JP2018105114A - Work vehicle and method for controlling the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hybrid type work vehicle capable of obtaining a large brake force while suppressing an excessive increase of an engine rotation speed during braking, and a method for controlling the same.SOLUTION: A pump brake torque control section increases a pump brake torque corresponding to a load of a hydraulic pump in pump brake control of generating a brake force by a load of a hydraulic pump. The pump brake torque control section determines a target braking power being a power to be absorbed by the work vehicle. The pump brake torque control section determines an engine regeneration power to be regenerated by the engine. When the target braking power is larger than the engine regeneration power, the pump brake torque control section controls the hydraulic pump so that a power value equal to a difference between the target braking power and the engine regeneration power is variably consumed.SELECTED DRAWING: Figure 11

Description

  The present invention relates to a work vehicle, particularly a hybrid work vehicle, and a control method thereof.

  2. Description of the Related Art In recent years, hybrid work vehicles that travel using a driving force from an engine and a driving force from a motor have been proposed. As a power transmission device for a hybrid work vehicle, for example, Patent Document 1 discloses an HMT (hydraulic-mechanical transmission) or an EMT (electro-mechanical transmission).

  The HMT has a planetary gear mechanism, and a first pump / motor and a second pump / motor connected to a rotating element of the planetary gear mechanism. The first pump / motor and the second pump / motor function as either a hydraulic motor or a hydraulic pump according to the traveling state of the work vehicle. The HMT is configured such that the rotation speed of the output shaft can be changed continuously by changing the rotation speed of these pumps / motors.

  In EMT, an electric motor is used instead of the hydraulic motor in HMT. That is, the EMT has a first generator / motor and a second generator / motor. The first generator / motor and the second generator / motor function as either an electric motor or a generator depending on the traveling state of the work vehicle. Similar to the HMT, the EMT is configured such that the rotation speed of the output shaft can be changed continuously by changing the rotation speed of the generator / motor.

  The hybrid power transmission device as described above is a power transmission device as compared with a conventional power transmission device having a torque converter and a multi-stage transmission (hereinafter referred to as a “torque converter transmission”). There is little internal loss. Therefore, the hybrid type power transmission device is superior in efficiency in that the driving force from the engine is transmitted to the traveling device to obtain the traction force, and there is an advantage that the fuel consumption is good.

JP 2006-329244 A

  In a conventional work vehicle including a torque converter type transmission, braking force by engine braking can be obtained. In this case, a part of the power absorbed by braking (hereinafter referred to as “braking power”) is discarded as heat in the torque converter, and the rest is absorbed in the engine.

  On the other hand, since the hybrid type power transmission device is efficient as described above, the amount of braking power discarded as heat is small. For this reason, most of the braking power is returned to the engine. In this case, if the engine tries to absorb all braking power, the engine speed may increase excessively.

  Further, if the braking power absorbed by the engine is limited, an excessive increase in the engine rotation speed can be suppressed, but in this case, there is a problem that the braking force obtained by the engine brake becomes small.

  In a hybrid work vehicle including a power storage device such as a capacitor, a part of the braking power can be stored as electric energy by generating the generator / motor with the braking power. However, when the power storage device is in a fully charged state, a part of the braking power cannot be stored as electric energy, so that also in this case, the braking force obtained by the engine brake becomes small.

  An object of the present invention is to provide a hybrid work vehicle that can obtain a large braking force while suppressing an excessive increase in engine rotation speed during braking, and a control method therefor.

  A first aspect of the present invention is a work vehicle, and includes an engine, a hydraulic pump, a travel device, a power transmission device, a power take-out device, and a control unit. The hydraulic pump is driven by the engine. The traveling device is driven by an engine. The power transmission device transmits the driving force from the engine to the traveling device. The power take-out device distributes the driving force from the engine to the hydraulic pump and the power transmission device. The control unit controls the hydraulic pump and the power transmission device.

  The power transmission device includes an input shaft, an output shaft, a gear mechanism, and a motor. The gear mechanism has a planetary gear mechanism and transmits the rotation of the input shaft to the output shaft. The motor is connected to the rotating element of the planetary gear mechanism. The power transmission device is configured to change the rotation speed ratio of the output shaft to the input shaft by changing the rotation speed of the motor.

  The control unit has a pump brake torque control unit. The pump brake torque control unit increases the pump brake torque corresponding to the load of the hydraulic pump in the pump brake control in which the braking force is generated by the load of the hydraulic pump. The pump brake torque control unit obtains a target braking power that is a power to be absorbed by the work vehicle. The pump brake torque control unit obtains engine regeneration power regenerated by the engine. When the target braking power is larger than the engine regeneration power, the pump brake torque control unit controls the hydraulic pump to variably consume a power value equal to the difference between the target braking power and the engine regeneration power.

  A second aspect of the present invention is a work vehicle control method, wherein the work vehicle includes an engine and a hydraulic pump. A travel device, a power transmission device, and a power take-out device are provided. The hydraulic pump is driven by the engine. The traveling device is driven by an engine. The power transmission device transmits the driving force from the engine to the traveling device. The power take-out device distributes the driving force from the engine to the hydraulic pump and the power transmission device.

  The power transmission device includes an input shaft, an output shaft, a gear mechanism, and a motor. The gear mechanism has a planetary gear mechanism and transmits the rotation of the input shaft to the output shaft. The motor is connected to the rotating element of the planetary gear mechanism. The power transmission device is configured to change the rotation speed ratio of the output shaft to the input shaft by changing the rotation speed of the motor. The work vehicle control method according to this aspect includes the following steps.

In the first step, a target braking power that is a power to be absorbed by the work vehicle is obtained. In the second step, engine regeneration power regenerated by the engine is obtained. In the third step, when the target braking power is larger than the engine regenerative power, the hydraulic pump of the work vehicle is controlled to change the power value equal to the difference between the target braking power and the maximum power consumed by the work vehicle. To consume.
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  A third aspect of the present invention is a work vehicle control method, wherein the work vehicle includes an engine and a hydraulic pump. A travel device, a power transmission device, and a power take-out device are provided. The hydraulic pump is driven by the engine. The traveling device is driven by an engine. The power transmission device transmits the driving force from the engine to the traveling device. The power take-out device distributes the driving force from the engine to the hydraulic pump and the power transmission device.

  The power transmission device includes an input shaft, an output shaft, a gear mechanism, and a motor. The gear mechanism has a planetary gear mechanism and transmits the rotation of the input shaft to the output shaft. The motor is connected to the rotating element of the planetary gear mechanism. The power transmission device is configured to change the rotation speed ratio of the output shaft to the input shaft by changing the rotation speed of the motor. The work vehicle control method according to this aspect includes the following steps.

  In the first step, the power consumption of the work vehicle is obtained. In the second step, when the power to be consumed by the work vehicle is larger than the power consumed by the work vehicle, the hydraulic pump is controlled to variably consume the power consumed exceeding the work vehicle's power consumption.

  In the work vehicle and the control method thereof according to one aspect of the present invention, a large braking force can be obtained while suppressing an excessive increase in the engine speed during braking.

It is a side view of the work vehicle concerning an embodiment. It is a schematic diagram which shows the structure of a work vehicle. It is a schematic diagram which shows the structure of a power transmission device. It is a figure which shows the change of the rotational speed of the 1st motor and 2nd motor with respect to a vehicle speed. It is a control block diagram which shows the determination process of the command torque to a motor. It is a graph which shows an example of a required tractive force characteristic. It is a schematic diagram which shows the flow of the braking power absorbed at the time of braking. It is a control block diagram which shows the process performed by the control part at the time of braking. It is a schematic diagram which shows the hydraulic circuit connected to the working machine pump. It is a flowchart which shows the determination process of execution of pump brake control. It is a control block diagram which shows the determination process of pump brake torque. It is a graph which shows an example of vehicle speed restriction brake torque information. It is a control block diagram which shows the determination process of the command value to a pump brake control valve. FIG. 6 is a schematic diagram showing a part of a hydraulic circuit provided in a work vehicle according to a first modification. FIG. 10 is a schematic diagram showing a configuration of a work vehicle according to a second modification.

  Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 is a side view of a work vehicle 1 according to an embodiment of the present invention. As shown in FIG. 1, the work vehicle 1 includes a body frame 2, a work implement 3, traveling wheels 4 and 5, and a cab 6. The work vehicle 1 is a wheel loader and travels when the traveling wheels 4 and 5 are rotationally driven. The work vehicle 1 can perform work such as excavation using the work machine 3.

