JP7163432B2 - Excavators and systems for excavators - Google Patents

Description

The present invention relates to excavators with hydraulic actuators.

There is known a construction machine that prevents a sudden increase in the amount of fuel injection due to a sudden decrease in engine speed when the discharge pressure of a hydraulic pump suddenly increases (see Patent Document 1).

This construction machine performs horsepower control so that the pump absorption torque of the hydraulic pump does not exceed the engine output torque even if the discharge pressure of the hydraulic pump changes. Specifically, when the discharge pressure of the hydraulic pump increases, the discharge amount is reduced so that the pump absorption torque represented by the product of the discharge pressure and the discharge amount becomes a predetermined torque (high torque). Further, when it is determined that the rate of increase in the discharge pressure of the hydraulic pump is equal to or higher than the predetermined speed, the predetermined torque is reduced, and then the pump absorption torque is discharged so as to become the predetermined torque (low torque) after the reduction. reduce the amount. This is to prevent the pump absorption torque from exceeding the engine output torque due to a delayed decrease in the discharge amount when the discharge pressure of the hydraulic pump rises sharply. As a result, wasteful fuel consumption can be suppressed and the operability of hydraulic actuators and the like can be improved.

JP-A-2005-54903

However, the construction machine described above still reduces the discharge amount after detecting a rapid increase in the discharge pressure of the hydraulic pump, and the time required from the reduction of the discharge amount until the discharge pressure actually decreases. is affected by the response delay time of the hydraulic pump. Therefore, depending on the setting of the predetermined torque (low torque) after reduction, the pump absorption torque of the hydraulic pump may increase beyond the engine output torque, or the pump absorption torque may be excessively reduced.

In view of the above, it would be desirable to provide an excavator capable of controlling a hydraulic pump with greater responsiveness.

An excavator according to an embodiment of the present invention includes a lower traveling body, an upper revolving body rotatably mounted on the lower traveling body, an attachment including a working body attached to the upper revolving body, and driving the attachment. a hydraulic actuator; a state quantity acquiring unit that acquires a state quantity of at least one of the hydraulic actuator and the working body; a model state resetting unit that derives a coefficient using the state quantity; a predicted value after a predetermined control cycle in the model based on the derived variable coefficient; and a command value based on the calculated predicted value and a model predictive control unit that calculates

By the means described above, it is possible to provide an excavator capable of controlling the hydraulic pump with higher responsiveness.

1 is a diagram showing a configuration example of an excavator equipped with a drive system according to an embodiment of the present invention; FIG. 1 is a schematic diagram showing a configuration example of a drive system according to an embodiment of the present invention; FIG. 4 is a negative control control diagram showing the relationship between pump flow rate and negative control pressure; FIG. 4 is a horsepower control diagram (PQ diagram) showing the relationship between pump flow rate and pump discharge pressure; FIG. It is a functional block diagram which shows the structural example of a control system. FIG. 4 is a graph showing temporal transitions of pump discharge pressure, displacement volume, and pump absorption torque;

Preferred embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a configuration example of a shovel as a construction machine according to an embodiment of the present invention. In the excavator of FIG. 1, an upper revolving body 3 is mounted on a crawler-type lower traveling body 1 through a revolving mechanism 2 so as to be rotatable around the X axis. In addition, the upper revolving body 3 has an excavation attachment at its front central portion. The excavation attachment includes a boom 4, an arm 5 and a bucket 6 as working bodies, and a boom cylinder 7, an arm cylinder 8 and a bucket cylinder 9 as hydraulic actuators.

A boom stroke sensor 7 s is attached to the boom cylinder 7 to detect the telescopic state of the boom cylinder 7 . An arm stroke sensor 8s is attached to the arm cylinder 8 to detect the extension/contraction state of the arm cylinder 8, and a bucket stroke sensor 9s is attached to the bucket cylinder 9 to detect the extension/contraction state of the bucket cylinder 9. At least one of the stroke sensors may be replaced with an angle sensor that detects the rotation angle of the working body.

FIG. 2 is a schematic diagram of a drive system 100 according to an embodiment of the invention. Drive system 100 mainly includes engine 11 , hydraulic pump 14 , control valve 17 , and controller 30 .

