EP0509077B1 - Piston pump, especially a radial piston pump - Google Patents

Piston pump, especially a radial piston pump Download PDF

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Publication number
EP0509077B1
EP0509077B1 EP91918718A EP91918718A EP0509077B1 EP 0509077 B1 EP0509077 B1 EP 0509077B1 EP 91918718 A EP91918718 A EP 91918718A EP 91918718 A EP91918718 A EP 91918718A EP 0509077 B1 EP0509077 B1 EP 0509077B1
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EP
European Patent Office
Prior art keywords
pressure
groove
piston
piston pump
control
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Lifetime
Application number
EP91918718A
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German (de)
French (fr)
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EP0509077A1 (en
Inventor
Manfred Kahrs
Gerhard Kunz
Franz Fleck
Hermann SCHÖLLHORN
Gerhard Schudt
Winfried Huthmacher
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Bayerische Motoren Werke AG
ITT Automotive Europe GmbH
Original Assignee
Bayerische Motoren Werke AG
ITT Automotive Europe GmbH
Alfred Teves GmbH
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Publication of EP0509077A1 publication Critical patent/EP0509077A1/en
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Publication of EP0509077B1 publication Critical patent/EP0509077B1/en
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Expired - Lifetime legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
    • F04B1/04—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement
    • F04B1/10—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement the cylinders being movable, e.g. rotary
    • F04B1/107—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement the cylinders being movable, e.g. rotary with actuating or actuated elements at the outer ends of the cylinders
    • F04B1/1071—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement the cylinders being movable, e.g. rotary with actuating or actuated elements at the outer ends of the cylinders with rotary cylinder blocks
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
    • F04B1/04—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement
    • F04B1/0404—Details or component parts
    • F04B1/0452—Distribution members, e.g. valves
    • F04B1/0456—Cylindrical

Definitions

  • the invention relates to a piston pump, in particular a radial piston pump with suction throttling.
  • Piston pumps are often driven by drive units, for example internal combustion engines, whose drive speed is subject to considerable fluctuations. However, the full flow requirement is often already available at a low drive speed and no longer increases as the drive speed increases.
  • drive units for example internal combustion engines
  • the full flow requirement is often already available at a low drive speed and no longer increases as the drive speed increases.
  • the pistons being designed as a throttle point, in each case a throttle disc is arranged between a collar in the eccentric end of the pistons and return springs.
  • Hydraulic oils are hardly compressible.
  • the pressures arising during the movement of the piston can therefore become very large, which leads to the pump on the one hand may be overloaded due to the material being overloaded or, secondly, the resistance of the rotor becomes so great that it stops.
  • a disadvantage of the known types of pumps is thus the non-uniform delivery with steep pressure flanks when the pressure valves and mechanical ones open during operation with effective throttling of the suction flow Noise when opening and closing the pressure valves.
  • These known pumps therefore work relatively loudly and are therefore unsuitable for a number of applications, for example for use in a passenger car.
  • a radial piston pump of the type specified at the outset is known from DE 37 00 573 A1.
  • the rotor of the known radial piston pump is rotatably mounted on a control pin which, in the plane of the piston bores, contains two control slots which have a large cross section which is substantially constant over their entire length compared to the piston bores.
  • a throttling connection leads to a pressure chamber formed in the control pin, from which a channel emerges, which is located at the web and approximately at the outer dead center between the low-pressure control slot and the high-pressure control slot, based on the direction of rotation of the rotor, opens and periodically has connection to the piston bores. This is intended to achieve an improved reversal from the low-pressure side to the high-pressure side at the dead center.
  • a similar piston pump which has a control body with suction and pressure channels, which are connected to control openings.
  • the control openings are separated from one another by webs, pressure control bores being arranged in the webs and being connected to the latter via a check valve opening towards the pressure channel.
  • a rotor rotating on the control body is provided with pistons sliding in piston bores, whereby working spaces are formed, the volume of which increases during operation of the piston pump on the suction channel side and decreases on the pressure channel side.
  • the check valves arranged in the pressure control bores have the effect that the medium to be pumped only reaches the pressure channel when there is a pressure in the working chamber which corresponds to that in the pressure channel.
  • the above-mentioned pump is also not suitable for being provided with a suction throttling, since a pressure build-up in the piston bores which is necessary to avoid pressure surges when the piston bores are only partially filled during the suction stroke cannot be achieved.
  • the invention is therefore based on a piston pump of the type resulting from the preamble of the main claim and has set itself the task of reducing the noise and power consumption of this pump with comparatively simple means.
  • the invention therefore consists in largely preventing a backflow from the pressure channel into the cylinder having a low or negative pressure at the beginning of the slot on the pressure side.
  • the core solution is to pump the incompressible pressure medium via a pressure control groove (hereinafter often referred to as the damping groove) and preferably a non-return valve to the pressure connection or to greatly reduce the backflow of the hydraulic pressure medium through a special design of the slot-shaped pressure control opening.
  • a third approach starts on the suction side in order to reduce noise and improve performance.
  • the damping groove used on the pressure side can also be successfully connected several times in succession by connecting several damping grooves separated from one another by separating webs to the pressure connection in each case via a check valve in the direction of movement of the rotor.
  • the damping grooves can also be individually connected to the pressure channel belonging to the pressure bore via check valves.
  • the invention provides a particularly simple structure for a pump with the features resulting from claim 3.
  • a pump is characterized in essentially characterized in that the pressure medium coming from the cylinder is collected in the pressure channels of a radially inner control pin and the pressure is built up there accordingly.
  • control opening directly connected to the pressure connection of the pump as a bore.
  • the power consumption of the pump can be better limited and the load on the pump components can be reduced by the features resulting from claim 4, since this increases the delivery rate of the pump and the pressure load in the cylinder is reduced.
  • the shape of the damping groove according to the invention is not critical, which leads to advantages in the production of such a groove.
  • a pump according to the invention which uses the features resulting from claim 9, has proven to be particularly effective. A further improvement here can be achieved by using the features according to claim 10.
  • damping channel can be drilled simultaneously with the operation of drilling the pressure and suction channels.
  • Another possibility can be to connect the pressure control groove and the pressure channel of the pressure bore to one another by means of an oblique bore which runs essentially in the radial direction and to insert the check valve into the oblique bore.
  • check valve is particularly advantageously achieved using the features of claim 12, since a backflow behavior is largely prevented here.
  • the check valve can also be used in a separate damping channel.
  • Another possibility according to claim 14 can advantageously consist in connecting the channels leading out of the control pins to the pump outlet only in the pump housing via a check valve.
  • the damping groove thus reduces the gradient of the pressure rise in the piston bores at speeds above the Regulation speed.
  • the piston bores in the areas of the pressure-side control opening are partly filled with pressure medium and partly with gas or with vacuum.
  • the damping groove dampens the backflow of the pressure medium from the pressure side into the piston bore, while the pressure medium-gas mixture is pre-compressed there by the retracting movement of the pistons. This leads to an improved pressure adjustment between the piston bores and the pressure connection, which significantly reduces pressure pulsations.
  • the relatively small cross-section of the damping groove can also cause considerable power losses, which are disadvantageous if, for example in the case of motor vehicles, the drive unit (motor vehicle engine) is limited in its performance or, for example, is to be designed to be as energy-saving as possible .
  • the cross section of the damping groove is preferably small.
  • the damping groove preferably extends over an angular range of 30 ° to 50 ° and can be designed as a triangular groove with an opening angle of approximately 60 °.
  • the design of the length and cross section of the damping groove forms one Compromise between the increased push-out resistance at low speeds and the desired return flow damping at higher speeds.
  • the pressure in the piston bores must not exceed the permissible maximum value in any operating phase.
  • the cross section of the pressure groove adjoining the damping groove is selected according to the invention only so large that the pistons can push out the suctioned volume without an impermissibly high pressure increase in the piston bores against the system pressure at the pressure connection. It has proven advantageous here if the cross section of the pressure groove is at least twice as large as the cross section of the damping groove. It has also proven to be advantageous if the distance from the end of the pressure groove to the entry dead center is equal to or less than the radius of the piston foot bores. This avoids pressure peaks at the end of the piston stroke. An additional damping effect is achieved on the pressure side according to the invention in that the pressure bore opens into the end of the pressure groove adjacent to the web.
  • the damping groove and the pressure groove are formed by a single groove with a continuously increasing cross section, which extends over a partial area or over the entire length of the control opening assigned to the pressure connection.
  • the inventive design of the suction-side control opening achieves a delivery flow character in a piston pump of the type specified, in which a high degree of filling is achieved below a shutdown speed, while above the shutdown speed the delivery rate is almost independent of the speed and constant.
  • the operating temperature of the pump is minimal due to the ambient temperature, the operating medium and changing operating pressures.
  • the favorable filling behavior at speeds below the cut-off speed enables, at least at a higher cut-off speed, a restriction of the means that support the extension of the pistons, such as springs or increased piston weight.
  • pressure pulsations in the suction area of the pump can be reduced to a minimum by the invention.
  • the ratio of the cross section of the throttle groove in mm to the stroke volume of a piston in mm 3 is 1: 700 to 1: 1200, in particular 1: 1000.
  • the throttle groove can be designed as a triangular groove with an opening angle of approximately 60 °. The throttle groove allows, in particular at low speeds, a defined partial filling of the piston bores in the first part of the suction stroke and thereby prevents an excessive pressure drop until the suction bore is reached.
  • the ends of the piston bores facing the control body are offset in the rotor and can be connected to the control openings via piston base bores of smaller diameter.
  • the diameter of the piston foot holes should be chosen so that the piston foot holes have the effect of a throttle orifice.
  • the ratio of the diameter of the piston foot bore and piston bore is preferably between 1: 4 and 1: 7.
  • the radial piston pump 1 shown in FIG. 1 has an essentially disk-shaped pump housing 2, with a continuous longitudinal bore 3 and a cylindrical recess 4 adjoining it.
  • a control pin 5 is fastened in the longitudinal bore 3, for example by being pressed in the recess 4 protrudes.
  • a rotor 6 is rotatably mounted on the control pin 5 in the recess 4, in which a plurality of radially aligned piston bores 7 are formed, in which pistons 8 slide.
  • the pistons 8 are supported with their ends protruding from the piston bores 7 on the inner surface of a cam ring 9, which is mounted eccentrically to the control pin 5 in the recess 4 by means of a roller bearing.
  • the radially inner ends of the piston bores 7 are offset in the rotor 6 and connected to piston base bores 10 which open into the central bearing bore 11 of the rotor 6.
  • control openings 12, 13 are formed in the plane of the piston base bores, which in turn connect to the piston base bores 10 when the rotor 6 rotates.
  • the control opening 12 is located in the suction area of the pistons 8 and is connected via a suction bore 14 to a suction channel 15 which runs in the longitudinal direction in the control pin 5 and which is connected to a suction connection 16.
  • the control opening 13 is in the Pressure range of the pistons 8 and is connected via the pressure bore 17 to a pressure channel 18 formed in the control pin 5 parallel to the suction channel 15.
  • the pressure channel 18 opens into an annular groove 19 which is connected to a pressure connection 20.
  • the rotor 6 is driven via a coupling 21 by a shaft 22 which is mounted in a cover 23 closing the recess 4.
  • the configuration of the control openings 12, 13 in the control pin 5 can be seen from FIGS. 3 and 4.
  • the control opening 12 is divided into three different areas.
  • the first area begins at a distance of approximately 30 ° in the direction of rotation of the rotor 6, indicated by arrow X, after the entry dead center ET, which results from the smallest distance between the control pin 5 and the cam ring 9.
  • This area is designed as a throttle groove 24 of small cross section.
  • the throttle groove 24 has the shape of a triangular groove with an opening angle of approximately 60 °.
  • the throttle groove 24 ensures a defined partial filling of the piston bores 7 and prevents an excessive reduction in pressure before reaching the suction bore 14, thereby reducing pressure pulsations.
  • the narrow throttle groove 24 opens directly into the suction bore 13 which forms the second region of the control opening 12 and which is arranged at a distance of approximately 140 ° from the entry dead center ET is.
  • the suction hole 14 is followed as a third area by a filling groove 26 with a larger cross section, which ends at the exit emergency point AT.
  • the effective regulating speed of the radial piston pump 1 is determined primarily by the position of the suction bore 14, the filling groove 26 with its comparatively large cross section mainly improving the degree of filling at speeds which are below the regulating speed.
  • a short filling groove 26, on the other hand, can largely dispense with a strong throttling of the suction flow in the piston base bores 10, thereby reducing the sensitivity of the pump to dirt. If a low regulation speed is to be achieved, the suction bore 14 can be arranged immediately before the exit emergency point AT and a filling groove 26 can be dispensed with.
  • the control opening 13 connected to the pressure connection 20 is separated from the filling groove 26 in the area of the exit emergency point AT by a web 27. It is divided into two areas, namely a damping groove 28 and a pressure groove 29.
  • the cross section of the damping groove 28 is small. Tests have shown that triangular grooves with an opening angle of approx. 60 ° and an opening width between 0.6 and 1.0 mm are sufficient in many applications.
  • the length of the damping groove 28 is 40 ° in the described embodiment.
  • the damping groove 28 primarily has the task of avoiding the gradient of the pressure increase in the piston bores 7 at speeds that are above the cut-off speed. At these speeds, the piston bores 7 are partly filled with pressure medium and partly with gas when the connection to the control opening 13 is opened.
  • the damping of pressure pulsations is further contributed by the cross section of the pressure groove 29 adjoining the damping groove 28, which cross section is significantly larger, but also limited to a minimum value.
  • the pressure groove 29 extends to the entry dead center ET and thereby allows the pistons 8 to be conveyed until the maximum entry position is reached.
  • the pressure bore 17 opens into the end of the pressure groove 29 which is adjacent to the entry dead center ET and thereby also contributes to the damping effect of the pressure groove 29.
  • Fig. 5 shows a processing corresponding to Fig. 4 for a preferred solution according to claims 1 to 14.
  • the main difference compared to Fig. 4 is that a throttle groove 24 has been omitted on the suction side and the damping groove 28 with on the pressure side
  • Check valve 32 (which roughly corresponds to the previously described damping groove) on the surface of the control pin 5 no longer merges into the pressure groove 29, but is separated from it by a separating web 30.
  • the connection is made via a pressure control bore 31 indicated in FIG. 5 and designed as a radial bore, which is symbolically indicated as line 31 A in FIG. 5.
  • the pressure control bore 31 and thus the damping groove 28 are connected to the pressure connection 20 via a check valve 32 and a damping channel D.
  • the pressure control opening is designed as a pressure groove 29, which is connected to the pressure connection 20 via the pressure bore 17 and a pressure channel 18, as already described in connection with FIG. 1.
  • the check valve 32 can be arranged in the radial bore 31, in the damping channel D, but also at the end of the damping channel D in the connection area to the pressure connection 20 in the housing.
  • the diameter of the pressure control bore 31 is shown here somewhat smaller than the diameter of the bores 14 and 17.
  • the pressure control bore can have the same diameter as the bores mentioned.
  • the width and the diameter of the radial groove shown in FIG. 5 is also largely uncritical and can therefore have the same width as the grooves 26 and 29. It is also possible to provide between the grooves 28 and 29 or instead of the groove 28 a plurality of individual grooves lying in line one behind the other, each of which is connected to the pressure connection 20 via its own check valve. This achieves improved performance and reduced noise.
  • the throttle groove 24 has also been omitted since this results in a considerable simplification of the design of the grooves, which now all have the same shape. The resulting reduction in performance or increase in noise is extremely low, so that this must be regarded as an advantageous solution compared to FIG. 4.
  • the position of the suction bore 14 relative to the filling groove 26 is largely uncritical, as long as only the suction bore 14 is in the region of the filling groove 26.
  • the length of the filling groove largely depends on the desired throttling effect, since the degree of filling of the respective pump cylinder increases with the length of the filling groove 26.
  • the pressure-side control opening 13 according to FIG. 4 has been divided into two grooves separated by a separating web 30, the offset pressure control groove 28 admitting pressure medium from the piston bore 7 (FIGS. 1 and 2) and thus contributes significantly to the pump performance, while a backflow of pressure medium via the channels 18 and D from the pressure groove 29 having a higher pressure into the pressure control groove 28 is prevented by the check valve 32.

