JP2019143474A - Shovel - Google Patents

Abstract

To provide a shovel which has control of preventing a delay of reduction in a discharge amount and preventing exceeding of a pump absorption torque from an engine output torque in quick rise of a discharge pressure of a hydraulic pump while keeping higher responsibility.SOLUTION: A shovel includes an arm cylinder, a hydraulic pump that supplies hydraulic oil to the arm cylinder and a controller 30 that controls the hydraulic pump, in which the controller 30 predicts a state amount in the future after a predetermined time based on an expansion and contraction velocity v of the arm cylinder.SELECTED DRAWING: Figure 5

Description

  The present invention relates to an excavator having a hydraulic actuator.

  There is known a construction machine that prevents a rapid increase in fuel injection amount due to a rapid 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). In addition, when it is determined that the rate of increase of the discharge pressure of the hydraulic pump is equal to or higher than the predetermined speed, the predetermined torque is reduced and 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 delay in the decrease in the discharge amount when the discharge pressure of the hydraulic pump suddenly increases. As a result, wasteful fuel consumption can be suppressed and the operability of the hydraulic actuator and the like can be improved.

JP 2005-54903 A

  However, the construction machine described above is affected by the response delay time.

  In view of the above, it is desirable to provide a shovel that can be controlled with higher responsiveness.

  An excavator according to an embodiment of the present invention predicts a state quantity in the future after a predetermined time based on a state quantity of a hydraulic actuator, a variable displacement hydraulic pump that supplies hydraulic oil to the hydraulic actuator, and a state quantity of the hydraulic actuator. A control device.

  The above-described means can provide a shovel that can be controlled with higher responsiveness.

It is a figure which shows the structural example of the shovel carrying the drive system which concerns on the Example of this invention. It is the schematic which shows the structural example of the drive system which concerns on the Example of this invention. It is a negative control chart showing the relationship between the pump flow rate and the negative control pressure. It is a horsepower control diagram (PQ diagram) showing the relationship between pump flow rate and pump discharge pressure. It is a functional block diagram which shows the structural example of a control system. It is a figure which shows the time transition 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. The excavator in FIG. 1 mounts an upper swing body 3 on a crawler type lower traveling body 1 via a swing mechanism 2 so as to be rotatable around the X axis. Further, the upper swing body 3 is provided with a drilling attachment in the front center portion. The excavation attachment includes a boom 4, an arm 5, and a bucket 6 as working bodies, and includes a boom cylinder 7, an arm cylinder 8, and a bucket cylinder 9 as hydraulic actuators.

  Further, a boom stroke sensor 7 s for detecting the expansion / contraction state of the boom cylinder 7 is attached to the boom cylinder 7. Further, an arm stroke sensor 8 s for detecting the expansion / contraction state of the arm cylinder 8 is attached to the arm cylinder 8, and a bucket stroke sensor 9 s for detecting the expansion / contraction state of the bucket cylinder 9 is attached to 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 work body.

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

  The hydraulic pump 14 is driven by the engine 11. In the present embodiment, the hydraulic pump 14 is a variable displacement swash plate hydraulic pump that can vary the discharge amount per one rotation (push-out volume). The displacement volume is controlled by the pump regulator 14a. Specifically, the hydraulic pump 14 includes a first hydraulic pump 14L whose displacement is controlled by a pump regulator 14aL and a second hydraulic pump 14R whose displacement is controlled by a pump regulator 14aR. In this embodiment, the rotation shaft of the hydraulic pump 14 is connected to the rotation shaft of the engine 11 and rotates at the same rotation speed as the rotation speed of the engine 11. Note that the rotation shaft of the hydraulic pump 14 is connected to the flywheel, and suppresses fluctuations in the rotation speed when the engine output torque fluctuates.

  The engine 11 is a shovel drive source. In this embodiment, the engine 11 is a diesel engine including a turbocharger as a supercharger and a fuel injection device, and is mounted on the upper swing body 3. The engine 11 may include a supercharger as a supercharger.

  The control valve 17 is a hydraulic control mechanism that supplies hydraulic oil discharged from 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, 175R. The hydraulic actuator includes 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 oil to the hydraulic oil tank 22 through the center bypass conduit 40L that communicates the control valves 171L, 172L, 173L, and 175L. Similarly, the hydraulic pump 14R circulates the hydraulic oil to the hydraulic oil tank 22 through the center bypass conduit 40R that communicates 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-side traveling hydraulic motor 2L and the hydraulic pump 14L.

