Work machine

Abstract

In the present invention, a control device obtains an estimated value of the load torque of a hydraulic pump on the basis of a target value for the discharge amount of the hydraulic pump and a measured value of the discharge pressure. The control device controls a prime mover on the basis of the estimated value of the load torque.

Description

Working machinery

[0001] This disclosure relates to a work machine. This application claims priority to Japanese Patent Application No. 2024-168904, filed in Japan on September 27, 2024, the contents of which are incorporated herein by reference.

[0002] Work machines that are driven by hydraulic pressure using hydraulic fluid are known (see, for example, Patent Document 1).

[0003] International Publication No. 2006 / 123704

[0004] In the supply of hydraulic fluid to a work machine, when the load torque increases relative to the output torque of the prime mover, the rotational speed of the prime mover decreases. In particular, since there is a response delay in the control of the hydraulic pump, when the hydraulic pump is controlled by feedback control based on the difference between the target rotational speed and the actual rotational speed, even if the actual rotational speed decreases and a gap is created with respect to the target rotational speed, the actual rotational speed does not quickly converge to the target rotational speed. In other words, once the actual rotational speed decreases, the effect of the response delay causes the actual rotational speed to remain below the target rotational speed for a non-negligible period of time. The purpose of this disclosure is to provide a work machine that can suppress the decrease in prime mover rotational speed due to the response delay.

[0005] According to a first aspect of the present invention, the work machine comprises a prime mover, a variable-capacity hydraulic pump driven by the power of the prime mover and discharging hydraulic fluid, an actuator driven by the hydraulic fluid discharged by the hydraulic pump, and a control device, wherein the control device determines an estimated value of the load torque of the hydraulic pump based on a target value of the discharge amount and a measured value of the discharge pressure of the hydraulic pump, and controls the prime mover based on the estimated value of the load torque.

[0006] According to a second aspect of the present invention, the work machine comprises a prime mover, a mechanism for changing the output characteristics of the prime mover, a variable-capacity hydraulic pump driven by the power of the prime mover and for discharging hydraulic fluid, an actuator driven by the hydraulic fluid discharged by the hydraulic pump, and a control device, wherein the control device determines a target value for the load torque of the hydraulic pump, limits the target value for the load torque to less than or equal to the upper limit torque that the prime mover can output according to the state of the mechanism, and controls the hydraulic pump based on the limited target value for the load torque.

[0007] According to a third aspect of the present invention, the work machine comprises a prime mover and a control device, the control device determines the acceleration torque required for accelerating and decelerating the prime mover based on the target angular acceleration of the prime mover, determines a target value for the load torque of the hydraulic pump, limits the target value for the load torque to less than or equal to a torque threshold obtained by subtracting the acceleration torque from the maximum torque of the prime mover, and controls the hydraulic pump based on the limited target value for the load torque.

[0008] According to a fourth aspect of the present invention, the work machine comprises a variable-capacity hydraulic pump for discharging hydraulic fluid and a control device, wherein the control device identifies the response characteristics of the load torque of the hydraulic pump based on the operating state of the hydraulic pump and controls the hydraulic pump based on an upper limit torque obtained by subtracting a margin corresponding to the response characteristics of the load torque from the maximum torque of the hydraulic pump.

[0009] According to the above embodiment, the working machine can suppress the decrease in the rotational speed of the prime mover due to the response delay.

[0010] This is a perspective view showing the configuration of a work machine according to the first embodiment. This is a schematic block diagram showing the configuration of a drive system according to the first embodiment. This is a schematic block diagram showing the configuration of a control device according to the first embodiment. This is a flowchart showing a method for predicting a target flow rate after a certain period of time according to the first embodiment. This is a flowchart showing a method for determining a target flow rate of a hydraulic pump according to the first embodiment. This is a flowchart showing a method for determining whether two main oil passages merge or separate according to the first embodiment. This is a block diagram showing engine control by the engine control unit according to the first embodiment. This is a diagram showing engine control when feedforward control based on future load torque is not performed. This is a diagram showing engine control when feedforward control based on future load torque is performed. This is a flowchart showing a method for determining the upper limit capacity of a hydraulic pump according to the first embodiment. This is a diagram showing the relationship between an isotorque curve function and a tangent function according to the first embodiment. This is a block diagram showing the control of a hydraulic pump by the pump control unit according to the first embodiment. This is a schematic block diagram showing the configuration of a drive system according to the second embodiment. This is a flowchart showing a method for determining the upper limit capacity of a hydraulic pump according to the second embodiment. This is a block diagram showing the configuration of a drive system for a work machine according to the third embodiment. This is a flowchart showing merge / separation control according to the third embodiment. This is a schematic block diagram showing the configuration of a computer according to at least one embodiment.

[0011] <First Embodiment> <Configuration of the Working Machine> The embodiment will be described in detail below with reference to the drawings. Figure 1 is a perspective view showing the configuration of the working machine 1 according to the first embodiment. The working machine 1 according to the first embodiment is a hydraulic excavator.

[0012] The work machine 1 comprises a traveling body 110, a rotating body 120, a work implement 130, a driver's cab 140, and a machine room 150. The work machine 1, which is a hydraulic excavator, excavates and levels soil and other materials at work sites. The traveling body 110 and the rotating body 120 constitute the vehicle body.

[0013] The vehicle body 110 supports the work machine 1 so that it can move. The vehicle body 110 has a pair of left and right tracks 111. The work machine 1 moves forward, turns, or moves backward by the rotation of the pair of tracks 111. Hereinafter, the drive motor 112 for driving the right track 111 will be called the first drive motor 112R, and the drive motor 112 for driving the left track 111 will be called the second drive motor 112L.

[0014] The slewing body 120 is supported by the traveling body 110 so that it can rotate. The slewing body 120 rotates relative to the traveling body 110 by a slewing motor 127. The slewing body 120 supports the work equipment 130, the operator's cab 140, and the machine room 150.

[0015] The work machine 130 is movably supported on the body (slewing body 120) of the work machine 1. The work machine 130 comprises a boom 131, an arm 132, and a bucket 133 which is a work tool. The base end of the boom 131 is rotatably attached to the slewing body 120. The base end of the arm 132 is rotatably attached to the tip of the boom 131. The bucket 133 is rotatably attached to the tip of the arm 132.

[0016] The work machine 130 is driven by a plurality of actuators. The plurality of actuators include, for example, a boom cylinder 131C, an arm cylinder 132C, and a bucket cylinder 133C.

[0017] The boom cylinder 131C is a hydraulic cylinder for driving the boom 131. The base end of the boom cylinder 131C is attached to the slewing body 120. The tip end of the boom cylinder 131C is attached to the boom 131. The arm cylinder 132C is a hydraulic cylinder for driving the arm 132. The base end of the arm cylinder 132C is attached to the boom 131. The tip end of the arm cylinder 132C is attached to the arm 132. The bucket cylinder 133C is a hydraulic cylinder for driving the bucket 133. The base end of the bucket cylinder 133C is attached to the arm 132. The tip end of the bucket cylinder 133C is attached to the bucket 133.

[0018] The operator's cab 140 is where the operator of the work machine 1 sits and operates and controls it. The operator's cab 140 is located, for example, on the left side of the front end of the slewing body 120. The operator's cab 140 of the work machine 1 is equipped with a driver's seat 141 for the operator to sit in, an operating device 142 for operating the work machine 1, and a control device 145 for controlling the work machine 1. The operating device 142 is operated by the operator to operate the work machine 1. The operating device 142 outputs an operation signal in response to the operator's operation.

[0019] The control device 145 controls the work machine 1. The control device 145 receives operation signals from the operating device 142. Based on the operation signals from the operating device 142, the control device 145 controls the movement of the traveling body 110, the driving of the work machine 130, and the rotational movement of the rotating body 120.

[0020] The machine room 150 houses the drive system 20, which will be described later. The machine room 150 is located, for example, behind the driver's cab 140. The machine room 150 forms a space for housing the drive system 20.

[0021] 《Configuration of the drive system 20》 Figure 2 is a schematic block diagram showing the configuration of the drive system 20 according to the first embodiment. The drive system 20 includes an engine 121, a front hydraulic pump 122A, a rear hydraulic pump 122B, a front main oil passage 123A, a rear main oil passage 123B, a main connecting oil passage 124, a front control valve 126A, a rear control valve 126B, a swing motor 127, a boom cylinder 131C, an arm cylinder 132C, a bucket cylinder 133C, a first travel motor 112R, and a second travel motor 112L. Hereinafter, the front hydraulic pump 122A and the rear hydraulic pump 122B will be collectively referred to as the hydraulic pump 122. Also, hereafter, the front main oil passage 123A and the rear main oil passage 123B will be collectively referred to as the main oil passage 123. Also, hereafter, the front control valve 126A and the rear control valve 126B will be collectively referred to as the control valve 126.

[0022] The engine 121 is a prime mover that drives the hydraulic pump 122. In the first embodiment, the engine 121 is a diesel engine. Examples of engines 121 in other embodiments may include a gasoline engine or an electric motor. The front hydraulic pump 122A and the rear hydraulic pump 122B are variable displacement pumps driven by the engine 121. The front hydraulic pump 122A supplies hydraulic fluid to the front main oil passage 123A. The rear hydraulic pump 122B supplies hydraulic fluid to the rear main oil passage 123B. The front hydraulic pump 122A is provided with a front swash plate angle sensor 1221A that measures the inclination angle of the swash plate of the front hydraulic pump 122A. The rear hydraulic pump 122B is provided with a rear swash plate angle sensor 1221B that measures the inclination angle of the swash plate of the rear hydraulic pump 122B. Since the capacity of the hydraulic pump 122 is determined by the inclination angle of the swash plate, the control device 145 can determine the discharge amount of the hydraulic pump 122 based on the measurement value of the swash plate angle sensor 1221. Note that the discharge amount of the hydraulic pump 122 also changes depending on the rotational speed in addition to the capacity, so if the work machine 1 according to other embodiments controls the discharge amount of the hydraulic pump 122 by rotational speed, the work machine 1 may be equipped with a sensor that measures the rotational speed of the hydraulic pump 122 instead of the swash plate angle sensor 1221.

[0023] The main oil passage 123 is equipped with a pump discharge pressure sensor 1231 and a bleed valve 1232. The pump discharge pressure sensor 1231 measures the pressure of the hydraulic fluid discharged by the hydraulic pump 122. The bleed valve 1232 discharges the hydraulic fluid supplied when the actuator is not being driven. Even when the actuator is not being driven, the hydraulic pump 122 discharges the minimum flow rate of hydraulic fluid. Hereinafter, the pump discharge pressure sensor 1231 provided in the front main oil passage 123A will also be called the front pump discharge pressure sensor 1231A, and the pump discharge pressure sensor 1231 provided in the rear main oil passage 123B will also be called the rear pump discharge pressure sensor 1231B. Furthermore, the bleed valve 1232 provided in the front main oil passage 123A will also be called the front bleed valve 1232A, and the bleed valve 1232 provided in the rear main oil passage 123B will also be called the rear bleed valve 1232B.

