EP0229330B1 - Swing gear driven by two separate cylinder-crank mechanisms - Google Patents

Description

The invention relates to a swivel mechanism with two separate cylinder crank drives as a drive, both swivel cylinders working double-acting and having a piston rod led out on one side to form a differential piston arrangement and a swivel angle of less than 360 degrees, as for use in crane and excavator systems, have and the dead centers of the cylinder crank drives offset from one another with respect to the swivel cylinder and the swivel cylinders can be actuated by means of a preferably electro-hydraulic control by a pump via control elements.

Swivel mechanisms of this type are able to cover a swivel range of 180 ° and more. For this you need a special design of the geometry and the hydraulic control if you want to ensure increased demands on the uniformity of the swiveling speed under all loads as well as safe entry into the end positions. Such swivel mechanisms are known according to DE-A-2 443 760, in which the cylinders are actuated from one side only when driving through a swivel direction, for example a cylinder over a certain swivel angle range from the rod space and the other cylinder over another swivel angle range from the floor space forth. Both areas partially overlap.

The control takes place either via a control valve for both cylinders with corresponding black-and-white distributor valves or one control valve per cylinder. The cylinder spaces that are not actively acted upon are generally activated by externally controllable valves, so that they can suck in oil from the oil reservoir or push oil back into this reservoir or exchange the oil between the two cylinder spaces. The suction function can of course also be achieved by a simple check valve between the cylinder and the oil tank.

Other swing mechanism controls additionally use one of the known load pressure sensing methods according to DE-A-3 413 866, in which the load pressure of each cylinder is fed back to a separate control element, such as a variable displacement pump or pressure compensator, which adjusts the pump pressure so that it is a constant amount higher than the load pressure. These controls are very complex.

Controls according to DE-A-2 611 085 have therefore already been proposed, in which a common control element is used for both cylinders, so that both cylinders are supplied with the same system pressure. Such a solution leads to non-uniform swiveling speeds, even if the dead center areas of the cylinders are excluded, as by activation.

It has been shown that the swiveling process cannot be adequately mastered with such swivel mechanism controls, in particular the tying up of a driving load at high swiveling speeds is not guaranteed. In addition, the oil can not be sucked in satisfactorily in the activated cylinders.

According to the aforementioned DE-A-2 443 760, a so-called steam engine reversal for a slewing gear has been proposed which, in deviation from the slewing gear, can perform any number of revolutions in succession. These known arrangements are not transferable for generic swiveling mechanisms, since they cannot be used for sensitive control of a swiveling movement. In addition, with this known design it is necessary to create a spatial unit between the cams arranged on the rotating gear shaft and the reversing valves. Therefore, there is also a lack that the valves are designed for the full flow and must therefore be large.

The object of the invention is to provide a swivel mechanism of the generic type with a simple swivel mechanism control, which supplies the swivel cylinders with oil in all operating states sufficiently and enables a uniform swivel movement and also ensures that a driving load is tied up at high swivel speeds.

This object is achieved according to the invention in that the rod spaces of the two swivel cylinders in the swivel areas in which they absorb oil are subjected to the full pump pressure and both cylinder spaces of the swivel cylinder are assigned flow restrictors and for switching over the pump or return pressures when the pistons of the swivel cylinders pass through the dead center sensors, and that when swiveling the swivel mechanism from one end position to the other end position or vice versa, the floor spaces of the cylinders in the swivel areas in which they both absorb oil simultaneously, either with a reduced pressure are acted upon, or that only one floor space is acted upon by the full pump pressure in these pivoting areas, while in the other cylinder, when the bottom dead center is reached, both cylinder spaces are connected to the return.

This ensures that the swivel cylinders are supplied with oil in all operating states and the oil flow is not stopped. An optimal configuration is achieved in that both pivot cytinder can be switched in the manner mentioned.

It is also proposed that a load pressure feedback signal is formed from the load pressures of both swivel cylinders and a control element (adjusting device of the variable pump) load pressure-independent speed control is supplied. It is thereby achieved that an active driving through of pivot cylinder dead centers with a corresponding switchover does not lead to the feared load pressure peaks and can therefore be carried out, even if both cylinders remain pressurized up to the dead centers. Of course, load-sensing methods are generally known, as is known from DE-A-3 413 866.

In order to avoid that both bottom sides of the swivel cylinder are pressurized with pressure oil when moving into the end positions and that this results in an excessively high braking pressure being generated on the rod sides because of the pressure-transmission ratio between the bottom and rod sides and the flywheel masses involved in fast swivel movements that the reduced pressurization of the floor spaces of both swivel cylinders can be carried out by a switchable pressure limitation of the load pressure feedback signal.

