EP0211096B1 - Capacity control or torque limitation - Google Patents

Capacity control or torque limitation Download PDF

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Publication number
EP0211096B1
EP0211096B1 EP19850109781 EP85109781A EP0211096B1 EP 0211096 B1 EP0211096 B1 EP 0211096B1 EP 19850109781 EP19850109781 EP 19850109781 EP 85109781 A EP85109781 A EP 85109781A EP 0211096 B1 EP0211096 B1 EP 0211096B1
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EP
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Prior art keywords
control
pressure
piston
power
spring
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EP19850109781
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German (de)
French (fr)
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EP0211096A1 (en
Inventor
David A. Jacombs
Gerald Warren
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Vickers Systems GmbH
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Vickers Systems GmbH
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Priority to DE8585109781T priority Critical patent/DE3578196D1/en
Priority to EP19850109781 priority patent/EP0211096B1/en
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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
    • F04B1/12—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinder axes coaxial with, or parallel or inclined to, main shaft axis
    • F04B1/26—Control
    • F04B1/30—Control of machines or pumps with rotary cylinder blocks
    • F04B1/32—Control of machines or pumps with rotary cylinder blocks by varying the relative positions of a swash plate and a cylinder block
    • F04B1/324—Control of machines or pumps with rotary cylinder blocks by varying the relative positions of a swash plate and a cylinder block by changing the inclination of the swash plate