  A work machine 3 and traveling wheels 4 and 5 are attached to the body frame 2. The work machine 3 is driven by hydraulic oil from a work machine pump 23 (see FIG. 2) described later. The work machine 3 includes a boom 11 and a bucket 12. The boom 11 is attached to the vehicle body frame 2. The work machine 3 includes a lift cylinder 13 and a bucket cylinder 14. The lift cylinder 13 and the bucket cylinder 14 are hydraulic cylinders. One end of the lift cylinder 13 is attached to the vehicle body frame 2. The other end of the lift cylinder 13 is attached to the boom 11. The boom 11 rotates up and down as the lift cylinder 13 expands and contracts with the hydraulic oil from the work implement pump 23. The bucket 12 is attached to the tip of the boom 11. One end of the bucket cylinder 14 is attached to the vehicle body frame 2. The other end of the bucket cylinder 14 is attached to the bucket 12 via a bell crank 15. As the bucket cylinder 14 expands and contracts with hydraulic oil from the work implement pump 23, the bucket 12 rotates up and down.

  A driver's cab 6 is attached to the body frame 2. The cab 6 is placed on the vehicle body frame 2. In the cab 6, a seat on which an operator is seated, an operation device to be described later, and the like are arranged. The vehicle body frame 2 has a front frame 16 and a rear frame 17. The front frame 16 and the rear frame 17 are attached so as to be rotatable in the left-right direction.

  The work vehicle 1 has a steering cylinder 18. The steering cylinder 18 is attached to the front frame 16 and the rear frame 17. The steering cylinder 18 is a hydraulic cylinder. As the steering cylinder 18 expands and contracts with hydraulic oil from a steering pump 30 described later, the traveling direction of the work vehicle 1 is changed to the left and right.

  FIG. 2 is a schematic diagram showing the configuration of the work vehicle 1. As shown in FIG. As shown in FIG. 2, the work vehicle 1 includes an engine 21, a power take-out device 22 (hereinafter referred to as “PTO22”), a power transmission device 24, a travel device 25, an operation device 26, a control unit 27, and the like. .

  The engine 21 is, for example, a diesel engine. The output of the engine 21 is controlled by adjusting the amount of fuel injected into the cylinder of the engine 21. The adjustment of the fuel amount is performed by the control unit 27 controlling the fuel injection device 28 attached to the engine 21. The work vehicle 1 includes an engine rotation speed detection unit 31. The engine rotation speed detection unit 31 detects the engine rotation speed and sends a detection signal indicating the engine rotation speed to the control unit 27.

  The work vehicle 1 includes a work machine pump 23, a steering pump 30, and a transmission pump 29. The work machine pump 23, the steering pump 30, and the transmission pump 29 are hydraulic pumps. PTO 22 (Power Take Off) transmits a part of the driving force from the engine 21 to these hydraulic pumps 23, 30, and 29. That is, the PTO 22 distributes the driving force from the engine 21 to these hydraulic pumps 23, 30, 29 and the power transmission device 24.

  The work machine pump 23 is driven by the driving force from the engine 21. The hydraulic oil discharged from the work implement pump 23 is supplied to the lift cylinder 13 and the bucket cylinder 14 described above via the work implement control valve 41. The work vehicle 1 includes a work machine pump pressure detection unit 32. The work machine pump pressure detection unit 32 detects the discharge pressure of hydraulic oil from the work machine pump 23 (hereinafter referred to as “work machine pump pressure”), and sends a detection signal indicating the work machine pump pressure to the control unit 27. .

  The work machine pump 23 is a variable displacement hydraulic pump. The discharge capacity of the work implement pump 23 is changed by changing the tilt angle of the swash plate or the oblique axis of the work implement pump 23. A first capacity control device 42 is connected to the work machine pump 23. The first capacity control device 42 is controlled by the control unit 27 and changes the tilt angle of the work implement pump 23. Thereby, the discharge capacity of the work implement pump 23 is controlled by the control unit 27. The work vehicle 1 includes a first tilt angle detection unit 33. The first tilt angle detection unit 33 detects the tilt angle of the work implement pump 23 and sends a detection signal indicating the tilt angle to the control unit 27.

  The steering pump 30 is driven by a driving force from the engine 21. The hydraulic oil discharged from the steering pump 30 is supplied to the above-described steering cylinder 18 via the steering control valve 43. The work vehicle 1 includes a steering pump pressure detection unit 34. The steering pump pressure detection unit 34 detects the discharge pressure of hydraulic oil from the steering pump 30 (hereinafter referred to as “steering pump pressure”), and sends a detection signal indicating the steering pump pressure to the control unit 27.

  The steering pump 30 is a variable displacement hydraulic pump. By changing the tilt angle of the swash plate or the oblique axis of the steering pump 30, the discharge capacity of the steering pump 30 is changed. A second capacity control device 44 is connected to the steering pump 30. The second capacity control device 44 is controlled by the control unit 27 and changes the tilt angle of the steering pump 30. Thereby, the discharge capacity of the steering pump 30 is controlled by the control unit 27. The work vehicle 1 includes a second tilt angle detection unit 35. The second tilt angle detection unit 35 detects the tilt angle of the steering pump 30, and sends a detection signal indicating the tilt angle to the control unit 27.

  The transmission pump 29 is driven by the driving force from the engine 21. The transmission pump 29 is a fixed displacement hydraulic pump. The hydraulic oil discharged from the transmission pump 29 is supplied to the clutches CF, CR, CL, and CH of the power transmission device 24 via clutch control valves VF, VR, VL, and VH described later.

  The PTO 22 transmits a part of the driving force from the engine 21 to the power transmission device 24. The power transmission device 24 transmits the driving force from the engine 21 to the traveling device 25. The power transmission device 24 shifts and outputs the driving force from the engine 21. The configuration of the power transmission device 24 will be described in detail later.

  The traveling device 25 includes an axle 45 and traveling wheels 4 and 5. The axle 45 transmits the driving force from the power transmission device 24 to the traveling wheels 4 and 5. As a result, the traveling wheels 4 and 5 rotate. The work vehicle 1 includes a vehicle speed detection unit 37. The vehicle speed detector 37 detects the rotational speed of the output shaft 63 of the power transmission device 24 (hereinafter referred to as “output rotational speed”). Since the output rotation speed corresponds to the vehicle speed, the vehicle speed detection unit 37 detects the vehicle speed by detecting the output rotation speed. Further, the vehicle speed detection unit 37 detects the rotation direction of the output shaft 63. Since the rotation direction of the output shaft 63 corresponds to the traveling direction of the work vehicle 1, the vehicle speed detection unit 37 detects the traveling direction of the work vehicle 1 by detecting the rotation direction of the output shaft 63. It functions as a part. The vehicle speed detection unit 37 sends a detection signal indicating the output rotation speed and the rotation direction to the control unit 27.

  The operating device 26 is operated by an operator. The operating device 26 includes an accelerator operating device 51, a work implement operating device 52, a shift operating device 53, a forward / reverse operating device 54 (hereinafter referred to as “FR operating device 54”), a steering operating device 57, and a brake operating device. 58.

  The accelerator operation device 51 includes an accelerator operation member 51a and an accelerator operation detection unit 51b. The accelerator operation member 51a is operated to set a target rotational speed of the engine 21. The accelerator operation detection unit 51b detects an operation amount of the accelerator operation member 51a (hereinafter referred to as “accelerator operation amount”). The accelerator operation detection unit 51b sends a detection signal indicating the accelerator operation amount to the control unit 27.

  The work machine operation device 52 includes a work machine operation member 52a and a work machine operation detection unit 52b. The work implement operating member 52a is operated to operate the work implement 3. The work machine operation detection unit 52b detects the position of the work machine operation member 52a. The work machine operation detection unit 52b outputs a detection signal indicating the position of the work machine operation member 52a to the control unit 27. The work machine operation detection unit 52b detects the operation amount of the work machine operation member 52a by detecting the position of the work machine operation member 52a.

  The shift operation device 53 includes a shift operation member 53a and a shift operation detection unit 53b. The operator can select the speed range of the power transmission device 24 by operating the speed change operation member 53a. The shift operation detecting unit 53b detects the position of the shift operation member 53a. The position of the speed change operation member 53a corresponds to a plurality of speed ranges such as first speed and second speed. The shift operation detection unit 53b outputs a detection signal indicating the position of the shift operation member 53a to the control unit 27.

  The FR operation device 54 includes a forward / reverse operation member 54a (hereinafter, “FR operation member 54a”) and a forward / reverse position detection unit 54b (hereinafter, “FR position detection unit 54b”). The operator can switch between forward and reverse travel of the work vehicle 1 by operating the FR operation member 54a. The FR operation member 54a is selectively switched between a forward position (F), a neutral position (N), and a reverse position (R). The FR position detector 54b detects the position of the FR operation member 54a. The FR position detection unit 54b outputs a detection signal indicating the position of the FR operation member 54a to the control unit 27.

  The steering operation device 57 includes a steering operation member 57a. The steering operation device 57 drives the steering control valve 43 by supplying pilot hydraulic pressure to the steering control valve 43 based on the operation of the steering operation member 57a. The steering operation member 57 may drive the steering control valve 43 by converting the operation of the steering operation member 57a into an electric signal. The operator can change the traveling direction of the work vehicle 1 to the left and right by operating the steering operation member 57a.

