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Publication numberUS3554095 A
Publication typeGrant
Publication dateJan 12, 1971
Filing dateJul 3, 1968
Priority dateJul 7, 1967
Also published asDE1703538A1
Publication numberUS 3554095 A, US 3554095A, US-A-3554095, US3554095 A, US3554095A
InventorsBobst Gerhard
Original AssigneeVon Roll Ag
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Axial piston unit
US 3554095 A
Abstract  available in
Images(3)
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Claims  available in
Description  (OCR text may contain errors)

United States Patent [72] Inventor Gerhard Bobst Oensingen, Switzerland [21 Appl. No. 742,396 [22] Filed July 3, 1968 [45] Patented Jan. 12, 1971 [73] Assignee Von Roll A.G.

' Gerlafingen, Switzerland a corporation of Switzerland [32] Priority July 7, 1967 [33] Switzerland [31] No. 9806/67 [54] AXIAL PISTON UNIT 9 Claims, 6 Drawing Figs.

[521' US. Cl 92/57, 1 91/486, 91/489, 92/156 [51] Int. Cl F04b 1/26, F01b13/04,F01b31/l0 [50] Field otSearch .L 103/162, 162V,162B;230/178;92/57, 156; 91/198, 199

[56] References Cited UNITED STATES PATENTS 2,064,299 12/ 1936 Ferris et al 103/ 161 2,155,455,, 4/1939 Thoma 103/162 2,781,775 2/ 1957 Merrill 2,871,798 2/1959 103/162 3,137,243 6/1969 103/162 3,253,551 5/1966 103/162 2,623,353 12/1952 308/9 2,941,853 l/1960 Bartholomaus 308/9 3,208,395 9/1965 Budzich 308/9 FOREIGN PATENTS 1 625,655 1011944 01661131116111 103/162 791,992 3/1958 01651131186111"... 103/162 1,142,912 9 l 9 5] France 308/9 Primary Examiner-William L. Freeh Attorney-Wemer W. Kleeman ABSTRACT: An axial piston unit of the type comprising a drive shaft and a cylinder having an inclined cylinder axis. Support means serve to support the drive shaft, such support means incorporating a substantially ring-shaped hydrostatic;

PATENTED JAN 1 219m sum 1 BF 3 INVENTOR BY {Mad ATTORNEYJ PATENTEBYJAN 12m?! 3,654,095

SHEET 2 OF 3 F ig. 2 24 INVENTOR Get/ 1'90 202.?7

BYMIM ATTORNEY;

PATENTED JAN 1 2 1971 SHEET 3 OF 3 INVENTOR Gav /Mn) B0 317 AT'TQRNEYS AXIAL PISTON UNIT BACKGROUND OF THE iNveNrioN The presentinvention relates to an improved axial piston unit, such as is employed in the hydraulic oil art both as a pump as well as also a=motor. I

- half of the pistons being subjected to such pressure..Con-

sequently, considerable forces are exerted upon the drive plateor flange, sometirnes referred to'as 'a wobble or swash plate. Thus, for iristance,',withj a seven-piston unit having a piston diameter of 25 millimeters, wherein altematingly three or four pistons are loaded by an operating pressure of 200 atmospheres absolute pressure, with a small stroke, that is to say, when the axis of the cylinder and the'axis'of the drive shaft are approximately aligned or in coincidence with one another, the thrust or axial force amounts to approximately 3 tons (for three Ioaded pistons)-or approximately 4 tons (for four loaded pistons). v

These considerable forces must be transmitted through the agency of suitable bearings to the housing. Because of the compact construction of these units only very little place is available for accommodating such bearings. Thus, special types of roller or antifriction bearings have been developed,

