DE2751663A1 - HYDROSTATIC POWER TRANSMISSION ARRANGEMENT WITH LOAD COMPENSATION - Google Patents

Description

Liedl, Köth, Zeitler " _{T (} " Patent Attorneys " ^CU CI 61. -. H I. P. «« »New phone number G 89 / a% 41 Munich 22 - Steίηsdοrfstra Be 21 - 22 Telephone 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, for example an internal combustion engine, which or which one or more hydraulic pumps, in particular with adjustable volume, drives each of which for the delivery of a pressure medium with a hydraulic Circuit is connected, to each of which a hydraulic actuator is connected, a hydraulically controlled Directional valve group, which controls the pressure medium flow of each circuit to the hydraulic actuator ^ and each with in the hydraulic circuits in the main flow lines between the hydraulic pump and the directional control valve group to reduce the maximum pressure in each hydraulic circuit, with each hydraulic pump at its maximum during normal operation Setting angle, i.e. with the maximum set volume, works while the speed of the drive motor is as low as possible is.

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

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

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can install a hydraulic machine.

In the case of the hydrostatic power transmission arrangement of the type mentioned at the outset, this object is achieved according to the invention in that that when the load on one or more actuating devices due to the movement load defined by a control pulse and / or the movement speed of the actuating devices a higher pressure of the pressure medium and / or a greater production of pressure medium requires ^ than is present in the respective hydraulic circuit ^ and after the control pressure of the or the corresponding directional control valve has reached or exceeded a certain preset value by means of the corresponding Directional control valve a connection is opened, which leads from the pressure line or the channel of the corresponding directional control valve to a pressure demand line, which in turn leads to a hydraulic Steuerbzw. Actuator is guided, which monitors the operation or the position of an adjusting member which is used to adjust the speed of the drive motor, so that due to the increased load automatically a load compensation pulse through the pressure demand line to adjust the speed of the drive motor depending on the hydraulic control or actuator and the Adjusting member is achieved, and that the speed of the drive motor is adjusted in dependence on the hydraulic circuit, in which the highest speed is required for the respective hydraulic pump, while overproduction by the others Pumping is avoided by a load balancing pulse control for each pump.

In the invention, each directional control valve, which of the respective hydraulic actuating device is assigned, a control pulse can be delivered. A comparison is then made between the desired

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th and the actual movement of the actuating device in the respective directional control valve. If there is a difference between the desired and the actual sequence of movements is, the directional control valve sends a load compensation pulse along the pressure demand line sent to the drive motor, for example to the injection pump of a diesel engine ^. The injection pump sets the speed of the drive motor as required.

The hydraulic pumps in the power transmission system can be used as controllable Pumps be designed which work in normal operation at maximum setting angles, so that the speed of the drive motor can be kept as low as possible. The control devices of the pumps contain a hydraulic maximum force control device, which prevents overloading of the drive motor torque. The control facilities also include a reduction in the maximum pressure in the hydraulic medium. Besides the pumps can have an adjusting device for adjusting the angle based on the load compensation pulse. These Angle adjustment is used to reduce production if the speed of the drive motor is currently too high, which can occur frequently in a system with multiple pumps. In a power transmission system with multiple pumps, the speed can the drive motor be controlled depending on the hydraulic pump circuit, in which the highest speed is required. Overproduction of the other pumps is then determined by the load balancing pulse control for each pump avoided. Therefore, if a high speed is required from a single pump, this does not lead to overproduction or overproduction. loss of power in the other pumps.

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

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

There is another advantage in terms of energy consumption in that, because of the low weight of the control devices, simultaneous control or regulation for the individual actuating devices is possible independently of one another. At acquaintances In power transmission systems, the operator must determine the results of four to five separate control situations that arise mutually influence, monitor and constantly individual control impulses depending on the changes in the load situation correct. In the hydrostatic power transmission arrangement according to the invention, on the other hand, it is no longer necessary that the operator constantly monitors the operation and adapts it to the requirements occurring during operation.

An exemplary embodiment of the invention is shown in the figures. The invention is to be explained in greater detail on the basis of these figures. Show it:

Fig. 1 is a block diagram of a hydrostatic force

bearing arrangement with load compensation;

2 shows the control sequence in a schematic representation

between the main components of the power transmission system;

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Fig. 3 in a schematic representation of a hydraulic ge

controlled directional valve group and

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

Direction of movement of an individual directional valve in the directional valve group of FIG. 3.

An essential part of the hydrostatic power transmission arrangement are the directional control valves. In this respect, the structure of an exemplary embodiment of such a directional control valve, which is used in the invention are explained.