  The control valve 171R is a spool valve as a traveling straight valve. The control valve 171R switches the flow of hydraulic fluid so that hydraulic fluid is supplied from the hydraulic pump 14L to the left traveling hydraulic motor 2L and the right traveling hydraulic motor 2R, respectively, in order to improve the straight traveling performance of the lower traveling body 2. Specifically, when the left traveling hydraulic motor 2L and the right traveling hydraulic motor 2R and any of the other hydraulic actuators are operated simultaneously, the hydraulic pump 14L includes the left traveling hydraulic motor 2L and the right traveling hydraulic pressure. Hydraulic fluid is supplied to both motors 2R. In other cases, the hydraulic pump 14L supplies hydraulic oil to the left traveling hydraulic motor 2L, and the hydraulic pump 14R supplies hydraulic oil 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 fluid 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 oil between the right-side 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 fluid between the boom cylinder 7 and the hydraulic pumps 14L and 14R, respectively. The control valve 173R operates when a boom operation lever as an operation device is operated, and the control valve 173L operates when the boom operation lever is operated in the boom raising direction with a predetermined lever operation amount or more.

  The control valve 174 </ b> R is a spool valve that controls the flow rate and flow direction of hydraulic oil between the hydraulic pump 14 </ b> R 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 operation lever as an operation device is operated, and the control valve 175R operates when the arm operation lever is operated at a predetermined lever operation amount or more.

  The center bypass pipelines 40L and 40R are respectively provided with negative control throttles 40L and 40R between the control valves 175L and 175R located on the most downstream side and the hydraulic oil tank 22. Hereinafter, the negative control is abbreviated as “negative control”. The negative control throttles 41L and 41R generate negative control pressure upstream of the negative control throttles 41L and 41R by restricting the flow of hydraulic oil discharged from the hydraulic pumps 14L and 14R.

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

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

  In this embodiment, the controller 30 determines the operation contents (for example, presence / absence of lever operation, lever operation direction, lever operation amount, etc.) of various operation devices 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 a pilot pressure generated when various operation devices such as an arm operation lever and a boom operation 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 for detecting the inclination of various operation levers.

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

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

  The pressure sensors S3 and S4 detect the discharge pressures 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 so that the operator can switch the engine speed in a plurality of stages, for example.

  Then, the controller 30 causes the CPU to execute programs corresponding to the various functional elements according to the operation contents of the various operation devices and the operating states of the various hydraulic actuators.

  Next, a process in which the controller 30 controls the discharge amount of the hydraulic pump 14 according to the negative control pressure will be described with reference to FIG. FIG. 3 is a negative 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. The vertical axis indicates the pump flow rate and the horizontal axis indicates the relationship. Distribute negative control pressure.

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

  Specifically, the hydraulic oil discharged from the hydraulic pump 14L passes through the center bypass conduit 40L, reaches the negative control throttle 41L, and generates a negative control pressure upstream of the negative control throttle 41L.

  For example, when the control valve 175L moves to operate the arm cylinder 8, the hydraulic oil 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 decreases or disappears, and the negative control pressure generated upstream of the negative control throttle 41L decreases.

  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. The pump regulator 14aL increases the displacement volume by increasing the swash plate tilt angle of the hydraulic pump 14L according to the increase in the control current from the controller 30. As a result, sufficient hydraulic oil is supplied to the arm cylinder 8, and the arm cylinder 8 is appropriately driven.

  Thereafter, when the control valve 175L is returned to the neutral position in order 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. The pump regulator 14aL reduces the displacement volume by reducing the tilt angle of the swash plate of the hydraulic pump 14L according to the reduction of the control current from the controller 30. 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 conduit 40L is suppressed.

  A pump control line represented by a solid line shows a tendency that the pump flow rate increases as the negative control pressure decreases. Hereinafter, the control of the pump flow rate based on the negative control pressure as described above is referred to as “negative control”. With the negative control, the drive system 100 can suppress wasteful energy consumption in a standby state where the hydraulic actuator is not operated. This is because the pumping loss generated by the hydraulic oil discharged from the hydraulic pump 14 can be suppressed. Further, when operating the hydraulic actuator, the drive system 100 can supply necessary and sufficient hydraulic fluid from the hydraulic pump 14 to the hydraulic actuator.