  The brake operation device 58 includes a brake operation member 58a and a brake operation detection unit 58b. The operator can operate the braking force of the work vehicle 1 by operating the brake operation member 58a. The brake operation detection unit 58b detects an operation amount of the brake operation member 58a (hereinafter referred to as “brake operation amount”). The brake operation detection unit 58b outputs a detection signal indicating the amount of brake operation to the control unit 27. The brake oil pressure may be used as the brake operation amount.

  The control unit 27 includes an arithmetic device such as a CPU and a memory such as a RAM and a ROM, and performs various processes for controlling the work vehicle 1. The control unit has a storage unit 56. The storage unit 56 stores various programs and data for controlling the work vehicle 1.

  The control unit 27 sends a command signal indicating a command throttle value to the fuel injection device 28 so that a target rotation speed of the engine 21 corresponding to the accelerator operation amount can be obtained. The control of the engine 21 by the control unit 27 will be described in detail later.

  The control unit 27 controls the hydraulic pressure supplied to the hydraulic cylinders 13 and 14 by controlling the work implement control valve 41 based on the detection signal from the work implement operation detecting unit 52b. Thereby, the hydraulic cylinders 13 and 14 expand and contract, and the work machine 3 operates.

  Further, the control unit 27 controls the power transmission device 24 based on detection signals from the respective detection units. The control of the power transmission device 24 by the control unit 27 will be described in detail later.

  Next, the configuration of the power transmission device 24 will be described in detail. FIG. 3 is a schematic diagram showing the configuration of the power transmission device 24. As shown in FIG. As shown in FIG. 3, the power transmission device 24 includes an input shaft 61, a gear mechanism 62, an output shaft 63, a first motor MG1, a second motor MG2, and a capacitor 64. The input shaft 61 is connected to the PTO 22 described above. The rotation from the engine 21 is input to the input shaft 61 via the PTO 22. The gear mechanism 62 transmits the rotation of the input shaft 61 to the output shaft 63. The output shaft 63 is connected to the traveling device 25 described above, and transmits the rotation from the gear mechanism 62 to the traveling device 25 described above.

  The gear mechanism 62 is a mechanism that transmits a driving force from the engine 21. The gear mechanism is configured to change the rotation speed ratio of the output shaft 63 to the input shaft 61 in accordance with the change in the rotation speed of the motors MG1, MG2. The gear mechanism 62 includes an FR switching mechanism 65 and a speed change mechanism 66.

  The FR switching mechanism 65 includes a forward clutch CF (hereinafter referred to as “F clutch CF”), a reverse clutch CR (hereinafter referred to as “R clutch CR”), and various gears (not shown). Yes. The F clutch CF and the R clutch CR are hydraulic clutches, and hydraulic fluid from the transmission pump 29 is supplied to the clutches CF and CR. The hydraulic fluid to the F clutch CF is controlled by the F clutch control valve VF. The hydraulic oil for the R clutch CR is controlled by the R clutch control valve VR. Each clutch control valve CF, CR is controlled by a command signal from the control unit 27.

  The direction of rotation output from the FR switching mechanism 65 is switched by switching between connection / disconnection of the F clutch CF and connection / disconnection of the R clutch CR. Specifically, when the vehicle moves forward, the F clutch CF is connected and the R clutch CR is disconnected. When the vehicle moves backward, the F clutch CF is disconnected and the R clutch CR is connected.

  The transmission mechanism 66 includes a transmission shaft 67, a first planetary gear mechanism 68, a second planetary gear mechanism 69, a Hi / Lo switching mechanism 70, and an output gear 71. The transmission shaft 67 is connected to the FR switching mechanism 65. The first planetary gear mechanism 68 and the second planetary gear mechanism 69 are arranged coaxially with the transmission shaft 67.

  The first planetary gear mechanism 68 includes a first sun gear S1, a plurality of first planetary gears P1, a first carrier C1 that supports the plurality of first planetary gears P1, and a first ring gear R1. . The first sun gear S1 is coupled to the transmission shaft 67. The plurality of first planetary gears P1 mesh with the first sun gear S1 and are rotatably supported by the first carrier C1. A first carrier gear Gc1 is provided on the outer periphery of the first carrier C1. The first ring gear R1 meshes with the plurality of planetary gears P1 and is rotatable. A first ring outer peripheral gear Gr1 is provided on the outer periphery of the first ring gear R1.

  The second planetary gear mechanism 69 includes a second sun gear S2, a plurality of second planetary gears P2, a second carrier C2 that supports the plurality of second planetary gears P2, and a second ring gear R2. . The second sun gear S2 is connected to the first carrier C1. The plurality of second planetary gears P2 mesh with the second sun gear S2 and are rotatably supported by the second carrier C2. The second ring gear R2 meshes with the plurality of planetary gears P2 and is rotatable. A second ring outer peripheral gear Gr2 is provided on the outer periphery of the second ring gear R2. The second ring outer peripheral gear Gr2 meshes with the output gear 71, and the rotation of the second ring gear R2 is output to the output shaft 63 via the output gear 71.

  The Hi / Lo switching mechanism 70 is a mechanism for switching the driving force transmission path in the power transmission device 24 between a high speed mode (Hi mode) where the vehicle speed is high and a low speed mode (Lo mode) where the vehicle speed is low. The Hi / Lo switching mechanism 70 has an H clutch CH connected in the Hi mode and an L clutch CL connected in the Lo mode. The H clutch CH connects or disconnects the first ring gear R1 and the second carrier C2. The L clutch CL connects or disconnects the second carrier C2 and the fixed end 72, and prohibits or allows the rotation of the second carrier C2.

  Each clutch CH, CL is a hydraulic clutch, and hydraulic oil from the transmission pump 29 is supplied to each clutch CH, CL. The hydraulic oil for the H clutch CH is controlled by the H clutch control valve VH. The hydraulic fluid to the L clutch CL is controlled by the L clutch control valve VL. Each clutch control valve VH, VL is controlled by a command signal from the control unit 27.

  The first motor MG1 and the second motor MG2 function as driving motors that generate driving force by electric energy. The first motor MG1 and the second motor MG2 also function as generators that generate electrical energy using the input driving force. When a command signal is given from control unit 27 so that torque in the direction opposite to the rotation direction acts on first motor MG1, first motor MG1 functions as a generator. A first motor gear Gm1 is fixed to the output shaft of the first motor MG1, and the first motor gear Gm1 meshes with the first carrier gear Gc1. A first inverter I1 is connected to the first motor MG1, and a command signal for controlling the motor torque of the first motor MG1 is given to the first inverter I1 from the control unit 27.

  The second motor MG2 has the same configuration as the first motor MG1. A second motor gear Gm2 is fixed to the output shaft of the second motor MG2, and the second motor gear Gm2 meshes with the first ring outer peripheral gear Gr1. Further, the second inverter I2 is connected to the second motor MG2, and a command signal for controlling the motor torque of the second motor MG2 is given to the second inverter I2 from the control unit 27.

  Capacitor 64 functions as an energy storage unit that stores energy generated in motors MG1 and MG2. That is, the capacitor 64 stores the electric power generated by the motors MG1 and MG2 when the total power generation amount of the motors MG1 and MG2 is large. Capacitor 64 discharges power when the total power consumption of motors MG1 and MG2 is large. In other words, each motor MG1, MG2 is driven by the electric power stored in capacitor 64. Alternatively, the motors MG1 and MG2 can be driven by the electric power stored in the capacitor 64. A battery may be used instead of the capacitor.

  The control unit 27 receives detection signals from various detection units and gives command signals indicating command torques to the motors MG1 and MG2 to the inverters I1 and I2. Note that the control unit 27 may output rotational speed commands for the motors MG1 and MG2. In this case, the inverters I1 and I2 calculate a command torque corresponding to the rotation speed command, and control the motors MG1 and MG2. Further, the control unit 27 gives a command signal for controlling the clutch hydraulic pressure of each clutch CF, CR, CH, CL to each clutch control valve VF, VR, VH, VL. Thereby, the gear ratio and output torque of the power transmission device 24 are controlled. Hereinafter, the operation of the power transmission device 24 will be described.