Definitions

  • the invention relates to a power-controlled hydraulic pump with a variable displacement (control pump).
  • Power-controlled control pumps are known and are used, among other things, when the drive motor (electric motor or internal combustion engine) should have a smaller output than the so-called basic performance value, which is calculated from the largest displacement at the highest permissible pressure. If one assumes a drive motor with an approximately constant output speed, then power-controlled actuating pumps also have the torque limitation feature.
  • the output of a control pump is calculated from the pump current multiplied by the working pressure.
  • the pump current and the working pressure have to be measured or corresponding values have to be formed and processed. While the working pressure can be used directly on the piston surfaces of a hydraulic control device, measuring the pump current requires a greater effort, which essentially determines the cost of the power control.
  • a hydraulic pump with variable displacement volume is already known, in which a control valve is provided with opposing first and second control piston surfaces, of which the first, larger control piston surface is acted upon by the working pressure and acts against the force of the valve spring, while the second , smaller control piston area is acted on by the control pressure.
  • the control valve is designed to switch and has a middle closed position, in which the control channel is neither connected to the working pressure nor to the outlet to the tank.
  • the control valve is used to regulate the displacement volume, but it does not appear from the document that this should be done with a view to a certain maximum power limitation.
  • the same control valve can also be used to regulate a double pump (US Pat. No. 3,093,081), the two pumps being different in size and successively adjusted in terms of their displacement volume. It does not appear from the writing that this should be done to limit performance.
  • control in particular for overload protection of the drive motor, can be achieved with significantly reduced technical outlay. It is particularly advantageous that the control valve does not have to be attached to the control pump itself.
  • the piston return device is designed so that each position of the actuating piston corresponds to a certain level of the control pressure.
  • This requirement is met for a spring or a spring assembly as a piston return device, regardless of whether there is direct action on the actuating piston or via an intermediate element, for example the swash plate of an axial piston pump.
  • the piston adjustment device should have a non-linear characteristic, with decreasing pump current e.g. increasing spring stiffness. The control pressure should therefore increase more than it corresponds to the decrease in the pump current.
  • a simple control valve with two piston surfaces is sufficient to carry out the invention, the one, smaller, piston surface being subjected to the working pressure and the other, larger, piston surface being subjected to the control pressure.
  • the control valve forms, so to speak, a product of the working pressure and the pump current, the size of the pump current being entered into the power control as a pressure signal in the form of the control pressure.
  • the control valve can be designed as a spring-loaded slide piston with two control edges. This not only represents an extremely economical solution to the performance problem, it is also extremely small in size (hardly larger than a conventional compensator with two slide pistons) and can be accommodated in the valve housing by other valves.
  • the axial piston pump 1 has a pump housing 2 with a housing cover 3, which is designed as a control plate, a swash plate 4, a cylinder drum 5 with a pump piston 6 mounted therein and an actuating piston 7 are also provided.
  • the cylinder drum 5 has splines for driving by a drive shaft 8, which is mounted in bearings in the housing 2 or the housing cover 3.
  • the pistons 6 are supported on the swash plate 4 via sliding shoes 9.
  • the swash plate 4 is pivotally mounted and is pushed via a return device 10 in the direction of the maximum inclined position against the possibility of the actuating piston 7 acting.
  • inlet and outlet channels 11, 12 are provided, each ending in kidney-shaped slots on the inside of the cover 3.
  • Hydraulic fluid is sucked in from a tank 13 via the inlet channel 11, moved to the pressure side by rotation of the cylinder drum 5 and displaced into the outlet channel 12.
  • a corresponding working pressure or system pressure builds up in the outlet channel.
  • Fig. 2 shows the return device 10 and the actuating piston 7 in an enlarged view.
  • the return device contains a spring assembly with two compression springs 15 and 16 which are fitted into one another and whose spring travel is of different sizes.
  • the spring 16 represents an additional spring which only comes into effect after the swash plate 4 has been pivoted to a certain extent and then exerts a force in addition to the spring 15.
  • a kinked spring characteristic of the overall arrangement is achieved in this way. Desirable is a spring characteristic that becomes progressively steeper in a force-displacement diagram.
  • the actuating piston 7 is guided in a cylinder 17 which is connected to a control valve 20 via a control line 18. This is connected to the pump outlet channel 12 via a pressure line 19. A drain line 14 connects to the tank 13.
  • the control valve 20 is accommodated in a housing 21 which can also accommodate further valves, for example a pressure relief valve 40.
  • the control valve 20 has a slide piston 22 which is guided in a bore 23 with the cross-sectional area A1 and has a control piston collar 24 with two control edges, which cooperate with corresponding edges between the valve bore 23 and a control bore 25.
  • the valve bore 23 is connected to the working pressure via the pressure line 19, while there is a connection to the drain line 14 to the right of the control piston collar 24.
  • this connection runs via a control channel 35 and a valve bore 33 of the pressure relief valve 40, which has a slide piston 32 and a control piston collar 44, which normally shuts off the system pressure supplied via 19 from the control channel 35.
  • a throttled leakage current flows from the pressure line 19 to the drain line 14, so that an average pressure is established in the control bore 25, which pressure is supplied to the actuating piston 7 as a control pressure via the control pressure line 18.
  • the control device operates as follows: When starting, the swash plate 4 is in its maximum swivel position, i.e. the axial piston pump is set to the maximum displacement. It is assumed that the drive motor, not shown, which engages the shaft 8 is brought to the desired speed. The maximum pump current Q1 (FIGS. 3 and 4) is then supplied to a hydraulic consumer, not shown, the properties of which determine the level of the working pressure which is established in the outlet line 12. It is assumed that a higher pressure than P1 (FIGS. 3 and 4) arises. An imbalance is reached on the spool 22, i.e.
  • the spool moves to the right in the drawing and releases a larger throttle cross section between its pressure supply side and the control bore 25, so that the control pressure in the control pressure line 18 increases and the actuating piston 7 is displaced.
  • the swash plate 4 is pivoted into a position for reduced pump current, which is represented by the falling characteristic curve in the QP diagram of FIGS. 3 and 4.
  • the increased control pressure is fed back to the piston 27, so that the valve 20 moves back into its equilibrium position, the filling of the cylinder 17 essentially not changing any further.
  • the pump then delivers a certain pump current at the requested working pressure. If a further increased working pressure is requested by the consumer, this leads to the control valve 20 responding again and the control pressure being increased further, with the result that the swash plate 4 is adjusted to a further reduced pump delivery flow.
  • the inclination of the QP characteristic in Fig. 3 or 4 is determined by the ratio of the cross-sectional areas of the control collar 24 to the control piston 27, i.e. determined according to the value A1 / A2.
  • the behavior of the control device can accordingly be determined by appropriate selection of the ratio A1: A2.
  • Another means of influencing the control characteristic is to influence the characteristic of the return device 10.
  • this is done by the spring 16, which may have a spring characteristic K2. It is assumed that the spring 16 comes into effect at the operating point Q2 P2 (FIG. 4), that is to say is acted upon as a result of the retracting swash plate 4.
  • the further swiveling of the swash plate 4 on a reduced pump flow is thus opposed by the two springs 15 and 16 with the characteristic curves K1 and K2. It therefore becomes a relative requires greater actuating piston force, which can only be generated by a relatively higher control pressure P c . According to the equation given, this leads to a higher working or system pressure P s compared to the case without the spring 16.
  • the "broken" characteristic according to FIG. 4 is therefore achieved.
  • the pressure relief valve 40 responds and increases the control pressure Pc to such an extent that the pump swivels into its zero stroke position. This is shown at P3 in FIG. 4.
  • the slope of the characteristic between P1 and P2 in FIG. 3 represents a rough approximation to a hyperbola.
  • the approximation to a hyperbola is better in the characteristic of the FIG.
  • the approach to a hyperbola can be driven further by appropriate design of the return device 10. In general, however, it is sufficient to achieve the approximation according to FIG. 3 or 4 to the ideal hyperbola when it comes to creating only overload protection for the drive motor.
  • the overload protection control valve 20 can be coupled to further valves in order to regulate the servomotor, because it only responds in the limit case for overload protection purposes.
  • Fig. 5 shows an axial piston pump 1a, in which two actuating pistons 7a, 7b are provided. While the - larger - control piston 7a is connected to the overload protection control valve 20 via the control line 18, the - smaller - control piston 7b is connected to a branch line 12b of the outlet line 12 and receives the system pressure.
  • the smaller actuating piston 7b forms, together with a return spring (not shown), the return device 10 of the swash plate 4.
  • an increasing control pressure in the cylinder 17 can be achieved as a function of the swivel position of the swash plate 4.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Control Of Positive-Displacement Pumps (AREA)
  • Reciprocating Pumps (AREA)

Description

Die Erfindung bezieht sich auf eine leistungsgeregelte Hydropumpe mit veränderlichem Verdrängungsvolumen (Stellpumpe).The invention relates to a power-controlled hydraulic pump with a variable displacement (control pump).