As can be seen from FIG. 1, the directional control valve contains two identical ones Operating units for, for example, the movements in both directions of a piston 51 of a double-acting piston-cylinder arrangement 48. One operating facility applies to one direction of movement and the other operating facility applies to the other movement device. Fig. 2 shows the structure of an operating device in a directional control valve. The operating equipment contains two rods 1 and 2 controlled in slide bearings as pistons. Cavities 7 and 8 at the left ends of rods 1 and 2 in the drawing are connected to a tank line via connections 9 and 10. 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 to an annular cavity 14, which is designed as a bore is and is used to control the piston 1. From this annular cavity extends a channel 15 within the rod 1 to Cavity 7 and from there to the tank line. There is also direct access from the annular cavity 14 to the cavity 11, in which the control pressure prevails, possible on the right side of the rod.

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This depends on the position of the rod 1. As can be seen from FIG. 2, a central part 16 of the rod 1 has the smallest cross section of the rod 1. At this point of the rod 1, a channel 4 opens. Another channel 3 is offset to the left in relation to channel 4. A support sleeve 17 is between the channel 3 and the channel 4 in Valve body provided. An annular flange 18, which is molded into the rod 1, interacts with the support surface of the support sleeve 17. The flange stands 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, with a spiral spring 21 around the rod 1 between the intermediate sleeve 20 and the annular flange 18 on the rod 1 is arranged. The spiral spring pushes the rod 1 towards the control pressure cavity 11, i.e. to the right in the drawing. The pressure reduction of the valve is brought about by a sleeve 22 (pressure reduction sleeve). In the illustrated embodiment, this is located on the rod 1 (pressure reducing piston) and the intermediate sleeve 20. The pressure reducing sleeve 22 is converted into a flow by a spiral spring 23 which is arranged around the sleeve cutting position pressed (to the right in the drawing). The inner diameter of the pressure reducing sleeve 22 is larger at its right end than in the area which is on the outer surface of the Intermediate sleeve 20 rests. In this way an annular surface 24 is formed. The pressure, which in a space 25 prevails 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 way that the pressure reducing valve is opened. The depressurization sleeve 22 of the depressurization portion is also Equipped with flow channels or the like (not shown in the drawing) in the region of the intermediate space 25, so that when pressed the pressure reduction sleeve 22 in the open direction due to the

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In the pressure prevailing space 25 a certain area is open, which leads from the space 25 into the channel 3. This specific open passage area is constantly formed in the intermediate space 25 at a specific pressure. This open one Area defines the flow velocity together with the pressure value.

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

The rod 1 is a depressurization piston with essentially four Print areas:

Al = the pressure area of the control pressure A2 = the tank pressure area
A3 - pressure area of the gap and A4 = the compensation pressure area of the return line back pressure.

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

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When this pressure tries to increase, the force compensation acting on the rod 1 presses the rod 1 against the control pressure and cuts off the connection between the channel 4 and the space 25. If there is pressure in the return line 3, the pressure value of the intermediate space 25 must increase to a corresponding equivalent value so that the desired pressure difference is present in the pressure reducing valve. This is achieved by having a Connection channel 26 from the return line 3 to an annular one Cavity 27 leads, which is formed between the rod 1 and a sleeve 28, which is placed around the right end of the rod. The sleeve has a larger diameter on its inner surface 30 on the left side of the annular space 27 than the inner surface on the right side of the annular space 27. The diameters of the rod 1 on both sides of the annular space 27 correspond to the diameters of the inner surfaces 30 and 29 the sleeve 28, as can be seen from FIG. In this way, the compensation pressure area A4 is obtained in the annular gap 27 on the rod 1. This pressure area is dimensioned so that it is equal to the printing area A3 of the space 25. In this way, the pressure value of the space 25 increases with the Pressure value in the return line 3. A pressure difference corresponding to the control pressure and the desired flow volume is produced in the pressure control valve which controls the flow rate.

The arrangement on the rod 2 is essentially 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 can be seen from FIG. One channel 6 is open in this area around rod 2. A channel 5 is arranged offset to the left to the channel 6. A support sleeve 32 is provided between the channels 5 and 6 in the valve body. A

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The annular flange 33 interacts with the support 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 rod end 34 of the rod 2 is guided through an intermediate sleeve 35 in the embodiment shown. A spiral spring 36 is arranged around the rod 2 between the intermediate sleeve 35 and the annular flange 33 on the rod 2. The spiral spring presses the rod 2 against the cavity 12 at the right end of the rod 2. The pressure reduction cross section of the valve is formed by a sleeve 37 which is arranged on the pressure reduction rod 2 and the intermediate sleeve 35. The pressure reducing sleeve 37 is pressed into a position in which the flow is cut off by a spiral spring 36 arranged around the sleeve (to the right in the drawing). The inner diameter of the pressure reducing sleeve 37 is larger at its right end than in the area which rests on the outer surface of the intermediate sleeve 35, so that an annular surface 39 is formed. The pressure which is generated in a space 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 channels or the like (not shown in the figure) in the area of the interspace 40 leads from the intermediate space 40 into the channel 5. This always happens when a certain pressure is present in the intermediate space 40. 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 area of the channel 6 are for both directions of movement of the rod 2