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  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Reciprocating Pumps (AREA)

Abstract

The invention relates to a piston pump with suction regulation, especially for motor vehicles, and preferably with a control stub for internal admission. It is the purpose of the invention to provide such a pump with a constant supply flow rate over a wide revolution speed range and the lowest possible losses and low noise level. To this end the present invention proposes several solutions. In a first, the control slot (13) on the pressure side is divided into several grooves (27, 28), which are at least partly connected to the pressure link (20) via non-return valves (32). In another, the shape of the control slot on the pressure side is adapted to the optimum operating mode by making its inlet end relatively narrow to reduce the noise level (noise reduction at high revolution speeds) and keeping an adequate width at the outlet end for the power required. In a third, the arrangement for the slot (12) on the intake side is similar to that of the second solution and its length also affects the degree of filling of the piston drilling.

Description

Die Erfindung betrifft eine Kolbenpumpe, insbesondere eine Radialkolbenpumpe mit Saugdrosselung.The invention relates to a piston pump, in particular a radial piston pump with suction throttling.

Kolbenpumpen werden häufig von Antriebsaggregaten, beispielsweise Verbrennungsmotoren, angetrieben, deren Antriebsdrehzahl erheblichen Schwankungen unterliegt. Der volle Förderstrombedarf ist aber oft schon bei niedriger Antriebsdrehzahl vorhanden und nimmt mit steigender Antriebsdrehzahl nicht mehr zu. Um die Fördercharakteristik diesem Bedürfnis anzupassen, ist es aus der DE-AS 20 61 960 bekannt, bei einer Radialkolbenpumpe mit in einem Gehäuse sternförmig etwa in einer Ebene angeordneten Zylindern und durch eine Exzenterwelle betätigten, federbelasteten Kolben, bei welcher das Pumpmedium über am Umfang des Exzenters angeordnete Nuten angesaugt, durch die hohlen Kolben gepumpt und über mindestens ein Rückschlagventil im Gehäuse weitergefördert wird, die Kolben als Drosselstelle auszubilden, indem zwischen einem Bund im exzenterseitigen Ende der Kolben und Rückstellfedern jeweils eine Drosselscheibe angeordnet ist. Durch diese Ausbildung wird dem Pumpmedium auf der Saugseite mit zunehmender Drehzahl ein zunehmender Widerstand entgegengesetzt, welcher dazu führt, daß von einer bestimmten Drehzahl ab die Fördermenge nicht mehr linear mit dieser Drehzahl ansteigt, sondern einen maximalen Wert erreicht, welcher nahezu unabhängig von einer weiteren Drehzahlsteigerung ist.Piston pumps are often driven by drive units, for example internal combustion engines, whose drive speed is subject to considerable fluctuations. However, the full flow requirement is often already available at a low drive speed and no longer increases as the drive speed increases. In order to adapt the delivery characteristics to this need, it is known from DE-AS 20 61 960, in a radial piston pump with cylinders arranged in a star shape in approximately one plane and by an eccentric shaft actuated, spring-loaded piston, in which the pumping medium extends over the circumference of the Eccentrically arranged grooves are sucked in, pumped through the hollow pistons and conveyed further via at least one check valve in the housing, the pistons being designed as a throttle point, in each case a throttle disc is arranged between a collar in the eccentric end of the pistons and return springs. With this design, the pump medium on the suction side with increasing speed is opposed by an increasing resistance, which means that from a certain speed, the flow rate no longer increases linearly with this speed, but reaches a maximum value that is almost independent of a further speed increase is.

Hydraulische Öle sind kaum kompressibel. Die bei der Bewegung des Kolbens entstehenden Drücke können daher sehr groß werden, was dazu führt, daß die Pumpe zum einen aufgrund der Überforderung des Materials überbelastet werden kann oder zum anderen der Widerstand des Rotors so groß wird, daß dieser stehenbleibt.Hydraulic oils are hardly compressible. The pressures arising during the movement of the piston can therefore become very large, which leads to the pump on the one hand may be overloaded due to the material being overloaded or, secondly, the resistance of the rotor becomes so great that it stops.

