EP1831573A1

EP1831573A1 - Hydraulic control system

Info

Publication number: EP1831573A1
Application number: EP05850157A
Authority: EP
Inventors: Wolfgang Kauss
Original assignee: Bosch Rexroth AG
Current assignee: Bosch Rexroth AG
Priority date: 2004-12-21
Filing date: 2005-12-14
Publication date: 2007-09-12
Anticipated expiration: 2025-12-14
Also published as: WO2006066548A1; JP2008524529A; DE102004061555A1; US20080053081A1; EP1831573B1; ATE554291T1; JP4801091B2; US7946114B2

Abstract

The invention relates to a hydraulic control system for providing at least one hydraulic consumer with a pressure medium. Said system comprises an LS pump system and a metering port for adjusting the pressure medium volume flow rate towards the consumer. The LS line is connected to a pressure medium sink via a current regulator. The invention is characterized in that the current regulator can be adjusted depending on the pump rate in order to modify the pressure drop at the metering port.

Description

Description Hydraulic control unit

The invention relates to a hydraulic control arrangement for pressure medium supply at least one hydraulic consumer.

Such control arrangements described, for example, in DE 199 30 618 A1 have a variable displacement pump or constant displacement pump with bypass pressure compensator, which in each case are controlled as a function of the highest load pressure of the actuated hydraulic consumers such that the pump pressure is above the highest load pressure by a specific pressure difference. The hydraulic consumers, the pressure medium flows through adjustable metering orifices, which are arranged between an outgoing of the variable displacement supply line and the hydraulic consumers. It is achieved by the pressure compensators associated with the metering orifices that a certain pressure difference across the metering orifices is created independently of the load pressures of the hydraulic consumers so that the quantity of pressure medium flowing to the respective consumer depends only on the opening cross section of the respective metering orifice. The pump regulator of the variable displacement pump or the bypass pressure compensator of the fixed displacement pump is adjusted in such a way that it delivers the required quantity of pressure medium - this is called demand flow control.

In the case of LUDV systems, the individual pressure compensator assigned to the metering orifice in the closing direction is usually controlled by the highest load pressure of the hydraulic consumers and, in the opening direction, by the pressure downstream of the metering orifice. If, during a simultaneous operation of several hydraulic consumers, the metering dazzle opened so far that the pressure medium supplied by the pump is smaller than the demand quantity, the pressure fluid quantities flowing to the individual hydraulic consumers are proportionally reduced independently of the respective load pressure of the hydraulic consumers. Such a LUDV control represents a special case of an LS control. This is referred to as a mere LS control if the individual pressure compensator is acted upon in the closing direction by the pressure before the metering orifice and in the opening direction by the pressure after the metering orifice, this pressure, when the metering orifice of the pressure compensator is downstream, corresponding to the individual load pressure, at the same time Operation of a plurality of hydraulic consumers and not enough of the variable pressure fluid supplied only reduces the pressure of the highest pressure hydraulic consumers flowing pressure medium quantity. In the known solution, an LS line carrying the highest load pressure is connected to the tank via a flow control valve.

Such hydraulic control arrangements are used, for example, to supply the consumers of construction machines, for example a slewing gear, a jib, a spoon or a dipper stick of a mobile implement. In such implements, the pump is often driven by an internal combustion engine, this pump is assigned to all consumers. The size of the pump is designed according to the available engine power, with the individual movements of the consumers to a large extent being coordinated with one another in view of good controllability. In some cases, for example, when operating the slewing gear, the entire pump flow rate is required for a single motion. Accordingly, the maximum opening cross section of the metering orifice must be designed for this quantity of pressure medium. If the opening cross-section of the Zutnessblende is designed for the entire amount of pump at maximum engine speed, with reduced or minimum engine speed, the control range of the slide is not fully utilized. For this amount of pressure medium, the metering orifice must be opened only to a part of the maximum opening cross-section, so that only a partial stroke of the metering orifice is available for controlling this amount of pressure medium. Accordingly, the resolution of the metering orifice is comparatively low, so that the accuracy of the consumer movement at low speed often does not meet the requirements.

