JPH02266161A - Hydraulic drive circuit - Google Patents

Abstract

PURPOSE:To prevent any undesired tire slip from occurring by controlling the displacement of a variable displacement hydraulic pump so as to be abated, so as to cause the sum of travel load pressure and work load pressure not to exceed a limit of output in a prime mover. CONSTITUTION:Compound operation on travel and work is performed in keeping an accelerator 13 operated intact, and when the sum of both values of load exceeds a limit of engine power, rotational frequency of an engine 1 is abated, and the tilting value of a travel hydraulic pump 2 is reduced. On the other hand, both travel and work load pressure PT, PD works on an on-off valve 42, and when the sum exceeds the setting pressure of a spring 42a, this valve is opened, and symmetrical chambers 8a, 8b of a tilting cylinder 8 come to the same pressure. Then, the tilting value of the variable displacement hydraulic pump 2 shifts to the neutral side, travel load pressure goes down and the on-off valve 42 is selected to the closed side, whereby the tilting value of the pump 2 grows larger, increasing an amount of discharge, and then this operation is repeated, keeping the travel load pressure constant. Thus, an undesired tire slip can be prevented.

Description

DETAILED DESCRIPTION OF THE INVENTION A. Field of Industrial Application The present invention relates to a hydraulic drive circuit for construction machinery, and in particular to a variable capacity hydraulic pump that limits load and controls running torque during combined operation of running and work. The pushing and pushing is performed effectively by controlling the volume.

B. Prior Art In FIG. 5, which shows an example of this type of hydraulic drive circuit, an engine 1 drives a variable displacement hydraulic pump 2. The charge pump 3 and the working hydraulic pump 4 rotate. When the forward/reverse switching valve 6 is in the neutral position, the oil discharged from the charge pump 3 flows from downstream of the throttle 5 to the forward/reverse switching valve 6. It is guided to the left and right cylinder chambers 8a, 8b of the tilting cylinder 8 via pipes 7A, 7B, respectively, and the cylinder chambers 8a, 8b are at the same pressure. Therefore, the piston 8C is in the neutral position, the displacement (hereinafter referred to as the tilting amount) of the variable displacement hydraulic pump 2 is set to zero, and the discharge amount is zero.

When the operating lever 10 is operated to switch the forward/reverse switching valve 6 to the α side, the upstream pressure of the throttle 5 acts on the cylinder chamber 8a, the downstream pressure of the throttle 5 acts on the cylinder chamber 8b, and the piston 8c
Only the differential pressure bone before and after the aperture 5 is displaced to the right. As a result, the amount of tilting of the variable displacement hydraulic pump 2 is set, and the variable displacement hydraulic pump 2 supplies pressure oil to the main pipe line at a flow rate corresponding to the amount of tilting.
1A, the hydraulic motor 1-2 rotates forward, and the vehicle moves forward. When the forward/reverse switching valve 6 is switched to the m side, the tilting of the variable displacement hydraulic pump 2 is set in the opposite direction, pressure oil is discharged to the main pipe line 11B, and the hydraulic motor 12 is rotated in the reverse direction.

The rotational speed of the engine 1 is adjusted by the accelerator pedal 13, and the discharge flow rate of the charge pump 3 is proportional to the engine rotational speed, so the differential pressure across the throttle 5 is proportional to the engine rotational speed. The amount of tilting is proportional to the engine speed.

Note that 14 is a crossover load relief valve, 15 is a flushing valve, and 16A and 16B are check valves, which are connected to the charge pump 3 via pipes 17a and 17b. Further, 18 is a relief valve for the charging system.

By using such a circuit, the sum of the loads of the traveling variable displacement hydraulic pump 2 and the working hydraulic pump 4 can be kept below the maximum output of the engine 1.

This point will be explained below.

Depress the accelerator pedal] 3 to move forward and make the front (packet) (not shown) penetrate into gravel or the like.

