Hydrostatic power transmission for adjustable volume pumps - incorporates automatic load compensation impulse to adjust drive motor speed

Liedl, Köth, Zeitler «

T (

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Patent Attorneys "

CU

CI 61. -. HI.

P. «« »New Tel. - No. G 89 / a% 41

Munich 22 - Steίηsdórfstra Be 21 - 22 Phone 089/29 84 62

RAUMA-REPOLA OY Lokomon Tehtaat, PL 306-307, 33 101 Tampere 10, Finland

Hydrostatic power transmission arrangement with load compensation

809821/0965

Hydrostatic power transmission arrangement with load compensation

The invention relates to a hydrostatic power transmission arrangement with load compensation with a drive motor, such as an internal combustion engine, which which drives one or more hydraulic pumps, in particular of adjustable volume, each of which is connected to a hydraulic circuit for supplying a pressure medium, to each of which a hydraulic actuating device is connected, a hydraulically controlled directional control valve group, which controls the pressure medium flow of each circuit to the hydraulic actuator ^ and arranged respectively in the hydraulic circuits in the main flow lines between the hydraulic pump and the directional control valve pressure reducing valves to reduce the maximum pressure in each hydraulic circuit, wherein in normal operation each hydraulic pump at its maximum setting angle, ie with the maximum volume set, operates while the speed of the drive motor is as low as possible.

In such a hydrostatic power transmission system, the efficiency of the system is quite low, and the speed of the drive motor of the system can either be set only manually or the speed can only be maintained at a constant value.

The object of the invention is therefore to provide a hydrostatic power transmission assembly with load compensation, which works taking into account the safety and efficiency requirements with reduced efficiency loss and which in

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can install a hydraulic machine.

This object is achieved in the hydrostatic power transmission arrangement of the aforementioned type according to the invention in that then, when the load of one or more actuators due to the defined by a control pulse movement load and / or the movement speed of the Belätigungs devices requires a higher pressure of the pressure medium and / or a greater generation of pressure medium ^ than in the respective hydraulic circuit is present ^ and after the control pressure or the corresponding directional valves has reached or exceeded a certain preset value, a connection is opened by the corresponding directional valve, which leads from the pressure line or the channel of the corresponding directional valve to a pressure demand line, which in turn to a hydraulic Steuerbzw. Actuator is guided, which the operation or the position of an adjusting member monitors, which serves to adjust the speed of the drive motor, so that due to the increased load automatically a load balancing pulse through the pressure supply line for adjusting the speed of the drive motor depending on the hydraulic control or Actuator and the adjusting member is achieved, and that the rotational speed of the drive motor is adjusted in response to the hydraulic circuit in which the highest speed for the respective hydraulic pump is required, while overproduction by the remaining pumps is avoided by a load balancing pulse control for each pump ,

In the invention, a control pulse can be supplied to each directional valve associated with the respective hydraulic actuator. There is then a comparison between the desired

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th and the actual movement of the actuator in the respective directional valve instead. If there is a difference between the desired and the actual motion sequence, a load balancing pulse to the drive motor, for example, to the injection pump of a diesel engine ^ sent from the directional control valve along the pressure demand line. The injection pump adjusts the speed of the drive motor according to demand.

The hydraulic pumps in the power transmission system can be designed as controllable pumps which operate in normal operation at maximum Einstellwinkeln, so that the speed of the drive motor can be kept as low as possible. The regulators of the pumps include a hydraulic maximum force control device which prevents overloading the drive motor torque. The control devices also include a reduction in the maximum pressure in the hydraulic medium. In addition, the pumps may have an adjusting device for adjusting the angle due to the load compensation pulse. This angle adjustment is used to reduce production when the speed of the drive motor is currently too high, which can often occur in a system with multiple pumps. In a power transmission system with multiple pumps, the speed of the drive motor may be controlled in response to the hydraulic pump circuit in which the highest speed is required. Overproduction of the other pumps is then avoided by the load balancing pulse control for each pump. Therefore, when a single pump requires a high speed, it does not result in overproduction or loss of power in the other pumps.

In the hydrostatic power transmission arrangement according to the invention, there is the advantage of increased efficiency

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In addition, the energy requirement of the drive motor is improved due to the correctly set engine speed. If an internal combustion engine is used as the drive motor, the speed can be reduced. During operation, this results in a reduction of the noise level, which can be detected by noise level measurements.