In the documents DE 100 06 659 Al, US 5, 085, 051, US 5, 481, 875 and US 5, 226, 800 LS-control arrangements are described in which the consumers are supplied via a variable displacement pump with pressure medium whose delivery volume in Dependent on the highest load pressure is adjustable. In these known solutions, the pump governor of the variable displacement pump can be charged electronically with an additional force, so that the pressure drop controlled by the pump regulator is reduced and the pressure medium volume flow delivered by the variable displacement pump is correspondingly changed. However, this electronic adjustment by means of an electromagnet is only suitable for the pump regulator of a variable displacement pump, since the larger forces required by the bypass pressure compensator of a fixed displacement pump can hardly be controlled by an electromagnet. Furthermore, a complex Elelektronik is required for the control of the required proportional solenoid, which is not present in many machines.

Furthermore, in LS systems, the motor speed of the pump drive has little effect on the speed of movement of the loads, since the volume current zυ. the consumer is limited by the metering orifice of the associated spool at the regulated by the pressure compensator pressure drop.

The invention has for its object to provide a hydraulic control arrangement, with which a sufficiently accurate control of a consumer is possible even at low flow rate of a variable displacement pump or a constant displacement pump.

This object is achieved by a control arrangement having the features of patent claim 1.

The control arrangement according to the invention has a controllable depending on the load pressure of a consumer pump assembly and a metering orifice for adjusting the pressure medium flow to the consumer. The load pressure is tapped via an LS line, which is connected by means of a current regulator with a pressure medium sink, for example a tank. In the solution according to the invention, the current regulator can be adjusted as a function of the delivery rate, preferably the pump speed. As a result, a volume flow dependent on the delivery quantity or pump speed flows out of the load-reporting line. This is increased with decreasing speed, so that due to the pressure drop in the load-sensing line, a lower pressure to the pump is reported and this adjusts accordingly. The pressure drop across the metering orifice and thus the pressure medium flow rate flowing over it is reduced, so that the metering orifice must be opened further and the control range of the metering orifice is better utilized.

This concept according to the invention can be applied both in LUDV systems and in the aforementioned LS systems (pressure compensator of Δp via metering orifice) also use in control arrangements in which only one consumer is controlled via a metering orifice (without pressure compensator).

In LUDV systems, a LUDV pressure compensator with an orifice plate is provided, which then generates a constant, larger pressure gradient when the flow controller is opened, as a result of which, as described above, the pressure drop across the metering orifice is reduced. For LS systems such an aperture is not required.

According to the invention, it is preferred if the current regulator is driven as a function of the engine speed of a pump drive. This engine is designed in a preferred embodiment as an internal combustion engine.

In a particularly preferred embodiment, an additional nozzle is arranged in the LS line downstream of the flow regulator, via which the above-described pressure drop can be generated, which then leads to lowering the Volumenεtroms on the metering orifice. This additional aperture makes it possible to more sensitively control several consumers of a control arrangement. Without this additional nozzle, on the other hand, only the consumer with the highest load pressure can be controlled more sensitively, because only with this does the fully opened individual pressure balance not influence the pressure downstream of the metering orifice since this pressure corresponds to the highest load pressure or the pressure set in the LS line. If the pump pressure changes, the pressure difference across the metering diaphragm also changes. On the other hand, at the load pressure-reduced consumers, the individual pressure balances downstream of the metering orifices regulate the lower pressure prevailing in the LS line. Accordingly, with the lower load consumers, the pressure before and after the metering orifices in the changed the same extent, if one adjusts the flow controller - the pressure difference across these orifices of low-load consumers then remains the same.

In a variant of the invention, the pressure in the LS line is limited by a LS pressure relief valve. This can be located either downstream or upstream of the flow regulator. An LS pressure relief valve located downstream of the additional nozzle limits the pressure reported to the pump. In the LS line upstream of this additional nozzle and thus on the backs of all individual pressure balances, there is then a slightly higher pressure which the individual pressure balances adjust downstream of the metering orifice. The pump, on the other hand, only works by the control Δp above the lower pressure determined by the pressure relief valve. This reduces the pressure difference across all metering orifices - in some cases even zero. There is a possibility that not only the consumer who is at the stop but all consumers stop.

