EP0954703B1 - Power- and moment-regulating system for a plurality of hydraulic pumps - Google Patents

Abstract

The invention concerns a power- and moment-regulating system for at least two adjustable hydraulic pumps (10, 11), said system comprising for each hydraulic pump (10, 11) a hydraulic servo control apparatus (22, 23) for continuously adjusting the delivery. To that end, the delivery of each hydraulic pump (10, 11) is determined as a function of the delivery pressure of the respective hydraulic pump (10, 11) in a delivery pressure line (16, 17) associated with the hydraulic pump (10, 11) and the control pressures in control lines (20, 21) provided for each hydraulic pump (10, 11). Each servo control apparatus (22, 23) is provided with a moment valve (42, 43) having a valve piston (62) which can move in a valve sleeve (61) and forms a valve seat with the latter and whose closure force is determined by a measuring spring arrangement (46, 47) which is connected to the pump actuator (37, 38) and is pretensioned as a function of the set delivery. To that end, each moment valve (42; 43) connects the control line (20, 21) of the associated hydraulic pump (10; 11) to an outlet (41) as a function of the control pressure in this control line (20; 21) and the control pressure in the control line (21; 20), that is the control pressures in the control lines of the other hydraulic pump(s) (11; 10) in each case, under the pretension of the measuring spring arrangement (46; 47). According to the claimed development, for a given hydraulic pump (10; 11), the control pressure in the control line (21; 20) or the control pressures in the control lines of the other hydraulic pump(s) (11; 10) in each case act both on the valve piston (62) and on the valve sleeve (61) of the moment valve (42; 43) associated with the given hydraulic pump (10; 11).

Description

The invention relates to a power or torque control device for at least two adjustable hydraulic pumps with one for each hydraulic pump hydraulic servo control unit for continuous adjustment of the delivery rate accordingly the preamble of claim 1.

A generic power or torque control device is e.g. from EP 0149 787 B2 known. In the known power or torque control device, the Delivery rate of each hydraulic pump depending on the delivery pressure of the respective hydraulic pump in a discharge pressure line assigned to the hydraulic pump and the control pressures in the control lines provided for each hydraulic pump. To the servo control unit has a pump actuator in the direction of the maximum delivery rate swiveling device and one on the pump actuator in the direction of a Delivery volume reduction acting piston, the piston area over a hydraulically operated control valve can be acted upon with the delivery pressure or with a Process is connectable. The control valve is actuated by the control pressure in the control line of the corresponding hydraulic pump. There is one for each servo control unit Torque valve provided with a valve piston movable in a valve sleeve, the forms a sealing seat with the valve sleeve and its closing force of one Measuring spring arrangement is determined, which is connected to the pump actuator and in Is biased depending on the set flow rate. This moment valve of the two hydraulic pumps connects the control line to the assigned one Hydraulic pump depending on the control pressure in this control line and the Control pressure in the control line of the other hydraulic pump under pre-tensioning the Measuring spring arrangement with the sequence.

The characteristic of this known power or torque control device is shown in Fig. 2 as a function of the high pressure pHD in in the discharge pressure line Dependence on the flow rate Q of the associated hydraulic pump illustrated. A ideal performance curve of one of the two hydraulic pumps when the consumer is switched off in the pressure circuit of the other hydraulic pump is 1 characterized. This hyperbolic ideal characteristic curve 1 has constant power the product of flow rate Q and pressure pHD in the high pressure line constant and the curve is therefore hyperbolic. In the case of EP 0149 787 B2 known control device, the ideal characteristic curve 1 is replaced by a real characteristic curve 1 ' approximated. The real characteristic curve 1 'has two linear sections. In each of the linear The closing force of the valve piston of the torque valve is determined by one of the sections determines two individual springs provided in the measuring spring arrangement of the torque valve. In this way, the hyperbolic profile of the ideal characteristic 1 for the sufficiently approximate practical needs.