Another advantage in terms of energy requirements is that because of the low weight of the control devices a simultaneous control or regulation for the individual controls is independent of each other possible. In known power transmission systems, the operator must monitor the results, which consist of four to five separate control situations, which influence each other, and constantly correct individual control pulses in response to the changes in the load situations. In contrast, in the hydrostatic power transmission arrangement according to the invention, it is no longer necessary for the operator to constantly monitor the operation and to adapt to the requirements occurring during operation.

In the figures, an embodiment of the invention is shown. With reference to these figures, the invention will be explained in more detail. Show it:

Fig. 1 is a block diagram of a hydrostatic force over

bearing arrangement with load compensation;

Fig. 2 in a schematic representation of the control sequence

between the main components of the power transmission system;

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Fig. 3 is a schematic representation of a hydraulically ge

controlled directional valve group and

Fig. 4 shows an operating situation of counter valves for a

Direction of movement of a single directional valve in the way valve group of Fig. 3rd

An essential component of the hydrostatic power transmission arrangement are the directional control valves. In this respect, the construction of an embodiment of such a directional valve, which is used in the invention, will be explained first.

The directional control valve includes, as can be seen from Fig. 1, two identical operating units for example, the movements in both directions of a piston 51 of a double-acting piston-cylinder assembly 48. Here comes the one operating device for the one direction of movement to the advantage and the other operating device for the other movement device. Fig. 2 shows the structure of an operating device in a directional control valve. The operating device contains as pistons two slide bearings controlled rods 1 and 2. Cavities 7 and 8 at the left ends of the rods 1 and 2 in the drawing are connected via connections 9 and 10 to a tank line. A cavity 11 at the right end of the rod 1 in the drawing is connected to a control pressure line. A cavity 12 at the right end of the piston 2 is connected via a transmission channel 13 with an annular cavity 14 which is formed as a bore and serves to control the piston 1. From this annular cavity, a channel 15 extends within the rod 1 to the cavity 7 and from there to the tank line. Also, a direct access from the annular cavity 14 to the cavity 11, in which the control pressure prevails, on the right side of the rod is possible.

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This depends on the position of the rod 1. As can be seen from Fig. 2, a central portion 16 of the rod 1 has the smallest cross-section of the rod 1. At this point of the rod 1 opens a channel 4. Another channel 3 is offset from the channel 4 to the left. A support sleeve 17 is provided between the channel 3 and the channel 4 in the valve body. An annular flange 18, which is formed in the rod 1, cooperates with the support surface of the support sleeve 17. The flange protrudes from the side surface of the rod as a projection. The left end 19 of the rod 1 is guided in an intermediate sleeve 20, wherein a coil spring 21 is arranged around the rod 1 between the intermediate sleeve 20 and the annular flange 18 on the rod 1. The coil spring pushes the rod 1 toward the control pressure cavity 11, ie to the right in the drawing. The pressure reduction of the valve is caused by a sleeve 22 (pressure reducing sleeve). This is located in the illustrated embodiment on the rod 1 pruckverringerungskolben) and the intermediate sleeve 20. The pressure reduction sleeve 22 is pressed by a coil spring 23 which is disposed around the sleeve in a flow intersecting position (in the drawing to the right). The inner diameter of the pressure reducing sleeve 22 is larger at its right end than in the region which rests on the outer surface of the intermediate sleeve 20. In this way, an annular surface 24 is formed. The pressure which prevails in a gap 25 is reduced by the sleeve 22. The pressure-reducing rod 1 and the support sleeve 17 act on the annular surface 24 in such a manner that the pressure-reducing valve is opened. The pressure reducing sleeve 22 of the pressure reducing portion is also provided with flow grooves or the like. (Not shown in the drawing) in the region of the gap 25, so that when pressing the pressure-reducing sleeve 22 in the open direction due to the

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In the intermediate space 25 prevailing pressure, a certain area is opened, which leads from the gap 25 in the channel 3. This particular open passage area is constantly formed at a certain pressure in the gap 25. This open area together with the pressure value defines the flow rate.

The pressurized areas of the rod 1, which are connected to the channel 4, are the same for both directions of movement of the rod 1, so that the pressure in the channel 4, the movements of the rod 1 is not affected. The control pressure and thus the force acting on the pressure area at the end of the rod 1 force the pressure in the space 25 by the pressure area formed in this gap 25 counter. This pressure area is formed by the difference of the diameter of the annular flange 18 on the rod 1 and the diameter of the rod end 19.