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equal, so that the pressure in the channel 6 the movements of the rod 2 unaffected. The pressure in the space 12 and thus the force that acts on the pressure area at the end of the rod 2 If there is an effect, the pressure in the intermediate space 40 opposes it. In this intermediate space 40, a pressure area is formed by the Difference between 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 is related has been described with the rod 1, the arrangement of the rod 2 is made in a corresponding manner. A connection 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 On its inner surface 45 on the left side of the annular cavity 42, the sleeve 43 has a larger diameter than it has the inner surface 44 on the right side of the annular cavity 42. The diameter 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 can be seen from FIG. In this way, a compensation pressure area is formed both in the annular cavity 42 and on the rod 2. This printing

The 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 a pressure difference corresponding to the control pressure and the desired flow volume in the pressure reducing valve is achieved and the Controlled flow rate.

In the valve body, a channel 46 is provided, which is opposite to a Cutout in the surface 45 of the sleeve 43 is open. A counter valve 47 is arranged in channel 46. This counter valve allows

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a flow away from the rod 2, but it prevents a flow towards the rod 2. When there is sufficient pressure in the cavity 12 is, the rod 2 moves to the left. A region of the annular cavity 42 moves into 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 valve has four pressure-compensated check valves with adjustable volume, which are connected to each other are connected in the manner shown schematically in Fig. 3.

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

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if 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 channel 4 to channel 3, so that the Pressure medium is removed from the cylinder cavity 49 of the piston-cylinder arrangement 48 through the intermediate space 25 of the channel 4 and through the channel 3 into the tank line. The filling of the cylinder cavity 50 on the other side of the piston-cylinder assembly 48 takes place through a suction valve - a check valve, which the free Allow flow from the tank line into channel 5, but prevent flow in the opposite direction - directly from the Tank line. In the case of a positive load, in other words that of the Load exerted force If the direction of movement is opposite to the desired direction, the normal opening movement of rod 1 is sufficient the check valve does not turn off. 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 in the cavity 12 comes into effect at the right end of the rod. The resulting pressure in the cavity 12 opens the check valve so that the Connection from channel 6 to channel 5 is open. If the pump then does not generate enough pressure medium, the rod 2 located on the pressure side of the check valve becomes larger as it grows of the control pressure over compensated, so that the rod 2 is shifted by the control pressure until the connection from the channel 5 through the connecting channel 41 to channel 46 is open and so a connection from channel 5 to the load compensation pulse line through the check valve 47 is formed. The production of the control pump is brought about by the load compensation pulse line so that the oil volume generated by the pump is sufficient for the actuation device or actuation devices. if for those in the Flg. 3 a movement of the piston rod 52 in the direction of the arrow is desired, i. E.

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when an outward direction of movement of the piston rod 52 is desired control pressure is applied to channel OA3 of the directional control valve. A check valve VT connects the outflow side to to obtain the desired direction of movement, i.e. the cylinder space 50 on the side of the piston rod 52 with the tank line. That Check valve VT has a controlled volume, i.e. an oil flow proportional to the control pressure can flow through it can. With negative load, in other words when the load supports the desired movement, it is necessary that the check valve VT is open to produce the desired movement. The other cylinder cavity 49 is filled by the Suction valve from the tank line. In the case of positive loading, in which the movement is achieved by connecting the cylinder cavity 50 on the outflow side is not supported by the tank line, there is an over-adjustment in the non-return valve VT due to the increased control pressure. The overlapping opens the channel of the cylinder cavity 49 on the inflow side to the pressure line, which via the valve VP of the Pump comes. In this way, the control pressure OA3 opens the valves VT and VP while the suction valve remains closed.

If there is not enough pressure in the pump line, it takes place an over setting in valve VP with increasing control pressure. This means that the channel B3 on the joint side of the cylinder with the load compensation impulse line is connected so that the pressure present in this line adjusts the pump production so that for the desired movement sufficient print medium is generated.

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

The following is an embodiment of a hydrostatic Power transmission arrangement described with load compensation according to the invention. A hydrostatic power transmission arrangement with load compensation is understood to mean one in which a desired direction of movement and the speed of movement are defined by control pulses and in which the arrangement depends on the force and speed of the drive motor automatically adjusts to the load situation.