If a consumer is now in the pressure circuit of the other hydraulic pump, e.g. a Excavator control, switched on, is known from the EP 0149 787 B2 Control device that provided in the servo control unit of the first hydraulic pump Torque valve through a control line connected to the second hydraulic pump additionally charged. This is done with the known from EP 0149 787 B2 Control device in such a way that the valve piston of the torque valve against Measuring spring arrangement is additionally acted upon in the opening direction. In the in Fig. 2nd P-Q diagram shown corresponds to a parallel shift of the characteristic 1 'in y direction, which is illustrated by the vector y. By connecting the Consumer in the pressure circuit of the second hydraulic pump is the original Characteristic curve 1 'of the first hydraulic pump is converted into characteristic curve 2'. In the area of one relatively low delivery rate Q or in the range of a relatively high pressure pHD in the However, this leads to a considerable deviation from the delivery pressure line corresponding ideal characteristic curve 2 constant power (Q x p = constant). It comes in this area to a torque overshoot, which is hatched in FIG. 2 Area is illustrated. This can overload the first hydraulic pump or of the drive unit. A better approximation would be through the characteristic curve 2 "linearized in sections, but which corresponds to that from EP 0 149 787 B2 known control device is not feasible.

The relatively large deviation from the ideal characteristic curve illustrated in FIG. 2 (Q x p = constant) can be achieved by using a very complex, so-called Hyperbolic power regulator or by electronically working, e.g. microprocessor-controlled power controllers can be avoided. The Construction costs for such solutions and the associated manufacturing costs are however significant and disproportionate to the relatively low cost for a generic power or torque control device such as. B. from the EP 0149 787 B2 or from US-A-4 613 286.

It is therefore the object of the present invention to provide a generic performance or To develop torque control device so that a better approximation to the ideal control characteristic is achieved.

The object is in connection with the characterizing features of claim 1 solved with the generic features.

The invention is based on the finding that a considerably improved approximation to the ideal characteristic can be achieved if by the from the delivery pressure second hydraulic pump or the delivery pressures provided in any number other hydraulic pumps derived control pressure or control pressures not only the Valve piston, but also the valve sleeve of the torque valve in a suitable manner is applied.

According to claims 2 to 10 there are advantageous developments of the invention.

According to claim 2 can on the valve piston of each torque valve for each Control line an associated measuring surface can be provided, which with the control pressure of the each assigned control line in the opening direction of the torque valve is acted upon. Furthermore, according to claim 3, a driving pin on the Pump actuator can be provided in the associated valve sleeve Torque valve for changing the bias of the measuring spring arrangement attacks.

According to claim 4 can in the control line of the other hydraulic pump prevailing control pressure act on a valve sleeve adjusting piston so that the Valve sleeve adjusting piston moves the valve sleeve against a return spring. According to claim 5, it is advantageous if the direction of movement of the valve sleeve adjusting piston directed essentially perpendicular to the direction of movement of the valve sleeve is because this is a particularly compact construction of the torque valve allows. Between the valve sleeve actuating piston and the valve sleeve, a Intermediate member may be provided according to claim 6. The contact area between the Valve sleeve actuating piston and the intermediate element can shift the valve sleeve Compensate perpendicular to the direction of movement of the valve sleeve adjusting piston if that Intermediate member is carried with the valve sleeve.

According to claim 7, the valve sleeve adjusting piston or the intermediate member can Have inclined surface, which is in engagement with the valve sleeve Bolt element attacks. The inclined surface deflects the Direction of movement of the valve sleeve adjusting piston to the direction of movement of the Valve sleeve reached. By appropriately dimensioning the angle of the inclined surface, in a reduction can easily be achieved. According to claim 8, the Driving pin of the pump actuator as a hollow body, in particular as a hollow cylinder be formed, the valve sleeve adjusting piston or the intermediate piece in the Driver pin of the pump actuator slidably engages and from the driver pin is enclosed. This measure also results in a particularly compact one Structure of the torque valve. To the system of the bolt element on the inclined surface of the To enable valve sleeve adjusting piston or the intermediate piece, the Driving pin in the area of the inclined surface corresponding to a suitable recess Claim 9.

Fig. 1

einen Hydraulikschaltplan der erfindungsgemäßen Leistungs- bzw. Momentenregeleinrichtung;

Fig. 2

die Regelcharakteristik einer Leistungs- bzw. Momentenregeleinrichtung nach dem Stand der Technik;

Fig. 3

die Regelcharakteristik einer erfindungsgemäß weitergebildeten Leistungs-bzw. Momentenregeleinrichtung;

Fig. 4

einen vertikalen Schnitt durch ein bei der erfindungsgemäßen Leistungs-bzw. Momentenregeleinrichtung verwendetes Momentenventil; und

Fig. 5

einen horizontalen Schnitt durch ein erfindungsgemäßes Momentenventil entsprechend Fig. 4.

