CH713006A1 - Hydraulic drive for a working machine, in particular for a slewing gear. - Google Patents

Abstract

The invention relates to a hydraulic drive circuit for driving a work machine, in particular for driving a slewing gear, with at least one hydraulic pump (1) which is connected via at least one directional control valve (3) with at least one hydraulic motor (4) via two feed pressure lines (A, B) and a pressure limiting function for limiting the positive and negative acceleration of the hydraulic motor (4), wherein the pressure limiting function is realized by at least one pressure limiting system (20) comprising at least one main stage, which allows relief of a feed pressure line in the second feed pressure line when a limiting pressure is exceeded , And at least one electrically controllable precursor for adjusting the limit pressure has.

Description

Description: The invention relates to a hydraulic drive circuit for driving a work machine, in particular for driving a rotating mechanism of the work machine, with at least one hydraulic pump, which is connected to at least one directional control valve with at least one hydraulic motor via two feed pressure lines and a pressure limiting function to limit the positive or has negative acceleration of the hydraulic motor.

Such hydraulic drive circuits are used specifically in applications which, on the one hand, require the full drive power to be provided as quickly as possible, but on the other hand, a limitation of the maximum acceleration is necessary for safety reasons. A special application of such drive circuits is the drive of a rotating mechanism of a mobile machine, in particular an excavator. In order to maximize the material turnover, achieving the desired / required change in the angular position of the superstructure is of high to very high importance.

The slewing gear cycle is composed of three movement phases, namely an acceleration starting from a standstill, a rotation at maximum angular velocity and braking to a standstill. The desired goal of short slewing gear working cycles is achieved when the maximum angular speed achieved when swiveling is as high as possible and secondly the time period until this maximum speed is reached is as short as possible and thirdly the time period within which the braking from the maximum angular speed to the slewing gear stop is as short as possible is. As a result, the slewing gear must be operated with high positive and negative angular accelerations, but on the other hand, the absolute values of the angular accelerations and angular velocity occurring must not be arbitrarily high in order to avoid overloading certain components actuated by the slewing gear. To avoid such overloads, there are various system solutions based on pressure relief valves.

Also, the centrifugal and inertial forces acting on the operator of the machine due to the rotational movement must be in an acceptable range. The constantly changing angular velocities of the superstructure with the cabin can lead to a significant burden on the well-being of the operator or driver, which is why it is also necessary to limit the angular accelerations that occur due to the use of pressure relief valves. In practice, the necessary limit is higher than the technical limit. In the further course, measures will be discussed that affect this further limitation.

Another problem with such applications is the standstill counter reaction. Due to the very high total mass moment of inertia of the driven components, when the slewing gear is braked to a standstill, the swinging movement decays with counter-rotating movements, which is due to the compressibility of the hydraulic oil used and, above all, the elasticity of the hydraulic lines used.

An example of a hydraulic drive circuit for a slewing gear known from the prior art is shown in Fig. 1. The circuit comprises an adjustable hydraulic pump 1, the negative pressure side of which is connected to the hydraulic oil storage tank 2 and the upper pressure side of which is connected to a controllable connection of a 4/3-way valve 3. The directional control valve is connected to the hydraulic motor 4 via the two valve outputs via two feed pressure lines A, B. The different switching states of the directional control valve 3 enable a bidirectional rotary movement or a complete decoupling of the hydraulic motor 4 from the variable displacement pump 1.

The control inputs of the directional valve 3 are indirectly coupled to the control lever for operating the hydraulic drive or the slewing gear. The deflection of the control lever leads to a corresponding height and direction of the acceleration acting on the slewing gear. A trivial pressure-limiting function to protect the hydraulic components from pressure overload is implemented by means of the pressure-limiting valves 5, 6, which connect the two feed pressure lines A, B and open them from a defined limit pressure and provide pressure relief.

Fig. 2 shows on the basis of the qualitative oil pressure time curves in the two feed pressure lines, the previously explained slewing gear cycle with the different movement phases. Phase I shows the acceleration of the slewing gear. As of a fixed pressure level, the pressure relief valve 5 opens on the A side, creating a bypass to the slewing gear hydraulic motor 4. During the decaying pressure pulsation, the pressure relief valve 5 changes between the open and closed states until a fixed pressure value is reached. As soon as the slewing gear has been accelerated to the desired maximum angular velocity, the required hydraulic nominal power becomes lower, since only the hydraulic power required to compensate the power loss has to be supplied to the slewing gear motor 4. In the diagram in FIG. 2, this is represented by phase II, in which the working pressure in line A drops to a comparatively low pressure level.

The dashed line in the movement phase III denotes the pressure curve in the feed pressure line B during the braking process. This pressure increase is based on the fact that the hydraulic motor 4 in this case works in the pump mode and, due to the high rotational energy at the beginning, ensures an increasing pressure level in the pressure line B. From a certain pressure level, the pressure relief valve 6 provides relief for the pressure line B into the pressure line A, pressure fluctuations here also being caused by an alternating opening or closing of the pressure relief valve 6. Movement phase IV characterizes the standstill counter-reaction. In this case there are counter rotating movements of the slewing gear due to the high moment of inertia and the

CH 713 006 A1

Elasticity of the feed pressure lines. The elasticity of the feed pressure lines and the compressibility of the hydraulic oil act as a parasitic pressure accumulator, which stores part of the rotational energy when the slewing gear is braked. The stored energy is returned to the hydraulic motor as hydraulic power and leads to a noticeable acceleration of the slewing gear in the opposite direction of rotation.

