application Priority Data July 17, 1970 Germany P 20 35 575.3

[52] U.S. Cl 418/71, 418/102, 418/109, 418/133, 418/170 [51 Int. Cl. F0lc 21/00, F03c 3/00, F04c 1/06 [58] Field of Search 418/71, 108, 109, 418/131, 133, 169471, 102

[56] References Cited UNITED STATES PATENTS 3,315,608 4/1967 Eckerle 418/169 1,719,640 7/1929 Wilsey 418/169 2,132,813 10/1938 Wahlmark 418/71 Dec. 18, 1973 1,970,146 8/1934 Hill ..41s/133 3,676,027 7/1972 Molly ..418/71 Primary ExaminerCarlton R. Croyle Assistant Examiner-John J. Vrablik Att0rneyArthur Schwartz et a1.

[57] ABSTRACT A high-pressure gear pump with a drive pinion, internal gear ring, and filler wedge, where both the axial thrust and radial thrust are substantially compensated; a deflecting axial thrust plate cooperates with an O- ring to define an axial compensation pressure field; a pair of transverse sealing members, either elastic or radially sliding in grooves, cooperates with the adjacent faces of the periphery of the internal gear assembly and of the housing bore to define a radial compensation pressure field. A bearing ring around the internal gear, together with means for directionally restricting the displacements of the internal gear assembly, allow maintenance of a fixed gear center distance and backlash.

21 Claims, 14 Drawing Figures PATENTEBuEcwms 1 3179.674

' SHEET ll]? 6 INVENTORS:

OTTO ECKERLE ROBERT JUNG IMENIEDUEc 13 m5 3779574 sum 2 OF 6 OTTO ECKERLE- ROBERT JUNG BY: I

SHEET 38F 6 PATENTEUBEE18 I975 INVENTORS:

OTTO ECKERLE R BERT JUNG ML.) AGENT PATENTEDUEC 18 I975 SHEET u UF 6 MM 'W i g m i 1 Q X 5 I INVENTORS I OTTO ECKER LE ROBERT HUN 6 AGENT PATENTEDUEE 18 I975 3; 779,674 sum 5 0F 6 INVENTORS:

OTTO ECKERLE ROHERT JUNG l HIGH-PRESSURE GEAIR PUMP BACKGROUND OF THE INVENTION 1. Field of the Invention The invention relates to gear pumps, and in particular to high-pressure gear pumps with axial and radial thrust compensation, where a drive pinion engages a driven internal gear member which is substantially hydrostatically balanced, and where a pivotably mounted filler wedge is arranged between the two gears.

2. Description of the Prior Art From the prior art are known various high-pressure gear pumps, where the axial thrust compensation is obtained by means of one or two axially movable thrust plates which are under the influence of pressure fields, while radial thrust compensation is obtained, for example, by the provision of a thrust compensation plunger which is likewise under the influence of one or several pressure fields and which acts against the internal gear member.

In view of the fact that complex sealing elements are necessary for the sealing of these axial and radial compensation fields, it is not always possible to provide these known sealing elements in simpler or smaller pumps, the design becoming either too complicated and costly or too difficult to assemble.

SUMMARY OF THE INVENTION It is a primary objective of the invention to provide a simplified, economical and yet operationally reliable high-pressure gear pump where, even under comparatively high pressures, at good volumetric efficiency is obtained, the pump operating satisfactorily even under elevated temperatures.

To attain this objective, the invention suggests a highpressure gear pump where the axial thrust compensation is obtained by providing at least one peripherally restrained relatively thin metal plate which is subjected to an axial compensation pressure field arranged in the housing cover, while the radial compensation pressure field is arranged on the periphery of the internal gear ring or on the periphery of the bearing ring which surrounds the internal gear member and deliniated by two sealing blocks which extend over at least the entire width of the gears.

BRIEF DESCRIPTION OF THE DRAWINGS Further special features and advantages of the invention will become apparent from the description following below, when taken together with the accompanying drawings which illustrate, by way of examples, several embodiments of the invention, represented in the various figures as follows:

FIG. 1 represents a first embodiment of the invention, shown in a radial cross-section along the line I-I of FIG. 2;

FIG. 2 shows the embodiment of FIG. I in an axial cross-section along the line Il-II of FIG. 1;

FIG. 3 represents a second embodiment of the invention, shown in a radial cross-section along the line III- -III of FIG. 4;

FIG. 4 shows the embodiment of FIG. 3 in an axial cross-section along the line IVIV of FIG. 3;

FIG. 5 shows a modified embodiment similar to that of FIG. 3;