A preferred embodiment of the invention is described below with reference to the drawing. The drawing shows:

Fig. 1

a hydraulic circuit diagram of the power or torque control device according to the invention;

Fig. 2

the control characteristic of a power or torque control device according to the prior art;

Fig. 3

the control characteristic of a performance or. Torque control device;

Fig. 4

a vertical section through a in the inventive power or. Torque control device used torque valve; and

Fig. 5

3 shows a horizontal section through a torque valve according to the invention corresponding to FIG. 4.

1 shows a hydraulic circuit diagram which shows the power or Torque control device illustrated schematically in one embodiment. at the exemplary embodiment shown in FIG. Torque control device for controlling two hydraulic pumps 10 and 11. Die However, the power or torque control device according to the invention is also for Suitable for controlling more than two hydraulic pumps in the same way.

The basic mode of operation of the power or torque control device is apart from the further development according to the invention known from EP 0149 787 B2 and described there in detail. This publication is therefore expressly referred to taken. However, the following is intended to facilitate understanding of the Present invention, the basic operation of the generic performance or. Torque control device are briefly described.

The hydraulic pumps 10 and 11 are each driven by a drive shaft 12 and 13 driven a drive unit, not shown. The hydraulic pumps 10 and 11 suck the pressure fluid, e.g. Oil from a pressurized fluid tank 41 via an intake pipe 14 and 15 and deliver the pressure fluid to a delivery pressure line 16 or 17, where it is available for a consumer that can be connected to port B. To the with Outputs B connected to the delivery pressure lines 16 and 17 are preferably via adjustable throttle elements 18 and 19 control lines 20 and 21 connected. Located behind the throttle elements 18 and 19 in the direction of flow also the work connections A1 and A2 of the Hydraulic pumps 10 and 11. The control line 20 of the first hydraulic pump 10 is connected to the Input X of the servo control unit 22 of the first hydraulic pump 10 and with the input P2 of the servo control unit 23 of the second hydraulic pump 11. The analog is Control line 21 of the second hydraulic pump 11 with the input X of the servo control unit 23 of the second hydraulic pump and with the input P2 of the servo control unit 22 of the first Hydraulic pump 10 connected. The control pressure prevailing in the control line 20 is in a control valve 25 designed as a pressure compensator with the one in the delivery pressure line 16 prevailing delivery pressure compared. For this, the control valve is over a Connection line 24 connected to the delivery pressure line 16. The control valve 25 is a pressure relief valve 26 is connected downstream to the pressure in the control pressure line 27 to limit. In the same way, the second hydraulic pump is in the servo control unit 23 11 a control valve 28, which operates in any case as a pressure compensator, which controls the pressure in the control line 21 compares with the delivery pressure in the delivery pressure line 17. To is the control valve 28 via a connecting line 29 with the delivery pressure line 17 second hydraulic pump 11 connected. The control valve 28 is also a Pressure relief valve 30 downstream to the pressure in the signal pressure line 50 to limit.

The first hydraulic pump 10 is moved in the direction of by a swiveling device 31 maximum delivery rate swung out, while the second hydraulic pump 11 by a Swinging device 32 also in the direction of maximum delivery is swung out. In the exemplary embodiment, the pivoting device 31 or 32 from a piston 35 or 36 which can be loaded against a spring 33 or 34 Swiveling device 31 or 32 engages the delivery rate of hydraulic pump 10 on one or 11 adjusting pump actuator 37 or 38. To reset the Pump actuator 37 or 38 in the direction of a delivery rate reduction serves Pistons 39 and 40, which can be acted upon hydraulically in the signal pressure line 27 or 50 prevailing signal pressure.

With an increase in the delivery pressure prevailing in the delivery pressure line 16 or 17 compared to the control pressure prevailing in the control line 20 or 21 is via the as a pressure compensating control valve 25 or 28 the signal pressure in the signal pressure line 27 or 50 increased and thus the hydraulic pump 10 or 11 in the direction of one Flow rate reduction until equilibrium is reached pivoted back.

