US3470825A - Hydrostatic radial piston pump - Google Patents

Description

Oct. 7, 1969 w. GSCHING HYDROSTATIC RADIAL PISTON PUMP Filed July 31, 1967 Fig. I

2 Sheets-Sheet 1 Oct. 7, 1969 w. GSCHING HYDROSTATIC RADIAL PISTON PUMP 2 Sheets-Sheet 2 Filed July 31, 1967 Fig. 5

United States Patent 3,470,825 HYDROSTATIC RADIAL PISTON PUMP Wilhelm Gsching, Heidenheim (Brenz), Germany, assignor to Voith Getriebe KG, Heidenheim (Brenz), Germany Filed July 31, 1967, Ser. No. 657,263 Claims priority, application Germany, Aug. 6, 1966, V 31,663; Feb. 2, 1967, V 32,891 Int. Cl. F04b 1/10 US. Cl. 103-161 2 Claims ABSTRACT OF THE DISCLOSURE Rotary fluid pressure device having radial pistons in radial cylinders in a rotor rotatable within a housing in which the rotor is divided into two axial sections with a chamber therebetween supplied with pressure from the pressure side of the device so the rotor sections are pressed in sealing engagement with the inner walls of the housing.

The present invention relates to a hydrostatic radial piston pump for increasing the pressure of a flowing medium, in which the radially arranged cylinders are provided in a rotor adapted to rotate together with the drive shaft, and in which the passages for the admission of a pressureless medium communicate with said cylinders from a suction conduit to the cylinders or for the withdrawal of the pressure medium from the cylinders to the pressure conduit lead into one of two, or into both lateral surfaces of the rotor and from there continue in the housing. Between the rotor and the housing there is provided an axially displaceable pressure member which has one of its lateral surfaces subjected to the influence of a pressure medium in sealing engagement with the corresponding lateral surface of the rotor. With such a pump, the pressure member journalled in the housing is likewise provided with passages which on one hand communicate with the suction or pressure conduit of the pump, and on the other hand communicate with the flow passages in the rotor. Depending on the type of operation of the pump, the suction and pressure passages are exchangeable. It is necessary that the rotor at the transition areas of the rotating passages to the passages is well sealed which are arranged in the housing, or in those portions which are non-rotatably connected to said housing, in order to prevent any leakage at said transition areas. This is realized by pressing the pressure member within the range of the mouth of the housing passages in axial direction against the rotor which forms a single piece with the shaft. This pressure member is pressed against said rotor by a pressure medium in such a way that at these sliding areas between the rotating and stationary parts a reliable seal will be effected. The pressure member has those side surfaces thereof which are remote from the sliding surface provided with two concentric annular passage surfaces, whereas the housing has two, concentric annular cylinders on corresponding diameters. One of the thus formed cylinder chambers communicates with the suction conduit, whereas the other one of said cylinder chambers communicates with the pressure conduit.

The operational possibilities of a radial piston pump of the described type are limited. In order to make possible an economic production of such radial piston pumps in large numbers, it is necessary so to design a pump that with a minimum of expense concerning different individual elements it can be safely used for all practical occurring conditions of operation without the necessity of structurally changing the pump itself. Similarly, it is required that the pump can in a very simple manner be converted to a motor. These requirements can, with the ICC abovementioned pump, be met only by additional parts and even then only to a limited extent. Moreover, a particular disadvantage consists in that the axial extension of the pressure members to be arranged in the immediate vicinity of the rotor requires an axial width. of the rotor which considerably exceeds the spacing between the bearings. This makes an oscillation-free operation of the pump at high speeds rather difficult and forces a considerable increase in the thickness of the driving shaft. This in turn requires a corresponding increase in the radial outer dimensions of the pump housing and thus brings about the employment of additional material, requires additional manufacturing time, and also increases the costs.

It is, therefore, an object of the present invention to provide a hydrostatic radial piston pump which will reduce and possibly completely overcome the abovementioned drawbacks.

These and other objects and advantages of the present invention will appear more clearly from the following specification in connection with the accompanying drawings, in which:

FIGURE 1 illustrates a longitudinal section of a double pump according to the present invention with two axially adjacent rows of cylinders, said section being taken in the direction of the eccentricity between the rotor and the housing.