' especially axialcylindrical roller bearings which are presently used quite, extensively. Still, it was not'possible to satisfy all the requirements by. means of these bearings. This was especially so if they were used in an environment where, on account of the size of the installation, even small standstill periods for the exchange of the bearings do not have to be accepted, or where because of. reasons of space greater loaded units had to be employed where the longevity or lifespan of the aforementioned bearingsis not sufficient. It was therefore. apparent to use sliding or friction bearings instead of-lthe roller or antifriction bearings. Hydrostatic sliding bearings, especially hydrostatic axial sliding bearings are known to the art; likewise their use in axial piston units while making use of the operating pressure. They are capable of satisfying increased operating conditions. A hydrostatic axial sliding bearing, consisting of an annular region or zone or a number of ring or annular segmentlike regions or zones, is supplied throughout-its entire periphery with a throttled high pressure which, however, at the unit itself only appears at the pressure side, that is to say, at half of the pistons. During idling when the supply pressure prevails at both pressure sides, the loading of the bearing is relatively twice as great than if a pronounced high pressure would prevail, since then the throttled pressure which is delivered to the axial sliding bearing, depending upon the size of the hearing, proportionally corresponds to the pressure prevailing at both pressure sides, that is to say, it appears at all piston members. Measurements taken of the bearing temperature during idling confirms this characteristic of hydrostatic bearings operated in this manner, in that the bearing temperature during idling exceeds that during the loaded condition.

A further feature or characteristic of the hydrostatic bearing resides in'the fact that it is only operationally reliable during the presence of a pressure. When starting, initially there is no pressure, which first then begins to build up during the start of the stroke movement of the piston members. The'resulting metallic contact of the bearing surfaces can be only then prevented if, prior to the start of the'operation, a special pressure source delivers pressurized oil to the hydrostatic bearings and thus overcomes the phenomena of contact of the bearing surfaces already during standstill. I

Accordingly, it is necessary to differentiate between three operating conditions for the hydrostatic bearings or support of a hydraulic unit, namely: Thestarting condition in which the pressure at the bearing has not yet built up, the idling condition where the bearing is loaded relatively great because of the same or equal pressure which prevails at both pressure sides. and the actual operation where the highpressure prevailing only at one pressure side is available for the entire bearing.

Failure to observe these factors or conditions can lead to considerable difficulties. It is quite possible that because of these reasons there existed a rather limited use of hydrostatic bearings, notwithstanding their unequivocal superiority with regard to their loading and longevity in contrast to roller or antifriction bearings.

SUMMARY OF THEINVEN'I'ION Accordingly, it is a primary object of the present invention to provide an improved axial piston unit which effectively overcomes the aforementioned drawbacks of the prior art 7 structures.

Another, more specific object of the present invention relates to a solution for effectively controlling the characteristics of hydrostatic sliding bearings used in hydraulic units, especially the features associated with axial sliding bearings. so that'they can be employed as operationally reliable and maintenance-free structural elements.

Now, in order to implement these and still further objects of the invention, which will become more readily apparent as the description proceeds, the invention contemplates delivering the oil pressure through the agency of at least one throttling device to the hydrostatic sliding bearing, the throttling action of which is at least partially suppressed during low oil pressure and becomes more pronounced with increasing oil.pressure.

BRIEF DESCRIPTION OF THE DRAWINGS The invention will be better understood, and objects other than those set forth above, will become apparent. when consideration is given to the following detailed description thereof. Such description makes reference to the annexed drawings wherein:

FIG. 1 is a partial sectional view through an axial pistonconstant unit employing a first embodiment of the inventive thrust or axial bearing;

FIG. 2 is a partial sectional view through a second embodiment of thrust or axial bearing employed with an axial pistonadjustable unit;

FIG. 3 is a partial sectional view of the arrangement of FIG. 2, taken substantially perpendicular to the sectional view shown in FIG. 2;

FIG. 4 is a partial sectional view through an embodiment of thrust or axial bearing unit substantially of the-type according to FIG. 2, but here used with an axial piston-constant unit;

FIG. 5 is a fragmentary cross-sectional view showing details of a throttling device; and 7 FIG. 6 is a similar fragmentary sectional view of a second embodiment of throttling device.