Als Abhilfe hierzu ist man dazu übergegangen, sowohl auf der Saugseite als auch auf der Druckseite Steuerschlitze vorzusehen, welche sich über einen größeren Winkelbereich längs der Bewegungsrichtung der Kolbenbohrung erstrecken, um somit den Saugvorgang als auch den Pumpvorgang zu verstetigen. Derartige Pumpen arbeiten durchaus zufriedenstellend. Erhebliche Probleme gibt es allerdings dann, wenn man versucht, derartige mit Steuerschlitzen versehene Pumpen im Saugdrosselbetrieb arbeiten zu lassen. Soweit man noch im niedrigen Umdrehungsbereich arbeitet, die Zylinder also wie bei Pumpen ohne Saugdrosselbetrieb auch, voll mit Druckmittel gefüllt werden, arbeitet eine derartige Pumpe wie eine Pumpe ohne Saugdrosselbetrieb. Wird aber die kritische Umdrehungszahl überschritten, so wird der jeweilige Zylinder während des Saugvorgangs nicht mehr voll mit Hydraulikmittel gefüllt, so daß in dem Zylinder ein sehr niedriger Druck oder ein Unterdruck herrscht, wenn der Kompressionsvorgang des Kolbens beginnt. Wenn nun ein derartiger, Unterdruck aufweisender Zylinder Zugang zu dem unter dem Ausgangsdruck der Pumpe stehenden Druck des druckseitigen Steuerschlitzes hat, wird der Zylinder schlagartig mit Druckmittel gefüllt, welches bei der weiteren Rotationsbewegung des Zylinders in der üblichen Weise verdichtet und vor Erreichen des Endes des druckseitigen Steuerschlitzes wieder aus dem Kolben herausgeschoben wird.As a remedy for this, it has become common practice to provide control slots on both the suction side and on the pressure side, which extend over a larger angular range along the direction of movement of the piston bore, in order to thus stabilize the suction process and the pumping process. Such pumps work quite satisfactorily. However, there are considerable problems when trying to make such pumps provided with control slots work in the suction throttle mode. Insofar as you are still working in the low speed range, that is to say the cylinders are filled with pressure medium as in pumps without suction throttle operation, such a pump works like a pump without suction throttle operation. However, if the critical number of revolutions is exceeded, the respective cylinder is no longer completely filled with hydraulic fluid during the suction process, so that there is a very low pressure or a negative pressure in the cylinder when the compression process of the piston begins. If such a cylinder, which has negative pressure, now has access to the pressure of the pressure-side control slot under the outlet pressure of the pump, the cylinder is suddenly filled with pressure medium, which compresses in the usual manner during the further rotational movement of the cylinder and before the end of the pressure-side is reached Control slot is pushed out of the piston.

Die beschriebenen Vorgänge führen zu einer erheblichen Geräuschbildung, die insbesondere dann sehr nachteilig ist, wenn die Arbeitsumgebung der Pumpe leise ist. Dies gilt beispielsweise für die hinsichtlich Geräuschbedämpfung mit immer mehr Komfort versehenen modernen Kraftfahrzeuge. Im übrigen ist durch die Bewegung des Druckmittels vom Druckkanal über den druckseitigen Steuerschlitz in den Zylinder und wieder zurück ein beachtlicher Leistungsverlust zu verzeichnen, der das Antriebsaggregat der Pumpe unnötig belastet.The processes described lead to considerable noise, which is particularly disadvantageous when the working environment of the pump is quiet. This applies, for example, to modern motor vehicles which are provided with more and more comfort with regard to noise reduction. In addition, due to the movement of the pressure medium from the pressure channel via the pressure-side control slot into the cylinder and back again, there is a considerable loss in performance, which unnecessarily stresses the drive unit of the pump.

Die hier beschriebenen Vorgänge gelten mit Abwandlung in Analogie auch saugseitig, so daß auch hier Maßnahmen zu treffen sind, die eine Geräuschverbesserung bedingen und Leistungsverluste mindern. Dabei ist allerdings zu beachten, daß das saugseitig entstehende Vakuum in dem Zylinder leichter zu beherrschen ist als druckseitig das inkompressible Hydraulikmedium. Es ist daher durchaus möglich, durch Verkürzen des saugseitigen Schlitzes eine Drosselwirkung zu erzielen, so daß auf eine gesonderte Drosselstelle verzichtet werden kann. Hierdurch läßt sich eine Leistungsverbesserung und eine Geräuschverminderung erreichen. Der saugseitige Schlitz beträgt hierbei in seiner Länge nur einen Bruchteil der Länge des druckseitigen Schlitzes. Gegebenenfalls läßt sich auf den saugseitigen Schlitz in seiner Gänze verzichten.The processes described here also apply on the suction side, with a modification in analogy, so that measures must also be taken here which result in an improvement in noise and reduce power losses. It should be noted, however, that the vacuum in the cylinder is easier to control than the incompressible hydraulic medium on the pressure side. It is therefore entirely possible to achieve a throttling effect by shortening the suction-side slot, so that a separate throttling point can be dispensed with. This can improve performance and reduce noise. The length of the suction-side slot is only a fraction of the length of the pressure-side slot. If necessary, the entire suction slot can be dispensed with.

Ein Nachteil der bekannten Pumpenarten ist somit die im Betrieb mit wirksamer Drosselung des Saugstroms auftretende, ungleichförmige Förderung mit steilen Druckflanken beim Öffnen der Druckventile und mechanischen Geräuschen beim Auf- und Zugehen der Druckventile. Diese bekannten Pumpen arbeiten daher verhältnismäßig laut und sind daher für eine Reihe von Anwedungen, beispielsweise für den Einsatz in einem Personenwagen, nicht geeignet.A disadvantage of the known types of pumps is thus the non-uniform delivery with steep pressure flanks when the pressure valves and mechanical ones open during operation with effective throttling of the suction flow Noise when opening and closing the pressure valves. These known pumps therefore work relatively loudly and are therefore unsuitable for a number of applications, for example for use in a passenger car.

Eine Radialkolbenpumpe der eingangs angegebenen Art ist aus der DE 37 00 573 A1 bekannt. Der Rotor der bekannten Radialkolbenpumpe ist auf einem Steuerzapfen drehbar gelagert, der in der Ebene der Kolbenbohrungen zwei Steuerschlitze von in Vergleich zu den Kolbenbohrungen großem, über ihre gesamt Länge im wesentlichen konstantem Querschnitt enthält. Zur Reduzierung von Ungleichförmigkeiten des Flüssigkeitstroms führt bei der bekannten Radialkolbenpumpe vom Hochdruck-Steuerschlitz eine drosselnde Verbindung zu einem im Steuerzapfen ausgebildeten Druckraum, von dem ein Kanal ausgeht, der am Steg und etwa am äußeren Totpunkt zwischen dem Niederdrucksteuerschlitz und dem Hochdrucksteuerschlitz, bezogen auf die Drehrichtung des Rotors, mündet und periodisch Verbindung zu den Kolbenbohrungen hat. Hierdurch soll eine verbesserte Umsteuerung von der Niederdruck- zur Hochdruckseite im Totpunkt erreicht werden. Diese Maßnahme ist jedoch nicht geeignet, um bei einer Regelung der Fördermenge durch Drosselung des Saugstroms Druckschläge beim Übergang der während des Saughubs nur teilweise gefüllten Kolbenbohrung auf das hohe Druckniveau des Steuerschlitzes der Druckseite zu vermeiden. Derartige Kolbenpumpen sind daher seither nicht mit einer Drosselregelung auf der Saugseite verwendet worden.A radial piston pump of the type specified at the outset is known from DE 37 00 573 A1. The rotor of the known radial piston pump is rotatably mounted on a control pin which, in the plane of the piston bores, contains two control slots which have a large cross section which is substantially constant over their entire length compared to the piston bores. In order to reduce non-uniformities in the liquid flow in the known radial piston pump from the high-pressure control slot, a throttling connection leads to a pressure chamber formed in the control pin, from which a channel emerges, which is located at the web and approximately at the outer dead center between the low-pressure control slot and the high-pressure control slot, based on the direction of rotation of the rotor, opens and periodically has connection to the piston bores. This is intended to achieve an improved reversal from the low-pressure side to the high-pressure side at the dead center. However, this measure is not suitable for avoiding pressure surges when regulating the delivery rate by throttling the suction flow when the piston bore, which is only partially filled during the suction stroke, reaches the high pressure level of the control slot on the pressure side. Piston pumps of this type have therefore not been used since then with a throttle control on the suction side.

Aus der US-A-2,529,309 ist eine ähnliche Kolbenpumpe bekannt, die einen Steuerkörper mit Saug- und Druckkanal aufweist, welche mit Steueröffnungen verbunden sind. Die Steueröffnungen sind durch Stege voneinander getrennt, wobei in den Stegen Drucksteuerbohrungen angeordnet sind, die über ein zum Druckkanal hin öffnendes Rückschlagventil mit diesem verbunden sind.From US-A-2,529,309 a similar piston pump is known, which has a control body with suction and pressure channels, which are connected to control openings. The control openings are separated from one another by webs, pressure control bores being arranged in the webs and being connected to the latter via a check valve opening towards the pressure channel.

Ein auf dem Steuerkörper umlaufender Rotor ist mit in Kolbenbohrungen gleitenden Kolben versehen, wodurch Arbeitsräume gebildet werden, deren Volumen im Betrieb der Kolbenpumpe saugkanalseitig zu- sowie druckkanalseitig abnimmt.A rotor rotating on the control body is provided with pistons sliding in piston bores, whereby working spaces are formed, the volume of which increases during operation of the piston pump on the suction channel side and decreases on the pressure channel side.

Die in den Drucksteuerbohrungen angeordeten Rückschlagventile bewirken, daß das zu fördernde Medium erst dann in den Druckkanal gelangen, wenn in der Arbeitskammer ein Druck herrscht, der demjenigen im Druckkanal entspricht.The check valves arranged in the pressure control bores have the effect that the medium to be pumped only reaches the pressure channel when there is a pressure in the working chamber which corresponds to that in the pressure channel.

Die oben genannte Pumpe ist ebenfalls nicht dazu geeignet, mit einer Saugdrosselung versehen zu werden, da ein zur Vermeidung von Druckschlägen bei während des Saughubs nur teilweise gefüllten Kolbenbohrungen erforderlicher Druckaufbau in den Kolbenbohrungen nicht erreicht werden kann.The above-mentioned pump is also not suitable for being provided with a suction throttling, since a pressure build-up in the piston bores which is necessary to avoid pressure surges when the piston bores are only partially filled during the suction stroke cannot be achieved.