Fig. 3 shows a further development of the system shown in Fig. 1, which is also known from the prior art. Here, so-called shockless pressure relief valves 7, 8 are used instead of the conventional pressure relief valves 5, 6 in order to achieve a greater limitation of the slewing gear acceleration in the interests of the operator. In the built-in shockless pressure relief valves 7, 8, the opening movement begins at a significantly lower pressure level than in the usual pressure relief valves 5, 6. In keeping with its name, the pressure fluctuations that occur in the shockless pressure relief valve 7, 8 are markedly less pronounced. However, the time required to reach the final state of the open valve is significantly longer. A good compromise must therefore be chosen between high handling performance and comfort requirements for the operator. A second disadvantage when using shockless pressure relief valves is that a change in the opening pressure or the valve characteristics is not possible during operation. A shockless pressure relief valve is usually selected when a comparatively hard response behavior when accelerating and braking is preferred for the particular application of the work machine. In addition, the hydraulic circuit according to FIG. 3 also contains an additional anti-reaction valve 9 for damping the standstill counter-reaction. The valve 9 assumes its central position in order to permit bidirectional pressure compensation between the pressure lines A, B via the throttle, provided 1) the pressure differences in these two pressure lines are small and b) the changes in this pressure difference take place comparatively slowly. As a result, the amplitudes of the vibration representing the standstill counter-reaction in phase IV are reduced, so that this is perceived far less by the operator.

Fig. 4 shows a further embodiment known from the prior art, in which proportional pressure relief valves 10, 11 are used and these can be operated, for example, as so-called 2-threshold pressure relief valves 10, 11. These 2-threshold pressure relief valves are acted upon by the same control pressure that is also applied to the directional valve 3. The pilot control of the pressure limiting valves 10, 11 used, which is linked to the operating lever, allows the limit pressure at which the pressure valves 10, 11 to open to be varied between two stages. With a slight deflection of the operating lever, the pressure limiting valve is operated with a comparatively lower limit pressure, while a greater limit pressure is set when the operating lever is strongly deflected. It is also disadvantageous here that the operating characteristics of the pressure relief valves installed subsequently, i.e. do not change or influence during operation.

It is therefore sought for an improved hydraulic drive circuit, with greater flexibility with regard to the operating behavior of the pressure limiting function.

This object is achieved by a hydraulic drive circuit according to the features of claim 1. Advantageous embodiments of the drive circuits are the subject of the dependent subclaims.

According to the generic hydraulic drive circuit is further developed such that a pressure limiting system is used to implement the pressure limiting function. The pressure relief system has a two-stage structure. It is composed of at least one main stage which, when a limit pressure is exceeded, relieves the pressure on one feed pressure line into the second feed pressure line. A preliminary stage can be controlled electrically in order thereby to be able to indirectly influence the limit pressure of the at least one main stage. As a result, the pressure limiting function can be flexibly expanded and modified using various pilot control characteristics. This offers the advantage that the pressure limitation function can be selected specifically for each application by the driver, taking into account his individual preference. It is advantageous, for example, if the selection change can be made conveniently via a human-machine interface of the work machine. In addition, certain advantages can be combined that do not occur in pairs in the solutions according to the prior art.

It is conceivable that the hydraulic drive circuit is an open or a closed hydraulic circuit. The built-in directional valve preferably allows any position of the control slide, whereby the level of the flow resistance along the passage edge is continuously adjustable.

It is particularly preferred if the at least one main stage comprises at least one proportional pressure relief valve which connects the two feed pressure lines to one another in the open state and consequently leads to a pressure relief.

In drive circuits for a direction of rotation of the hydraulic motor, a main stage for the pressure relief from one pressure line into the second pressure line is sufficient. If operation and acceleration limitation when operating with two directions of rotation are desired, a second main stage must be provided, which ensures pressure relief from the second feed pressure line into the first feed pressure line, while the first main stage provides pressure relief from the first feed pressure line into the second feed pressure line. In this case, both main stages can each comprise a unidirectional proportional pressure relief valve.

CH 713 006 A1 The opening behavior of the one or more main stages, in particular the proportional pressure relief valve of the respective main stage, can be influenced by the preliminary stage and consequently depends indirectly on the electrical control of the preliminary stage. It is particularly preferred if both main stages can be precontrolled by a common preliminary stage, i.e. the limit pressure of the respective main stage can be individually adjusted through the common preliminary stage. Alternatively, however, a separate, electrically controllable preliminary stage could also be used for each main stage. For the sake of simplicity, the following always speaks of a main stage, but the explanations apply without restriction to drive circuits with two or more main stages, which can be controlled in pairs by a common preliminary stage or, however, by means of separate preliminary stages.

According to a preferred embodiment of the invention, the opening and / or closing pressure of the proportional pressure relief valve for opening the through-channel of the at least one main stage can be set by the preliminary stage. The pilot-controlled pressure relief valve of the main stage opens its passage channel, for example, when the pressure level of the feed pressure line at the valve inlet of the main stage overcomes the closing force of the pressure relief valve of the main stage. The variation of the closing force influences the operating characteristics of the entire pressure limiting function.

It is particularly preferred if a first control input of the pressure relief valve of the at least one main stage, at which the pressure level of the feed pressure line is present, via at least one flow valve, in particular at least one throttle, with the valve input of the preliminary stage, which can also be a proportional pressure relief valve , connected is. Furthermore, the output of the flow valve or the throttle output is connected to a second control input of the proportional pressure relief valve of the main stage. The closing force of the main stage can be influenced via this second control input, in particular the pressure level applied there interacts with the restoring force of a valve spring. The pressure level at the second control input can be influenced by targeted control of the pre-stage.

Particularly in the case of a preliminary stage designed as a proportional pressure relief valve, depending on the degree of opening, a certain pressure drop is effected at the second control input of the main stage, as a result of which the closing force decreases and the limit pressure level of the main stage can be reduced. In contrast, when the preliminary stage is closed, the closing force is at a maximum with increasing pressure in the feed lines.