FIGURE 2 is a vertical section through a modified double pump according to the invention, said section being taken perpendicularly with regard to the eccentricity between the runner and the housing.

FIGURES 3 and 4 respectively illustrate two further modifications of the gap chamber or chambers between the two rotor portions.

FIGURE 5 illustrates a section through a modified pump according to the invention in which the resultant of the pressing force employed for sealing purposes is in alignment with the resultant of the pump pressure of the sealing surface.

The pump according to the present invention is characterized primarily in that the rotor comprises two axially arranged adjacent parts which are non-rotatable with regard to the driving shaft but are axially displaceable relative to each other. The pump according to the invention is furthermore characterized in that between the two rotor parts there is provided at least one gap chamber which communicates with the respective pressure conduit, the arrangement being such that the pressure medium intorduced into the gap chamber presses the rotor portion or the rotor portions against the respective sealing surfaces of the housing.

Such an arrangement of the pump will, while yielding a proper seal at the transition area of the'suction and pressure conduits from the stationary housing part to the rotating rotor and vice versa, reduce the manufactuing costs for the pump, especially an expensive machining of the pump housing. Furthermore, the dimensions of the pump, especially in axial direction, are reduced, and the journalling of the driving shaft is improved. The gap chamber filled with pressure medium for pressing the rotor parts against the axial lateral surfaces of the "stationary housing parts (or, if desired, a plurality of such gap chamber) is provided between the rotor parts Therefore, the design of the housing is completely independent of the design of said gap chamber'or "gap chambers. Even with a pump provided with one single cylinder row only, in which these cylinders are arranged in one rotor part, wherea the other rotor part is designed merely as a disc and may, if desired, be even rigidly connected to a shaft or formed of one piece therewith, the axial space requirement for said last-mentioned additional rotor part is irrelevant with regard to the space requirement for the pressure member. With a so-called double pump povided in each runner part with a series of cylinders, the axial space requirement for the space gap is no longer material because the said space chamber is formed by turned-out portions or by one or more pairs of turned-out portions and protrusions of the two rotor parts.

Advantageously, according to a further feature of the present invention, each space chamber is formed primarily in the form of an annular recess provided on at least one of those lateral surfaces of the runner parts which face each other. On the outer circumference of said annular recess and on the inner circumference thereof there are respectively provided piston seals of any standard type, for instance, a highly elastic sealing ring.

According to a further development of the present invention, the gap chamber communicates through a preferably narrow base with an annular groove Connected to the pressure conduit, said annular groove being arranged on the axially outer lateral surface of at least one of the runner parts and being coaxial to the axis of rotation. In view of this structure, also with a plurality of cylinders in one or each runner part, it will be assured that the gap chamber is always in communication with that cylinder which produces the highest pressure of the medium. In order to be sure that if the supply of pressure medium should fail, or not exist, for instance when starting the pump against a pressureless consumer, the axial outer lateral surfaces of the runner parts be sufficiently pressed against the corresponding lateral surfaces of the housing parts, pressure springs are, in conformity with a further development of the invention arranged between the two runner parts.

According to a further development of the invention, between the axially displaceable runner part of each axially displaceable runner part on one hand, and the respective lateral surface of the housing on the other hand, there is provided a fitting with passages therethrough. That lateral surface of said fitting which is adjacent said rotor is plane while the lateral surface adjacent the housing is spherical. In view of this arrangement of the fitting members, without any material increase in the axial length of the pump, there will on one hand be assured the seal of the sealing surfaces even under unfavorable installing conditions, and on the other hand there will be created an easily exchangeable wear element which comprises that sliding surface which is subjected most to wear.

According to a further feature of the present invention,'the two runner parts which are axially displaceable on a'splined shaft profileof the driving shaft are in the axial range of said splined shaft profile near their axially outer lateral surface provided with a cylindrical recess each. Each of said recesses has associated therewith a cylindrical groove arranged at the respective areas of the drive shaft, and in each of said recesses and grooves there-is inserted a highly elastic sealing ring and axially outwardly there is provided a transversely divided supporting ring carrying said sealing ring. Such a construction yields a highly reliable radially inner seal of the gap chamber whch can easily be serviced. Advantageously, the said recess has a larger axial width than the groove because such a design makes possible an effortless assembly and disassembly of the sealing rings and of the supporting rings.

