JPH076521B2 - Load sensing hydraulic drive circuit controller - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a load sensing hydraulic drive circuit of a hydraulic machine including a plurality of hydraulic actuators such as a hydraulic excavator and a hydraulic crane. More specifically, the present invention relates to a load sensing regulator. Controls the pump discharge flow rate so that the discharge pressure of the hydraulic pump becomes higher than the maximum load pressure of those hydraulic actuators by a predetermined value, while controlling the flow rate of the pressure oil supplied to the hydraulic actuator by the flow control valve with pressure compensation. Load sensing hydraulic drive circuit control device.

(Prior Art) Conventionally, a load sensing hydraulic drive circuit is used in a hydraulic machine including a plurality of hydraulic actuators such as a hydraulic excavator and a hydraulic crane. This hydraulic drive circuit is provided between a hydraulic pump and each hydraulic actuator. And a flow control valve with pressure compensation for controlling the flow rate of pressure oil supplied to the hydraulic actuator according to the operation signal of the operating means, and the discharge pressure of the hydraulic pump is higher than the maximum load pressure of the hydraulic actuators. Equipped with a load-sensing regulator that controls the pump discharge flow rate so that the value increases only by this value, this load-sensing regulator reduces the discharge flow rate of the hydraulic pump to the minimum necessary amount that complies with the required consumption flow rate of the hydraulic actuator, and saves energy while also compensating pressure. Difference Before and After Required for Pressure Compensation Control of Flow Control Valve Device The pressure is secured, and the pressure compensation control of the flow rate of the pressure oil supplied to the hydraulic actuator is surely performed.

(Problems to be Solved by the Invention) However, the above-mentioned conventional techniques have the following problems.

In the load sensing hydraulic drive circuit, the hydraulic pump is usually equipped with a horsepower control device. In this case, if the pump discharge pressure rises during operation and the engine that drives the hydraulic pump is likely to stall, the load is reduced. The horsepower control device functions in priority to the sensing regulator, the discharge flow rate (maximum discharge flow rate) of the hydraulic pump decreases, and the pump discharge flow rate becomes smaller than the required consumption flow rate of the hydraulic actuator. When operating the operating levers to simultaneously drive a plurality of hydraulic actuators, the sum of the required consumption flow rates of the plurality of hydraulic actuators commanded by the operating means is the maximum capacity of the hydraulic pump (maximum discharge flow rate). May exceed.

In the above conventional technology, in any of these cases, the pump discharge flow rate is insufficient with respect to the total required consumption flow rate. Therefore, even if the load sensing regulator operates, the discharge pressure of the hydraulic pump becomes higher than the maximum load pressure by a predetermined value. There is a problem that it becomes impossible to maintain it, most of the pump discharge flow rate flows to the low pressure side hydraulic actuator, pressure oil is not supplied to the high pressure side hydraulic actuator, and the work element cannot be controlled in a desired manner.

The object of the present invention is to control the supply of pressure oil to each hydraulic actuator regardless of the load pressure, even when the operating state is such that the pump discharge flow rate is insufficient with respect to the required consumption flow rate. It is an object of the present invention to provide a load sensing hydraulic drive circuit control device capable of improving composite operability.

(Means for Solving the Problems) The above-mentioned object is to connect at least one hydraulic pump, a plurality of hydraulic actuators driven by pressure oil discharged from the hydraulic pump, and between the hydraulic pump and each hydraulic actuator. And a flow control valve with pressure compensation for controlling the flow rate of the pressure oil supplied to the hydraulic actuator according to the operation signal of the operating means, and the discharge pressure of the hydraulic pump is a predetermined value higher than the maximum load pressure of the plurality of hydraulic actuators. In the control device of the load sensing hydraulic drive circuit having a load sensing regulator for controlling the pump discharge flow rate so as to be higher, the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators is detected, A differential pressure detecting means for outputting a differential pressure signal corresponding thereto, an operation signal of the operating means, and the differential pressure detecting means. The differential pressure signal of the output means is input to obtain the flow rate command value of each of the hydraulic actuators, and the control means controls the flow control valve with pressure compensation based on the flow rate command value. The control coefficient becomes 1 in a range where the differential pressure signal is larger than a predetermined value, and when the differential pressure signal becomes a predetermined value or less, the control coefficient becomes smaller than 1 as the differential pressure signal decreases. The function relation of is set in advance,
Based on this functional relation, a control coefficient is obtained from the input differential pressure signal, and a flow rate signal for each hydraulic actuator obtained from each operation signal is multiplied by the control coefficient to obtain the flow rate command value. This is achieved by adopting a configuration having

(Operation) When the required consumption flow rate of the hydraulic actuator becomes larger than the maximum discharge flow rate of the hydraulic pump and the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure becomes small, the differential pressure signal becomes equal to or less than a predetermined value, and the control means A value smaller than 1 is obtained as the control coefficient, and each flow rate command value is reduced at the same rate. With this, while distributing the discharge flow rate of the hydraulic pump according to the ratio of the operation amount of the operating means, the supply flow rate to each hydraulic actuator is reduced by the same ratio, and the pressure oil of each hydraulic actuator is reduced regardless of the load pressure. Supply control can be performed, and composite operability can be improved. In addition, the differential pressure signal is input as an electronic signal to perform flow rate distribution control based on a preset functional relationship,
By performing the flow rate distribution control by controlling the flow rate control valve with pressure compensation, it is possible to control the differential pressure between the pump discharge pressure and the maximum load pressure to always approach a specified value. Can be stable.