Hydraulic pump 14 is driven by engine 11 . In this embodiment, the hydraulic pump 14 is a variable displacement swash plate type hydraulic pump that varies the discharge amount (displacement volume) per rotation. Displacement volume is controlled by pump regulator 14a. Specifically, the hydraulic pump 14 includes a first hydraulic pump 14L whose displacement volume is controlled by a pump regulator 14aL, and a second hydraulic pump 14R whose displacement volume is controlled by a pump regulator 14aR. Further, in this embodiment, the rotary shaft of the hydraulic pump 14 is connected to the rotary shaft of the engine 11 and rotates at the same rotational speed as the engine 11 . The rotary shaft of the hydraulic pump 14 is connected to a flywheel to suppress fluctuations in rotational speed when the engine output torque fluctuates.

The engine 11 is the driving source of the shovel. In this embodiment, the engine 11 is a diesel engine having a turbocharger as a supercharger and a fuel injection device, and is mounted on the upper revolving body 3 . Note that the engine 11 may be provided with a supercharger as a supercharger.

The control valve 17 is a hydraulic control mechanism that supplies hydraulic fluid discharged by the hydraulic pump 14 to various hydraulic actuators. In this embodiment, the control valve 17 includes control valves 171L, 171R, 172L, 172R, 173L, 173R, 174R, 175L and 175R. The hydraulic actuators include a left traveling hydraulic motor 2L, a right traveling hydraulic motor 2R, a boom cylinder 7, an arm cylinder 8, a bucket cylinder 9, and a turning hydraulic motor 21.

Specifically, the hydraulic pump 14L circulates the hydraulic fluid to the hydraulic fluid tank 22 through the center bypass line 40L communicating with the control valves 171L, 172L, 173L, and 175L. Similarly, the hydraulic pump 14R circulates hydraulic fluid to the hydraulic fluid tank 22 through a center bypass line 40R communicating with the control valves 171R, 172R, 173R, 174R, and 175R.

The control valve 171L is a spool valve that controls the flow rate and flow direction of hydraulic fluid between the left traveling hydraulic motor 2L and the hydraulic pump 14L.

The control valve 171R is a spool valve as a straight travel valve. The control valve 171R switches the flow of hydraulic oil so that hydraulic oil is supplied from the hydraulic pump 14L to each of the left traveling hydraulic motor 2L and the right traveling hydraulic motor 2R in order to improve the straightness of the lower traveling body 2. Specifically, when the left travel hydraulic motor 2L and the right travel hydraulic motor 2R and any other hydraulic actuator are operated at the same time, the hydraulic pump 14L operates to operate the left travel hydraulic motor 2L and the right travel hydraulic pressure. Hydraulic oil is supplied to both motors 2R. In other cases, the hydraulic pump 14L supplies hydraulic fluid to the left traveling hydraulic motor 2L, and the hydraulic pump 14R supplies hydraulic fluid to the right traveling hydraulic motor 2R.

The control valve 172L is a spool valve that controls the flow rate and flow direction of hydraulic oil between the turning hydraulic motor 21 and the hydraulic pump 14L.

The control valve 172R is a spool valve that controls the flow rate and flow direction of the hydraulic fluid between the right traveling hydraulic motor 2R and the hydraulic pumps 14L and 14R.

The control valves 173L and 173R are spool valves that control the flow rate and flow direction of hydraulic oil between the boom cylinder 7 and the hydraulic pumps 14L and 14R, respectively. The control valve 173R operates when a boom operating lever as an operating device is operated, and the control valve 173L operates when the boom operating lever is operated in the boom raising direction by a predetermined lever operation amount or more.

The control valve 174R is a spool valve that controls the flow rate and flow direction of hydraulic oil between the hydraulic pump 14R and the bucket cylinder 9 .

The control valves 175L and 175R are spool valves that control the flow rate and flow direction of hydraulic fluid between the arm cylinder 8 and the hydraulic pumps 14L and 14R, respectively. The control valve 175L operates when an arm operating lever as an operating device is operated, and the control valve 175R operates when the arm operating lever is operated by a predetermined lever operation amount or more.

The center bypass lines 40L, 40R are provided with negative control throttles 40L, 40R between the most downstream control valves 175L, 175R and the hydraulic oil tank 22, respectively. The negative control is abbreviated as "negative control" below. The negative control throttles 41L and 41R restrict the flow of hydraulic oil discharged from the hydraulic pumps 14L and 14R, thereby generating negative control pressure upstream of the negative control throttles 41L and 41R.

The relief valves 19L and 19R are valves that limit the negative control pressure to a predetermined relief pressure or less. In this embodiment, the relief valves 19L, 19R are connected in parallel to the negative control throttles 41L, 41R in the center bypass lines 40L, 40R, respectively.