A double pump designed in the above-outlined manner accordingto the present invention works highly satisfactorily when the pressure passages and the suction passages of the two cylinder rows are brought together in the housing and when the common pressure line is connected to a single place of consumption. In many instances where such double pumps are employed, it is also highly desirable to connect each cylinder row to a place of consumption of its own.

Whenoperating a double pump connected in this way to two different consumers, it is unavoidable that the two 4 cylinder rows will occasionally operate at considerably different pressures. Since both cylinder rows communicate with the gap chamber common thereto, it will be appreciated that at non-uniform load of said cylinder rows, undesired compensating fiows will occur from the conduit connected to the cylinder row with the higher pressure, through annular grooves, the annular bores and the gap chamber to the pressure conduit which is connected to the cylinder row with the lower pressure. These compensating flows cause losses which are all the higher the greater the pressure drop between the two pressure conduits.

It may also be mentioned that in such an instance the pressure in the gap chamber adjusts itself to a medium value between the pressures in the two pressure conduits. In this connection, it may occur that when greatly relieving one cylinder row, the medium pressure forming in the gap chamber will no longer be sufficient for properly sealing the sliding surfaces on the pump side which has the higher pressure. This drawback can, in conformity with a further development of the invention be eliminated by providing each of the bores between an annular groove and the gap chamber with a check valve opening toward said gap chamber, preferably in an axis parallel blind bore. In this way, the pressure in the gap chamber will always be maintained at the respective highest pressure prevailing in one of the two pressure conduits, but the flow-through of pressure medium toward the pressure conduit will be interrupted by the lower pressure.

When, during operation of a double pump designed in the above-mentioned manner the pressure drops in both pressure conduits, the pressure prevailing prior to the pressure drop will be maintained in the gap chamber. In other words, the check valves will prevent the automatic pressure compensation between the gap chamber and the respective pressure conduit having the higher pressure. This is harmless, as a rule. If, however, the working pressures vary considerably, it may occur that the higher energy consumption caused by the higher pressure in the gap chamber will, in view of the friction in the Sliding surfaces between the runner parts and the housing parts, be too high with regard to the useful output. This drawback can be overcome, according to the present invention, by providing a compensating bore in the or each runner part or provided with a check valve in the bore leading to the gap chamber. Said compensating bore didectly interconnects one of the cylinder chambers of this runner part wth the gap chamber, preferably through one of the lines bores. This compensating bore merely has to have a rather narrow cross-section because only minute quantities of pressure medium have to pass through for compensating the pressure. In view of this compensating bore, it will be appreciated that when the working pressure drops in the cylinder rows of the two runner parts, a direct pressure compensation will be obtained between the gap chamber and the pressure conduit with the higher working pressure.

The design of the double pump is furthermore advantageously so selected that each of the rotor sections by means of a common splined shaft profile of the drive shaft, which profile extends over the axial extension of the two rotor sections, is non-rotatably connected to said drive shaft but is axially displaceable thereon.

In order to prevent that such an arrangement require the employment of combined axial radial bearings for journalling the drive shaft, the splined shaft profile of the drive shaft is, in conformity with a further development of the invention interrupted within the range of the gap between the two rotor sections by a turned-out portion for inserting a transversely divided fitting ring which preferably has the same outer diameter as said splined shaft profile. The axial end faces of this fitting ring are engaged by the axial end faces directed toward the gap chamber and pertaining to protruding portions of the splined shaft profile on said rotor sections. In this way,

an axial movement back and forth of the rotor sections will be prevented.

According to a further development of the present invention, the gap chamber on the radially inner side is sealed by a highly elastic sealing ring which rests directly on the outer surface of the transversely divided fitting ring. Since the sealing ring is under outer overpressure, it simultaneously prevents a radial escape of the two sections of the transversely divided fitting ring under the influence of the centrifugal forces occurring during the rotation of the rotor. I