(Examples) Hereinafter, preferred examples of the present invention will be described with reference to the drawings.

FIG. 1 shows the entire configuration of a load sensing hydraulic drive circuit and a control device of the present invention. First, the load sensing hydraulic drive circuit will be described. The hydraulic drive circuit includes, for example, a swash plate type variable displacement hydraulic pump 1 and first and second hydraulic pumps 1 and 2 driven by pressure oil from the hydraulic pump 1.
First hydraulic actuators 2 and 3 arranged between the hydraulic pump 1 and the first hydraulic actuator 2 to control the flow rate and flow direction of the pressure oil supplied from the hydraulic pump 1 to the first hydraulic actuator. Flow control valve with negative pressure compensation 4
And a second hydraulic pressure control unit arranged between the hydraulic pump 1 and the second hydraulic actuator 3 for controlling the flow rate and the flow direction of the pressure oil supplied from the hydraulic pump 1 to the second hydraulic actuator.
Flow control valve 5 with pressure compensation. The first flow control valve 4 has, on its input side, a pressure oil supply line.
Connected to the hydraulic pump 1 via 6, 7 and a return line for pressure oil
It is connected to the tank 10 via 8, 9 and is connected to the first hydraulic actuator 2 on the output side via the main pipelines 11 and 12. On the input side, the second flow control valve 5 is connected to the hydraulic pump 1 via pressure oil supply pipelines 6 and 13, and is connected to the tank 10 via pressure oil return pipelines 8 and 14. The output side is connected to the second hydraulic actuator 3 via the main pipes 15 and 16.

The flow control valves 4 and 5 are pilot operated valves in the illustrated embodiment, and have pilot lines 17a and 17b at both ends.
And 18a, 18b are connected to pilot chambers, and are controlled by pilot hydraulic pressures transmitted to these pilot lines in response to operation signals from operation levers 19, 20 as described later.

The first and second flow control valves 4,5 also have a first
And the pilot lines 21a, 21b and 22a, 22b for picking up the load pressure of the second hydraulic actuators 2, 3 are connected, and the flow control valves 4,5 have the pilot lines 21a, 22a respectively inside the valves. , One of the two main ports on the output side is communicated with, and the pilot conduits 21b, 22b are respectively communicated with the other main ports on the output side inside the valves. The pilot lines 21a, 21b and 22a, 22b have check valves 23a, 23b for selecting the maximum pressure, respectively.
And 24a, 24b are provided, and the downstream side thereof is connected to a common pilot line 25, and the common pilot line 25
Is further connected to the return line 8 via a throttle 26.
As a result, the maximum pressure of the load pressure of the first and second hydraulic actuators is transmitted to the pilot line 25, and it is used as will be described later.

The hydraulic pump 1 is provided with a load sensing regulator 27 that controls the pump discharge flow rate so that the discharge pressure of the hydraulic pump 1 becomes higher than the maximum load pressure of the hydraulic actuators 2 and 3 by a predetermined value.

The load sensing regulator 27 has a pilot operated control valve 31 connected by pipelines 29, 30 between the supply pipeline 6 and a hydraulic cylinder 28 that controls the amount of swash plate tilt of the hydraulic pump 1. One of the pilot chambers is connected to a pipe 29 via a pilot pipe 32, the other pilot chamber is connected to a pilot pipe 25 for maximum load pressure transmission via a pilot pipe 33, and this pilot chamber is also connected. A spring 34 for pressure setting is provided on the operation section on the side where the pressure is present.