FIG. 6 represents a further embodiment derived from the embodiment of FIG. 3, shown in a radial crosssection along the line VII3 VI of FIG. 7;

FIG. 7 shows the embodiment of FIG. 6 in an axial cross-section taken along the line VII-VII of FIG. 6;

FIG. 8 represents a still further embodiment of the invention shown in a radial cross-section taken along the line VIIIVIII of FIG. 11;

FIG. 9 shows an enlarged detail of the gear pump of FIG. 8;

FIG. 10 shows the detail of FIG. 9, viewed in the direction of arrow X;

FIG. 11 shows the embodiment of FIG. 8 in an axial cross-section taken along the line XIXI of FIG. 8;

FIGS. Ill and 12 show, respectively, cross-sectional details of FIG. 11; and

FIG. 14 represents a still further embodiment of the invention, shown in an axial cross-section.

DESCRIPTION OF THE PREFERRED EMBODIMENTS The first embodiment of the invention, as illustrated in FIGS. I and 2, shows a particularily simple version of the high-pressure gear pump of the invention. The drive shaft I is rotatably supported in the

cover plates

2 and 3, and the pinion 4 is in driving connection with the shaft 1. A filler wedge 5 is pivotably mounted by means of a dowel pin 6 which is positioned for rotation in oppositely arranged bores in the

cover plates

2 and 3. The internal gear ring 7 is so arranged that a clearance remains between it and the inner diameter of the housing 8. In the separation plane between the central housing 8 and the cover plate 2 is arranged a flexible thrust plate 9. The

various pump elements

2, 3, 8, and 9 are firmly clamped against one another by means of bolts 10.

The inner space of the gear pump is sealed off against the outside by means of O-rings Ill. Another O-ring l2 deliniates the axial compensation pressure field 13 which receives its pressure fluid through the bore 14. According to the invention, the radial compensation pressure field is deliniated by two sealing blades 16 and I7 which, when no pressure is present, are pushed against the periphery of the internal gear member 7 by means of springs 18. The pressure medium flows into the pump via the intake connection I9 and leaves the pump through the pressure connection 20. The diagonal bores 21a and 21b allow the oil from the bearings to return to the intake side.

The axial thrust against the

cover plates

2 and 3 created by the operational pressure results in a deflection of the

cover plates

2 and 3, thus normally creating an increased axial gap. However, in the invention the creation of such an axial gap alongside the gears 4 and 7 is prevented by the provision of the flexible thrust plate 9. The gap instead is created between the cover plate 2 and the thrust plate 9, but it is sealed off by the O-ring 12. A similar situation exists with respect to the radial pressure conditions. The radial thrust causes a deflection of the drive shaft 1 and of the tiller wedge pin 6. Because the radial compensation pressure field 15 is larger than the pumping pressure field, the internal gear member 7 moves in the same direction as the filler wedge 5 and the deflecting shaft 1, thereby preventing that a gap opens between the pinion 4 and the filler wedge 5 and the internal gear ring 7. In spite of these displacements, the sealing blades 16 and I7 maintain their contact against the periphery of the internal gear ring 7, so that no gap exists here either. In terms of design, this first embodiment per FIGS. 1 and 2 is the most simple one, but it has the disadvantage that the pump must operate without clearance on its mating tooth flanks. This shortcoming is avoided in further improved embodiments of the invention as shown in FIGS. 3 through 7.

In the embodiment of FIGS. 3 and 4, the internal gear ring 22 is supported inside a bearing ring 23 which is retained by means of two

dowel pins

24 and 25 arranged in the plane determined by the two axes of the gears. Thus, the bearing ring 23 has its mobility restricted in the direction perpendicular to the above plane in order to maintain a fixed center distance between the gears 4 and 7. Along the end lines of the compensation pressure field, either in the periphery of the bearing ring 23 or in the adjacent surface of the housing bore, are arranged two grooves containing two sealing blocks 26 of either rubber or plastic material, the unmounted blocks being axially longer than the width of the bearing ring 23. Inbetween the two sealing blocks 26 is further arranged a recess 27 which improves the pressure distribution in the compensation pressure field. Between the bearing ring 23 and the central housing 28 is arranged such a radial clearance that the filler wedge 29 can wear during running-in in accordance with the distance e between the dowel pin 24 and the groove by which it retains the bearing ring. This clearance further insures that, under deflection of the pinion shaft 30 and corresponding deflection of the dowel pin 24, the

elements

22 and 29, and 23 can follow this displacement without restriction, meaning that the bearing ring 23 must not abut against the inner wall of the central housing 28. In connection with the above, it is to be understood that the dowel pin must be of such a diameter that its deflection corresponds to the deflection of the pinion shaft.