In order to be able to employ a double pump according to the invention in any desired manner, which means not only with a desired direction of rotation and to the same extent as a pump or as a motor, but also with any desired direction of flow, with an eccentricity adjustable to either side and with separate connection of the two cylinder rows to dilferent consumers, a design with two concentrically arranged gap chambers has proved advantageous of which one gap chamber is connected to the two flow passages leading to the pressure side of the pump whereas the other gap chamber is connected to the two flow passages connected to the suction side of the pump. Such a design makes it possible to exchange pressure passages and suction passages. To this end, according to a further feature of the invention, the two rotor sections have those axial lateral surfaces which face each other provided with a turned-out portion and/or a protrusion which in radial direction are confined by coaxial cylinder surfaces. The turned-out portion of one rotor section has associated therewith a protrusion of the other runner section located at the same medium distance from the axis of rotation. The radial extension of said protrusion is less than the associated turned-out portion so that between each two confining cylindrical surfaces of the turned-out portion which face each other and the associated protrusion, a highly elastic sealing ring can be arranged. For purposes of reducing the radial space requirement of these gap chambers and in order firmly to hold the transversely divided fitting ring in the turned-out portion of the splined shaft profile, the radially inner sealing ring snugly engages the fitting ring. The outside of this sealing ring is surrounded by the radially inner cylindrical confining surface of the protrusion of one rotor section.

With all above-mentioned features according to the invention, the occurring resultant pressing force acts along the axis whereas the resultant force of the pump pressure acts outside said axis and, more specifically, in conformity with the center of gravity of the flow-through openings for the respective conduits (pressure and suction passages) in the stationary control level or control deflection. In this Way, a tilting moment is exerted upon the rotor and the sealing surfaces perpendicular to the axis of rotation to bring about an edging of the rotor. For this reason, it would be possible to provide additional pressure surfaces for countering the tilting moment. This, however, would involve additional costs and space. Partioularly, additional space requirement is a rather important drawback with compact pumps aside from the fact that also the friction losses increase.

According to a further suggestion of the present invention, also this drawback can be avoided. Starting from the finding that for frictional reasons it is less important to press the respective rotor section in its entirety against the housing, and that it is more important to press that portion of the respective rotor section against the housing in which the conduits are under the pump pressure, the resultant of the pump pressure should be at least approximately in alignment at the sealing surface so that little or no tilting moment can occur. To this end, it is suggested that the pressure chamber is divided into a number, cora conduit while being provided with one sealing piston each which rests against the other rotor section. This feature brings about that merely at the required portions of the control level, a corresponding sealing pressure occurs. It will be appreciated that the pressure in the respective cylinder passage is transferred into the respective pressure chamber pertaining thereto and merely from there brings about a pressing of the corresponding sealing surfaces. Within the range of the suction passages, in order to prevent the portions on the control level from being lifted off, only a slight pressing pressure is required at the sealing surfaces of the control level which, for instance, can be realized by a number of springs arranged between the rotor sections. In this way, the friction losses on the sealing surfaces are kept to a minimum.

According to a further development of the present invention, the same number of cylinders is provided for the two rotor sections while two partial pressure chambers each which are located adjacent to each other in axial direction are aligned with each other and are provided with a sealing piston common thereto. Such an arrangement greatly simplifies the construction and manufacture of the pump. The size of the surfaces of the pressure chambers acted upon by the pressure medium will determine the pressing force exerted upon the sealing surfaces. This is of great importance for the economy of the pump. It is suggested to make the total of the surfaces of the sealing pistons directly acted upon by the working pressure (when seen in axial direction) approximately from 1.4 to 1.8 times as great as the total of the surfaces of the pressure passages in the control level or control deflection of the housing. Such surface ratio takes into consideration the fact that the working pressure in the control level is present and effective also outside the flow passages, i.e. in the gap, and, more specifically, within the limits between the maximum value at the outer surface of the flow passages and zero in a range of the gap remote from said outer surfaces of said flow passages. In this connection, in addition to the force produced by the direct action, there is obtained a force which tends to lift the respective rotor section of the control level.

Referring now to the drawings in detail, the double pump shown in FIG. 1 comprises a drive shaft 1 having a splined profile section 2 engaged by two axially adjacently arranged rotor section 3, 4 forming the pump rotor for rotation with said shaft. The said pump rotor is somewhat axially displaceable on said splined section 2. The rotor sections 3, 4 have radially bored cylinders S and 6 with pistons 7 and 8 respectively arranged therein. The pistons 7 and 8 have linked thereto guiding members 9 engaging the inner surface of a barrel ring 10. Said ring 10 is by means of lateral members 11 journalled in bearings 12, 13 in housing 19 with a fixed or adjustable eccentricity. In this way, it will be assured that the barrel ring 10 and the lateral members 11 Will rotate together with the drive shaft 1 so that the guiding members 9 will on ring 10 carry out only a slight sliding movement, in conformity with the respective eccentricity.