The control valve 31 can be continuously provided between a first position for communicating the pipe line 30 with the pipe line 29 and a second position for communicating the pipe line 30 with the tank 10 as shown in the figure. Pilot line
When the pressing force due to the discharge pressure of the hydraulic pump 1 guided by 32 becomes larger than the sum of the pressing force due to the maximum load pressure guided to the pilot line 33 and the force of the spring 34, the control valve 34 is directed to the first position. The hydraulic cylinder 28 by displacing it and supplying pressure oil to the hydraulic cylinder 28 to reduce the tilt amount of the swash plate of the hydraulic pump 1 and to reduce the discharge flow rate thereof. When the sum of the force of 34 and the force of the discharge pressure of the hydraulic pump 1 guided by the pilot pipe 32 becomes larger, the control valve 34 is displaced toward the second position and the pressure oil in the hydraulic cylinder 28 is displaced. The oil is discharged to the tank 10, and the tilt amount of the swash plate of the hydraulic pump 1 is increased to increase the discharge flow rate. As a result, when the load acting on the hydraulic actuators 2 and 3 becomes lighter and the maximum load pressure becomes smaller, the discharge pressure of the hydraulic pump 1 decreases, and when the load becomes heavier and the maximum load pressure becomes higher, the discharge pressure increases and the hydraulic pressure increases. The discharge pressure of the pump 1 is a hydraulic actuator.
The pressure is kept higher than the maximum load pressure of a few by a set pressure corresponding to the strength of the spring 34. In other words, the pump discharge flow rate is controlled so that the differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the hydraulic actuators 2 and 3 is maintained at a constant value set by the spring 34.

As described above, the flow control valves 4 and 5 are valves with pressure compensation. By using the load sensing regulator 27 together with such a flow control valve with pressure compensation, the discharge flow rate of the hydraulic pump 1 can be suppressed to a necessary minimum and energy can be saved, and pressure compensation control of the flow control valve with pressure compensation can be performed. The required minimum differential pressure before and after that can be secured, pressure compensation control of the flow rate of the pressure oil supplied to the hydraulic actuator can be reliably performed, and stable control of the hydraulic actuators 2 and 3 can be performed.

As the flow control valves 4 and 5 with pressure compensation, various conventionally known structures can be used, but in the illustrated embodiment, as an example, the spool valve described in JP-A-56-66567 is used. is doing. Roughly speaking, this valve uses a valve at the corner of the spool land so that the outflow angle decreases as the spool opening (spool stroke amount) increases.
The main notch that is inclined at a predetermined angle and the auxiliary notch that has a predetermined edge angle and that is narrower than the main notch are provided, which increases the load pressure and the differential pressure across the valve to increase the fluid force. When the spool opening is about to decrease due to an increase in the flow rate, the outflow angle is decreased to reduce the rate of increase in the fluid force, and the decrease in the opening area of the valve orifice is suppressed, reducing the differential pressure across the valve by half. The amount of increase in the flow rate proportional to and the amount of decrease in the flow rate due to the decrease in the opening area of the orifice portion cancel each other out, and the flow rate is kept substantially constant.

The load-sensing hydraulic drive circuit configured as described above is controlled by the control device of the present invention shown by reference numeral 40.

The control device 40 is connected to the supply pipeline 6 and the pilot pipeline 25 via pipelines 41 and 42, respectively, and guides the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the hydraulic actuators 2 and 3 and the difference between them. A differential pressure gauge 43 that outputs a differential pressure signal at a level proportional to the pressure, and four electromagnetic proportional control valves 44a, 44b and 45a, 45b connected to the pilot lines 17a, 17b and 18a, 18b, respectively.
And the operation signals of the operation levers 19 and 20 and the differential pressure gauge 43 described above.
Differential pressure signal is input, the flow rate command value for the first and second hydraulic actuators 2, 3 is calculated based on these signals, and the flow rate command value signal is output to the electromagnetic proportional control valves 44a, 44b and 45a, 45b. And a control device body 46 for

The differential pressure gauge 43 has a structure as shown in FIG. 2 as an example. That is, the differential pressure gauge 43 includes a body 50 having pressure oil supply ports 47 and 48 connected to the pipelines 41 and 42 and a pressure oil discharge port 49 connected to the tank, and a cylinder 51 attached to the body 50. And a piston 52 housed in the cylinder 51 and opposed to receive two pressures from the supply ports 47, 48 and having pressure receiving portions 52a, 52b of the same area,
A shaft 53 which is made of a non-magnetic material and transmits the displacement and force of the piston 52, a spring 54 which is housed in the cylinder 51 and receives a force of the piston 52 and gives a displacement proportional to the force to the piston 52, and a non-magnetic material. The case 55 attached to the cylinder 51, the core 56 made of a magnetic material, attached to the tip of the shaft 53 and accommodated in the case 55, and having the same displacement as the piston 52 in the case 55, and the case 55. A displacement sensor 57 that is fixed to the outer periphery and converts the displacement of the core 56 into an electric signal, and an amplifier 59 that is housed in a cover 58 attached to the cylinder 51, amplifies the electric signal from the displacement sensor 57, and outputs the amplified electric signal to the outside. And a spring 60 mounted between the piston 52 and the body 50.