DESCRIPTION OF THE PREFERRED EMBODIMENTS Describing now the drawings, in FIG. 1 reference numeral 1 bores for the infeed and outfeed of the pressurized fluid medium, typically oil, are arranged atthe control level cover member 3, and with which bores there are connected appropriate conduits. The drive shaft 4, advantageously formed of one piece or integral with the drive plate 5, is mounted by means of two radial bearings, namely the inner bearing 6 and the outer bearing 7, both shown as cylindrical roller bearings, at the housing 1 and at the drive plate cover member 2, respectively. By means of the end plate or further cover member '8, having a recess 9 in which there is mounted a nonillustrated standard lip-sealing device bearing by means of its sealing lips at the traveling ring 10, the housing 1 is sealed against loss of oil.

The axial or thrust bearing unit, consists of a first rotor ring 11, the cylindrical roller bearing means 12, the second rotor ring 13, and the stator plate 14 which is defined by the portion of the drive plate cover member 2 directed towards the interior of the housing 1. By means of the connection or duct 15, the throttled operating medium flows into the annular or ring zone 16 of the hydrostatic bearing. This ring zone 16 can also be subdivided into individual ring segmentlike areas or zones wherein each possesses a separate infeed for the throttled operating medium beginning at the connection 15.

By means of suitable ball-and-socket joints 100 (FIG. 4) it is possible to hingedly connect the support rod 18 mounted at the cover member 3 and which guides the cylinder 17 as well as the piston rods 19 and their associated piston members 20, with the drive plate 5.

The sliding thrust or friction bearing functions in the following manner: The operating medium flows throttled through the connection 15, in a manner which will be described more fully hereinafter, into the ring zone or area 16, raising the second rotor ring 13 to such an extent until there appears at this ring zone 16 that pressure which is in equilibrium with the thrust or axial force exerted by the piston members 20. This ring member 13 can therefore rotate without contacting the a stator 14, so that for the frictional force there is only operative the shearing force of the operating medium which is generated by such operating medium.

The axial force or thrust of the piston members is transmitted by the first rotor ring 11, fixedly seated upon the drive shaft 4, to the roller bearing unit 12 and from this location to the second rotor ring 13.

The frictional force of the roller bearing 12 is dependent upon the axial force or thrust of the piston members 20 and, accordingly, during idling is very small, in fact, smaller than the frictional force of the hydrostatic sliding bearing 13, 14. This means that during this operating condition the second rotor ring 13 acts as a stator, in other words is at standstill or only rotates with a very small rotational speed. Now, if the operating pressure increases, then also the friction of the roller bearing 12 increases and, at a certain operating pressure, it becomes greater than the shear force of the operating medium at the hydrostatic slide bearing 13, 14, this shearing force essentially only being dependent upon the viscosity and the size or height of the gap. It therefore increases considerably slower or hardly with the operating pressure, so that at a certain operating pressure the second rotor ring 13 begins to rotate and the roller bearing remains at standstill or only rotates very slowly with respect to both rotors.

The change from the one to the other of both-mentioned operating conditions occurs when going from idle to load. Initially, the axial force or thrust is small, so that only the roller bearing is effective. Then, when during the operation the pressure builds up, the moment occurs where the hydrostatic sliding or friction bearing becomes completely effective and the roller bearing practically is at standstill. Theoretically, the situation could occur where each of both bearings assumes one-half of the rotational speed. However, in practical operation always one or both operating conditions exists, where at one time the roller bearing and at one time the hydrostatic slide bearing has imparted to it almost the entire rotational speed. I i

In the arrangement shown in FIGS. 2 and 3 there is again depicted a drive shaft 4, a drive plate or flange 5, an inner radial bearing 6 and an outer radial bearing 7. The construction of the axial or thrust bearing arrangement is similar to that of FIG. 1, however, the individual elements are arranged somewhat differently. Reference numeral 11 designates the first rotor ring which in this case bears directly against the drive flange or plate 5, reference numeral 12 the cylindrical roller bearing, and reference numeral 13 the second rotor ring which is mounted together with the inner radial bearing 6 upon the drive shaft 4. The inner radial bearing 6 has an inner ring or race 6', but no external ring or race, and therefore travels in the second rotor ring 13 which, for this reason, must be manufactured of ball-bearing steel or a similar material.

