JP2014055635A - Method for controlling shovel - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a method for controlling a shovel configured to appropriately adjust a discharge flow rate of a hydraulic pump immediately after starting an operation of a hydraulic actuator.SOLUTION: A method for controlling a shovel according to an embodiment of the invention performs an initial motion control immediately after starting an operation of a hydraulic actuator. The initial motion control takes priority over a negative control and a horse power control. The initial motion control is carried out on the basis of a discharge flow rate command of a hydraulic pump in the horsepower control. The discharge flow rate command of the hydraulic pump is determined on the basis of a discharge pressure of the hydraulic pump and a control current that determines a horsepower curve.

Description

  The present invention relates to a method for controlling an excavator.

  Conventionally, a construction machine pump control circuit is known that reduces the discharge flow rate of a hydraulic pump as the pressure of the hydraulic pump increases so as not to exceed the maximum horsepower of the engine (see, for example, Patent Document 1).

  For this purpose, the pump control circuit includes first and second regulators that control the discharge flow rate of the hydraulic pump, and an electromagnetic proportional control valve that can adjust the pressure at each control port of the first and second regulators. With.

JP 2010-48359 A

  However, since adjustment by the electromagnetic proportional control valve is slow in response, the pump control circuit temporarily increases the discharge flow rate when the hydraulic actuator is first moved. For this reason, an overload torque is generated with respect to the engine, and the operation is deteriorated.

  In view of the above, an object of the present invention is to provide a shovel control method for appropriately adjusting the discharge flow rate of a hydraulic pump immediately after the start of operation of a hydraulic actuator.

  In order to achieve the above-described object, a method for controlling an excavator according to an embodiment of the present invention is characterized in that initial-time control is performed immediately after the operation of a hydraulic actuator is started.

  By the above-described means, the present invention can provide a shovel control method for appropriately adjusting the discharge flow rate of the hydraulic pump immediately after the start of the operation of the hydraulic actuator.

1 is a side view of a hydraulic excavator according to an embodiment of the present invention. It is a block diagram which shows the structure of the drive system of the hydraulic shovel of FIG. It is a figure which shows the principal part structural example of the hydraulic circuit mounted in the hydraulic shovel of FIG. It is a PQ diagram showing the relationship between the discharge pressure P of the main pump and the first discharge flow rate command Q. It is a flowchart which shows the flow of the control processing at the time of initial movement. It is a figure which shows the time transition of various physical quantities immediately after starting operation of a hydraulic actuator. It is a figure which shows another principal part structural example of the hydraulic circuit mounted in the hydraulic shovel of FIG. It is a block diagram which shows the structural example of the drive system of the hybrid type shovel which concerns on the Example of this invention. It is a block diagram which shows the structure of the electrical storage system of the hybrid type shovel which concerns on the Example of this invention. It is a figure which shows the principal part structural example of the hydraulic circuit mounted in the hybrid type shovel which concerns on the Example of this invention.

  FIG. 1 is a side view showing a hydraulic excavator according to an embodiment of the present invention.

  An upper swing body 3 is mounted on the lower traveling body 1 of the hydraulic excavator via a swing mechanism 2. A boom 4 is attached to the upper swing body 3. An arm 5 is attached to the tip of the boom 4, and a bucket 6 is attached to the tip of the arm 5. The boom 4, the arm 5, and the bucket 6 constitute an attachment, and are hydraulically driven by a boom cylinder 7, an arm cylinder 8, and a bucket cylinder 9, which are hydraulic cylinders. The upper swing body 3 is provided with a cabin 10 and is mounted with a power source such as an engine.

  FIG. 2 is a block diagram showing the configuration of the drive system of the hydraulic excavator shown in FIG. In FIG. 2, the mechanical power system is indicated by a double line, the high-pressure hydraulic line is indicated by a thick solid line, the pilot line is indicated by a broken line, and the electric drive / control system is indicated by a thin solid line.

  A main pump 14 and a pilot pump 15 as hydraulic pumps are connected to an output shaft of the engine 11 as a mechanical drive unit. A control valve 17 is connected to the main pump 14 via a high pressure hydraulic line 16. An operation device 26 is connected to the pilot pump 15 via a pilot line 25.

  The control valve 17 is a device that controls a hydraulic system in the hydraulic excavator. Hydraulic actuators such as the hydraulic motors 1A (for right) and 1B (for left), the boom cylinder 7, the arm cylinder 8, the bucket cylinder 9, and the swing hydraulic motor 21B for the lower traveling body 1 are connected to the control valve 17 via a high-pressure hydraulic line. It is connected to the.

  The operating device 26 includes a lever 26A, a lever 26B, and a pedal 26C. The lever 26A, the lever 26B, and the pedal 26C are connected to the control valve 17 and the pressure sensor 29 via hydraulic lines 27 and 28, respectively. The pressure sensor 29 is connected to a controller 30 that performs drive control of the electric system.

  The controller 30 is a controller as a main control unit that performs drive control of the hydraulic excavator. The controller 30 includes a CPU (Central Processing Unit) and an arithmetic processing unit including an internal memory, and is realized by the CPU executing a drive control program stored in the internal memory.