[0024] The main connecting oil passage 124 connects the front main oil passage 123A and the rear main oil passage 123B. The main connecting oil passage 124 is equipped with a main confluence / separation valve 1241. The main confluence / separation valve 1241 is a shut-off valve that controls the opening and closing of the main connecting oil passage 124. This allows the main confluence / separation valve 1241 to switch between confluence or separation of the hydraulic fluid flowing through the front main oil passage 123A and the hydraulic fluid flowing through the rear main oil passage 123B.

[0025] The front control valve 126A is connected to the front main oil passage 123A. The front control valve 126A distributes the hydraulic fluid supplied from the front main oil passage 123A to the swing motor 127, the arm cylinder 132C, and the first travel motor 112R. Hereinafter, the set of the swing motor 127, arm cylinder 132C, and first travel motor 112R connected to the front control valve 126A will also be referred to as the front actuator unit UA. In other embodiments, the front actuator unit UA may consist of a single actuator. The swing motor 127 is provided with a swing load pressure sensor 127P that measures the pressure (load pressure) on the inlet side of the swing motor 127. The arm cylinder 132C is provided with an arm load pressure sensor 132CP that measures the pressure (load pressure) on the inlet side of the arm cylinder 132C. The first travel motor 112R is equipped with a front travel load pressure sensor 112RP that measures the pressure (load pressure) on the inlet side of the first travel motor 112R.

[0026] The rear control valve 126B is connected to the rear main oil passage 123B. The rear control valve 126B distributes the hydraulic fluid supplied from the rear main oil passage 123B to the boom cylinder 131C, the bucket cylinder 133C, and the second travel motor 112L. Hereinafter, the set of boom cylinder 131C, bucket cylinder 133C, and second travel motor 112L connected to the rear control valve 126B will also be referred to as the rear actuator unit UB. In other embodiments, the rear actuator unit UB may consist of a single actuator. The boom cylinder 131C is provided with a boom load pressure sensor 131CP that measures the pressure (load pressure) on the inlet side of the boom cylinder 131C. The bucket cylinder 133C is provided with a bucket load pressure sensor 133CP that measures the pressure (load pressure) on the inlet side of the bucket cylinder 133C. The second travel motor 112L is equipped with a rear travel load pressure sensor 112LP that measures the pressure (load pressure) on the inlet side of the second travel motor 112L. The front control valve 126A or the rear control valve 126B may be equipped with a port that supplies hydraulic fluid to actuators (such as breakers, grapples, and tilt rotators) provided on the work tool (attachment).

[0027] Each control valve 126 has a flow rate adjustment valve 1261 and a pressure compensation valve 1262 for each actuator. The flow rate adjustment valve 1261 adjusts the flow rate of the hydraulic oil supplied to the corresponding actuator according to a command from the control device 145. The pressure compensation valve 1262 prevents the hydraulic oil from being unevenly distributed to the actuators on the low load side even if the load pressures of the actuators provided in the same control valve 126 (main oil passage 123) are different. The pressure compensation valve 1262 according to the first embodiment applies a pressure loss to the shaft with a low load pressure so that the outlet pressure of the flow rate adjustment valve 1261 of the actuator with a low load pressure becomes equal to the outlet pressure of the flow rate adjustment valve 1261 of the actuator with the maximum load pressure. Thereby, since the differential pressure between the inflow side and the outflow side of each flow rate adjustment valve 1261 becomes equal, it is possible to realize flow rate distribution using the meter-in opening area ratio. Therefore, an LS (load sensing) oil passage 128 that shares the hydraulic oil at the outlet of each flow rate adjustment valve 1261 is provided in the same control valve 126. Each pressure compensation valve 1262 is connected to the LS oil passage 128. Thereby, the pressure compensation valve 1262 can detect the maximum load pressure. Hereinafter, the LS oil passage 128 of the front control valve 126A is referred to as a front LS oil passage 128A, and the LS oil passage 128 of the rear control valve 126B is referred to as a rear LS oil passage 128B. Note that the pressure compensation valve 1262 according to another embodiment may be provided on the upstream side of the flow rate adjustment valve 1261.

[0028] The LS connection oil passage 129 connects the front LS oil passage 128A and the rear LS oil passage 128B. An LS merging and separating valve 1291 is provided in the LS connection oil passage 129. The LS merging and separating valve 1291 is a shut-off valve that controls the opening and closing of the LS connection oil passage 129. Thereby, the LS merging and separating valve 1291 can switch whether to merge or separate the hydraulic oil flowing through the front LS oil passage 128A and the hydraulic oil flowing through the rear LS oil passage 128B. When the front LS oil passage 128A and the rear LS oil passage 128B merge, the maximum load pressures of the actuators connected to the front control valve 126A and the actuators connected to the rear control valve 126B are applied to each pressure compensation valve 1262.

[0029] "Configuration of Control Device 145" Figure 3 is a schematic block diagram showing the configuration of the control device 145 according to the first embodiment. The control device 145 according to the first embodiment includes a measurement value acquisition unit 51, an operation amount acquisition unit 52, a target flow rate determination unit 53, a load torque estimation unit 54, a confluence separation control unit 55, an engine control unit 56, a pump control unit 57, and a valve control unit 58.

[0030] The measurement value acquisition unit 51 acquires sensor data indicating measurement values from various sensors (front pump discharge pressure sensor 1231A, rear pump discharge pressure sensor 1231B, swing load pressure sensor 127P, arm load pressure sensor 132CP, front travel load pressure sensor 112RP, boom load pressure sensor 131CP, bucket load pressure sensor 133CP, rear travel load pressure sensor 112LP). If there is a port for supplying hydraulic oil to the attachment in the front control valve 126A or the rear control valve 126B, the work machine 1 may include a load pressure sensor for the attachment. The operation amount acquisition unit 52 receives an operation command from the operation device 142 and specifies the operation amount of each actuator. When the work machine 1 has an automatic control function, the operation amount acquisition unit 52 may acquire the operation amount calculated by the automatic control function.

[0031] The target flow rate determination unit 53 determines the target value of the flow rate of hydraulic oil required to drive each actuator according to the operation amount acquired by the operation amount acquisition unit 52. The load torque estimation unit 54 obtains an estimated value of the load torque of the hydraulic pump 122. The load torque of the hydraulic pump 122 means the load torque applied to the engine 121 by the hydraulic pump 122 (the torque consumed by the hydraulic pump 122). The estimated value of the load torque estimated by the load torque estimation unit 54 is the value of the load torque consumed by the hydraulic pump 122 after a certain period of time based on the operation amount acquired by the operation amount acquisition unit 52 and the operating state of the hydraulic pump 122. The method for estimating the load torque will be described later.

[0032] The confluence separation control unit 55 controls the main confluence separation valve 1241 and the LS confluence separation valve 1291 based on the target value of the hydraulic oil determined by the target flow rate determination unit 53.

[0033] The engine control unit 56 controls the fuel injection amount and other parameters of the engine 121 so that the engine speed approaches the target speed through feedback control based on the measured rotational speed of the engine 121. The target speed is determined, for example, by the value indicated by the rotational speed adjustment knob acquired by the manipulated variable acquisition unit 52.

[0034] The pump control unit 57 controls the swash plate angle of each hydraulic pump 122 based on the target flow rate determined by the target flow rate determination unit 53 and the merging and separation state of the main oil passage 123. The valve control unit 58 controls the opening degree of the flow rate adjustment valve 1261 and bleed valve 1232 of each control valve 126 based on the target flow rate determined by the target flow rate determination unit 53 and the merging and separation state of the main oil passage 123.

[0035] 《Determination of Target Flow Rate》 The target flow rate determination unit 53 determines the required flow rate corresponding to the operating amount of each actuator based on a conversion table that shows the relationship between the operating amount of each actuator and the required flow rate, which is predetermined. On the other hand, since the discharge amount of the hydraulic pump 122 does not change instantaneously, it takes time for the flow rate of the hydraulic fluid to reach the required flow rate. Therefore, the target flow rate determination unit 53 predicts the target value of the flow rate in the actuator unit U after a certain period of time, based on the current transient target flow rate and the steady-state target flow rate. The certain period of time is the time equivalent to the response delay of the controlled equipment.

[0036] Figure 4 is a flowchart showing the method for predicting the target flow rate after a certain period of time according to the first embodiment. The target flow rate determination unit 53 predicts the target flow rate after a certain period of time for each actuator unit U and each actuator in the following procedure.

[0037] The target flow rate determination unit 53 obtains the manipulated amount of each actuator from the manipulated amount acquisition unit 52 and determines the required flow rate for each actuator from the acquired manipulated amounts according to a predetermined conversion table (step S1). The target flow rate determination unit 53 determines whether the sum of the required flow rates of the actuators is greater than the sum of the maximum flow rates of the hydraulic pump 122, or the total upper limit flow rate of the actuator unit U obtained by dividing the upper limit horsepower of the multiple actuators by the pump discharge pressure (step S2). The pump discharge pressure used in calculating the total upper limit flow rate of the actuator unit U is determined based on the measured value of the pump discharge pressure sensor 1231. If the sum of the required flow rates of the actuators is greater than the sum of the maximum flow rates of the hydraulic pump 122 or the total upper limit flow rate (step S2: YES), the target flow rate determination unit 53 determines a target value for the required flow rate of each actuator so that its sum matches the sum of the maximum flow rates of the hydraulic pump 122 or the total upper limit flow rate, while maintaining the ratio of the required flow rates of each actuator (step S3). If the sum of the requested flow rates of the actuators is less than or equal to the sum of the maximum flow rates of the hydraulic pumps 122 and the total upper limit flow rate (step S2: NO), the target flow rate determination unit 53 determines the requested flow rate obtained in step S1 as the target value for the requested flow rate of each actuator (step S4).

[0038] The target flow rate determination unit 53 performs the following steps S5 to S8 for each of the front main oil passage 123A and the rear main oil passage 123B. First, the target flow rate determination unit 53 determines the target flow rate for the next calculation cycle based on the current target flow rate and the target required flow rate identified in step S3 or step S4 (step S5). Specifically, the target flow rate determination unit 53 determines the target flow rate after one calculation cycle by adding the current target flow rate to a value obtained by applying a predetermined filter to the difference between the current target flow rate and the target required flow rate. The filter may be, for example, a low-pass filter or a limiter that limits the rate of change. By applying a filter to the difference between the current target flow rate and the target required flow rate, the target flow rate determination unit 53 can prevent abrupt changes in the target flow rate. Next, the target flow rate determination unit 53 estimates the target flow rate for the next calculation cycle (the nth calculation cycle from the current point in time) based on the target flow rate for the next calculation cycle (the nth calculation cycle from the current point in time) and the target required flow rate identified in step S3 or step S4 (step S6). The target flow rate determination unit 53 repeats the calculation in step S6 until the time taken for the n+1th calculation cycle coincides with a time a certain period of time after the current point in time (step S7). If the time taken for the n+1th calculation cycle coincides with a time a certain period of time after the current point in time (step S7: YES), the target flow rate determination unit 53 determines the target flow rate obtained in the n+1th calculation as the target flow rate after the certain period of time (step S8). For example, if one calculation cycle is 10 milliseconds and the certain period of time is 1 second, the target flow rate determination unit 53 can estimate the target flow rate after 1 second by repeating the calculation in step S6 99 times. Hereinafter, the target flow rate value of the front actuator unit UA after a certain period of time will be referred to as target value Tua, and the target flow rate value of the rear actuator unit UB after a certain period of time will be referred to as target value Tub.