Connected to cylinders 5 and 6 are flow passages 15 and 20, respectively. These passages are in communication with the suction and pressure side of the pump, depending on the respective position occupied by each cylinder during a revolution of the drive shaft 1. Between the stationary housing 9 of the pump and the rotor sections 3 and 4 there are in axial direction arranged fitting members 17 and 22 each of which is provided with passages 16, 21, respectively. The passages 16, 21 establish communication between the fiow passages 15 and 20 and the suction conduits 18, 18' and the pressure conduits 23, 23. According to a simplified design of the pump, the said fitting members may be omitted and the rotor sections may be pressed directly against the axially inwardly directed lateral surfaces of the housing 19. Furthermore, the said fitting members may be connected to the housing for rotation therewith or may be connected to the rotor for rotation therewith.

For purposes of sealing the sliding surfaces between the rotating rotor sections and the fitting members which will be assumed as being non-rotating, as well as the Outer lateral surfaces which face each other and pertain to the fitting members and the inner lateral surfaces of the housing, the rotor sections, 3, 4 by means of the pressure medium are pressed against the fitting members 17, 22, which in turn are pressed against the respective associated axial inner lateral surfaces of the housing 19. To this end, between the rotor sections 3, 4 in the radially inner range adjacent the driving shaft 1 there is provided an annular gap chamber 24 receiving a pressure medium from the passage 21 which leads to the pressure conduit 23 and pertains to the fitting member 22 through a recess 25 and through an annular groove 26 in the rotor section by means of the bore 28. In this way, in the gap chamber 24 a high pressure builds up by means of which the rotor sections 3, 4 are pressed away from each other and against the respective associated fitting members 17, 22. The gap chamber 24 is radially outwardly sealed by a highly elastic sealing ring 30 and on the radially inner side of which rotor section is sealed by a highly elastic sealing ring 34 The sealing rings 34 are placed into grooves 31 in the axial range of the splined shaft profile 2 near the axial outer lateral surfaces of the rotor sections 3, 4 and into recesses 32 forming pairs with the grooves 31 and pertaining to rotor sections 3, 4, and are held on their axially outer side by a transversely divided supporting ring 33.

For purposes of maintaining a pressure in gap chamber 24 which corresponds to the operational pressure of the fluid medium for purposes of obtaining a good seal of the fluid passages at the sliding areas, only a small quantity of pressure fluid is required. The bore 28 may, therefore, be narrow.

The fitting members 17, 22 which are non-rotatably connected to the housing 19 have plane inner lateral surfaces facing the rotor sections 3, 4 and also have spherical outer lateral surfaces facing the housing 19. In view of the last-mentioned feature, lack of precision with regard to the assembly, especially aligning errors and deforming influence will be compensated for, whereas the radial escape of the central portion of the driving shaft 1 in view of radial forces will be absorbed in the plane sealing surfaces perpendicular to the axis without affecting the sealing effect, for instance by edging or increasing the friction at the sliding surfaces because at said sliding surfaces a sliding will occur during the operation of the pump.

Pressure springs 35 inserted into axial bores 36 and 37 of the rotor sections 3, 4 will at a lack of pressure in the gap chamber '24, for instance when starting the pump, be pressed against the sealing surfaces.

The construction according to FIG. 2 has a wider field of employment than the construction of FIG. 1 without sacrificing the advantageous space-saving overall structure and without any restricting conditions. The construction shown in FIG. 2 of a hydrostatic radial piston pump differs from the embodiment of FIG. 1 primarily only with regard to the design of the two- sectional rotor 3, 4. Therefore, it will suffice to describe only said rotor and the respective parts of the housing 9 which in axlal direction are immediately adjacent thereto. The remaining housing parts correspond to those of the embodlment of FIG. 1. In contrast to the illustration of FIG. 1, in FIG. 2 the sectional plane is located perpendicular to the eccentricity between the housing and the rotor so that the pistons 7, 8 in cylinders 5, 6 occupy their central stroke position.