A modification of the previously described embodiment is shown in FIG. 5, where the dowel pins 31 and 32 are arranged in radial alignment and include a shoulder for the positioning of the bearing ring 33. In this case, it is necessary that the dowel pins are sealed against the outside. This embodiment has the advantage of providing a more stable bearing ring 33.

A further improved embodiment is shown in FIGS. 6 and 7. Here, the central housing 34 includes oppositely arranged guide noses 35 which engage matching grooves 36 in the bearing ring 37, so as to restrict the beearing ring motion to the horizontal direction only. At a 90 angle to these guide noses is arranged a flexible dowel pin 38 whose cross section and restrained length is selected to correspond with the elasticity of the pinion shaft 39.

In this context, it is of particular importance that the dowel pin, or dowel pins, are of such a cross-section and length that their elasticity gives them a deflection under pressure increase which is identical to the deflections of the pinion shaft and of the filler wedge pin.

A still further embodiment is shown in FIGS. 8 through 11. Here, the internal gear member is represented schematically only by a chain line, while the filler wedge and the drive pinion are omitted, in order to show the structural details of the thrust plate 41.

The bore 42 in the housing 43 is not circular, but slightly oval so as to retain the internal gear ring at the points 44 and 45 against displacement in the direction of the center distance along the double arrow 46. This permits the internal gear ring to engage the pinion with a fixed, predetermined center distance and with a predetermined amount of backlash, while allowing the internal gear ring to shift in compensation for the deflection of the pinion shaft and for the wear on the filler wedge in the direction of the double arrow 47. It should be evident that this displacement possibility also insures compensation of any tolerance deviations on the pinion shaft, as well as on the filler wedge and on the internal gear ring. The oval contour of the bore 42 in the housing 43 can be obtained in a simple manner, for example, by first producing a circular bore which is offcenter with respect to the center of the internal gear member by the amount of a with the radius R, and by producing the same bore oppositely off-center by the amount of a with the same radius R. Another simple manner of producing such a bore is by broaching. The broaching operation, in this case, has the further advantage that the guide slots 48 for the sealing blocks 49 can be produced in one and the same operation.

A further inventive contribution consists in providing for the thin thrust plate 41 to be die-cut, whereby all the control contours are obtained in a simple blanking operation. With pressure-compensated gear pumps it is particularly important that the pressure areas be clearly deliniated, so as to avoid the possibility that, under varying operational conditions, the areas which are subjected to the operational pressure should increase or decrease. With the thrust plate 41 of the invention, the two passages 50 prevent the possibility that the pressure between the tooth gaps increases in this area already, as these passages connect the tooth gaps to the intake side of the pump. Another pair of passages 51 assure that the tooth gaps are clearly communicating at this point with the pressure area in the axial pressure field 52 which is deliniated in FIG. 8 by a broken line.

In order to avoid that the pressure between the tooth gaps increases too rapidly when the tooth gaps travel from the passages 50 to the passages 51, a further pair of small nozzle passages 53 is arranged between the above passages, so as to insure that the pressure increases gradually over a distance somewhat less than one tooth pitch.

With gear pumps, there always exists the danger that, in the area of the tooth engagement on the pressure side, a certain residual volume of fluid is traped between the teeth, this volume of fluid being compressed by the motion of the teeth and squeezed through a very narrow passage to either the pressure side or the intake side. Conventional gear pumps, in order to avoid the above situation, provide so-called squeeze fluid discharge grooves or pockets which are milled into the face plate. This milling operation is very expensive, particularly because of the fact that is requires great accuracy. With the thrust plate 41 of the invention, this squeeze fluid discharge groove can be obtained in a most simple manner and with high accuracy (see reference 54 in FIG. 8). The fluid passing through the passage 54 can flow back to the pressure side of the pump, via the bore 55, without requiring an increased consumption of energy for this purpose.