  Next, a method for controlling the discharge flow rate of the main pump 14 mounted on the hydraulic excavator shown in FIG. 1 will be described with reference to FIG. FIG. 3 shows a configuration example of a main part of a hydraulic circuit mounted on the hydraulic excavator shown in FIG. In the following, when referring to each of the pair of left and right components, “L” is added to the end of the reference symbol of the left component, and “R” is added to the end of the reference symbol of the right component. Attached. Further, when referring to both of the pair of left and right components, “L” and “R” at the end are omitted.

  3 mainly includes a main pump 14 (main pump 14L, main pump 14R), a horsepower proportional valve 31, a flow proportional valve 32 (flow proportional valve 32L, a flow proportional valve 32R), a regulator 50 (regulator 50L, Regulator 50R). In the following, a method for controlling the discharge flow rate of the main pump 14L will be mainly described, but the same description applies to the main pump 14R.

  The main pump 14L is connected to the tank via the center bypass pipe C1L, the control valve 17, and the negative control throttle 33L. Hereinafter, the negative control is referred to as “negative control”.

  The center bypass conduit C1L is connected to the discharge pressure sensor 34L upstream of the control valve 17. The discharge pressure sensor 34L detects the discharge pressure of the main pump 14L and outputs the detection result to the controller 30.

  The center bypass conduit C1L is connected to the negative control conduit C2L upstream of the negative control restrictor 33L. The negative control conduit C2L is connected to the regulator 50L via the shuttle valve 35L. Further, the negative control line C2L is connected to the negative control pressure sensor 36L between the shuttle valve 35L and the regulator 50L. The negative control pressure sensor 36L detects the negative control pressure acting on the regulator 50L and outputs the detection result to the controller 30.

  The horsepower proportional valve 31 is a valve for generating a control pressure used for the horsepower control of the main pump 14. The discharge pressure of the pilot pump 15 is set as the primary pressure, and the secondary pressure is adjusted according to the control current from the controller 30. A control pressure that is a side pressure is generated.

  The “horsepower control” is control for adjusting the input horsepower of the main pump 14 so that the input horsepower of the main pump 14 does not exceed the output horsepower of the engine 11.

  The flow rate proportional valve 32 is a valve that generates a control pressure used for initial control of the main pump 14 </ b> L. The discharge pressure of the pilot pump 15 is set as the primary pressure, and the secondary pressure is controlled according to the control current from the controller 30. The control pressure which is the pressure of is generated.

  “Initial control” is control for adjusting the discharge flow rate of the main pump 14 immediately after the start of operation of the hydraulic actuator (at the time of initial operation). In this embodiment, the initial operation control is executed using a hydraulic circuit portion for negative control.

  “Negative control” is control for increasing the discharge flow rate of the main pump 14 when the hydraulic actuator is operated and decreasing the discharge flow rate of the main pump 14 when the hydraulic actuator is not operated.

  In this embodiment, the initial operation control is executed by replacing the negative control pressure in the negative control with the control pressure generated by the flow rate proportional valve 32. More specifically, the initial operation control temporarily disables the negative control immediately after the start of the operation of the hydraulic actuator, and prevents the discharge flow rate of the main pump 14L from increasing due to the negative control.

  The regulator 50L is a device that controls the discharge flow rate of the main pump 14L, and mainly includes a horsepower control mechanism 51L, a negative control mechanism 52L, a spool valve mechanism 53L, and a swash plate tilting mechanism 54L.

  The horsepower control mechanism 51L is a mechanism that determines the discharge flow rate of the main pump 14L so that the input horsepower of the main pump 14 does not exceed the output horsepower of the engine 11. In the present embodiment, the horsepower control mechanism 51L mainly includes a horsepower control piston 510L, a first pressure chamber 511L, a second pressure chamber 512L, and a third pressure chamber 513L.

  The first pressure chamber 511L is a pressure chamber that receives the pressure on the secondary side of the horsepower proportional valve 31 via the conduit C3L. Further, the second pressure chamber 512L is a pressure chamber that receives the discharge pressure of the main pump 14R via the pipe line C4L. The third pressure chamber 513L is a pressure chamber that receives the discharge pressure of the main pump 14L via the conduit C5L.

  The horsepower control piston 510L moves left and right in the drawing according to the total thrust determined based on the pressures in the first pressure chamber 511L, the second pressure chamber 512L, and the third pressure chamber 513L. The 53L spool valve 530L is slid left and right. Specifically, the horsepower control piston 510L has a tendency to move to the right, that is, to decrease the discharge flow rate of the main pump 14L as the sum of the thrusts increases.

  The negative control mechanism 52L is a mechanism that determines the discharge flow rate of the main pump 14L according to the negative control pressure. In this embodiment, the negative control mechanism 52L mainly includes a negative control piston 520L and a fourth pressure chamber 521L.

  The fourth pressure chamber 521L is a pressure chamber that receives the larger of the negative control pressure upstream of the negative control throttle 33L and the pressure on the secondary side of the flow rate proportional valve 32L as the effective negative control pressure.