[0039] Figure 5 is a flowchart illustrating the method for determining the target flow rate of the hydraulic pump 122 according to the first embodiment. Hereinafter, of the front actuator unit UA and the rear actuator unit UB, the one with the larger target flow rate will be called the first actuator unit U1, and the one with the smaller target flow rate will be called the second actuator unit U2. Also, of the front hydraulic pump 122A and the rear hydraulic pump 122B, the one corresponding to the first actuator unit will be called the first hydraulic pump, and the one corresponding to the second actuator unit will be called the second hydraulic pump. For example, if the target flow rate Qua of the front actuator unit is greater than the target flow rate Qub of the rear actuator unit, then the first actuator unit U1 is the front actuator unit UA, the second actuator unit U2 is the rear actuator unit UB, the first hydraulic pump is the front hydraulic pump 122A, and the second hydraulic pump is the rear hydraulic pump 122B. In this case, the target flow rate of the first actuator unit Qu1 = Qua, and the target flow rate of the second actuator unit Qu2 = Qub.

[0040] The target flow rate determination unit 53 determines whether the target value Qu1 of the flow rate of the first actuator unit, obtained by the process shown in Figure 4, is less than or equal to the flow rate threshold Qt1 (first flow rate threshold) of the first hydraulic pump, and whether the target value Qu2 of the flow rate of the second actuator unit is less than or equal to the flow rate threshold Qt2 (second flow rate threshold) of the second hydraulic pump (step S13). The flow rate threshold Qt may be a value obtained by subtracting a margin based on a predetermined safety factor from the maximum flow rate Qmax of the hydraulic pump 122. For example, the flow rate threshold Qt may be 90% of the maximum flow rate Qmax of the hydraulic pump 122.

[0041] If the target flow rate Qu1 of the first actuator unit is less than or equal to the flow rate threshold Qt1 of the first hydraulic pump and the target flow rate Qu2 of the second actuator unit is less than or equal to the flow rate threshold Qt2 of the second hydraulic pump (step S13: YES), the target flow rate determination unit 53 determines the target flow rate Qp1 of the first hydraulic pump to be the target flow rate Qu1 of the first actuator unit, and determines the target flow rate Qp2 of the second hydraulic pump to be the target flow rate Qu2 of the second actuator unit (step S14).

[0042] If the target flow rate of the first actuator unit exceeds the flow rate threshold of the first hydraulic pump, or if the target flow rate of the second actuator unit exceeds the flow rate threshold of the second hydraulic pump (step S13: NO), the target flow rate determination unit 53 determines whether the main connecting oil passage 124 is in a merged state or not (step S15). If the main connecting oil passage 124 is in a separated state (step S15: NO), the target flow rate determination unit 53 determines the target flow rate Qp1 of the first hydraulic pump to be the smaller of the target flow rate Qu1 of the first actuator unit and the maximum flow rate Qmax1 of the first hydraulic pump (step S16). The target flow rate determination unit 53 determines the target flow rate Qu2 of the second actuator unit to be the target flow rate Qp2 of the second hydraulic pump (step S17).

[0043] If the main connecting oil passage 124 is in a merging state (step S15: YES), the target flow rate determination unit 53 calculates a determined flow rate Qs by adding the larger of zero and the value obtained by subtracting the maximum flow rate Qmax1 of the first hydraulic pump from the required flow rate Qu1 of the first actuator unit to the target value Qu2 of the flow rate of the second actuator unit (step S18). The target flow rate determination unit 53 determines whether the determined flow rate Qs is smaller than the minimum flow rate Qmin2 of the second hydraulic pump (step S19).

[0044] If the determined flow rate Qs is smaller than the minimum flow rate Qmin2 of the second hydraulic pump (step S19: YES), the target flow rate determination unit 53 determines the target value Qp2 of the flow rate of the second hydraulic pump to be the minimum flow rate Qmin2 of the second hydraulic pump (step S20). Next, the target flow rate determination unit 53 determines the target value Qp1 of the first hydraulic pump to be the flow rate obtained by subtracting the minimum flow rate Qmin2 of the second pump from the sum of the target value Qu1 of the flow rate of the first actuator unit and the target value Qu2 of the flow rate of the second actuator unit (step S21). In this case, a support flow rate (Qp2 - Qu2), which is the total flow rate of the hydraulic fluid discharged by the second hydraulic pump minus the flow rate consumed by the second actuator unit, is supplied to the first actuator unit. The support flow rate at this time is the difference between the target value Qp1 of the flow rate of the first hydraulic pump and the target value Qu1 of the flow rate of the first actuator unit.

[0045] If the determined flow rate Qs is greater than or equal to the minimum flow rate Qmin2 of the second hydraulic pump (step S19: NO), the target flow rate determination unit 53 determines the target value Qp2 of the flow rate of the second hydraulic pump to be the determined flow rate Qs obtained in step S18 (step S22). Next, the target flow rate determination unit 53 determines the target value Qp1 of the first hydraulic pump to be the smaller of the maximum flow rate Qmax1 of the first hydraulic pump and the required flow rate Qu1 of the first actuator unit (step S23). In this case, the support flow rate (Qp2 - Qu2), which is the total flow rate of hydraulic fluid discharged by the second hydraulic pump minus the flow rate consumed by the second actuator unit, is supplied to the first actuator unit. The support flow rate at this time is the difference between the minimum flow rate Qmin2 of the second hydraulic pump and the target value Qp2 of the flow rate of the second actuator unit. As a result, the control device 145 can use the difference between the minimum flow rate Qmin2 output by the second hydraulic pump and the flow rate Qp2 consumed by the second actuator unit as a support flow rate for the first actuator unit without releasing it through the bleed valve 1232.

[0046] As a result, the target flow rate determination unit 53 can determine the target value Qa for the flow rate of each actuator, the target value Qu for the flow rate of each actuator unit U, and the target value Qp for the flow rate of each hydraulic pump 122. By determining the target value Qp for the flow rate of the hydraulic pump 122 using the procedure described above, the supply flow rates of each hydraulic pump 122 can be made different even if the main oil passage 123 is merged. As a result, the control device 145 can supply a support flow rate from the second hydraulic pump to the first actuator unit if the target value Qu1 for the flow rate of the first actuator unit cannot be met by the maximum flow rate of the first hydraulic pump. The support flow rate according to the first embodiment is smaller than the flow rate passing through the main merge / separation valve 1241 when the supply flow rates of each hydraulic pump 122 are the same. As a result, the control device 145 can suppress the pressure loss that occurs in the main merge / separation valve 1241.

[0047] 《Estimation of Load Torque of Hydraulic Pump 122》 The mechanical energy output by the engine 121 is transmitted to the hydraulic pump 122. When the load torque applied to the engine 121 by the hydraulic pump 122 becomes excessive and the rotational speed of the hydraulic pump 122 becomes zero, it becomes impossible to transmit energy to the actuator. Hereinafter, the load torque applied to the engine 121 by the hydraulic pump 122 will also be referred to as the load torque of the hydraulic pump 122. The angular velocity of the hydraulic pump 122 correlates with the balance between the sum of the load torques of the hydraulic pump 122 and the torque output by the engine 121. That is, the hydraulic pump 122 accelerates when the torque balance, obtained by subtracting the load torque of the hydraulic pump 122 from the torque output by the engine 121, is positive, and the hydraulic pump 122 decelerates when the torque balance is negative. When the hydraulic pump 122 decelerates, the engine 121 tries to recover from deceleration by increasing its output torque, but due to the effect of response delay, the state of negative torque balance continues for a considerable period of time. Since the hydraulic pump 122 continues to decelerate while the torque balance is negative, there is a risk that the rotational speed of the hydraulic pump 122 will drop to zero before the torque balance discrepancy is resolved. Therefore, the control device 145 can prevent the deceleration of the hydraulic pump 122 by controlling the engine 121 and the hydraulic pump 122 based on an estimated value of the load torque after a certain period of time from the hydraulic pump 122, thereby compensating for the effect of the response delay.

[0048] The load torque estimation unit 54 calculates the estimated load torque after a certain period of time using the front hydraulic pump 122A by following the procedure below. The same procedure can be used to calculate the estimated load torque after a certain period of time for the rear hydraulic pump 122B.

[0049] The load torque estimation unit 54 obtains an estimated value of the discharge pressure of the front hydraulic pump 122A after a certain period of time from the measured value of the front pump discharge pressure sensor 1231A obtained by the measurement value acquisition unit 51. For example, the load torque estimation unit 54 can obtain an estimated value of the discharge pressure after a certain period of time by obtaining the differential value of the discharge pressure, i.e., the rate of change of the discharge pressure, from the time series of the measured value of the front pump discharge pressure sensor 1231A, multiplying this rate of change by a certain period of time, and adding it to the current discharge pressure. Alternatively, for example, the load torque estimation unit 54 may obtain an estimated value of the discharge pressure after a certain period of time by applying phase lead compensation to the time series of the measured value of the front pump discharge pressure sensor 1231A. Based on the estimated value of the discharge pressure after a certain period of time and the target value Qpa of the flow rate of the front hydraulic pump 122A, the load torque estimation unit 54 obtains an estimated value of the load torque by the front hydraulic pump 122A after a certain period of time. In other words, the load torque estimation unit 54 can obtain an estimated value of the load torque after a certain period of time by the front hydraulic pump 122A by dividing the product of the estimated value of the discharge pressure after a certain period of time and the target value of the flow rate Qpa by 2π and multiplying it by the mechanical efficiency of the front hydraulic pump 122A.

[0050] In the first embodiment, the load torque estimation unit 54 determines the estimated load torque after a certain period of time by multiplying the target value of the flow rate after a certain period of time determined by the target flow rate determination unit 53 by the estimated value of the discharge pressure after a certain period of time obtained from the measured value of the discharge pressure. However, it is not limited to this. For example, in another embodiment, the load torque estimation unit 54 may determine the target value of the flow rate after a certain period of time by multiplying the current target value of the flow rate of the hydraulic pump 122 by a phase lead compensation for a certain period of time, and then determine the estimated load torque after a certain period of time by multiplying this by the estimated value of the discharge pressure after a certain period of time. Alternatively, in another embodiment, the load torque estimation unit 54 may determine the current load torque from the current target value of the flow rate of the hydraulic pump 122 and the current measured value of the discharge pressure, and then determine the estimated load torque after a certain period of time by multiplying this by a phase lead compensation for a certain period of time.