Although the running load increases, generally variable displacement pump 2 for running
Since the single maximum load of the engine 1 is set smaller than the engine output, the engine 1 will not stall. In this state, operate the front operation control valve (not shown),
When the bucket is raised to scoop up gravel, etc., a load is applied to the working hydraulic pump 4. When the sum of the loads of the traveling and working hydraulic pumps 2.4 exceeds the engine output, the engine speed decreases. As a result, the differential pressure across the throttle 5 decreases, the screw 1 hem 8C of the tilting cylinder 8 moves toward the neutral side, and the amount of tilting of the traveling variable displacement hydraulic pump 2 decreases.

Due to this action, the engine continues to rotate at a rotation speed commensurate with the sum of the load of traveling and work while preventing engine stall.

FIG. 6 shows another example of a conventional hydraulic drive circuit, and the same parts as in FIG. 5 are given the same reference numerals and will be explained.

As the engine 1 rotates, the variable displacement hydraulic pump 2 and the working hydraulic pump 4 rotate, and the hydraulic power source 21- discharges pressure oil. When the operating lever 22 is in the neutral position as shown, the first servo valve 23 is also in the neutral position, and the left and right cylinder chambers 8a and 8b of the tilting cylinder 8 are both in communication with the tank to prevent the tilting of the variable displacement hydraulic pump 2. The amount of rotation is zero, and the hydraulic motor 12 is stopped.

When the operation lever 22 is operated in the on direction, the first
The servo valve 23 is switched to the right, pressure oil is guided to the cylinder chamber 8a, the piston 8c moves to the right, and the hydraulic pump 2 discharges pressure oil. This movement of the screw 1-8C causes the feedback link 25 to move in parallel to the right, and the first
The sleeve 23b of the servo valve 23 is moved to the right.

As a result, the positional relationship between the spool 23a and the sleeve 23b shifts to a neutral state, and the left and right cylinder chambers 8a and 8b of the tilting cylinder 8 are brought into communication again. Therefore, when the amount of tilting of the variable displacement hydraulic pump 2 decreases, the feedback link 25 moves in parallel to the left, and the cylinder chamber 8
Pressure oil is introduced to a, increasing the amount of tilting. By repeating such operations, the piston 8C of the tilting cylinder 8, in other words, the amount of tilting of the hydraulic pump 2 is maintained at a position corresponding to the amount of operation of the operating lever 22. Further, due to the parallel movement of the link 25, the link 26 is moved to the rotation axis 27. Swinging clockwise around the fulcrum 28, the sleeve 24b is displaced to the left,
Sleeve 24. b and the spool 24a are in a positional relationship that is shifted by the absolute amount of the tilting amount.

The discharge pressure of the variable displacement hydraulic pump 2 is transmitted to the spool 2 of the second servo valve 24 via the high pressure selection valve 29 and the pipe line 30, and the discharge pressure of the working hydraulic pump 4 is transmitted to the spool 2 of the second servo valve 24 via the pipe line 31.
One end of 4a is guided.

Therefore, when the sum of the pressures of both hydraulic pumps 2 and 4 overcomes the spring force of the spring 24c of the second servo valve 24a, the spool 24a moves to the right, the pressure oil of the hydraulic source 21 is cut off, and the The communication pipe line 32 with the first servo valve 23 is communicated with the tank, the pressure in the cylinder chamber 8a of the tilting cylinder 8 is reduced, and the amount of tilting of the variable displacement hydraulic pump 2 is reduced.

Therefore, the amount of tilting of the traveling variable displacement hydraulic pump 2 is controlled as shown in the P-q diagram in FIG. 7 so that the sum of the loads of the traveling and working hydraulic pumps 2 and 4 does not exceed the engine output. Ru.

C6 Problems to be Solved by the Invention In the above two conventional examples, when the sum of the traveling and working loads exceeds the engine output, the amount of tilting of the variable displacement hydraulic pump 2 is reduced accordingly to limit the load. However, the running circuit pressure is not limited in any way and is determined by the running load. In other words, if the running resistance is large, the running circuit pressure will rise to the relief pressure. Since the motor output torque is determined by the product of volume and pressure, the motor output torque will not decrease unless the travel circuit pressure decreases.