Furthermore, it makes sense to connect the control inputs of the main stage to the preliminary stage via at least one check valve, in particular via a spring-loaded check valve, in order to block the backflow from the preliminary stage to the control input of the main stage. This is particularly necessary if two main stages are confirmed by a common preliminary stage. The pressure difference between the control lines means that only one main stage pressurizes the preliminary stage, i.e. the pressure level at the second control input of this main stage is also at the valve input of the preliminary stage. An influence of this pressure level via the external circuit of the other main stage is prevented by another check valve.

The preliminary stage is expediently an electrically pilot-controlled proportional pressure relief valve, the preliminary stage being able to be controlled by means of a proportional magnet in accordance with a preferred embodiment. It is conceivable that the electrical control of the preliminary stage acts either against the closing force of the preliminary stage or with the closing force of the preliminary stage, i.e. in the case of a spring-loaded proportional pressure relief valve acts either against the spring restoring force or with the spring restoring force. As a result, either a positive or negative characteristic curve of the pre-stage can be generated by means of the electrical pre-control.

[0024] According to a preferred embodiment of the invention, the hydraulic drive circuit comprises at least one control device in order to correspondingly control the preliminary stage electrically, as a result of which a predefined behavior of the entire pressure limiting function can be implemented. In particular, the pre-stage can be controlled by the control device in order to implement a predefined characteristic curve of the pressure limiting system. It is particularly advantageous if the control unit has several characteristic curves available for changing the characteristic of the pressure limiting function, i.e. the characteristic behavior of the pressure limitation function can be flexibly changed by the application-specific selection of the appropriate characteristic.

Such characteristic curves represent, for example, the opening pressure of the pressure limiting function or the pressure drop across a proportional pressure limiting valve of the at least one main stage, depending on the position of the operating lever for actuating the drive. The electrical control capability of the preliminary stage can be used to generate any characteristic curves, for example a linear and / or parabolic curve. As a result, it is possible, depending on the application, to obtain a characteristic curve according to the shockless pressure limiting function known from the prior art and / or a characteristic curve curve of the 2-threshold pressure limiting valve and / or any other pressure limiting valve. Depending on the application and depending on the advantages of the operator, the desired characteristic of the pressure limiting function can be set flexibly during operation at the machine runtime, in order to choose focal points with regard to handling efficiency or, alternatively, working comfort.

In addition to a selection option of permanently stored characteristic curves within the control device, the possibility is also created to influence the user in a targeted manner by scaling the characteristic curves. Additionally or alternatively, it can also be provided that the user selectively se4 one or more parameters influencing the characteristic

CH 713 006 A1 ready or can change together to adapt the characteristic to his needs. It can also be provided that the user can freely define an individual characteristic curve via the control device and can save it for later use within the control device.

It is also conceivable that the control device is designed in such a way that it can manipulate at least one predetermined characteristic curve by superimposing one or more input variables and thus adapt the system to other system parameters. Suitable input variables for consideration within the control unit for the pressure limiting function are, for example, the speed of the hydraulic motor and / or the speed of the driven component, for example the speed of a rotating mechanism, and / or the pressure level within the supply pressure lines to the hydraulic motor and / or the pressure difference between the supply pressure lines and / or the switching position of the directional control valve and / or the pressure within the control line or control lines of the directional control valve and / or the sensor position of a control lever for drive control.

The inventive design of the hydraulic drive also has the advantage that, for example, an observer function can be implemented within the control unit, which can take into account any cross effects or environmental conditions that may have an influence on the drive control for the implementation of the pressure limiting function. Such influences can be compensated for or adapted to advantage, for example, by adapting the characteristic curve used. For example, in addition to the simple circular movement, the turning mechanism also changes its height during rotation when the slewing gear is tilted, so that this has an influence on the braking behavior. This condition could also be used to control the preliminary stage of the pressure limiting system. In addition, the current temperature of the hydraulic medium and the resulting viscosity of the hydraulic oil can also influence the pressure limiting function. Consideration of the sensor-sensed medium temperature can be used specifically to compensate for this cross-effect.

Furthermore, the occurrence of leaks in the drive system can be recognized by means of the observer function, which is ideally signaled to the driver in order at least to be able to avoid the occurrence of consequential damage.

Furthermore, the control device can be network-compatible so that it can be communicatively networked with other control devices or components of a work machine. For example, in the event of a fault in the working machine that is recognized by an external control device, a shutdown or drastic limitation of the maximum acceleration could also be triggered.

[0031] Furthermore, according to a particularly preferred embodiment, the control device and the pressure limiting system can be used to dampen a standstill counter-reaction when the hydraulic drive is braked. In contrast to the prior art, no additional anti-reaction valve is required for this, but this function can also be performed by the pressure limiting function according to the invention. To implement the function, it is advantageous to supply the control device with the sensor-detected rotational speed of the hydraulic motor or that of the driven machine component and / or the measured pressure level within the feed pressure lines to the hydraulic motor. Here, too, there is the possibility of maintaining different characteristic curves within the control unit for the user-side selection of the desired operating characteristic of the standstill counter-reaction function.

In addition to the hydraulic drive circuit according to the invention, the present invention also relates to a working machine, in particular an excavator, with a hydraulic drive circuit according to the present invention. The working machine is consequently distinguished by the same advantages and properties as have already been explained above with reference to the drive circuit according to the invention.

The drive circuit according to the invention is used particularly advantageously when driving a slewing gear between the upper and lower carriage of the mobile machine. However, it is also conceivable to use it as a hydraulic drive for the work machine, since similar drive conditions, i.e. there may be a quick provision of the maximum drive power with simultaneous limitation of the maximum acceleration and the need for a strong braking in the regular work cycle.

[0034] It is also conceivable to use the hydraulic drive circuit according to the invention for a winch drive of a work machine, preferably in stationary structures or ships. As is well known, the pull rope threatens to break if the initial acceleration is too strong.