Each cylinder 5, 6 has a fluid passage 15, 20 leading axially to the lateral surface of the respective rotor sections 3, 4 which lateral surface faces the housing 19. The mouth of said passages 15, 20, will, during rotation of the rotor pass over the mouth of the passages 16, 20 which are located in the fitting members 17, 22 and communicate with the housing passages 18, 18 and 23, 23', Thus, over a short period of time communication will exist between rotor passages and the housing passages. Radially outside and radially inside the mouth of the rotor passages 15, 20 the axially outer lateral surfaces of the rotor sections 3, 4 are provided with two annular grooves 26, 27 which are coaxial with the axis of rotation. Depending on the adjustment of the eccentricity between the rotor and the housing, one annular groove 26 communicates with the pressure side facing passage 16 and the other annular groove 27 communicates with the suction side facing passage 21 of the fitting members 17, 22 or vice versa, and, more specifically, through a recess 25, 25.

Similar to the embodiment of FIG. 1, between the two rotor sections 3, 4 there is provided an annular gap chamber 24 which extends from the driving shaft 1 radially outwardly up to approximately the radial spacing of the cylinder bottoms from the axis of rotation. This gap chamber 24 is formed by turned-out portions in the two rotor sections 3, 4. Instead of this design of the gap chamber 24 as illustrated in FIG. 2, it is also possible to form the gap chamber by providing a turned-out portion only in one of the two rotor sections, whereas the other rotor section is provided with a smooth inner side wall.

The gap chamber 24 is sealed with regard to the separating slot which extends in radial direction up to the outer rim of the two rotor sections by means of an inserted highly elastic sealing ring 30. Radially inside the sealing ring 30 in axial direction, the end faces of the rotor sections which face each other are provided with blind bores 36, 37. From the bottom of said bores 36, 37 relatively narrow bores 28, 29 lead to the annular grooves 26, 27 in the axially outer lateral surfaces of the rotor sections 3, 4. These narrow bores 28, 29 are closed with regard to the gap chamber 24 by check valves which, in the particular embodiment shown for formed by balls 40. These balls 40 are by means of pressure springs 35' pressed against the bottom of said blind bores 36, 37. The said pressure springs 35 simultaneously serve at a lack of medium in gap chamber 24 for pressing the rotor sections 3, 4 against the fitting members 17, 22 and against those lateral surfaces of housing 19 which are adjacent to the rotor.

On the radially inner side, the gap chamber 24 is sealed in the same manner as with the embodiment of FIG. 1, by sealing rings 34 which are inserted into grooves 31 in driving shaft 1 and in corresponding recesses 32 in the rotor sections 3, 4 and which in axial direction toward the outside are supported by transversely divided supporting rings 33. For purposes of directly compensating for the pressure in the gap chamber 24 when the pressure drops in the conduit 23 leading to the consumer station, between the blind bore 37 starting from the gap chamber 24 on one hand and the sealing chamber 6 on the other hand, there is provided a narrow compensating bore 42. A similar compensating bore 41 is also provided between the cylinder chamber 5 of the rotor section 3 and the blind bore 36.

By means of this construction of the rotor of the hydrostatic radial piston pump according to FIG. 2, it is made possible that with this double pump not only the cylinders of one rotor section can be connected to another consumer station than the cylinders of the other rotor section without leakage losses between the cylinders of one row and the cylinders of the other row when the two consumer stations are under a different load, but it is also possible with this construction to exchange the pressure side and the suction side. Thus, the eccentricity of the rotor relative to the housing can from the central position be adjusted in both directions, and the delivery of the pump can be varied. Also, the employment of the pump as motor is possible.

The necessity of inserting check valves in bores 36, 37 in order to prevent the occurence of leakage flow between the two cylinders at different loads thereof has,

however, the undesired effect that the pressure prevailing in the gap chamber 24 will always respond to the highest pressure produced in one of the connecting openings 16, '21 of the fitting- members 17, 22, and thus to the highest pressure in the cylinders 5,6. During a reduction in the pressure in thepressure line' or in one of the pressure lines, the pressure prevailing in the gap chamber 24 prior'to the pressure drop will be maintained because the check valves will, when the working pres sure drops in the passages leading to the consumer station or stations prevent an automatic pressure equalization. As a result, a higher force is exerted upon the lateral surfaces of the runner parts which act as sealing surfaces than is required for maintaining the sealing efiectbet-ween said surfaces and the adjacent lateral surfaces of the fitting members. Consequently, the friction losses occurring on said sealing surfaces during the rotation of the rotor become, relative to the pump, output, too high so that in an undesired manner the degree of efficiency of the pump drops. This drawback can be overcome by the compensating bores 41, 42 through which the gap chamber 24 directly communicates with one of the cylinder chambers 5, 6 of each cylinder row.