A further inventive contribution is illustrated in FIG. 8 with respect to the lubrication of the bearing of the pinion shaft. Normally, hydraulic gear pumps are lubricated by using the pumping medium for this purpose. A particular problem arises in this regard when the pump has to operate under a wide range of different temperatures. As the gear pump of the invention is intended to be used particularly for mobile machinery, and as such vehicles may be operated in zones of hot as well as cold climate, it may have to be started at a temperature as low as C as well as under operational temperatures of +IOOC. These extreme operating conditions require a system of lubrication for the shaft bearing which guarantees that, at any viscosity of the hydraulic fluid within the indicated temperature range, more oil is supplied to the lubrication groove than is consumed by the bearing. This requirement is met by providing a helical groove 56 in the bearing surface, and by having this groove communicate with the pocket 57 which is blanked out of the thrust plate 411. This pocket 57 coincides with the area in which the teeth of the two gears leave their corresponding tooth gaps, thereby creating a vacuum. When the viscosity of the fluid is very high, such as under start-up of the hydraulic system at low temperatures, the suction effect across the lubrication groove is not very great, as the pump operates under practically absolute vacuum anyway. On the other hand, however, the entrainment effect in the helical groove 56 at a high viscosity is a very marked one, so that the supply of lubricating oil to the bearing under these operational conditions is assured. Conversely, at high temperatures and low viscosity, the entrainment effect inside the helical groove 56 is not sufficient. However, this is then compensated by the suction effect on the lubricating oil from the vacuum present in the pocket 57.

FIG. 8 also shows a particularly advantageous arrangement of the sealing elements which are represented by the blades 16 and 17 in FIG. I. Here, these blades are replaced by sealing blocks 49. In FIGS. 9 and 10, these sealing blocks 49 are illustrated at an enlarged scale, where it can be seen that each pair of sealing blocks is supplied with fluid from the axial compensation pressure field through a die-cut bore 58 in the face plate 41. Thus, the pressure fluid passes through the passages 51 and 55, and via the bore 58, into the space delineated by the sealing blocks 49 and by the adjacent faces of the outer periphery of the internal gear member and of the housing bore 42 with the radius R. The sealing blocks 49 include a connecting groove 59 through which the pressure fluid passes into the pressure-balancing recess 60 and, via the bores 61, into the space 62 provided between the sealing block 49 and the housing 43. This arrangement assures that the sealing blocks 49 are always pressed against the internal gear member. In the idling condition, however, when the operational pressure is zero, this hydrostatic force resulting from the overcompensation is not effective. For this purpose, additional leaf springs 63 are provided in the space between the sealing blocks 49 and the housing 43 so that, even during idling, the sealing blocks are pressed against the internal gear ring, while the latter is pressed against the filler wedge, and the filler wedge in turn is pressed against the crown circle of the pinion. Due to the fact that the space between the sealing blocks 49 is under pressure, each of the sealing blocks is pressed against the lateral retaining face 64, where it forms a clearance-free metallic seal. For this reason, this design does not require any rubber sealing elements. One advantage of this is that the temperature sensitivity of this kind of seal is of no concern.

As can be seen from FIG. 8, each pair of sealing blocks defines a compensation pressure field of the length L. The space 65 between the two inner sealing blocks is connected to the intake side of the pump via a central groove 66, a peripheral channel 67, and via the end grooves 68. Normally, it would suffice to use only the two exterior sealing blocks, omitting the two interior sealing blocks. In this case, the layout of the compensation pressure field would correspond to that shown in FIG. 1. The latter layout, however, has the disadvantage that the internal gear member is subjected to a very high alternating bending stress. This situation is due to the requirement that the outer compensation pressure field, i.e. the area enclosed between the sealing blades should be only slightly larger than the pressure area on the inside of the internal gear member. However, because the outer diameter of the internal gear member is considerably larger than its internal diameter, it becomes necessary to have the angle enclosed between the sealing blades correspondingly smaller than the angle within which the pressure acts on the inner diameter of internal gear member. This condition results in a bending moment which tends to flatten the affected ring portion of the internal gear member in the pressure area. By properly selecting the pressure-free distance between the two inner sealing blocks, the above mentioned bending moment can be reduced to such an extent that it becomes negligible.

The embodiment of FIG. 11 also shows the arrangement of positioning pins 83 which were not shown in FIG. 8. Form this figure, it can also be seen that the thrust plate 41, in contrast to the earlier described FIGS. 2, 4 and 7, is not clamped between the

housing parts

84 and 43, but has a minute axial clearance inside the recess 69 of the housing part 84. In this recess, the thrust plate 41 is fixed against rotation but free to make a small axial displacement. This arrangement has the advantage that, when temperature differentials are present between either the pinion shaft, the filler wedge, the internal gear assembly, or the housing, such as may occur during start-up at low temperatures, the warmest part of the face plate 41, which is then the thickest, can yield in the direction of the shaft axis, so as to avoid accidental seizure between the thrust plate 41 and the moving parts.