The controller 30 is a functional element that controls the shovel, and is, for example, a computer that includes a CPU, RAM, ROM, NVRAM, and the like.

In this embodiment, the controller 30 determines the operation details of various operation devices (for example, presence/absence of lever operation, lever operation direction, lever operation amount, etc.) based on the output of a pilot pressure sensor (not shown). Detect electrically. The pilot pressure sensor is an example of an operation content detection unit that measures pilot pressure generated when various operating devices such as an arm operating lever and a boom operating lever are operated. However, the operation content detection unit may be configured using a sensor other than the pilot pressure sensor, such as an inclination sensor that detects the inclination of various operation levers.

Also, the controller 30 electrically detects the operating conditions of various hydraulic actuators based on the outputs of the sensors S1 to S4.

The pressure sensors S1 and S2 detect negative control pressures generated upstream of the negative control throttles 41L and 41R, and output the detected values to the controller 30 as electrical negative control pressure signals.

The pressure sensors S3 and S4 detect the discharge pressure of the hydraulic pumps 14L and 14R and output the detected values to the controller 30 as electrical discharge pressure signals.

The engine speed setting dial 75 is a dial for setting the engine speed, and is provided in the cabin, for example, so that the operator can switch the engine speed in a plurality of steps.

The controller 30 causes the CPU to execute programs corresponding to various functional elements in accordance with the operation contents of various operating devices and the operating conditions of various hydraulic actuators.

Next, referring to FIG. 3, the process of controlling the discharge amount of the hydraulic pump 14 according to the negative control pressure by the controller 30 will be described. FIG. 3 is a negative control control diagram showing the relationship between the discharge amount of the hydraulic pump 14 (hereinafter referred to as "pump flow rate") and the negative control pressure. Apply negative control pressure.

In this embodiment, the controller 30 increases or decreases the displacement volume of the hydraulic pump 14L by increasing or decreasing the control current to the pump regulator 14aL to increase or decrease the tilt angle of the swash plate of the hydraulic pump 14L. For example, the lower the negative control pressure, the more the controller 30 increases the control current to increase the displacement volume of the hydraulic pump 14L. Although the displacement volume of the hydraulic pump 14L will be described below, the same explanation applies to the displacement volume of the hydraulic pump 14R.

Specifically, the hydraulic oil discharged by the hydraulic pump 14L passes through the center bypass pipe 40L and reaches the negative control throttle 41L to generate negative control pressure upstream of the negative control throttle 41L.

For example, when the control valve 175L moves to operate the arm cylinder 8, hydraulic fluid discharged from the hydraulic pump 14L flows into the arm cylinder 8 via the control valve 175L. Therefore, the amount reaching the negative control throttle 41L is reduced or eliminated, and the negative control pressure generated upstream of the negative control throttle 41L is reduced.

The controller 30 increases the control current for the pump regulator 14aL according to the decrease in the negative control pressure detected by the pressure sensor S1. As the control current from the controller 30 increases, the pump regulator 14aL increases the tilt angle of the swash plate of the hydraulic pump 14L to increase the displacement volume. As a result, sufficient working oil is supplied to the arm cylinder 8, and the arm cylinder 8 is properly driven.

After that, when the control valve 175L is returned to the neutral position to stop the operation of the arm cylinder 8, the hydraulic oil discharged from the hydraulic pump 14L reaches the negative control throttle 41L without flowing into the arm cylinder 8. Therefore, the amount reaching the negative control throttle 41L increases, and the negative control pressure generated upstream of the negative control throttle 41L increases.

The controller 30 reduces the control current for the pump regulator 14aL according to the increase in the negative control pressure detected by the pressure sensor S1. As the control current from the controller 30 is reduced, the pump regulator 14aL reduces the tilt angle of the swash plate of the hydraulic pump 14L to reduce the displacement volume. As a result, the pump flow rate is reduced, and the pressure loss (pumping loss) when the hydraulic oil discharged from the hydraulic pump 14L passes through the center bypass pipe 40L is suppressed.

A pump control line represented by a solid line shows the tendency of the pump flow rate to increase as the negative control pressure decreases. Hereinafter, the control of the pump flow rate based on the negative control pressure as described above will be referred to as "negative control". By negative control control, the drive system 100 can suppress wasteful energy consumption in a standby state in which the hydraulic actuator is not operated. This is because the pumping loss caused by the hydraulic oil discharged by the hydraulic pump 14 can be suppressed. In addition, the drive system 100 can supply necessary and sufficient working oil to the hydraulic actuator from the hydraulic pump 14 when operating the hydraulic actuator.