Leistungsgeregelte Stellpumpen sind bekannt und werden unter anderem dann eingesetzt, wenn der Antriebsmotor (Elektromotor oder Verbrennungsmotor) eine kleinere Leistung aufweisen soll, als es dem sogenannten Leistungseckwert entspricht, der sich aus dem größten Verdrängungsvolumen bei höchstem zulässigem Druck errechnet. Wenn man einen Antriebsmotor mit etwa konstanter Abtriebsdrehzahl voraussetzt, dann weisen leistungsgeregelte Stellpumpen gleichzeitig das Merkmal der Drehmomentbegrenzung auf.Power-controlled control pumps are known and are used, among other things, when the drive motor (electric motor or internal combustion engine) should have a smaller output than the so-called basic performance value, which is calculated from the largest displacement at the highest permissible pressure. If one assumes a drive motor with an approximately constant output speed, then power-controlled actuating pumps also have the torque limitation feature.

Die Leistung einer Stellpumpe errechnet sich bekanntlich aus dem Pumpenstrom, multipliziert mit dem Arbeitsdruck. Um zu einer Leistungsbegrenzung zu gelangen, müssen der Pumpenstrom und der Arbeitsdruck gemessen oder entsprechende Werte gebildet und verarbeitet werden. Während der Arbeitsdruck unmittelbar an Kolbenflächen einer hydraulischen Steuereinrichtung verwendet werden kann, macht das Messen des Pumpenstroms einen größeren Aufwand erforderlich, der die Kosten der Leistungsregelung im wesentlichen bestimmt.As is well known, the output of a control pump is calculated from the pump current multiplied by the working pressure. In order to limit the output, the pump current and the working pressure have to be measured or corresponding values have to be formed and processed. While the working pressure can be used directly on the piston surfaces of a hydraulic control device, measuring the pump current requires a greater effort, which essentially determines the cost of the power control.

Für diesen Zweck sind in der Praxis eine mechanische Rückführung der Stellung des Stellkolbens der Hydropumpe auf den Leistungsregler (Volvo in "Ölhydraulik und Pneumatik", 29 (1985), Nr. 1, S. 31 bis 34) oder eine Pilotsteuerung in Abhängigkeit vom Stellkolbenweg (Linde, Schrift H, Nr. 5.0015/2.80e) in Gebrauch. Bekannt (DE-A 3 232 695) ist auch, eine mit dem Stellkolben verbundene Meßfeder zu benutzen, um die Lage des Stellkolbens zu messen und als Kraftsignal dem Steuerventil zuzuführen. Das Steuerventil muß deshalb in den vorerwähnten Fällen in der Nähe des Stellkolbens angeordnet werden, was unerwünscht sein kann und immer noch einen beträchtlichen Aufwand erforderlich macht.For this purpose, a mechanical feedback of the position of the hydraulic pump's control piston to the power controller (Volvo in "Oil hydraulics and pneumatics", 29 (1985), No. 1, p. 31 to 34) or a pilot control depending on the control piston travel are in practice (Linde, font H, No. 5.0015 / 2.80e) in use. It is also known (DE-A 3 232 695) to use a measuring spring connected to the actuating piston in order to measure the position of the actuating piston and to supply it to the control valve as a force signal. The control valve must therefore be arranged in the aforementioned cases in the vicinity of the actuating piston, which can be undesirable and still requires considerable effort.

Aus US-A 2 921 560 ist bereits eine Hydropumpe mit veränderlichem Verdrängungsvolumen bekannt, bei der ein Steuerventil mit gegeneinanderwirkenden ersten und zweiten Steuerkolbenflächen versehen ist, wovon die erste, größere Steuerkolbenfläche vom Arbeitsdruck beaufschlagt ist und entgegen der Kraft der Ventilfeder wirkt, während die zweite, kleinere Steuerkolbenfläche vom Steuerdruck beaufschlagt ist. Das Steuerventil ist schaltend ausgebildet und weist eine mittlere geschlossene Stellung auf, in welcher der Steuerkanal weder mit dem Arbeitsdruck noch mit dem Ablauf zum Tank verbunden ist. Das Steuerventil dient zur Regelung des Verdrängungsvolumens, daß dies aber im Hinblick auf eine bestimmte maximale Leistungsbegrenzung erfolgen soll, geht aus der Schrift nicht hervor. Das gleiche Steuerventil kann auch zur Regelung einer Doppelpumpe verwendet werden (US-A 3 093 081), wobei die beiden Pumpen unterschiedlich groß und nacheinander hinsichtlich ihres Verdrängungsvolumens verstellt werden. Daß dies zur Leistungsbegrenzung erfolgen soll, geht aus der Schrift nicht hervor.From US-A 2 921 560 a hydraulic pump with variable displacement volume is already known, in which a control valve is provided with opposing first and second control piston surfaces, of which the first, larger control piston surface is acted upon by the working pressure and acts against the force of the valve spring, while the second , smaller control piston area is acted on by the control pressure. The control valve is designed to switch and has a middle closed position, in which the control channel is neither connected to the working pressure nor to the outlet to the tank. The control valve is used to regulate the displacement volume, but it does not appear from the document that this should be done with a view to a certain maximum power limitation. The same control valve can also be used to regulate a double pump (US Pat. No. 3,093,081), the two pumps being different in size and successively adjusted in terms of their displacement volume. It does not appear from the writing that this should be done to limit performance.