The rod 1 is a pressure reducing piston having substantially four pressure areas:

Al = the pressure area of ​​the control pressure A2 = the tank pressure area A3 - pressure area of ​​the clearance and A4 = the compensation pressure area of ​​the back pressure of the return pipe.

The pressure area ratio Al: A3 defines the pressure in the intermediate space 25 with a certain control pressure value if one does not take into account the compensation of the pressure in the return line 3. At a pressure area ratio of 4: 1, for example, four times the control pressure in the space 25 is formed.

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When this pressure tends to increase, the force acting on the rod 1 force balance pushes the rod 1 against the control pressure and cuts off the connection between the channel 4 and the gap 25 from. If there is a pressure in the return line 3, the pressure value of the gap 25 must increase to a corresponding counter value, so that the desired pressure difference in the pressure reducing valve is present. This is achieved in that a connecting channel 26 leads from the return line 3 to an annular cavity 27, which is formed between the rod 1 and a sleeve 28 which is placed around the right end of the rod. The sleeve has on its inner surface 30 on the left side of the annular gap 27 a larger diameter than the inner surface on the right side of the annular gap 27th The diameters of the rod 1 on both sides of the annular gap 27 correspond to the diameters of the inner surfaces 30 and 29 of the sleeve 28, as shown in FIG. 4 can be seen. In this way, one obtains the compensation pressure area A4 in the annular space 27 on the bar 1. This pressure area is dimensioned to be equal to the pressure area A3 of the space 25. In this way, the pressure value of the intermediate space 25 increases with the pressure value in the return line 3. In this case, a pressure difference corresponding to the control pressure and the desired flow volume in the pressure control valve, which controls the flow velocity, caused.

The arrangement on the rod 2 is substantially the same as on the rod 1. A central part 31 of the rod 2 has the smallest cross section of the rod 2, as shown in FIG. 4 can be seen. A channel 6 is opened on this area around the rod 2. A channel 5 is arranged offset to the left of the channel 6. A support sleeve 32 is provided between the channels 5 and 6 in the valve body. One

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Ring flange 33 cooperates with the bearing surface of the support sleeve 32. The annular flange is arranged on the rod 2 so that it protrudes from the rod sides as a projection. The left end of the rod 34 of the rod 2 is guided in the illustrated embodiment by an intermediate sleeve 35. A coil spring 36 is disposed about the rod 2 between the intermediate sleeve 35 and the annular flange 33 on the rod 2. The coil spring pushes the rod 2 against the cavity 12 at the right end of the bar rod second The pressure reducing cross section of the valve is formed by a sleeve 37 which is arranged on the pressure reducing rod 2 and the intermediate sleeve 35. The pressure reducing sleeve 37 is urged by a coil spring 36 disposed around the sleeve into a flow cutting position (to the right in the drawing). The inner diameter of the pressure reducing sleeve 37 is larger at its right end than in the region which rests on the outer surface of the intermediate sleeve 35, so that an annular surface 39 is formed. The pressure generated in a gap 40 formed by the sleeve 37, the rod 2 and the support sleeve 32 acts on the annular surface 39 in such a way that the pressure-reducing valve is opened. The pressure reducing sleeve 37 is also provided with flow grooves or the like (not shown in the figure) in the area of ​​the space 40. Therefore, when the pressure in the clearance 40 presses the pressure reducing sleeve 37 in the open direction, there is an opened area leads from the gap 40 in the channel 5. This always happens when a certain pressure in the space 40 is present. This open area together with the pressure value determines the flow velocity.

The pressure areas at the connection points of the rod 2 or in the region of the channel 6 are for both directions of movement of the rod second

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equal, so that the pressure in the channel 6 does not affect the movements of the rod 2. The pressure in the space 12 and thus the force which comes to act on the pressure area at entspredienden end of the rod 2, the pressure in the space 40 counteracts. In this intermediate space 40, a pressure area is formed by the difference of the diameter of the annular flange 33 on the rod 2 and the diameter of the rod end 34.