In the case where the LS pressure relief valve is connected to the LS line upstream of the additional nozzle, the LS pressure relief valve limits the pressure on the backs of the individual pressure compensators. The pump pressure is higher than the pressure on the backs of the individual pressure compensators by one through the additional nozzle, the adjustment of the flow regulator and the adjustment of the pump regulator or the bypass pressure balance (constant pump), so that the pressure difference across the metering orifices of the lower-load consumer is maintained, even if a consumer is pending an attack. The control arrangement according to the invention is used particularly advantageously in a construction machine, for example an excavator, wherein a slewing gear is to be moved at a comparatively low speed.

Other advantageous developments of the invention are the subject of further subclaims.

In the following a preferred embodiment of the invention will be explained in more detail with reference to schematic drawings. Show it:

FIG. 1 shows a hydraulic circuit diagram of a LUDV control arrangement of a mobile working device;

Figure 2 is a detail of the control arrangement of Figure 1;

Figure 3 is a sectional view of a current regulator of Figure 1;

Figure 4 shows another embodiment of the control arrangement of Figure 1 and

Figure 5 shows a third embodiment of the control arrangement according to the invention;

1 shows a circuit diagram of a working according to the LUDV principle control arrangement 1, as used in a construction machine, such as an excavator. With such a LUDV control arrangement 1 consumers of the excavator, such as the cylinders or hydraulic motors one slewing 2, a spoon 4, a spoon handle 6 and a boom 8 are supplied in response to the activation of a directional control valve block 10 with pressure medium. This is described in the present embodiment. promoted by a constant pump 12, which is driven by an internal combustion engine 14. The control of the internal combustion engine 14 by means of a control lever (throttle / accelerator pedal) 16 which is connected via a throttle cable 20 in operative connection with the motor 14 to adjust its speed. The mobile control block 10 is composed of a plurality of directional valve sections, wherein each of the consumer 2, 4, 6, 8 is associated with a directional valve section with a LUDV valve assembly 22, 24, 26, and 28 respectively. In an input section of the mobile control block 10, a bypass pressure compensator 30 and an LS pressure relief valve 32 are provided.

About the constant pump 12 pressure fluid is sucked from a tank T and conveyed via a feed line 30 to a port P of the mobile control block. Connected to this connection P is an inlet channel 36, via which the consumers 2, 4, 6, 8 can be supplied with pressure medium in the manner described in more detail below. The limited by the LS-pressure compensator 32 maximum load pressure is applied to a connected to a LS port of the mobile control block 10 LS channel 38 at. To the LS connection, an LS tank line 40 is connected, in which an adjustable current regulator 42 is arranged. The adjustment of this current regulator 42 via the throttle cable 20 such that when reducing the speed of the motor 14, the opening of the current regulator 42 is increased. Thus, a comparatively low control oil volume flow to the tank T always flows through this flow regulator 42.

The running of the consumers 2, 4, 6, 8 pressure medium is returned via a tank channel 44, a tank port T and a tank line 46 to the tank T. Further details of the circuit will be shown with reference to the enlarged view in Figure 2, which the pump-side part, the input section and the Consumer 8 associated directional valve section shows the other sections have an identical structure.

Accordingly, the bypass pressure compensator 30 is arranged in a bypass channel 48, via which the inlet channel 36 is connected to the tank channel 44. The bypass pressure compensator 30 is acted upon in the closing direction by the force of a spring and by the pressure present in the LS channel 38, which is picked off via an LS control channel 50. In the opening direction of the pressure at the entrance of the bypass pressure compensator 30, d. H. , the pressure in the supply line 36. The spring of the bypass pressure compensator is selected so that in the supply line 36, a pressure is adjusted by a pump Δ p (for example, 10 bar) above the load pressure in the LS channel 38.

About the LS pressure relief valve 32, the pressure in the LS channel 38 is limited to a maximum value. For this purpose, the LS pressure relief valve 32 is acted upon in the closing direction by the adjustable force of a spring, in the opening direction, the pressure acts at the input of the LS pressure relief valve 32, which is connected via a channel 52 to the LS channel 38.