Mit der Erfindung ist eine Leistungsregelung, insbesondere für einen Überlastschutz des Antriebsmotors, mit wesentlich verringertem technischem Aufwand erzielbar. Von besonderem Vorteil ist, daß das Steuerventil nicht an der Stellpumpe selbst angebracht sein muß.With the invention, power control, in particular for overload protection of the drive motor, can be achieved with significantly reduced technical outlay. It is particularly advantageous that the control valve does not have to be attached to the control pump itself.

Bei der erfindungsgemäßen Leistungsregelung wird vorausgesetzt, daß die Kolbenrückholeinrichtung so ausgebildet ist, daß jeder Stellung des Stellkolbens eine bestimmte Höhe des Steuerdrucks entspricht. Diese Voraussetzung ist bei einer Feder oder einem Federpaket als Kolbenrückstelleinrichtung gegeben, und zwar unabhängig davon, ob eine unmittelbare Einwirkung auf den Stellkolben stattfindet oder über ein Zwischenglied, beispielsweise die Schrägscheibe einer Axialkolbenpumpe. Die Kolbenverstelleinrichtung sollte eine nichtlineare Kennlinie aufweisen, und zwar mit abnehmendem Pumpenstrom z.B. sich erhöhende Federsteifigkeit. Der Steuerdruck sollte also stärker zunehmen, als es der Abnahme des Pumpenstroms entspricht.In the power control according to the invention, it is assumed that the piston return device is designed so that each position of the actuating piston corresponds to a certain level of the control pressure. This requirement is met for a spring or a spring assembly as a piston return device, regardless of whether there is direct action on the actuating piston or via an intermediate element, for example the swash plate of an axial piston pump. The piston adjustment device should have a non-linear characteristic, with decreasing pump current e.g. increasing spring stiffness. The control pressure should therefore increase more than it corresponds to the decrease in the pump current.

Wenn die besprochenen Verhältnisse gegeben sind, genügt zur Durchführung der Erfindung ein einfaches Steuerventil mit zwei Kolbenflächen, wobei die eine, kleinere, Kolbenfläche mit dem Arbeitsdruck und die andere, größere, Kolbenfläche mit dem Steuerdruck beaufschlagt werden. Das Steuerventil bildet gewissermaßen ein Produkt aus Arbeitsdruck und Pumpenstrom, wobei die Größe des Pumpenstroms als Drucksignal in Form des Steuerdrucks in die Leistungsregelung eingegeben wird.If the conditions discussed are given, a simple control valve with two piston surfaces is sufficient to carry out the invention, the one, smaller, piston surface being subjected to the working pressure and the other, larger, piston surface being subjected to the control pressure. The control valve forms, so to speak, a product of the working pressure and the pump current, the size of the pump current being entered into the power control as a pressure signal in the form of the control pressure.

, Das Steuerventil kann als federbelasteter Schieberkolben mit zwei Steuerkanten ausgebildet sein. Dies stellt nicht nur eine außerordentlich ökonomische Lösung des Leistungsproblems dar, es ist auch die erforderliche Baugröße außerordentlich klein (kaum größer als ein konventioneller Kompensator mit zwei Schieberkolben) und kann im Ventilgehäuse von anderen Ventilen mit untergebracht werden., The control valve can be designed as a spring-loaded slide piston with two control edges. This not only represents an extremely economical solution to the performance problem, it is also extremely small in size (hardly larger than a conventional compensator with two slide pistons) and can be accommodated in the valve housing by other valves.

Die Erfindung wird anhand der Zeichnung beschrieben. Dabei zeigt:

  • Fig. 1 eine Axialkolbenpumpe mit Leistungsregelung im Querschnitt,
  • Fig. 2 eine vergrößerte Einzelheit aus Fig. 1,
  • Fig. 3 ein Pumpenstrom-Arbeitsdruck-Diagramm,
  • Fig. 4 ein weiteres Pumpenstrom-Arbeitsdruck-Diagramm Und
  • Fig. 5 eine abgeänderte Axialkolbenpumpe.
The invention is described with reference to the drawing. It shows:
  • 1 is an axial piston pump with power control in cross section,
  • 2 shows an enlarged detail from FIG. 1,
  • 3 is a pump current working pressure diagram,
  • Fig. 4 shows another pump current-working pressure diagram
  • Fig. 5 shows a modified axial piston pump.