The use of the functional principle is also conceivable as a damping function in cylindrical drives. Use as a fan drive is also conceivable in order to implement a speed setting via pressure control. This can be particularly useful for high-performance fans.

Further advantages and properties of the invention will be explained in more detail below with reference to an embodiment shown in the figures. Show it:

1: a conventional drive circuit for a slewing gear;

2: shows a diagram to illustrate the different operating phases of a slewing gear;

CH 713 006 A1

3: a further drive circuit known from the prior art for a slewing gear with a pressure limiting function;

4: a further drive circuit known from the prior art for a rotating mechanism with a pressure limiting function;

5: the hydraulic drive circuit according to the invention for implementing a flexible pressure limiting function;

FIG. 6: different types of characteristic curve which can be implemented by means of the drive circuit according to the invention from FIG. 5 and

7: shows a diagram to illustrate the flexible modification of the characteristic operating behavior of the pressure relief valve according to the invention.

As an alternative to the prior art, the use of a pressure limiting system is proposed according to the present invention, the operating characteristics of which can be flexibly adjusted to application-related needs via an electrical control. In this way, the pressure limiting function can be adapted through the availability of various pilot control curves. This offers the advantage that the pressure limitation function can be selected specifically for each application by the driver, taking into account his individual preference. The selection change is conveniently possible via the human-machine interface and also during a work assignment. In addition, certain advantages can be combined that do not occur in pairs in the solutions according to the prior art.

5 shows the circuit diagram of the hydraulic drive circuit according to the invention. The circuit comprises an adjustable hydraulic pump 1, the negative pressure side of which is connected to the hydraulic oil storage tank 2 and the upper pressure side of which is connected to a controllable connection of a 4/3-way valve 3. The directional control valve is connected to the hydraulic motor 4 via the two valve outputs via two feed pressure lines A, B. The different switching states of the directional control valve 3 enable a bidirectional rotary movement or a complete decoupling of the hydraulic motor 4 from the variable displacement pump 1.

The control inputs 3a, 3b of the directional control valve 3 are indirectly coupled to the operating lever 13 for operating the hydraulic drive or the slewing gear, the deflection of the operating lever 13 leads to the displacement of the control slide of the directional control valve 3. To limit the maximum acceleration of the hydraulic motor 4, the new electrically controllable pressure limiting system 20 is integrated into the circuit. This pressure limiting system 20 is composed of two main stages, each main stage comprising a pilot-operated proportional pressure limiting valve 21, 22. Both main stages 21, 22, i.e. Both proportional pressure relief valves 21, 22 can be precontrolled by a common preliminary stage, which is itself formed by an electrically controllable proportional pressure relief valve 23.

The first main stage 21 or the pressure relief valve 21 opens at an adjustable limit pressure to relieve the feed pressure line A in line B, while the second main stage 22, i.e. the pressure relief valve 22 can relieve the feed pressure line B into line A from a certain limit pressure.

The pressure relief valve 23 of the hydraulic preliminary stage can be controlled via an electrical actuator contained therein, for example in the form of a proportional magnet. As an output signal, the pre-stage generates a pressure level with which a control input 21a, 22a of the proportional pressure-limiting valves 21, 22 acting as the unidirectional main stage is applied. The output signal of the electrically controlled hydraulic preliminary stage has no influence on the passage behavior of the main stage located in the blocking direction. Clearly, a reversal of the pressure difference must be recognized in a sufficiently short time and the system must be able to act quickly enough to reverse the pressure difference.

The pressure level Pa present on the A side is present at the control input 21b of the main stage 21. The control input 21b is also connected via a throttle 21c to the further control input 21a of the main stage 21. The throttle output is also connected via a spring-loaded check valve 24 to the inlet 23c of the valve 23 of the hydraulic preliminary stage. The connection of the main stage 22 relates to the pressure level p _B present on the B side. With regard to this, the pin assignment of the main stage 22 corresponds to that of the main stage 21. The pressure level p _B present on the B side is present at the control input 22b. The control input 22b is also connected via a throttle 22c to the further control input 22a of the main stage 22. The throttle output is also connected via a spring-loaded check valve 25 to the inlet 23c of the valve 23 of the hydraulic preliminary stage. Consequently, both valves 21, 22 are connected to the preliminary stage 23 via the same valve input 23c.

To explain the operation of the main stages 21, 22, an overpressure on the A-side is assumed as an exemplary example; this has the relative value p _A. The relative pressure on the B side has the value 0. The main stage 21 is thus in the forward direction and the main stage 22 is in the blocking direction. Since the check valve 25 in

CH 713 006 A1

Is in the blocking direction, there is no pressure on the main stage 22. There is therefore no leakage or pressure loss due to the pressure level Pa existing on the A side.

Case 1 The passage channel of the hydraulic preliminary stage 23 is completely closed: there is no pressure drop along the throttle 21c, since the closed hydraulic preliminary stage 23 results in a significantly higher (virtually the entire) pressure drop from Pa to the pressure level of the hydraulic oil storage tank 26 results. Via the control input 21b, the control surface of the main stage 21 is reached by the pressure level Pa, which exerts a right-hand force on the control slide. A pressure level is applied to the control surface via the control input 21 a, which pressure level approximately reaches the value p _A and exerts a left-hand force on the control slide. The size ratio of the two control surfaces is balanced such that, up to a certain threshold value of the pressure level p _A, the compression spring 21 d on the right-hand side of the valve symbol keeps the passage of the main stage 21 closed. It is clear that this must open when a certain threshold value of p _{A is} exceeded, so that the bursting of a hydraulic line due to overpressure is avoided in any case, even in the event of incorrect operation. Falling below the threshold value closes the main stage 21.