[0051] 《Merging and Separation Determination》 The merging and separation control of the work machine 1 according to the first embodiment will be described. The drive system 20 according to the first embodiment has a front main oil passage 123A and a rear main oil passage 123B separated by a main merging and separation valve 1241. When the main merging and separation valve 1241 is closed, hydraulic fluid discharged from the front hydraulic pump 122A flows through the front main oil passage 123A, and hydraulic fluid discharged from the rear hydraulic pump 122B flows through the rear main oil passage 123B. In this case, the front hydraulic pump 122A supplies hydraulic fluid at the required flow rate to the actuator connected to the front control valve 126A, and the rear hydraulic pump 122B supplies hydraulic fluid at the required flow rate to the actuator connected to the rear control valve 126B. Here, the flow rate of hydraulic fluid required by one of the control valves 126 may exceed the maximum discharge amount of one hydraulic pump 122. In such situations, the control device 145 opens the main confluence separation valve 1241 to control the actuator using the hydraulic fluid discharged by the two hydraulic pumps 122, thereby avoiding a power reduction due to insufficient hydraulic fluid.

[0052] Figure 6 is a flowchart showing the method for determining whether the two main oil passages 123 merge or separate according to the first embodiment. The measurement value acquisition unit 51 determines whether there is an abnormality in at least one of the front pump discharge pressure sensor 1231A and the rear pump discharge pressure sensor 1231B (step S41). For example, the measurement value acquisition unit 51 determines that there is an abnormality in the pump discharge pressure sensor 1231 if the measured value of the pump discharge pressure sensor 1231 is greater than a threshold for determining a sensor fault. Alternatively, for example, the measurement value acquisition unit 51 determines that there is an abnormality in the pump discharge pressure sensor 1231 if the measured value of the pump discharge pressure sensor 1231 is less than a threshold for determining a sensor fault.

[0053] If there is an abnormality in the pump discharge pressure sensor 1231 (step S41: YES), the confluence / separation control unit 55 decides to confluence the front main oil passage 123A and the rear main oil passage 123B (step S42). By converging the front main oil passage 123A and the rear main oil passage 123B, the pressures in the front main oil passage 123A and the rear main oil passage 123B become equal. As a result, for example, if there is an abnormality in the front pump discharge pressure sensor 1231A and the rear pump discharge pressure sensor 1231B is normal, the measured value of the rear pump discharge pressure sensor 1231B can be considered as the measured value of the pressure in the front main oil passage 123A. Therefore, if there is an abnormality in either the front pump discharge pressure sensor 1231A or the rear pump discharge pressure sensor 1231B, the measurement value acquisition unit 51 substitutes the measured value of the pump discharge pressure sensor 1231 that is not abnormal for the measured value of the pump discharge pressure sensor 1231 that is abnormal. If both the front pump discharge pressure sensor 1231A and the rear pump discharge pressure sensor 1231B are malfunctioning, the measurement value acquisition unit 51 substitutes a predetermined fixed value for the measured value of each pump discharge pressure sensor 1231. The fixed value may be, for example, the set pressure of a relief valve (safety valve) not shown. Since the pump discharge pressure does not generally exceed the set pressure by a large margin, substituting the set pressure for the measured value of the pump discharge pressure sensor 1231 allows control to continue even if there is a malfunction in the pump discharge pressure sensor 1231. Alternatively, the fixed value may be the pump discharge pressure value during high-load operation determined by prior experiments (for example, approximately μ (mean value) + σ (standard deviation)). This allows the control device 145 to continue controlling the work machine 1 even if there is a malfunction in the pump discharge pressure sensor 1231.

[0054] If there is no abnormality in the pump discharge pressure sensor 1231 (step S41: NO), the confluence / separation control unit 55 determines whether or not the traveling body 110 is being operated based on the operation amount acquired by the operation amount acquisition unit 52 (step S43). If the traveling body 110 is being operated (step S43: YES), the confluence / separation control unit 55 determines whether or not there is any operation other than that of the traveling body 110 (step S44). If there is no operation other than that of the traveling body 110 (step S44: NO), the confluence / separation control unit 55 decides to separate the front main oil passage 123A and the rear main oil passage 123B (step S48). This allows the flow rate of the hydraulic fluid supplied to the first traveling motor 112R to be controlled by the front hydraulic pump 122A, and the flow rate of the hydraulic fluid supplied to the second traveling motor 112L to be controlled by the rear hydraulic pump 122B. On the other hand, if there is an operation other than that of the traveling body 110 (step S44: YES), the merging / separating control unit 55 decides to merge the front main oil passage 123A and the rear main oil passage 123B (step S42).

[0055] If there is no operation of the traveling body 110 (step S43: NO), the merging / separating control unit 55 determines the operation amount of the rotating body 120 to a predetermined rotating operation amount threshold Th based on the operation amount acquired by the operation amount acquisition unit 52. ｓ１ The above conditions are met, and the amount of control of the boom 131 is the boom control threshold Th ｂ１ It is determined whether or not the above is true (step S45). The amount of operation of the slewing body 120 is the slewing operation threshold Th ｓ１ The above conditions are met, and the amount of control of the boom 131 is the boom control threshold Th ｂ１ If the above conditions are met (step S47: YES), the merger / separation control unit 55 decides to merge the front main oil passage 123A and the rear main oil passage 123B (step S42).

[0056] The amount of operation of the slewing body 120 is the slewing operation threshold Th ｓ１ If the amount of the boom 131 is not greater than or equal to the boom operating amount threshold Th ｂ１ If the above is not the case (Step S45: NO), the confluence / separation control unit 55 determines that the amount of operation of the boom 131 is equal to the boom operation threshold Th based on the amount of operation acquired by the operation amount acquisition unit 52. ｂ２is satisfied and the operation amount of the arm 132 is equal to or greater than a predetermined arm operation amount threshold Th ａ１ is satisfied and it is determined whether or not the front main oil passage 123A and the rear main oil passage 123B are separated (step S46). When the operation amount of the boom 131 is equal to or greater than the boom operation amount threshold Th ｂ２ is satisfied and the operation amount of the arm 132 is equal to or greater than a predetermined arm operation amount threshold Th ａ１ is satisfied and the front main oil passage 123A and the rear main oil passage 123B are separated (step S46: YES), the merging / separation control unit 55 determines to separate the front main oil passage 123A and the rear main oil passage 123B (step S48).

[0057] When the operation amount of the boom 131 is not equal to or greater than the boom operation amount threshold Th ｂ２ and the operation amount of the arm 132 is not equal to or greater than a predetermined arm operation amount threshold Th ａ１ or the front main oil passage 123A and the rear main oil passage 123B are merged (step S46: NO), the merging / separation control unit 55 determines whether or not the target value of the flow rate after a certain time determined by the target flow rate determination unit 53 exceeds the flow rate threshold Th Ｑ for at least one of the front actuator unit and the rear actuator unit (step S47). The flow rate threshold Th Ｑ may be a value obtained by subtracting a margin based on a predetermined safety factor from the maximum flow rate of the hydraulic pump 122. For example, the flow rate threshold Th Ｑ may be 90% of the maximum discharge amount of the hydraulic pump 122. Further, the flow rate threshold Th Ｑ may have hysteresis. That is, the flow rate threshold Th Ｑ for the merging determination and the flow rate threshold Th Ｑ for the separation determination may be different values.

[0058] In either the front actuator unit or the rear actuator unit, the sum of the target values of the flow rates after a certain time is the flow rate threshold Th ＱIf the pressure does not exceed (step S47: NO), the confluence separation control unit 55 decides to separate the front main oil passage 123A and the rear main oil passage 123B (step S48). In other words, if the supply of hydraulic fluid to the main oil passage 123 can be provided by the corresponding hydraulic pump 122, the work efficiency can be improved by separating the front main oil passage 123A and the rear main oil passage 123B and preventing pressure loss caused by the main confluence separation valve 1241.

[0059] In at least one of the front actuator unit and the rear actuator unit, the target value of the flow rate after a certain period of time is set to the flow rate threshold Th Ｑ If it exceeds this value (Step S47: YES), there is a possibility that the required flow rate of the main oil passage 123 will exceed the maximum flow rate of the hydraulic pump 122 in the near future. Therefore, by merging the front main oil passage 123A and the rear main oil passage 123B in advance (Step S42), the effects of pressure and flow rate changes of the hydraulic fluid associated with the merging can be suppressed.

[0060] Step S46 in Figure 6 will be explained. Depending on the work performed by the work machine 1, the merging of the lines may reduce work efficiency. Specifically, in work involving the combined operation of the boom 131 and the arm 132, the difference in load pressure between the boom cylinder 131C and the arm cylinder 132C becomes large, and when the front main oil passage 123A and the rear main oil passage 123B are merged, the drive system 20 has a pressure compensation valve 1262, resulting in a loss corresponding to the difference in load pressure. Examples of work involving the simultaneous operation of the boom 131 and the arm 132 include lightly leveling the ground with the tip of the bucket 133, excavation work, and extending the arm 132 while lowering the boom 131. In the work of leveling the ground with the tip of the bucket 133, the boom 131 is raised while the arm 132 is pulled. At this time, the reaction force from the ground is small, so the load pressure on the arm cylinder 132C is small, while the load pressure on the boom cylinder 131C increases in order to raise the boom 131 so that the tip of the bucket 133 does not fall. During excavation work, the arm 132 is pulled back significantly while the boom 131 is raised to prevent over-excavation. At this time, the load pressure on the arm cylinder 132C increases as it excavates into the ground, while the work machine 130 receives an upward force due to the reaction force from the ground. The boom cylinder 131C is lifted by this reaction force, so the load pressure on the boom cylinder 131C becomes very small. When the boom 131 is lowered while the arm 132 is extended, the boom 131 is lowered in the direction of gravity, so the load pressure on the boom cylinder 131C becomes very small. On the other hand, the arm 132 is being operated against gravity, so the load pressure on the arm cylinder 132C increases. Therefore, in the first embodiment, the control device 145 maintains the separation of the front main oil passage 123A and the rear main oil passage 123B without merging them in advance during certain operations where a decrease in work efficiency is expected.

[0061] 《Merging and Separation Control》 The merging and separation control unit 55 decides whether to separate or merge the main oil passages 123 using the procedure described above. On the other hand, if the two main oil passages 123 are merged while there is a pressure difference between them, the hydraulic fluid from the high-pressure main oil passage 123 may flow excessively into the low-pressure main oil passage 123, potentially causing the actuator's behavior to become temporarily unstable. Therefore, the control device 145 according to the first embodiment, when merging the separated main oil passages 123, equalizes the pressures of the two main oil passages 123 before merging them. Specifically, when the merging and separation control unit 55 decides to merge the main oil passages 123, it first opens the LS merging and separation valve 1291 at a predetermined rate. This equalizes the pressures of the front LS oil passage 128A and the rear LS oil passage 128B. The confluence / separation control unit 55 opens the main confluence / separation valve 1241 after a certain period of time has elapsed since opening the LS confluence / separation valve 1291. At this time, the confluence / separation control unit 55 opens the main confluence / separation valve 1241 at a rate corresponding to the amount of operation of each actuator (target flow rate). In other words, the confluence / separation control unit 55 increases the rate and opens the main confluence / separation valve 1241 more quickly as the total amount of operation of the actuators increases. This allows for a rapid change in flow rate in situations where a rapid change in speed is required. On the other hand, when a rapid change in speed is not required, the rate can be lowered to suppress pressure fluctuations associated with the switching between confluence and separation.

