JP4764018B2 - Traveling work machine - Google Patents

Description

  The present invention relates to a traveling work machine such as a wheel loader, and more particularly to a traveling work machine that includes a traveling body and a work machine and that may perform traveling and work simultaneously.

  2. Description of the Related Art Conventionally, construction machines such as a wheel loader and a hydraulic excavator generally have a configuration in which a working machine such as a boom and a bucket, and a traveling body such as a wheel and a crawler are operated using an engine as a drive source.

  However, in order to cope with environmental pollution and oil resource depletion in recent years, vehicles that use an engine and an electric motor together as a drive source or replace the engine with an electric motor are not only used in the field of general automobiles but also in construction machinery. It is also being developed in the field of traveling work machines.

In Patent Document 1 to be described later, an electric motor is attached to each drive shaft such as a boom, an arm, a bucket, a traveling body, and an upper swing body of a hydraulic excavator, and these boom, arm, bucket, traveling The invention in which the drive shafts of the body and the upper swing body are operated is described.
JP 2001-3398 A

  As described above, when an electric motor is individually provided for each work machine and for each drive shaft of the traveling body and configured to operate individually, a capacity capable of outputting horsepower beyond the required horsepower of the drive shaft. The electric motor must be prepared for each drive shaft. For this reason, the sum total of the capacities of individual electric motors (total horsepower) must be larger than the engine horsepower in the case of the conventional configuration in which each drive shaft is driven by one engine. This increases the size of the electric motor, increases the cost of the apparatus, and causes a problem that energy efficiency decreases.

  On the other hand, as in the conventional hydraulic system, the engine driving force is distributed to each driving shaft via a hydraulic pump, and the engine is replaced with an electric motor so that the electric motor driving force is It is also conceivable to arrange to distribute to the drive shaft.

  However, when the driving force of the electric motor is distributed to each drive shaft in this way, the configuration and control content are complicated compared to the conventional configuration and control content distributed to each drive shaft by the hydraulic system, The equipment cost increases.

  The present invention has been made in view of such a situation, and when operating a traveling body and a working machine of a traveling work machine using an electric motor as a driving source, the total capacity of the electric motor is reduced and driving of the electric motor is performed. The problem to be solved is to reduce the apparatus cost and improve the energy efficiency by simplifying the configuration and control contents for distributing the force to the drive shafts.

The first invention is
In the traveling work machine including the traveling body (13) and the work machines (18, 19),
A first electric motor (5);
A second electric motor (6);
The drive torques of the first and second electric motors (5, 6) are input, and the drive torques of the first and second electric motors (5, 6) work with the drive shaft (11a) of the traveling body (13). And a torque distribution transmission mechanism (10) distributed and transmitted to the drive shaft (12a) of the machine (18, 19).

The second invention is the first invention,
The traveling work machine is a wheel loader (1),
The torque distribution transmission mechanism (10) transmits the entire driving torque of the first electric motor (5) to the driving shaft (11a) of the traveling body (13), and the driving torque of the second electric motor (6). Is configured to be transmitted to the drive shaft (11a) of the traveling body (13).

The third invention is the first invention,
The torque distribution / transmission mechanism (10) is a differential gear mechanism.

A fourth invention is the first invention,
The torque distribution / transmission mechanism (10) is a planetary gear mechanism.

A fifth invention is the first invention,
The torque distribution transmission mechanism (10) is constituted by a differential gear mechanism,
The output shaft (6a) of the second electric motor (6) is connected to the second input shaft (10b) of the torque distribution transmission mechanism (10),
The second output shaft (10d) of the torque distribution transmission mechanism (10) is connected to the drive shaft (12a) of the work machine (18, 19),
A clutch (61) for engaging and disengaging the second input shaft (10b) and the second output shaft (10d) is provided.

A sixth invention is the first invention,
The torque distribution transmission mechanism (10) is composed of a planetary gear mechanism,
The output shaft (6a) of the second electric motor (6) is connected to the second input shaft (10b) of the torque distribution transmission mechanism (10),
The second output shaft (10d) of the torque distribution transmission mechanism (10) is connected to the drive shaft (12a) of the work machine (18, 19),
A clutch (61) that engages and disengages the second input shaft (10b) and the second output shaft (10d) is provided.
A clutch (62) for connecting and releasing the connection between the second input shaft (10b) and the planetary gear (110) is provided.

A seventh invention is the sixth invention,
A brake (63) for braking the ring gear (10R) of the planetary gear (110) is provided.