Case 2 The passage channel of the hydraulic preliminary stage 23 has assumed its maximum opening width. The entire pressure level p _A drops along the throttle 21c. At the inlet 21a there is a relative pressure level of 0 bar. Consequently, a comparatively low pressure level Pa is sufficient to overcome the restoring force of the compression spring 21 d until the main stage 21 releases an oil connection from the A to the B side.

Case 3 The passage channel of the hydraulic preliminary stage 23 has assumed a certain arbitrary position between the minimum and maximum opening width. The throttle 21c and the passage channel of the hydraulic preliminary stage 23 form a pressure divider between the pressure p _A present on the A side and the relative pressure of 0 bar. While the pressure level Pa is supplied to the control input 21b, the control input 21a is acted upon by a pressure level reduced by the pressure loss present at the throttle 21c. As a result of the pressurization of these two control inputs 21a, 21b, a force counteracts the restoring force of the compression spring 21d. If this force reaches a certain threshold value, the main stage 21 is opened.

The actuation of the preliminary stage 23 can be implemented in different ways.

(i) Use of an electrically pilot-controlled proportional pressure relief valve with a positive characteristic curve. The compression spring 23d of the valve 23 exerts a force on the control slide of the valve 23 in the direction “reduce the opening width of the control edge”. If there is a relative pressure level of or approximately 0 bar at the hydraulic control input, the control slide is pressed against an end stop by the restoring force of the prestressed compression spring 23d and is thus fixed in a position in which there is a closed control edge. The electrical control of the hydraulic preliminary stage 23 takes place by energizing a proportional magnet 23b. As the electrical current increases, the force emanating from the tappet of the proportional magnet 23b increases, which supports the restoring force of the compression spring 23d. This sum of forces is counteracted by a force which is exerted by the oil pressure at the hydraulic control input 23a of the preliminary stage 23.

The control surface and the compression spring 23d of the preliminary stage 23 are designed in such a way that the control slide is moved at a comparatively low pressurization at the hydraulic control input 23a without the involvement of the proportional magnet 23b. With increasing current supply to the proportional magnet 23b, the pressure level required for such a movement of the control slide increases. An opening movement or an increase in the opening width of the corresponding control edge of the valve 23 leads to the onset or an increase in the oil flow via the passage channel of the preliminary stage 23 into the tank return 26 and consequently leads, depending on which of the two main stages 21, 22 in the forward direction is located, to lower the control pressure at the control input 21a or 22a. The hydraulic preliminary stage 23 is thus self-regulating.

(ii) Use of an electrically pilot-controlled proportional pressure relief valve with a negative characteristic curve In this arrangement, the energization of the proportional magnet 23b leads to a force which counteracts the restoring force of the compression spring 23d. The control surface and the compression spring 23d of the preliminary stage 23 are designed such that, without the proportional magnet 23b cooperating, the control slide is only moved when the hydraulic control input 23a is subjected to a comparatively high pressure level. With increasing current supply to the proportional magnet 23b, the pressure level required for such a movement of the control slide decreases. All of the valves contained in the feed lines A, B of the slewing gear hydraulic motor 4 can be used as seat valves and thus

CH 713 006 A1

Leakage-free valves are designed. Implementing this option again offers the great advantage of reducing the oil leakage to that of the slewing gear hydraulic motor 4.

The specification of the electrical control of the hydraulic preliminary stage 23 is implemented by its own software, which is implemented on its own control unit 27 or as additional software on a control unit already present in the mobile machine. An important input variable for the software is the positioning of the joystick 13 that is relevant for the rotary movement. The joystick position is digitized at the beginning in relation to the signal path present in the control unit 27. Various characteristic curves are stored in the parameter set of the software. A superimposition of the digitized joystick position and the characteristic curve selected by the operator / driver of the mobile machine takes place in the control unit and is influenced by further input variables and is predetermined by the software. A driver stage, which can provide the necessary electrical power for energizing the proportional magnet 23b, is controlled by a digital-analog conversion of the superimposition signal.

The following can serve as further input variables in the control device 27 for controlling the preliminary stage 23:

- The speed of the slewing gear hydraulic motor 4

- the speed of the superstructure

- The pressure values in the hydraulic feed lines A, B of the slewing gear hydraulic motor 4 on the A and B sides

- The pressure value measured at its outlet in the presence of a shuttle valve 29 installed between the A and B sides

- The slide position of the 4/3 way valve 3 In the embodiment of the invention shown in FIG. 5, the two connecting lines A, B of the rotary hydraulic motor 4 are connected to the inputs of a shuttle valve 29. A pressure sensor 28 is installed at its outlet, which is acted upon by the (in terms of amount) higher pressure level of the two connecting lines A, B of the slewing gear hydraulic motor 4. The output value of the pressure sensor 28 is made available to the control unit 27 via a corresponding signal feed. With this arrangement, only a single pressure sensor 28 is required.

By the software of the control unit 27, the speed of the slewing gear hydraulic motor 4 or that of the uppercarriage and the position of the 4/3 directional control valve 3 are provided, taking into account their sign, algorithmic conclusions can be drawn unambiguously whether that of the software available pressure level is on the A or B side of the connecting lines of the slewing gear hydraulic motor 4. The necessary speed and position transmitter and the corresponding signal connections to the control unit are not shown in FIG. 5.

Due to the software-based setting option of the valve characteristic of the pressure-limiting function, the invention offers the advantage that the advantages of different types of pressure-limiting valves can be combined. The use of shockless pressure relief valves is preferred when a comparatively hard response when accelerating and braking is desirable for the particular use of the mobile work machine, i. H. is not uncomfortable for the driver / operator. With the proposed implementation of the pressure limitation by means of the electrically controllable pressure limitation system 20, in which a main stage is carried out by a proportional pressure relief valve based on a seat valve, the response behavior of a shockless pressure relief valve can be simulated by the invention, but without the disadvantage of a shockless pressure relief valve existing oil leakage.