The second rotor ring 13, the bearing or support surface 13' of which is spherical as in FIG. 1, which, however, is not in any way an absolute necessity for its functionality although advantageous, travels in the stator 14. This stator 14 possesses the form of a bowl having a spherical convex external surface 21. This stator 14 is stationary with respect to the rotor ring 13, yet it is pivotable with respect to the housing 22. For this purpose, a slot or recess 23 is formed in the housing 22, as best shown by referring to FIG. 3, and in which the shaft 4 can move from the illustrated extreme position for maximum stroke through the null-stroke position until reaching the other extreme position of maximum stroke.

A flange sleeve 14' is located at the stator 14 which forms a gap or slot 101 with the shaft member 4 through which the leakage oil of the hydrostatic sliding bearing, arranged between the rotor ring 13 and the stator 14, can flow away.

The hydrostatic sliding bearing is impinged by the momentary high pressure through the agency of the bore 24, the check or nonreturn valve 25, here shown in the form of a ball, and the displacement of which can be limited, or regulated by the screw 26, and the channel 27. One or a number of nonillustrated channels leads from the outside region 21 of the stator 14 to the hydrostatic bearing and supply such throughout the entire annular circular region. In so doing, the momentary low-pressure side is blocked by the check valve 25 from the high-pressure side.

In FIG. 4 there is depicted a hydrostatic axial slide bearing which has been mounted at a constant unit. The arrangement is essentially similar to that of FIGS. 2 and 3, but it is possible to dispense with the stator bowl 14 of FIG. 2 because of the fixed pivot angle. Reference numeral 11 designates the first rotor ring upon which travels the cylindrical roller bearing 12. A further bearing ring 12' is situated in the second rotor ring 13, the rear portion of which together with the stator surface 14 forms the hydrostatic slide bearing. The stator surface 14, similar to FIG. 1, is arranged in the drive plate cover member 2 at the inside thereof, which provides a portion of the housing l of the unit. The hydrostatic slide bearing 13, 14 is supplied via the conduit 15, and the pressurized medium is distributed by means of nonillustrated branch conduits into the spherical bearing ring surface arranged and distributed about the drive plate 5. The admission of the momentary high pressure to the hydrostatic slide bearing takes place externally of the bearing by means of a nonillustrated conventional reversing valve.

The stability of the hydrostatic sliding bearing, that is to say, the maintaining of a gap suitable for the function of the bearing, even when the load changes, is insured in known manner in that a stationary or fixed prethrottling device is connected before the hydrostatic slide bearing, whereas the bearing gap provides a variable throttle. If, for example, the bearing gap is too large with regard to the equilibrium condition, then the pressure in the bearing drops and the gap becomes momen tarily smaller. On the other hand, in reverse fashion with a gap which is too small the pressure in the bearing increases and the rotor ring is raised until the equilibrium condition has been maintained.

The prethrottling valve depicted in FIG. 5 provides a throttling arrangement which is dependent upon pressure, and specifically, in such a manner that with increasing pressure the throttling is increased. This valve consists of a hollow throttling piston 28 having a conically tapered portion 28' and which is pressed into its cylindrical bore by the spring 29 towards the infeed mouth 30. With small operating pressure or during idling, the spring 29 presses the piston 28 against the mouth portion 30 and therefore at least partially frees the throttle cross section, so that the operating medium can flow with very little throttling through the channel 32 into the compartment 33 and through the through bored nut member 34, against which bears the spring 29, into the compartment 35,