Die Axialkolbenpumpe 1 weist ein Pumpengehäuse 2 mit einem Gehäusedeckel 3 auf, der als Steuerplatte ausgebildet ist, ferner sind eine Schrägscheibe 4, eine Zylindertrommel 5 mit darin angebrachtem Pumpenkolben 6 und ein Stellkolben 7 vorgesehen. Die Zylindertrommel 5 weist eine Keilverzahnung zum Antrieb durch eine Antriebswelle 8 auf, die in Lagern des Gehäuses 2 bzw. des Gehäusedeckels 3 gelagert ist. Die Kolben 6 stützen sich über Gleitschuhe 9 an der Schrägscheibe 4 ab. Die Schrägscheibe 4 ist schwenkbar gelagert und wird über eine Rückholeinrichtung 10 in Richtung maximaler Schrägstellung entgegen der Einwirkungsmöglichkeit des Stellkolbens 7 gedrängt. In dem Gehäusedeckel 3 sind Einlaß- und Auslaßkanäle 11, 12 vorgesehen, die jeweils in nierenförmigen Schlitzen an der Innenseite des Deckels 3 enden. Über den Einlaßkanal 11 wird Hydraulikflüssigkeit von einem Tank 13 angesaugt, durch Drehung der Zylindertrommel 5 auf die Druckseite geschafft und in den Auslaßkanal 12 hineinverdrängt. Je nach dem Schluckvermögen eines angeschlossenen hydraulischen Verbrauchers baut sich im Auslaßkanal ein entsprechender Arbeitsdruck oder Systemdruck auf.The axial piston pump 1 has a pump housing 2 with a housing cover 3, which is designed as a control plate, a swash plate 4, a cylinder drum 5 with a pump piston 6 mounted therein and an actuating piston 7 are also provided. The cylinder drum 5 has splines for driving by a drive shaft 8, which is mounted in bearings in the housing 2 or the housing cover 3. The pistons 6 are supported on the swash plate 4 via sliding shoes 9. The swash plate 4 is pivotally mounted and is pushed via a return device 10 in the direction of the maximum inclined position against the possibility of the actuating piston 7 acting. In the case Cover 3, inlet and outlet channels 11, 12 are provided, each ending in kidney-shaped slots on the inside of the cover 3. Hydraulic fluid is sucked in from a tank 13 via the inlet channel 11, moved to the pressure side by rotation of the cylinder drum 5 and displaced into the outlet channel 12. Depending on the swallowing capacity of a connected hydraulic consumer, a corresponding working pressure or system pressure builds up in the outlet channel.

Fig. 2 zeigt die Rückholeinrichtung 10 und den Stellkolben 7 in vergrößerter Darstellung. Die Rückholeinrichtung enthält ein Federpaket mit zwei Druckfedern 15 und 16, die ineinander gefügt sind und deren Federweg unterschiedlich groß ist. Die Feder 16 stellt eine Zusatzfeder dar, die erst nach einer gewissen Schwenkung der Schrägscheibe 4 zur Wirkung kommt und dann eine Kraft zusätzlich zur Feder 15 ausübt. Es wird so eine geknickte Federcharakteristik der Gesamtanordnung erzielt. Erstrebenswert ist eine Federcharakteristik, die in einem Kraft-Weg-Diagramm progressiv steiler wird.Fig. 2 shows the return device 10 and the actuating piston 7 in an enlarged view. The return device contains a spring assembly with two compression springs 15 and 16 which are fitted into one another and whose spring travel is of different sizes. The spring 16 represents an additional spring which only comes into effect after the swash plate 4 has been pivoted to a certain extent and then exerts a force in addition to the spring 15. A kinked spring characteristic of the overall arrangement is achieved in this way. Desirable is a spring characteristic that becomes progressively steeper in a force-displacement diagram.

Der Stellkolben 7 ist in einem Zylinder 17 geführt, der über eine Steuerleitung 18 mit einem Steuerventil 20 verbunden ist. Dieses steht über eine Druckleitung 19 mit dem Pumpenausfaßkanal 12 in Verbindung. Eine Abflußleitung 14 stellt eine Verbindung zum Tank 13 her.The actuating piston 7 is guided in a cylinder 17 which is connected to a control valve 20 via a control line 18. This is connected to the pump outlet channel 12 via a pressure line 19. A drain line 14 connects to the tank 13.

Das Steuerventil 20 ist in einem Gehäuse 21 untergebracht, welches noch weitere Ventile aufnehmen kann, beispielsweise ein Druckbegrenzungsventil 40. Das Steuerventil 20 besitzt einen Schieberkolben 22, der in einer Bohrung 23 mit der Querschnittsfläche A1 geführt ist und einen Steuerkolbenbund 24 mit zwei Steuerkanten aufweist, die mit entsprechenden Kanten zwischen der Ventilbohrung 23 und einer Steuerbohrung 25 zusammenarbeiten. In der Zeichnung links vom Steuerkolbenbund 24 steht die Ventilbohrung 23 über die Druckleitung 19 mit dem Arbeitsdruck in Verbindung, während rechts von dem Steuerkolbenbund 24 eine Verbindung zur Abflußleitung 14 gegeben ist. Diese Verbindung läuft im vorliegenden Fall über einen Steuerkanal 35 und eine Ventilbohrung 33 des Druckbegrenzungsventils 40, welches einen Schieberkolben 32 und einen Steuerkolbenbund 44 besitzt, der normalerweise den über 19 zugeführten Systemdruck vom Steuerkanal 35 absperrt. Wenn der Steuerbund 24 seine Mittelstellung zur Steuerbohrung 25 einnimmt, fließt ein gedrosselter Leckstrom von der Druckleitung 19 zur Abflußleitung 14, so daß sich in der Steuerbohrung 25 ein mittlerer Druck einstellt, der als Steuerdruck über die Steuerdruckleitung 18 dem Stellkolben 7 zugeführt wird.The control valve 20 is accommodated in a housing 21 which can also accommodate further valves, for example a pressure relief valve 40. The control valve 20 has a slide piston 22 which is guided in a bore 23 with the cross-sectional area A1 and has a control piston collar 24 with two control edges, which cooperate with corresponding edges between the valve bore 23 and a control bore 25. In the drawing to the left of the control piston collar 24, the valve bore 23 is connected to the working pressure via the pressure line 19, while there is a connection to the drain line 14 to the right of the control piston collar 24. In the present case, this connection runs via a control channel 35 and a valve bore 33 of the pressure relief valve 40, which has a slide piston 32 and a control piston collar 44, which normally shuts off the system pressure supplied via 19 from the control channel 35. When the control collar 24 assumes its central position to the control bore 25, a throttled leakage current flows from the pressure line 19 to the drain line 14, so that an average pressure is established in the control bore 25, which pressure is supplied to the actuating piston 7 as a control pressure via the control pressure line 18.