Die Erfindung geht daher aus von einer Kolbenpumpe der sich aus dem Oberbegriff des Hauptanspruch ergebenden Gattung und hat sich zur Aufgabe gestellt, die Geräuschbildung und Leistungsaufnahme dieser Pumpe mit vergleichsweise einfachen Mitteln herabzusetzen.The invention is therefore based on a piston pump of the type resulting from the preamble of the main claim and has set itself the task of reducing the noise and power consumption of this pump with comparatively simple means.

Diese Aufgabe wird durch eine Kombination von Merkmalen gelöst, wie sie sich aus Anspruch 1 und dem kennzeichnenden Teil des nebengeordneten Anspruchs 15 ergibt.This object is achieved by a combination of features as it results from claim 1 and the characterizing part of the independent claim 15.

Die Erfindung besteht im Prinzip also darin, druckseitig einen Rückstrom vom Druckkanal in den einen niedrigen oder Unterdruck aufweisenden Zylinder am Beginn des druckseitigen Schlitzes weitgehend zu verhindern. Dabei besteht die eine Lösung im Kern darin, das inkompressible Druckmedium über eine Drucksteuernut (nachfolgend vielfach als Dämpfungsnut bezeichnet) und vorzugsweise ein Rückschlagventil zu dem Druckanschluß zu pumpen oder durch eine besondere Ausgestaltung der schlitzförmigen Drucksteueröffnung den Rückfluß des hydraulischen Druckmittels stark herabzusetzen. Ein dritter Lösungsansatz setzt auf der Saugseite an, um hier eine Geräuschverminderung und Leistungsverbesserung zu erreichen.In principle, the invention therefore consists in largely preventing a backflow from the pressure channel into the cylinder having a low or negative pressure at the beginning of the slot on the pressure side. The core solution is to pump the incompressible pressure medium via a pressure control groove (hereinafter often referred to as the damping groove) and preferably a non-return valve to the pressure connection or to greatly reduce the backflow of the hydraulic pressure medium through a special design of the slot-shaped pressure control opening. A third approach starts on the suction side in order to reduce noise and improve performance.

Die auf der Druckseite angewendete Dämpfungsnut läßt sich auch mehrfach mit Erfolg hintereinanderschalten, indem in Bewegungsrichtung des Rotors mehrere durch Trennstege voneinander getrennte Dämpfungsnuten jeweils über ein Rückschlagventil mit dem Druckanschluß verbunden sind. Selbstverständlich können auch die Dämpfungsnuten über Rückschlagventile einzeln mit dem zur Druckbohrung gehörenden Druckkanal verbunden sein.The damping groove used on the pressure side can also be successfully connected several times in succession by connecting several damping grooves separated from one another by separating webs to the pressure connection in each case via a check valve in the direction of movement of the rotor. Of course, the damping grooves can also be individually connected to the pressure channel belonging to the pressure bore via check valves.

Die Erfindung ergibt einen besonders einfachen Aufbau für eine Pumpe mit den sich aus Anspruch 3 ergebenden Merkmalen. Eine derartige Pumpe zeichnet sich im wesentlichen dadurch aus, daß das Druckmittel, vom Zylinder kommend, in den Druckkanälen eines radial innen liegenden Steuerzapfens gesammelt und dort dementsprechend der Druck aufgebaut wird.The invention provides a particularly simple structure for a pump with the features resulting from claim 3. Such a pump is characterized in essentially characterized in that the pressure medium coming from the cylinder is collected in the pressure channels of a radially inner control pin and the pressure is built up there accordingly.

An sich ist es möglich, die direkt mit dem Druckanschluß der Pumpe verbundene Steueröffnung als eine Bohrung vorzusehen. Die Leistungsaufnahme der Pumpe läßt sich aber besser begrenzen und die Belastung der Pumpenbauteile läßt sich herabsetzen durch die sich aus Anspruch 4 ergebenden Merkmale, da hierdurch die Fördermenge der Pumpe erhöht und die Druckbelastung im Zylinder herabgesetzt wird.As such, it is possible to provide the control opening directly connected to the pressure connection of the pump as a bore. The power consumption of the pump can be better limited and the load on the pump components can be reduced by the features resulting from claim 4, since this increases the delivery rate of the pump and the pressure load in the cylinder is reduced.

Wie weiter oben schon erläutert, läßt sich eine weitere Herabsetzung der Geräuschbildung und eine Verbesserung des Wirkungsgrades durch Maßnahmen erreichen, wie sie in Anspruch 5 dargelegt sind. Hierbei ist allerdings ein etwas höherer Fertigungsaufwand notwendig. In der Praxis kommt man durchaus mit einer einzigen Drucksteuernut aus.As already explained above, a further reduction in noise generation and an improvement in efficiency can be achieved by measures as set out in claim 5. However, a somewhat higher manufacturing effort is necessary here. In practice you can get by with just one pressure control groove.

Für eine korrekte Arbeitsweise der erfindungsgemäßen Dämpfungsnut ist die Berücksichtigung der in Anspruch 6 angegebenen Merkmale wichtig, da hier eine optimale Arbeitsweise erreicht wird.For a correct operation of the damping groove according to the invention, it is important to take into account the features specified in claim 6, since an optimal operation is achieved here.

Folgt man diesen Maßnahmen nicht, so muß befürchtet werden, daß über den Zylinder die Druckunterschiede in den einzelnen Dämpfungsnuten kurzgeschlossen werden, so daß hier mit einer zusätzlichen Geräuschbildung und mit einer Leistungsaufnahme zu rechnen ist.If you do not follow these measures, it must be feared that the pressure differences in the individual damping grooves will be short-circuited via the cylinder, so that additional noise and power consumption can be expected here.

Weiterhin ist es besonders vorteilhaft, zur Optimierung der Arbeitsweise der erfindungsgemäßen Pumpe die in Anspruch 7 angegebenen Merkmale zu berücksichtigen, da andernfalls der Druckunterschied in zwei nacheinander folgenden Zylindern über die Dämpfungsnut selbst, möglicherweise aber auch über die zu lang gewählte druckseitige Steueröffnung, ausgeglichen wird, was wiederum zu Geräuschbildung und zu Leistungsverlust führt.Furthermore, it is particularly advantageous to take into account the features specified in claim 7 in order to optimize the mode of operation of the pump according to the invention, since otherwise the pressure difference in two successive cylinders is compensated for via the damping groove itself, but possibly also via the pressure-side control opening which is selected too long, which in turn leads to noise and loss of performance.

An sich ist die Form der erfindungsgemäßen Dämpfungsnut unkritisch, was zu Vorteilen bei der Fertigung einer derartigen Nut führt. In Verbesserung der Erfindung empfiehlt es sich hierbei, die Merkmale nach Anspruch 8 anzuwenden, da hier durch einen einfachen Fräsvorgang die Dämpfungsnut hergestellt werden kann.As such, the shape of the damping groove according to the invention is not critical, which leads to advantages in the production of such a groove. In an improvement of the invention, it is recommended here to apply the features of claim 8, since the damping groove can be produced here by a simple milling process.

Als besonders wirkungsvoll hat sich eine erfindungsgemäße Pumpe erwiesen, die die sich aus Anspruch 9 ergebenden Merkmale anwendet. Eine weitere Verbesserung hierbei läßt sich durch Nutzung der Merkmale gemäß Anspruch 10 erreichen.A pump according to the invention, which uses the features resulting from claim 9, has proven to be particularly effective. A further improvement here can be achieved by using the features according to claim 10.

Eine weitere Vereinfachung ergibt sich durch Anwendung der Merkmale nach Anspruch 11, indem mit dem Arbeitsgang des Bohrens des Druck- und des Saugkanals gleichzeitig auch der Dämpfungskanal gebohrt werden kann. Eine andere Möglichkeit kann darin bestehen, durch eine schräge, im wesentlichen in radialer Richtung verlaufende Bohrung Drucksteuernut und Druckkanal der Druckbohrung miteinander zu verbinden und das Rückschlagventil in die Schrägbohrung einzusetzen.A further simplification results from the application of the features according to claim 11, in that the damping channel can be drilled simultaneously with the operation of drilling the pressure and suction channels. Another possibility can be to connect the pressure control groove and the pressure channel of the pressure bore to one another by means of an oblique bore which runs essentially in the radial direction and to insert the check valve into the oblique bore.

Die Wirkung des Rückschlagventiles wird besonders vorteilhaft unter Anwendung der Merkmale nach Anspruch 12 erreicht, da hier ein Rückströmverhalten weitgehend verhindert wird. In vorteilhafter Weiterbildung der Erfindung kann man auch das Rückschlagventil in einen gesonderten Dämpfungskanal gemäß Anspruch 13 einsetzen. Eine andere Möglichkeit gemäß Anspruch 14 kann vorteilhaft darin bestehen, die aus den Steuerzapfen herausführenden Kanäle erst in dem Pumpengehäuse über ein Rückschlagventil miteinander zum Pumpenausgang hin zu verbinden.The effect of the check valve is particularly advantageously achieved using the features of claim 12, since a backflow behavior is largely prevented here. In an advantageous development of the invention, the check valve can also be used in a separate damping channel. Another possibility according to claim 14 can advantageously consist in connecting the channels leading out of the control pins to the pump outlet only in the pump housing via a check valve.

