DE4423572A1 - Load-sensing hydraulic system for use with mobile excavator - Google Patents

Abstract

The hydraulic system is used for one or more loads (700) with the hydraulic control provided by a hydraulic pump (100), a pump regulator (300), a control valve (400) and pre-control unit for a pre-control valve (800). The pressure difference obtained across the control valve at high pressures is variable in dependence on the pressure at the load, this pressure supplied to a regulating valve (900) via a control line, for modification of the pre-control pressure.

Description

The present invention relates to such a called load-sensing hydraulic system and in particular to a load-sensing hydraulic system in which the Hydraulic pump regulating pump regulator (flow regulator) is unable to increase the flow rate of the pump regulate that the differential pressure at the control valve of the Consumer remains constant.

So-called load-sensing hydraulic systems are in the Technology known. A load sensing circuit is in the Principle of a combined flow and pressure control basically consists of a variable displacement pump with pumps regulator and pressure relief valve and one or more ver users. The control takes place with a special throttling directional control valve (control valve), which has a additional control line the pump controller with the the respective load pressure.

The pump controller regulates the flow of the pump so that the differential pressure Δp at the control valve for each opening cross section remains constant, regardless of Consumer pressure needs. When the turned on pressure set responds to the pressure relief valve and adjusts the pump to a smaller stroke volume.

If the opening cross section of the control valve ver increases, the pressure difference Δp and the decreases Pump controller increases the stroke volume of the pump long until the pressure difference Δp on the control valve again corresponds to the spring force set on the pump controller.

The flow rate can be directly controlled by the control valve Adjust consumers, the pump pressure only by Δp is above the load pressure.  

With load-sensing hydraulic systems in particular several consumers, for different requirements and requirements have to be met, it can happen that the pump controller is unable to follow the Δp curve (Pressure-difference curve) of the control valve constantly increases hold. For example, Δp increases with decreasing High pressure on the consumer.

This problem occurs when a consumer for the maximum speed or speed does not have the maximum pump delivery Q _max. required but less than Q _max. .

For example, in a mobile excavator, it may be that the hydraulic traction motor (hydraulic motor) is operated with a maximum quantity (swallowing volume) that is less than the maximum pump delivery quantity Q _max. . The maximum swallowing volume of the traction motor is limited by the control piston of the control valve, provided the Δp curve is constant. The valve curve of the control valve is designed for a constant Δp, so that the maximum swallowing volume of the engine flows when the control piston is fully switched.

If the Δp increases - which should not be - then will the flow rate through the control valve is greater, and in proportion

In the example above, this causes a problem that when driving downhill the driving pressure drops motor increases the Δp and thus the delivery rate for the Hydromotor is bigger than wanted. Thus the High pressure on the traction motor slower than wanted and that  The engine brake valve is prevented from engaging and speaks later than intended. On the one hand, this has Consequence that the hydraulic motor is turned over, what life duration of the hydraulic motor is detrimental, and on the other the driving speed increases, which is undesirable and may also be prohibited and / or dangerous.

The problem is solved according to the present invention solved that the high pressure line of the drive motor with a Control valve is connected, which in turn with the front control of the control valve is connected. The input tax pressure of the control valve is dependent on the Δp- Increase by the control valve is reduced and thus the control piston of the control valve with a lower Stroke and the Δp increase is compensated.

The high pressure dependent on the consumer Change in the pressure difference occurring at the control valve Δp is therefore compensated with a variable measuring orifice. According to an embodiment of the invention, the variable orifice plate a control valve and a control valve on, the control valve preferably a throttling 6/3 proportional directional valve with hydraulic control and the control valve is preferably a regulated, adjustable pressure relief valve.

Further advantages, aims and details of the invention result from the following description of an off exemplary embodiment of the invention with reference to the drawing; in the drawing shows:

Fig. 1 shows schematically a hydraulic circuit diagram of a conventional load-sensing hydraulic control;

Figure 2 is a hydraulic circuit diagram for an excavator control with load-sensing hydraulics.