  The negative control piston 520L moves left and right in the drawing according to the thrust determined based on the pressure in the fourth pressure chamber 521L, and slides the spool valve 530L of the spool valve mechanism 53L to the left and right. Specifically, the negative control piston 520L has a tendency to move to the right as the thrust increases, that is, to decrease the discharge flow rate of the main pump 14L.

  The spool valve mechanism 53L is a mechanism that drives the swash plate tilting mechanism 54L. In the present embodiment, the spool valve mechanism 53L mainly includes a 3-port 3-position spool valve 530L.

  The spool valve 530L slides in accordance with the movement of the piston with the larger amount of movement in the right direction in the figure among the horsepower control piston 510L of the horsepower control mechanism 51L and the negative control piston 520L of the negative control mechanism 52L. Move.

  The spool valve 530L has a first position 531L, a second position 532L, and a third position 533L. The first position 531L is a spool position that is effective when the spool valve 530L is at a predetermined neutral position. The second position 532L is a spool position that is effective when the spool valve 530L moves in the right direction in the figure with respect to the neutral position. The third position 533L is a spool position that is effective when the spool valve 530L moves in the left direction in the drawing with respect to the neutral position.

  The swash plate tilt mechanism 54L is a mechanism that controls the swash plate tilt angle of the main pump 14L. In the present embodiment, the swash plate tilting mechanism 54L mainly includes a servo piston 540L, a fifth pressure chamber 541L, and a sixth pressure chamber 542L.

  The servo piston 540L is based on a rightward thrust expressed by the product of the pressure of the fifth pressure chamber 541L and the pressure receiving area A1, and a leftward thrust expressed by the product of the pressure of the sixth pressure chamber 542L and the pressure receiving area A2. Slide left and right. Specifically, the servo piston 540L slides rightward when the rightward thrust is greater than the leftward thrust, and changes the swash plate tilt angle so that the discharge flow rate of the main pump 14L decreases. On the contrary, the servo piston 540L slides leftward when the leftward thrust is larger than the rightward thrust, and changes the swash plate tilt angle so that the discharge flow rate of the main pump 14L increases. The pressure in the fifth pressure chamber 541L is adjusted between the discharge pressure of the main pump 14L and the tank pressure by the spool valve mechanism 53L. The pressure in the sixth pressure chamber 542L is equal to the discharge pressure of the main pump 14L.

  When the spool valve 530L enables the first position 531L, the spool valve mechanism 53L causes the flow of hydraulic oil between the pipe line C5L and the fifth pressure chamber 541L, and between the tank and the fifth pressure chamber 541L. Shut off the flow of hydraulic oil in between. That is, the pressure in the fifth pressure chamber 541L is prevented from changing.

  Further, when the spool valve 530L enables the second position 532L, the spool valve mechanism 53L communicates between the pipe line C5L and the fifth pressure chamber 541L, and between the tank and the fifth pressure chamber 541L. Shut off the flow of hydraulic oil in between. That is, the pressure in the fifth pressure chamber 541L is increased in the range from the tank pressure to the discharge pressure of the main pump 14L.

  Further, when the spool valve 530L enables the third position 533L, the spool valve mechanism 53L communicates between the tank and the fifth pressure chamber 541L, and connects the pipe C5L and the fifth pressure chamber 541L. Shut off the flow of hydraulic oil in between. That is, the pressure in the fifth pressure chamber 541L is reduced in the range from the tank pressure to the discharge pressure of the main pump 14L.

  With the above configuration, the regulator 50L allows the discharge pressure of the main pump 14L, the discharge pressure of the main pump 14R, and the secondary side of the horsepower proportional valve 31 when the horsepower control piston 510L slides the spool valve 530L. The larger the total thrust based on each of the pressures, the smaller the discharge flow rate of the main pump 14L. That is, when the horsepower control piston 510L is on the right side of the negative control piston 520L, the controller 30 increases or decreases the pressure on the secondary side of the horsepower proportional valve 31 to thereby discharge the main pump 14L. Can be reduced or increased.

  When the negative control piston 520L slides the spool valve 530L, the regulator 50L has an effective negative control pressure (the original negative control pressure upstream of the negative control throttle 33L and the pressure on the secondary side of the flow proportional valve 32L). As the thrust based on the larger one of the two is larger, the discharge flow rate of the main pump 14L is decreased. That is, when the negative control piston 520L is on the right side of the horsepower control piston 510L, the controller 30 increases or decreases the pressure on the secondary side of the flow proportional valve 32 to thereby discharge the main pump 14L. Can be reduced or increased.

  In the above-described configuration, the horsepower control mechanism 51L has a slower response than the negative control mechanism 52L. The main reason is that the pressure range acting on the horsepower control piston 510L is larger than the pressure range acting on the negative control piston 520L.

  Therefore, immediately after the operation of the hydraulic actuator is started, the moving speed in the left direction of the negative control piston 520L accompanying the decrease of the effective negative control pressure is the horsepower control piston 510L accompanying the increase of the discharge pressure of the main pump 14L. Exceeds moving speed in the right direction. As a result, since the spool valve 530L of the spool valve mechanism 53L is in contact with the negative control piston 520L, it slides to the left together with the negative control piston 520L that moves to the left at a relatively high speed. Thereafter, the spool valve 530L contacts the horsepower control piston 510L that moves to the right at a relatively low speed, moves away from the negative control piston 520L that moves to the left at a relatively high speed, and moves to the right together with the horsepower control piston 510L. To slide. The above-described series of movements means that the discharge flow rate of the main pump 12L increases suddenly although temporarily.

