EP0509077A1 - Pompe a pistons, notamment pompe a pistons radiaux. - Google Patents

Pompe a pistons, notamment pompe a pistons radiaux.

Info

Publication number
EP0509077A1
EP0509077A1 EP91918718A EP91918718A EP0509077A1 EP 0509077 A1 EP0509077 A1 EP 0509077A1 EP 91918718 A EP91918718 A EP 91918718A EP 91918718 A EP91918718 A EP 91918718A EP 0509077 A1 EP0509077 A1 EP 0509077A1
Authority
EP
European Patent Office
Prior art keywords
pressure
groove
piston pump
piston
pump according
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.)
Granted
Application number
EP91918718A
Other languages
German (de)
English (en)
Other versions
EP0509077B1 (fr
Inventor
Manfred Kahrs
Gerhard Kunz
Franz Fleck
Hermann Schoellhorn
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
Alfred Teves GmbH
Priority date (The priority date 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 date listed.)
Filing date
Publication date
Application filed by Alfred Teves GmbH filed Critical Alfred Teves GmbH
Publication of EP0509077A1 publication Critical patent/EP0509077A1/fr
Application granted granted Critical
Publication of EP0509077B1 publication Critical patent/EP0509077B1/fr
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

Links

Classifications

    • 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

  • Piston pump in particular radial piston pump
  • 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.
  • 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 pump 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.
  • the pump medium on the suction side with increasing speed is opposed by an increasing resistance, which means that from a certain speed the
  • 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 - 2 -
  • 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 is known from DE 37 00 573 AI.
  • 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 from the high-pressure control slot to a pressure chamber formed in the control pin, from which a channel extends, 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.
  • the invention is therefore based on a piston pump - 5 -
  • 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.
  • one solution is to pump the incompressible pressure medium to the pressure connection via a pressure control groove (hereinafter often referred to as the damping groove) and preferably a check valve 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 pressure control groove (or damping groove) used on the pressure side can also be successively connected several times in that, in the direction of movement of the rotor, a plurality of pressure control grooves separated by separating webs are each connected to the pressure connection via a check valve.
  • the pressure control grooves can also be operated individually using check valves be connected to the pressure channel belonging to the pressure bore.
  • the invention provides a particularly simple structure for a pump with the features resulting from claim 3.
  • a pump is 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.
  • the shape of the pressure control 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.
  • 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 that are above the cut-off 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, as in the case of motor vehicles, for example, 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 .
  • Cross section of the damping groove is preferably small.
  • Area of use is a ratio of the cross section of the
  • 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 a 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 sucked-in 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.
  • damping groove and the pressure groove by a single groove with continuously increasing Cross-section are formed, which extends over a partial area or the entire length of the control opening assigned to the pressure connection.
  • This solution can be used in parallel or alternatively to the solutions according to claim 1 and claim 15.
  • 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 allows, at least at a higher cut-off speed, a restriction of the means that support the extension of the pistons, e.g. Springs or increased piston weight.
  • pressure pulsations in the suction area of the pump can be reduced to a minimum by the invention.
  • the throttle groove can according to the invention as
  • Triangular groove with an opening angle of approximately 60 °.
  • the throttle groove allows, especially with 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 with piston foot bores of smaller diameter connectable to the control openings.
  • 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 invention is based on
  • FIG. 2 shows a cross section through the rotor of the radial piston pump according to FIG. 1,
  • Fig. 3 shows a cross section in the plane of the control openings through the control pin of the radial piston pump according to - 13 -
  • FIG. 6 in a symbolic representation a section through the control pin with recording advantageous angular extent for the embodiment of FIGS. 5 and
  • 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. - 14 -
  • 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 lies in the pressure region 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 in the recess
  • 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 about 30 c 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 °. Your opening width is preferably between 0.7 and 1.2 mm.
  • 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 is arranged at a distance of approximately 140 ° from the entry dead center ET.
  • 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 14 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 that are below the regulating speed.
  • 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.
  • 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. Due to the high system pressure prevailing in the control opening 13, pressure medium flows back into the piston bores 7, as a result of which 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 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 - 17th
  • FIG. 5 shows a development 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 on the suction side has been dispensed with and the pressure control groove 28 on the pressure side Check valve (which roughly corresponds to the previously described damping groove) at 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 radial bore 31 indicated in FIG. 5, which is symbolically indicated as line 31 A in FIG. 5.
  • the radial 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 18, in the pressure channel D, but also at the end of the pressure channel D in the connection area to the pressure connection 20 in the housing.
  • the diameter of the radial bore 31 is shown here somewhat smaller than the diameter of the bores 14 and 17.
  • the radial 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 dispensed with, since this considerably simplifies 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 viewed as an advantageous solution compared to FIG. 4.
  • the position of the suction bore 14 with respect 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 from the groove 29 via the channels 18, D from the groove 29 having a higher pressure into the Pressure control groove 28 is prevented by the check valve 32.