The in each case a consumer 2, 4, 6, 8 associated LUDV valve assemblies - in Figure 2, reference numeral 22 - consist essentially of a continuously variable directional control valve 54 and a LUDV pressure compensator 56. In the directional control valve 54, a directional part 58 and a speed part are formed , which has a variable metering orifice 60, which are formed by the same spool. If such a directional control valve 54 is displaced from its spring-biased basic position into one of its two lateral working positions (a), (b), pressure medium coming from the inlet channel 36 flows from an inlet chamber 62 via the metering orifice 60 into an intermediate chamber 64, from there via an opening cross-section of the LUDV pressure compensator 56 into a second intermediate chamber 66 and then via the directional part 54 into a consumer chamber 68 or 70 and from there via a feed channel 72 and a return channel 74 to two working ports A, B of the directional valve section. The working port A is then connected via a supply line 76 with a bottom-side cylinder chamber 78 and the working port B via a return line 80 with an annular space of the consumer 8, that is connected to the boom actuating the lifting cylinder. A control piston of the LUDV pressure compensator 56 is designed so that, when this pressure compensator 56 is fully opened, it creates a throttled connection between the intermediate chamber 64 and the LS channel 38. This is the case when the associated hydraulic consumer is operated alone or when in a simultaneous operation of several hydraulic consumers of the LUDV pressure compensator 56 associated hydraulic consumer has the highest load pressure. The control piston of the LUDV pressure compensator 56 is provided with a diaphragm 84, via which the line section connected to the intermediate chamber 64 is connected to a rear chamber 86 of the LUDV pressure compensator 56, which is connected to the LS channel 38 via a signaling channel 88. Accordingly, the control piston of the LUDV pressure compensator 56 is acted upon in the closing direction by the pressure in the LS pressure channel 38, generally the highest load pressure and in the opening direction by the pressure in the intermediate chamber 64. As described above, the pressure drop across the metering orifice 60 is kept constant with respect to the load pressure via this LUDV valve arrangement 22 with the metering orifice 60 and the downstream LUDV pressure compensator 84. With regard to the further function, reference is made to the aforementioned DE 199 30 618 A1 for the sake of simplicity. FIG. 3 shows a section through a specific exemplary embodiment of the current regulator 42. The basic structure of such a current regulator 42 is known, so that only the components essential for understanding are described here. The current regulator 42 essentially consists of a variable metering orifice 90 and a pressure compensator 92 connected upstream of it, which is shown in a control position in FIG. The metering orifice 90 and the pressure compensator 92 are received in a housing 94 on which an input port 96 and an output port 98 are formed.

The metering orifice 90 has an orifice bore 100 formed by a radially recessed portion of a housing bore 102 closed on one side. The opening cross-section of the aperture bore 100 can be changed by means of a metering orifice valve 104, which is guided rotatably and sealed in a vertical bore 106 of the housing 94. The upper end section 108 of the metering gate valve, which is at the top in FIG. 3, protrudes from the housing and is connected to the throttle cable 20 via connecting means, not shown, so that an actuation of the throttle cable is converted into a rotation of the metering gate valve 104. This is thus executed in the illustrated embodiment as a rotary valve, wherein according to the rotation of the opening guerschnitt the aperture bore 100 is changed. Alternatively, the metering orifice 104 may be received axially displaceable.

The pressure balance 92 has a pressure compensator piston 110, which is biased by a pressure balance spring 112 against a stop screw 113 screwed into a pressure balance bore 114. In FIG. 3, the pressure balance piston 110 is in one of its control positions. The pressure balance piston 110 has two annular grooves 116, 120, the are separated from each other by a control collar forming a control edge 122. In the right, stop-screw-side end face in FIG. 3, an angular bore 124 opens, which on the other hand opens into the annular groove 116 via a short radial limb, which is hydraulically connected to the housing bore 102 and the vertical bore crossing the pressure balance bore 114. The input port 96 opens in the region of the annular groove 120, the output port 98 is connected on the one hand to the housing bore 102 and on the other hand to a spring chamber for the spring 112 of the pressure compensator 92. Accordingly, the pressure balance piston 110 in the opening direction (to the stopper screw 113) by the force of the pressure balance spring 112 and the pressure at the outlet 98, ie, the pressure downstream of the aperture bore 100 and in the closing direction by the pressure in the space between the right end face of the pressure compensator piston 110 and the stop screw 113 is acted upon, which corresponds to the pressure in the vertical bore 106 and thus upstream of the metering orifice 100. The pressure medium volume flow through the metering orifice 100 is determined by the setting of the metering orifice slide 104, wherein the pressure drop across the metering orifice 90, more precisely above the orifice bore 100, is kept constant independent of the load pressure. Ie. , as the pump pressure increases, this pressure increase is throttled by the pressure compensator 92.