Fig. 3 is a simplified schematic hydraulic circuit diagram of an inventive load sensing control with additional control valve;

Fig. 4 shows a detail of the circuit diagram of Figure 2, but with the inventive control valve.

Fig. 5 is a diagram showing the Δp curve depending on the high pressure at the consumer.

In Fig. 1, a variable displacement pump 100 is shown, whose delivery rate is adjustable by means of an adjusting 200th The adjustment device 200 is actuated by a pump controller or flow controller 300 via a control line. The pump controller 300 is a 3/2-way valve.

The high pressure line of the pump 100 is connected to a throttling directional control valve 400 as well as to a connection and a control input of a pressure relief valve 500 and to a control input of the pump regulator 300 . The control input on the opposite side of the pump regulator 300 is connected via a control line to the high pressure reduced by Δp and present at the consumer, that is to say there is a pressure difference of Δp between the control inputs of the pump regulator 300 . This pressure difference is compensated for by a spring 600 so that the pump regulator 300 is in equilibrium if there is a pressure difference of Δp between the control inputs.

The pressure difference Δp results from the throttling effect of the throttling directional control valve (control valve) 400 . The pump controller 300 regulates the flow rate of the pump 100 so that the differential pressure Δp at the control valve 400 remains constant for every opening cross section, regardless of the pressure requirement of the consumer. When the set maximum pressure is reached, the pressure relief valve 500 responds and adjusts the pump 100 to a smaller stroke volume.

If the opening cross section of the control valve 400 increases, the pressure difference Δp decreases and the pump regulator 300 increases the stroke volume of the pump 100 until the pressure difference Δp on the control valve 400 again corresponds to the spring force set on the pump regulator 300 .

Via the control valve 400 , the flow rate can be directly adapted to the consumer 700 , the pump pressure being only Δp above the load pressure.

Fig. 2 shows the hydraulic control of an excavator in detail. The individual components and connections are readily apparent to those skilled in the art from the drawing. With regard to the details of FIG. 2, reference is also made to DE-OS 44 05 472, the content of which is hereby made the subject of the present disclosure by reference.

The drive motor 1 is preferably a diesel engine with a nominal output of 101 kW at a nominal speed of 2000 min ^-1 .

A pump 41 supplies pressure for a steering device of the vehicle, which is not shown here. Another pump 37 provides pressure for a braking device control 38 . The pumps 41 and 37 preferably have a delivery rate of 19 l / min. or 38 l / min. at the nominal speed of the drive motor of 2000 revolutions per minute. The pump 41 for the steering could, however, just in case a delivery volume of 38 l / min. have.

The drive motor 1 also drives a work pump designed as a load-sensing pump (LS pump) in the work pump unit 3 and a slewing gear pump in the slewing gear pump unit 4 .

The working pump or LS pump is preferably one Axial piston pump in swash plate design in an open circle for driving and excavator operation. The work pump is differential pressure controlled, their maximum operating pressure is preferably 320 bar, their maximum delivery rate preferred way 260 l / min. The pump output can be or depending on the mode and depending on the setting 52, 72 or 101 kW.

A load-sensing system for the work pump and the drive factory or tools has the characteristic of load pressure independent flow distribution. The load pressure independent Flow distribution is a special kind of Load sensing control. It enables one of the momen tanen load state independent, sensitive Control of consumer movements.

The ver. Controlled by the highest load pressure (LS signal) control pump only provides the delivery volume, the requested via an orifice plate of the control block becomes. If the requested volume flow exceeds the maxi Male pump delivery rate (e.g. when actuating several Consumer), the pump volume flow becomes load pressure independent and proportional across the orifice cross cuts allocated to consumers.

If all consumers are in neutral position, the pump in standby pressure (approx. 20 bar) only pumps the leakage flow. The LS line is relieved to the tank via a 2-way flow controller (Q _max = 0.5 l / min.). The pressure compensators are in their closed position via their retaining spring and are also unloaded from the tank. The pump is separated from the consumers. Unnecessary circulation and throttling losses are avoided.