  According to the first invention, as shown in FIG. 1, the drive torques of the first and second electric motors 5 and 6 are input to the torque distribution transmission mechanism, and the first and second electric motors 5 and 6 are driven. Torque is distributed and transmitted to the drive shaft 11a of the traveling body 13 and the drive shaft 12a of the work implement.

  Thereby, as shown in FIG. 11, the sum total of the capacity | capacitance of the 1st and 2nd electric motors 5 and 6 can be made small. Further, as shown in FIG. 2, a torque distribution transmission mechanism 10 and a controller 20 including an electric motor 5, 6, an inverter 3, 4, a battery 2, a planetary gear mechanism, and the like are added to an existing in-vehicle system. It is possible to simplify the configuration and control contents for distributing the driving force of the first and second electric motors 5 and 6 to the driving shafts 11a and 12a. This reduces device costs and improves energy efficiency.

  According to the second invention, as shown in FIG. 1, in the wheel loader 1, the entire drive torque of the first electric motor 5 is transmitted to the drive shaft 11 a of the traveling body 13, and the second electric motor 6 is driven. All or part of the torque is transmitted to the drive shaft 11 a of the traveling body 13. When a part of the driving torque of the second electric motor 6 is transmitted to the driving shaft 11 a of the traveling body 13, the remaining driving torque of the second electric motor 6 is driven by the working machines 18 and 19. It is transmitted to the shaft 12a.

  The wheel loader 1 requires a large amount of power for traveling power during excavation. According to the second invention, not only the driving torque of the first electric motor 5 but also all of the driving torque of the second electric motor 6 that is normally distributed to the work machine power can be distributed to the traveling power. In addition, when excavation work is performed with the wheel loader 1, a large traveling power can be obtained, and work efficiency is improved.

  According to the third invention, the torque distribution transmission mechanism 10 is constituted by a differential gear mechanism using a differential gear such as the planetary gear 110 or the differential gear 130 shown in FIGS. 22 (a) and 22 (b), for example.

  According to the fourth invention, the torque distribution transmission mechanism 10 is constituted by the planetary gear mechanism using the planetary gear 110 shown in FIG.

  According to the fifth aspect of the present invention, as shown in FIG. 23A, the clutch 61 that engages and disengages the second input shaft 10b and the second output shaft (work machine output shaft) 10d is provided. In the case where only traveling power is required and work machine power is not required, the clutch 61 is disengaged, and all of the driving torques of the first and second electric motors 5 and 6 are supplied to the first output. It is controlled so that it is transmitted to the shaft (travel output shaft) 10c and not transmitted to the second output shaft (worker output shaft) 10d. For this reason, in a situation where only the traveling power is required and the work machine power is not required, the work machine output shaft 10d is not driven to rotate, and the pressure oil discharged from the hydraulic pump 12 is wastefully discharged to the tank, resulting in energy loss. It is avoided that it occurs.

  According to the sixth aspect of the invention, as shown in FIG. 23 (a), in addition to the clutch 61, the clutch 62 that connects and disconnects the second input shaft 10b and the gear 10e is provided. When the driving power is not required, the clutch 62 is disengaged, and the second electric motor 6 is controlled so that the ring gear 10R of the planetary gear 110 is not rotationally driven. For this reason, in a situation where only work machine power is required and travel power is not required, the travel output shaft 10c is not driven by the second electric motor 6 via the ring gear 10R of the planetary gear 110. Energy loss is avoided.

  According to the seventh invention, as shown in FIG. 23A, in addition to the clutches 61 and 62, the brake 63 for braking the ring gear 10R of the planetary gear 110 is provided. As a result, for example, as shown in FIG. 23B, the engagement and release of the clutches 61 and 62 and the operation and release of the brake 63 can be controlled according to the work modes A to E, and as a result. As a result, the consumption of extra power can be suppressed and the energy efficiency can be increased.

  As will be described later, particularly in a wheel loader, there is a frequent change in work ratio between excavation and traveling. For this reason, in order to obtain the optimum power transmission in each work, sequentially selecting the work mode leads to a great improvement in fuel consumption.

  Embodiments of a traveling work machine according to the present invention will be described below with reference to the drawings. In the following description, a wheel loader is assumed as the traveling work machine. However, the present invention can also be applied to other traveling work machine machines such as a hydraulic excavator.

  Fig.1 (a) has shown the internal structure of the wheel loader 1 of 1st Example.