Here, the schematic operation of the power transmission device 24 when the vehicle speed is accelerated from 0 to the forward side while the rotation speed of the engine 21 is kept constant will be described with reference to FIG. FIG. 4 shows the rotational speeds of the motors MG1 and MG2 with respect to the vehicle speed. When the rotational speed of the engine 21 is constant, the vehicle speed changes according to the rotational speed ratio of the power transmission device 24. The rotation speed ratio is the ratio of the rotation speed of the output shaft 63 to the rotation speed of the input shaft 61. Therefore, in FIG. 4, the change in the vehicle speed coincides with the change in the rotational speed ratio of the power transmission device 24. That is, FIG. 4 shows the relationship between the rotational speeds of the motors MG1 and MG2 and the rotational speed ratio of the power transmission device 24. In FIG. 4, the solid line indicates the rotation speed of the first motor MG1, and the broken line indicates the rotation speed of the second motor MG2.
In a region where the vehicle speed is between 0 and V1, the L clutch CL is connected and the H clutch CH is disconnected (Lo mode). In this Lo mode, since the H clutch CH is disconnected, the second carrier C2 and the first ring gear R1 are disconnected. Further, since the L clutch CL is connected, the second carrier C2 is fixed.

  In the Lo mode, the driving force from the engine 21 is input to the first sun gear S1 via the transmission shaft 67, and this driving force is output from the first carrier C1 to the second sun gear S2. On the other hand, the driving force input to the first sun gear S1 is transmitted from the first planetary gear P1 to the first ring gear R1, and is output to the second motor MG2 via the first ring outer peripheral gear Gr1 and the second motor gear Gm2. . The second motor MG2 mainly functions as a generator in the Lo mode, and a part of the electric power generated by the second motor MG2 is stored in the capacitor 64. Further, part of the electric power generated by the second motor MG2 is consumed for driving the first motor MG1.

  In the Lo mode, the first motor MG1 mainly functions as an electric motor. The driving force of the first motor MG1 is output to the second sun gear S2 through the path of the first motor gear Gm1 → first carrier gear Gc1 → first carrier C1 →. The driving force output to the second sun gear S2 as described above is transmitted to the output shaft 63 through the path of the second planetary gear P2, the second ring gear R2, the second ring outer peripheral gear Gr2, and the output gear 71.

  In the region where the vehicle speed exceeds V1, the H clutch CH is connected and the L clutch CL is disconnected (Hi mode). In this Hi mode, since the H clutch CH is connected, the second carrier C2 and the first ring gear R1 are connected. Further, since the L clutch CL is disconnected, the second carrier C2 is disconnected. Accordingly, the rotation speeds of the first ring gear R1 and the second carrier C2 coincide.

  In the Hi mode, the driving force from the engine 21 is input to the first sun gear S1, and this driving force is output from the first carrier C1 to the second sun gear S2. The driving force input to the first sun gear S1 is output from the first carrier C1 to the first motor MG1 via the first carrier gear Gc1 and the first motor gear Gm1. In the Hi mode, the first motor MG1 mainly functions as a generator, so that part of the electric power generated by the first motor MG1 is stored in the capacitor 64. A part of the electric power generated by the first motor MG1 is consumed for driving the second motor MG2.

  Further, the driving force of the second motor MG2 is output to the second carrier C2 through the path of the second motor gear Gm2 → the first ring outer peripheral gear Gr1 → the first ring gear R1 → the H clutch CH. The driving force output to the second sun gear S2 as described above is output to the second ring gear R2 via the second planetary gear P2, and the driving force output to the second carrier C2 is the second planetary gear. Output to the second ring gear R2 via P2. The driving force combined by the second ring gear R2 in this way is transmitted to the output shaft 63 via the second ring outer peripheral gear Gr2 and the output gear 71.

  Although the above description is for forward drive, the same operation is performed for reverse drive. Further, at the time of braking, the roles of the first motor MG1 and the second motor MG2 as generators and motors are opposite to those described above.

  Next, control of the power transmission device 24 by the control unit 27 will be described. The control unit 27 controls the output torque of the power transmission device 24 by controlling the motor torque of the first motor MG1 and the second motor MG2. That is, the control unit 27 controls the traction force or the braking force of the work vehicle 1 by controlling the motor torque of the first motor MG1 and the second motor MG2.

  First, a method for determining a motor torque command value (hereinafter referred to as “command torque”) for the first motor MG1 and the second motor MG2 will be described.

  FIG. 5 is a control block diagram showing processing executed by the control unit 27. As shown in FIG. 5, the control unit 27 includes a transmission request determination unit 84, an energy management request determination unit 85, and a work implement request determination unit 86.

  The transmission request determination unit 84 determines the required traction force Tout based on the accelerator operation amount Aac and the output rotation speed Nout. Specifically, the transmission request determination unit 84 determines the required tractive force Tout from the output rotation speed Nout based on the required tractive force characteristic information D1 stored in the storage unit 56. The required tractive force characteristic information D1 is data indicating a required tractive force characteristic that defines the relationship between the output rotation speed Nout and the required tractive force Tout. Further, the required tractive force characteristic is changed according to the accelerator operation amount. The required traction force characteristic corresponds to a predetermined vehicle speed-traction force characteristic. The transmission request determination unit 84 determines the required tractive force Tout from the output rotational speed Nout using the required tractive force characteristic according to the accelerator operation amount, and determines the transmission required horsepower Htm from the product of the output rotational speed Nout and the required tractive force Tout. To do.

  Specifically, as shown in FIG. 6, the storage unit 56 stores data Lout1 indicating the required traction force characteristic that is a reference (hereinafter referred to as “reference traction force characteristic Lout1”). The reference traction force characteristic Lout1 is a required traction force characteristic when the accelerator operation amount Aac is the maximum value, that is, 100%. The reference traction force characteristic Lout1 is determined according to the speed range selected by the speed change operation member 53a. The transmission request determination unit 84 determines the current required tractive force characteristic Lout2 by multiplying the reference tractive force characteristic Lout1 by a predetermined ratio corresponding to the accelerator operation amount Aac.

  The required tractive force characteristic information D1 defines a required tractive force Tout that increases as the output rotational speed Nout decreases. Further, when the above-described shift operation member 53a is operated, the transmission request determination unit 84 changes the required tractive force characteristic in accordance with the speed range selected by the shift operation member 53a. For example, when a downshift is performed by the speed change operation member 53a, the required tractive force characteristic information is changed from Lout2 to Lout2 '. Thereby, the upper limit value of the output rotation speed Nout is reduced. That is, the upper limit value of the vehicle speed is reduced.

  Further, the required tractive force characteristic information D1 defines a negative required tractive force Tout with respect to an output rotation speed Nout equal to or higher than a predetermined speed. For this reason, when the output rotation speed Nout is larger than the upper limit value of the output rotation speed in the selected speed range, the required tractive force Tout is determined to be a negative value. When the required tractive force Tout is a negative value, a braking force is generated. As a result, the EMT type power transmission device 24 realizes the same behavior as the engine brake generated in the torque converter type transmission. Control during braking by the engine brake will be described later.

  The energy management request determination unit 85 shown in FIG. 5 determines the energy management request horsepower Hem based on the remaining amount of power in the capacitor 64. The energy management required horsepower Hem is the horsepower required for the power transmission device 24 to charge the capacitor 64. For example, the energy management request determination unit 85 determines the current capacitor charge amount from the voltage Vca of the capacitor 64. The energy management request determination unit 85 increases the energy management request horsepower Hem as the current capacitor charge amount decreases.

  The work implement request determining unit 86 determines the work implement required horsepower Hpto based on the work implement pump pressure Pwp and the operation amount Awo of the work implement operating member 52a (hereinafter referred to as “work implement operation amount Awo”). In the present embodiment, the work machine required horsepower Hpto is a horsepower distributed to the work machine pump 23. However, the work machine required horsepower Hpto may include horsepower distributed to the steering pump 30 and / or the transmission pump 29.

  Specifically, the work implement request determination unit 86 determines the required flow rate Qdm of the work implement pump 23 from the work implement operation amount Awo based on the request flow rate information D2. The required flow rate information D2 is stored in the storage unit 56, and defines the relationship between the required flow rate Qdm and the work implement operation amount Awo. The work implement request determining unit 86 determines the work implement required horsepower Hpto from the request flow rate Qdm and the work implement pump pressure Pwp.

  The control unit 27 includes a target output shaft torque determination unit 82, a target input shaft torque determination unit 81, and a command torque determination unit 83.

  The target output shaft torque determining unit 82 determines the target output shaft torque To_ref. The target output shaft torque To_ref is a target value of torque output from the power transmission device 24. The target output shaft torque determining unit 82 determines the target output shaft torque To_ref based on the required traction force Tout determined by the transmission request determining unit 84. Specifically, the target output shaft torque To_ref is determined by multiplying the required traction force Tout by a predetermined distribution rate. The predetermined distribution ratio is set so that, for example, the total of the work machine required horsepower Hpto, the transmission required horsepower Htm, and the energy management required horsepower Hem does not exceed the output horsepower from the engine 21.