In the differential pressure gauge 43 configured as described above, the supply port 4
The pump discharge pressure Pp and the maximum load pressure Pam act on the pressure receiving portions 52a and 52b of the piston 52 through 7,48. At this time, if the pressure receiving area is A, then Pp> Pam, so piston 52 has A
The force of × (Pp-Pam) acts on the upper part of the figure. This force causes the piston 52 to displace against the springs 54, 60 that elastically support the piston in a pre-compressed state, and the core 56
Is similarly displaced. Set the spring constants of springs 54 and 60 to K ₁ and K ₂
Then, this displacement S is S = A × (Pp−Pam) / (K ₁ / K ₂ ). The displacement sensor 57 converts this displacement into an electric signal,
Amplified by amplifier 59 and output. The displacement sensor 57 is preferably a non-contact type because oil is present in the portion of the core 56 that is displaced, and is, for example, a differential transformer type or a magnetoresistive element type. For this reason, the shaft 53 and the case 55 are made of a non-magnetic material. Further, in any of these types of displacement sensors, the relationship of the displacement signal level E with respect to the displacement S has good linearity and is in a linear proportional relationship. Therefore, the proportional constant is K
When an electric signal level E is, E = K · S = { K · A / (K 1 + K 2)} a (Pp-Pam). Here, since A, K ₁ and K ₂ are all constants, the electric signal level E is the pressure difference (Pp−−) between the pump discharge pressure and the maximum load pressure.
It becomes a value proportional to Pam). In this way, since the pressure difference between the two pressures is applied by the pressure receiving portions of the piston 52 facing each other, the respective pressures are led to different pressure sensors and the respective electric signals are obtained, and then the difference between them is obtained and is equivalent to the pressure difference. It is possible to measure the differential pressure with high accuracy even when the differential pressure is increased, without causing an error due to the nonlinearity of the output with respect to the pressure in the pressure sensor and the hysteresis with respect to the increase and decrease of the pressure as in the case of obtaining an electric signal for .

In the illustrated embodiment, since the differential pressure gauge 43 only needs to measure the differential pressure when Pp> Pam, the spring 60 may be omitted, and in this case, the structure is simplified and the output electric signal level E at this time is reduced. The relation with the differential pressure is as follows: E = {K · A / K ₁ } (Pp-Pam).

The electromagnetic proportional control valves 44a, 44b and 45a, 45b can be configured as shown in FIG. 3 as an example. This electromagnetic proportional control valve is an example of an electromagnetic proportional pressure reducing valve, and includes a proportional solenoid portion 62 and a pressure reducing valve portion 63. Proportional solenoid
62 has a known structure including a proportional solenoid and an iron core (neither is shown), and the proportional solenoid has terminals 64a and 64b.

The pressure reducing valve portion 63 includes a pressure oil supply port 67 connected to the auxiliary pump 65 via a supply pipeline 66, a pressure oil discharge port 69 connected to the tank 10 via a return pipeline 68, and a pilot pipeline. Pressure oil output port connected to 17a, 17b and 18a, 18b
A body 71 having 70, a spool 72 disposed in the body 71, having opposite end surfaces 72a, 72b and having an internal passage 72c formed therein, one end of which is engaged with an iron core of a proportional solenoid 62, and the other end of which is The push rod 73 comes into contact with the end surface 72a of the spool 72.

When current is supplied to the proportional solenoid from the terminals 64a and 64b, a force proportional to this is applied to the iron core of the proportional solenoid 62, and this force is applied to the iron core and the end face 72a of the spool 72 via the engaging lower push rod 73. Be transmitted to. This makes the spool 72
Moves to the right from the position shown in the figure to connect the internal passage 72c with the supply port 67, and the supply port 67 and the output port 70 communicate with each other via the internal passage 72c. This results in output port 7
The oil pressure in 0 rises, and the force acting on the end surface 72b of the spool 72 also rises. When this force becomes larger than the pressing force of the push rod 73 (force applied to the iron core of the proportional solenoid 62), the spool 72 moves to the left, the internal passage 72c communicates with the discharge port 69, and the output port 70 And the discharge port 69 communicate with each other through the internal passage 72c. As a result, the oil pressure of the output port 70 is reduced, and the force received by the end surface 72b is also reduced. When this force becomes smaller than the pressing force of the push rod 73, the spool 72 moves to the right in the figure again.

In this way, the spool 72 of the pressure reducing valve portion 63 operates by receiving the force applied to the iron core of the proportional solenoid portion 62, so that the hydraulic pressure generated at the output port 70 eventually becomes equal to the level of the current supplied to the proportional solenoid. It will be proportional.