Therefore, if the accelerator is pressed while the packet penetrates gravel or the like, the motor output torque will reach the maximum, causing tire slip. If you try to raise the packet by operating the front in this state, the anti-stall control described above will work, but this will only reduce the amount of tilting of the travel hydraulic pump, so although the speed of tire slip will decrease, it will still reach its maximum speed. Continues to slip under running torque. Furthermore, excessive pressing force may make it impossible to lift the packet. In other words, with conventional devices, the output horsepower for driving is controlled, but the output torque is not controlled, so there is a problem that tire slip cannot be prevented and the front and driving forces cannot be matched. .

A technical problem of the present invention is to control the displacement of the hydraulic motor during combined operation of traveling and work, and also to control the traveling torque.

00 Means for Solving the Problems The present invention provides a traveling variable displacement hydraulic pump and a working hydraulic pump driven by a prime mover, and the sum of the loads of the traveling variable displacement hydraulic pump and the working hydraulic pump is the same as that of the prime mover. Pushing the variable displacement hydraulic pump so that the output is not exceeded is applied to a hydraulic drive circuit comprising load limiting means for controlling the displacement.

The above technical problem is solved by the following configuration.

A running load pressure detection means for detecting running load pressure, a working load pressure detecting means for detecting working load pressure, a running torque that is small as the working load pressure is large, and a running torque that is large as the working load pressure is small. so that you can
The driving torque control means is provided to reduce the displacement of the variable displacement hydraulic pump when the sum of both load pressures exceeds a predetermined value.

If the sum of the E0 acting running load and the working load exceeds the output horsepower of the prime mover, the engine will stall, so if the sum of the surface loads exceeds a predetermined value, the load limiting means will reduce the pressure on the variable displacement hydraulic pump for running. This reduces the volume and reduces its pump absorption per herk. Further, =8= The running torque limiting means limits the running torque to a smaller value as the work load pressure increases, and tire slip is effectively prevented.

F. Embodiment 1 First Embodiment FIG. 1 shows a first embodiment of a hydraulic drive circuit that performs a combined movement of traveling and work, and the same parts as in FIG. 5 are given the same reference numerals. explain.

The variable displacement hydraulic pump 2 and the hydraulic motor ]-2 are the main pipe 1
An open circuit connection is made by 1A and 11B. Main pipeline 1], A and 11. B is connected to the cross-over rotary relief valve 14, and the flushing valve 1
5 to the tank.

In addition, the main pipes 11A and IIB are connected to the check valve 16A.
and a pipe line 1 connected to the discharge port of a charging gear pump (hereinafter referred to as charge pump) 3 via 16B.
7a.

The discharge volume of the variable displacement hydraulic pump 2 (hereinafter referred to as the tilting amount in this embodiment) is controlled by the tilting cylinder 8.

A spring is installed in each of the left and right cylinder chambers 8a and 8b of the tilting cylinder 8, and the piston 8c is always held at a neutral position by the spring force of these springs. The cylinder chamber 8a and the cylinder chamber 8b are respectively connected to the forward/reverse switching valve 6 via a conduit 7A and a conduit 7B. On the other hand, each of the conduits 7A and 7B is the conduit 41A,
41B and an on-off valve 42. At one end of the spool of the on-off valve 42, there is a main pipe 11A taken out by the high pressure selection valve 44.

The running load pressure of either one of the hydraulic pumps 11B is led through the pipe 43, and the discharge pressure of the working hydraulic pump 4, in other words, the working load pressure is led through the pipe 45. Further, a spring 42a is provided at the other end of the spool of the on-off valve 42 to counteract the same pressure, and when the on-off valve 42 is normally in the closed position, the sum of the above pressures overcomes the spring force of the spring 4.2a. Switch to open position. The forward/reverse switching valve 6 is a manual three-position switching valve, and its P1 port is connected to the upstream pipe line 1 of the throttle 5.
-7b, and the P2 port is connected to the downstream pipe line 1 of the throttle 5.
7a.

The rotation speed of the engine 1 is controlled by an accelerator pedal 13, and the engine 1 operates three hydraulic pumps 2, 3, .
4 is driven. Further, the relief pressure of the charging pump 3 is regulated by a relief valve 18.

In the above embodiment, charging pump 3. .