For axially guiding the two axially displaceable. runner sections 3, 4 in the splined profile 2 of the drive shaft, the splined shaft profile is at the separating gap of the two runner sections interrupted by a turned-out portion 38 of rectangular crossgsection which extends. to the bottom of the splined profile. A fitting ring 39 of approximately the same outer'diameter as the splined shaft profile 2 is inserted into said turned-out portion 38.

Instead of the embodiment illustrated in FIG. 1 concerning the radially outer seal of the gap chamber 24, it is advantageous for facilitating the assembly to design the seal in conformity with FIG. 3. With this design, one rotor section, for instance the rotor section 3, is provided with a turned-out portion 43 coaxial to the axis of rotation, whereas the other rotor section, in the illustrated example the rotor section 4, is provided with an annular protrusion 44 which, in assembled condition of the two rotor sections 3, 4 extends with axial and radial play into the turned-out portion 43 of the rotor section 3.

Between oppositely located radially outer cylindrical confining surfaces of the turned-out portion 43 and the protrusion 44, a highly elastic sealing ring 45 is inserted between the two rotor sections 3 and 4. A further highly elastic sealing ring 46 is arranged between the cylindrical outer surface of the transversely divided fitting ring 39 which is inserted in the turned-out portion 38 of the splined profile 2 of shaft 1, and the radially inwardly directed cylindrical confining surface of protrusion 4.

The gap chamber 49 between the two sealing rings 45, 46 has an annular surface which is so dimensioned that there will be exerted a sealing force corresponding to the prevailing highest working pressure. This force is conveyed through the axially outer lateral surfaces of the runner sections of the fitting members and the associated lateral surfaces of the housing which are adjacent said rotor. The said gap chamber 49 communicates through bores 28, 29 with the non-illustrated annular grooves in the axially outer lateral surfaces of the runner sections 3', 4.

These bores 28, 29, similar to the embodiment of FIG. 2 are safeguarded by check valves which comprise compression spring-loaded balls 40 arranged in axial blind bores 47, 48. For purposes of pressure equalization when the working pressure drops in one or in both cylinder rows, the gap chamber 49 communicates with one cylinder chamber of each row by means of a compensating bore. In FIG. 3 there is inserted a compensating bore 42 leading from the axial blind bore 48 to the cylinder chamber 6.

When designing the seal in conformity with FIG. 4,

the axially outer lateral surfaces of the rotor sections 3, 4 are provided with two annular grooves 26, 27 each. Between said annular grooves there extend the connectpassages in the lateral surfaces which connecting passages lead from the cylinders to the housing conduits. Of these connecting passages, FIG. 4 shows the passage leading from cylinder 5 toward the outside through connecting passage 15.

The rotor section 3 has that lateral surface thereof which faces the rotor section 4 in the radially inner range provided with a turned-out portion for receiving a sealing ring 50 and is radially outwardly provided with a protrusion 44 and with a further turned-out portion 43. The rotor section 4 has that lateral surface thereof which faces the rotor section 3 provided with a turned-out portion 43 in the radially inner range of said lateral surface and is radially outwardly provided with a further turnedout portion 43'. The rotor section 4 is furthermore radially outwardly provided with a protrusion 44. Protrusions and turned-out portions of one and the same rotor part are separated from each other by coaxial cylindrical surfaces against which sealing rings 50, 51 and 52 rest. Each protrusion of one rotor section has associated therewith a turned-out portion of the other rotor section in such a way that both have the same intermediate distance from the axis of rotation, and the radial extension of which protrusion will be less to an extent determined by the adjacent sealing rings than the radial extension of the associated turned-out portion. Also the height of the protrusion or the depth of the turned-out portions corresponds to the thickness of the adjacent sealing rings.