Returning to FIG. 1 again, the control device main body 46 has the operation lever 1
9, 20 and the differential pressure gauge 43 are electrically connected to the multiplexer 80 for inputting the operation signal of the operation lever and the differential pressure signal of the differential pressure gauge, and switching, selecting and outputting them.
An A / D converter 81 for converting the operation signal and the differential pressure signal output through the digital signal into a digital signal, and a memory 82 storing a control program for executing a processing procedure described later,
A central processing unit or CPU 83 for processing digital signals from the A / D converter 81 in accordance with a control program stored in the memory 82, and solenoid proportional control valves 44a, 44b and 45a, 45.
An output device or driver 84 that is electrically connected to b and amplifies the output from the CPU 83 and outputs it to these electromagnetic proportional control valves.
It consists of

FIG. 4 is a conceptual representation of the contents of control performed by the control device body 46. That is, in the range where the differential pressure signal ΔPp is larger than the predetermined value ΔPp ₀ , the control coefficient ΔK
Based on the predetermined functional relationship 90 between the differential pressure signal ΔPp and the control coefficient ΔKp, p becomes 1 and the control coefficient ΔKp becomes smaller than 1 as the differential pressure signal ΔPp decreases below a predetermined value ΔPp _0. , Control coefficient Δ from the input differential pressure signal ΔPp
Kp is calculated, and flow rate signals Qi ₁ and Qi ₂ and control coefficient ΔKp for each hydraulic actuator 2 and 3 obtained from each operation signal X _L1 and X _L2
Are multiplied as indicated by reference numerals 91 and 92, and the flow rate command values Q ₁ and Q ₂ are determined through such a procedure. Here, the predetermined value ΔPp ₀ of the differential pressure signal in the functional relationship 90 between the differential pressure signal and the control coefficient is a value corresponding to the set pressure of the spring 34 in the load sensing regulator 27 described above.

And, in particular in the illustrated embodiment, preferably each operation signal
After multiplying the flow rate signals Qi ₁ and Qi ₂ obtained from X _L1 and X _L2 for the hydraulic actuators 2 and 3 by the control coefficient ΔKp, the calculation results Q′i ₁ , The maximum flow rate change limiting control is performed to limit the flow rate change amount indicated by Q′i ₂ to an appropriate maximum flow rate change amount suitable for the operation of the corresponding hydraulic actuators 2, 3.

Further, in the illustrated embodiment, the operation amounts X _L1 and X _{L2 of the} operation levers 19 and 20
With respect to the respective hydraulic actuators 2, 3, the metering characteristics 95, 96 that give the optimum speed are predetermined, and the operating levers 19, 96 are based on the metering characteristics 95, 96.
The flow rate signals Qi ₁ and Qi ₂ for the respective hydraulic actuators 2 and 3 are obtained from the 20 manipulated variables X _L1 and X _L2 .

Flow rate command value Q obtained through maximum flow rate change limiting control 93, 94
_As shown by reference numerals 97 and 98, ₁ and Q ₂ are sorted by positive and negative, respectively, and the corresponding flow rate proportional control valves 44a and 44b.
And 45a and 45b.

The above-described functional relationship 90 between the differential pressure signal ΔPp and the control coefficient ΔKp and the metering characteristics 95 and 96 are stored in the memory 82 as a function table and a metering characteristic map, respectively.

Next, the operation of the control device 40 configured as described above will be described according to the processing procedure represented by the flowcharts shown in FIGS.

First, in step S ₁ , the operation signals X _L1 and X _L2 of the operation levers 19 and 20
And captures the differential pressure signal ΔPp the differential pressure gauge 43, in step S _2, the flow rate signal Qi ₁ from the operation from the metering characteristic map signal X _L1, X _L2 being set metering characteristics 95 and 96 described above, Qi ₂ In step S ₃ , the control coefficient ΔKp is read from the differential pressure signal ΔPp from the function table in which the functional relationship 90 between the differential pressure signal ΔPp and the control coefficient ΔKp is set, and in step S ₄ , the flow rate signal is read. Qi ₁ and Qi ₂ are multiplied by the control coefficient ΔKp to obtain flow rate command values Q′i ₁ and Q′i ₂ .

At this time, the sum of the required consumption flow rates of the first and second hydraulic actuators 2 and 3 is smaller than the maximum discharge flow rate of the hydraulic pump 1, and the load sensing regulator 27 described above functions effectively, thereby When the pump discharge flow rate is controlled so that the discharge pressure Pp of the pump 1 becomes higher than the maximum load pressure Pam of the first and second hydraulic actuators 2 and 3 by the set pressure determined by the strength of the spring 34, , The differential pressure signal ΔPp indicates a value near a predetermined value ΔPp ₀ ,
The control coefficient ΔKp obtained from the functional relation 90 in step S ₃
The value of is a constant 1. Therefore, in step S ₄ , the flow rate signals Qi ₁ and Qi ₂ obtained from the metering characteristics 95 and 96 become the flow rate command values Q′i ₁ and Q′i _{2 as they are} , and the operation signals X _{L1 of the} operating levers 19 and 20 are transmitted. , X _L2 is indicated.