  The target input shaft torque determining unit 81 determines the target input shaft torque Te_ref. The target input shaft torque Te_ref is a target value of torque input to the power transmission device 24. The target input shaft torque determining unit 81 determines the target input shaft torque Te_ref based on the transmission required horsepower Htm and the energy management required horsepower Hem. More specifically, the target input shaft torque determination unit 81 adds the value obtained by multiplying the transmission request horsepower Htm by a predetermined distribution ratio and the energy management request horsepower Hem and multiplies the engine rotation speed to obtain the target input shaft torque. Calculate Te_ref. The transmission required horsepower Htm is calculated by multiplying the above-described required traction force Tout by the current output rotational speed Nout.

  The command torque determining unit 83 determines command torques Tm1_ref and Tm2_ref for the motors MG1 and MG2 based on torque balance information from the target input shaft torque Te_ref and the target output shaft torque To_ref. The torque balance information defines the relationship between the target input shaft torque Te_ref and the target output shaft torque To_ref so as to satisfy the torque balance in the power transmission device 24. The torque balance information is stored in the storage unit 56.

As described above, the transmission path of the driving force in the power transmission device 24 differs between the Lo mode and the Hi mode. Therefore, the command torque determining unit 83 determines the command torques Tm1_ref and Tm2_ref to the motors MG1 and MG2 using different torque balance information in the Lo mode and the Hi mode. Specifically, the command torque determination unit 83 determines the command torques Tm1_Low and Tm2_Low for the motors MG1 and MG2 in the Lo mode using the first torque balance information shown in the following Equation 1. In the present embodiment, the first torque balance information is a formula for balance of torque in the power transmission device 24.
[Equation 1]
Ts1_Low = Te_ref * r_fr
Tc1_Low = Ts1_Low * (-1) * ((Zr1 / Zs1) + 1)
Tr2_Low = To_ref * (Zod / Zo)
Ts2_Low = Tr2_Low * (Zs2 / Zr2)
Tcp1_Low = Tc1_Low + Ts2_Low
Tm1_Low = Tcp1_Low * (-1) * (Zp1 / Zp1d)
Tr1_Low = Ts1_Low * (Zr1 / Zs1)
Tm2_Low = Tr1_Low * (-1) * (Zp2 / Zp2d)

Further, the command torque determining unit 83 determines command torques Tm1_Hi and Tm2_Hi for the motors MG1 and MG2 in the Hi mode using the second torque balance information expressed by the following formula 2. In the present embodiment, the second torque balance information is a formula of torque balance in the power transmission device 24.
[Equation 2]
Ts1_Hi = Te_ref * r_fr
Tc1_Hi = Ts1_Hi * (-1) * ((Zr1 / Zs1) + 1)
Tr2_Hi = To_ref * (Zod / Zo)
Ts2_Hi = Tr2_Hi * (Zs2 / Zr2)
Tcp1_Hi = Tc1_Hi + Ts2_Hi
Tm1_Hi = Tcp1_Hi * (-1) * (Zp1 / Zp1d)
Tr1_Hi = Ts1_Hi * (Zr1 / Zs1)
Tc2_Hi = Tr2_Hi * (-1) * ((Zs2 / Zr2) + 1)
Tcp2_Hi = Tr1_Hi + Tc2_Hi
Tm2_Hi = Tcp2_Hi * (-1) * (Zp2 / Zp2d)
Here, the parameters of each torque balance information are as shown in Table 1 below.

  Next, control of the engine 21 by the control unit 27 will be described. As described above, the control unit 27 controls the engine 21 by sending a command signal to the fuel injection device 28. Hereinafter, a method for determining the command throttle value for the fuel injection device 28 will be described. The control unit 27 includes an engine request determination unit 87 and a request throttle determination unit 89.

  The engine request determination unit 87 determines the engine request horsepower Hdm based on the work machine request horsepower Hpto, the transmission request horsepower Htm, and the energy management request horsepower Hem. Specifically, the engine request determination unit 87 determines the engine request horsepower Hdm by adding the work machine request horsepower Hpto, the transmission request horsepower Htm, and the energy management request horsepower Hem.

  The requested throttle determining unit 89 determines a command throttle value Th_cm from the engine required horsepower Hdm and the accelerator operation amount Aac. The requested throttle determining unit 89 uses the engine torque line Let and the matching line Lma stored in the storage unit 56 to determine the command throttle value Th_cm. The engine torque line Let defines the relationship between the output torque of the engine 21 and the engine rotational speed Ne. The matching line Lma is information for determining the first required throttle value from the engine required horsepower Hdm.

  The required throttle determining unit 89 determines the first required throttle value so that the engine torque line Let and the matching line Lma are matched at the matching point Pma1 where the output torque of the engine 21 is a torque corresponding to the engine required horsepower Hdm. To do. The requested throttle determining unit 89 determines the smaller one of the first requested throttle value and the second requested throttle value corresponding to the accelerator operation amount Aac as the command throttle value Th_cm.

  Next, control during braking by engine braking will be described. FIG. 7 is a schematic diagram showing a flow of braking power absorbed during braking. As shown in FIG. 7, a part of the braking power absorbed by the traveling device 25 is stored in the capacitor 64 as electric energy by driving the first motor MG1 and / or the second motor MG2. A part of the braking power is distributed to the work implement pump 23, the steering pump 30, and the transmission pump 29 via the PTO 22.

  FIG. 8 is a control block diagram showing processing executed by the control unit 27 during braking. As shown in FIG. 8, the control unit 27 includes a pump brake control determination unit 91 and a pump brake torque control unit 92. The pump brake control determination unit 91 determines execution of pump brake control for generating a braking force by the load of the work implement pump 23 during braking. The pump brake torque control unit 92 increases the pump brake torque by increasing the load of the work implement pump 23 in the pump brake control. The pump brake torque corresponds to the load of the work implement pump 23.

  As described above, the braking force is generated when the required tractive force Tout is a negative value. However, in this embodiment, the increase or decrease or magnitude relationship of the braking force or the brake torque depends on the braking force or the brake torque. It means an increase or decrease in absolute value or a magnitude relationship. The same applies to other parameters in braking control such as engine regeneration torque described later.

  FIG. 9 is a schematic diagram showing a hydraulic circuit connected to the work implement pump 23. As shown in FIG. As shown in FIG. 9, the work implement control valve 41 described above includes a boom control valve 41a and a bucket control valve 41b. The boom control valve 41a controls the hydraulic oil supplied to the lift cylinder 13. The bucket control valve 41b controls the hydraulic oil supplied to the bucket cylinder 14.

  The hydraulic circuit connected to the work implement pump 23 has a pump brake control valve 47 and a relief valve 48. The work machine pump 23 is connected to a relief valve 48 via a pump brake control valve 47. The relief valve 48 is provided in parallel with the lift cylinder 13 and the bucket cylinder 14 in the hydraulic circuit. The pump brake control valve 47 controls the hydraulic oil supplied to the relief valve 48. The pump brake control valve 47 is an electromagnetic control valve, and controls hydraulic oil supplied to the relief valve 48 based on a command signal input from the pump brake torque control unit 92. The pump brake torque control unit 92 controls the pump brake control valve 47 to increase the load on the work implement pump 23.

  The first capacity control device 42 includes a load sensing valve 46 (hereinafter referred to as “LS valve 46”). The LS valve 46 is configured so that the differential pressure between the discharge pressure of the work implement pump 23 and the outlet hydraulic pressure of the boom control valve 41a, the bucket control valve 41b, and the pump brake control valve 47 becomes a predetermined value. The discharge flow rate is controlled. Specifically, the maximum outlet hydraulic pressure (hereinafter referred to as “LS pressure”) among the outlet hydraulic pressure of the boom control valve 41a, the outlet hydraulic pressure of the bucket control valve 41b, and the outlet hydraulic pressure of the pump brake control valve 47 is the LS valve 46. Is input. The LS valve 46 controls the discharge flow rate of the work implement pump 23 so that the differential pressure between the discharge pressure of the work implement pump 23 and the LS pressure becomes a predetermined value. The boom control valve 41a, the bucket control valve 41b, and the pump brake control valve 47 are each provided with a pressure compensation valve (not shown) on the inlet side. The pressure compensation valve generates a pressure difference corresponding to the differential pressure between the LS pressure and each outlet pressure. In FIG. 9, broken lines connected to the left side of the boom control valve 41a, the bucket control valve 41b, and the pump brake control valve 47 indicate that the LS pressure is input to the control valves 41a, 41b, 47 for pressure compensation. Indicates that The control unit 27 controls the boom control valve 41a, the bucket control valve 41b, and the pump brake control valve 47, whereby the discharge flow rate of the work implement pump 23 is controlled according to the command signal from the control unit 27.

  FIG. 10 is a flowchart showing a determination process of execution of pump brake control by the pump brake control determination unit 91. First, in step S101, it is determined whether or not the engine rotational speed Ne is equal to or higher than a predetermined rotational speed threshold Ne_th. When the engine rotation speed Ne is equal to or higher than the predetermined rotation speed threshold value Ne_th, the process proceeds to step S102.