Based on the pressure compensation, which has been described in connection with the rod 1, the arrangement of the rod 2 is carried out in a corresponding manner. A connecting channel 41 leads to an annular cavity 42 between the rod 2 and a sleeve 43, which is arranged around the right end of the rod. The sleeve 43 has on its inner surface 45 on the left side of the annular cavity 42 has a larger diameter than it has the inner surface 44 on the right side of the annular cavity 42. The diameters of the rod 2 on both sides of the annular cavity 42 correspond to the diameters of the inner surfaces 45 and 44 of the sleeve 43, as shown in FIG. 4 can be seen. In this way, a compensating pressure area is formed both in the annular cavity 42 and on the rod 2. This pressure gauge

area is dimensioned so that it is equal to the pressure area of ​​the gap 40. In this respect, the pressure value of the intermediate space 40 increases with the pressure value in the return line 5. In this case, a pressure difference is achieved according to the control pressure and the desired flow volume in the pressure reducing valve and the flow rate controlled.

In the valve body, a channel 46 is provided, which is opposite to a cutout in the surface 45 of the sleeve 43 is opened. A counter-valve 47 is disposed in the channel 46. This counter valve allows

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a flow away from the rod 2, but prevents flow towards the rod 2. When sufficient pressure is present in the cavity 12, the rod 2 moves to the left. A portion of the annular cavity 42 moves in a position on the channel 46, so that a connection from the channel 5 via the connecting channel 41 to the channel 46 is present.

In the following, the operation of the directional control valve with reference to FIGS. 3 and explained. Each directional control valve has four pressure compensated check valves with adjustable volume, which are connected together in the manner shown schematically in Fig. 3.

By the check valve a certain direction of movement and a certain speed of the associated hydraulic actuator is achieved. This happens regardless of whether the load applied to the actuator is positive or negative. The directional control valve, which is used in the illustrated embodiment of the invention also has a load balancing pulse line through which a necessary volume of oil, which is guided by the control pump to the actuator is controlled. To explain the operation, with the valve as the actuator, a piston-cylinder device 48 is shown. When the piston 51 and the piston rod 52 of the piston cylinder assembly 48 are moved to the left and pressure oil from the channel 5 into the cavity 50 of the piston cylinder arrangement 48 flows and is retracted from the cylinder cavity 49 of the piston cylinder assembly 48 through the channel 4 into the tank line, the control pressure supplied in the control pressure cavity 11 at one end of the rod 1 in the check valve in the line OB3 in Fig. 3. If the load is negative, in other words,

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when the desired movement is increased or accelerated by the load, it is sufficient that the control pressure in the control cavity 11 opens the connection from the channel 4 to the channel 3, so that the pressure medium from the cylinder cavity 49 of the piston cylinder assembly 48 through the gap 25 of the channel 4 and is eliminated through the channel 3 in the tank line. The filling of the cylinder cavity 50 on the other side of the piston cylinder arrangement 48 is effected by a suction valve - a check valve, which allows the free flow from the tank line into the channel 5, but prevents flow in the opposite direction - directly from the tank line. In the case of a positive load, in other words the force emanating from the load is opposite to the desired direction of movement, the normal opening movement of the rod 1 of the check valve is not sufficient. The control pressure must be increased so that the rod 1 is moved until the control pressure in the control pressure cavity through the transmission channel 13 also comes into effect in the cavity 12 at the right end of the rod. The resulting pressure in the cavity 12 opens the check valve, so that the connection from the channel 6 to the channel 5 is opened. If the pump then does not generate enough pressure medium, located on the pressure side of the check valve rod 2 is compensated for further increase in the control pressure, so that the rod 2 is moved by the control pressure until the connection from the channel 5 through the connecting channel 41 for Channel 46 is open and so a connection of channel 5 to the load balancing pulse line is formed by the check valve 47. The production of the control pump is caused by the load balancing pulse so that the volume of oil produced by the pump is sufficient for the actuator or actuators. If for the in the Flg. 3, piston movement 52 in the direction of the arrow is desired, ie

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when a direction of movement of the piston rod 52 is desired to the outside, the channel OA3 of the directional control valve is subjected to the control pressure. A check valve VT connects the outflow side to obtain the desired direction of movement, ie the cylinder space 50 on the side of the piston rod 52 with the tank line. The check valve VT has a controlled volume, that is, that an oil flow proportional to the control pressure can flow therethrough. In the case of a negative load, in other words when the load supports the desired movement, it is necessary for the check valve VT to be open in order to produce the desired movement. The filling of the other cylinder cavity 49 takes place through the suction valve from the tank line. In a positive load, in which the movement is not supported by connection of the cylinder cavity 50 on the outflow side with the tank line, an adjustment takes place in the check valve VT due to the increased control pressure. The conformation opens the channel of the cylinder cavity 49 on the inflow side to the pressure line, which comes from the pump via the valve VP. In this way, the control pressure OA3 opens the valves VT and VP while the suction valve remains closed.