In order to brake the entry into the end position on the rod side of the swivel cylinder, it is proposed that the swivel cylinders be provided with arrangements for end position damping on their rod side. For this purpose, it is also provided that the rod spaces of the swivel cylinders are assigned lowering brake valves which can be controlled by the pressure prevailing in the base space of the cylinder in question.

Furthermore, the rapid swiveling behavior under load is positively influenced by the fact that throttle check valves with discharge throttling are assigned to the floor spaces of the swiveling cylinders.

Favorable conditions are achieved when using a load pressure sensing method in that each swivel cylinder can be controlled via an associated proportional valve, which has load pressure sensing connections and that the mean value of both sensed load pressures can be fed to the control element as a control variable.

Furthermore, it is alternatively also advantageous that the lower of the two sensed load pressures of the swivel cylinder serves as a control variable of the control element, provided that both swivel cylinders are activated.

A further alternative is that the higher of the two sensed load pressures of the swivel cylinders, reduced to a value proportional to the activation of the proportional valve, serves as the control variable of the control element.

• Fig. 1 eine Draufsicht auf ein Zwei-Zylinder-Schwenkwerk,
• Fig. 2 ein hydraulisches Ansteuerschema für ein Schwenkwerk,
• Fig. 3 und 4 zwei Funktionsablaufdiagramme,
• Fig. 5 ein Schaltschema für eine elektrische Ansteuerung der Proportionalventile.

In the drawing, an embodiment of the invention is shown schematically. Show it:

• 1 is a plan view of a two-cylinder slewing gear,
• 2 shows a hydraulic control diagram for a swivel mechanism,
• 3 and 4 two functional flow diagrams,
• Fig. 5 is a circuit diagram for electrical control of the proportional valves.

The two swivel cylinders 1 and 2 shown are fastened on the bottom side to a common pivot point 3 of an abutment 4. Your piston rods 5 and 6 are attached to a swivel platform 8 at points 9 and 10 arranged offset by an angle 7. Both swivel cylinders 1 and 2 are double-acting and can therefore be pressurized with pressure oil from the floor space side as well as from the rod side. They have internal end position dampers 11 on the rod side, which are intended to reduce the swiveling speed when approaching the end stops. The swivel cylinder 1 determines the one end position 8, the swivel cylinder 2 the other end position 8 '. The total pivot angle 12 corresponds to the pivot angles 13 and 14 of the articulation points 9 and 10 of the piston rods 5 and 6. It is therefore greater than 180 °. The bottom dead center of the swivel cylinder 1 is reached when the swivel stage center has reached point 15. The bottom dead center of the pivot cylinder 2 is correspondingly given at point 16 of the pivot stage 8.

At dead center 15, a transmitter 60 and at dead center 16, a transmitter 61 is fixed to the foundation. Both transmitters 60, 61 are actuated by a cam 62 attached to the swivel platform 8 during passage through the respective dead centers 15 and 16, respectively. The corresponding functions are described in connection with the explanation of the circuit diagram according to FIG. 5.

The pivot cylinders 1 and 2 are connected to a hydraulic system on the bottom side via lines 17 and 18 and on the rod side on lines 19 and 20 according to the control diagram in FIG. 2. This receives its hydraulic energy from a variable displacement pump 21, which sucks in oil from a tank 22 and supplies two 4/3 proportional valves 25 and 26 via lines 23 and 24. These are controlled by electrical sensors 60, 61, 63. Each proportional valve 25, 26 actuates a swivel cylinder 1, 2. In the middle position, the oil inflows and outflows are blocked. When the valves are deflected to the left, the bottom sides of the cylinders 1 and 2 are pressurized with oil via lines 17 and 18 and the rod sides are connected to a return line 27 via lines 19 and 20.

When the proportional valves 25 and 26 are deflected to the right, the cylinders 1 and 2 are pressurized in the opposite direction.

In addition, assigned 4/2-way valves 28 and 29 with two fixed switching positions for stopping or starting the hydraulic flow enable the pivoting cylinders 1, 2. For example, the pivoting cylinder 1 is connected to the return line 27 on the base side and on the rod side when the valve 28 is electrical is activated.

Furthermore, load pressure sensing lines 30 and 31 determine which load pressure is present of the respective pressure line 17, 18, 19 or 20 on the cylinder side at the proportional valve 25, 26. This is transferred to the variable displacement pump 21 via a pressure change valve 32 and the line 33. This enables them to adjust their flow rate so that it remains approximately constant regardless of the load pressure. This means that the flow rate to the swivel cylinders 1 and 2 is load-independent and is only determined by the control of the proportional valves 25, 26.