In step S102, it is determined whether or not the engine regenerative torque Te_ref is equal to or greater than a predetermined regenerative torque threshold Tth1. The engine regenerative torque Te_ref corresponds to the target input shaft torque Te_ref described above, and is a torque regenerated from the traveling device 25 to the engine 21 via the power transmission device 24 during braking.
[Equation 3]
Te_ref = (Htm-Hem) / Nout
Htm is the transmission required horsepower described above. When the required tractive force Tout is determined to be a negative value, the required tractive force Tout corresponds to a target braking force that is a target value of the braking force absorbed by the output shaft 63 of the power transmission device 24 during braking by the engine brake. In this case, Htm corresponds to a target braking power that is a target value of the braking power absorbed by the output shaft 63 of the power transmission device 24 during braking. Therefore, the work implement request determining unit 86 functions as a target braking power determining unit that determines the target braking power Htm during braking by engine braking. In the above equation 3, Htm may be multiplied by a predetermined efficiency.

  Hem is the horsepower required for energy management described above, and corresponds to the charging power in the capacitor 64. Therefore, at the time of braking by engine brake, the energy management request determination unit 85 functions as a storage power calculation unit that calculates storage power.

  When the engine regenerative torque Te_ref is equal to or greater than the predetermined regenerative torque threshold Tth1 in step S102, it is determined in step S103 that the pump brake control is to be executed.

  When the engine regeneration torque Te_ref is not equal to or greater than the predetermined regeneration torque threshold Tth1 in step S102, the process proceeds to step S104. In step S104, it is determined whether or not the engine output torque Te is equal to or less than a predetermined output torque threshold value Tth2. The engine output torque Te may be an estimated value or a command value. For example, the engine output torque Te may be calculated from the command throttle value Th_cm to the engine. When the engine output torque Te is less than or equal to the predetermined output torque threshold value Tth2, it is determined in step S103 that the pump brake control is to be executed.

  When the engine rotation speed Ne is not equal to or higher than the predetermined rotation speed threshold value Ne_th in step S101, it is determined in step S105 that the pump brake control is not executed. Further, when the engine output torque Te is not less than or equal to the predetermined output torque threshold Tth2 in step S104, it is determined in step S105 that the pump brake control is not executed.

  As described above, when the engine rotational speed Ne is equal to or higher than the predetermined rotational speed threshold Ne_th and the engine regenerative torque Te_ref is equal to or higher than the predetermined regenerative torque threshold Tth1, the pump brake control determination unit 91 performs the pump brake control. Is determined to be executed. The pump brake control determination unit 91 also executes pump brake control when the engine rotational speed Ne is equal to or higher than the predetermined rotational speed threshold Ne_th and the engine output torque Te is equal to or lower than the predetermined output torque threshold Tth2. Judge that.

  When executing the pump brake control, the pump brake torque control unit 92 controls the pump brake control valve 47 to increase the load on the work implement pump 23. Further, when the pump brake control is not executed, the pump brake torque control unit 92 does not increase the load of the work implement pump 23. That is, the pump brake torque Tpto_ref described later is set to zero. Next, a method for controlling the pump brake control valve 47 will be described in detail. As shown in FIG. 8, the pump brake torque control unit 92 includes a pump brake torque determination unit 93 and a pump brake valve command unit 94.

  The pump brake torque determining unit 93 determines the pump brake torque Tpto_ref. The pump brake torque Tpto_ref is the pump brake torque converted to the output shaft of the engine 21, and is generated by the work machine pump 23, the steering pump 30, the transmission pump 29, and other auxiliary equipment (not shown) during braking by the pump brake control. It is the total value of the load torque to be caused. FIG. 11 is a control block diagram illustrating a process of determining the pump brake torque Tpto_ref by the pump brake torque determining unit 93. As shown in FIG. 11, the pump brake torque control unit 92 includes a first pump brake torque calculation unit 95, a second pump brake torque calculation unit 96, a third pump brake torque calculation unit 97, and a vehicle speed limit brake torque calculation. A section 98 and a maximum value selection section 99.

The first pump brake torque calculation unit 95 calculates the first pump brake torque Tpto1 based on the engine regeneration torque Te_ref. Specifically, the first pump brake torque calculating unit 95 calculates the first pump brake torque Tpto1 by the following equation (4).
[Equation 4]
Tpto1 = (Te_ref-Te_loss) * k1
Te_loss is an engine loss and corresponds to the braking power that can be absorbed by the engine 21. The engine loss Te_loss may be a fixed value. Alternatively, the engine loss Te_loss may be determined from the engine rotational speed Ne by a table or a mathematical expression. k1 is a predetermined coefficient, which is larger than 0 and smaller than 1. An upper limit value and a lower limit value may be set for the first pump brake torque Tpto1.

Ne_targetは、ポンプブレーキ制御中における目標エンジン回転速度である。目標エンジン回転速度Ne_targetは、固定値であってもよい。或いは、目標エンジン回転速度Ne_targetは、車速からテーブル或いは数式等により決定されてもよい。kpは、PI制御で用いるPゲインである。kiは、PI制御で用いるIゲインである。なお、第2ポンプブレーキトルクTpto2に上限値と下限値とが設定されてもよい。積分項∫(ΔNe)dtに上限値及び／又は下限値が設定されてもよい。積分は、ポンプブレーキ制御開始時にリセットし、ゼロから行うことが好ましい。 The second pump brake torque calculation unit 96 calculates the second pump brake torque Tpto2 based on the engine rotation speed Ne. Specifically, the second pump brake torque calculating unit 96 calculates the second pump brake torque Tpto2 by the following equation (5).
[Equation 5]

Ne_target is a target engine rotation speed during pump brake control. The target engine rotation speed Ne_target may be a fixed value. Alternatively, the target engine rotation speed Ne_target may be determined from the vehicle speed by a table or a mathematical expression. kp is a P gain used in PI control. ki is an I gain used in PI control. An upper limit value and a lower limit value may be set for the second pump brake torque Tpto2. An upper limit value and / or a lower limit value may be set for the integral term ∫ (ΔNe) dt. It is preferable that the integration is reset at the start of pump brake control and is performed from zero.

  When the determination by the pump brake control determining unit 91 is “true”, that is, when it is determined that the pump brake control is to be executed, the third pump brake torque calculating unit 97 performs the first pump brake torque Tpto1 and the second pump brake torque The third pump brake torque Tpto3 is calculated by adding Tpto2. Accordingly, the pump brake torque control unit 92 determines the pump brake torque based on the engine regenerative torque, and also determines the pump brake torque by feedback control so that the engine rotation speed Ne becomes the target engine rotation speed Ne_target.

  The third pump brake torque calculation unit 97 sets the third pump brake torque Tpto3 to 0 when the determination by the pump brake control determination unit 91 is “false”, that is, when it is determined not to execute the pump brake control.

  The vehicle speed limit brake torque calculation unit 98 calculates a vehicle speed limit brake torque Tpto_limit based on the vehicle speed. Specifically, the vehicle speed limit brake torque calculation unit 98 refers to the vehicle speed limit brake torque information and determines the vehicle speed limit brake torque Tpto_limit from the output rotation speed Nout. For example, the vehicle speed limit brake torque information is a table that defines the relationship between the output rotation speed Nout and the vehicle speed limit brake torque Tpto_limit.

  FIG. 12 is a graph showing the relationship between the output rotation speed Nout defined by the vehicle speed limit brake torque information and the vehicle speed limit brake torque Tpto_limit. As shown in FIG. 12, in the vehicle speed limit brake torque information, the vehicle speed limit brake torque Tpto_limit is 0 when the output rotation speed Nout is 0 or more and less than the predetermined speed threshold value Nout_th. When the output rotation speed Nout is equal to or higher than a predetermined speed threshold value Nout_th, the vehicle speed limit brake torque Tpto_limit becomes a predetermined value Ta.

  The maximum value selection unit 99 determines the larger one of the third pump brake torque Tpto3 and the vehicle speed limit brake torque Tpto_limit as the pump brake torque Tpto_ref. Therefore, when it is determined to execute the pump brake control and the output rotation speed Nout is smaller than the predetermined speed threshold Nout_th, the third pump brake torque Tpto3 becomes the pump brake torque Tpto_ref. In this case, the pump brake torque control unit 92 determines the pump brake torque Tpto_ref based on the engine regeneration torque Te_ref and controls the pump brake torque Tpto_ref so that the engine rotation speed Ne becomes the target engine rotation speed Ne_target.