  In contrast, the horsepower proportional valve 31 increases the pressure on the secondary side in order to minimize the amount of movement of the spool valve 530L in the left direction at the time of initial movement, that is, the increase amount of the discharge flow rate of the main pump 14L as much as possible. Increasing the pressure on the secondary side means creating an increase in input horsepower in the main pump 14 in a pseudo manner.

  Here, the effect when the secondary pressure of the horsepower proportional valve 31 is increased will be described with reference to FIG. FIG. 4 is a PQ diagram showing the relationship between the discharge pressure P of the main pump 14L and the first discharge flow rate command Q. FIG. 4 shows the horsepower curve CV1 after increasing the pressure on the secondary side of the horsepower proportional valve 31 with a solid line, and the horsepower curve CV2 before increasing the pressure on the secondary side of the horsepower proportional valve 31 with a broken line. Represented by

  The “first discharge flow rate command” is a discharge flow rate of the main pump 14L to be realized by the horsepower control mechanism 51L. In the present embodiment, the value of the first discharge flow rate command corresponds to the position of the horsepower control piston 510L of the horsepower control mechanism 51L. Specifically, the value of the first discharge flow rate command decreases as the horsepower control piston 510L in FIG. 3 moves to the right.

  The discharge flow rate of the main pump 14L is adjusted to the flow rate indicated by the first discharge flow rate command when the horsepower control piston 510L slides the spool valve 530L. On the other hand, the discharge flow rate of the main pump 14L is not adjusted to the flow rate indicated by the first discharge flow rate command when the negative control piston 520L slides the spool valve 530L. This is because the flow rate is adjusted to the flow rate indicated by the second discharge flow rate command.

  The “second discharge flow rate command” is a discharge flow rate of the main pump 14L to be realized by the negative control mechanism 52L. In the present embodiment, the value of the second discharge flow rate command corresponds to the position of the negative control piston 520L of the negative control mechanism 52L. Specifically, the value of the second discharge flow rate command decreases as the negative control piston 520L in FIG. 3 moves to the right.

  As shown in FIG. 4, when the secondary pressure is increased by decreasing the control current for the horsepower proportional valve 31, the region represented by the horsepower curve is reduced. This means that the input horsepower of the main pump 14L is limited to be smaller. This is because an increase in the pressure on the secondary side means to artificially create an increase in input horsepower in the main pump 14. As a result, if the discharge pressure P of the main pump 14L is constant, the discharge flow rate command Q of the main pump 14L decreases. The horsepower proportional valve 31 may be configured such that the pressure on the secondary side increases when the control current increases.

  As described above, the hydraulic circuit of FIG. 3 limits the input horsepower by the horsepower proportional valve 31 in order to prevent a sudden increase in the discharge flow rate at the initial movement due to the difference in response speed between the horsepower control mechanism 51L and the negative control mechanism 52L. Temporarily strengthen. As a result, a sudden increase in the discharge flow rate can be minimized.

  In addition, the hydraulic circuit of FIG. 3 uses the flow rate proportional valve 32L to more reliably prevent a sudden increase in the discharge flow rate at the initial operation.

  Here, a process for controlling the discharge flow rate when the flow proportional valve 32L is initially moved (hereinafter referred to as “initial-time control process”) will be described with reference to FIG. FIG. 5 is a flowchart showing the flow of the initial motion control process, and the controller 30 executes the initial motion control process immediately after the operation of the hydraulic actuator is started. In the present embodiment, the flow proportional valve 32L already supplies a predetermined control pressure higher than the negative control pressure when no hydraulic actuator is operated to the fourth pressure chamber 521L of the negative control mechanism 52L. ing. However, the flow rate proportional valve 32L may start supplying a predetermined control pressure to the fourth pressure chamber 521L immediately after the operation of the hydraulic actuator is started.

  First, the controller 30 determines whether or not the first discharge flow rate command Q is equal to or less than a predetermined value (step S1). That is, the controller 30 determines whether or not the horsepower control piston 510L of FIG. 3 is positioned on the right side of the predetermined position. In the present embodiment, the controller 30 calculates the first discharge flow rate command Q based on the output of the discharge pressure sensors 34L and 34R and the value of the control current for the horsepower proportional valve 31. Note that the controller 30 may determine the first discharge flow rate command Q based on a reference map indicating the relationship between the discharge pressure and control current and the first discharge flow rate command Q. Further, the controller 30 may determine the first discharge flow rate command Q based on the output of a sensor (not shown) that detects the position of the horsepower control piston 510L.

  When it is determined that the first discharge flow rate command Q is greater than the predetermined value (NO in step S1), the controller 30 continues the determination in step S1 until the first discharge flow rate command Q becomes equal to or less than the predetermined value.