The "shockless characteristic" is imprinted as follows: As soon as the pressure level at which the movement of the valve spool of the valves 21, 22 is to be reached is reached, the control unit 27 triggers / triggers the superimposition of the joystick position Corresponding digital signal with the pressure-time curve "shockless characteristic" as a characteristic. Such a combination is possible by superimposing the signal emanating from the joystick 13 with a characteristic curve stored in the control unit 27, which has the behavior of a shockless pressure relief valve. The superposition itself is triggered when the pressure limit has reached a state in which a reaction of a real shockless pressure limit valve would occur.

By means of the digital data processing carried out in the control device 27, the pressure limitation system 20 can be electrically controlled such that there is, for example, a linear scaling between the joystick position and the opening pressure of the valves 21, 22. This is also possible when countering, i.e. by deflecting the operating lever 13, which triggers an acceleration opposite to the current direction of rotation of the uppercarriage, in order to force the braking of the slewing gear to a greater extent. Alternatively, different scales could be used for the position of the joystick 13 and the opening pressure, e.g. a characteristic curve resembling a parabola or a hyperbola is available. See, for example, FIG. 6. The diagram shows possible adjustable characteristic curves that can be implemented by the electrically controllable pressure limiting system 20. The characteristic curves represent the pressure drop along the passage channel of the valves 21, 22 as a function of the joystick position or the resulting control pressure at the directional control valve 3. The characteristic curve a symbolizes the behavior of a conventional 2-threshold pressure relief valve, which is also related to the pressure relief system 20 of the invention can be simulated. However, the actual characteristic curve is much steeper. Curve b shows a linear characteristic curve, while curve c represents a parabolic characteristic curve. The pressure p1 represents the threshold value above which the valve 21, 22 located in the forward direction is automatically opened in the pressure limiting system 20. Except in the event that the printing area 8

CH 713 006 A1 limit system 20 to emulate the characteristic of a 2-threshold pressure relief valve, the opening pressure p can be continuously specified by the respective position of the joystick 13.

In a further development, a pilot control curve could first be stored, the rotation of which, when used, behaves as if the pressure limiting function was implemented by a real shockless pressure limiting valve, which the driver / operator perceives as particularly pleasant. Building on this basis, various virtual shockless pressure relief valves could be implemented in the same way by adding further characteristic curves. In a further development, the specification of the opening pressure of the selected virtual shockless pressure relief valve could be determined by different scaling of the joystick position.

In a further development, it is not only possible to select one from a set of different shockless pressure relief valves, each of which is clearly predetermined, but certain individual parameters of the characteristic curve can be varied, regardless of whether such a pressure-time profile can be achieved with a real pressure relief valve is. The curve drawn with a thin line in FIG. 7 shows a simplified, schematically illustrated pressure-time curve of a shockless limiting valve. The pressure-time curve shown there with a thick dashed line is not achievable with the real pressure relief valves mentioned so far, but would be possible with the pressure relief system 20 according to the invention by entering the following two parameters:

a) the pressure from which an opening process of the simulated shockless pressure relief valve begins is designated in FIG. 7 as pressure p ₀

b) the pressure curve as a function of the time with which the opening process takes place. In Fig. 7, this circumscribed section of the characteristic curve can be defined by specifying the pressure increase over time (i.e. the derivation of the pressure over time) and the duration of the increase.

Ultimately, such characteristics could also be created, the use of which results in a response behavior of the slewing gear, which would not be possible at all by using only hydraulically controlled valves or only by using valves whose manufacturing costs are very high.

If the characteristic of a pilot-controlled 2-threshold pressure relief valve is desired, this would be possible by selecting an appropriate characteristic (see FIG. 6, curve a). In contrast to the system, which contains real, purely hydraulically pilot-operated 2-threshold pressure relief valves, the opening pressures could be carried out by the operator / driver himself by selecting another or adjusting the characteristic directly at the driver's workplace on the human-machine interface.

There are three different procedures for the selection and provision of characteristic curves.

a) The driver / operator has a choice from a certain amount of clearly defined characteristic curves stored in the control unit 27.

b) as a) In addition, the driver / operator can change certain parameters of existing characteristics within predetermined limits.

c) Any characteristic curves can be defined and saved in the parameter set of the slewing gear software for later retrieval.

Further expansion stages and advantages of the invention The invention also offers the possibility of integrating or compensating for external influences on the drive system. Due to the possibility of sensory detection - by real sensors and / or virtual sensors, so-called observer functions - such over-effects and ambient conditions can be taken into account in the data processing carried out in control unit 27. On the one hand, this allows the influence of such cross effects to be compensated and, on the other hand, the desired influence of the ambient conditions on the slewing gear movement can be achieved.

Example of an environmental condition, the consideration of which for the slewing gear movement is advantageous. In a further development, an observer function implemented in the control unit 27 can take into account the early stages of the slewing gear angular velocity and the oil pressure level on the A and B sides of the slewing gear brake operation IV can be deduced from the energy currently to be dissipated. This energy is made up of the rotational energy and any change in the positional energy. A relevant change in the position energy is present when the mobile machine is in an inclined position.

As previously explained, the rotational energy is dependent on the bulk loading and the position of the implement - in the case of an excavator these are the angular positions of the stroke, the style and the bucket - the direct determination of which would be very complex during a work assignment. Through the influence of a possible situation energy situation9

CH 713 006 A1 change during the braking process - depending on whether the respective slewing gear movement includes an upward or downward movement - this must either be removed from the system in addition to the rotational energy or this causes an additional braking of the rotary movement. In the case of the power supply to the slewing gear hydraulic motor 4 required to accelerate the slewing gear or to maintain the slewing gear angular velocity, the situation is correspondingly the same when the mobile working machine is in an inclined position.