On the other hand, the sum of the required consumption flow rates of the first and second hydraulic actuators 2 and 3 becomes larger than the maximum discharge flow rate of the hydraulic pump 1, and the operation of the load sensing regulator 27 causes the discharge pressure Pp of the hydraulic pump 1 to reach the maximum load. When the pressure drops by a set pressure higher than the pressure Pam, the differential pressure signal Δ
Pp becomes a value smaller than the predetermined value ΔPp _0, and the control coefficient ΔKp obtained from the functional relationship 90 in step S ₃ becomes a value smaller than 1. Therefore, in step S ₄ , the metering characteristics
The flow rate signals Qi ₁ and Qi ₂ obtained from 95 and 96 are respectively reduced as flow rate command values Q′i ₁ and Q′i ₂ by the same rate according to the value of the control coefficient ΔKp. , Operating lever
A flow rate smaller than the flow rate indicated by the operation signals X _L1 and X _{L2 of} 19, 20 is indicated.

As a case where such an operating state occurs, as described above, when the hydraulic pump 1 is further provided with the horsepower control device, the pump discharge pressure increases during operation,
When the engine that drives the hydraulic pump 1 is about to stall, the horsepower control device functions by giving priority to the control of the load sensing regulator 27 to reduce the discharge flow rate of the hydraulic pump 1, or when the two hydraulic actuators 2, There is a case where the total of the required consumption flow rates instructed by the operating means exceeds the maximum capacity (maximum discharge flow rate) of the hydraulic pump 1 when 3 are simultaneously driven.

Next, in step S ₅ , maximum flow rate change limiting control is performed to obtain flow rate command values Q ₁ and Q ₂ , and in step S ₆ the flow rate command value Q 1
₁ and Q ₂ are output to the above-mentioned electromagnetic proportional control valves 44a, 44b and 45a, 45b to issue a flow rate command.

The maximum flow rate change limiting control performed in step S ₅ can be performed as shown in the flowchart of FIG. 6 as an example, and in this flowchart, the same applies to the first and second hydraulic actuators 2 and 3. Since the contents are controlled, for convenience of explanation, the two flow rate command values Q′i ₁ and Q′i ₂ are collectively represented by Q′i, and the two flow rate command values Q ₁ and Q ₂ are collectively represented by Q. It is represented.

First, in Step S _5-1, obtains a difference Z between the flow rate command Q _-1 to flow rate command value Q'i the previous electromagnetic proportional control valve of the previous calculation result, in step S _5-2, 1 sampling time ΔT
The appropriate maximum change amount C of the flow rate change ΔQ per hit is selected for the corresponding hydraulic actuator. Here, the flow rate change amount ΔQ corresponds to the acceleration of the hydraulic actuator, and the appropriate maximum flow rate change amount C is set to a different value for each hydraulic actuator so as not to cause a shock in the hydraulic actuator. That is, a small value is set for the hydraulic actuator that drives a large load, and a large value is set for the hydraulic actuator that drives a small load.

Then in step S _5-3, whether the absolute value of the difference Z of the command value obtained in Step _5-1 is smaller than the flow rate change amount ΔQ determined in Step S _5-2 is determined, is determined to be smaller when moved to Step S _5-4, the flow rate command value Q'i the current calculation result is output as the flow rate command value Q. If the difference Z is determined to be greater than the flow rate change amount ΔQ is moved to Step S _5-5, wherein the sign of the difference Z of the command value is determined, the procedure S _5-6 in the case of a positive Z If negative, go to step S _5-7 .
In step S _5-6 , the flow rate change amount ΔQ obtained in step S _5-2 is added to the flow rate command value Q _-1 to the previous solenoid proportional control valve, and this is set as the flow rate command value Q, and the procedure is performed. In S _5-7 , the flow rate change value ΔQ obtained in step S _5-2 is subtracted from the previous flow rate command value Q ₋₁ to the solenoid proportional control valve, and the result is subtracted from the flow rate command value Q 1.
And As a result, the flow rate change amount is limited to an appropriate maximum flow rate change amount suitable for the operation of the corresponding hydraulic actuator.

Therefore, according to the control device of the illustrated embodiment, when the total of the required consumption flow rates of the first and second hydraulic actuators 2 and 3 is smaller than the maximum discharge flow rate of the hydraulic pump 1, the control is performed as described above. The coefficient ΔKp has a value of 1, and the flow rate signals Qi ₁ and Qi ₂ obtained from the metering characteristics 95 and 96 become the flow rate command values Q′i ₁ and Q′i _{2 as they are} , and the flow rate control valves 4,5 have their flow rates. The flow rates indicated by the signals Qi ₁ and Qi ₂ are controlled so that the hydraulic actuators 2 and 3 are controlled in speed according to the operation signals X _L1 and X _L2 of the operation levers 19 and 20, respectively. Also at this time,
When an operation that changes the operating speed of the hydraulic actuators 2 and 3 is performed, such as when the operating levers 19 and 20 are turned on and returned, the maximum flow rate change limit control is used to adjust the hydraulic actuator Since the amount of change in the optimum maximum flow rate suitable for the operation is limited, shock is not generated in the hydraulic actuator, and the operation is controlled at an appropriate maximum acceleration or deceleration.