Further, drive system 100 executes horsepower control in parallel with negative control control. Horsepower control reduces the pump flow rate in accordance with an increase in the discharge pressure of the hydraulic pump 14 (hereinafter referred to as "pump discharge pressure"). This is to prevent the occurrence of overtorque. That is, this is to prevent the absorption horsepower (pump absorption torque) of the hydraulic pump, which is represented by the product of the pump discharge pressure and the pump flow rate, from exceeding the output horsepower (engine output torque) of the engine.

FIG. 4 is a horsepower control diagram (PQ diagram) showing the relationship between the pump flow rate and the pump discharge pressure, in which the vertical axis represents the pump flow rate and the horizontal axis represents the pump discharge pressure. The horsepower control line shows the trend of increasing pump flow as pump discharge pressure decreases. Also, the horsepower control line is determined according to the target pump absorption torque Tt, and shifts to the upper right in the drawing as the target pump absorption torque Tt increases. FIG. 4 shows that the target pump absorption torque Tt1 corresponding to the horsepower control line represented by the solid line is greater than the target pump absorption torque Tt2 corresponding to the horsepower control line represented by the dashed line. Note that the target pump absorption torque Tt is a value that is set in advance as the allowable maximum value of the pump absorption torque that the hydraulic pump 14 can use. In this embodiment, the target pump absorption torque Tt is preset as a fixed value, but may be a variable value.

In this embodiment, when operating the hydraulic pump 14L at the target pump absorption torque Tt, the controller 30 controls the displacement volume of the hydraulic pump 14L according to the horsepower control line as shown in FIG. Specifically, the target displacement volume is derived from the pump flow rate corresponding to the pump discharge pressure detected by the pressure sensor S3. The controller 30 then outputs a control current corresponding to the target displacement volume to the pump regulator 14aL. The pump regulator 14aL increases or decreases the tilt angle of the swash plate according to the control current to bring the displacement volume to the target displacement volume. With such feedback control of the pump absorption torque, the controller 30 can operate the hydraulic pump 14L at the target pump absorption torque Tt even if the pump discharge pressure fluctuates due to fluctuations in the load on the hydraulic actuator. The same applies to the pump 14R.

However, as long as such feedback control is used, the controller 30 cannot eliminate the response delay time required from detecting a change in the pump discharge pressure to actually changing the pump flow rate.

Therefore, the controller 30 employs model predictive control to eliminate this response delay time. In this embodiment, the controller 30 predicts the state quantity of the hydraulic pump 14 after a predetermined time based on the state quantity of the hydraulic actuator and the state quantity of the hydraulic pump 14 and derives the command value for the hydraulic pump 14 . The state quantity of the hydraulic actuator is a state quantity that responds to fluctuations in the load of the hydraulic actuator faster than the discharge pressure of the pump. etc. The state quantities of the hydraulic pump 14 include pump discharge pressure, displacement volume, swash plate tilt angle, and the like. The state quantity of the hydraulic pump 14 after a predetermined time is a state quantity to be predicted, and includes pump discharge pressure and the like. Also, the command value for the hydraulic pump 14 includes a displacement command value and the like.

Here, a configuration example of a control system employed when executing model predictive control will be described with reference to FIG. Note that FIG. 5 is a functional block diagram showing a configuration example of the control system.

Specifically, the control system of FIG. 5 includes a controller 30 , an actuator state quantity acquisition section 31 and a pump discharge pressure detection section 32 . The controller 30 also includes a horsepower control section 33 , a model predictive control section 34 and a model state resetting section 35 .

The actuator state quantity acquisition unit 31 is a functional element that acquires the state quantity of the hydraulic actuator, and includes, for example, a displacement sensor that detects the expansion/contraction state of the hydraulic cylinder as the hydraulic actuator. The displacement sensor acquires the expansion/contraction velocity v of the hydraulic cylinder based on the detected displacement, and outputs the value to the model predictive control unit 34 as an electrical expansion/contraction velocity signal. In this embodiment, the displacement sensors include boom stroke sensor 7s, arm stroke sensor 8s, and bucket stroke sensor 9s. The actuator state quantity acquisition unit 31 may be an angle sensor that detects the rotation angle of the working body. In this case as well, the actuator state quantity acquisition unit 31 acquires the expansion/contraction speed v of the hydraulic cylinder based on the detected angle, and outputs the value to the model prediction control unit 34 as an electrical expansion/contraction speed signal. .