In FIGS. 11, 12, and 13 is indicated a further inventive contribution which concerns itself with the accurate positioning of the two

cover plates

84 and 70 relative to the housing 43. With gear pumps where the shaft 71 is supported in two different housing parts, there is always the problem that the two

cover plates

84 and 70 must be aligned very accurately relative to one another. Normally, this positioning is obtained by means of dowel pins or by means of tapered pins which require the corresponding bores in the housing parts to be produced at very close tolerances as regards their diameter and their alignment. A well-known procedure to obtain this accurate alignment consists of using a clamping fixture to simultaneously bore all the housing parts, the separate parts being later cleaned of shavings and the pins being mounted during final assembly. However, because this common boring operation of the parts must be performed as a part of assembly operations, there exists the danger that, during the subsequent final assembly, some of the metal shavings may get back into the pump. This type of simultaneous boring operation has the further shortcoming that, when one housing part is to be replaced, the bores are no longer in alignment so that the other two housing parts must either be bored again or discarded.

These shortcomings are obviated in the gear pump as suggested in FIGS. 11, 12, and 13. Here, the housing or center part 43, is provided on both axial faces with a pair of positioning bores 72 which are slightly tapered and therefore need not be positioned very accurately. The two

cover plates

84 and 70, likewise, have each a pair of cylindrical bores which are only roughly predrilled. During assembly, the

parts

43, 84 and 70 are clamped together in a fixture, whereby the latter are aligned with reference to the pinion shaft bore and the bore for the filler wedge pin 74 (see also FIG. 8). Subsequently, a pair of cylindrical pins 83, as shown in FIG. 12, is introduced into the bores 73 and, by means of a press with opposing punches, forcibly deformed as shown in FIG. 13. To facilitate this operation, the pins 83 are of a material which can be readily deformed by cold-flowing. Now, when one of the three

parts

84, 43, or 70 need replacement, only the deformed pin 83 needs to be forced 'cutand discarded. The other, more expensive housing parts can bereused.

In pumps of th plate type design, as is the case in FIGS. 1 through 3, the radial forces inside the pump must be transmitted through frictional contact between the

housing parts

84, 43, and 70. This means that the bolt 75 (see FIG. 8) must be preloaded to such an extent that the friction obtained between the housing parts as a result of this preloading prevents any relative shifting of the housing parts, even under the highest pressures. For gear pumps which have to operate under high and extremely fluctuating pressures, a great number of bolts 75 will be necessary. In FIG. 14 is shown a design which circumvents this problem. This figure shows a pump structure where the central housing 78 has an outer shoulder on each side which engages a corresponding recess in the

adjacent cover plates

77 and 78. In this case, the bolt 79 only needs to resist the axial load from the compensation pressure field 80. However, this version requires that thrust plates 81 and 82 are provided on either side of the housing, because the machining of a recessed face in the cover plates with the surface quality necessary for this application requires machining operations which are too costly.

I claim:

l. A high-pressure gear pump comprising in combination:

a closed housing assembly including a housing body with a housing bore, an intake connection for the pumping medium, a discharge connection for same, and at least one removable lateral cover plate;

an internal gear assembly arranged inside the housing bore, the internal gear assembly having a diameter which is smaller than the housing bore so as to define a gap therewith which permits limited displacements of the internal gear assembly relative to the housing assembly, the internal gear assembly comprising a rotatable internal gear ring;

a drive pinion and drive shaft arranged inside the internal gear ring and mating therewith to create the pumping action, the drive shaft being journalled in the housing assembly;

a radially adjustable arcuate filler wedge arranged between the two gears on their pumping side, the narrower end of the wedge delimiting the field within which the gear teeth are exposed to the pumping pressure;

and

at least one radial compensation pressure field arranged in the gap between the adjacent faces of the housing bore and the periphery of the internal gear assembly and communicating with the pumping pressure field so as to overcome the radially outward directed thrust on the internal gear assembly resulting from the pumping pressure field, in order to establish forcible radial contact between the teeth of the internal gear ring and the outer arc of the filler wedge and between the inner arc of the filler wedge and the teeth of the drive pinion,

the peripheral length of the radial compensation pressure field, or fields, being defined by transverse sealing members which sealin gly interrupt the gap over its entire axial width, the transverse sealing members yielding radially independently of each other in response to changes in the gap opening, when the internal gear assembly changes its position inside the housing bore.

2. A gear pump as defined in claim 1, wherein the housing bore includes transverse guide grooves wherein the transverse sealing members are received;

the sealing members are in the form of wiper blades which are radially movable in the guide grooves so that they can follow the displacement of the internal gear assembly; and

the transverse sealing members include means for urging them against the periphery of the internal gear assembly.

3. A gear pump as defined in claim 1 comprising two separate radial compensation pressure fields, the peripheral length of each one being defined by a pair of transverse sealing members, the peripheral gap between the separate radial compensation pressure fields communicating with the intake side of the pump.