Further, the sum of the required consumption flow rates of the first and second actuators 2 and 3 becomes larger than the maximum discharge flow rate of the hydraulic pump 1, and the discharge pressure Pp of the hydraulic pump 1 becomes greater than the maximum load pressure Pam by the operation of the load sensing regulator 27. When the set pressure drops from a high pressure, the control coefficient ΔKp is 1
The flow rate command values Q′i ₁ and Q′i ₂ are smaller than the flow rate signals Qi ₁ and Qi ₂ obtained from the metering characteristics 95 and 96.
It becomes a small value that decreases at the same rate according to the value of the control coefficient ΔKp. Therefore, the flow control valves 4 and 5 are controlled so that the small opening indicated by the flow command values Q′i ₁ and Q′i ₂ , respectively, are obtained, and the supply flow rates to the hydraulic actuators 2 and 3 are reduced at the same rate. To do. That is, while the discharge flow rate of the hydraulic pump 1 is distributed according to the ratio of the operation amounts of the operation levers 19 and 20, the supply flow rates to the hydraulic actuators 2 and 3 are reduced at the same rate. This is generally called non-saturation control of the discharge flow rate of a hydraulic pump. As a result, when it is no longer possible to maintain the discharge pressure of the hydraulic pump higher than the maximum load pressure by the set pressure, most of the pump discharge flow will flow to the low-pressure side hydraulic actuator, and pressure oil will be supplied to the high-pressure side hydraulic actuator. It is possible to avoid the situation of disappearing, and the composite operability of work elements is greatly improved.

In such a non-saturation control, when the discharge pressure of the hydraulic pump 1 is greatly reduced and steep, and the control coefficient ΔKp changes to a small value of 1 or less for a short time, the flow rate command values Q′i ₁ , Q are obtained. The flow rate change of ′ i ₂ becomes large. Therefore, if the flow rate control valve is directly controlled by the flow rate command values Q′i ₁ and Q′i ₂ , the hydraulic actuators 2, 3
The change in speed will increase and shock will be generated in the hydraulic actuator, which is dangerous. In such a case, in the illustrated embodiment, the maximum flow rate change limiting control described above limits the flow rate change amount ΔQ to an appropriate maximum flow rate change amount C suitable for the operation of the corresponding hydraulic actuator. The hydraulic actuator is controlled at an appropriate maximum acceleration or deceleration to prevent shock from occurring in the hydraulic actuator. As a result, smoother non-saturation control can be performed.

As described above, the optimum maximum flow rate change amount C is set to an optimum value for each hydraulic actuator.

Although the embodiments of the present invention have been described with reference to the drawings, the present invention is not limited to these embodiments, and various modifications and changes can be made within the scope of the spirit of the present invention. For example, in the above embodiments, the maximum flow rate change limiting control is also performed. However, this is preferable, and even without this, the basic control in the control means of the present invention, that is, the operating lever. It is possible to reduce the supply flow rate to each hydraulic actuator 2, 3 at the same rate while distributing the discharge flow rate of the hydraulic pump 1 according to the ratio of the manipulated variables 19, 20.

Regarding the maximum flow rate change limiting control itself, in the above-described embodiment, the sampling time ΔT is used to control the flow rate change amount.
Although ΔT is kept constant and ΔQ is changed, it is also possible to set ΔQ as a constant unit change amount and change ΔT.

Although the hydraulic pump 1 has been described as a variable displacement type, the hydraulic pump 1 is not limited to this and may be a fixed displacement type hydraulic pump. In this case, the load sensing regulator may be arranged such that the relief valve normally provided on the discharge side of the hydraulic pump is modified, and the maximum discharge pressure of the plurality of hydraulic actuators is guided to the spring side thereof in opposition to the pump discharge pressure. As a result, the discharge pressure of the hydraulic pump can be kept higher than the maximum load pressure of the plurality of hydraulic actuators by a predetermined value corresponding to the spring.

In the above embodiment, the case where the two hydraulic actuators 2 and 3 are driven by the hydraulic pump 1 has been described.
The present invention is not limited to this, and can be applied to a hydraulic drive circuit having three or more hydraulic actuators.

Further, as described above, the load compensating valve devices 4 and 5 with pressure compensation are not limited to the spool valve described in JP-A-56-66567, and various types can be used. For example, a classically well-known flow control throttle valve that changes the throttle opening with a half-moon shaped groove and a pressure compensating valve that controls the variable orifice by displacing the spool by introducing the differential pressure before and after it. A flow control valve with pressure compensation may be used.