The pressure change valve 32 has the effect that the higher of the two load pressures present at the valves 25 and 26 acts on the variable displacement pump 21. This would increase up to the maximum pressure of the variable displacement pump 21 - secured by the valve 34 - if both swivel cylinders 1 and 2 reached a dead center at the same time or if only one cylinder was connected to the pressure line 23 or 24 and reached the dead center. In these cases, the volume flow would be braked to zero and the variable displacement pump 21 would be informed of a high load via line 33, which it countered by increasing the delivery rate and maximum pressure increase.

This increase in pressure is avoided in the two-cylinder swivel mechanism with swivel cylinders 1, 2 working offset against one another if both cylinders 1 and 2 are pressurized at the same time and one cylinder runs through the dead center and the pressure supply in the bottom dead centers from the rod side to the Bottom side or vice versa with the help of proportional valves 25 and 26 and dedr encoder 60 and 61. In this case, the oil flow to the cylinder 1 pivoting through the dead center is reduced to zero. The oil flow to the other swivel cylinder 2 remains fully intact. The sensed load pressure at the swivel cylinder 1 or at the valve 25 cannot therefore increase appreciably beyond the load pressure from the swivel cylinder 2 present at the valve 26, because both valves are connected on the pressure side via the lines 23 and 24 to the same pressure source in the form of the variable displacement pump 21.

An additional possibility of reducing sensed load pressure peaks only at a proportional valve 25 or 26 is that the double check valve 32 is replaced by a valve which forwards the mean value from both sensed load pressures or the smaller of both load pressures to the variable displacement pump 21. A time-controllable reduction of high load pressure peaks is of course also possible with the aid of a pressure relief valve 36, which can be activated actively in critical cases by a 2/2-way valve 35.

The lowering brake valves 37 and 38 are intended to brake the entry of the swivel cylinders 1, 2 into the rod-side end positions in addition to the internal damping 11. They are unlocked when the bottom sides are depressurized. In this case, the oil on the rod side can only flow out via the restrictors 39 and 40.

The throttles 41 and 42 on the floor side serve to limit the swiveling speed under external load when the piston rods 5, 6 are inserted into the swiveling cylinders 1, 2. The throttles 41, 42 are bridged by the check valves 43, 44 when they are released by pressure that is simultaneously present on the rod sides (lines 19, 20).

3 shows the functional sequence for swiveling the swivel platform 8 into position 8 ', swivel direction according to arrow 45. Lines 46 and 47 indicate the dead centers of the cylinders 1 and 2 on the bottom. First, the rod spaces swivel cylinder 1 (48) and swivel cylinder 2 (49) pressurized up to dead centers 46 and 47, respectively, while the floor spaces swivel cylinder 1 (50) and swivel cylinder 2 (51) are connected to the return line 27. At the dead center 47, the pivot cylinder 2 is acted upon by pressure 51 'on the bottom side and 49' is relieved on the rod chamber side. The associated lowering brake valve 38 is unlocked by the pressure 51 '. At the same time, the pivot cylinder 1 remains connected to the pump pressure 48 on the rod side. The oil flow is not interrupted.

In the dead position 46 of the pivot cylinder 1, the rod chamber is depressurized 48 ', while the bottom chamber pivot cylinder 2 is still subjected to pressure 51' and maintains the oil flow from the variable displacement pump 21. The floor space pressure 50 and the rod space pressure 48 'of the swivel cylinder 1 remain activated from the dead center 46 by switching the associated proportional valve 25 to the middle position and the black-and-white valve 28 to the flow position. This relieves the pressure in line 17 and the lowering brake valve 37 switches to the blocking position. The oil running off on the rod side is braked by the throttle 39 (compare 52 in FIG. 3). The swiveling speed decreases and is braked to zero shortly before reaching the end position 8 'by the end position damping 11 additionally acting in the swiveling cylinder 2 (compare 53 in FIG. 3).

4 analogously shows the pivoting back process from the end position 8 'to the end position 8, arrow 59. In this case, both rod spaces are again subjected to pressure 54 and 56 in the areas in which they absorb oil. From the dead position 47 of the swivel cylinder 2, its bottom space pressure 57 is not kept at zero, as in FIG. 3, the pressure 50, but is applied with a reduced value 57 '. This can be done, for example, by activating valves 35 and 36, provided a load pressure sensing method is used.

At the same time, this measure also results in a reduction in the ground pressure 55 of the swivel cylinder 1 from the high value 55 'to the value 55 ", which is equal to the value 57' Damping pressure of the internal end position damping 58 (11 in FIGS. 1 and 2 for pivot cylinder 1) increases too high. This means that the swivel platform 8 can be braked to zero until the end position.