4. A gear pump as defined in claim 1, wherein the housing assembly further includes a plurality of centering pins of cold-flowable material engaging adjacent parts of the housing assembly, the latter being thereby permanently centered relative to one another after the pins have been deformed during assembly.

5. A gear pump as defined in claim 1, wherein the internal gear assembly further includes a bearing ring surrounding and guiding the internal gear ring, the outer periphery of the bearing ring thus being the periphery of the internal gear assembly and forming one of the two adjacent faces which define the gap for the radial compensation pressure field;

at least one of these adjacent faces is provided with transverse grooves for the accommodation of the transverse sealing members;

the transverse sealing members are of a resiliently compressible material; and

the bearing ring includes means for restricting it against rotation while permitting limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.

6. A gear pump as defined in claim 5, wherein the bearing ring restricting means are arranged to prevent the relative displacement of the internal gear assembly in the direction of the center distance between the two gears so as to maintain a fixed distance between their axes.

7. A gear pump as defined in claim 6, wherein the bearing ring restricting means includes in the bearing ring periphery two opposite transverse recesses in alignment with the plane defined by the two gear axes, and two restricting pins secured in the housing assembly and so arranged with respect to the transverse recesses that, while no radial clearance exists between the pins and the bottoms of the recesses so as to maintain a fixed center distance between the gear axes, lateral clearances between the pins and the sides of the recesses permit other limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.

8. A gear pump as defined in claim 6, wherein the bearing ring restricting means includes in the bearing ring periphery two opposite recesses in alignment with the plane defined by the two gear axes, and two radially aligned restricting pins secured in the housing body, the restricting pins including means for radially abutting against the housing body so adjusted that, while no radial clearance exists between the pin ends and the bottoms of the recesses so as to maintain a fixed center distance between the gear axes, lateral clearances between the pins and the sides of the recesses permit other limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.

9. A gear pump as defined in claim 6, wherein the bearing ring restricting means includes in the bearing ring periphery and housing bore, near their most distant points from the plane defined by the two gear axes, cooperating guide noses and guide grooves whose flanks are perpendicular to the gear axes plane, so as to restrict the displacements of the internal gear assembly to displacements in the direction perpendicular to the gear axes plane.

10. A gear pump as defined in claim 6, wherein the bearing ring restricting means includes in the plane defined by the two gear axes, on the side of tooth engagement, a restricting pivot connection betweenthe bearing ring and the housing assembly, the bearing ring restricting means further including a means for limiting the pivoting displacements of the internal gear assembly around the pivot connection so that, in the new condition of the gear pump, a pivoting displacement commensurate with the running-in wear of the tiller wedge against the teeth of both gears is permitted, so that the radial contact pressure between the filler wedge and the teeth of both gears is substantially reduced after running-in, when the limiting means becomes effective.

11. A gear pump as defined in claim 1, wherein the adjacent peripheral faces of the internal gear assembly and of the housing bore are not congruent in outline so that the gap defined therebetween is substantially zero in the area of the two points which coincide with the plane defined by the two gear axes so as to maintain a fixed center distance between the gears, while being large enough along the remaining portions of the outline to permit limited displacements of the internal gear assembly relative to the housing assembly.

12. A gear pump as defined in claim 11, wherein the outer periphery of the internal gear ring is cylindrical and forms one of the two adjacent faces defining the gap; and

the housing bore is slightly oval in outline, the smallest diameter of the bore being equal to the diameter of the internal gear ring and aligned with the plane defined by the two gear axes.

13. A gear pump as defined in claim 1, wherein the outer periphery of the internal gear ring is cylindrical and forms one of the two adjacent faces de fining the gap for the radial compensation pressure field.

14. A gear pump as defined in claim 13, wherein the housing bore includes guide grooves for the transverse sealing members, the radial depth of the grooves being less than their arcuate width; and

the transverse sealing members are in the form of flat sliding blocks and are radially movable inside the guide grooves so as to follow the displacements of the axis of the internal gear ring.

15. A gear pump as defined in claim 14, wherein each sliding block has a sliding face for contacting the periphery of the rotating internal gear ring and a back face substantially parallel thereto;

a major portion of the area of the sliding face is recessed and the space thereby created as well as the space between the back face and the bottom of the guide groove are in communication with the adjacent radial compensation pressure field, so as to minimize the radial hydrostatic load on the sliding block in the direction toward the internal gear ring; and

each sliding block includes a means for maintaining the contact between its sliding face and the internal gear ring in the absence of hydraulic pressure in the radial compensation pressure field during startup on the pump.