the pressure of which is identical with the pressure in the hydrostatic slide bearing. I l I I I The throttle cross section is formed by theconically tapered throttle piston 28 and the edge portion 31. However, this throttle piston 28 can be also cylindrical and possess at least one groove situated at its external or outer surface, the effective length of which increases with increasing operating pressure. The groove could also be provided at the cylindrical wall of the compartment 33, in which case, then, the throttle piston 28 would possess a smooth wall construction The entire prethrottling valve can be, as shown in FIG. 5, mounted in a bore which is situated at the stator or in a special in FIG. 6 there is shown a slight variant embodiment of the prethrottling valve depicted in FIG. 5. in order that the spring member 29 can be dimensioned smaller, a small measuring piston 36 is serially arrangedbefore the throttle piston 28 and the end face of the throttle piston 28 is connected via a channel 37 to a pressureless compartment, so that the pressure a'cting upon the measuring piston 36, on the one hand, is in equilibrium with the pressure acting in the compartment 33 upon the throttle piston 28 and, onthe other hand, with the force of the spring member 29.

The hydrostatic sliding or friction hearing which is impinged by the operating pressure of an axial piston unit and having the prethrottle valve arrangement provide s a bearing which is eflicient and operationally reliable forthe majority of operating conditions. During starting, where the pressure has not yet built up at the slide bearing and during idling, the roller bearing arranged in series with the slide bearing provides a supplementary device by means of which it is possible to prevent the completely insignificant starting wear at the slide bearing, and further, to additionally reduce or, in fact, prevent the temperature increases which occur during idling, which already are effectively reduced by the prethrottling valve due to the taking over of a portion of the rotational speed.

it should be apparent from the foregoing detailed'descrip tion, that the objects set forth at the outset of the specification have been successfully achieved. I

1. An axial piston unit of the type comprisin a drive shaft and a cylinder having an inclined cylinder axis, means defining an outlet for said axial piston unit, support means for said drive shaft, said support means incorporating a substantially ring-shaped hydrostatic axial sliding bearing means acted upon by the pressure of the fluid vmedium, said hydrostatic axial sliding bearing means including rotor means and stator means, and at least one throttle means for delivering the pressure of the fluid medium to said hydrostatic sliding bearing means, said throttle means communicating with said outlet and operably communicating with .said axial sliding bearing, the throttling action of said throttling means being at least partially suppressed during low pressure of the fluid medium and becoming more pronounced during increasing pressure of the fluid medium. I

2. An axial piston unit as defined in .claim 1 wherein said support means for said drive shaft'further includes an axial roller bearing means arranged in series with said hydrostatic axial sliding bearing means;

tapered at one end region thereof.

6. An axial piston unit of the type comprising a drive shaft and a cylinder having an inclined cylinder axis, support means for said drive shaft, said support means incorporating a substantially ring-shaped hydrostatic axial sliding through the agency of said measuring piston, bearing means acted upon by the pressure of the fluid medium, said hydrostatic axial sliding bearing means including rotor means and stator means, and at least one throttle means for delivering the pressure of the fluid medium to said hydrostatic sliding bearing means, the throttling action of said throttling means being at least partially suppressed during low pressure of the fluid medium and becoming more pronounced during increasing pressure of the fluid medium, said throttle means including a substantially cylindrical throttle piston and provided with a throttle cross section which is formed by at least one groove means, the effective length of which is dependent upon the position of said cylindrical throttle piston, said position of said cylindrical throttle piston being regulated as a function of the operating pressure of the fluid medium.

7. An axial piston unit of the type comprising a drive shaft and a cylinder having an inclined cylinder axis, support means for said drive shaft, said support means incorporating a substantially ring-shaped hydrostatic axial sliding bearing means acted upon by the pressure of the fluid medium, said hydrostatic axial sliding bearing means including rotor means and stator means, and at least one throttle means for delivering the pressure of the fluid medium to saidhydrostatic sliding bearing means, the throttling action of said throttling means being at least partially suppressed during low pressure of the fluid medium and becoming more pronounced during increasing pressure of the fluid medium, said throttle means incorporating a throttle piston and a measuring piston, said operating pressure acting upon said throttle piston through the agency of said measuring piston.