  When it is determined that the first discharge flow rate command Q is equal to or less than the predetermined value (YES in step S1), the controller 30 decreases the control current for the flow rate proportional valve 32L (step S2). This is because the control pressure acting on the fourth pressure chamber 521L of the negative control mechanism 52L is reduced. That is, the negative control piston 520L in FIG. 3 is moved in the left direction. The degree of decrease in the control current may be set in advance, or may be dynamically determined based on the operation amount of the hydraulic actuator.

  By this initial control process, the hydraulic circuit of FIG. 3 moves the negative control piston 520L in contact with the spool valve 530L until the horsepower control piston 510L moves rightward and reaches a predetermined position. It stays at a position on the right side of the piston 510L. Thereafter, when the horsepower control piston 510L reaches a predetermined position, the hydraulic circuit of FIG. 3 gradually moves the negative control piston 520L and the spool valve 530L to the left. The spool valve 530L contacts the horsepower control piston 510L when the negative control piston 520L moves to the left of the horsepower control piston 510L, and thereafter moves together with the horsepower control piston 510L.

  In this manner, the hydraulic circuit of FIG. 3 can more reliably prevent a sudden increase in the discharge flow rate at the time of initial operation using the flow rate proportional valve 32L.

  Next, the effect of the initial operation control process will be described with reference to FIG. FIG. 6 is a diagram showing temporal transitions of various physical quantities immediately after the operation of the hydraulic actuator is started. FIG. 6 (A) shows the transition of the discharge flow rate command, and FIG. 6 (B) shows the discharge flow rate. 6C shows the transition of the effective negative control pressure, and FIG. 6D shows the transition of the control current for the flow proportional valve 32L.

  In FIG. 6A, the transition of the first discharge flow rate command is indicated by a one-dot chain line, and the transition of the second discharge flow rate command when the initial operation control processing is not executed is indicated by a broken line, and the initial operation control processing is executed. The transition of the second discharge flow rate command is shown by a solid line.

  FIG. 6B shows the change in the discharge flow rate when the initial operation control process is not executed with a broken line, and the change in the discharge flow rate when the initial operation control process is executed with a solid line.

  FIG. 6C shows the transition of the effective negative control pressure when the initial control process is not executed with a broken line, and the transition of the effective negative control pressure when the initial control process is executed with a solid line.

  When the initial-time control process is not executed, when the operation of the hydraulic actuator is started at time t1, the first discharge flow rate command decreases relatively slowly as shown by the one-dot chain line in FIG. This represents that the horsepower control piston 510L of FIG. 3 moves to the right relatively slowly. On the other hand, the second discharge flow rate command increases relatively abruptly as shown by the broken line in FIG. This represents that the negative control piston 520L in FIG. 3 moves to the left relatively rapidly.

  The discharge flow rate follows the smaller one of the first discharge flow rate command and the second discharge flow rate command with a slight delay. This represents that the spool valve 530L in FIG. 3 is in contact with the more right side of the horsepower control piston 510L and the negative control piston 520L and slides according to the movement of the piston on the right side. .

  Therefore, as shown by the broken line in FIG. 6B, the discharge flow rate increases rapidly after the time t1, and then gradually decreases following the decrease in the first discharge flow rate command. It gradually increases following the increase in the discharge flow rate command.

  As shown by the broken line in FIG. 6C, the effective negative control pressure decreases rapidly to the minimum level by the center bypass pipe C1L being blocked by the control valve 17, and then remains at the minimum level.

  As described above, when the initial-time control process is not executed, the discharge flow rate of the main pump 14L temporarily increases immediately after the operation of the hydraulic actuator is started, as indicated by hatched hatching in FIG. 6B. . Although this rapid increase is suppressed because the input horsepower is temporarily restricted by the horsepower proportional valve 31, it causes overload torque and deterioration of operability.

  On the other hand, when the initial-time control process is executed, even when the operation of the hydraulic actuator is started at time t1, as shown in FIG. Is maintained at a predetermined value. This means that even if the original negative control pressure decreases as shown by the broken line in FIG. 6C, the effective negative control pressure changes at the same level as shown by the solid line in FIG. 6C. That is, it means that the initial control is executed in preference to the negative control.

  As a result, as shown by the solid line in FIG. 6A, the second discharge flow rate command changes at the same level without increasing. This represents that the position of the negative control piston 520L in FIG. 3 is maintained as it is.

  Therefore, as shown by the solid line in FIG. 6 (B), the discharge flow rate does not increase temporarily and stays at the same level. This means that it is possible to further suppress the rapid increase in the discharge flow rate that is suppressed by temporarily limiting the input horsepower by the horsepower control. That is, it means that the initial operation control is executed in preference to the horsepower control.

  Thereafter, when the first discharge flow rate command becomes less than the predetermined value at time t2, as shown in FIG. 6D, the control current for the flow proportional valve 32L is decreased at a predetermined rate until time t4.