Eine weitere Lösung der erfindungsgemäßen Aufgabe läßt sich in einfacher Weise durch die sich aus Anspruch 15 ergebende Merkmalskombination erreichen. Hierbei wird die Dämpfungsnut quasi einstückig, ohne gesonderte Verbindung zum Druckanschluß, mit der Drucksteueröffnung verbunden, wobei die so gebildete Dämpfungsnut aber einen sehr viel geringeren Querschnitt als die der Drucksteueröffnung zugeordnete Nut aufweisen sollte, die nachfolgend vielfach als Drucknut bezeichnet wird. Es ergibt sich hierdurch ein sehr viel einfacherer Aufbau der Pumpe, der allerdings zwei Einschränkungen unterliegt. Zum einen ist die Dimensionierung der einzelnen Nuten von der Pumpenleistung und der gewählten Drehzahl abhängig, ab der die Fördermenge nicht mehr erhöht wird (Abregeldrehzahl). Somit muß zur Optimierung des Geräusch- und Leistungsverhaltens die Abmessung der Nuten an die jeweilige Pumpe angepaßt werden. (Die anfangs beschriebene Lösung gemäß Anspruch 1 ist hierbei vergleichsweise unkritisch). Die Dämpfungsnut vermindert somit den Gradienten des Druckanstiegs in den Kolbenbohrungen bei Drehzahlen, die über der Abregeldrehzahl liegen. In diesem Drehzahlbereich sind die Kolbenbohrungen bei Bereichen der druckseitigen Steueröffnung teilweise mit Druckmedium und teilweise mit Gas bzw. mit Vakuum gefüllt. Die Dämpfungsnut dämpft die Rückströmung des Druckmediums von der Druckseite in die Kolbenbohrung, während dort durch die Einfahrbewegung der Kolben das Druckmedium-Gasgemisch vorkomprimiert wird. Dies bewirkt eine verbesserte Druckangleichung zwischen den Kolbenbohrungen und dem Druckanschluß, wodurch Druckpulsationen entscheidend vermindert werden. Dabei ist allerdings zu beachten, daß durch den relativ kleinen Querschnitt der Dämpfungsnut auch beachtliche Leistungsverluste bewirkt werden können, die dann nachteilig sind, wenn, wie beispielsweise bei Kraftfahrzeugen, das Antriebsaggregat (Kraftfahrzeugmotor) in seiner Leistungsfähigkeit beschränkt ist oder z.B. möglichst energiesparend ausgestaltet werden soll.Another solution to the object of the invention can be achieved in a simple manner by the combination of features resulting from claim 15. Here, the damping groove is connected in one piece, without a separate connection to the pressure connection, to the pressure control opening, but the damping groove formed in this way should have a much smaller cross section than the groove assigned to the pressure control opening, which is often referred to below as the pressure groove. This results in a much simpler construction of the pump, which is, however, subject to two restrictions. On the one hand, the dimensioning of the individual grooves depends on the pump output and the selected speed, from which the flow rate is no longer increased (regulation speed). Thus, the dimensions of the grooves must be adapted to the respective pump in order to optimize the noise and performance behavior. (The initially described solution according to claim 1 is comparatively uncritical). The damping groove thus reduces the gradient of the pressure rise in the piston bores at speeds above the Regulation speed. In this speed range, the piston bores in the areas of the pressure-side control opening are partly filled with pressure medium and partly with gas or with vacuum. The damping groove dampens the backflow of the pressure medium from the pressure side into the piston bore, while the pressure medium-gas mixture is pre-compressed there by the retracting movement of the pistons. This leads to an improved pressure adjustment between the piston bores and the pressure connection, which significantly reduces pressure pulsations. It should be noted, however, that the relatively small cross-section of the damping groove can also cause considerable power losses, which are disadvantageous if, for example in the case of motor vehicles, the drive unit (motor vehicle engine) is limited in its performance or, for example, is to be designed to be as energy-saving as possible .

In vorteilhafter Weiterbildung empfiehlt sich hierbei die Anwendung der Merkmalskombination nach Anspruch 16. Der Querschnitt der Dämpfungsnut ist vorzugsweise klein. Versuche haben gezeigt, daß je nach Pumpengröße und Einsatzgebiet ein Verhältnis des Querschnitts der Dämpfungsnut in mm gemessen zum Hubvolumen eines Kolbens in mm³ gemessen 1:1000 bis 1:1600, vorzugsweise 1:1300 zweckmäßig ist.In an advantageous further development, the use of the combination of features according to claim 16 is recommended. The cross section of the damping groove is preferably small. Experiments have shown that depending on the pump size and area of use, a ratio of the cross section of the damping groove in mm to the stroke volume of a piston in mm 3, measured 1: 1000 to 1: 1600, preferably 1: 1300, is expedient.

Die Dämpfungsnut erstreckt sich vorzugsweise über einen Winkelbereich von 30° bis 50° und kann als Dreiecksnut mit ca. 60° Öffnungswinkel ausgebildet sein. Die Auslegung von Länge und Querschnitt der Dämpfungsnut bildet dabei einen Kompromiß zwischen dem erhöhten Ausschiebewiderstand bei niedrigen Drehzahlen und der angestrebten Rückströmdämpfung bei höheren Drehzahlen. Dabei darf der Druck in den Kolbenbohrungen in keiner Betriebsphase den zulässigen Höchstwert überschreiten.The damping groove preferably extends over an angular range of 30 ° to 50 ° and can be designed as a triangular groove with an opening angle of approximately 60 °. The design of the length and cross section of the damping groove forms one Compromise between the increased push-out resistance at low speeds and the desired return flow damping at higher speeds. The pressure in the piston bores must not exceed the permissible maximum value in any operating phase.

Der Querschnitt der sich an die Dämpfungsnut anschließenden Drucknut wird erfindungsgemäß nur so groß gewählt, daß die Kolben das angesaugte Volumen ohne unzulässig hohen Druckanstieg in den Kolbenbohrungen gegen den Systemdruck am Druckanschluß ausschieben können. Hierbei hat es sich als vorteilhaft erwiesen, wenn der Querschnitt der Drucknut wenigstens doppelt so groß ist wie der Querschnitt der Dämpfungsnut. Es hat sich weiterhin als vorteilhaft erwiesen, wenn der Abstand vom Ende der Drucknut bis zum Einfahrtotpunkt gleich oder kleiner ist als der Radius der Kolbenfußbohrungen. Hierdurch werden Druckspitzen am Ende des Einfahrhubs der Kolben vermieden. Eine zusätzliche dämpfende Wirkung wird auf der Druckseite erfindungsgemäß dadurch erreicht, daß die Druckbohrung in das dem Steg benachbarte Ende der Drucknut mündet.The cross section of the pressure groove adjoining the damping groove is selected according to the invention only so large that the pistons can push out the suctioned volume without an impermissibly high pressure increase in the piston bores against the system pressure at the pressure connection. It has proven advantageous here if the cross section of the pressure groove is at least twice as large as the cross section of the damping groove. It has also proven to be advantageous if the distance from the end of the pressure groove to the entry dead center is equal to or less than the radius of the piston foot bores. This avoids pressure peaks at the end of the piston stroke. An additional damping effect is achieved on the pressure side according to the invention in that the pressure bore opens into the end of the pressure groove adjacent to the web.

In einer alternativen Ausgestaltung der Erfindung kann weiterhin vorgesehen sein, daß die Dämpfungsnut und die Drucknut durch eine einzige Nut mit stetig zunehmendem Querschnitt gebildet sind, die sich über einen Teilbereich oder die gesamt Länge der dem Druckanschluß zugeordneten Steueröffnung erstreckt.In an alternative embodiment of the invention it can further be provided that the damping groove and the pressure groove are formed by a single groove with a continuously increasing cross section, which extends over a partial area or over the entire length of the control opening assigned to the pressure connection.

Eine weitere Möglichkeit zur Lösung der gestellten Aufgabe ergibt sich durch Anwendung der Merkmale nach Anspruch 24. Diese Lösung kann parallel oder auch alternativ zu den Lösungen gemäß Anspruch 1 und Anspruch 15 angewendet werden. Durch die erfindungsgemäße Ausgestaltung der saugseitigen Steueröffnung wird bei einer Kolbenpumpe der angegebenen Art ein Förderstromcharakter erzielt, bei der unterhalb einer Abregeldrehzahl ein hoher Füllungsgrad erreicht wird, während oberhalb der Abregeldrehzahl die Fördermenge nahezu drehzahlunabhängig und konstant ist. Durch die Umgebungstemperatur, das Betriebsmedium und wechselnde Betriebsdrücke bedingte Einflüsse auf die Betriebseigenschaften der Pumpe sind gering. Das günstige Füllverhalten bei Drehzahlen unterhalb der Abregeldrehzahl ermöglicht, zumindest bei höherer Abregeldrehzahl, eine Einschränkung der Mittel, die das Ausfahren der Kolben unterstützen, wie z.B. Federn oder erhöhtes Kolbengewicht. Weiterhin lassen sich durch die Erfindung Druckpulsationen im Saugbereich der Pumpe auf ein Minimum reduzieren.Another possibility for solving the problem arises by using the features according to claim 24. This solution can be in parallel or alternatively to Solutions according to claim 1 and claim 15 are applied. The inventive design of the suction-side control opening achieves a delivery flow character in a piston pump of the type specified, in which a high degree of filling is achieved below a shutdown speed, while above the shutdown speed the delivery rate is almost independent of the speed and constant. The operating temperature of the pump is minimal due to the ambient temperature, the operating medium and changing operating pressures. The favorable filling behavior at speeds below the cut-off speed enables, at least at a higher cut-off speed, a restriction of the means that support the extension of the pistons, such as springs or increased piston weight. Furthermore, pressure pulsations in the suction area of the pump can be reduced to a minimum by the invention.

Günstig ist es, wenn das Verhältnis des Querschnitts der Drosselnut in mm gemessen zum Hubvolumen eines Kolbens in mm³ gemessen 1:700 bis 1:1200, insbesondere 1:1000 beträgt. Die Drosselnut kann erfindungsgemäß als Dreiecksnut mit einem Öffnungswinkel von ca. 60° ausgebildet sein. Die Drosselnut erlaubt, insbesondere bei kleinen Drehzahlen, eine definierte Teilbefüllung der Kolbenbohrungen im ersten Teil des Saughubs und verhindert dadurch ein zu starkes Druckgefälle bis zum Erreichen der Saugbohrung.It is advantageous if the ratio of the cross section of the throttle groove in mm to the stroke volume of a piston in mm 3 is 1: 700 to 1: 1200, in particular 1: 1000. According to the invention, the throttle groove can be designed as a triangular groove with an opening angle of approximately 60 °. The throttle groove allows, in particular at low speeds, a defined partial filling of the piston bores in the first part of the suction stroke and thereby prevents an excessive pressure drop until the suction bore is reached.

Um die Drosselquerschnitte besser variieren und die Abregeldrehzahl entsprechend den jeweiligen Anforderungen festlegen und um die Kolbenbohrungen von den Druckschwingungen im Saugkanal entkoppeln zu können, sind nach einem weiteren Vorschlag der Erfindung die dem Steuerkörper zugekehrten Enden der Kolbenbohrungen im Rotor abgesetzt und über Kolbenfußbohrungen von geringerem Durchmesser mit den Steueröffnungen verbindbar. Der Durchmesser der Kolbenfußbohrungen ist dabei so zu wählen, daß die Kolbenfußbohrungen die Wirkung einer Drosselblende haben. Vorzugsweise liegt das Verhältnis der Durchmesser von Kolbenfußbohrung und Kolbenbohrung zwischen 1:4 und 1:7.In order to vary the throttle cross-sections better and determine the speed limit according to the respective requirements and the piston bores from the To be able to decouple pressure vibrations in the suction channel, according to a further proposal of the invention, the ends of the piston bores facing the control body are offset in the rotor and can be connected to the control openings via piston base bores of smaller diameter. The diameter of the piston foot holes should be chosen so that the piston foot holes have the effect of a throttle orifice. The ratio of the diameter of the piston foot bore and piston bore is preferably between 1: 4 and 1: 7.