8. An axial piston unit of the type comprising a drive shaft and a cylinder having an inclined cylinder axis, support means for said drive shaft, said support means incorporating a substantially ring-shaped hydrostatic axial sliding bearing means acted upon by the pressure of the fluid medium, and an axial roller-bearing means arranged in series with said hydrostatic axial sliding bearing means, said hydrostatic axial sliding bearing means including rotor means and stator means, and at least one throttle means for delivering the pressure of the fluid medium to said hydrostatic sliding bearing means, the throttling action of said throttling means being at least partially suppressed during low pressure of the fluid medium and becoming more pronounced during increasing pressure of the fluid medium, said throttle means including a substantially cylindrical throttle piston and provided with a throttle cross section which is formed-by at least one groove means, the effective length of which is dependent upon the position of said cylindrical throttle piston, said position of said cylindrical 3. An axial piston unit as defined in claim 2, further includ ing a housing, said rotor means including a first rotor element and a second rotor element, said stator means incorporating a pivotably mounted stator bowllike element arranged between said housing and said second rotor element.

4. An axial piston unit as defined inclaim 2, wherein said support means further including at least an inner radial bearing means, said rotor means encompassing a first rotor element and a second rotor element, said second rotor element carrying both said axial sliding bearing means as well as said inner radial roller bearing means.

5. An axial piston unit as defined in claim 1, wherein said I throttle means includes a throttle piston which is conically throttle piston being regulated as a function of the operating pressure of the fluid medium.

9. An axial piston unit of the type comprising a drive shaft and a cylinder having an inclined cylinder axis, support means for said drive shaft, said support means incorporating a substantially ring-shaped hydrostatic axial sliding bearing means acted upon by the pressure of the fluid medium, and an axial roller-bearing means arranged in series with said hydrostatic axial sliding bearing means, said hydrostatic axial sliding bearing means including rotor means and stator means, and at least one throttle means for delivering the pressure of the fluid medium to said hydrostatic sliding bearing means, the throttling action of said throttling means being at least partially suppressed during low pressure of the fluid medium and becoming more pronounced during increasing pressure of the fluid medium, said throttle means incorporating a throttle piston and a measuring piston, said operating pressure acting upon said throttle piston through the agency of said measuring piston.

Patent Citations
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US2064299 *Aug 9, 1933Dec 15, 1936Oilgear CoPump or motor
US2155455 *Apr 20, 1935Apr 25, 1939Hans ThomaHydraulic motor and pump
US2623353 *Jul 8, 1950Dec 30, 1952Paul GerardCombined fluid pressure and mechanical bearing for gas turbine engines
US2781775 *Sep 1, 1950Feb 19, 1957Oilgear CoLimited capacity check valve
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US2941853 *Apr 22, 1955Jun 21, 1960Daimler Benz AgThrust bearing
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Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3858487 *May 26, 1972Jan 7, 1975Hitachi LtdRadial piston type hydraulic pump or motor
US4024799 *Mar 14, 1975May 24, 1977Linde AktiengesellschaftAxial-piston machine with improved bearing arrangement
US4991492 *Nov 28, 1989Feb 12, 1991Stig BrattHydrostatic axial piston machine
US6288848 *Dec 31, 1996Sep 11, 2001Canon Kabushiki KaishaLens barrel and optical equipment
US8276503Jul 9, 2007Oct 2, 2012Robert Bosch GmbhAxial piston machine
WO2008017557A1 *Jul 9, 2007Feb 14, 2008Bosch Gmbh RobertAxial piston machine
Classifications
U.S. Classification92/57, 91/489, 92/156, 91/486
International ClassificationF03C1/00, F04B1/24, F04B1/20, F01B3/00, F03C1/06, F03C1/40
Cooperative ClassificationF01B3/0073, F04B1/2085
European ClassificationF04B1/20C7B, F01B3/00B4G2