Becomes a consumer, for example the highest load When operated, the oil flow flows from the pump via the orifice plate, the downstream pressure compensator and the load holding valve to the consumer. For the LS message the pressure from the Load holding valve used. This measure becomes a Falling back of the load (by taking volume from the Consumer line) in the LS line with sensitive Prevents movement. When the load pressure is reached opens the load holding valve and the consumer moves in the selected direction with that through the orifice plate given speed.

If several consumers with low pressure level in Operated in parallel, the highest LS Signal to the differential pressure regulator and to all Section pressure compensators reported. The arrangement of the pressure scales between orifice plate and consumer guaranteed same differential pressure (Δp2) across all piston orifice plates. The differential pressure (Δp3) between the highest and lowest load is at the control edge of the Pressure compensated.

In parallel operation with undersaturation, i.e. H. if over the piston orifice plates a larger volume flow Q is indicated requests than the pump can provide the control Δp2 (undersaturation) and the Differential pressure regulator goes into neutral position. The LS The control function of the pump is not active. The pump will only by the power controller according to the Performance target set. The load pressure independent Flow distribution is down to a minimum Δp maintain approx. 3 bar.  

If a consumer drives to a stop, d. H. in the end position, the LS pressure relief valve responds and fulfills one Print cut function. The pump swings to minimal swivel angle and holds the pressure cut value constant. Only the system leakage oil and the LS exit flow as well as the pressure cutoff valve flow applied at 20 bar.

If one or more consumers are also in operation, the pump swings out. The Δp required for this is generated by the volume flow flowing through the pressure cut-off valve. The rest of the available volume flow Q _R , which is reached at the control end of the power setting, can be distributed independently of the load pressure and proportionally in the ratio of the orifice openings.

This results in the following relationship:

P _loss = (p _Bmax - p _v2 ) × Q₂ + (p _Bmax - p _v3 ) × Q₃.

The working pump unit 3 supplies pressure fluid to a valve control unit 50 , which supplies the pressure fluid to the various consumers. The valve control unit 50 is controlled by a pilot control unit 60 . The operating pressure for the hydraulic pilot control unit 60 is branched off in a valve unit or control oil supply 25 from a line coming from the brake arrangement 38 . A control line also runs from the pilot control unit 60 to the rotary pump unit 4 for controlling it.

The consumers are designated in Figure 2 as follows:
Boom 12 , stick 13 , bucket 14 , hydraulically adjustable boom (VAH / HAB) 15 , hammer 16 , additional function (Aux) 17 , chassis 19 , undercarriage support 21 and dozer blade 22 . Other and further functions are possible.

Fig. 2 shows the diesel engine 1 and the driven thereby working pump or LS pump in the working pump unit 3. Further, in the working pump unit 3, a load-sensing control is shown, including using two 3/2-way proportional directional valves.

Also shown is the slewing gear pump unit 4 , which also includes the control hydraulics for the slewing gear 18 . The slewing gear pump has working connections which are connected to the slewing gear motor 18 a using closed-loop technology, ie in a closed circuit.

The LS valves or load-sensing valves in the valve control unit 50 are 6/3 proportional directional valves with hydraulic control. The control fluid coming from the respective control elements of the pilot control unit 60 acts on the 6/3 directional control valve, which is spring-centered, via a throttle check valve. The pressure fluid left by the 6/3-way valve is passed through a 3/3-way proportional valve and then further via non-return valves and again through the 6/3-way valve to the respective consumer. Furthermore, a Druckab cut is provided, which is designed by means of two pressure relief valves and a throttle valve and acts for the entire valve control unit.

In the pilot control unit 60 , the pilot control or operating elements are the right and left pilot control device (joystick) 27 and 28 , the accelerator pedal 29 , the boom adjustment pedal (VA pedal) 30 , the pilot control device 31 for the hammer and the pilot control device 32 for the Dozer blade and / or support.

The control levers or pedals each work in opposite directions on two proportional valves which are in the neutral position of the Control lever or the pedals are closed and ent speaking their deflection are opened to the pressure present at port P is proportional create the respective outlet.