  As a result, as shown by the solid line in FIG. 6C, the effective negative control pressure decreases at a predetermined rate according to the decrease in the control current, and the second discharge flow rate command is shown by the solid line in FIG. Thus, it increases at a predetermined rate according to the decrease in the effective negative control pressure. Further, as shown by the solid line in FIG. 6B, the discharge flow rate increases at a predetermined rate in accordance with the increase in the second discharge flow rate command. Further, as shown in FIG. 6A, after the second discharge flow rate command exceeds the first discharge flow rate command at time t3, the discharge flow rate is controlled to follow the first discharge flow rate command. This indicates that the position of the horsepower control piston 510L is on the right side of the position of the negative control piston 520L, and the spool valve 530L moves together with the horsepower control piston 510L.

  Thus, the initial-time control process according to the embodiment of the present invention prevents a temporary sudden increase in the discharge flow rate of the main pump 14L immediately after the operation of the hydraulic actuator is started, and smoothes the discharge flow rate of the main pump 14L. Can be launched. As a result, it is possible to prevent deterioration in operability such as a decrease in engine speed, vibration of the vehicle body due to a shock caused by the decrease, generation of overload torque, and uncomfortable feeling related to acceleration of the attachment.

  Note that the hydraulic circuit according to the embodiment of the present invention may execute the initial operation control by the flow rate proportional valve 32 after omitting the temporary restriction of the input horsepower by the horsepower proportional valve 31 at the initial operation.

  Next, another configuration example of the hydraulic circuit mounted on the hydraulic excavator according to the embodiment of the present invention will be described with reference to FIG. FIG. 7 is a diagram illustrating a configuration example of the regulator 50AL in the regulator 50A in the hydraulic circuit mounted on the hydraulic excavator in FIG. Hereinafter, the regulator 50AL will be described, but the same description applies to a regulator 50AR (not shown).

  The regulator 50AL is different from the regulator 50L of FIG. 3 in that the negative control mechanism 52L and the initial operation control mechanism 55L are separated, but is common in other points. Therefore, the difference will be described in detail while omitting the description of the common points.

  In the present embodiment, the fourth pressure chamber 521L of the negative control mechanism 52L is configured as a pressure chamber that receives only the negative control pressure upstream of the negative control throttle 33L. Then, the negative control piston 520L moves left and right in the drawing according to the thrust determined based on the pressure in the fourth pressure chamber 521L, and slides the spool valve 530L of the spool valve mechanism 53L to the left and right. Specifically, the negative control piston 520L has a tendency to move to the right as the thrust increases, that is, to decrease the discharge flow rate of the main pump 14L.

  The initial action control mechanism 55L is a mechanism that determines the discharge flow rate of the main pump 14L according to the pressure on the secondary side of the flow rate proportional valve 32L. In the present embodiment, the initial motion next control mechanism 55L mainly includes an initial motion control piston 550L and a seventh pressure chamber 551L.

  The seventh pressure chamber 551L is configured as a pressure chamber that receives only the pressure on the secondary side of the flow proportional valve 32L. Then, the initial-movement control piston 550L moves left and right in the drawing according to the thrust determined based on the pressure in the seventh pressure chamber 551L, and slides the spool valve 530L of the spool valve mechanism 53L to the left and right. Specifically, the initial-movement control piston 550L has a tendency to move to the right as the thrust increases, that is, to decrease the discharge flow rate of the main pump 14L.

  The spool valve 530L slides in accordance with the movement of the piston having the largest amount of movement in the right direction in the drawing among the horsepower control piston 510L, the negative control piston 520L, and the initial motion control piston 550L.

  With the above configuration, the hydraulic circuit including the regulator 50AL can achieve the same effect as the hydraulic circuit including the regulator 50L.

  Next, a hybrid excavator according to an embodiment of the present invention will be described with reference to FIG. FIG. 8 is a block diagram illustrating a configuration example of a drive system of a hybrid excavator. In FIG. 8, the mechanical power system is indicated by a double line, the high-pressure hydraulic line is indicated by a thick solid line, the pilot line is indicated by a broken line, and the electric drive / control system is indicated by a thin solid line. 8 includes a motor generator 12, a transmission 13, an inverter 18, and a power storage system 120. Instead of the swing hydraulic motor 21B, the inverter 20, the swing electric motor 21, the resolver 22, and the mechanical brake are provided. 2 and a drive system shown in FIG. 2 in that a load drive system including a turning speed reducer 24 is provided. However, the other points are common to the drive system of FIG. Therefore, the differences will be described in detail while omitting the description of the common points.

  In FIG. 8, an engine 11 as a mechanical drive unit and a motor generator 12 as an assist drive unit that also generates electric power are connected to two input shafts of a transmission 13, respectively. A main pump 14 and a pilot pump 15 as hydraulic pumps are connected to the output shaft of the transmission 13.

  The motor generator 12 is connected via an inverter 18 to a power storage system (power storage device) 120 including a capacitor as a power storage device.

  The power storage system 120 is disposed between the inverter 18 and the inverter 20. Thereby, when at least one of the motor generator 12 and the turning electric motor 21 is performing a power running operation, the power storage system 120 supplies electric power necessary for the power running operation, and at least one of them is performing a power generation operation. In this case, the power storage system 120 stores the electric power generated by the power generation operation as electric energy.