The hydraulic flow diagram in FIG. 1 shows an exemplary embodiment in which a diesel engine 53 drives two hydraulic pumps 54 and 55 with adjustable volumes. Actuators 68 and 69 maintain the setting angles of 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 through the hydraulically controlled ones Directional valves 70 and 71, the actuating devices 68 and 69, throttle points 72 and 73, a channel 74 which opens into a tank 75 from which the pumps 54 and 55 in turn draw pressure oil. The main flow lines 60 and 61 lead from the hydraulic circuits to valve groups 56 and 57, through which, as from Fig. 1 can be seen that from the main flow lines 60 and incoming pressure oil is supplied to the actuating devices which are connected to the transmission system. The oil supply will controlled by valves 58 and 59 of the directional control valve group assigned to each circuit. In this way, the direction of movement and the speed of movement of the actuating device, which are defined by control pulses from valves 58 and 59 is caused.

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The hydrostatic power transmission system with load compensation According to the invention, a control pulse is provided which is applied to the directional control valves 58 and 59 of the respective actuating device is. A comparison is made between the desired and the actual movement in these valves. If a difference occurs during this comparison, a control pulse is sent from the directional control valve 58 and / or 59 through the load compensation pulse line 66 and / or 67 supplied to the control or actuating member 65, which the speed of the drive motor 53 by appropriate setting of an adjusting member 64, for example the injection pump of the diesel engine adjusted accordingly. In this respect, the system provides the speed of the drive motor based on the existing load situation With the help of the load compensation pulses, which are passed through the lines and / or 67, automatically. The load compensation impulses are dependent on the respective load situation generated. A throttle point 76 is provided between the power supply lines 66 and 67 and the tank. Each line 66 and also has a check valve 77 and 78, which the pressure oil flow in the power lines 66 and 67 only in the direction of the directional valve groups 56 and 57 to the control or actuating member 65 allow.

The control devices of the pumps 54 and 55 contain a hydraulic one Control unit for controlling the maximum force, which prevents the torque of the drive machine 53 from being overloaded will. The control devices also include a reduction in the maximum pressure of the hydraulic fluid. In this respect, pressure reducing valves are 62 and 63 are provided in the main flow lines 66 and 61. The valves 62 and 63 are for controlling actuators 68 and 69 of the pumps 54 and 55 are provided, as can be seen from the hydraulic flow diagram of FIG. For example, if the production of the pump 54 in the upper hydraulic circuit of the

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

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

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Claims (1)

2751653 Liedl, Moth, ZeiNer Munich 22 - S t e i η sdo rf st ra B e 21 - 22 ■ Telephone 089/29 84 62 Claim Hydrostatic power transmission arrangement with load compensation with a drive motor, ζ. B. an internal combustion engine, which or which one or more hydraulic pumps, in particular with adjustable volume, drives, each of which for delivery a pressure medium is connected to a hydraulic circuit, to each of which a hydraulic actuating device is connected, a hydraulically controlled directional valve group, which the Pressure medium flow of each circuit to the hydraulic actuator controls pund with each in the hydraulic circuits in the main flow line between the hydraulic pump and the directional control valve group arranged pressure reduction valves for reduction the maximum pressure in each hydraulic circuit, with each hydraulic pump in normal operation at its maximum setting angle, i.e. with the maximum set volume, working at the lowest possible speed of the drive motor, characterized in that when the load on one or more actuating devices is due to the movement load defined by a control pulse and / or speed of movement of the actuating devices a higher pressure of the pressure medium and / or a stronger generation of pressure medium than in the respective hydraulic circuit is present, and after the control pressure of the or the corresponding Directional control valve (58, 59) has reached or exceeded a certain preset value through the corresponding directional control valve (58 or 59) 809821/0965 B 8526 N / Li ORIGINAL INSPECTED a connection is opened which leads from the pressure line or the channel of the corresponding directional valve to a power requirement line (66, 67), which in turn is led to a hydraulic control or actuating member (65) which controls the operation or the position of a Adjusting member (64), which is used to adjust the speed of the drive motor (53), so that, due to the increased load, a load compensation pulse is automatically sent through the power requirement line to adjust the speed of the drive motor as a function of the hydraulic control or actuating member (65 ) and the adjusting member (64) is achieved and that the speed of the drive motor (53) is set depending on the hydraulic circuit in which the highest speed for the respective hydraulic pump (54 or 55) is required, while an overproduction of remaining hydraulic pumps (54, 55) is avoided by a load compensation pulse control for each pump . 809821 / 096S 8526