  Moreover, the drive system 100 performs horsepower control in parallel with the negative control. In the horsepower control, the pump flow rate is reduced 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, the absorption horsepower (pump absorption torque) of the hydraulic pump represented by the product of the pump discharge pressure and the pump flow rate does not exceed the engine output horsepower (engine output torque).

  FIG. 4 is a horsepower control diagram (PQ diagram) showing the relationship between the pump flow rate and the pump discharge pressure, where the vertical axis represents the pump flow rate and the horizontal axis represents the pump discharge pressure. The horsepower control line shows a tendency that the pump flow rate increases as the pump discharge pressure decreases. The horsepower control line is determined according to the target pump absorption torque Tt, and shifts to the upper right in the figure 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 larger than the target pump absorption torque Tt2 corresponding to the horsepower control line represented by the broken line. The target pump absorption torque Tt is a value set in advance as the allowable maximum value of the pump absorption torque that can be used by the hydraulic pump 14. 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 with 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 that is the detection value of the pressure sensor S3. Then, the controller 30 outputs a control current corresponding to the target displacement volume to the pump regulator 14aL. The pump regulator 14aL increases / decreases the swash plate tilt angle in accordance with the control current to set the displacement volume to the target displacement volume. By such feedback control of the pump absorption torque, the controller 30 can operate the hydraulic pump 14L with the target pump absorption torque Tt even if the pump discharge pressure varies due to the variation of the load related to 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 the detection of the change in the pump discharge pressure to the actual change in the pump flow rate.

  Therefore, the controller 30 employs model predictive control in order to eliminate this response delay time. In this embodiment, the controller 30 derives a command value for the hydraulic pump 14 by predicting 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. The state quantity of the hydraulic actuator is a state quantity that reacts faster than the pump discharge pressure with respect to the load fluctuation related to the hydraulic actuator. Etc. The state quantity of the hydraulic pump 14 includes 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 a pump discharge pressure and the like. The command value for the hydraulic pump 14 includes a displacement volume 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. FIG. 5 is a functional block diagram illustrating a configuration example of the control system.

  Specifically, the control system of FIG. 5 includes a controller 30, an actuator state quantity acquisition unit 31, and a pump discharge pressure detection unit 32. The controller 30 includes a horsepower control unit 33, a model prediction control unit 34, and a model state resetting unit 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 speed v of the hydraulic cylinder based on the detected displacement, and outputs the value to the model prediction control unit 34 as an electric expansion / contraction speed signal. In the present embodiment, the displacement sensor includes a boom stroke sensor 7s, an arm stroke sensor 8s, and a bucket stroke sensor 9s. The actuator state quantity acquisition unit 31 may be an angle sensor that detects the rotation angle of the work body. Similarly in this case, 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 electric 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 value as an electric pump discharge pressure signal to the model prediction control unit 34.

  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 controls the pump absorption torque by outputting a control current corresponding to the displacement volume command value to the pump regulator 14a to control the pump flow rate of the hydraulic pump 14.

  The model predictive control unit 34 is a functional element that performs control (model predictive control) based on the 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. The model predictive control of the hydraulic circuit is control using a plant model of the hydraulic circuit. The plant model of the hydraulic circuit is a model that can derive the output of the hydraulic circuit from the input to the hydraulic circuit. In the present embodiment, the model prediction control unit 34 determines that the pump discharge pressure Pd, the hydraulic cylinder extension speed v, the displacement volume Vd of the hydraulic pump 14L, and the minute change ΔVt of the displacement volume command value within a finite time. The predicted value of pump discharge pressure in the future can be derived. The extension / contraction speed v includes the extension speed of the boom cylinder 7, the extension speed of the arm cylinder 8, the extension / contraction speed of the bucket cylinder 9, and the like. The minute change ΔVt is a difference between the previous displacement volume command value Vt and the previous displacement volume command value Vtp. The displacement volume Vd reaches a value corresponding to the displacement volume command value with a slight delay from the time when the horsepower control unit 33 outputs a 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, the current displacement volume Vd is set to the same value as the displacement volume command value. When the displacement volume command value has changed most 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 time constant of the primary delay. .