Example of a cross effect, the influence of which is advantageously compensated for.

In a further development, the corresponding hydraulic oil temperature can be measured and its value can be taken into account in the pre-control characteristics of the anti-reaction function. As is known, the oil temperature has an influence on the oil viscosity and thus on the flow resistance in the hydraulic system. The consequence of this is that in the main stage, with the same actuation of the control slide, a lower oil mass flow flows through the passage duct than is the case with the hydraulic oil that is at the operating temperature.

Networking of the pressure limitation function In combination with other monitoring, error detection and safety functions present in the mobile work machine, the choice of the pilot control characteristic or its form could be influenced. Alternatively, a further signal could be calculated or taken from another map. The output signal relevant for the electrical control of the hydraulic proportional pressure relief valves could then be generated by superimposing the following signals:

a) The digital signal corresponding to the joystick position

b) The characteristic curve chosen by the operator / driver for determining the slewing gear response characteristic

c) Another correction characteristic curve formed by the recorded environmental conditions and Ouereffects

Coordination of the pressure limiting function with the operation of the hydraulic pump 1 in the control unit. If the driver / operator moves the joystick 13 into a position, this information can, for. B. the control unit of the diesel engine, so that there is already proactively an increase in the fuel injection is carried out and this is not left only afterwards because of a drop in speed.

Networking the pressure limiting function could also have the effect that, in certain error cases, acceleration of the slewing gear is generally prevented by the software.

[0070] The combination of a specific pressure limitation characteristic and a specific slewing cycle could trigger pressure pulsations. The pressure monitoring in the feed lines A, B of the slewing gear hydraulic motor 4 could be used in order to inform the driver of this by means of a display provided for this purpose. Alternatively, the software of the control unit 27 could autonomously make a small change in the pressure limitation characteristic, which is not / hardly perceptible to the driver, in order to prevent the pressure pulsations from continuing.

Detection and reaction to a leak on the A or B side. Assuming that the slewing gear is braked, this results in a pressure build-up in the hydraulic line on the A side. If there was a leak in the oil connection on the A-side in the event of a fault, hydraulic oil would emerge from the system under consideration virtually unimpeded and the slewing gear hydraulic motor 4 operating in pump mode would continue to draw hydraulic oil from the B-side and supply it to the A-side. As a result, the opening pressure of the pressure relief valve 21 concerned could never be reached.

With a corresponding expansion of the observer function, this fault or the leakage can be detected. For example, a corresponding leakage can be inferred from the angular velocity of the slewing gear and the knowledge that there is no pressure rise on the A side or a much smaller pressure rise than the expected pressure despite the valve 21 being closed. In this case, the opening pressure of the valve 21 could be reduced in order to avoid the occurrence of the direct damage to the hydraulic motor 4 and to indicate the presence of the leak to the driver.

Damping of the standstill counter reaction by the electronically / electrically controlled

Pressure limitation system 20 A standstill counter reaction is composed of several successive, qualitatively identical individual reactions. Due to the anti-reaction function of the control unit 27, when the first individual reactions of a braking process occur, the predominant portion of the energy leading to the triggering of the counter-reaction at a standstill is dissipated. Accordingly, the first individual reaction is only in a weakened form. Although weakened further by the occurrence of sustained power losses - such as mechanical friction and flow losses, etc. - there is still a certain amount of energy in the system under consideration, which is a second individual reaction

CH 713 006 A1 causes etc. Complete damping without the occurrence of a standstill counter-reaction - which is referred to in physics as an aperiodic limit case - is physically and technically possible, but would require a comparatively very long braking time and is therefore not practical for implementation , The standstill counter reaction corresponds to a decaying vibration. The respective vibration amplitudes are significantly reduced by the presence of an anti-reaction functionality.

The state of motion of the slewing gear can be tracked by means of a sensor system which detects the slewing gear angular velocity and the oil pressures on the A and B sides. Braking of the slewing gear as such can thus be recognized by the control software of control unit 27 and the time of occurrence of the standstill counter-reaction or of the individual reactions can be predicted in good time. This enables a suitable electrical control of the preliminary stage 23, which triggers the opening-closing movement sequence of the pressure-limiting valves 21, 22, which is required to dampen these torsional vibrations. A coordinated opening-closing movement sequence of the pressure limiting valves 21, 22 then results, taking into account the instantaneous oil pressure values on the A and B sides.

An individualization of the damping of the standstill counter-reaction is possible by storing various pre-control characteristics in the control unit 27 for the anti-reaction function as well. From this, the operator / driver of the mobile work machine can select the characteristic line he prefers, depending on which the electrical control of the hydraulic preliminary stage 23 takes place, which in turn specifies the opening-closing movement sequence of the pressure limiting valves 21, 22 and thus its influence on the damping counteracting the standstill.

In addition, it is not only possible to fine-tune the output variables, but also to use the variability of the detection features, which can be individually determined by fine-tuning. In the case of an electronically triggered antireaction function based on a pre-control characteristic curve, the influence thereof can be specifically excluded for each individual variable or consciously acting in a predefinable dependency. On the other hand, when using a real anti-reaction valve 9, the detection features are predetermined by its construction, e.g. B. by the throttles at the control inputs, the size of the control surfaces, the moment of inertia of the slide, etc. (see. Fig. 3, 4). Thus, a real anti-reaction valve 9 only reacts to the primary size, the pressure difference between the A and B sides and to the difference as to whether the change in pressure difference between the A and B sides changes quickly, as if the joystick changed position 13 or comparatively slowly as when the standstill counter-reaction occurred. The effect of parasitic variables on the reaction of the anti-reaction valve 9 is also determined by its design and cannot be individualized.