The pump discharge pressure detector 32 is a functional element that detects the pump discharge pressure. In this embodiment, the pump discharge pressure detector 32 is pressure sensors S3 and S4 that detect the pump discharge pressure Pd. The pressure sensors S3 and S4 output the detected values to the model predictive control section 34 as electrical pump discharge pressure signals.

The horsepower control unit 33 is a functional element that controls the absorption horsepower (pump absorption torque) of the hydraulic pump 14 . In this embodiment, the horsepower control unit 33 outputs a control current corresponding to the displacement command value to the pump regulator 14a to control the pump flow rate of the hydraulic pump 14, thereby controlling the pump absorption torque.

The model predictive control unit 34 is a functional element that performs control (model predictive control) based on optimal control theory in real time using a model that predicts the behavior of the hydraulic circuit including the hydraulic actuator and the hydraulic pump 14 . Model predictive control of a hydraulic circuit is control using a plant model of the hydraulic circuit. Also, the plant model of the hydraulic circuit is a model that allows the output of the hydraulic circuit to be derived from the input to the hydraulic circuit. In this embodiment, the model predictive control unit 34 calculates the displacement within a finite time from the pump discharge pressure Pd, the extension speed v of the hydraulic cylinder, the displacement volume Vd of the hydraulic pump 14L, and the minute change ΔVt of the displacement volume command value. A prediction of the pump discharge pressure in the future can be derived. The telescopic speed v includes the telescopic speed of the boom cylinder 7, the telescopic speed of the arm cylinder 8, the telescopic speed of the bucket cylinder 9, and the like. Further, the minute change ΔVt is the difference between the previous displacement volume command value Vt and the displacement volume command value Vtp the time before last. Further, the displacement volume Vd reaches a value corresponding to the displacement volume command value with a slight delay after the horsepower control unit 33 outputs the control current corresponding to the displacement volume command value to the pump regulator 14a. Therefore, the displacement volume Vd is derived from the past transition of the displacement volume command value. For example, if the displacement volume command value has not changed over a predetermined period of time, the current displacement volume Vd is set to the same value as the displacement volume command value. Further, when the displacement volume command value has changed recently, the current displacement volume Vd is calculated using the displacement volume command value before the change, the displacement volume command value after the change, and the first-order lag time constant. .

The model state resetting unit 35 is a functional element that constants variable coefficients in the plant model used by the model predictive control unit 34 . In this embodiment, the model state resetting unit 35 sets the displacement command value Vt as a variable coefficient to a function f(Pd) of the pump discharge pressure Pd, and sets the displacement command value Vt according to the pump discharge pressure Pd as an input. It outputs (=f(Pd)=2π×Tt/Pd). More specifically, the model state resetting unit 35 refers to the horsepower control line in the horsepower control and derives the displacement command value Vt corresponding to the pump discharge pressure Pd. That is, the displacement command value Vt corresponding to the pump discharge pressure Pd is derived so that the target pump absorption torque Tt is obtained by dividing the product of the pump discharge pressure Pd and the displacement command value Vt by 2π.

Next, calculations performed by the model predictive control unit 34 and the model state resetting unit 35 will be described. The model predictive control unit 34 derives the predicted value of the pump discharge pressure using the following matrix equation representing the state of the hydraulic circuit. Note that _A and B are coefficient matrices representing the structural features of the hydraulic circuit, and v ₁ , v ₂ , . Further, the coefficient matrix A includes the above-mentioned function f as a component, and the multiplication of f and Pd in this matrix equation is to calculate the function f using the pump discharge pressure Pd as an argument. This means that the displacement command value Vt is calculated based on the discharge pressure Pd. Furthermore, the coefficient matrix A contains as components the ratio of the number of revolutions of the hydraulic pump 14 to the cylinder volume and the reciprocal of the time constant (1/T) representing the pump response delay, and the ratio in this matrix equation and the displacement volume Vd means to calculate the pressure change of the hydraulic oil, and multiplication of its reciprocal by the displacement volume Vd means to calculate the volume change of the hydraulic oil. Further, the coefficient matrix B is the ratio of the piston pressure receiving area RA to the cylinder volume RV in each of the n hydraulic cylinders (RA ₁ /RV ₁ , RA ₂ /RV ₂ , . . . RA _n /RV _n ). as a component, and the multiplication of the ratio RA _n /RV _n and the extension speed v _n in this matrix equation means that the model predictive control unit 34 calculates the hydraulic fluid pressure change due to the extension and retraction of the n-th hydraulic cylinder. do. Further, the coefficient matrix B contains the reciprocal of the time constant (1/T) representing the response delay of the pump as a component, and the multiplication of the reciprocal of the time constant and the minute change ΔVt in this matrix equation is performed by the model predictive control unit 34 with n means to calculate the volume change of the hydraulic oil due to expansion and contraction of the hydraulic cylinder.