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

Next, calculation executed by the model prediction control unit 34 and the model state resetting unit 35 will be described. The model prediction control unit 34 derives a predicted value of the pump discharge pressure using the following matrix equation representing the state of the hydraulic circuit. Incidentally, A, B is a coefficient matrix representing the structural characteristics of the hydraulic _{circuit, v 1, v 2, ···} , v n are each the extension rate of the n hydraulic cylinder. The coefficient matrix A includes the above-described 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, that is, the model state resetting unit 35 performs pumping. It means that the displacement volume command value Vt is calculated based on the discharge pressure Pd. Further, the coefficient matrix A includes the ratio of the rotational speed of the hydraulic pump 14 to the cylinder volume and the inverse of the time constant (1 / T) representing the pump response delay as components, and the ratio and displacement volume Vd in this matrix equation. Multiplication means that the change in pressure of the hydraulic oil is calculated, and multiplication of the reciprocal thereof and the displacement volume Vd means that the volume change of the hydraulic oil is calculated. Further, the coefficient matrix B, the ratio of the pressure receiving area RA and the cylinder volume RV of the piston in each of the n of the hydraulic cylinder _{(RA 1 / RV 1, RA} 2 / RV 2, ···, RA n / RV n) Is multiplied by the ratio RA _n / RV _n and the extension speed vn in this matrix equation means that the model prediction control unit 34 calculates the pressure change of the hydraulic oil due to the expansion and contraction of the nth hydraulic cylinder. . The coefficient matrix B includes as a component the reciprocal (1 / T) of the time constant representing the response delay of the pump, and the model prediction control unit 34 performs n multiplications of the reciprocal in this matrix equation and the minute change ΔVt. This means that the volume change of hydraulic oil due to the expansion and contraction of the hydraulic cylinder is calculated.

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 prediction 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 pushes it away from the model state resetting unit 35. Receive Vt. The model predictive control unit 34 also includes the pump discharge pressure Pd, the displacement volume Vd, the expansion / contraction speeds v _{1 to} v _{n of the n} hydraulic cylinders, the minute change ΔVt of the displacement volume command value, the matrix equation, Based on the above, the differential values of the pump discharge pressure Pd and the displacement volume Vd are derived. Then, the differential value is added to the current pump discharge pressure Pd to derive the predicted value Pd ′ of the pump discharge pressure after one control cycle. Similarly, for the displacement volume Vd, a predicted value Vd ′ of the displacement volume after one control cycle is derived. Note that the displacement volume Vd at the present time is a value calculated in consideration of the primary delay from the past displacement volume command value Vt.

Thereafter, the model prediction control unit 34 inputs the predicted value Pd ′ to the model state resetting unit 35. The model state resetting unit 35 calculates the displacement volume command value Vt ′ using the predicted value Pd ′ and the horsepower control diagram. The model prediction control unit 34 receives the displacement volume command value Vt ′ calculated by the model state resetting unit 35. Further, the predicted value Pd 'and Vd', each of the telescopic velocity v ₁ to v _n and displacement minute change in volume command value ΔVt of n of the hydraulic cylinder at the present time, the predicted value on the basis of the above matrix equation Pd ' And the differential values of 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. Similarly, the predicted value Vd ′ after two control cycles is derived for the predicted value Vd ′.

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

  Further, the model predictive control unit 34 uses a plurality of minute change values set with the minute change ΔVt as a reference, and the pump discharge pressure and displacement after the n control period when continuously used over the n control period. Each predicted value of 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 the predetermined value from the minute change ΔVt.

  In addition, the model prediction control unit 34 determines 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. The minute change ΔVtc that minimizes the difference between the two 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.

  Then, the model prediction control unit 34 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 prediction control using the minute change ΔVtc selected by the model prediction control unit 34.

  As described above, the model prediction control unit 34 calculates the predicted value of the pump discharge pressure and the displacement volume after the n control periods when each of a plurality of minute change values is used for each control period. 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 the n control period and the displacement volume after the n control period is selected. The horsepower control unit 33 calculates a displacement volume command value Vtc to be adopted this time from the minute change ΔVtc selected by the model prediction control unit 34 for each control cycle, and supplies a control current corresponding to the displacement volume command value Vtc to the pump regulator. 14a is output. The horsepower controller 33 calculates a displacement volume command value Vtc to be adopted this time, for example, by adding a minute change ΔVtc to the previous displacement volume command value Vt. The pump regulator 14a changes the swash plate tilt angle 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, time transitions of the pump discharge pressure, the displacement volume, and the pump absorption torque when the excavation load rapidly increases during execution of the excavation operation including the arm closing operation will be described with reference to FIG. FIG. 6 is a diagram showing temporal transitions of the pump discharge pressure, displacement, and pump absorption torque. In addition, a broken line in FIG. 6 indicates a temporal transition when the model predictive control is not employed, and a solid line indicates a temporal transition when the model predictive control is employed. During the excavation operation, the hydraulic pump 14 is controlled by horsepower control.