[0062] As a result, the confluence / separation control unit 55 can confluence the main oil passages 123 only after sufficiently reducing the pressure difference between them, thereby preventing instability caused by pressure differences. Note that a certain amount of time elapsed after opening the LS confluence / separation valve 1291 is one example of a condition under which the pressure difference between the front main oil passage 123A and the rear main oil passage 123B can be considered sufficiently small.

[0063] On the other hand, when the confluence separation control unit 55 decides to separate the confluenced main oil passages 123, it first closes the main confluence separation valve 1241. At this time, the confluence separation control unit 55 closes the main confluence separation valve 1241 at a rate corresponding to the operating state of each actuator. After a certain period of time has elapsed since closing the main confluence separation valve 1241, the confluence separation control unit 55 closes the LS confluence separation valve 1291 at a predetermined rate.

[0064] 《Engine Control》 When the rotational speed of engine 121 becomes zero, energy transmission becomes impossible, so the engine control unit 56 controls engine 121 to prevent its rotational speed from decreasing. The engine control unit 56 controls the fuel injection amount of engine 121 and other parameters to bring the rotational speed of engine 121 closer to the target value through feedback control based on the measured rotational speed of engine 121. On the other hand, since feedback control does not function until the rotational speed has decreased, it is possible that feedback control alone may not be able to adequately suppress the decrease in rotational speed. Here, the load torque estimated by the load torque estimation unit 54 after a certain period of time corresponds to the load torque required for engine 121 in the near future. Therefore, in the first embodiment, the engine control unit 56 determines a target value for the load torque of engine 121 by feedforward control based on the load torque estimated by the load torque estimation unit 54, in addition to feedback control based on the measured rotational speed of engine 121. The load torque required for the engine 121 includes not only the load torque of the hydraulic pump 122, but also the load torque of other auxiliary equipment (e.g., cooling fan, generator (alternator), etc.) and rotational friction torque. However, torques other than the load torque of the hydraulic pump 122 do not usually fluctuate significantly. Therefore, the engine control unit 56 may treat the load torque of the engine 121 as the sum of the load torque of the hydraulic pump determined by the load torque estimation unit 54 and the other torques that have been previously determined as integral terms, or it may ignore the other torques.

[0065] Figure 7 is a block diagram showing the control of the engine 121 by the engine control unit 56 according to the first embodiment. The engine control unit 56 comprises a first subtraction point 561, a feedback control unit 562, a summation point 563, and a command value generation block 564. The first subtraction point 561 calculates the difference between the target rotational speed and the measured rotational speed of the engine 121. The feedback control unit 562 obtains a feedback term for the target torque by performing feedback control using the difference between the target value and the measured value input from the first subtraction point 561.

[0066] The summing point 563 calculates the target torque by adding the calculation result of the feedback control unit 562 and the estimated load torque after a certain period of time estimated by the load torque estimation unit 54. The command value generation block 564 generates commands for the engine 121, such as the fuel injection amount and throttle opening, based on the target torque.

[0067] The load torque estimated by the load torque estimation unit 54 represents the load torque after a time has elapsed equivalent to the response delay of the engine 121. This makes it possible to shorten the time during which the rotational speed of the engine 121 decreases. Figure 8 shows the difference in engine control when feedforward control based on future load torque is performed and when it is not. Figure 8A shows the engine control when feedforward control based on future load torque is not performed. Figure 8B shows the engine control when feedforward control based on future load torque is performed. When the engine control unit 56 does not perform feedforward control based on future load torque, that is, when the engine control unit 56 controls the engine 121 based on the current load torque, the engine 121 changes after a predetermined delay time from the control command by the engine control unit 56. If the load torque of the hydraulic pump 122 changes at a predetermined rate, the torque of the engine 121 cannot keep up with the load torque of the hydraulic pump 122, so the rotational speed may continue to decrease while the load torque of the hydraulic pump 122 is changing.

[0068] On the other hand, in the case where the engine control unit 56 performs feedforward control based on future load torque, as in the first embodiment, if the load torque of the hydraulic pump 122 is increasing, the estimated value of the load torque will be greater than the current load torque, as shown in Figure 8B. As a result, the engine control unit 56 can quickly bring the torque of the engine 121 closer to the load torque of the hydraulic pump 122, and as shown in Figure 8B, the time during which the rotational speed of the hydraulic pump 122 decreases can be shortened.

[0069] 《Control of Hydraulic Pump 122》 The pump control unit 57 determines the upper limit capacity of the hydraulic pump 122 based on the maximum torque of the engine 121. Based on the upper limit capacity and the target flow rate determined by the target flow rate determination unit 53, the pump control unit 57 determines a target value for the capacity of the hydraulic pump 122 and controls the swash plate of the hydraulic pump 122 according to the target value for capacity.

[0070] Figure 9 is a flowchart showing the method for determining the upper limit capacity of the hydraulic pump 122 according to the first embodiment. The pump control unit 57 determines the maximum torque of the engine 121 from the rotational speed of the engine 121 (step S61). The relationship between the rotational speed of the engine 121 and the maximum torque is determined in advance by experiments or other means.

[0071] The pump control unit 57 determines the acceleration torque required for the change in rotational speed from the time rate of change of the target rotational speed (target angular acceleration) (step S62). The time rate of change of the target rotational speed is determined, for example, by the amount of operation of the rotational speed adjustment knob acquired by the operation amount acquisition unit 52, or by whether or not deceleration control is performed. Note that if there is no change in the rotational speed of the engine 121, the acceleration torque is zero. The pump control unit 57 calculates the total upper limit torque that the front hydraulic pump 122A and the rear hydraulic pump 122B can consume by subtracting the load torque of the auxiliary equipment, the rotational friction torque, and the acceleration torque from the maximum torque of the engine 121 (step S63). However, torques other than the load torque of the hydraulic pump 122 do not usually fluctuate significantly. Therefore, if the engine control unit 56 eliminates the consumption of the load torque and rotational friction torque of the auxiliary equipment by feedback control or feedforward control such as separate integral control, the engine control unit 56 may control the engine 121 based on the torque obtained by subtracting the sum of the load torque of the auxiliary equipment and the acceleration torque from the maximum torque of the engine 121.

[0072] The pump control unit 57 determines the upper limit torque of the front hydraulic pump 122A and the rear hydraulic pump 122B by distributing the total upper limit torque calculated in step S63 according to the ratio of the load torques of the front hydraulic pump 122A and the rear hydraulic pump 122B estimated by the load torque estimation unit 54 (step S64). As a result, if the acceleration torque is not zero, the sum of the upper limit torques of the hydraulic pumps 122 will be less than the maximum torque of the engine 121. As a result, the pump control unit 57 does not use up all of the maximum torque of the engine 121 with the hydraulic pumps 122, leaving room for acceleration in the engine 121.

[0073] Next, the pump control unit 57 determines an equitorque curve function that shows the relationship between the discharge pressure of the hydraulic pump 122 and the capacity of the hydraulic pump 122, based on the determined upper limit torque of the hydraulic pump 122 (step S65). In the equitorque curve function, discharge pressure and capacity are inversely proportional. That is, the rate of change of capacity increases as the discharge pressure decreases. Therefore, if the capacity is determined from the discharge pressure according to the equitorque curve function, when pressure fluctuations occur in the low-pressure region, the capacity fluctuation may become excessive, and the swash plate control of the hydraulic pump 122 may not be able to keep up, and swash plate hunting may occur.

[0074] The pump control unit 57 identifies a discharge pressure threshold in the constant torque curve function where the slope exceeds a predetermined slope threshold (for example, a rate of change in capacity that can be achieved in consideration of the responsiveness of the swash plate), and finds a tangent function, which is a linear function that has contact with the constant torque curve function at the discharge pressure threshold (step S66). Figure 10 is a diagram showing the relationship between the constant torque curve function and the tangent function according to the first embodiment. The pump control unit 57 determines whether the measured value of the discharge pressure is equal to or greater than the discharge pressure threshold (step S67: YES). If the measured value of the discharge pressure is equal to or greater than the discharge pressure threshold (step S67: YES), the pump control unit 57 determines the upper limit capacity (first upper limit capacity) of the hydraulic pump 122 according to the measured value of the discharge pressure and the constant torque curve function (step S68). Control according to the upper limit capacity determined from the constant torque curve function is control according to the upper limit torque of the hydraulic pump 122.

[0075] If the measured discharge pressure is less than the discharge pressure threshold (step S67: NO), the pump control unit 57 determines the upper limit capacity (first upper limit capacity) of the hydraulic pump 122 according to the measured discharge pressure and the tangent function (step S69). As shown in Figure 10, the upper limit capacity determined from the tangent function is always smaller than the upper limit capacity determined from the constant torque curve function. In other words, control according to the upper limit capacity determined from the tangent function is control according to a torque smaller than the upper limit torque of the hydraulic pump 122. In this way, the pump control unit 57 can control the swash plate according to a rate of change that does not exceed a predetermined slope threshold in the low-pressure region, thus preventing swash plate hunting.

[0076] In other embodiments, the function that determines the upper limit capacity when the measured discharge pressure is below the discharge pressure threshold does not have to be a tangent function. For example, in other embodiments, when the measured discharge pressure is below the discharge pressure threshold, a horizontal function, which is a linear function with a slope of zero that always shows the value of the upper limit capacity corresponding to the discharge pressure threshold in the constant torque curve function, may be used. In other embodiments, any function that lies between the tangent function and the horizontal function may be used as long as the rate of change is not excessive. With such a function, when the measured discharge pressure is below the discharge pressure threshold, control can be performed according to a torque that is smaller than the upper limit torque corresponding to the measured discharge pressure and greater than or equal to the upper limit torque at the discharge pressure threshold.

[0077] Next, the pump control unit 57 identifies the response characteristics of the swash plate of the hydraulic pump 122 based on the operating state of the hydraulic pump 122 (capacity, load pressure, and rotational speed), and determines the transient upper limit capacity by adding these response characteristics to the first upper limit capacity determined in step S68 or step S69 (step S70).

[0078] (Regarding transient upper capacity limits) Since the control of the swash plate of the hydraulic pump 122 toward the upper torque limit of the hydraulic pump 122 is performed by feedback control, there is a possibility that the load torque of the hydraulic pump 122 may overshoot the upper torque limit of the hydraulic pump 122. If an overshoot of load torque occurs, it becomes an overload and causes a decrease in the rotational speed of the engine 121. In order to prevent overshoot, the pump control unit 57 in the first embodiment determines a transient upper capacity limit by applying a filter according to the response characteristics to the first upper capacity limit for achieving the upper torque limit of the hydraulic pump 122.

[0079] The pump control unit 57 controls the current actual torque T of the hydraulic pump 122. ａｃｔ and upper limit torque T ｌｉｍ From a certain time t ｃ Later, upper limit torque T ｌｉｍ The allowable rate of change R of torque to reach ｌｉｍ Identify the allowable rate of change of torque R. ｌｉｍThis can be expressed by equation (1). Current actual torque T ａｃｔ The allowable rate of change R of time Δt and torque ｌｉｍ The sum of the products of these factors equals the transient upper limit torque of the hydraulic pump 122.