It is now assumed that a consumer, for example, the slewing gear 2 is to be moved at a comparatively low speed and that the highest load pressure is applied to the slewing gear 2 or that only the slewing gear 2 is activated. According to the LUDV principle, the LUDV pressure compensator 56 of the LÜDV valve arrangement 22 (slewing gear 2) is then completely opened - the corresponding load pressure of the slewing gear 2 then lies in the load signaling channel 38. In a conventional control arrangement With a flow regulator whose volumetric flow can not be adjusted, the fixed displacement pump 12 would only rotate at comparatively low speed due to the small pivoting of the control lever 16 and accordingly only a small pressure medium volume flow via the metering orifice 60 and the fully opened LUDV pressure compensator 56 to the slewing gear 2 flow and run from this over the directional control valve 54 and the tank channel 44 to the tank T out. The control range of the valve spool of the directional control valve 54 would - as described above - not fully utilized. This is inventively prevented that the current regulator 42 is adjusted in response to the setting of the control lever 16 via the throttle cable 20 so that the control oil volume flow is increased via the current regulator 42. This control oil volume flow generates a pressure gradient across the orifice 84 of the LUDV pressure compensator, so that correspondingly a lower load pressure is reported to the bypass pressure compensator 30. Since the pump pressure is always around the control Δp above reported pressure, the pressure drop across the metering orifice changes accordingly with the adjustment of the flow regulator 42. The pressure medium flow flowing through the metering orifice 60 is reduced due to the smaller pressure difference and the driver must use the metering orifice 60 not shown Vorstellgerät readjust so that the consumer is moved at the desired low speed - the control range of the slide of the directional control valve 54 is thus utilized much better than the above-mentioned prior art.

In the case in which, in addition to the slewing gear 2, further, lower-load consumers are actuated, the pressure difference across the metering orifices associated with this consumer remains constant, as explained in detail above, since the lower pressure also builds up downstream of the metering orifices. The Adjustment of the current regulator 42 thus has no effect on the load-pressure-reduced consumers in the exemplary embodiment illustrated in FIG.

FIG. 4 shows an improved embodiment in which a further nozzle 118 is provided in the region between the flow regulator 42 and the LUDV pressure compensators 56. In the embodiment shown in Figure 4, this nozzle 118 is located downstream of the branch of the channel 52, in which the LS pressure relief valve 32 is located. At the back of the pressure compensators is still the highest load pressure, d. H. , the pressure upstream of the other nozzle 118. When the consumers are actuated, a constant pressure gradient is generated via this nozzle 118 as a function of the setting of the flow regulator 42, so that the pressure reported to the bypass pressure regulator 30 is lower than the highest load pressure or the load pressure of the single consumer. In the manner described above, the pump pressure is then adjusted via the predetermined control Δp on this reduced pressure, so that correspondingly, the pressure drop across the metering orifice 60 and the pressure medium flow rate flowing thereabove is reduced. Accordingly, the pressure drop across the metering orifices of the lower-load consumer is reduced so that all consumers can be controlled more sensitively.

It is now assumed that several consumers are driven and drives one of these consumers to the attack. The LS pressure limiting valve 32 is then opened the connection to the tank channel 44 when the preset maximum load pressure is exceeded, thus limiting the pressure upstream of the nozzle 118. This limited pressure is applied to the rearward spaces 86 of the LUDV pressure compensators 56. The pump pressure then adjusts according to the pressure drop across the additional nozzle, the setting of the flow controller and the control Δp the bypass pressure compensator to a higher value than it rests on the backs of the LUDV pressure compensator 56, so that the pressure difference across the metering orifice 60 of the lower-load consumer is maintained, even if the load pressurehighest consumers on a Stop stands.