  When it is determined to execute the pump brake control, the output rotation speed Nout is equal to or higher than the predetermined speed threshold Nout_th, and the vehicle speed limit brake torque Tpto_limit is greater than the third pump brake torque Tpto3, the vehicle speed limit brake torque Tpto_limit is equal to the pump brake torque Tpto_ref. Become. Therefore, the pump brake torque control unit 92 increases the pump brake torque Tpto_ref when the output rotation speed Nout becomes equal to or higher than the speed threshold value Nout_th. That is, the pump brake torque control unit 92 increases the pump brake torque Tpto_ref when the vehicle speed becomes equal to or higher than a predetermined vehicle speed threshold corresponding to the speed threshold Nout_th. Accordingly, as shown in FIG. 6, the pump brake torque Tpto_ref can be increased in accordance with the increase in the target braking force Tout when the output rotation speed Nout is equal to or higher than the predetermined speed threshold value Nout_th. Thereby, an excessive increase in the engine speed can be suppressed.

  When it is determined not to execute the pump brake control, the third pump brake torque Tpto3 is zero. Further, when the output rotation speed Nout is smaller than the predetermined speed threshold value Nout_th, the vehicle speed limit brake torque Tpto_limit is also zero. Therefore, the pump brake torque Tpto_ref is 0, and the pump brake torque control unit 92 does not generate the pump brake torque.

  However, even if it is determined not to execute the pump brake control, the pump brake torque Tpto_ref becomes the predetermined value Ta when the output rotation speed Nout is equal to or higher than the predetermined speed threshold Nout_th. Therefore, even when it is determined not to execute the pump brake control, when the output rotational speed Nout is high, an excessive increase in the engine rotational speed can be suppressed by generating the pump brake torque.

  The pump brake valve command unit 94 shown in FIG. 8 determines a command value PTOB_EPC for the pump brake control valve 47 based on the pump brake torque Tpto_ref. As described above, the pump brake control valve 47 is an electromagnetic control valve, and the command value PTOB_EPC to the pump brake control valve 47 is a command current value. FIG. 13 is a control block diagram showing a process of determining the command value PTOB_EPC for the pump brake control valve 47 by the pump brake valve command unit 94. As shown in FIG. 13, the pump brake valve command unit 94 includes a necessary pump flow rate calculation unit 101, a pump brake valve flow rate calculation unit 102, and a pump brake valve command value calculation unit 103.

The necessary pump flow rate calculation unit 101 determines the required flow rate Q_Lo_ref of the work implement pump 23 based on the pump brake torque Tpto_ref. Specifically, the necessary pump flow rate calculation unit 101 determines the required flow rate Q_Lo_ref of the work implement pump 23 using the following equation (6).
[Equation 6]
Q_Lo_ref = ((Tpto_ref-Tpto_fix) / Pwp) * Ne
Tpto_fix is a fixed load in terms of the output shaft of the engine 21 and is a load that does not perform adjustment for controlling the braking force. For example, Tpto_fix is the sum of the load torque of the steering pump 30, the load torque of the transmission pump 29, and the load torque of other auxiliary equipment (not shown). Pwp is the work implement pump pressure described above.

The pump brake valve flow rate calculation unit 102 calculates the pump brake valve flow rate PTOB_Q_ref. The pump brake valve flow rate PTOB_Q_ref is a flow rate at the pump brake control valve 47. Specifically, the pump brake valve flow rate calculation unit 102 calculates the pump brake valve flow rate PTOB_Q_ref using the following equation (7).
[Equation 7]
PTOB_Q_ref = Q_Lo_ref-Qdm
Qdm is the required flow rate of the work implement pump 23 described above. That is, the pump brake torque control unit 92 generates a pump based on the required flow rate Q_Lo_ref of the work implement pump 23 for obtaining the pump brake torque Tpto_ref and the required flow rate Qdm of the work implement pump 23 for driving the work implement 3. The brake valve flow rate PTOB_Q_ref is determined.

  The pump brake valve command value calculation unit 103 determines the pump brake valve command value PTOB_EPC based on the pump brake valve flow rate PTOB_Q_ref. For example, the pump brake valve command value calculation unit 103 refers to a table that defines the relationship between the pump brake valve flow rate PTOB_Q_ref and the pump brake valve command value PTOB_EPC, and determines the pump brake valve command value PTOB_EPC from the pump brake valve flow rate PTOB_Q_ref. To do.

  When the pump brake valve command value PTOB_EPC is output from the pump brake torque control unit 92 to the pump brake control valve 47, the first capacity control device 42 increases the displacement of the work implement pump 23 by the action of the LS valve 46. As a result, the torque generated by the work implement pump 23 increases, and the pump brake torque can be increased. In the hydraulic circuit shown in FIG. 9, the discharge pressure of the work implement pump 23 is equal to or higher than the LS pressure even when the work implement control valve 41 operates simultaneously with the pump brake valve 47 by the action of the LS valve 46 and a pressure compensation valve (not shown). The required pump brake torque can be generated.

  The work vehicle according to the present embodiment has the following features.

  When the pump brake control determination unit 91 determines that the pump brake control is to be executed during braking, the pump brake torque control unit 92 increases the pump brake torque by increasing the load of the work implement pump 23. Therefore, by increasing the braking power distributed to the work implement pump 23 in the PTO 22, a large braking force can be obtained while suppressing an excessive increase in the engine rotation speed during braking.

  The pump brake control determination unit 91 determines to execute the pump brake control when the engine regenerative torque is equal to or greater than a predetermined regenerative torque threshold Tth1. Accordingly, when an engine regeneration torque larger than the torque that can be absorbed by the engine 21 is generated, an excessive increase in the engine rotation speed can be suppressed by increasing the pump brake torque.

  The pump brake control determination unit 91 determines to execute the pump brake control when the engine rotation speed is equal to or higher than a predetermined rotation speed threshold value Ne_Th. For this reason, an excessive increase in engine speed can be suppressed.

  The pump brake torque control unit 92 determines the pump brake torque based on the engine regeneration torque. For this reason, the pump brake torque can be appropriately controlled according to the magnitude of the engine regeneration torque.

  The pump brake torque control unit 92 determines the engine regeneration torque by subtracting the stored power from the target braking power. Since the target braking power is defined by the required tractive force characteristic information D1, the braking force can be appropriately controlled based on the required tractive force characteristic information D1. Further, the power charged in the capacitor 64 can be secured by subtracting the stored power.

  The pump brake torque control unit 92 determines the pump brake torque so that the engine rotation speed becomes the target engine rotation speed during the pump brake control. For this reason, an excessive increase in engine speed can be suppressed.

  The pump brake torque control unit 92 increases the pump brake torque by determining the vehicle speed limit brake torque to a predetermined value Ta when the vehicle speed becomes equal to or higher than a predetermined vehicle speed threshold. For this reason, the braking force can be increased when the vehicle speed becomes equal to or higher than a predetermined vehicle speed threshold. Thereby, an excessive increase in the vehicle speed can be prevented.

  The pump brake torque control unit 92 controls the pump brake control valve 47 to increase the load on the work implement pump 23. Therefore, the load increment of the work machine pump 23 is discarded as the heat of the hydraulic oil in the relief valve 48. For this reason, it is possible to increase the load on the work implement pump 23 while suppressing the influence on the operation of the work implement 3.

  The pump brake torque control unit 92 determines the pump brake valve flow rate PTOB_Q_ref based on the required flow rate Q_Lo_ref of the work implement pump 23 and the required flow rate Qdm of the work implement pump 23. For this reason, the required flow rate Qdm required for the operation of the work machine 3 can be secured. As a result, the braking force can be increased while supplying the necessary working oil to the work implement 3. Moreover, in order to obtain the required flow rate Q_Lo_ref of the work machine pump 23, the flow rate at the pump brake control valve 47 can be suppressed by setting the pump flow rate PTOB_Q_ref as the shortage of the required flow rate Qdm. Thereby, the temperature rise of hydraulic fluid can be suppressed. As a result, the braking force due to the load of the work implement pump 23 can be increased for a long time.

  The present invention is not limited to the above-described embodiments, and various changes or modifications can be made without departing from the scope of the present invention.

  The present invention is not limited to the wheel loader described above, and may be applied to other types of work vehicles such as a bulldozer, a tractor, a forklift, or a motor grader.

  The present invention is not limited to EMT and may be applied to other types of transmissions such as HMT. In this case, the first motor MG1 functions as a hydraulic motor and a hydraulic pump. The second motor MG2 functions as a hydraulic motor and a hydraulic pump. The first motor MG1 and the second motor MG2 are variable displacement pumps / motors, and the displacement is controlled by the control unit 27 controlling the tilt angle of the swash plate or the oblique shaft. Then, the capacities of the first motor MG1 and the second motor MG2 are controlled so that the command torques Tm1_ref and Tm2_ref calculated in the same manner as in the above embodiment are output.

  The configuration of the power transmission device 24 is not limited to the configuration of the above embodiment. For example, the connection and arrangement of the elements of the two planetary gear mechanisms 68 and 69 are not limited to the connection and arrangement of the above embodiment. Further, the number of planetary gear mechanisms provided in the power transmission device 24 is not limited to two. The power transmission device 24 may have only one planetary gear mechanism. Alternatively, the power transmission device 24 may have three or more planetary gear mechanisms.