If, for example, only the traction motor 19 is to be actuated as the only consumer, an operator will actuate the corresponding proportional valves of the pilot control unit 60 via the accelerator pedal 29 . Thus, pressure is applied via a pilot control line to the control piston of the control valve, which is deflected in accordance with the actuation of the accelerator pedal 29 and opens or releases a corresponding passage cross section for high pressure fluid from the pump for the drive motor via control edges. In accordance with the open passage cross section, a certain amount of fluid flows via the high pressure line to the traction motor 19 , the delivery rate of the pump being readjusted via the pump controller (included in the working pump unit 3 ) in such a way that Δp is essentially maintained across the control valve.

Since the Δp is flow-dependent and not load-pressure-dependent, a load-pressure-independent delivery quantity control is thus provided, so that the consumer, here the traction motor 19 , moves independently of the load at a speed or speed according to the pilot control setting, provided the hydraulic pump output is sufficient in order to deliver the corresponding delivery rate at the existing load pressure in order to always provide Δp across the control valve for any control valve opening. Then a certain modulation or a certain actuation path of the accelerator pedal 29 will correspond to a certain driving speed, regardless of the load on the driving motor 19 of the excavator.

Ideally, the pressure difference Δp across the control valve is constant at any load pressure (high pressure). For example, in a mobile excavator that has several consumers, such as boom 12 , stick 13 , bucket 14 , undercarriage 19 , slewing gear 18, etc., for which various requirements and specifications have to be met, it may happen that the pump controller is not able is to keep the Δp curve (pressure difference curve) of the control valve constant. For example, Δp increases with decreasing high pressure at the consumer.

In the mobile excavator according to Fig. 2, the hydraulic driving motor (hydraulic motor) 19 with a maximum amount (displacement volume) of 200 l / min. operated, which is less than the maximum pump delivery Q _max. from 260 l / min. The maximum swallowing volume of the traction motor is limited by the control piston of the control valve, provided the Δp curve is constant. The valve curve of the control valve is designed for a constant Δp of 20 bar, so that the maximum swallowing volume of the engine flows when the control piston is fully switched.

Now increases the Δp, then the flow rate through the control valve larger, in proportion

In the example above, this causes a problem that when driving downhill the driving pressure drops motor, for example when the high pressure is below 100 bar falls, the Δp increases, for example to approximately 25 bar, and thus the delivery rate for the hydraulic motor is larger  is wanted because the pump is only then Pump controller is reduced when the increased Δp on Control valve is present. Thus, the high pressure drops on Traction motor slower than intended when driving downhill and that The engine brake valve is prevented from engaging and speaks later than intended. On the one hand, this has Consequence that the hydraulic motor turns over at full throttle what is detrimental to the life of the hydraulic motor, and on the other hand the driving speed increases what is undesirable and also inadmissible and / or dangerous can be.

Fig. 3 is a greatly simplified hydraulic circuit diagram compared to Fig. 2, showing only the parts and components that are essential for the description of the invention.

In particular, the hydraulic pump 100 is connected with its working connection via a high-pressure pump line 101 to the control valve 400 and to the pump regulator 300 . The control valve 400 is also connected to the tank T by a working connection. A control line 102 is led from the control valve 400 to the pump controller 300 , so that the Δp occurring at the control valve is also present at the pump controller, which regulates the delivery rate of the pump 100 .

The control valve 400 is controlled via two pilot lines 103 , 104 , the pressure or pressure difference of which determines the displacement of the control piston of the control valve 400 . The pilot lines 103 , 104 are pressurized via pilot valves 800 in accordance with the deflection of the respective pilot device, which in the present case is the accelerator pedal.

Two pressure lines 105 , 106 lead from the control valve 400 to the traction motor 700 , with forward travel one pressure line 106 being under high pressure and with reverse travel the other pressure line 105 being under high pressure.

As shown in FIG. 3, an additional control valve 900 is now provided according to the present invention, which is connected with a control connection to the high-pressure line 106 of the traction motor 700 and with a working connection (input) and with a further control connection with the pilot control line 104 of the control valve 400 is connected, specifically at a point between a throttle 108 with a diameter of preferably 1.2 mm and the control connection of the control valve 400 . The control valve 900 is set in such a way that it reduces the pilot pressure of the control valve 400 as a function of the Δp increase and thus the modulation or the stroke of the control piston of the control valve 400 is reduced. The Δp increase at low high pressure (for example <100 bar) caused by the pump controller 300 is thus compensated for by the action of the control valve 900 .