より具体的には、モデル予測制御部３４は、ポンプ吐出圧検出部３２から受けた現時点におけるポンプ吐出圧Ｐｄをモデル状態再設定部３５に入力してモデル状態再設定部３５から押し退け容積指令値Ｖｔを受ける。また、モデル予測制御部３４は、現時点におけるポンプ吐出圧Ｐｄ、押し退け容積Ｖｄ、ｎ個の油圧シリンダのそれぞれの伸縮速度ｖ_１～ｖ_ｎ、及び押し退け容積指令値の微小変化ΔＶｔと上記行列方程式とに基づいてポンプ吐出圧Ｐｄ及び押し退け容積Ｖｄのそれぞれの微分値を導き出す。そして、現時点におけるポンプ吐出圧Ｐｄにその微分値を加算して１制御周期後のポンプ吐出圧の予測値Ｐｄ'を導き出す。押し退け容積Ｖｄについても同様にして１制御周期後の押し退け容積の予測値Ｖｄ'を導き出す。なお、現時点における押し退け容積Ｖｄは、過去の押し退け容積指令値Ｖｔから一次遅れを考慮して算出される値である。

More specifically, the model predictive control unit 34 inputs the current pump discharge pressure Pd received from the pump discharge pressure detection unit 32 to the model state resetting unit 35 and outputs the displacement command value from the model state resetting unit 35. receive Vt. In addition, the model predictive control unit 34 calculates the current pump discharge pressure Pd, the displacement volume Vd, the expansion and contraction speeds v ₁ to v _n of each of the n hydraulic cylinders, and the minute change ΔVt in the displacement command value and the above matrix equation. Differential values of the pump discharge pressure Pd and the displacement volume Vd are derived based on . Then, the differential value is added to the current pump discharge pressure Pd to derive a predicted value Pd' of the pump discharge pressure after one control cycle. For the displacement volume Vd, similarly, a predicted value Vd' of the displacement volume after one control cycle is derived. Note that the current displacement volume Vd is a value calculated from the past displacement volume command value Vt in consideration of the first-order lag.

After that, the model prediction control unit 34 inputs the prediction value Pd′ to the model state resetting unit 35 . The model state resetting unit 35 calculates the displacement command value Vt' using the predicted value Pd' and the horsepower control diagram. Then, the model predictive control unit 34 receives the displacement command value Vt′ calculated by the model state resetting unit 35 . Further, the predicted values Pd' and Vd', the expansion and contraction velocities v ₁ to v _n of the n hydraulic cylinders at the present time, the small changes ΔVt in the displacement command values, and the matrix equation, the predicted value Pd' and Vd' are derived. Then, the differential value is added to the predicted value Pd' to derive the predicted value Pd'' of the pump discharge pressure after two control cycles. For the predicted value Vd', similarly, the predicted value Vd'' after two control cycles is derived.

In this way, the model predictive control unit 34 calculates the pump discharge pressure and the displacement after n control cycles when the minute change ΔVt is continuously used (that is, when the displacement command value changes by ΔVt every control cycle). Derive respective predictions of volume.

Furthermore, the model predictive control unit 34 determines the pump discharge pressure and the displacement after n control cycles in the case of continuous use over n control cycles with respect to a plurality of small change values set based on the small change ΔVt. Each predicted value for volume is derived in the manner described above. Each of the plurality of minute change values is derived, for example, by adding a predetermined value to the minute change ΔVt or subtracting a predetermined value from the minute change ΔVt.

Then, the model predictive control unit 34 calculates the displacement volume command value derived by the model state resetting unit 35 based on the pump discharge pressure after n control cycles (for example, after 10 control cycles), and the displacement volume after n control cycles. A minute change ΔVtc that minimizes the difference between and is selected from a plurality of minute change values. Specifically, one of a plurality of minute change values including the minute change ΔVt is selected as the minute change Vtc to be adopted this time.

The model predictive control unit 34 then outputs the selected minute change ΔVtc to the horsepower control unit 33 . The horsepower control unit 33 controls the displacement volume of the hydraulic pump 14 by model predictive control using the minute change ΔVtc selected by the model predictive control unit 34 .