Die Erfindung wird nachfolgend anhand von Ausführungsbeispielen näher erläutert, die in der Zeichnung dargestellt sind. Es zeigen

Fig. 1
einen Axialschnitt einer Radialkolbenpumpe gemäß der Erfindung,
Fig. 2
einen Querschnitt durch en Rotor der Radialkolbenpumpe gemäß Fig. 1,
Fig. 3
einen Querschnitt in der Ebene der Steueröffnungen durch den Steuerzapfen der Radialkolbenpumpe gemäß Fig. 1 und
Fig. 4
eine Abwicklung der Steueröffnungen gemäß Fig. 3,
Fig. 5
die Abwicklung einer weiteren Ausführungsform der Erfindung mit getrennter Dämpfungsnut,
Fig. 6
in symbolischer Darstellung einen Schnitt durch den Steuerzapfen mit Aufzeichnung vorteilhafter Winkelerstreckung für die Ausführungsform nach Fig. 5 und
The invention is explained in more detail below on the basis of exemplary embodiments which are illustrated in the drawing. Show it
Fig. 1
an axial section of a radial piston pump according to the invention,
Fig. 2
2 shows a cross section through the rotor of the radial piston pump according to FIG. 1,
Fig. 3
a cross section in the plane of the control openings through the control pin of the radial piston pump according to FIG. 1 and
Fig. 4
a processing of the control openings according to FIG. 3,
Fig. 5
the handling of a further embodiment of the invention with a separate damping groove,
Fig. 6
a symbolic representation of a section through the control pin with recording advantageous angular extent for the embodiment of FIGS. 5 and

Die in Fig. 1 dargestellte Radialkolbenpumpe 1 weist ein im wesentlichen scheibenförmiges Pumpengehäuse 2 auf, mit einer durchgehenden Längsbohrung 3 und einer sich an diese anschließenden, zylindrischen Ausnehmung 4. In der Längsbohrung 3 ist ein Steuerzapfen 5, beispielsweise durch Einpressen, befestigt, der in die Ausnehmung 4 hineinragt. Auf dem Steuerzapfen 5 ist in der Ausnehmung 4 ein Rotor 6 drehbar gelagert, in dem mehrere, radial ausgerichtete Kolbenbohrungen 7 ausgebildet sind, in denen Kolben 8 gleiten. Die Kolben 8 stützen sich mit ihren aus den Kolbenbohrungen 7 herausragenden Enden an der Innenfläche eines Hubrings 9 ab, der mittels eines Wälzlagers exzentrisch zum Steuerzapfen 5 in der Ausnehmung 4 gelagert ist. Die radial inneren Enden der Kolbenbohrungen 7 sind im Rotor 6 abgesetzt und an Kolbenfußbohrungen 10 angeschlossen, die in die mittige Lagerbohrung 11 des Rotors 6 münden.The radial piston pump 1 shown in FIG. 1 has an essentially disk-shaped pump housing 2, with a continuous longitudinal bore 3 and a cylindrical recess 4 adjoining it. A control pin 5 is fastened in the longitudinal bore 3, for example by being pressed in the recess 4 protrudes. A rotor 6 is rotatably mounted on the control pin 5 in the recess 4, in which a plurality of radially aligned piston bores 7 are formed, in which pistons 8 slide. The pistons 8 are supported with their ends protruding from the piston bores 7 on the inner surface of a cam ring 9, which is mounted eccentrically to the control pin 5 in the recess 4 by means of a roller bearing. The radially inner ends of the piston bores 7 are offset in the rotor 6 and connected to piston base bores 10 which open into the central bearing bore 11 of the rotor 6.

Im Steuerzapfen 5 sind in der Ebene der Kolbenfußbohrungen 10 Steueröffnungen 12,13 ausgebildet, die bei Drehung des Rotors 6 nacheinander mit den Kolbenfußbohrungen 10 in Verbindung treten. Die Steueröffnung 12 befindet sich im Saugbereich der Kolben 8 und ist über eine Saugbohrung 14 an einen in Längsrichtung im Steuerzapfen 5 verlaufenden Saugkanal 15 angeschlossen, der mit einem Sauganschluß 16 in Verbindung steht. Die Steueröffnung 13 liegt im Druckbereich der Kolben 8 und ist über die Druckbohrung 17 an einen parallel zum Saugkanal 15 im Steuerzapfen 5 ausgebildeten Druckkanal 18 angeschlosen. Der Druckkanal 18 mündet in eine Ringnut 19, die mit einem Druckanschluß 20 in Verbindung steht. Der Rotor 6 wird über eine Kupplung 21 von einer Welle 22 angetrieben, die in einem die Ausnehmung 4 verschließenden Deckel 23 gelagert ist.In the control pin 5 10 control openings 12, 13 are formed in the plane of the piston base bores, which in turn connect to the piston base bores 10 when the rotor 6 rotates. The control opening 12 is located in the suction area of the pistons 8 and is connected via a suction bore 14 to a suction channel 15 which runs in the longitudinal direction in the control pin 5 and which is connected to a suction connection 16. The control opening 13 is in the Pressure range of the pistons 8 and is connected via the pressure bore 17 to a pressure channel 18 formed in the control pin 5 parallel to the suction channel 15. The pressure channel 18 opens into an annular groove 19 which is connected to a pressure connection 20. The rotor 6 is driven via a coupling 21 by a shaft 22 which is mounted in a cover 23 closing the recess 4.

Die Ausgestaltung der Steueröffnungen 12,13 im Steuerzapfen 5 ist aus den Fig. 3 und 4 ersichtlich. Durch die Auslegung der Strömungsquerschnitte der im Bereich des Saughubs der Kolben 8 liegenden Steueröffnung 12 wird das maximale Fördervolumen und der Füllungsgrad bestimmt sowie eine Dämpfung der Druckpulsationen auf der Saugseite erreicht. Die Steueröffnung 12 ist in drei unterschiedliche Bereiche gegliedert. Der erste Bereich beginnt in einem Abstand von etwa 30° in der durch Pfeil X gekennzeichneten Drehrichtung des Rotors 6 gesehen nach dem Einfahrtotpunkt ET, der sich aus dem geringsten Abstand zwischen dem Steuerzapfen 5 und dem Hubring 9 ergibt. Dieser Bereich ist als Drosselnut 24 von geringem Querschnitt ausgebildet. Die Drosselnut 24 hat die Form einer Dreiecksnut mit einem Öffnungswinkel von etwa 60°. Ihre Öffnungsbreite liegt vorzugsweise zwischen 0,7 und 1,2 mm. Vor allem bei kleinen Drehzahlen sorgt die Drosselnut 24 für eine definierte Teilbefüllung der Kolbenbohrungen 7 und sie verhindert eine zu starke Druckabsenkung vor dem Erreichen der Saugbohrung 14, wodurch Druckpulsationen vermindert werden. Die enge Drosselnut 24 mündet unmittelbar in die den zweiten Bereich der Steueröffnung 12 bildende Saugbohrung 13, die in einem Abstand von etwa 140° vom Einfahrtotpunkt ET angeordnet ist. An die Saugbohrung 14 schließt sich als dritter Bereich eine Füllnut 26 mit größerem Querschnitt an, die im Ausfahrtotpunkt AT endet. Vor allem durch die Lage der Saugbohrung 14 wird die wirksame Abregeldrehzahl der Radialkolbenpumpe 1 bestimmt, wobei die Füllnut 26 mit ihren vergleichsweise großen Querschnitt hauptsächlich den Füllungsgrad bei Drehzahlen verbessert, die unter der Abregeldrehzahl liegen. Durch eine kurze Füllnut 26 kann andererseits auf eine starke Drosselung des Saugstroms, in den Kolbenfußbohrungen 10 weitgehend verzichtet werden, wodurch eine Schmutzempfindlichkeit der Pumpe reduziert wird. Soll eine niedrige Abregeldrehzahl erreicht werden, so kann die Saugbohrung 14 unmittelbar vor dem Ausfahrtotpunkt AT angeordnet sein und auf eine Füllnut 26 verzichtet werden.The configuration of the control openings 12, 13 in the control pin 5 can be seen from FIGS. 3 and 4. By designing the flow cross-sections of the control opening 12 located in the area of the suction stroke of the pistons 8, the maximum delivery volume and the degree of filling are determined and damping of the pressure pulsations on the suction side is achieved. The control opening 12 is divided into three different areas. The first area begins at a distance of approximately 30 ° in the direction of rotation of the rotor 6, indicated by arrow X, after the entry dead center ET, which results from the smallest distance between the control pin 5 and the cam ring 9. This area is designed as a throttle groove 24 of small cross section. The throttle groove 24 has the shape of a triangular groove with an opening angle of approximately 60 °. Their opening width is preferably between 0.7 and 1.2 mm. Especially at low speeds, the throttle groove 24 ensures a defined partial filling of the piston bores 7 and prevents an excessive reduction in pressure before reaching the suction bore 14, thereby reducing pressure pulsations. The narrow throttle groove 24 opens directly into the suction bore 13 which forms the second region of the control opening 12 and which is arranged at a distance of approximately 140 ° from the entry dead center ET is. The suction hole 14 is followed as a third area by a filling groove 26 with a larger cross section, which ends at the exit emergency point AT. The effective regulating speed of the radial piston pump 1 is determined primarily by the position of the suction bore 14, the filling groove 26 with its comparatively large cross section mainly improving the degree of filling at speeds which are below the regulating speed. A short filling groove 26, on the other hand, can largely dispense with a strong throttling of the suction flow in the piston base bores 10, thereby reducing the sensitivity of the pump to dirt. If a low regulation speed is to be achieved, the suction bore 14 can be arranged immediately before the exit emergency point AT and a filling groove 26 can be dispensed with.