If there is insufficient pressure in the pump line, there is an over adjustment in the valve VP with increasing control pressure. This means that the channel B3 is connected to the impulse side of the cylinder with the load equalization impulse line, so that the existing pressure in this line, the pump production is adjusted so that sufficient pressure medium is generated for the desired movement.

The check valves, which are connected to the tank line of the directional control valve, constantly contain the part which controls the flow volume. However, this is not always the case for the check valves on the pressure side. The pressure control takes place via the transmission channel 13_f but for this purpose, a mechanical device may be present. 809821/0965 8526

In the following an embodiment of a hydrostatic power transmission assembly with load compensation according to the invention will be described. Under a hydrostatic power transmission arrangement with load compensation is meant such in which a desired direction of movement and the speed of movement are defined by control pulses and in which the arrangement automatically adjusts the power and speed of the drive motor in response to the load situation.

From the hydraulic flow chart in Fig. 1, an embodiment can be seen in which a diesel engine 53 drives two hydraulic pumps 54 and 55 with adjustable volumes. Actuators 68 and 69 hold the setting angle of the pumps 54 and 55 in normal operation in the maximum position, so that the diesel engine runs at the lowest possible speed. The hydraulic pumps 54 and 55 move pressurized oil in the associated hydraulic circuits, in normal operation by the hydraulically controlled straightening valves 70 and 71, the actuators 68 and 69, throttles 72 and 73, a passage 74 which opens into a tank 75 from which the pumps 54 and 55 in turn remove pressure oil. The main flow lines 60 and 61 lead from the hydraulic circuits to way valve groups 56 and 57, through which, as shown in FIG. 1 It can be seen that is supplied from the main flow lines 60 and incoming pressure oil to the actuators, which are connected to the transmission system. The oil supply is controlled by valves 58 and 59 of the directional control valve group, which is assigned to each circuit. In this way, the direction of movement and the speed of movement of the actuator, which is defined by control pulses of the valves 58 and 59, caused.

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The hydrostatic power transmission system with load compensation according to the invention provides a control pulse applied to the directional control valves 58 and 59 of the respective actuator. There is a comparison between the desired and the actual movement in these valves. If a difference occurs in this comparison, a control pulse from the directional control valve 58 and / or 59 through the load equalization pulse line 66 and / or 67 the control or Actuator 65 is supplied, which adjusts the rotational speed of the drive motor 53 by corresponding adjustment of an adjusting member 64, for example, the injection pump of the diesel engine accordingly. In this respect, the system sets the speed of the drive motor due to the existing load situation using the load compensation pulses, which are passed through the lines and / or 67, automatically. The load balancing pulses are generated as a function of the respective load situation. A throttle point 76 is provided between the power demand lines 66 and 67 and the tank. Each line 66 and also has a check valve 77 and 78, which allow the pressure oil flow in the power supply lines 66 and 67 only in the direction of the directional control valve groups 56 and 57 to the control or actuator 65 out.

The control means of the pumps 54 and 55 include a hydraulic control unit for controlling the maximum force, whereby an overload of the torque of the engine 53 is prevented. The control devices also include a reduction in the maximum pressure of the hydraulic fluid. As such, pressure reducing valves 62 and 63 are provided in the main flow lines 66 and 61. The valves 62 and 63 are provided to control actuators 68 and 69 of the pumps 54 and 55, as shown in the hydraulic flow chart of FIG. For example, if the production of the pump 54 in the upper hydraulic circuit of

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Fig. 1 is too high and the pressure in this circuit increases to the maximum pressure value of the pressure reducing valve 62, opens the valve 62, so that the pressure oil can flow through it. The pressure in the conduit between the valve 62 and the restriction 72 then increases and the actuator 68 reduces the setting angle of the pump 54, thus reducing the production of the pump.

The pumps 54 and 55 have a setting possibility for the setting angle on the basis of the load compensation pulse. This adjustment is used to reduce production when the rotational speed of the drive motor is currently too high or if the system has multiple pumps. Overproduction of the other pumps is thereby prevented by the load balance pulse control for each pump.

809821/0965 ⁸⁵²⁶ ORIGINAL INSPECTED