In this manner, the model predictive control unit 34 calculates the predicted values of the pump discharge pressure and the displacement volume after n control cycles when each of the values of the plurality of minute changes is used for each control cycle. Then, a minute change ΔVtc that minimizes the difference between the displacement volume command value derived by the model state resetting unit 35 based on the pump discharge pressure after n control cycles and the displacement volume after n control cycles is selected. The horsepower control unit 33 calculates a displacement command value Vtc to be adopted this time from the minute change ΔVtc selected by the model predictive control unit 34 for each control cycle, and outputs a control current corresponding to the displacement command value Vtc to the pump regulator. 14a. The horsepower control unit 33 calculates the displacement command value Vtc to be adopted this time, for example, by adding a minute change ΔVtc to the previous displacement command value Vt. The pump regulator 14a changes the tilting angle of the swash plate of the hydraulic pump 14 to an angle corresponding to the control current. As a result, the displacement volume Vd of the hydraulic pump 14 is increased or decreased to the displacement volume that provides the target pump absorption torque Tt.

Next, with reference to FIG. 6, temporal changes in pump discharge pressure, displacement volume, and pump absorption torque when the excavation load suddenly increases during execution of an excavation operation including an arm closing operation will be described. FIG. 6 is a graph showing temporal transitions of pump discharge pressure, displacement volume, and pump absorption torque. Also, the dashed line in FIG. 6 indicates the temporal transition when the model predictive control is not adopted, and the solid line indicates the temporal transition when the model predictive control is adopted. Further, the hydraulic pump 14 is controlled by horsepower control during execution of this excavation operation.

When the model predictive control is not adopted, the pump discharge pressure Pd and the pump absorption torque Tp begin to increase when the excavation load increases sharply at time t1. This is because the extension speed of the arm cylinder 8 slows down even though the flow rate of the hydraulic oil directed from the hydraulic pump 14 to the arm cylinder 8 is maintained for a while due to the delayed response of the hydraulic pump 14 .

Specifically, the pump absorption torque Tp begins to decrease at time t2 after increasing beyond the target pump absorption torque Tt, then decreases until it falls below the target pump absorption torque Tt, and then increases again at time t3. turn around. After that, the pump absorption torque Tp converges to the target pump absorption torque Tt while repeating increases and decreases across the target pump absorption torque Tt.

In this way, the controller 30 increases or decreases the displacement volume command value according to the fluctuation of the pump discharge pressure Pd in a state including the response delay, and adjusts the displacement volume Vd so that the pump absorption torque Tp becomes the target pump absorption torque Tt. Control by feedback control. Therefore, the pump absorption torque Tp temporarily exceeds the target pump absorption torque Tt, and in some cases, the pump absorption torque Tp may exceed the engine output torque. Conversely, when returning the pump absorption torque Tp that has exceeded the target pump absorption torque Tt to the level of the target pump absorption torque Tt, the pump absorption torque Tp is excessively reduced to a level significantly below the target pump absorption torque Tt. Sometimes.

On the other hand, when the model predictive control was adopted, unlike the case where the control system of FIG. 5 was not adopted, the pump absorption torque Tp did not greatly exceed the target pump absorption torque Tt and maintained the target pump absorption torque Tt remain unchanged. This is because the deviation of the pump absorption torque Tp is preventively suppressed by changing the displacement command value before the pump absorption torque Tp deviates from the target pump absorption torque Tt by the model predictive control.

Specifically, based on the pump discharge pressure Pd, the displacement volume Vd, the expansion and contraction speed of the arm cylinder 8, and the above matrix equation, the controller 30 determines the pump displacement after a predetermined time when various displacement volume command values are adopted, based on the current pump discharge pressure Pd, displacement volume Vd, and the matrix equation. Predict discharge pressure and displacement volume. The expansion/contraction speed of the arm cylinder 8 is the change in the length of the arm cylinder 8, which is a state quantity that reacts faster than the pump discharge pressure to fluctuations in the excavation load. Therefore, although the extension speed of the arm cylinder 8 has already decreased at this time, the pump discharge pressure of the hydraulic pump 14 has not yet increased.

Then, the controller 30 sets the displacement volume that minimizes the difference between the displacement command value derived by the model state resetting unit 35 based on the predicted value of the pump discharge pressure after a predetermined time and the predicted value of the displacement after a predetermined time. Derive the command value. Then, the displacement command value is used to control the pump flow rate of the hydraulic pump 14 so that the pump absorption torque Tp is maintained at the target pump absorption torque Tt.