Die mit dem Druckanschluß 20 verbundene Steueröffnung 13 ist im Bereich des Ausfahrtotpunkts AT durch einen Steg 27 von der Füllnut 26 getrennt. Sie ist in zwei Bereiche, nämlich eine Dämpfungsnut 28 und eine Drucknut 29 unterteilt. Der Querschnitt der Dämpfungsnut 28 ist klein. Versuche haben gezeigt, daß Dreiecksnuten mit ca. 60° Öffnungswinkel und einer Öffnungsbreite zwischen 0,6 und 1,0 mm in vielen Anwendungsfällen ausreichend sind. Die Länge der Dämpfungsnut 28 beträgt bei der beschriebenen Ausführungsform 40°. Die Dämpfungsnut 28 hat in erster Linie die Aufgabe, den Gradienten des Druckanstiegs in den Kolbenbohrungen 7 bei Drehzahlen zu vermeiden, die über der Abregeldrehzahl liegen. Bei diesen Drehzahlen sind die Kolbenbohrungen 7 bei Öffnung der Verbindung zur Steueröffnung 13 teils mit Druckmedium und teils mit Gas gefüllt. Durch den in der Steueröffnung 13 herrschenden hohen Systemdruck strömt Druckmedium in die Kolbenbohrungen 7 zurück. wodurch diese gefüllt werden. Hierbei kommt es zu einem Druckabfall und kurz darauf durch die Verdrängungsarbeit der Kolben 8 erneut zu einem Druckanstieg auf das Niveau des Systemdrucks. Durch die Drosselwirkung der Dämpfungsnut 28 wird die Rückströmung in die Zylinderbohrung 7 gedämpft, während dort durch die Einfahrbewegung der Kolben 8 das Druckmedium komprimiert wird. Auf diese Weise wird eine vergleichsweise langsame Druckangleichung zwischen den Kolbenbohrungen 7 und dem Druckanschluß 20 erreicht, und die Druckpulsationen werden erheblich vermindert.The control opening 13 connected to the pressure connection 20 is separated from the filling groove 26 in the area of the exit emergency point AT by a web 27. It is divided into two areas, namely a damping groove 28 and a pressure groove 29. The cross section of the damping groove 28 is small. Tests have shown that triangular grooves with an opening angle of approx. 60 ° and an opening width between 0.6 and 1.0 mm are sufficient in many applications. The length of the damping groove 28 is 40 ° in the described embodiment. The damping groove 28 primarily has the task of avoiding the gradient of the pressure increase in the piston bores 7 at speeds that are above the cut-off speed. At these speeds, the piston bores 7 are partly filled with pressure medium and partly with gas when the connection to the control opening 13 is opened. By the prevailing in the control opening 13 high system pressure, pressure medium flows back into the piston bores 7. whereby these are filled. This results in a pressure drop and shortly thereafter, due to the displacement work of the pistons 8, the pressure rises again to the level of the system pressure. The backflow into the cylinder bore 7 is damped by the throttling action of the damping groove 28, while the pressure medium is compressed there by the retracting movement of the pistons 8. In this way, a comparatively slow pressure adjustment between the piston bores 7 and the pressure connection 20 is achieved, and the pressure pulsations are considerably reduced.

Zur Dämpfung von Druckpulsationen trägt weiterhin der zwar deutlich größere, jedoch ebenfalls auf einen Minimalwert begrenzte Querschnitt der sich an die Dämpfungsnut 28 anschließenden Drucknut 29 bei. Die Drucknut 29 erstreckt sich bis zum Einfahrtotpunkt ET und erlaubt dadurch ein Fördern der Kolben 8 bis zur Erreichung der maximalen Einfahrposition. Die Druckbohrung 17 mündet in das dem Einfahrtotpunkt ET benachbarte Ende der Drucknut 29 und trägt dadurch ebenfalls zur Dämpfungswirkung der Drucknut 29 bei.The damping of pressure pulsations is further contributed by the cross section of the pressure groove 29 adjoining the damping groove 28, which cross section is significantly larger, but also limited to a minimum value. The pressure groove 29 extends to the entry dead center ET and thereby allows the pistons 8 to be conveyed until the maximum entry position is reached. The pressure bore 17 opens into the end of the pressure groove 29 which is adjacent to the entry dead center ET and thereby also contributes to the damping effect of the pressure groove 29.

Fig. 5 zeigt eine der Fig. 4 entsprechende Abwicklung für eine bevorzugte Lösung gemäß den Ansprüchen 1 bis 14. Der wesentliche Unterschied gegenüber Fig. 4 besteht darin, daß auf eine Drosselnut 24 auf der Saugseite verzichtet wurde und auf der Druckseite die Dämpfungsnut 28 mit Rückschlagventil 32 (die grob gesehen der vorher beschriebenen Dämpfungsnut entspricht) an der Oberfläche des Steuerzapfens 5 nicht mehr in die Drucknut 29 übergeht, sondern von dieser durch einen Trennsteg 30 getrennt ist. Die Verbindung erfolgt über eine in Fig. 5 angedeutete, als Radialbohrung ausgebildete Drucksteuerbohrung 31, die als Linie 31 A in Fig. 5 symbolisch angedeutet ist. Die Drucksteuerbohrung 31 und damit die Dämpfungsnut 28 sind über ein Rückschlagventil 32 und einen Dämpfungskanal D mit dem Druckanschluß 20 verbunden. Die Drucksteueröffnung ist als Drucknut 29 ausgestaltet, die über die Druckbohrung 17 und einen Druckkanal 18, wie schon im Zusammenhang mit Fig. 1 beschrieben, mit dem Druckanschluß 20 in Verbindung steht.Fig. 5 shows a processing corresponding to Fig. 4 for a preferred solution according to claims 1 to 14. The main difference compared to Fig. 4 is that a throttle groove 24 has been omitted on the suction side and the damping groove 28 with on the pressure side Check valve 32 (which roughly corresponds to the previously described damping groove) on the surface of the control pin 5 no longer merges into the pressure groove 29, but is separated from it by a separating web 30. The connection is made via a pressure control bore 31 indicated in FIG. 5 and designed as a radial bore, which is symbolically indicated as line 31 A in FIG. 5. The pressure control bore 31 and thus the damping groove 28 are connected to the pressure connection 20 via a check valve 32 and a damping channel D. The pressure control opening is designed as a pressure groove 29, which is connected to the pressure connection 20 via the pressure bore 17 and a pressure channel 18, as already described in connection with FIG. 1.

Das Rückschlagventil 32 kann dabei in der Radialbohrung 31, in dem Dämpfungskanal D, aber auch am Ende des Dämpfungskanals D im Verbindungsbereich zu dem Druckanschluß 20 im Gehäuse angeordnet sein.The check valve 32 can be arranged in the radial bore 31, in the damping channel D, but also at the end of the damping channel D in the connection area to the pressure connection 20 in the housing.

Der Durchmesser der Drucksteuerbohrung 31 ist hier etwas kleiner dargestellt als der Durchmesser der Bohrungen 14 und 17. Die Drucksteuerbohrung kann aber den gleichen Durchmesser wie die genannten Bohrungen besitzen. Auch die Breite und der Durchmesser der in Fig. 5 gezeigten Radialnut ist weitgehend unkritisch und kann somit die gleiche Breite wie die Nuten 26 und 29 besitzen. Es ist auch möglich, zwischen die Nuten 28 und 29 oder anstatt der Nut 28 mehrere einzelne, in Linie hintereinander liegende Nuten vorzusehen, welche jeweils über ein eigenes Rückschlagventil mit dem Druckanschluß 20 verbunden sind. Es wird hierdurch eine verbesserte Leistung und eine verminderte Geräuschentwicklung erreicht. Gegenüber Fig. 4 wurde weiterhin noch auf die Drosselnut 24 verzichtet, da sich hierdurch eine erhebliche Vereinfachung der Ausgestaltung der Nuten ergibt, die nunmehr alle die gleiche Form haben. Die hierdurch bedingte Leistungsverminderung bzw. Geräuscherhöhung ist äußerst gering, so daß dies als vorteilhafte Lösung gegenüber Fig. 4 angesehen werden muß.The diameter of the pressure control bore 31 is shown here somewhat smaller than the diameter of the bores 14 and 17. However, the pressure control bore can have the same diameter as the bores mentioned. The width and the diameter of the radial groove shown in FIG. 5 is also largely uncritical and can therefore have the same width as the grooves 26 and 29. It is also possible to provide between the grooves 28 and 29 or instead of the groove 28 a plurality of individual grooves lying in line one behind the other, each of which is connected to the pressure connection 20 via its own check valve. This achieves improved performance and reduced noise. Compared to FIG. 4, the throttle groove 24 has also been omitted since this results in a considerable simplification of the design of the grooves, which now all have the same shape. The resulting reduction in performance or increase in noise is extremely low, so that this must be regarded as an advantageous solution compared to FIG. 4.

Die Lage der Saugbohrung 14 gegenüber der Füllnut 26 ist dabei weitgehend unkritisch, solange sich nur die Saugbohrung 14 im Bereich der Füllnut 26 befindet. Die Länge der Füllnut hängt weitgehend von der erwünschten Drosselwirkung ab, da der Füllungsgrad des jeweiligen Pumpenzylinders mit der Länge der Füllnut 26 zunimmt.The position of the suction bore 14 relative to the filling groove 26 is largely uncritical, as long as only the suction bore 14 is in the region of the filling groove 26. The length of the filling groove largely depends on the desired throttling effect, since the degree of filling of the respective pump cylinder increases with the length of the filling groove 26.

Vom Prinzip her ist aus Fig. 5 deutlich ersichtlich, daß die druckseitige Steueröffnung 13 gemäß Fig. 4 in zwei durch einen Trennsteg 30 getrennte Nuten unterteilt wurde, wobei die abgesetzte Drucksteuernut 28 zwar Druckmittel aus der Kolbenbohrung 7 (Fig. 1 und 2) übernimmt und damit erheblich zur Pumpenleistung beiträgt, während ein Rückströmen von Druckmittel über die Kanäle 18 und D von der einen höheren Druck aufweisenden Drucknut 29 in die Ducksteuernut 28 durch das Rückschlagventil 32 verhindert wird.In principle, it can be clearly seen from FIG. 5 that the pressure-side control opening 13 according to FIG. 4 has been divided into two grooves separated by a separating web 30, the offset pressure control groove 28 admitting pressure medium from the piston bore 7 (FIGS. 1 and 2) and thus contributes significantly to the pump performance, while a backflow of pressure medium via the channels 18 and D from the pressure groove 29 having a higher pressure into the pressure control groove 28 is prevented by the check valve 32.