[0080]

[0081] The pump control unit 57 determines the upper limit capacity (second upper limit capacity) for the next control based on equation (1) above. For example, the pump control unit 57 can obtain the second upper limit capacity by the following calculation. First, the pump control unit 57 determines the rate of change of the actual torque over time based on the pump capacity and the pump discharge pressure. Next, the pump control unit 57 determines a target value for the rate of change of the upper limit capacity over time by feedback control based on the difference between the allowable rate of change of torque obtained by equation (1) and the rate of change of the actual torque over time. The pump control unit 57 then determines the second upper limit capacity based on the determined current pump capacity.

[0082] Therefore, the pump control unit 57 determines the smaller of the first upper limit capacity and the second upper limit capacity as the transient upper limit capacity. The difference between the first upper limit capacity and the second upper limit capacity corresponds to a margin according to the response characteristics of the hydraulic pump 122.

[0083] (Static Determination Adjustment) In the first embodiment, the hydraulic pump 122 can maintain a constant supply flow rate regardless of the load pressure by adjusting the swash plate angle by electronic control. On the other hand, since the hydraulic fluid has a bulk modulus, it compresses and expands with the acceleration and deceleration of the actuator. This causes fluctuations in the actuator's speed. The pump control unit 57 estimates the oscillation of the actuator's speed based on the flow rate balance, which shows the difference between the flow rate of the hydraulic fluid supplied to the actuator and the change in the volume of the hydraulic fluid on the actuator's inlet side, and corrects the target value of the flow rate based on the estimation result to cancel out the fluctuations in the actuator's speed.

[0084] Figure 11 is a block diagram showing the control of the hydraulic pump 122 by the pump control unit 57 according to the first embodiment. The pump control unit 57 includes a flow rate balance estimation unit 571, a filter 572, an inverse phase calculation unit 573, a phase compensation block 574, a first summation point 575, a proportional block 576, a limit block 577, a subtraction point 578, a feedback control unit 579, a feedforward control unit 580, and a second summation point 581.

[0085] The flow rate balance estimation unit 571 estimates the flow rate balance, which is the difference between the flow rate of the hydraulic fluid supplied to the actuator and the change in the volume of the hydraulic fluid at the actuator's inlet side. The flow rate balance estimation unit 571 can estimate the flow rate balance, for example, by following the procedure below. Integrating the actuator's flow rate balance over time yields pressure. Therefore, the flow rate balance estimation unit 571 can estimate the actuator's flow rate balance by differentiating the measured pressure at the actuator's inlet side. Alternatively, for example, the flow rate balance estimation unit 571 may estimate the change in the volume of the hydraulic fluid at the actuator's inlet side and estimate the difference between this change and the target value of the actuator's flow rate determined by the target flow rate determination unit 53 as the flow rate balance. The change in the volume of the hydraulic fluid at the actuator's inlet side can be estimated, for example, from the cylinder's velocity and pressure-receiving area, or from the hydraulic motor's angular velocity and motor capacity. The flow rate balance estimation unit 571 can estimate the velocity or angular velocity by obtaining the change in position from the encoder, cylinder stroke sensor, and IMU. Alternatively, the flow rate balance estimation unit 571 may estimate the displacement of the actuator by image recognition of an image in which the actuator is captured, and then estimate the velocity or angular velocity.

[0086] The filter 572 extracts vibration components from the flow balance estimated by the flow balance estimation unit 571. The filter 572 may be a high-pass filter that extracts high-frequency components. Alternatively, the filter 572 may be a band-pass filter. The frequency of the rate of change in the flow balance required for acceleration and deceleration by actuator control is sufficiently low compared to the frequency of vibration components. Therefore, the filter 572 can remove the change due to actuator control and extract vibration components due to the elasticity of the hydraulic fluid. By having the filter 572 extract vibration components and the pump control unit 57 controlling the hydraulic pump 122 to cancel out these vibration components, vibrations can be suppressed without hindering the acceleration and deceleration of the actuator. On the other hand, in the pump control unit 57 of another embodiment, instead of the calculation by the filter 572, the pump control unit 57 may multiply by a gain greater than 0 and less than 1. In this case, the pump control unit 57 cannot cancel out all vibration components, but it can reduce the amount of cancellation of the acceleration and deceleration of the actuator. Furthermore, since the hydraulic pump 122 system being controlled has damping characteristics, vibrations can be suppressed to a degree that does not cause discomfort to the operator by appropriately setting the gain, even without completely canceling out the vibration components.

[0087] The inverse phase calculation unit 573 determines the inverse phase of the oscillation component of the flow rate balance output by the filter 572. That is, the inverse phase calculation unit 573 multiplies the oscillation component of the flow rate balance by -1. Alternatively, instead of multiplying by -1, the inverse phase calculation unit 573 may multiply the second derivative or second integral of the flow rate balance (the phase is inverse of the flow rate balance, but the gain is different) by the gain. The inverse phase of the oscillation component is added to the target flow rate at the first summing point 575 as a correction term for the pump discharge rate.

[0088] The phase compensation block 574 performs phase compensation processing on the discharge amount correction term output by the inverse phase calculation unit 573 to compensate for the response delay of the hydraulic pump 122 system. The compensation parameters of the phase compensation block 574 are determined from the natural frequency and response characteristics of the hydraulic pump 122 system and are determined experimentally in advance. By having the phase compensation block 574, the pump control unit 57 can prevent the discharge amount correction term from resonating due to the response delay of the hydraulic pump 122 system. The phase compensation block 574 may perform either phase lead compensation or phase lag compensation. Note that since the hydraulic pump 122 system has damping characteristics, in other embodiments, if the resonance component due to the response delay can be sufficiently absorbed by the damping characteristics, the pump control unit 57 does not need to have the phase compensation block 574. The first summing point 575 calculates the sum of the target value of the flow rate and the discharge amount correction term output by the phase compensation block 574.

[0089] The proportional block 576 determines the target value of the hydraulic pump 122's capacity by dividing the flow rate output by the first summing point 575 by the rotational speed N. The limit block 577 limits the target value of the capacity output by the proportional block 576 to less than or equal to the transient upper limit capacity determined by the procedure shown in Figure 9. The subtraction point 578 calculates the difference between the target value of the capacity limited by the limit block 577 and the actual capacity of the hydraulic pump 122 obtained from the measurement value of the swash plate angle sensor. The feedback control unit 579 determines the feedback term for the target value of the swash plate angle of the hydraulic pump 122 by feedback control based on the difference between the target value of the capacity and the actual capacity. The feedforward control unit 580 determines the feedforward term for the target value of the swash plate angle of the hydraulic pump 122 by feedforward control based on the target value of the capacity. The second summing point 581 calculates the target value of the swash plate angle of the hydraulic pump 122 by adding the calculation result of the feedback control unit 579 and the calculation result of the feedforward control unit 580.

[0090] The pump control unit 57 controls the hydraulic pump 122 by outputting a command value for the swash plate angle corresponding to a target value of the capacity limited to below the transient upper limit capacity. Furthermore, while the main confluence separation valve 1241 is changing from a closed state to an open state, the pump control unit 57 changes the target value of the flow rate at the same rate as the rate of change of the opening degree of the main confluence separation valve 1241.

[0091] 《Valve Control》 The valve control unit 58 controls the opening degree of the control valve 126 according to the target flow rate of the actuator determined by the target flow rate determination unit 53. As a result, each actuator is supplied with hydraulic fluid at a flow rate according to the target flow rate value.

[0092] Furthermore, when the main oil passage 123 is separated, if the target flow rate of at least one actuator unit U is smaller than the minimum flow rate of the corresponding hydraulic pump 122, the valve control unit 58 controls the opening degree of the bleed valve 1232 to discharge a flow rate equal to the difference between the minimum flow rate and the target flow rate. Specifically, the valve control unit 58 controls the opening degree A of the bleed valve 1232 of the main oil passage 123. ｂｌｅｅｄ This can be determined according to the following equation (2).

[0093]

[0094] Q ｐ This represents the discharge volume of the hydraulic fluid from the hydraulic pump 122. Q ｖａｌｖｅ This represents the target value of the flow rate of actuator unit U. ｔｇｔ This represents the target value of the load pressure of the actuator unit U. In step S19 of the flowchart shown in Figure 5, if the determined flow rate Qs is greater than or equal to the minimum flow rate Qmin2 of the second hydraulic pump, the bleed valve 1232 is fully closed tp.

[0095] When the main oil passage 123 is merged by the merge / separation control unit 55, if there is a difference in the discharge pressure of the two pumps, a large amount of hydraulic fluid may flow from the high-pressure oil passage to the low-pressure oil passage, potentially causing the actuator of the high-pressure oil passage to decelerate rapidly and the actuator of the low-pressure oil passage to accelerate rapidly. In the drive system 20 according to the first embodiment, since it is equipped with a pressure compensation valve 1262, the valve control unit 58 can equalize the discharge pressure of the two pumps in advance by opening the LS merge / separation valve before starting to open the main merge / separation valve 1241. In the third embodiment, which is not equipped with a pressure compensation valve as described later, the discharge pressure of the two pumps is equalized in advance by increasing the pressure loss by throttling the flow control valve 1261 and the bleed valve 1232 of the low-pressure oil passage before starting to open the main merge / separation valve 1241. Specifically, the valve control unit 58 restricts the flow rate control valve 1261 when the pump discharge flow rate of the low-pressure oil passage is equal to or greater than the minimum flow rate, and restricts the bleed valve 1232 when the pump discharge flow rate of the low-pressure oil passage is less than the minimum flow rate.

[0096] 《Operation and Effects》 According to the work machine 1 of the first embodiment, when the requested flow rate of the front actuator unit UA exceeds the flow rate threshold of the front hydraulic pump 122A, the control device 145 determines the target flow rate of the rear hydraulic pump 122B to be the requested flow rate of the rear actuator unit UB plus the support flow rate to be supplied to the front actuator unit UA, and opens the main confluence separation valve 1241. As a result, the control device 145 can reduce the pressure loss occurring in the main confluence separation valve 1241 by suppressing the support flow rate supplied from the rear hydraulic pump 122B through the main confluence separation valve 1241 to the front actuator unit UA. Similarly, when the requested flow rate of the rear actuator unit UB exceeds the flow rate threshold of the rear hydraulic pump 122B, the control device 145 determines the target flow rate of the front hydraulic pump 122A to be the requested flow rate of the front actuator unit UA plus the support flow rate to be supplied to the rear actuator unit UB, and opens the main confluence separation valve 1241. As a result, the control device 145 can reduce the pressure loss occurring in the main confluence separation valve 1241 by suppressing the support flow rate supplied from the front hydraulic pump 122A through the main confluence separation valve 1241 to the rear actuator unit UB. Furthermore, according to the work machine 1 of the first embodiment, when switching the main oil passage 123 from separation to confluence, the LS confluence separation valve 1291 is opened first, followed by the opening of the main confluence separation valve 1241. This reduces the difference in pump discharge pressure between the front actuator unit UA and the rear actuator unit UB before the main oil passage 123 merges. Therefore, it is possible to prevent the actuator's behavior from becoming unstable due to the difference in pump discharge pressure between the front actuator unit UA and the rear actuator unit UB at the time of confluence.