The control line 20 or 21 is via a torque valve 42 or 43 with the Pressurized fluid tank 41 connected. The valve piston 44 or 45 of the torque valve 42 or 43 is on the one hand of the control pressure in the control line 20 or 21 respectively assigned hydraulic pump 10 and 11 and on the other hand by the control pressure in the Control line 21 or 20 of the other hydraulic pump 11 or 10 in Direction of opening applied. A measuring spring arrangement 46 or 47 on the valve piston 44 or 45, which in the exemplary embodiment consists of two Individual springs exist, in sections, around the one already explained with reference to FIG. 2 to generate a linear course of the characteristic. The preload of the measuring spring arrangement 46 or 47 is determined by the position of the pump actuator 37 or 38.

Reaches the control pressure in the control line 20 or the control pressure in the control line 21 the value set on the torque valve 42 or 43, it starts To open torque valve 42 and 43 and on the throttle element 18 and 19 there is a Pressure gradient. Consequently, the control valve 25 or 28 is opened further and supplies the Piston 39 or 40 with an increased signal pressure, so that it strives to Pump actuator 37 or 38 to adjust in the direction of a reduced flow rate. The measuring spring of the measuring spring arrangement 46 or 47 of the torque valve 42 or 43 biased. In this way, constant power control is achieved.

According to the development according to the invention, the valve sleeve 48 or 49 is also of the one in the control line 21 or 20 of the other hydraulic pump 11 or 10 prevailing control pressure. Through this development according to the invention there is an improved approximation of the control characteristics of the power or Torque control device on the ideally hyperbolic course. this will described below with reference to FIG. 3.

FIG. 3 shows, similarly to FIG. 2, the one prevailing in the delivery pressure line 16 Delivery pressure pHD as a function of the delivery rate Q of the first hydraulic pump 10 or 11. If the connected to the delivery pressure line 17 of the second hydraulic pump 11 Consumer has only a low power consumption and therefore the second hydraulic pump 11 is only slightly loaded, the first hydraulic pump 10 is connected to the ideal characteristic curve 1 approximate real characteristic curve 1 'adjusted to an approximately constant power. If the second hydraulic pump 11 has a significant power output, the power output by the first hydraulic pump 10 can be reduced so that the Hydraulic pumps 10 and 11 output total power a predetermined maximum value does not exceed and a drive unit driving the hydraulic pumps 10 and 11 is not overloaded. By acting on the valve piston 44 of the Torque valve 42 becomes the parallel displacement illustrated in vector y y direction, i.e. a reduction in the delivery pressure of the hydraulic pump 10. By simultaneously acting on the valve sleeve 48 of the torque valve 42 however, a reduction in the delivery rate of the hydraulic pump 10 also realizes what a parallel displacement illustrated with the vector x leads in the x direction.

As can be seen from a comparison of the control characteristics of a generic 2 with the one in FIG. 3 illustrated characteristic of the performance or Torque control device results directly, leads the development according to the invention to an improved approximation of the control curve 2 "to the ideal control characteristic curve 2.

An exemplary embodiment of the further developed according to the invention is described below Torque valve 42 and 43 described with reference to FIGS. 4 and 5. 4 shows a vertical longitudinal section through the torque valve 42, while Fig. 5 a horizontal longitudinal section through the torque valve 42 shows. The torque valves 42 and 43 are formed in the same way, so that the following description the torque valve 42 is limited.

The torque valve 42 includes a valve housing 60, one in the valve housing 60 axially movably arranged valve sleeve 61 and a movable with respect to the valve sleeve 61 Valve piston 62. The valve piston 62 is via a spring plate 63 through the Measuring spring arrangement 46 acted upon in the closing direction. The measuring spring arrangement 46 in the exemplary embodiment consists of two individual springs 64 and 65 arranged one inside the other, which leads to the sectionally linear control characteristic shown in FIG. 3. The preload of the spring assembly 46 is adjustable by means of a spring bolt 66. For the Control line 20 is a first pressure medium connection P1 and for control line 21 is a second pressure medium connection P2 is provided in the valve housing 60. The one with the Control line 20 connected pressure medium connection P1 is via a connection channel 75 connected to a first pressure chamber 67. When the first one is applied Pressure chamber 67 with the control pressure prevailing in the control line 20 becomes a first measuring surface 68 with the control pressure prevailing in the control line 20 in Opening direction of the torque valve 42 is applied. Once the tip 69 the Control edge 70 reached, the torque valve 46 opens the control line 20 to the Pressure fluid tank 41 out. For this purpose, the stepped bore 71 is via a connecting channel 72 connected to the transverse bore 73 so that the pressure fluid flows into the leakage space 74 can.