16. A gear pump as defined in claim 1, wherein the housing assembly further includes:

on at least one side of the gears, and laterally covering at least the lateral area of the pumping pressure field and radial compensation pressure field, a flexibly yielding thrust plate arranged between the gears and the housing assembly; and

an axial thrust compensation pressure field between the thrust plate and the housing assembly so arranged as to overcome the axially outward directed thrust on the thrust plate resulting from the pumping pressure, in order to establish forcible lateral contact between the thrust plate on the one side, and the drive pinion, filler wedge and internal gear assembly on the other side.

17. A gear pump as defined in claim 16, wherein the housing assembly further includes an axial recess for the accommodation of the thrust plate, the latter being slightly thinner than the depth of the axial recess.

18. A gear pump as defined in claim 16, wherein the thrust plate includes a suction pocket in the area in which the gear teeth start separating so as to create a suction effect; and

the housing assembly further includes a bearing bore inside which the pinion drive shaft is joumalled, and a helical lubricating groove in the bearing bore, one end of the groove communicating with the suction pocket, the other end of the groove communicating with the intake side of the pumps, the direction of the groove helix being such that the rotation of the drive shaft against the groove causes the pumping medium to be frictionally entrained along the groove toward the suction pocket.

19. A gear pump as defined in claim 16, wherein the housing assembly further includes two removable cover plates, one on each side of the housing body;

two thrust plates on opposite sides of the gears, the outside diameter of the thrust plates being larger than the outside diameter of the internal gear assembly;

each cover plate has on its inner side an axial recess of a diameter corresponding to the diameter of the thrust plate and of a depth in excess of the thickness of the thrust plate;

the housing body has on each side a protruding shoulder which fits into the axial recess of the adjacent cover plate, thereby centering the parts of the housing assembly relative to each other, each shoulder, when positioned in the recess, leaving a slight axial play for the thrust plate; and

an axial compensation pressure field between each cover plate and the adjacent thrust plate, the pressure field being delineated by a groove in the cover plate with an O-ring received therein in contact with the thrust plate.

20. A gear pump as defined in claim 16, wherein the thrust plate includes for each row of gear teeth and in alignment with their paths:

an axial pressure relief passage communicating with the intake side of the pump and located in the area ahead of the opposing pumping and axial compensation pressure fields; and

an axial entry passage linking the pumping pressure field with the axial compensation pressure field.

21. A gear pump as defined in claim 20, wherein the thrust plate further includes:

between each pressure relief passage and pressure entry passage an axial pressure build-up nozzle linking the axial compensation pressure field with the gear teeth at a point shortly before the latter are exposed to the pumping pressure field.

Claims

1. A high-pressure gear pump comprising in combination: a closed housing assembly including a housing body with a housing bore, an intake connection for the pumping medium, a discharge connection for same, and at least one removable lateral cover plate; an internal gear assembly arranged inside the housing bore, the internal gear assembly having a diameter which is smaller than the housing bore so as to define a gap therewith which permits limited displacements of the internal gear assembly relative to the housing assembly, the internal gear assembly comprising a rotatable internal gear ring; a drive pinion and drive shaft arranged inside the internal gear ring and mating therewith to create the pumping action, the drive shaft being journalled in the housing assembly; a radially adjustable arcuate filler wedge arranged between the two gears on their pumping side, the narrower end of the wedge delimiting the field within which the gear teeth are exposed to the pumping pressure; and at least one radial compensation pressure field arranged in the gap between the adjacent faces of the housing bore and the periphery of the internal gear assembly and communicating with the pumping pressure field so as to overcome the radially outward directed thrust on the internal gear assembly resulting from the pumping pressure field, in order to establish forcible radial contact between the teeth of the internal gear ring and the outer arc of the filler wedge and between the inner arc of the filler wedge and the teeth of the drive pinion, the peripheral length of the radial compensation pressure field, or fields, being defined by transverse sealing members which sealingly interrupt the gap over its entire axial width, the transverse sealing members yielding radially independently of each other in response to changes in the gap opening, when the internal gear assembly changes its position inside the housing bore.

2. A gear pump as defined in claim 1, wherein the housing bore includes transverse guide grooves wherein the transverse sealing members are received; the sealing members are in the form of wiper blades which are radially movable in the guide grooves so that they can follow the displacement of the internal gear assembly; and the transverse sealing members include means for urging them against the periphery of the internal gear assembly.

5. A gear pump as defined in claim 1, wherein the internal gear assembly further includes a bearing ring surrounding and guiding the internal gear ring, the outer periphery of the bearing ring thus being the periphery of the internal gear assembly and forming one of the two adjacent faces which define the gap for the radial compensation pressure field; at least one of these adjacent faces is provided with transverse grooves for the accommodation of the transVerse sealing members; the transverse sealing members are of a resiliently compressible material; and the bearing ring includes means for restricting it against rotation while permitting limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.