Die in Fig. 5 gezeigte Winkellage der Nuten und Bohrungen ist nicht zwingend. Vielmehr hat sich auch eine Lageverteilung bewährt, wie sie in Fig. 6 dargestellt ist. Darin sind entsprechend Fig. 3 die senkrecht zur Betrachterebene laufenden Kanäle 15,18 und D gezeigt, wobei die einzelnen dargestellten Winkel folgende Werte haben: a=110°; b=70°; c,d=20°.The angular position of the grooves and bores shown in FIG. 5 is not mandatory. Rather, a position distribution, as shown in FIG. 6, has also proven itself. 3, the channels 15, 18 and D running perpendicular to the viewer plane are shown, the individual angles shown having the following values: a = 110 °; b = 70 °; c, d = 20 °.

Claims (29)

  1. A piston pump, in particular, a radial piston pump, comprising pistons (8) sliding within piston bores (7) in a rotor (6) and having one of their respective ends supported on a stroke-generating member (9), and a control member (5), such as a control pin (5), being in operative connection with the rotor (6), in which control pin are formed control orifices (12, 13) communicating with a suction channel (15) and a pressure channel (18), respectively, and being separated by webs, with the control orifices, upon rotation of the rotor (6), successively communicating with the piston bores (7), and wherein the piston pump includes a restrictor on the suction side and the control orifice (13) on the pressure side includes a pressure groove (29) and at least one pressure control groove (28), and the pressure connection (20) is in communication not only with the pressure groove (29) but also, through a directionally dependent resistance element (32) permeable toward the pressure connection (20), with at least one pressure control groove (28) which, in the direction of rotation of the rotor (6), is disposed ahead of the pressure groove (29) and, through separating webs (27,30), is substantially separated from the pressure groove and from the suction-side control orifice (12).
  2. A piston pump as claimed in claim 1, characterized in that the resistance member is a check valve (32).
  3. A piston pump as claimed in claims 1 or 2, characterized in that it is a pump provided with a control pin (5) and having its pressure chambers inwardly provided.
  4. A piston pump as claimed in any one of the preceding claims, characterized in that the pressure control orifice (29) is a pressure groove (29) extending in the direction of rotation of the piston (8).
  5. A piston pump as claimed in any one of the preceding claims, characterized in that a plurality of pressure control grooves (28) separated from one another and arranged in series in the direction of rotation of the rotor (6) are provided, which, through check valves (32) respectively associated firmly with the control grooves, are in communication with the pressure connection (20).
  6. A piston pump as claimed in any one of the preceding claims, characterized in that the length of the respective separating web (25, 27, 30), between the pressure control grooves (28) or groove (28), respectively, ahead of the pressure control orifice (29), and the pressure control orifice (29), in the direction of movement of the piston (8), is greater than the diameter of the piston stem bore.
  7. A piston pump as claimed in any one of the preceding claims, characterized in that the length of the pressure control groove (28) is shorter than the shortest distance of the edges of two successive piston stem bores (10).
  8. A piston pump as claimed in any one of the preceding claims, characterized in that the cross-section of the pressure control groove (28) is of rectangular configuration.
  9. A piston pump as claimed in any one of the preceding claims, characterized in that four pistons (8) equidistantly arranged in the rotor (6) are provided and that one pressure control groove (28) is provided which extends at an angle of approximately 70°, that the pressure control orifice (29), preferably, extends at an angle of about 45° and the separating webs (27, 30) between the pressure control orifice (29) and the pressure control groove (28) and between the latter and the suction control orifice (26), preferably, are 20°.
  10. A piston pump as claimed in any one of the preceding claims, characterized in that the angle between the compressed-mode dead center (AT) and the edge of the pressure control groove (28) directed toward the pressure orifice (29) is about 110°.
  11. A piston pump as claimed in any one of claims 3 to 10, characterized in that the rotor (6) is rotatably mounted on a control pin (5) wherein the pressure channel (18) and the suction channel (15) extend coaxially, and in that a damping channel (D) extends in parallel thereto and is in communication with the pressure control groove (28) through a damping bore (31), preferably of radial configuration.
  12. A piston pump as claimed in any one of the preceding claims, characterized in that the resistance member (32) is provided directly adjacent to the pressure control groove (28).
  13. A piston pump as claimed in claims 11 and 12, characterized in that the damping channel (D) serves to accommodate the check valve (32).
  14. A piston pump as claimed in claim 11, characterized in that the check valve (32) is located in the transitory area between the control pin (5) and the housing (2) surrounding the control pin.
  15. A piston pump, in particular, a radial piston pump, comprising pistons (8) sliding within piston bores (7) in a rotor (6) and having one of their respective ends supported on a stroke-generating member (9), and a control member (5), such as a control pin (5), being in operative connection with the rotor (6), in which control pin are formed control orifices (12, 13) communicating with a suction channel (15) and a pressure channel (18), respectively, and being separated by webs, with the control orifices, upon rotation of the rotor (6), successively communicating with the piston bores (7), characterized in that the piston pump includes a restrictor on the suction side, and a pressure groove (29) of the pressure-side control opening (13) associated with the pressure connection (20), against the direction of rotation of the rotor, passes into a neighbouring damping groove (28) of considerably smaller diameter.
  16. A piston pump as claimed in claim 15, characterized in that the pressure control orifice is formed by a pressure groove (29), in which terminates a pressure bore (17) leading to the pressure channel (20), and is of a substantially larger cross-section than the damping groove.
  17. A piston pump as claimed in claim 16, characterized in that the ratio of the cross-section of the damping groove (28), measured in mm, to the stroke volume of a piston (8), measured in mm³, is 1 : 1000 to 1 : 1600, preferably is 1 : 1300.
  18. A piston pump as claimed in claims 15 to 17, characterized in that the damping groove (28) is in the form of a triangular groove having an aperture angle of about 60°.
  19. A piston pump as claimed in any one of claims 15 to 18, characterized in that the damping groove (28) extends across an angular range of between 30° to 50°.
  20. A piston pump as claimed in any one of claims 15 to 19, characterized in that the cross-section of the pressure groove (29) is at least twice as large as the cross-section of the damping groove (28).
  21. A piston pump as claimed in any one of claims 15 to 20, characterized in that the distance from the end of the pressure groove (29) to the suction-mode dead center (ET) is equal to or smaller than the radius of the piston stem bores (10).
  22. A piston pump as claimed in any one of claims 15 to 21, characterized in that the pressure bore (17) terminates into the end of the pressure groove (29) adjacent to the web (25).
  23. A piston pump as claimed in any one of claims 15 to 22, characterized in that the damping groove (28) and the pressure groove (29) are formed by a groove (13) of gradually increasing cross-section, extending across a partial area or throughout the length of the control orifice associated with the pressure connection.
  24. A piston pump as claimed in any one of the preceding claims, characterized in that the control orifice (12) associated with the suction connection is a restriction groove (24) of small cross-section which, viewed in the direction of rotation, begins at an angle of 20° to 60°, with respect to the suction-mode dead center (ET), and includes a suction bore (14) joining the restriction groove (24) and disposed between an angle of 120°, with respect to the suction-mode dead center (ET), and the compressed-mode dead center (AT), and in that, with the suction bore (14) spaced from the compressed-mode dead center (AT), a filling groove (26) of larger cross-section extends from the suction bore (14) to the compressed-mode dead center (AT).
  25. A piston pump as claimed in claim 24, characterized in that the ratio of the cross-section of the restriction groove (24), measured in mm, to the stroke volume of a piston (8), measured in mm³, is 1 : 700 to 1 : 1200, preferably is 1 : 1000.
  26. A piston pump as claimed in claim 24 or 25, characterized in that the restriction groove (24) is in the form of a triangular groove having an aperture angle of about 60°.
  27. A piston pump as claimed in any one of claims 24 to 26, characterized in that the ends of the piston bores (7) facing the control member (5) are stepped within the rotor (6) and, through piston stem bores (10) of smaller diameter, are connectable with the control orifices (12, 13).
  28. A piston pump as claimed in claim 27, characterized in that sleeves are provided in the ends of the piston bores (7) facing the control member (5), which sleeves include a respective piston stem bore.
  29. A piston pump as claimed in claim 4 or claim 5, characterized in that the ratio of the diameters of piston stem bore (10) and piston bore (7) is between 1 : 4 and 1 : 7.
EP91918718A 1990-11-06 1991-11-05 Piston pump, especially a radial piston pump Expired - Lifetime EP0509077B1 (en)

Applications Claiming Priority (5)

Application Number Priority Date Filing Date Title
DE4035180 1990-11-06
DE4035180 1990-11-06
DE4135904A DE4135904A1 (en) 1990-11-06 1991-10-31 PISTON PUMP, PARTICULARLY RADIAL PISTON PUMP
DE4135904 1991-10-31
PCT/EP1991/002085 WO1992008051A1 (en) 1990-11-06 1991-11-05 Piston pump, especially a radial piston pump

Publications (2)

Publication Number Publication Date
EP0509077A1 EP0509077A1 (en) 1992-10-21
EP0509077B1 true EP0509077B1 (en) 1996-05-15

Family

ID=25898277

Family Applications (1)

Application Number Title Priority Date Filing Date
EP91918718A Expired - Lifetime EP0509077B1 (en) 1990-11-06 1991-11-05 Piston pump, especially a radial piston pump

Country Status (5)

Country Link
US (1) US5295797A (en)
EP (1) EP0509077B1 (en)
JP (1) JPH05503336A (en)
DE (2) DE4135904A1 (en)
WO (1) WO1992008051A1 (en)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102019110762A1 (en) * 2019-04-25 2020-10-29 Hoerbiger Automotive Komfortsysteme Gmbh Slot-controlled radial piston pump

Families Citing this family (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE4139611A1 (en) * 1991-11-30 1993-06-03 Zahnradfabrik Friedrichshafen TRANSMISSION WITH A DISPLACEMENT PUMP
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Also Published As

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EP0509077A1 (en) 1992-10-21
DE4135904A1 (en) 1992-05-21
JPH05503336A (en) 1993-06-03
WO1992008051A1 (en) 1992-05-14
DE59107817D1 (en) 1996-06-20
US5295797A (en) 1994-03-22

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