In the exemplary embodiment illustrated in FIG. 5, the additional nozzle 118 is arranged upstream of the LS pressure-limiting valve 32, ie. H. , the channel 52 branches off downstream of this nozzle 118 from the LS channel 38. This embodiment does not differ from the above-described embodiment with "normal" control of the consumer. A difference only arises when one of the consumers drives to the limit. In this case, the control valve pressure reported to the pump 12 or more precisely to the bypass pressure compensator 30 is then limited via the LS pressure limiting valve 32. The pressure upstream of the additional nozzle 118 and thus the pressure in the rear chambers 86 of the LUDV pressure balances 56 is then higher than the limited pump pressure. This higher pressure is adjusted via the LUDV pressure compensators 56. The pump pressure, however, only adjusts by the control Δp above the comparatively low pressure determined by the LS pressure limiting valve 32. Thus, the pressure difference across all orifices 60 is smaller, there is even the possibility that this pressure difference is 0 and all consumers stop.

In the embodiment described above, the control oil volume flow flowing through the flow regulator 42 is conducted back to the tank T. In principle, however, it is also possible to supply a control circuit to this control oil volume flow and to utilize it there, so that the losses are reduced. The above-described embodiments show LUDV systems. However, the concept according to the invention can also be used in LS systems. Instead of the constant pump and a variable displacement pump with pump regulator can be used, which is adjusted in dependence on the pressure in the LS channel (38).

Disclosed is a hydraulic control arrangement for pressure medium supply at least one hydraulic consumer, with a LS pump arrangement and a metering orifice for adjusting the pressure medium volume flow to the consumer. The LS line is connected via a current regulator with a pressure medium sink. According to the invention, the current regulator is adjustable as a function of the pump speed in order to change the pressure drop across the metering orifice.

LIST OF REFERENCES:

1 control arrangement

2 lathe

4 spoons

6 dipperstick

8 outriggers

10 mobile control block

12 constant pump

14 engine

16 levers

20 throttle cable

22 LUDV valve arrangement

24 LUDV valve arrangement

26 LUDV valve arrangement

28 LUDV valve arrangement

30 Bypass pressure balance

32 LS pressure relief valve1

34 supply line

36 inlet channel

38 LS channel

40 LS tank line

42 current regulator

44 tank channel

46 tank line

48 bypass channel

50 LS control channel

52 channel

54 way valve

56 LUDV pressure balance

58 direction part

60 metering orifice

62 inlet chamber

64 intermediate chamber

68 Consumer Chamber consumer chamber

forward channel

Return channel

supply line

Zylinderräum

Return line

annulus

Aperture rear space

signaling channel

orifice

pressure compensator

casing

input port

output port

B1endenbohrung

housing bore

Measuring shutters

vertical drilling

end

Pressure regulator piston

Compensator spring

stop screw

Pressure compensator bore

ring groove

jet

ring groove

control edge

Claims

claims:

1. Hydraulic control arrangement for supplying pressure medium to at least one hydraulic consumer (2, 4, 6,

8), with a depending on a via a LS line (38, 40) tapped consumer load pressure controlled pump assembly (12, 30) and a metering orifice (60) for adjusting the pressure fluid flow to the consumer (2, 4, 6 , 8), wherein the LS line (38,

40) is connected via a flow regulator (42) to a pressure medium sink, characterized in that the flow regulator (42) is adjustable in dependence on the flow rate of the pump arrangement (12, 30), preferably on the pump speed.

2. Control arrangement according to claim 1, wherein the pump (12) by a motor (14) is driven and the current regulator (42) is adjustable in dependence on the engine speed.

3. Control arrangement according to claim 2, wherein the motor (14) is an internal combustion engine.

4. Control arrangement according to one of the preceding claims, wherein in the LS line (38, 40) upstream of the flow regulator (42) has a nozzle (84, 118) is arranged.

5. Control arrangement according to claim 4, with a LS- pressure relief valve (32) for limiting the LS- pressure, which is arranged either between the flow regulator (42) and the nozzle (118) or upstream of the nozzle (118).

6. Control arrangement according to one of the preceding claims, wherein the metering orifice (60) is associated with a LUDV or LS-pressure compensator (56).

7. Control arrangement according to claim 6, wherein the LUDV pressure is provided with an aperture (84) (56) are connected via the two limited! ^¬ of pressure balance piston control surfaces th pressure compensator spaces.

8. Control arrangement according to one of the preceding claims, wherein the pump is a variable displacement pump or a variable-speed fixed displacement pump (12) with a bypass pressure compensator (30).

9. Control arrangement according to one of the preceding claims, wherein the consumer is a slewing gear (2) of a mobile working device.