Pipe lines 17b, 17a, throttle 5. Pipe lines 7a and 7b serve as load limiting means, pipe line 43 serves as running load pressure detection means, and pipe line 4
5 is a work load pressure detection means, an on-off valve 42. Spring 42a
, conduit 4. LA, 4.1. B constitutes a traveling torque control means, respectively.

Next, the operation of the embodiment configured as above will be explained.

When performing a combined operation of driving and working (hereinafter referred to as front) while keeping the accelerator pedal 13 depressed, if the sum of the driving load and the front load exceeds the engine output, the rotational speed of the engine 1 is reduced. The amount of tilting of the traveling hydraulic pump 2 decreases. This operation is as described above.

On the other hand, the running circuit pressure (running load pressure) PT and the front circuit pressure (working load pressure) Pp are acting on the on-off valve 42, and when (PT+PP) exceeds the pressure Pr set by the springs 4 and 2a, The on-off valve 42 is opened, and the pipe line 4LA,
Pipe lines 7A and 7B are communicated via 4.1B. As a result, the left and right cylinder chambers 8a and 8b of the tilt cylinder 8 have the same pressure, and the tilt amount of the variable displacement hydraulic pump 2 begins to decrease toward neutral, that is, zero. As a result, the traveling circuit pressure PT decreases, and the force of pushing the on-off valve 42 due to the pressure in the pipe line 43 also decreases. When (PT+PF)≦PR, the on-off valve 42 switches to the closed position, the amount of tilting of the variable displacement hydraulic pump 2 increases, the discharge amount increases, and the travel circuit pressure increases. When (PT+PP)>PR again, the on-off valve 42
is switched to the closed position, and the discharge amount of the variable displacement hydraulic pump 2 is reduced. By repeating such operations, the running circuit pressure is kept constant.

Such an operation is determined by the magnitude relationship between (PT0PP) and PR, and the relationship between the front circuit pressure and the running circuit pressure is as shown in FIG.

In Fig. 2, p Tmax is the maximum pressure of the running circuit;
PS is the circuit pressure when the tire is slipping, P pmax and P pmin are the maximum pressure of the front circuit and the pressure when not operating, P 7Low is P p
It shows the running circuit pressure at max. When the front is not operated, the running circuit pressure tries to rise to P TmaX, but the tires slip beyond a certain limit. Here, when the front is operated, the front circuit pressure PP increases, and the traveling circuit pressure PT decreases while maintaining the relationship shown in the figure. If the front circuit pressure is PFf3 or more, the running circuit pressure is controlled to be less than ps, and tire slip is stopped.

Further, FIG. 3 shows a Pq diagram in the above embodiment, and the cutoff pressure Pc depends on the front circuit pressure PF. For example, the cutoff pressure when Ppmjn is P crn
ax (= P 7max: Fig. 2), Ppma
When x, P c mjn (= P7Low: 2nd
), and the p-q diagram is set between Max and Min according to the front circuit pressure.

-Second Embodiment- Fig. 4 shows a second embodiment, and the same parts as in Figs. 1 and 6 are given the same reference numerals and will be described.

The 11 pressure source 21 generates pressure oil for controlling the tilting amount of the traveling variable capacity 4 (1 pressure pump 2). The first servo valve 23 that functions as a reverse switching valve and the travel circuit pressure P
A second servo valve 24 that determines the P-q diagram of the variable displacement hydraulic pump 2 by the sum of T and Freon 1 - circuit pressure I)F
and a control valve 142 that controls the cutoff pressure of the Pq diagram determined by the servo valve 24 in accordance with the front circuit pressure PP are connected in series. The running circuit pressure PT is introduced to one end of the spool of the control valve 142 via the pipe 30, and the front circuit pressure PF is introduced via the pipe 31, and (PT+PF) is applied to the spring 142a.
When the pressure exceeds the pressure P1 set by communicate with. Further, when completely switched to the right, the pressure oil is cut off and the pipe line 143 is communicated with the tank.

The feedback link 25 is movable in parallel to the left and right in the figure according to the amount of tilting of the variable displacement hydraulic pump 2, and controls the positional relationship between the spool 23a and the sleeve 23b of the first servo valve 23. Further, the link 26 is connected to the feedback link 25 with the pivot shaft 27 as the center and the fulcrum 28 as the fulcrum.
It is configured to swing clockwise or counterclockwise depending on the amount of left and right momentum. Therefore, the forked portion 33a
The lever 33, which cooperates with the link 26 via the lever 33, displaces the sleeve 24b of the second servo valve 24 by the absolute amount of the displacement movement of the variable displacement hydraulic pump 2.