The invention can also be used for other hydraulically driven implements and, in particular, for hydraulic travel drives, since here too, on the one hand, it should be possible to utilize the available drive powers under suitable ambient and boundary conditions, but the positive and negative accelerations must also be limited depending on the situation ,

Other applications:

[0078]

- Winch drives in mobile machines, in stationary structures, in ships. (As is known, if the initial acceleration is too strong, the pulling rope may break.)

- Damping function for cylindrical drives

- Speed specification for fans via pressure control (In the case of high-performance fans, it is advantageous to use a pressure control valve that is electrically pilot-controlled as described.)

Claims (15)

claims 1. Hydraulic drive circuit for driving a working machine, in particular for driving a slewing gear, with at least one hydraulic pump, which is connected via at least one directional valve to at least one hydraulic motor via two feed pressure lines and has a pressure limiting function for limiting the positive and negative acceleration of the hydraulic motor, characterized that the pressure limiting function is implemented by at least one pressure limiting system, which comprises at least one main stage, which enables a supply pressure line to be relieved into the second supply pressure line when a limit pressure is exceeded, and has at least one electrically controllable preliminary stage for setting the limit pressure. 2. Hydraulic drive circuit according to claim 1, characterized in that at least one main stage comprises at least one proportional pressure relief valve, which connects the two feed pressure lines in the open state. 3. Hydraulic drive circuit according to one of the preceding claims, characterized in that a main stage is provided per direction of rotation of the hydraulic motor in order to enable pressure relief from the first to the second feed pressure line or from the second to the first feed pressure line, the opening behavior of the main stages preferably by a common preliminary stage can be influenced. CH 713 006 A1 4. Hydraulic drive circuit according to claim 2 or 3, characterized in that a control input of the at least one main stage, in particular the proportional pressure relief valve of the at least one main stage, is connected to the valve input of the preliminary stage via at least one flow valve, in particular at least one throttle, the Throttle output is connected to a second control input of the at least one main stage, in particular the proportional pressure relief valve, and wherein the second control input interacts with the closing force of the main stage, and the pressure level at the second control input is preferably controllable via the preliminary stage. 5. Hydraulic drive circuit according to one of the preceding claims 2 to 4, characterized in that the control input or inputs of the at least one main stage are connected to the preliminary stage via at least one non-return valve, in particular a spring-loaded non-return valve, in order thereby to reflux from the preliminary stage to block the control inputs of all main stages connected with such a check valve. 6. Hydraulic drive circuit according to one of the preceding claims, characterized in that the preliminary stage comprises at least one electrically pilot-controlled proportional pressure relief valve which relieves the control input of the at least one main stage depending on the degree of opening of the preliminary stage, the preliminary stage being preferably controllable by means of a proportional magnet, and the electrical one Control either increases the closing force of the preliminary stage or alternatively acts against the closing force. 7. Hydraulic drive circuit according to one of the preceding claims, characterized in that a control device for electrically controlling the preliminary stage is provided in order to implement a predefined behavior of the pressure limiting function, in particular in order to implement a predefined characteristic curve of the pressure limiting system, the control device ideally having a plurality of characteristic curves Changes in the characteristics of the pressure limiting function ready. 8. Hydraulic drive circuit according to claim 6, characterized in that at least one characteristic curve characterizes the position of the operating lever for actuating the drive as a function of the opening pressure of the pressure limiting function, the characteristic curve profile being linear, square or parabolic, for example. 9. Hydraulic drive circuit according to claim 6 or 7, characterized in that an operating means is provided to select a characteristic curve for the operation of the drive circuit from a selection of characteristic curves stored in the control unit of the drive circuit and / or to vary the scaling of the characteristic curve and / or to change one or more parameters influencing the characteristic curve, preferably within predetermined limits, and / or to freely define an individual characteristic curve and to save it for later use. 10. Hydraulic drive according to one of claims 6 to 8, characterized in that preferably at least one characteristic curve simulates the behavior of a shockless pressure relief valve and / or at least one characteristic curve simulates the behavior of a 2-threshold proportional pressure relief valve. 11. Hydraulic drive circuit according to one of claims 6 to 9, characterized in that the control device is designed such that at least one characteristic curve can be manipulated by superimposing one or more input variables, the input variables preferably being the speed of the hydraulic motor and / or the speed of the driven component and / or the pressure values in the pressure lines to the hydraulic motor and / or the pressure difference between the pressure lines and / or the switching position of the directional control valve and / or the pressure within the control line or control lines of the directional control valve and / or the sensor-sensed position of a control lever to operate the drive can be fed to the control unit. 12. Hydraulic drive circuit according to one of the preceding claims, characterized in that the control device has an implemented observer function in order to be able to compensate for cross-effects and / or ambient conditions influencing the pressure-limiting function on the basis of sensor-determined operating conditions, for example the temperature of the hydraulic medium and / or that Leakage in the drive system and / or changes in the positional energy of the driven component can be recognized by the observer function of the control device and the pressure limiting function can be modified accordingly. 13. Hydraulic drive circuit according to one of the preceding claims, characterized in that the control device can be networked with other components or control devices of a work machine in order to trigger a safety shutdown of the hydraulic drive by targeted pressure relief, in particular by a, when certain events, in particular an error, occur to prevent further acceleration of the drive. 14. Hydraulic drive circuit according to one of the preceding claims, characterized in that the control unit controls the pressure limiting system in such a way as to achieve damping of the standstill counter-reaction when braking the hydraulic drive, in particular taking into account the rotational speed or the translatory speed of the driven machine component and the the prevailing pressures in the feed pressure lines to the hydraulic motor, ideally different characteristics that can be selected by the user for adjusting the damping behavior are present in the control device. 15. Working machine, in particular excavator, with a hydraulic drive circuit according to one of the present claims, in particular for driving a slewing gear between the superstructure and / or undercarriage and / or as a drive for the working machine. CH 713 006 A1 FIG. 1 State of the art CH 713 006 A1