Abstract

Afin d'obtenir une pompe peu bruyante à pistons à régulation de l'aspiration, notamment pour véhicules à moteur, ayant de préférence un téton de commande d'admission interne, avec un courant de refoulement constant dans une large plage de régimes et de pertes en puissance aussi réduites que possible, plusieurs solutions sont décrites. La première solution consiste à subdiviser la fente de commande (13) en plusieurs rainures (27, 28) du côté de la pression, reliées au moins en partie par des soupapes de non retour (32) au raccord de refoulement (20). Une autre solution consiste à adapter la forme de la fente de commande du côté de la pression au mode optimal de fonctionnement; son extrémité d'admission est relativement étroite afin de réduire le bruit (réduction du bruit à des régimes élevés) et son extrémité de sortie a une largeur suffisante pour assurer la puissance requise. Une troisième solution consiste à prévoir une fente (12) du côté d'aspiration ayant une forme analogue et correspondant à celle de la deuxième solution, qui influe en outre par sa longueur sur le degré de remplissage de l'alésage du piston.
EP91918718A 1990-11-06 1991-11-05 Pompe a pistons, notamment pompe a pistons radiaux Expired - Lifetime EP0509077B1 (fr)

Applications Claiming Priority (5)

Application Number Priority Date Filing Date Title
DE4035180 1990-11-06
DE4035180 1990-11-06
DE4135904A DE4135904A1 (de) 1990-11-06 1991-10-31 Kolbenpumpe, insbesondere radialkolbenpumpe
DE4135904 1991-10-31
PCT/EP1991/002085 WO1992008051A1 (fr) 1990-11-06 1991-11-05 Pompe a pistons, notamment pompe a pistons radiaux

Publications (2)

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

Family

ID=25898277

Family Applications (1)

Application Number Title Priority Date Filing Date
EP91918718A Expired - Lifetime EP0509077B1 (fr) 1990-11-06 1991-11-05 Pompe a pistons, notamment pompe a pistons radiaux

Country Status (5)

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

Families Citing this family (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE4139611A1 (de) * 1991-11-30 1993-06-03 Zahnradfabrik Friedrichshafen Getriebe mit einer verdraengerpumpe
DE4239202A1 (en) * 1992-01-25 1993-07-29 Naumann Ulrich Dr Ing Valveless pump or compressor - has control journal with rotating body moving over inlet and outlet passages
NL9301011A (nl) * 1993-06-11 1995-01-02 Applied Power Inc Radiale-plunjerpomp.
DE19504220A1 (de) * 1995-02-09 1996-08-14 Bosch Gmbh Robert Verstellbare hydrostatische Pumpe
DE19521574A1 (de) * 1995-06-14 1996-12-19 Rexroth Mannesmann Gmbh Hydrostatische Maschine
US5975864A (en) * 1998-02-19 1999-11-02 Jetech, Inc. Pump with self-reciprocating pistons
US10683854B2 (en) * 2015-05-21 2020-06-16 Eaton Intelligent Power Limited Radial piston device with reduced pressure drop
DE102019110762A1 (de) * 2019-04-25 2020-10-29 Hoerbiger Automotive Komfortsysteme Gmbh Schlitzgesteuerte Radialkolbenpumpe

Family Cites Families (18)

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Publication number Priority date Publication date Assignee Title
US2371078A (en) * 1942-09-14 1945-03-06 Hydraulic Dev Corp Radial pump with trunnion mounting of shift ring
GB570252A (en) * 1943-07-21 1945-06-28 Rudolph William Glasner Improvements in or relating to hydraulic pumps and motors
US2529309A (en) * 1946-03-11 1950-11-07 Hpm Dev Corp Fluid operable apparatus
DE1528613A1 (de) * 1966-03-28 1970-10-22 Ind Karl Marx Stadt Veb Hydraulische Kolbenpumpe
DE2061960B2 (de) * 1970-12-16 1974-01-17 Fichtel & Sachs Ag, 8720 Schweinfurt Radialkolbenpumpe mit Drosseleinrichtung zur Begrenzung des Fördervolumens
DE2061961C3 (de) * 1970-12-16 1974-10-17 Dieter Dr.-Ing. 8720 Schweinfurt Lutz Radialkolbenpumpe mit Saugsteuerung am Exzenter für reversiblen Betrieb
DE2251792A1 (de) * 1972-10-21 1974-04-25 Bosch Gmbh Robert Radialkolbenmotor
SU513167A1 (ru) * 1973-11-20 1976-05-05 Ордена Трудового Красного Знамени Институт Горного Дела Имени А.А.Скочинского Объемный гидродвигатель
DE2601970A1 (de) * 1976-01-20 1977-07-21 Linde Ag Steuerspiegel einer hydrostatischen maschine
GB1567100A (en) * 1977-10-03 1980-05-08 Caterpillar Tractor Co Flow control assembly for multi-piston pumps
DE2828022A1 (de) * 1978-06-26 1980-01-03 Danfoss As Rotationskolbenpumpe, insbesondere radialkolbenpumpe
DE2946746A1 (de) * 1979-11-20 1981-05-27 Fichtel & Sachs Ag, 8720 Schweinfurt Radialkolbenpumpe mit druckabhaengiger ansaugdrosselung
JPS60182365A (ja) * 1984-02-28 1985-09-17 Nippon Denso Co Ltd ラジアルプランジヤポンプ
DE3628769A1 (de) * 1986-08-25 1988-03-10 Teves Gmbh Alfred Radialkolbenpumpe
DD255966A1 (de) * 1986-11-10 1988-04-20 Karl Marx Stadt Ind Werke Steuerspiegel fuer hydrostatische kolbenpumpen
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US5049039A (en) * 1988-06-29 1991-09-17 Pneumotor, Inc. Radial piston and cylinder compressed gas motor

Non-Patent Citations (1)

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Title
See references of WO9208051A1 *

Also Published As

Publication number Publication date
DE4135904A1 (de) 1992-05-21
WO1992008051A1 (fr) 1992-05-14
DE59107817D1 (de) 1996-06-20
EP0509077B1 (fr) 1996-05-15
US5295797A (en) 1994-03-22
JPH05503336A (ja) 1993-06-03

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