Also in the second embodiment, as described above, the Pq diagram in FIG. 3 is determined by the action of the second servo valve 24, and the traveling circuit pressure is determined by the action of the newly added control valve 142. The cutoff pressure is changed by the front circuit pressure, and when the front circuit pressure becomes more than PPS, the running circuit pressure becomes less than Ps, and tire slip is prevented.

In the first embodiment described above, when the sum of the running circuit pressure and the front circuit pressure exceeds a predetermined value, the left and right cylinder chambers 8a of the tilting cylinder 8.

8b is connected, but the throttle 5 is an electrically controlled variable throttle, and when the sum of the above two pressures exceeds a predetermined value, the throttle area is maximized and the differential pressure across the front and rear is zero, thereby guiding the tilting cylinder to neutral. You can do it like this.

Further, in the second embodiment, the cut-off pressure is controlled by throttling or cutting off the supply of pressure oil to one cylinder chamber, but it may also be controlled in a similar manner.

The operation lever 22 is an electric lever that outputs an electric signal according to the amount of operation thereof, and the -th servo valve 23
is an electromagnetic proportional servo valve driven by the direction (target value) of the electric lever, and the amount of tilting of the variable displacement hydraulic pump is controlled according to the target value. Then, the running circuit pressure and the front circuit pressure may be electrically detected, and when the sum thereof is equal to or greater than a predetermined value, the target value may be set to zero to perform cutoff. Further, the tilting cylinder 8 can be similarly controlled using an electric actuator.

Although the above two embodiments have been described with respect to the case where the variable displacement hydraulic pump and the hydraulic motor are connected in a closed circuit, the present invention can also be applied to an open circuit.

G1 Effects of the Invention According to the present invention, the displacement of the variable displacement hydraulic pump is controlled so that the sum of the loads of the two hydraulic pumps does not exceed the output of the prime mover during combined operation of travel and work, and the displacement of the variable displacement hydraulic pump is controlled during the combined operation of traveling and work. Since the running torque is limited to a smaller value as the load pressure increases, undesired tire slip can be prevented and tire damage can be suppressed. Additionally, the fuel efficiency is improved compared to the conventional example. Furthermore, the combination of front working force and traveling pressing force can be set arbitrarily, improving efficiency as a working machine.

[Brief explanation of drawings]

FIG. 1 is a hydraulic circuit diagram showing a first embodiment. FIG. 2 is a diagram showing the relationship between running circuit pressure and front circuit pressure. FIG. 3 is a p-q diagram. FIG. 4 is a hydraulic circuit diagram showing a second embodiment. FIGS. 5 and 6 are diagrams showing conventional hydraulic circuits. FIG. 7 is a p-q diagram of the conventional circuit. 2: Hydraulic pump 3: Charge pump 4: Front hydraulic pump 5: Throttle 6: Forward/forward switching valves 7A, 7B:
Conduit 8: Tilting shifter 11A, IIB: Main conduit
12: Hydraulic motor 42: Open/close valve 142: Control valve

Claims (1)

[Scope of Claims] A traveling variable displacement hydraulic pump and a working hydraulic pump driven by the prime mover, and the sum of the load of the traveling variable displacement hydraulic pump and the load of the working hydraulic pump does not exceed the output of the prime mover. In the hydraulic drive circuit, the hydraulic drive circuit includes load limiting means for controlling the displacement of the variable displacement hydraulic pump, comprising: traveling load pressure detecting means for detecting traveling load pressure; and working load pressure detecting means for detecting working load pressure. and, when the sum of the two load pressures exceeds a predetermined value, the displacement volume of the variable displacement hydraulic pump is reduced so that the larger the working load pressure is, the smaller the running torque is, and the smaller the working load pressure is, the larger the running torque is. 1. A hydraulic drive circuit for construction machinery, comprising a traveling torque control means for reducing running torque.