A brake valve (shown in FIG. 2 for the traction motor 19 ) responds and develops its hydraulic braking effect when the high pressure at the consumer falls below 40 bar, for example. Without the control valve 900 according to the invention, the high pressure will reach this response value of the brake valve only later or not at all due to the increase in Δp, as a result of which the brake valve is prevented from engaging and its braking action is only delayed, with reduced action or does not develop at all. When using the control valve according to the invention, the brake valve will respond properly.

It should be noted that the control connection of the control valve 900 in the present case is connected to that pressure line 106 to the traction motor 700 which is under high pressure when the excavator is moving forward, since the above-mentioned problems occur mainly when driving forward at full throttle downhill. Of course, a corresponding control valve could instead or in addition be provided in connection with the high-pressure line which is under high pressure when reversing in order to compensate for an increase in Δp when reversing.

Of course, the present invention is also not for use with a traction motor one Wheeled excavators limited. Rather, it is straightforward possible, the hydraulic control according to the invention to use any hydraulic consumer where the problems mentioned above arise.

Fig. 4 shows a section of Fig. 2, but in addition the control valve according to the invention is shown in the vicinity of the control valve for driving.

Fig. 5 shows a diagram in which the pressure difference Δp is drawn on the control valve against the high pressure on the consumer. The dash-dot line denotes the desired, constant Δp setpoint curve of, for example, 20 bar. The solid line denotes the actual Δp curve, which increases as the high pressure decreases. The actual Δp begins to rise above the setpoint of Δp when the high pressure drops below, for example, 100 bar. At higher high pressures, the actual Δp is slightly below the setpoint of Δp.

Claims (9)

1. Load-sensing hydraulic control for one or more consumers ( 700 ), the hydraulic control comprising:
a hydraulic pump ( 100 ),
a pump controller ( 300 ),
a control valve ( 400 ) designed as a throttling directional control valve,
a pilot control unit for controlling the control valve ( 800 ),
wherein the pressure difference (Δp) across the control valve ( 400 ) is not constant when the high pressure changes,
characterized,
that the high pressure applied to the consumer ( 700 ) is led via a control line to a control valve ( 900 ) which modifies the pilot pressure. 2. Load-sensing hydraulic control according to claim 1, characterized in that the pilot pressure is reduced if the high pressure is below one certain value drops. 3. Hydraulic control according to one of claims 1 or 2, characterized in that the consumer ( 700 ) is a drive motor of a mobile excavator. 4. Hydraulic control according to one of the preceding claims, characterized in that the working pump ( 100 ) is reset when the maximum operating pressure is reached, ie the swash plate is pivoted back, thereby preventing power loss, while maintaining the pressure (pressure cut-off). 5. Hydraulic control according to one of the preceding claims, characterized in that a demand current control is provided in which the hydraulic pump ( 100 ) is reset to approximately zero delivery without a controlled working movement, ie without actuation of operating devices ( 60 ). 6. Hydraulic control according to one of the preceding claims, characterized in that a load-independent control is provided in which the actuation path to be undertaken on the operating devices ( 60 ) remains the same even at different load pressures. 7. Load-sensing hydraulic control for one or more consumers ( 700 ), a pressure difference (Δp) occurring on the control valve ( 400 ) which changes as a function of the high pressure on the consumer ( 700 ), characterized in that the change in Pressure difference (Δp) is compensated with a variable orifice ( 400 , 900 ). 8. Hydraulic control according to claim 7, characterized in that the variable measuring orifice has a control valve ( 400 ) and a control valve ( 900 ). 9. Hydraulic control according to claim 8, characterized in that the control valve ( 400 ) is a throttling 6/3-proportional directional valve with hydraulic control and that the control valve ( 900 ) is a regulated, adjustable pressure relief valve.