Fluchtend zur Ventilbohrung 23 ist eine Steuerkammerbohrung 26 vorgesehen, die zur Aufnahme eines Steuerkolbens 27 und einer Ventilfeder 28 dient. Der Steuerraum 26 ist über einen Verbindungskanal 29 mit der Steuerleitung 18 verbunden, so daß die eine Seite des Steuerkolbens 27 unter Steuerdruck steht. Die andere Seite des Kolbens 27 ist mit der Abflußleitung 14 verbunden. Der Steuerkolben 27 wird demnach entsprechend seiner Querschnittsfläche A2 und dem Steuerdruck Pc sowie durch die Kraft F1 der Ventilfeder 28 gegen den Schieberkolben 22 gedrängt , der seinerseits auf der .Fläche A1 unter dem Arbeits- oder Systemdruck Ps steht. In der Gleichgewichtsstellung gilt somit:

  • Ps. A1 = Po. A2 + F1 .
A control chamber bore 26 is provided in alignment with the valve bore 23 and serves to receive a control piston 27 and a valve spring 28. The control chamber 26 is connected to the control line 18 via a connecting channel 29, so that one side of the control piston 27 is under control pressure. The other side of the piston 27 is connected to the drain line 14. The control piston 27 is accordingly pressed according to its cross-sectional area A2 and the control pressure P c and by the force F1 of the valve spring 28 against the slide piston 22, which in turn is on the .Area A1 under the working or system pressure Ps. In the equilibrium position, the following therefore applies:
  • Ps. A1 = P o . A2 + F1.

Die Betriebsweise der Regeleinrichtung ist wie folgt: Beim Anlauf befindet sich die Schrägscheibe 4 in ihrer maximalen Schwenkstellung, d.h. die Axialkolbenpumpe ist auf maximales Verdrängungsvolumen eingestellt. Es sei angenommen, daß der nicht gezeichnete, an der Welle 8 angreifende Antriebsmotor auf Solldrehzahl gebracht wird. Es wird dann der maximale Pumpenstrom Q1 (Fig. 3 und 4) an einen nicht gezeigten hydraulischen Verbraucher geliefert, dessen Eigenschaften die Höhe des sich einstellenden Arbeitsdruckes in der Auslaßleitung 12 bestimmt. Es sei angenommen, daß sich ein höherer Druck als P1 (Fig. 3 und 4) einstellt. Dabei wird ein Ungleichgewicht am Schieberkolben 22 erreicht, d.h. der Schieberkolben wandert in der Zeichnung nach rechts und gibt einen größeren Drosselquerschnitt zwischen seiner Druckzuführseite und der Steurbohrung 25 frei, so daß der Steuerdruck in der Steuerdruckleitung 18 ansteigt und der Stellkolben 7 verschoben wird. Dadurch wird die Schrägscheibe 4 in eine Stellung für herabgesetzten Pumpenstrom verschwenkt, was durch die fallende Kennlinie im QP-Diagramm der Fig. 3 und 4 dargestellt ist. Der erhöhte Steuerdruck ist andererseits auf den Kolben 27 rückgeführt, so daß sich das Ventil 20 in seine Gleichgewichtslage rückbewegt, wobei im wesentlichen die Füllung des Zylinders 17 sich nicht weiter verändert. Die Pumpe liefert dann einen bestimmten Pumpenstrom bei dem angeforderten Arbeitsdruck. Wenn ein nochmals erhöhter Arbeitsdruck von dem Verbraucher angefordert wird, dann führt dies zum erneuten Ansprechen des Steuerventils 20 und einer weiteren Erhöhung des Steuerdrucks mit der Folge der Verstellung der Schrägscheibe 4 auf einen nochmals verringerten Pumpenförderstrom.The control device operates as follows: When starting, the swash plate 4 is in its maximum swivel position, i.e. the axial piston pump is set to the maximum displacement. It is assumed that the drive motor, not shown, which engages the shaft 8 is brought to the desired speed. The maximum pump current Q1 (FIGS. 3 and 4) is then supplied to a hydraulic consumer, not shown, the properties of which determine the level of the working pressure which is established in the outlet line 12. It is assumed that a higher pressure than P1 (FIGS. 3 and 4) arises. An imbalance is reached on the spool 22, i.e. the spool moves to the right in the drawing and releases a larger throttle cross section between its pressure supply side and the control bore 25, so that the control pressure in the control pressure line 18 increases and the actuating piston 7 is displaced. As a result, the swash plate 4 is pivoted into a position for reduced pump current, which is represented by the falling characteristic curve in the QP diagram of FIGS. 3 and 4. The increased control pressure, on the other hand, is fed back to the piston 27, so that the valve 20 moves back into its equilibrium position, the filling of the cylinder 17 essentially not changing any further. The pump then delivers a certain pump current at the requested working pressure. If a further increased working pressure is requested by the consumer, this leads to the control valve 20 responding again and the control pressure being increased further, with the result that the swash plate 4 is adjusted to a further reduced pump delivery flow.