  FIG. 9 is a block diagram illustrating a configuration example of the power storage system 120. The storage system 120 includes a capacitor 19 as a storage battery, a step-up / down converter 100, and a DC bus 110. The DC bus 110 serving as the second battery controls the transfer of electric power among the capacitor 19 serving as the first battery, the motor generator 12, and the turning motor 21. The capacitor 19 is provided with a capacitor voltage detector 112 for detecting a capacitor voltage value and a capacitor current detector 113 for detecting a capacitor current value. The capacitor voltage value and the capacitor current value detected by the capacitor voltage detection unit 112 and the capacitor current detection unit 113 are supplied to the controller 30. In the above description, the capacitor 19 is shown as an example of the capacitor. However, instead of the capacitor 19, a rechargeable secondary battery such as a lithium ion battery, a lithium ion capacitor, or another form of power source capable of transmitting and receiving power. May be used as a capacitor.

  The step-up / step-down converter 100 performs control to switch between the step-up operation and the step-down operation so that the DC bus voltage value falls within a certain range according to the operation state of the motor generator 12 and the turning electric motor 21. The DC bus 110 is disposed between the inverters 18 and 20 and the step-up / down converter 100, and transfers power between the capacitor 19, the motor generator 12, and the turning electric motor 21.

  Returning to FIG. 8, the inverter 20 is provided between the turning electric motor 21 and the power storage system 120, and performs operation control on the turning electric motor 21 based on a command from the controller 30. Thereby, the inverter 20 supplies necessary electric power from the power storage system 120 to the turning electric motor 21 when the electric turning motor 21 is performing a power running operation. Further, when the turning electric motor 21 is in a power generation operation, the electric power generated by the turning electric motor 21 is stored in the capacitor 19 of the power storage system 120.

  The turning electric motor 21 may be an electric motor capable of both power running operation and power generation operation, and is provided for driving the turning mechanism 2 of the upper turning body 3. In the power running operation, the rotational driving force of the turning electric motor 21 is amplified by the speed reducer 24, and the upper turning body 3 is subjected to acceleration / deceleration control to perform rotational movement. In addition, during the power generation operation, the inertial rotation of the upper-part turning body 3 is transmitted to the turning electric motor 21 with the number of rotations increased by the speed reducer 24, and regenerative power can be generated. Here, the turning electric motor 21 is an electric motor that is AC-driven by the inverter 20 using a PWM (Pulse Width Modulation) control signal. The turning electric motor 21 can be constituted by, for example, a magnet-embedded IPM motor. Thereby, since a larger induced electromotive force can be generated, the electric power generated by the turning electric motor 21 at the time of regeneration can be increased.

  The charging / discharging control of the capacitor 19 of the power storage system 120 is performed in the charging state of the capacitor 19, the operating state of the motor generator 12 (power running operation or power generating operation), and the operating state of the turning motor 21 (power running operation or regenerative operation). Based on the controller 30.

  The resolver 22 is a sensor that detects the rotation position and rotation angle of the rotation shaft 21 </ b> A of the turning electric motor 21. Specifically, the resolver 22 detects the difference between the rotation position of the rotation shaft 21A before the rotation of the turning electric motor 21 and the rotation position after the left rotation or the right rotation, thereby rotating the rotation angle of the rotation shaft 21A. And detecting the direction of rotation. By detecting the rotation angle and rotation direction of the rotating shaft 21A of the turning electric motor 21, the rotation angle and rotation direction of the turning mechanism 2 are derived.

  The mechanical brake 23 is a braking device that generates a mechanical braking force, and mechanically stops the rotating shaft 21 </ b> A of the turning electric motor 21. This mechanical brake 23 is switched between braking and release by an electromagnetic switch. This switching is performed by the controller 30.

  The turning transmission 24 is a transmission that decelerates the rotation of the rotating shaft 21 </ b> A of the turning electric motor 21 and mechanically transmits it to the turning mechanism 2. Thereby, during the power running operation, the rotational force of the turning electric motor 21 can be increased, and a larger rotational force can be transmitted to the upper turning body 3. On the contrary, in the regenerative operation, the rotation generated in the upper swing body 3 can be accelerated and mechanically transmitted to the swing electric motor 21.

  The turning mechanism 2 can turn in a state where the mechanical brake 23 of the turning electric motor 21 is released, whereby the upper turning body 3 is turned leftward or rightward.

  The controller 30 performs operation control of the motor generator 12 (switching between electric assist operation or power generation operation) and also performs charge / discharge control of the capacitor 19 by drivingly controlling the step-up / step-down converter 100 as the step-up / step-down control unit. The controller 30 is configured to control the step-up / down converter 100 based on the charge state of the capacitor 19, the operation state of the motor generator 12 (electric assist operation or power generation operation), and the operation state of the turning motor 21 (power running operation or regenerative operation). Switching control between the step-up operation and the step-down operation is performed, and thereby charge / discharge control of the capacitor 19 is performed. The controller 30 also controls the amount (charging current or charging power) charged in the capacitor.

  The switching control between the step-up / step-down operation of the step-up / step-down converter 100 is performed by controlling the DC bus voltage value detected by the DC bus voltage detection unit 111, the capacitor voltage value detected by the capacitor voltage detection unit 112, and the capacitor current detection unit 113. Is performed based on the capacitor current value detected by.