  When model predictive control is not employed, the pump discharge pressure Pd and the pump absorption torque Tp start to increase when the excavation load increases rapidly at time t1. This is because, due to the response delay of the hydraulic pump 14, the flow rate of the hydraulic oil from the hydraulic pump 14 toward the arm cylinder 8 is maintained for a while, but the extension speed of the arm cylinder 8 slows down.

  Specifically, the pump absorption torque Tp increases after exceeding the target pump absorption torque Tt, then starts decreasing at time t2, and then decreases until it falls below the target pump absorption torque Tt, and then increases again at time t3. Turn. Thereafter, the pump absorption torque Tp converges to the target pump absorption torque Tt while repeatedly increasing and decreasing across the target pump absorption torque Tt.

  As described above, 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 sets 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. In contrast, when the pump absorption torque Tp that has exceeded the target pump absorption torque Tt is returned to the target pump absorption torque Tt level, the pump absorption torque Tp is excessively reduced to a level that is significantly lower than the target pump absorption torque Tt. There is a case.

  On the other hand, when the model predictive control is adopted, the pump absorption torque Tp is maintained at the target pump absorption torque Tt without significantly exceeding the target pump absorption torque Tt, unlike the case where the control system of FIG. 5 is not adopted. It remains unchanged. This is because the displacement control command value is changed before the pump absorption torque Tp deviates from the target pump absorption torque Tt by the model predictive control, and the deviation of the pump absorption torque Tp is proactively suppressed.

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

  The controller 30 then pushes the displacement volume that minimizes the difference between the displacement volume 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 displacement value after the predetermined time. The command value is derived. Then, the pump flow rate of the hydraulic pump 14 is controlled using the displacement displacement command value so that the pump absorption torque Tp is maintained at the target pump absorption torque Tt.

  In this way, the controller 30 increases or decreases the displacement volume command value, and controls the displacement volume Vd by model predictive control so that the target pump absorption torque Tt is obtained. Therefore, even when the excavation load increases rapidly, the pump absorption torque Tp is temporarily suppressed or prevented from exceeding the target pump absorption torque Tt, and the pump absorption torque Tp exceeds the engine output torque. Can be suppressed or prevented.

  In addition, the control system of FIG. 5 can reduce the influence of the response delay of the hydraulic pump 14 on the excavation load fluctuation, compared to the 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 the specific embodiments, and various modifications and changes can be made within the scope of the gist of the present invention described in the claims. Is possible.

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

  In the above-described embodiment, the model prediction control unit 34 has been described as being integrated with the controller 30, but may be separate from the controller 30.

  DESCRIPTION OF SYMBOLS 1 ... Lower traveling body 2 ... Turning mechanism 2L ... Left side traveling hydraulic motor 2R ... Right traveling hydraulic motor 3 ... Upper turning body 4 ... Boom 5 ... Arm 6. .. Bucket 7 ... Boom cylinder 8 ... Arm cylinder 9 ... Bucket cylinder 11 ... Engine 14, 14L, 14R ... Hydraulic pumps 14a, 14aL, 14aR ... Pump regulator 17 ... Control valve 21 ... Hydraulic 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 predictive control unit 35 ... Model state resetting unit 40L, 40R ... Center bypass pipeline 41L, 41R ... Negative control 75 ... engine rotational speed setting dial 100 ... driving system 171L, 171R, 172L, 172R, 173L, 173R, 174R, 175L, 175R ··· control valve S1 to S4 ... sensor

Claims (4)

A hydraulic actuator;
A variable displacement hydraulic pump for supplying hydraulic oil to the hydraulic actuator;
A control device that predicts a state quantity in the future after a predetermined time based on a state quantity of the hydraulic actuator,
Excavator. The state quantity of the hydraulic actuator is an expansion / contraction speed of a hydraulic cylinder as the hydraulic actuator, or a rotation angle of a work element that the hydraulic cylinder rotates.
The excavator according to claim 1.   The shovel according to claim 1, wherein the control device obtains a predicted value after an n control period after the predetermined time and controls the shovel.   The excavator according to claim 3, wherein the control device calculates the predicted value using a plant model of a hydraulic circuit of the excavator.

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