Unter der Voraussetzung einer linearen Federkennlinie der Rückholeinrichtung 10 wird die Neigung der QP-Kennlinie in Fig. 3 oder 4 durch das Verhältnis der Querschnittsflächen des Steuerbundes 24 zum Steuerkolben 27, d.h. entsprechend dem Wert A1/A2 bestimmt. Je größer die Fläche A2 ist, umso flacher wird die QP-Kennlinie. Durch entsprechende Wahl des Verhältnisses A1:A2 kann man demnach das Verhalten der Regeleinrichtung bestimmen.Assuming a linear spring characteristic of the return device 10, the inclination of the QP characteristic in Fig. 3 or 4 is determined by the ratio of the cross-sectional areas of the control collar 24 to the control piston 27, i.e. determined according to the value A1 / A2. The larger the area A2, the flatter the QP characteristic. The behavior of the control device can accordingly be determined by appropriate selection of the ratio A1: A2.

Ein weiteres Mittel der Beeinflussung der Regelcharakteristik besteht darin, die Kennlinie der Rückholeinrichtung 10 zu beeinflussen. Im dargestellten Ausführungsbeispiel geschieht dies durch die Feder 16, der eine Federkennlinie K2 zukommen mag. Es sei angenommen, daß die Feder 16 im Betriebspunkt Q2 P2 (Fig. 4) zur Wirkung kommt, d.h. infolge der zurückweichenden Schrägscheibe 4 beaufschlagt wird. Der weiteren Schwenkung der Schrägscheibe 4 auf verringertem Pumpenstrom stehen somit die beiden Federn 15 und 16 mit den Kennlinien K1 und K2 entgegen. Es wird deshalb eine relativ größere Stellkolbenkraft benötigt, die nur durch einen relativ höheren Steuerdruck Pc erzeugt werden kann. Gemäß der angeführten Gleichung führt dies zu einem höheren Arbeits- oder Systemdruck Ps, verglichen mit dem Fall ohne die Feder 16. Es wird deshalb die "gebrochene" Charakteristik nach Fig. 4 erreicht.Another means of influencing the control characteristic is to influence the characteristic of the return device 10. In the illustrated embodiment, this is done by the spring 16, which may have a spring characteristic K2. It is assumed that the spring 16 comes into effect at the operating point Q2 P2 (FIG. 4), that is to say is acted upon as a result of the retracting swash plate 4. The further swiveling of the swash plate 4 on a reduced pump flow is thus opposed by the two springs 15 and 16 with the characteristic curves K1 and K2. It therefore becomes a relative requires greater actuating piston force, which can only be generated by a relatively higher control pressure P c . According to the equation given, this leads to a higher working or system pressure P s compared to the case without the spring 16. The "broken" characteristic according to FIG. 4 is therefore achieved.

Bei unzulässig hohen Drücken spricht das Druckbegrenzungsventil 40 an und erhöht den Steuerdruck Pc so weit, daß die Pumpe in ihre Nullhubstellung schwenkt. Dies ist bei P3 in Fig. 4 dargestellt.At impermissibly high pressures, the pressure relief valve 40 responds and increases the control pressure Pc to such an extent that the pump swivels into its zero stroke position. This is shown at P3 in FIG. 4.

Die Neigung der Kennlinie zwischen P1 und P2 in Fig. 3 stellt eine rohe Annäherung an eine Hyperbel dar. Die Annäherung an eine Hyperbel ist in der Kennlinie der Fig. besser . Durch entsprechende Gestaltung der Rückholeinrichtung 10 kann die Annäherung an eine Hyperbel noch weiter getrieben werden. Im allgemeinen ist es aber ausreichend, die Annäherung gemäß Fig. 3 oder 4 an die ideale Hyperbel zu erzielen, wenn es darum geht, nur einen Überlastschutz für den Antriebsmotor zu schaffen.The slope of the characteristic between P1 and P2 in FIG. 3 represents a rough approximation to a hyperbola. The approximation to a hyperbola is better in the characteristic of the FIG. The approach to a hyperbola can be driven further by appropriate design of the return device 10. In general, however, it is sufficient to achieve the approximation according to FIG. 3 or 4 to the ideal hyperbola when it comes to creating only overload protection for the drive motor.