The control line 21 connected to the pressure medium connection P2 is via a Connection channel 76 and via further connection channels 77 and 78 with a second Pressure chamber 79 connected, on which a second measuring surface 80 is formed. The in control pressure prevailing in control line 21 therefore acts on valve piston 72 also in the opening direction of the torque valve 42.

As already described, the control pressure prevailing in the control line 21 does not apply only on the valve piston 62, but also on the valve sleeve 61 to this depending on the control pressure prevailing in the control line 21 against a Return spring 81 and the measuring spring assembly 46 to move axially. There is a third Pressure chamber 82 via a connecting channel 90, only partially shown, with the second pressure medium connection P2 connected. The one prevailing in the pressure chamber 82 Control pressure of the second control line 21 thus acts on a valve sleeve actuating piston 83. In the preferred embodiment shown in FIGS. 4 and 5 is the direction of movement of the valve sleeve adjusting piston 83 perpendicular to the Direction of movement of the valve sleeve 61 aligned. This leads to a special compact construction of the torque valve 42 according to the invention. The valve sleeve adjusting piston 83 acts on an intermediate member 84, which has a plate-like, has end face 85. At its plate-like, end 85 opposite end, the intermediate member 84 has an inclined surface 86 which engages a bolt element 87 formed on the valve sleeve 61. With a suitable, flat angle of inclination of the inclined surface 87 can, if necessary, a Reduction between the movement of the valve sleeve adjusting piston 83 and the movement the valve sleeve 61 can be reached. The intermediate link 84 is shown in FIG Embodiment within a driving pin 88 designed as a hollow cylinder, which is connected to the pump actuator 37 in a suitable manner. The Driving pin 88 has a recess 89 for receiving the bolt element 87, so that the bolt element 87 is flush with the inclined surface 86 of the intermediate member 84.

The valve sleeve adjusting piston 83 is located on its opposite the driving pin 88 End biased by an adjusting spring 100 so that the valve sleeve adjusting piston 83 without 4 by the control pressure prevailing in the control line 21 is pressed at the top. In this way, a reset of the valve sleeve adjusting piston 83 reached. The preload of the actuating spring 100 is achieved by adjusting the spring plate 101 adjustable. This is the adjustment of the spring plate 101 after removal of a housing sleeve 102 accessible from the outside.

The pump actuator 3 acts by a horizontal displacement of the driving pin 88 also on the valve sleeve 61. By the plate-like conclusion 85 of the Intermediate member 84 ensures that the valve sleeve actuating piston 83 despite the in Fig. 4 horizontal movement of the driving pin 88 in constant engagement with the intermediate link 84 stands. Due to the vertical orientation in Fig. 4 Direction of movement of the valve sleeve adjusting piston 83 perpendicular to the direction of movement of the valve sleeve 61 and the driving pin 88 can shift the valve sleeve 61st through the driving pin 88 on the one hand and through the valve sleeve adjusting piston 83 on the other hand, take place independently of one another.

The invention is not limited to the illustrated embodiment. The Torque valve can be designed in many other ways. In particular, it can Torque valve further measuring surfaces for the control lines further from the power or. Have torque control device controlled hydraulic pumps. In corresponding The way is then for each additional hydraulic pump to be connected to the valve sleeve actuating piston 83 a separate pressure chamber for each additional one that can be connected Provide hydraulic pump or there are a corresponding number of valve sleeve actuating pistons 83 to be arranged in parallel.