  The control of the power transmission device 24 is not limited to the control in the above embodiment. In other words, in the above embodiment, the target input shaft torque Te_ref and the target output shaft torque To_ref are determined so that a predetermined vehicle speed-traction force characteristic in which the traction force continuously changes according to the vehicle speed is obtained. However, the target input shaft torque Te_ref and the target output shaft torque To_ref can be arbitrarily set.

  The torque balance information is not limited to the torque balance equation as in the above embodiment. For example, the torque balance information may be in the form of a table or a map.

  In the above embodiment, the pump brake torque is generated by increasing the load of the work machine pump 23. However, the pump brake torque may be generated by increasing the load of the hydraulic pump other than the work implement pump 23. Although the pump brake control operation has been described with reference to FIG. 9, the present invention is not limited to this method. That is, in a hydraulic circuit having a hydraulic pump connected to the engine shaft or PTO and a relief valve that relieves hydraulic oil discharged from the hydraulic pump, whichever of the flow rate discharged from the hydraulic pump and the relief pressure of the relief valve Or both. For example, the hydraulic pump may be a fixed displacement pump and the relief valve may be a variable relief valve. Alternatively, the hydraulic pump may be a variable displacement pump, and the relief valve may be a variable relief valve or a fixed relief valve.

  For example, FIG. 14 is a schematic diagram showing a part of a hydraulic circuit provided in the work vehicle according to the first modification. As shown in FIG. 14, the work vehicle according to the first modification includes a radiator 36, a cooling fan 38, a fan motor 39, and a fan pump 40. Cooling water for the engine 21 flows through the radiator 36. The cooling fan 38 cools the cooling water in the radiator 36. The fan motor 39 is a hydraulic motor and drives the cooling fan 38. The fan pump 40 is a hydraulic pump and discharges hydraulic oil for driving the fan motor 39. The fan pump 40 is connected to the engine 21 via the PTO 22, similarly to the work machine pump 23 described above. The fan pump 40 is a variable capacity pump, and a third capacity control device 49 is connected to the fan pump 40. The third capacity control device 49 is controlled by the control unit 27 and changes the tilt angle of the fan pump 40. Thus, the discharge capacity of the fan pump 40 is controlled by the control unit 27. Other configurations of the work vehicle according to the first modification are the same as those of the work vehicle 1 according to the above-described embodiment.

  In the pump brake control, the pump brake torque control unit 92 increases the load of the fan pump 40 by increasing the discharge capacity of the fan pump 40 and increasing the rotational speed of the fan motor 39. In this case, the pump brake torque can be increased by increasing the load of the fan pump 40.

  FIG. 15 is a schematic diagram illustrating a configuration of a work vehicle according to a second modification. As shown in FIG. 15, the work vehicle includes a warm-up hydraulic circuit 59. The warm-up hydraulic circuit 59 is connected to the transmission pump 29 described above. The warm-up hydraulic circuit 59 has, for example, a warm-up relief valve, and the control unit 27 increases the discharge pressure of the pump 29 by controlling the opening degree of the warm-up relief valve. Increase the temperature of the hydraulic fluid that passes through the relief valve. Thereby, the warm-up operation by the warm-up hydraulic circuit 59 is executed. Other configurations of the work vehicle according to the second modification are the same as those of the work vehicle 1 according to the above-described embodiment.

  In the pump brake control, the pump brake torque control unit 92 increases the load on the transmission pump 29 by executing a warm-up operation by the warm-up hydraulic circuit 59. In this case, the pump brake torque can be increased by increasing the load of the transmission pump 29. The hydraulic pump connected to the warm-up hydraulic circuit 59 is not limited to the transmission pump 29, and may be another hydraulic pump.

  The pump brake torque control unit 92 may determine a predetermined vehicle speed threshold based on the speed range selected by the speed change operation member 53a. For example, not only the vehicle speed threshold corresponding to the highest speed range but also the vehicle speed threshold corresponding to a low speed range such as the first speed or the second speed may be set. In this case, the braking force can be increased when the vehicle speed becomes equal to or higher than the vehicle speed threshold corresponding to the speed range selected by the speed change operation member 53a. Thus, the braking force can be increased when the vehicle speed exceeds the speed range selected by the speed change operation member 53a.

  The pump brake torque control unit 92 may determine a predetermined vehicle speed threshold based on the selection by the FR operation member 54a. That is, the vehicle speed threshold value for forward travel and the vehicle speed threshold value for reverse travel may be set to different values. In this case, the braking force can be increased when the vehicle speed becomes equal to or higher than the vehicle speed threshold value corresponding to the traveling direction selected by the FR operation member 54a. Note that the vehicle speed threshold value for each speed range during forward movement and the vehicle speed threshold value for each speed range during backward travel may be set to different values.

  According to the present invention, it is possible to provide a hybrid work vehicle that can obtain a large braking force while suppressing an excessive increase in engine rotation speed during braking, and a control method therefor.

21 ... Engine, 25 ... Running device 25, 24 ... Power transmission device 24, 61 ... Input shaft, 63 ... Output shaft, 68 ... First planetary gear mechanism, 69 ... Second planetary gear mechanism, 62 ... Gear mechanism, MG1 ... 1st motor, MG2 ... 2nd motor, 27 ... control unit, 91 ... pump brake control determination unit, 92 ... pump brake torque control unit, 53a ... speed change operation member, 54a ... forward / reverse operation member, 3 ... work machine, 23 ... Work machine pump, 48 ... Relief valve, 47 ... Pump brake control valve, 52a ... Work machine operation member, 101 ... Required pump flow rate determining unit, 86 ... Work machine request determining unit, 38 ... Cooling fan, 39 ... Fan motor, 40… fan pump, 59… warm-up hydraulic circuit, 29… transmission pump

Claims (3)

Engine,
A hydraulic pump driven by the engine;
A traveling device driven by the engine;
A power transmission device for transmitting driving force from the engine to the traveling device;
A power take-out device that distributes the driving force from the engine to the hydraulic pump and the power transmission device;
A control unit for controlling the hydraulic pump and the power transmission device;
With
The power transmission device is
An input shaft;
An output shaft;
A gear mechanism having a planetary gear mechanism and transmitting the rotation of the input shaft to the output shaft;
A motor connected to the rotating element of the planetary gear mechanism;
Have
The power transmission device is configured to change a rotation speed ratio of the output shaft to the input shaft by changing a rotation speed of the motor.
The control unit includes a pump brake torque control unit that increases a pump brake torque corresponding to the load of the hydraulic pump in pump brake control in which a braking force is generated by the load of the hydraulic pump.
The pump brake torque control unit is
Determining the target braking power, which is the power to be absorbed by the work vehicle,
Obtaining engine regeneration power regenerated by the engine,
When the target braking power is larger than the engine regenerative power, a work vehicle that variably consumes a power value equal to a difference between the target braking power and the engine regenerative power by controlling the hydraulic pump. Engine,
A hydraulic pump driven by the engine;
A traveling device driven by the engine;
A power transmission device for transmitting driving force from the engine to the traveling device;
A power take-out device that distributes the driving force from the engine to the hydraulic pump and the power transmission device;
A method for controlling a work vehicle comprising:
The power transmission device is
An input shaft;
An output shaft;
A gear mechanism having a planetary gear mechanism and transmitting the rotation of the input shaft to the output shaft;
A motor connected to the rotating element of the planetary gear mechanism;
Have
The power transmission device is configured to change a rotation speed ratio of the output shaft to the input shaft by changing a rotation speed of the motor.
Obtaining a target braking power which is a power to be absorbed by the work vehicle;
Obtaining engine regenerative power regenerated by the engine;
When the target braking power is greater than the engine regenerative power, the hydraulic pump is controlled to variably consume a power value equal to the difference between the target braking power and the maximum power consumed by the work vehicle. When,
A method for controlling a work vehicle. Engine,
A hydraulic pump driven by the engine;
A traveling device driven by the engine;
A power transmission device for transmitting driving force from the engine to the traveling device;
A power take-out device that distributes the driving force from the engine to the hydraulic pump and the power transmission device;
A method for controlling a work vehicle comprising:
The power transmission device is
An input shaft;
An output shaft;
A gear mechanism having a planetary gear mechanism and transmitting the rotation of the input shaft to the output shaft;
A motor connected to the rotating element of the planetary gear mechanism;
Have
The power transmission device is configured to change a rotation speed ratio of the output shaft to the input shaft by changing a rotation speed of the motor.
Obtaining power consumption of the work vehicle;
When the power to be consumed by the work vehicle is larger than the consumed power, the hydraulic pump is controlled to variably consume the consumed power exceeding the consumed power of the work vehicle;
A method for controlling a work vehicle.

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