[0097] According to the working machine 1 of the first embodiment, when switching between merging and separating the main oil passage 123, the opening degree of the main merging / separating valve 1241 is gradually changed according to a predetermined rate. This prevents the flow rate to the actuator from changing abruptly during the switching between merging and separation, which would cause the actuator's behavior to become unstable.

[0098] According to the work machine 1 of the first embodiment, if the support flow rate to be supplied to the front actuator unit UA is less than the minimum flow rate of the rear hydraulic pump 122B, the control device 145 sets the target flow rate of the front hydraulic pump 122A to a flow rate obtained by reducing the required flow rate of the front actuator unit UA by the difference between the required flow rate and the minimum flow rate of the rear actuator unit UB. This allows the difference between the required flow rate and the minimum flow rate of the rear actuator unit UB to be utilized by the front actuator unit UA without being released through the bleed valve 1232. Similarly, if the support flow rate to be supplied to the rear actuator unit UB is less than the minimum flow rate of the front hydraulic pump 122A, the control device 145 sets the target flow rate of the rear hydraulic pump 122B to a flow rate obtained by reducing the required flow rate of the rear actuator unit UB by the difference between the required flow rate and the minimum flow rate of the front actuator unit UA. This allows the difference between the required flow rate and the minimum flow rate of the front actuator unit UA to be utilized by the rear actuator unit UB without being released through the bleed valve 1232.

[0099] According to the first embodiment of the work machine 1, the control device 145 determines a predicted value of the required flow rate of the actuator unit after a certain period of time based on the required flow rate of the actuator unit, and opens the main confluence / separation valve 1241 when the predicted value of the required flow rate exceeds an upper threshold that is less than or equal to the flow rate threshold of the hydraulic pump 122. This makes it possible to switch between confluence and separation of the main oil passage 123 by compensating for the response delay of the hydraulic pump 122.

[0100] According to the work machine 1 of the first embodiment, the control device 145 maintains the closed state of the main confluence separation valve 1241 when the main confluence separation valve 1241 is closed, the boom 131 and arm 132 are operated simultaneously, the amount of operation of the boom 131 is greater than or equal to the boom operation threshold, the amount of operation of the arm 132 is greater than or equal to the arm operation threshold, and the traveling body 110 is not operated. This makes it possible to maintain the separation of the main oil passage 123 in work where efficiency would decrease if the main oil passage 123 were to merge.

[0101] According to the first embodiment of the work machine 1, when an abnormality occurs in the pump discharge pressure sensor 1231, the main oil passages 123 are merged. In other words, the control device 145 merges the main oil passage 123 in which the abnormal pump discharge pressure sensor 1231 is located with another main oil passage 123 having another pump discharge pressure sensor 1231. As a result, the pressures of the two main oil passages 123 become equal, and the measured value of the abnormal pump discharge pressure sensor 1231 is read as the measured value of the pump discharge pressure sensor 1231 without abnormality, thereby allowing control of the work machine 1 to continue.

[0102] In the first embodiment of the work machine 1, the control device 145 estimates the load torque of the hydraulic pump 122 from the target value of the discharge volume and the measured value of the discharge pressure of the hydraulic pump 122, and controls the engine 121 based on this. This allows the control device 145 to suppress the decrease in the rotational speed of the engine 121 caused by response delay. In other embodiments, the control device 145 may control the engine 121 based on a load torque value calculated from the pump capacity measured by the swash plate angle sensor 1221 and the pump discharge pressure measured by the pump discharge pressure sensor 1231, instead of an estimated load torque.

[0103] According to the first embodiment of the work machine 1, the control device 145 controls the hydraulic pump 122 based on a target value obtained by subtracting the acceleration torque required for accelerating the engine 121 from the maximum torque of the hydraulic pump 122. In other words, the control device 145 suppresses the control target of the hydraulic pump 122 when acceleration of the engine 121 is necessary, and does not suppress the control target of the hydraulic pump 122 when acceleration is not necessary. As a result, the control device 145 does not use up all of the maximum torque of the engine 121 with the hydraulic pump 122, leaving room for acceleration in the engine 121.

[0104] According to the first embodiment of the work machine 1, the control device 145 identifies the response characteristics of the hydraulic pump 122 based on the operating state of the hydraulic pump 122 and controls the hydraulic pump based on the upper limit torque of the hydraulic pump 122, which is obtained by subtracting a margin corresponding to the response characteristics from the maximum torque of the hydraulic pump 122. As a result, the control device 145 can appropriately control the capacity of the hydraulic pump 122 according to the response characteristics.

[0105] According to the first embodiment of the work machine 1, the control device 145 obtains an estimated value related to the vibration of the actuator's speed and determines the control amount of the hydraulic pump 122 based on the estimated value. As a result, the control device 145 can suppress the vibration of the actuator caused by the elasticity of the hydraulic fluid.

[0106] According to the work machine 1 of the first embodiment, the control device 145 determines the upper limit torque of the hydraulic pump 122 corresponding to the measured discharge pressure as a target value when the measured discharge pressure of the hydraulic pump 122 is equal to or greater than a predetermined pressure, and determines a torque smaller than the upper limit torque of the hydraulic pump 122 corresponding to the measured discharge pressure as a target value for the discharge pressure when the measured discharge pressure of the hydraulic pump 122 is less than the predetermined pressure. As a result, the control device 145 can prevent the swash plate of the hydraulic pump 122 from hunting when the discharge pressure is low.

[0107] <Second Embodiment> Figure 12 is a schematic block diagram showing the configuration of the drive system 20 according to the second embodiment. The engine 121 of the work machine 1 according to the second embodiment has a turbocharger 1211. The turbocharger 1211 compresses the intake air of the engine 121 by rotating a turbine with the exhaust gas of the engine 121. In other words, the turbocharger 1211 changes the work efficiency of the engine 121 according to the magnitude of the thermal energy of the exhaust gas of the engine 121. The turbocharger 1211 is an example of a mechanism that changes the output characteristics of the prime mover. In other embodiments, the work machine 1 may have a supercharger, a variable intake manifold, a variable valve, a variable compression ratio mechanism, etc., as mechanisms that change the output characteristics of the prime mover.

[0108] An engine 121 equipped with a mechanism such as a turbocharger 1211 cannot always output maximum torque. For engine 121 to produce maximum torque, the turbocharger 1211 needs to be rotating at a sufficient speed due to the exhaust from engine 121. The time required from when engine 121 with a turbocharger 1211 outputs a command to produce maximum torque according to the engine speed of engine 121 until it can actually produce maximum torque is called turbo lag.

[0109] Therefore, if the control device 145 controls the hydraulic pump 122 based on the maximum torque regardless of the state of the engine 121, the load torque will become excessive, causing the rotational speed of the engine 121 to decrease. In the second embodiment, the pump control unit 57 of the control device 145 determines the upper limit torque that the engine 121 can output according to the state of the turbocharger 1211, and limits the target value of the load torque to less than or equal to that upper limit torque. The relationship between the state of the turbocharger 1211 and the upper limit torque that the engine 121 can output is determined in advance by experiments or other means.

[0110] In the second embodiment, since the engine 121 is a diesel engine, the relationship between the state of the engine 121 and turbocharger 1211 and the upper limit of the fuel injection amount of the engine 121 is determined in advance by experimentation or other means. The pump control unit 57 identifies the upper limit of the fuel injection amount from the state of the engine 121 and turbocharger 1211, and estimates the upper limit of the torque that the engine 121 can output according to the current state of the turbocharger 1211 by multiplying the ratio of the upper limit of the fuel injection amount to the current fuel injection amount by the current output torque estimate. In other embodiments, if the engine 121 is a gasoline engine, the engine control unit 56 can optimize the ignition timing of the engine 121 from the state of the engine 121 and turbocharger 1211. Therefore, the pump control unit 57 may estimate the upper limit of the torque that the engine 121 can output according to the current state of the turbocharger 1211 based on the current estimated output torque of the hydraulic pump 122, the current ignition timing of the engine 121, and the output torque when the ignition timing is optimized. In a given operating state of engine 121, the relationship between the amount of deviation of the ignition timing toward the advance or retard side relative to the ignition timing that outputs maximum torque, and the torque reduction rate corresponding to the amount of deviation, can be determined in advance through experiments. In another embodiment, the pump control unit 57 may estimate the upper limit torque that engine 121 can output based on the current rotational speed of engine 121, intake air volume, intake air temperature, coolant temperature, and catalyst state, according to the current state of turbocharger 1211. The intake air volume changes according to the operating state of turbocharger 1211.

[0111] Figure 13 is a flowchart illustrating the method for determining the upper limit capacity of the hydraulic pump 122 according to the second embodiment. In the first embodiment, the pump control unit 57 determines the maximum torque of the engine 121 in step S61 of Figure 9. In contrast, the pump control unit 57 according to the second embodiment instead estimates the upper limit torque that the engine 121 can output according to the current state of the turbocharger 1211 (step S161). The pump control unit 57 determines the acceleration torque required for the change in rotational speed from the amount of change in target rotational speed (angular acceleration) (step S62).

[0112] The pump control unit 57, instead of performing the calculation in step S63 of the first embodiment, calculates a torque value obtained by subtracting the acceleration torque and other torques from the estimated upper limit torque of the engine 121, and sets this value as the total upper limit torque that the front hydraulic pump 122A and the rear hydraulic pump 122B can consume (step S163).

[0113] When controlling the hydraulic pump 122 based on the upper limit torque that the engine 121 can output according to the current state of the turbocharger 1211, the rotational speed of the engine 121 may deviate from the target value due to factors such as estimation errors in the upper limit torque that the engine 121 can output and differences in the responsiveness of the engine 121 and the hydraulic pump 122. If the rotational speed of the engine 121 deviates in the decreasing direction, the decrease in rotational speed may continue and the engine may stall. Also, if the rotational speed of the engine 121 deviates in the excessive direction, the output torque of the engine 121 will decrease due to rotational speed feedback control. As a result, exhaust energy will decrease, the activity of the turbocharger 1211 will decrease, and the upper limit torque that the engine 121 can output may decrease. Therefore, the pump control unit 57 according to the second embodiment adds a feedback control amount corresponding to the difference between the target rotational speed of the engine 121 and the measured value to the total upper limit torque that the front hydraulic pump 122A and the rear hydraulic pump 122B can consume (step S164). This feedback control amount is determined by integral control based on the difference between the target rotational speed of the engine 121 and the measured value. The pump control unit 57 calculates a feedback control amount corresponding to the difference between the target rotational speed and the measured value using integral control that can respond retrospectively and eliminate steady-state errors, thereby appropriately canceling out estimation errors that change depending on the operating conditions. When the upper limit torque that the engine 121 can output and the maximum torque of the engine 121 coincide, the feedback control amount is set to zero to prevent exceeding the maximum torque of the engine 121. In this case, the pump control unit 57 may gradually bring the feedback control amount closer to zero to suppress the effect of changes in the feedback control amount on the rotation of the engine 121. As a result, in the hydraulic pump 122 according to the second embodiment, as the turbocharger 1211 becomes more active over time and the upper limit torque that the engine 121 can output increases over time, the upper limit torque that the hydraulic pump 122 can consume becomes possible.