Das Überlastschutz-Steuerventil 20 kann mit weiteren Ventilen gekoppelt sein, um zu einer Regelung des Stellmotors zu gelangen, denn es spricht nur im Grenzfall zu Überlastschutzzwecken an.The overload protection control valve 20 can be coupled to further valves in order to regulate the servomotor, because it only responds in the limit case for overload protection purposes.

Fig. 5 zeigt eine Axialkolbenpumpe 1a, bei der zwei Stellkolben 7a, 7b vorgesehen sind. Während der - größere - Stellkolben 7a über die Steuerleitung 18 mit dem Überlastschutz-Steuerventil 20 verbunden ist, ist der - kleinere - Stellkolben 7 b mit einer Abzweigleitung 12b der Auslaßleitung 12 verbunden und empfängt den Systemdruck. Der kleinere Stellkolben 7b bildet, zusammen mit einer nicht dargestellten Rückholfeder, die Rückholeinrichtung 10 der Schrägscheibe 4. Auch hierbei läßt sich ein steigender Steuerdruck im Zylinder 17 in Abhängigkeit von der Schwenkstellung der Schrägscheibe 4 erzielen.Fig. 5 shows an axial piston pump 1a, in which two actuating pistons 7a, 7b are provided. While the - larger - control piston 7a is connected to the overload protection control valve 20 via the control line 18, the - smaller - control piston 7b is connected to a branch line 12b of the outlet line 12 and receives the system pressure. The smaller actuating piston 7b forms, together with a return spring (not shown), the return device 10 of the swash plate 4. Here too, an increasing control pressure in the cylinder 17 can be achieved as a function of the swivel position of the swash plate 4.

Claims (4)

  1. ' 1. Power-controlled hydraulic pump having variable displacement volume and the following features:
    an adjusting member (4, 7, 10) having a servo piston (7) is provided for adjusting the displacement volume (Q);
    the servo piston (7) is displaceable by control pressure (Pc) against the effect of a return means (10);
    the return means (10) is designed in such a way that a certain level of the control pressure (Pc) corresponds to each position of the servo piston (7);
    a control valve (20) controls the level of the control pressure (Pc) and to this end has a control member (22) having a first control piston area (A1) which is acted upon by the working pressure (Ps) against the force (F1) of a valve spring (28), and a second control piston area (A2) which is acted upon by the control pressure (Pc) against the effect of the first control piston area (A2), characterized in that the return means (10) contains one or more springs (15,16) in order to produce a progressively increasing total return force during adjustment for smaller displacement volumes (Q), and in that the second control piston area (A2) is larger than the first control piston area (A1).
  2. 2. Power-controlled pump according to Claim 1, characterized in that the springs (15, 16), in their spring travel, come into effect one after the other.
  3. 3. Power-controlled pump according to Claim 2, characterized in that the ratio of the sizes of the first and second control piston area (A1 : A2) to one another as well as the spring characteristics (K1, K2) and the spring travel of the springs (15, 16) of the return means (10) are selected in such a way that an approximation to a constant power limitation in accordance with a hyperbola in a QP-diagram is obtained.
  4. 4. Power-controlled pump according to one of Claims 1 to 3, characterized in that the control member (22) of the control valve (20) has two control edges on a spool collar (24) which interact with a control bore (25) and, in the centre position of the spool collar (24), allow a restricted leakage flow to pass from a pressure line (19) to a discharge line (14).
EP19850109781 1985-08-03 1985-08-03 Capacity control or torque limitation Expired - Lifetime EP0211096B1 (en)

Priority Applications (2)

Application Number Priority Date Filing Date Title
DE8585109781T DE3578196D1 (en) 1985-08-03 1985-08-03 PERFORMANCE CONTROL OR TORQUE LIMIT.
EP19850109781 EP0211096B1 (en) 1985-08-03 1985-08-03 Capacity control or torque limitation

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP19850109781 EP0211096B1 (en) 1985-08-03 1985-08-03 Capacity control or torque limitation

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EP0211096A1 EP0211096A1 (en) 1987-02-25
EP0211096B1 true EP0211096B1 (en) 1990-06-13

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DE (1) DE3578196D1 (en)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102021201409A1 (en) 2021-02-15 2022-08-18 Robert Bosch Gesellschaft mit beschränkter Haftung Displacement machine with displacement volume measuring device

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Publication number Priority date Publication date Assignee Title
DE29922559U1 (en) 1999-12-22 2000-03-02 Thueringer Bauholding Gmbh Tent scaffolding
US10570893B2 (en) * 2015-05-29 2020-02-25 Kanzaki Kokyukoki Mfg. Co., Ltd. Hydraulic pump and detachable servo unit

Family Cites Families (3)

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Publication number Priority date Publication date Assignee Title
US2921560A (en) * 1957-09-23 1960-01-19 New York Air Brake Co Engine control
US3093081A (en) * 1959-01-29 1963-06-11 New York Air Brake Co Pumping device
SE345501B (en) * 1967-01-16 1972-05-29 Wytwornia Sprzetu Komunikacyjn

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102021201409A1 (en) 2021-02-15 2022-08-18 Robert Bosch Gesellschaft mit beschränkter Haftung Displacement machine with displacement volume measuring device

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DE3578196D1 (en) 1990-07-19
EP0211096A1 (en) 1987-02-25

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