Claims (10)

1. Capacity and/or moment regulating device for at least two adjustable hydraulic pumps (10, 11) comprising in each case one hydraulic servo-control unit (22, 23) for each hydraulic pump (10, 11) for infinitely adjusting the delivery rate,

   wherein the delivery rate of each hydraulic pump is determined in dependence upon the delivery pressure of the respective hydraulic pump (10, 11) in a delivery pressure line (16, 17) associated with the hydraulic pump (10, 11) and upon the control pressures in control lines (20, 21) provided for each hydraulic pump (10, 11),

   wherein each servo-control unit (22, 23) comprises a swing-out device (31, 32) for adjusting a pump actuator (37, 38) in the direction of maximum delivery rate and a piston (39, 40), which acts upon the pump actuator (37, 38) in the direction of a delivery rate reduction and the piston area of which by means of a hydraulically actuable control valve (25, 28) is loadable with the delivery pressure or connectable to an outlet (41), and the actuation of the control valve (25, 28) is effected by the control pressure in the control line (20; 21) of the respective associated hydraulic pump (10; 11),

   wherein there is provided for each servo-control unit (22, 23) a moment valve (42, 43) having a valve piston (62), which is movable in a valve sleeve (61) and forms a sealing seat with the valve sleeve (61) and the closing force of which is determined by a measuring spring arrangement (46, 47), which is connected to the pump actuator (37, 38) and biased in dependence upon the adjusted delivery rate,

   and wherein each moment valve (42; 43) connects the control line (20; 21) of the respective associated hydraulic pump (10; 11) in dependence upon the control pressure in said control line (20; 21) and the control pressure in the control line (21; 20) and/or the control pressures in the control lines of the respective other hydraulic pump(s) (11; 10) to the outlet (41), while simultaneously biasing the measuring spring arrangement (46; 47),

   characterized in that for a specific hydraulic pump (10; 11) the control pressure in the control line (21 ; 20) and/or the control pressures in the control lines of the respective other hydraulic pump (s) (11; 10) acts and/or act both upon the valve piston (62) as well as upon the valve sleeve (61) of the moment valve (42; 43) associated with the specific hydraulic pump (10; 11).
2. Capacity and/or moment regulating device according to claim 1,
   characterized in that on the valve piston (62) of each moment valve (42, 43) there is provided for each control line (20, 21) an associated measuring face (68, 80), which is loadable with the control pressure of the respective associated control line (20, 21) in opening direction of the moment valve (42, 43).
3. Capacity and/or moment regulating device according to claim 1 or 2,
   characterized in that a driving pin (88) of the pump actuator (37, 38) acts upon the valve sleeve (61) of the associated moment valve (42, 43) for varying the bias of the measuring spring arrangement (46, 47).
4. Capacity and/or moment regulating device according to one of claims 1 to 3,
   characterized in that the capacity and/or moment regulating device controls two adjustable hydraulic pumps (10, 11) and in the moment valve (22; 23) of a specific hydraulic pump (10, 11) the control line (21; 20) of the respective other hydraulic pump (11; 10) loads a valve sleeve adjusting piston (83) with the control pressure prevailing in said control line (21; 20) so that the valve sleeve adjusting piston (83) displaces the valve sleeve (61) inside the valve housing (60) against a restoring spring (81) and/or the measuring spring arrangement (46; 47).
5. Capacity and/or moment regulating device according to claim 4,
   characterized in that the direction of motion of the valve sleeve adjusting piston (83) is directed substantially at right angles to the direction of motion of the valve sleeve (61).
6. Capacity and/or moment regulating device according to claim 5,
   characterized in that provided between the valve sleeve adjusting piston (83) and the valve sleeve (61) is an intermediate member (84), which is in frictional connection both with the valve sleeve adjusting piston (83) and with the valve sleeve (61).
7. Capacity and/or moment regulating device according to one of claims 5 or 6,
   characterized in that the valve sleeve adjusting piston (62) and/or the intermediate member (84) has an oblique face (86), which acts upon a bolt element (87) in engagement with the valve sleeve (61).
8. Capacity and/or moment regulating device according to claim 7,
   characterized in that the driving pin (88) takes the form of a hollow body and the valve sleeve adjusting piston (83) and/or the intermediate member (84) engages displaceably into the driving pin (88) of the pump actuator.
9. Capacity and/or moment regulating device according to claim 8,
   characterized in that the driving pin (88) has a recess (86) in the region of the oblique face (86) of the valve sleeve adjusting piston (61) and/or of the intermediate piece (84) to enable the bolt element (87) to be applied against the oblique face (86) of the valve sleeve adjusting piston (61) and/or intermediate piece (84) enclosed by the driving pin (88).
10. Capacity and/or moment regulating device according to one of claims 4 to 9,
    characterized in that the restoring spring (81) acts upon the end of the valve sleeve (61) remote from the measuring spring arrangement (46).

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