(Effects of the Invention) As is apparent from the above, in the controller for the load sensing hydraulic drive circuit of the present invention, the difference indicating the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second hydraulic actuators. The control coefficient becomes 1 in the range where the pressure signal is larger than the predetermined value, and the control coefficient becomes smaller than 1 as the differential pressure signal decreases below the predetermined value. Since the control coefficient is determined from the input differential pressure signal based on the functional relationship, and the flow rate signal for each hydraulic actuator obtained from each operation signal is multiplied by the control coefficient, the flow rate command value is determined. When the required consumption flow rate of the hydraulic pump becomes larger than the maximum discharge flow rate of the hydraulic pump and the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure becomes small, it is input to the control means. The generated differential pressure signal becomes equal to or less than a predetermined value, the control means obtains a value smaller than 1 as a control coefficient from a predetermined functional relationship, and the flow rate command value for each hydraulic actuator is reduced at the same rate. Since the opening of the control valve is controlled, the flow rate supplied to each hydraulic actuator is uniformly reduced while the discharge flow rate of the hydraulic pump is distributed according to the ratio of the operation amount of the operating means. Even when the flow rate becomes larger than the maximum discharge flow rate of the hydraulic pump, the pressure oil can be supplied to each hydraulic actuator regardless of the magnitude of the load pressure, and the composite operability can be greatly improved. In addition, by inputting the differential pressure signal as an electric signal and performing flow rate distribution control based on a preset functional relationship, and performing the flow rate distribution control by controlling the flow control valve with pressure compensation, the pump discharge pressure and maximum It is possible to control the pressure difference with the load pressure so that it always approaches the specified value.
This allows stable pump control by the load sensing regulator.

[Brief description of drawings]

FIG. 1 is a schematic diagram showing a control device of a load sensing hydraulic drive circuit according to an embodiment of the present invention together with a load sensing hydraulic drive circuit, and FIG. 2 is a sectional view of a differential pressure gauge of the control device shown in FIG. FIG. 3 is a cross-sectional view of an electromagnetic proportional control valve of the control device, FIG. 4 is a conceptual diagram schematically showing the control content of the control device, and FIG.
FIG. 6 is a flowchart showing the processing means performed by the control device, and FIG. 6 is a flowchart showing the processing procedure of the maximum flow rate change limiting control shown in FIG. In the figure, reference numeral 1 ... hydraulic pump, 2,3 ... hydraulic actuator, 4,5 ... pressure compensating flow control valve, 19,20 ... operating lever (operating means), 27 ... load sensing regulator, 40 ...... Control device, 43 …… Differential pressure gauge (differential pressure detection means),
44a, 44b, 45a, 45b ... Electromagnetic proportional control valve, 46 ... Control device body (control means), 90 ... Functional relationship between differential pressure signal and control coefficient, Pp ... Discharge pressure, Pam ... Maximum load Pressure, ΔPp ...
… Differential pressure, X _L1 , X _L2 …… Operation signal, Qi ₁ , Qi ₂ …… Flow signal,
Q ₁ , Q ₂ ... Flow rate command value

Claims (2)

[Claims] 1. At least one hydraulic pump, a plurality of hydraulic actuators driven by the pressure oil discharged from the hydraulic pumps, and a connection between the hydraulic pumps and the respective hydraulic actuators, and an operating signal for operating means. Flow control valve with pressure compensation for controlling the flow rate of the pressure oil supplied to the hydraulic actuators, and the pump discharge flow rate so that the discharge pressure of the hydraulic pumps is higher than the maximum load pressure of the hydraulic actuators by a predetermined value. In a control device of a load sensing hydraulic drive circuit including a load sensing regulator for controlling, a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of a plurality of hydraulic actuators is detected, and a differential pressure signal corresponding thereto is output. Inputting the differential pressure detecting means, the operation signal of the operating means, and the differential pressure signal of the differential pressure detecting means. A flow rate command value of each of the hydraulic actuators, and control means for controlling the flow rate control valve with pressure compensation based on the flow rate command value, wherein the control means has a range in which the differential pressure signal is larger than a predetermined value. Then, the control coefficient becomes 1, and when the differential pressure signal becomes equal to or less than a predetermined value, the control coefficient becomes smaller than 1 as the differential pressure signal decreases, and the functional relationship between the differential pressure signal and the control coefficient is preset. Based on the function coefficient, a control coefficient is obtained from the input differential pressure signal, the flow rate signal for each hydraulic actuator obtained from each operation signal is multiplied by the control coefficient, and first calculating means for determining the flow rate command value is provided. A control device for a load-sensing hydraulic drive circuit, which comprises: 2. The control means is preset with an appropriate maximum change amount of the flow rate command value for each actuator, and the flow rate signal is multiplied by a control coefficient by the first computing means, and then the computation is performed. When the resulting change amount of the flow rate command value is larger than the appropriate maximum change amount, second calculation means for correcting the flow rate command value is further provided so as to limit the change amount of the flow rate command value to the appropriate maximum change amount. The control device for a load sensing hydraulic drive circuit according to claim 1, characterized in that the control device has a load sensing hydraulic drive circuit.