5, an electrical control of the two proportional valves 25 and 26 for the pivot cylinders 1 and 2 is shown. These are shifted into the positions corresponding to the control slide by a manually operated control slide 63 via control magnets 64 to 67. The electrical energy required for this is supplied by a battery 68. For example, in the position shown, a line 69 is connected to the negative pole, a line 70 corresponding to a wiper position 71 on the electrical resistor 72 is connected to part of the positive battery voltage. Accordingly, the actuating magnets 65 and 67 are activated and deflect the valves in proportion to the position of the control slide 63. 2, the bottom sides of both swivel cylinders 1 and 2 are pressurized with pressure oil. If the control slide 63 is moved into the position 71 ', the actuating magnets 64 and 66 are conversely acted upon with electrical energy via the resistor 72' and thus the rod side of the two swivel cylinders 1 and 2 are supplied with pressure oil.

If the swivel platform now moves through the bottom dead center of the cylinder 1, the transmitter 60 (see FIG. 1) is actuated and activates the magnet 65 and with it the floor space of the cylinder 1. With further swivel movement, the transmitter 61 is also actuated, so that the bottom space of the cylinder 2 is pressurized. The encoders 60 and 61 can be both digital changeover switches and a rotary angle encoder (analog encoder), the signal of which is dependent on the angle of rotation is processed in a downstream signal processing unit to produce the two required changeover signals for the proportional valves 25, 26.

In order to limit the swiveling forces in this case, one floor space (cylinder 1) can be relieved via the switch 73 by moving the valve 25 back into the neutral position and the magnet 74 of the valve 28 (FIG. 2) is excited. This connects the cylinder 1 on the floor and rod side to the return 27 and thus unlocks it.

Claims (9)

1. Slewing mechanism with two separate cylinder crank drives by way of drive, whereby both slewing cylinders (1, 2) operate in double- acting manner and possess a unilaterally extending piston rod (5, 6) so as to form a differential piston arrangement and exhibit a slewing angle of less than 360 degrees, as for use with crane and excavator systems, the dead centres (15, 16) of the cylinder crank drives being mutually offset in respect of the slewing angle and the slewing cylinders (1, 2) being so designed that they can be operated in meterable manner by means of a preferably electrohydraulic drive from a pump (21), characterised in that,

- each of the rod chambers of the two slewing cylinders (1, 2) can be subjected, within the slewing ranges, in which they take in oil, to the full pressure of pump (21), in that discharge throttles (39, 41 and 40, 42) are associated with the two cylinder chambers of slewing cylinders (1,2), and in that transducers (60, 61) are used for changing over the pump- or return pressures as the pistons of slewing cylinders (1, 2) pass through the dead centres,

- in that, as the slewing mechanism slews from one end position (8) to the other end position (8') or vice versa, the head chambers of cylinders (1, 2) are, within the slewing ranges in which both of them take in oil simultaneously, subject either to a reduced pressure (55"), or in that, within said slewing ranges, only one head chamber is subject to the full pump pressure (55'), whereas in the other cylinder (1) both cylinder chambers are connected with the return line (27) when the dead centre on the head-side (46) is reached.

2. Slewing mechanism according to claim 1, characterised in that from the load pressures of the two slewing cylinders (1, 2) a load pressure return signal is derived and supplied to a controller (displacement device of displacement pump 21) for toad-independent speed control.

3. Slewing mechanism according to claim 2, characterised in that the reduced pressure (55", 57) to which the head chambers of the two slewing cylinders (1, 2) are subject, can be brought about by controllable pressure limitation (35, 36) of the load-pressure return signal.

4. Slewing mechanism according to one of claims 1 to 3, characterised in that the slewing cylinders (1, 2) are provided on their rod sides with arrangements (11) for end position attenuation.

5. Slewing mechanism according to one of claims 1 to 4, characterised in that the rod chambers of slewing cylinders (1, 2) are provided with lowering brake valves capable of being controlled by the pressure within the head chamber of the cylinder in question.

6. Slewing mechanism according to one of claims 1 to 5, characterised in that the head chambers of slewing cylinders (1, 2) are provided with throttle-non-return valves (41, 43; 42, 44) with discharge throttling action.

7. Slewing mechanism according to one of claims 2 to 6, characterised in that each slewing cylinder (1, 2) can be controlled via an associated proportional valve (25, 26) with load-pressure sensing connections, and in that the mean value of the two load pressures sensed can be supplied to controller (21) by way of control value.

8. Slewing mechanism according to one of claims 2 to 6, characterised in that the lower of the two sensed load pressures of slewing cylinders (1, 2) serves as a control value for controller (21), provided that the two slewing cylinders (1, 2) are activated.

9. Slewing mechanism according to one of claims 2 to 6, characterised in that the higher of the two sensed load pressures of slewing cylinders (1, 2) reduced to a value which is proportional to the drive of proportional valve (25 or 26, respectively) serves as control value for controller (21).

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