Thus, two gap chambers 53, 54 are separated from the space between the two rotor sections 3, 4 by a total of three highly elastic sealing rings 50, 51 and 52 concentrically arranged with regard to the driving shaft 1. The outer gap chamber communicates through the check valve 55 provided with a valve ball 40, and a bore 57 with the radially outer annular grooves 26, whereas the inner gap chamber 53 communicates through a similar check valve 51 and a bore 58 with the radially inwardly located annular groove 27. The radially innermost sealing ring 50 by means of its innermost sealing surface rests against a transversely divided fitting ring 39 mounted on the drive shaft 1 in a turned-out portion 38 which interrupts the splined shaft profile. By means of its radially outer sealing surface, said ring 50 rests against the radially inner cylindrical confining surface of the protrusion 44 of the rotor section 3. By means of the fitting ring 39', similar to the above-mentioned embodiment, the two rotor sections 3, 5 are held against displacement in axial direction. For purposes of pressure compensation when the working pressure drops in the cylinders, the gap chambers 53, 54 communicate through compensating bores with a cylinder of the respective associated row. Of these bores, FIG. 4 shows a compensating bore 41 arranged between the blind bore 47 starting at the gap chamber 53 on one hand and the cylinder 5 on the other hand. The bores 57 and 58 which connect the blind bores 47, 48 with the annular grooves 26 and 27 are with this embodiment relatively Wide and merely at that portion directly adjacent to the entry into the blind bore have a considerably less free width for limiting the flow-through of the pressure medium.

In order to assure that the sealing pressure will always occur only at the necessary stations, according to the embodiment of FIG. 5, to each pair of cylinders 5, 6 there pertains a common sealing piston 124 which is journalled in the bores 125 and 126 arranged in the rotor sections 3 and 4 and serving as partial pressure chambers. These bores communicate through an opening 127, 128 with the flow passages 15, 20 respectively so that the pressure medium presses the rotor sections 3 and 4 laterally apart and thus against the sealing surfaces between the rotor sections and the fitting members and between the latter and the housing, which means toward the left and toward the right, respectively. Seals 131 and 132 on the sealing piston 124 see to it that the pressure does not escape into the chamber between the rotor sections.

A number of springs 133 presses the two rotor sections 3 and 4 continuously away from each other and assures not only a continuous engagement with the sealing surfaces but also brings about a minimum sealing pressure when the pump pressure has not yet developed, for instance, during the starting of the pump.

What is claimed is:

1. A fluid pressure device having a housing, rotor means rotatable in the housing and having radial cylinders, radial pistons in the cylinders and means for causing the pistons to reciprocate in the cylinders when the rotor means rotates in the housing, said rotor means having lateral surfaces and said housing having surfaces adjacent the lateral surfaces of said rotor means, respective high pressure and low pressure passage means extending from outside the housing into the housing and to the cylinders to supply fluid to and receive fluid from the cylinders, said passage means extending through said lateral surfaces of the rotor means and the adjacent surfaces of the housing, said rotor means comprising two axial sections having closed chamber means therebetween, bores being provided in the face of each rotor section which faces the other rotor section, there being one of said bores for each cylinder in the pertaining rotor section, passages connecting each of said bores with the pertaining cylinder, and a sealing piston in each bore operatively engaging the other rotor section, said bores in the respective rotor sections being aligned with each other, and the piston in each bore being integral with the piston in the aligned bore in other rotor section, a drive shaft extending into the housing and on which said rotor sections are non-rotatably mounted while being axially movable thereon relative to each other, and means for supplying pressure fluid from said high pressure passage means to the said chamber means between said rotor sections for distributing force to plural pistons urging the rotor sections laterally into sealing engagement with the respective adjacent surfaces of said housing.

2. A fluid pressure device according to claim 1, in which the eifective area of all of said sealing pistons projected on a plane perpendicular of the axis of rotation of the motor means is about 1.5 to about 1.8 times as large as the total area of the corresponding projection of said passage means where said passage means passes from the rotor means into the housing.

References Cited UNITED STATES PATENTS 2,891,797 2/1959 Bourassa et al. 103-161 2,895,426 7/1959 Orshansky 103161 3,044,412 7/1962 Orshansky 103-461 3,094,077 6/1963 Cadiou 103--161 3,122,104 2/1964 Byers 103-161 3,304,883 2/1967 Eickmann 103161 3,357,362 12/1967 Orr 103161 WILLIAM L. FREEH, Primary Examiner

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