  The electric power generated by the motor generator 12, which is an assist motor, is supplied to the DC bus 110 of the power storage system 120 via the inverter 18 and supplied to the capacitor 19 via the step-up / down converter 100. The regenerative power generated by the regenerative operation of the turning electric motor 21 is supplied to the DC bus 110 of the power storage system 120 via the inverter 20 and supplied to the capacitor 19 via the step-up / down converter 100.

  The present invention can also be applied to a hybrid excavator having the above-described configuration.

  Here, a control method of the discharge flow rate of the main pump 14 mounted on the hybrid excavator according to the embodiment of the present invention will be described with reference to FIG. FIG. 10 shows a configuration example of a main part of a hydraulic circuit mounted on the hybrid excavator according to the embodiment of the present invention.

  The hydraulic circuit of FIG. 10 is different from the hydraulic circuit of FIG. 3 in that the flow proportional valve 32R is omitted and the secondary side of the flow proportional valve 32L is connected to both the shuttle valves 35L and 35R. Common.

  This configuration is based on the fact that in the hydraulic excavator, hydraulic oil discharged from the main pump 14L is supplied to the swing hydraulic motor 21B, whereas in the hybrid excavator, the swing mechanism is independent of the hydraulic circuit.

  Specifically, in the hydraulic circuit of FIG. 3, the main pump 14L is configured so that the flow proportional valve 32L can execute the swing flow rate cut control for restricting the flow rate of the hydraulic oil flowing from the main pump 14L to the swing hydraulic motor 21B. The related flow proportional valve 32L and the flow proportional valve 32R related to the main pump 14R are made independent. In contrast, the hydraulic circuit of FIG. 10 does not require the turning flow rate cut control, and it is not necessary to make the flow rate proportional valve 32L and the flow rate proportional valve 32R independent. Therefore, the function of the flow rate proportional valve 32R is changed to the flow rate proportional valve 32L. The flow rate proportional valve 32R is omitted as a whole.

  With the above configuration, the hydraulic circuit in FIG. 10 can achieve the same effect as the effect by the hydraulic circuit in FIG. 3. In the hydraulic circuit of FIG. 10, the flow rate proportional valve 32L may be omitted instead of the flow rate proportional valve 32R. Further, the flow rate proportional valve 32L and the flow rate proportional valve 32R may be made independent in the hydraulic circuit of FIG. Further, the functions of the flow proportional valve 32L and the flow proportional valve 32R may be integrated in the hydraulic circuit of FIG. 3, and one of them may be omitted. In this case, the turning flow rate cut control may be omitted or may be realized using another mechanism.

  Although the preferred embodiments of the present invention have been described in detail above, the present invention is not limited to the above-described embodiments, and various modifications and substitutions can be made to the above-described embodiments without departing from the scope of the present invention. Can be added.

  DESCRIPTION OF SYMBOLS 1 ... Lower traveling body 1A, 1B ... Traveling hydraulic motor 2 ... Turning mechanism 3 ... Upper turning body 4 ... Boom 5 ... Arm 6 ... Bucket 7 ... Boom Cylinder 8 ... Arm cylinder 9 ... Bucket cylinder 10 ... Cabin 11 ... Engine 12 ... Motor generator 13 ... Transmission 14 ... Main pump 15 ... Pilot pump 16 ..High pressure hydraulic line 17 ... Control valve 18 ... Inverter 19 ... Capacitor 20 ... Inverter 21 ... Rotary motor 21A ... Rotating shaft 21B ... Swivel hydraulic motor 22 ... Resolver 23 ... Mechanical brake 24 ... Swivel transmission 25 ... Pilot line 26 ... Operating devices 26A, 26B ... Rebar -26C ... Pedal 27, 28 ... Hydraulic line 29 ... Pressure sensor 30 ... Controller 31 ... Horsepower proportional valve 32 ... Flow proportional valve 33 ... Negative control throttle 34 ... Discharge Pressure sensor 35 ... Shuttle valve 36 ... Negative control pressure sensor 50, 50A ... Regulator 51 ... Horsepower control mechanism 52 ... Negative control control mechanism 53 ... Spool valve mechanism 54 ... Swash plate tilt Rotation mechanism 55 ... Initial motion control mechanism 100 ... Buck-boost converter 110 ... DC bus 111 ... DC bus voltage detector 112 ... Capacitor voltage detector 113 ... Capacitor current detector 120 ..Power storage system

Claims (6)

Perform initial motion control immediately after the start of hydraulic actuator operation,
Excavator control method. The initial control is prioritized over negative control and horsepower control.
The shovel control method according to claim 1. The initial operation control is performed based on a discharge flow rate command of a hydraulic pump in horsepower control.
The shovel control method according to claim 1 or 2. The discharge flow command of the hydraulic pump is determined based on the discharge pressure of the hydraulic pump and a control current that defines a horsepower curve.
The shovel control method according to claim 3. The initial action control is executed using a negative control piston.
The shovel control method according to any one of claims 1 to 4. The initial movement control is executed using a piston different from the negative control piston and the horsepower control piston.
The shovel control method according to any one of claims 1 to 4.

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