10. A gear pump as defined in claim 6, wherein the bearing ring restricting means includes in the plane defined by the two gear axes, on the side of tooth engagement, a restricting pivot connection between the bearing ring and the housing assembly, the bearing ring restricting means further including a means for limiting the pivoting displacements of the internal gear assembly around the pivot connection so that, in the new condition of the gear pump, a pivoting displacement commensurate with the running-in wear of the filler wedge against the teeth of both gears is permitted, so that the radial contact pressure between the filler wedge and the teeth of both gears is substantially reduced after running-in, when the limiting means becomes effective.

12. A gear pump as defined in claim 11, wherein the outer periphery of the internal gear ring is cylindrical and forms one of the two adjacent faces defining the gap; and the housing bore is slightly oval in outline, the smallest diameter of the bore being equal to the diameter of the internal gEar ring and aligned with the plane defined by the two gear axes.

13. A gear pump as defined in claim 1, wherein the outer periphery of the internal gear ring is cylindrical and forms one of the two adjacent faces defining the gap for the radial compensation pressure field.

14. A gear pump as defined in claim 13, wherein the housing bore includes guide grooves for the transverse sealing members, the radial depth of the grooves being less than their arcuate width; and the transverse sealing members are in the form of flat sliding blocks and are radially movable inside the guide grooves so as to follow the displacements of the axis of the internal gear ring.

15. A gear pump as defined in claim 14, wherein each sliding block has a sliding face for contacting the periphery of the rotating internal gear ring and a back face substantially parallel thereto; a major portion of the area of the sliding face is recessed and the space thereby created as well as the space between the back face and the bottom of the guide groove are in communication with the adjacent radial compensation pressure field, so as to minimize the radial hydrostatic load on the sliding block in the direction toward the internal gear ring; and each sliding block includes a means for maintaining the contact between its sliding face and the internal gear ring in the absence of hydraulic pressure in the radial compensation pressure field during startup on the pump.

16. A gear pump as defined in claim 1, wherein the housing assembly further includes: on at least one side of the gears, and laterally covering at least the lateral area of the pumping pressure field and radial compensation pressure field, a flexibly yielding thrust plate arranged between the gears and the housing assembly; and an axial thrust compensation pressure field between the thrust plate and the housing assembly so arranged as to overcome the axially outward directed thrust on the thrust plate resulting from the pumping pressure, in order to establish forcible lateral contact between the thrust plate on the one side, and the drive pinion, filler wedge and internal gear assembly on the other side.

18. A gear pump as defined in claim 16, wherein the thrust plate includes a suction pocket in the area in which the gear teeth start separating so as to create a suction effect; and the housing assembly further includes a bearing bore inside which the pinion drive shaft is journalled, and a helical lubricating groove in the bearing bore, one end of the groove communicating with the suction pocket, the other end of the groove communicating with the intake side of the pumps, the direction of the groove helix being such that the rotation of the drive shaft against the groove causes the pumping medium to be frictionally entrained along the groove toward the suction pocket.

19. A gear pump as defined in claim 16, wherein the housing assembly further includes two removable cover plates, one on each side of the housing body; two thrust plates on opposite sides of the gears, the outside diameter of the thrust plates being larger than the outside diameter of the internal gear assembly; each cover plate has on its inner side an axial recess of a diameter corresponding to the diameter of the thrust plate and of a depth in excess of the thickness of the thrust plate; the housing body has on each side a protruding shoulder which fits into the axial recess of the adjacent cover plate, thereby centering the parts of the housing assembly relative to each other, each shoulder, when positioned in the recess, leaving a slight axial play for the thrust plate; and an axial compensation pressure field between each cover plate and the adjacent thrust plate, the pressuRe field being delineated by a groove in the cover plate with an O-ring received therein in contact with the thrust plate.

20. A gear pump as defined in claim 16, wherein the thrust plate includes for each row of gear teeth and in alignment with their paths: an axial pressure relief passage communicating with the intake side of the pump and located in the area ahead of the opposing pumping and axial compensation pressure fields; and an axial entry passage linking the pumping pressure field with the axial compensation pressure field.

21. A gear pump as defined in claim 20, wherein the thrust plate further includes: between each pressure relief passage and pressure entry passage an axial pressure build-up nozzle linking the axial compensation pressure field with the gear teeth at a point shortly before the latter are exposed to the pumping pressure field.