In this manner, the controller 30 increases or decreases the displacement volume command value, and controls the displacement volume Vd by model predictive control so as to achieve the target pump absorption torque Tt. Therefore, even if the excavation load increases rapidly, the pump absorption torque Tp is suppressed or prevented from temporarily exceeding the target pump absorption torque Tt, and the pump absorption torque Tp does not exceed the engine output torque. can be suppressed or prevented.

Moreover, the control system of FIG. 5 can reduce the influence of delay in response of the hydraulic pump 14 to fluctuations in the excavation load, compared to feedback control of the pump flow rate based on the pump discharge pressure.

Although the embodiments of the present invention have been described in detail above, the present invention is not limited to specific embodiments, and various modifications and changes can be made within the scope of the gist of the invention described in the claims. is possible.

For example, in the above embodiment, the horsepower control unit 33 calculates the displacement command value Vtc to be adopted this time from the minute change ΔVtc selected by the model predictive control unit 34 in each control cycle. However, the model predictive control unit 34 may calculate the displacement command value Vtc from the minute change ΔVtc and then output the displacement command value Vtc to the horsepower control unit 33 .

Further, in the above embodiment, the model predictive control unit 34 was described as being integrated with the controller 30, but it may be separate from the controller 30. FIG.

REFERENCE SIGNS LIST 1 lower running body 2 turning mechanism 2L left running hydraulic motor 2R right running hydraulic motor 3 upper rotating body 4 boom 5 arm 6. Bucket 7 Boom cylinder 8 Arm cylinder 9 Bucket cylinder 11 Engine 14, 14L, 14R Hydraulic pump 14a, 14aL, 14aR Pump regulator 17 Control valve 21 Turning hydraulic motor 22 Hydraulic oil tank 30 Controller 31 Actuator state quantity acquisition unit 32 Pump discharge pressure detection unit 33 Horsepower control unit 34 Model prediction control section 35 Model state resetting section 40L, 40R Center bypass pipe 41L, 41R Negative control throttle 75 Engine speed setting dial 100 Drive system 171L, 171R , 172L, 172R, 173L, 173R, 174R, 175L, 175R... Control valves S1 to S4... Sensors

Claims (8)

a lower running body;
an upper revolving body rotatably mounted on the lower traveling body;
an attachment including a working body attached to the upper revolving body;
a hydraulic actuator that drives the attachment;
a state quantity acquisition unit that acquires a state quantity of at least one of the hydraulic actuator and the working body ;
a controller;
The control device is
a model state resetting unit that uses the state quantity to derive a variable coefficient in a model that predicts the behavior of the hydraulic actuator;
a model predictive control unit that calculates a predicted value after a predetermined control cycle in the model based on the derived variable coefficient, and calculates a command value based on the calculated predicted value;
Excavator. The model predictive control unit recalculates the predicted value until after a predetermined control cycle based on the variable coefficient re-derived by the model state resetting unit using the calculated predicted value.
Shovel according to claim 1 . The model predictive control unit calculates the predicted value after a predetermined control cycle when continuously used over the control cycle.
A shovel according to claim 1 or 2. The model state resetting unit constants the variable coefficients in the model used by the model predictive control unit.
A shovel according to any one of claims 1 to 3. The model predictive control unit calculates the predicted value when using a magnitude of change in the command value calculated for each predetermined control cycle,
A shovel according to any one of claims 1 to 4. The model predictive control unit derives the predicted value from a matrix equation including the velocities of a plurality of cylinders constituting the hydraulic actuator.
A shovel according to any one of claims 1 to 5. an attachment including a lower traveling body, an upper revolving body mounted on the lower traveling body, a working body attached to the upper revolving body, a hydraulic actuator for driving the attachment, and at least the hydraulic actuator and the working body A system for an excavator comprising a state quantity acquisition unit that acquires any state quantity,
having a controller,
The control device is
a model state resetting unit that uses the state quantity to derive a variable coefficient in a model that predicts the behavior of the hydraulic actuator;
a model predictive control unit that calculates a predicted value after a predetermined control cycle in the model based on the derived variable coefficient, and calculates a command value based on the calculated predicted value;
system for excavators. The model predictive control unit recalculates the predicted value until after a predetermined control cycle based on the variable coefficient re-derived by the model state resetting unit using the calculated predicted value.
A system for an excavator according to claim 7.

Images