[0114] In the second embodiment, the engine control unit 56 requests torque output from the engine 121 based on the load torque estimated by the load torque estimation unit 54, similar to the first embodiment, regardless of the state of the turbocharger 1211. If the output of the engine 121 is suppressed according to the limited load torque, the output torque of the engine 121 and the load torque of the hydraulic pump 122 will balance out in a state where the turbocharger 1211 is not activated, and the state in which the turbocharger 1211 is not activated will continue. Therefore, the engine control unit 56 in the second embodiment can increase the rotational speed of the engine 121 and accelerate the activation of the turbocharger 1211 by requesting a torque output determined regardless of the state of the turbocharger 1211. If the rotational speed of the engine 121 becomes larger than the target value due to this control, the load torque of the hydraulic pump 122 can be increased by the feedback control amount in step S164 described above, thereby preventing a decrease in the output torque of the engine 121 and a decrease in the activation of the turbocharger due to feedback control.

[0115] 《Third Embodiment》 Figure 14 is a block diagram showing the configuration of the drive system 20 of the work machine 1 according to the third embodiment. The drive system 20 of the work machine 1 according to the first and second embodiments has a pressure compensation valve 1262 that compensates for the flow rate distribution to each actuator regardless of the difference in load pressure of the actuators. In contrast, the drive system 20 according to the third embodiment does not have a pressure compensation valve 1262, an LS oil passage 128, and an LS connecting oil passage 129. The confluence separation control unit 55 according to the first embodiment equalizes the pressure before the main oil passage 123 merges by opening the LS confluence separation valve 1291 and then opening the main confluence separation valve 1241. The control device 145 according to the third embodiment equalizes the pressure before the main oil passage 123 merges, even in a drive system 20 that does not have a load sensing function.

[0116] Figure 15 is a flowchart illustrating the confluence and separation control according to the third embodiment. When the confluence and separation control unit 55 according to the third embodiment decides to confluence the main oil passages 123, it identifies the pressure of each main oil passage 123 from the measurement value of the pump discharge pressure sensor 1231 (step S201). The confluence and separation control unit 55 determines whether the pressure difference between the front main oil passage 123A and the rear main oil passage 123B is below a predetermined threshold (step S202). The threshold may be a value that can be considered to indicate that the pressure difference between the main oil passages 123 is sufficiently small. In other words, the pressure difference between the front main oil passage 123A and the rear main oil passage 123B being below a predetermined threshold is equivalent to satisfying a predetermined condition that can be considered to indicate that the pressure difference is sufficiently small.

[0117] If the pressure difference between the front main oil passage 123A and the rear main oil passage 123B exceeds a threshold (step S202: NO), the valve control unit 58 controls the opening degree of each flow control valve 1261 by feedback control and feedforward control to match the pressure difference between the inlet and outlet sides of each flow control valve 1261 to a predetermined target differential pressure (step S203). At this time, the pressure on the inlet side of each flow control valve 1261 may be approximated by the pump pressure of the corresponding main oil passage 123. As a result, the valve control unit 58 controls the flow control valve 1261 so that the pressure difference between the two main oil passages 123 becomes smaller. This allows the valve control unit 58 to increase the pressure in the main oil passage 123 with relatively lower pressure. The confluence / separation control unit 55 then returns to step S201 and continues to monitor the pressure difference. In addition, in other embodiments, the valve control unit 58 may control the flow control valve 1261 as follows when the pressure difference between the front main oil passage 123A and the rear main oil passage 123B exceeds a threshold. The valve control unit 58 identifies the shaft with the maximum outlet pressure among the shafts of the multiple actuators as the maximum load pressure shaft, and determines the differential pressure between the inlet pressure and outlet pressure at the maximum load pressure shaft as the target differential pressure for all shafts. The valve control unit 58 controls the meter-in opening of the flow control valve 1261 corresponding to each actuator to achieve the determined target differential pressure.

[0118] On the other hand, if the pressure difference between the front main oil passage 123A and the rear main oil passage 123B is below a threshold in step S202 (step S202: YES), the confluence separation control unit 55 opens the main confluence separation valve 1241 at a predetermined rate (step S204) and terminates the confluence control.

[0119] Through the above process, the control device 145 according to the third embodiment can equalize the pressure before the main oil passage 123 merges in a drive system 20 that does not have a load sensing function.

[0120] <Other Embodiments> Although one embodiment has been described in detail above with reference to the drawings, the specific configuration is not limited to that described above, and various design changes are possible. That is, in other embodiments, the order of the above-described processes may be changed as appropriate. Also, some processes may be executed in parallel.

[0121] The work machine 1 in the above-described embodiment is a hydraulic excavator, but is not limited to this, and the work machine 1 in other embodiments may be other work machines such as a wheel loader or a bulldozer.

[0122] The work machine 1 according to the above embodiment is equipped with a swashplate type variable displacement pump as the hydraulic pump 122, but is not limited to this. For example, the hydraulic pump according to another embodiment may be a pump that changes the discharge amount by individually varying the rotation speed. In this case, the pump control unit 57 can control the rotation speed of the hydraulic pump 122.

[0123] The drive system 20 of the work machine 1 according to the above-described embodiment comprises two hydraulic pumps 122 and two main oil passages 123, but is not limited thereto. The drive system 20 of the work machine 1 according to other embodiments may comprise three or more hydraulic pumps 122 and three or more main oil passages 123.

[0124] <Computer Configuration> Figure 16 is a schematic block diagram showing the configuration of a computer according to at least one embodiment. The computer 90 includes a processor 91, main memory 92, storage 93, and interface 94. The control device 145 described above is implemented in the computer 90. The operation of each of the above-described processing units is stored in the storage 93 in the form of a program. The processor 91 reads the program from the storage 93, loads it into the main memory 92, and executes the above-described processing according to the program. The processor 91 also allocates storage areas in the main memory 92 corresponding to each of the above-described storage units according to the program. Examples of the processor 91 include a CPU (Central Processing Unit), a GPU (Graphic Processing Unit), and a microprocessor.

[0125] The program may be for implementing a part of the functions to be performed by the computer 90. For example, the program may perform functions in combination with other programs already stored in storage, or in combination with other programs implemented on other devices. In other embodiments, the computer 90 may include a custom LSI (Large Scale Integrated Circuit) such as a PLD (Programmable Logic Device) in addition to, or instead of, the above configuration. Examples of PLDs include PAL (Programmable Array Logic), GAL (Generic Array Logic), CPLD (Complex Programmable Logic Device), and FPGA (Field Programmable Gate Array). In this case, some or all of the functions implemented by the processor 91 may be implemented by the integrated circuit. Such an integrated circuit is also included as an example of a processor. In other embodiments, the computer 90 may be virtualized on one or more computers.

[0126] Examples of storage 93 include magnetic disks, magneto-optical disks, optical disks, and semiconductor memory. Storage 93 may be an internal medium directly connected to the bus of the computer 90, or it may be an external medium connected to the computer 90 via an interface 94 or a communication line. Furthermore, if this program is distributed to the computer 90 via a communication line, the computer 90 that receives the distribution may expand the program into the main memory 92 and execute the above processing. In at least one embodiment, storage 93 is a tangible storage medium that is not temporary.

[0127] Furthermore, the program may be intended to implement some of the functions described above. In addition, the program may be a so-called differential file (differential program) that implements the functions described above in combination with other programs already stored in the storage 93.

[0128] The control device 145 according to the above-described embodiment may be composed of a single computer 90, or the configuration of the control device 145 may be divided and arranged among multiple computers 90, with the multiple computers cooperating with each other to function as the control device 145. In this case, some of the computers 90 constituting the control device 145 may be mounted inside the work machine 1, and the other computers 90 may be provided outside the work machine 1. For example, if the work machine 1 is a remotely operated vehicle that receives operation commands from a remotely provided operating device and is driven accordingly, the control device 145 may be provided separately for the work machine 1 and the remote operating device. Also, for example, if the work machine 1 is a remotely operated vehicle that autonomously drives according to operation commands transmitted from a control device, the control device 145 may be provided separately for the work machine 1 and the control device.

[0129] According to the above embodiment, the working machine can suppress the decrease in the rotational speed of the prime mover due to the response delay.

[0130]

Claims

1. A working machine comprising: a prime mover; a variable-capacity hydraulic pump driven by the power of the prime mover and discharging hydraulic fluid; an actuator driven by the hydraulic fluid discharged by the hydraulic pump; and a control device, wherein the control device determines an estimated value of the load torque of the hydraulic pump based on a target value of the discharge volume and a measured value of the discharge pressure of the hydraulic pump, and controls the prime mover based on the estimated value of the load torque.

2. The control device determines an estimated value of the load torque of the hydraulic pump after a certain period of time, as described in claim 1.

3. The work machine according to claim 2, wherein the control device determines an estimated value of the load torque after a certain period of time based on the time change of the target value of the discharge amount, the time change of the measured value of the discharge pressure, or the time change of the estimated value of the current load torque.

4. The work machine according to claim 2, wherein the control device obtains an estimated value of the load torque after a certain period of time by applying phase lead compensation to the target value of the discharge volume, the measured value of the discharge pressure, or the estimated value of the current load torque.

5. The work machine according to claim 2, wherein the control device determines a target value for the discharge amount after a certain period of time from the discharge amount corresponding to the amount of operation of the actuator and the current discharge amount of the hydraulic pump, and estimates the load torque after a certain period of time based on the target value for the discharge amount after a certain period of time and the measured value of the discharge pressure.

6. A working machine comprising: a prime mover; a mechanism for changing the output characteristics of the prime mover; a variable-capacity hydraulic pump driven by the power of the prime mover and for discharging hydraulic fluid; an actuator driven by the hydraulic fluid discharged by the hydraulic pump; and a control device, wherein the control device determines a target value for the load torque of the hydraulic pump, limits the target value for the load torque to less than or equal to the upper limit torque that the prime mover can output according to the state of the mechanism, and controls the hydraulic pump based on the limited target value for the load torque.

7. The work machine according to claim 6, wherein the control device controls the prime mover based on a target value of the load torque of the hydraulic pump that is not limited by the upper limit torque.

8. The work machine according to claim 6, wherein the control device modifies the upper limit torque by integral control based on the difference between the measured rotational speed of the prime mover and the target rotational speed when the upper limit torque that the prime mover can output is less than the maximum torque of the prime mover and the target value of the load torque is limited by the upper limit torque.

9. A work machine comprising a variable-capacity hydraulic pump for discharging hydraulic fluid, and a control device, wherein the control device identifies the response characteristics of the load torque of the hydraulic pump based on the operating state of the hydraulic pump, and controls the hydraulic pump based on an upper limit torque obtained by subtracting a margin corresponding to the response characteristics of the load torque from the maximum torque of the hydraulic pump.

10. The work machine according to claim 9, wherein the control device determines an upper limit value of the discharge amount of the hydraulic pump based on the upper limit torque, limits the target value of the discharge amount by the upper limit value, and controls the hydraulic pump based on the limited target value of the discharge amount.

11. The work machine according to claim 9, wherein the control device specifies the amount of change in load torque after a certain period of time as the margin, based on the rate of change of the load torque of the hydraulic pump based on the response characteristics of the hydraulic pump and the rate of change of the load torque of the hydraulic pump based on pressure fluctuations.

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