  As shown in FIG. 1A, the drive unit 7 includes a battery 2, inverters 3 and 4, and first and second electric motors 5 and 6. The battery 2 is electrically connected to the first electric motor 5 via the inverter 3 and is electrically connected to the second electric motor 6 via the inverter 4. The battery 2 includes a storage battery such as a capacitor, a lead battery, a nickel metal hydride battery, or a lithium ion battery, and a fuel cell.

  The drive unit 7 indicated by a broken line in FIG. 1A may be replaced with the configuration shown in FIG. 1B, the output shaft of the engine 8 is connected to the drive shaft of the generator 9, and the generator 9 is connected to the first and second electric motors 5 and 6 via the inverters 3 and 4, respectively. It is comprised so that it may be electrically connected to. Furthermore, the drive unit 7 may be configured by combining FIGS.

  The output shafts 5a and 6a of the first and second electric motors 5 and 6 are connected to the first input shaft 10a and the second input shaft 10b of the planetary gear mechanism 10 as a torque distribution transmission mechanism, respectively. The first output shaft (travel output shaft) 10 c of the planetary gear mechanism 10 is connected to the input shaft 11 a of the power transmission mechanism 11 as the drive shaft 11 a of the travel body 13. A second output shaft (work machine output shaft) 10d of the planetary gear mechanism 10 is connected to a drive shaft 12a of a variable displacement hydraulic pump 12 as a drive shaft 12a of the work machine.

  The planetary gear mechanism 10 inputs the driving torque of the first and second electric motors 5 and 6, and uses the driving torque of the first and second electric motors 5 and 6 as the driving shaft 11 a of the traveling body 13 and the work machine. The drive shaft 12a is distributed and transmitted. In the planetary gear mechanism 10, all of the driving torque of the first electric motor 5 is transmitted to the driving shaft 11 a of the traveling body 13, and all or part of the driving torque of the second electric motor 6 is transmitted to the traveling body 13. It is configured to be transmitted to the drive shaft 11a.

  FIG. 12A shows the planetary gear 110 in a front view, and FIG. 12B shows an enlarged side view of the planetary gear 110 corresponding to FIG. Referring to FIG. 12 and FIG. 1 together, the output shaft 5a of the first electric motor 5 is connected to the sun gear 10S of the planetary gear 110 via the first input shaft 10a of the planetary gear mechanism 10. Yes. The output shaft 6 a of the second electric motor 6 is connected to the gear 10 e via the second input shaft 10 b of the planetary gear mechanism 10. The gear 10e meshes with the ring gear 10R of the planetary gear 110.

  The planetarium gear 10P of the planetary gear 110 is connected to the planetarium carrier 10PC, and the planetarium carrier 10PC is connected to the drive shaft 11a of the traveling body 13 via the first output shaft 10c of the planetary gear mechanism 10.

  The gear 10 e of the planetary gear mechanism 10 is connected to the drive shaft 12 a of the work machine via the second output shaft 10 d of the planetary gear mechanism 10.

  The power transmission mechanism 11 includes a drive shaft 11a, a reduction gear 11b connected to the drive shaft 11a, a propeller shaft 11c connected to the reduction gear 11b, a differential gear 11d connected to the propeller shaft 11c, and a differential. It consists of an axle shaft 11e connected to the gear 11d and a final gear 11f connected to the axle shaft 11e. The final gear 11f is connected to a wheel 13 as a traveling body. The wheel 13 corresponds to a front wheel of the wheel loader 1, for example. The configuration from the propeller shaft 11c to the wheel 13 shown in FIG. 1 can be implemented as a four-wheel drive vehicle with the same configuration for both the front wheels and the rear wheels.

  FIG. 2 is a block diagram showing the configuration of various hydraulic devices including the hydraulic pump 12, these various hydraulic devices, and a control system for driving and controlling the first and second electric motors 5 and 6.

  That is, the discharge port 12 b of the hydraulic pump 12 communicates with the cylinder chambers of the hydraulic cylinders 16 and 17 through the discharge oil passage 25, the direction switching valve 14, and the direction flow control valve 15. The hydraulic cylinders 16 and 17 are connected to working machines 18 and 19, respectively. The work machines 18 and 19 are, for example, the boom and bucket of the wheel loader 1.

  The cylinder chambers of the hydraulic cylinders 16 and 17 communicate with the suction port 12 c of the hydraulic pump 12 through the directional flow control valve 15, the directional switching valve 14, the return oil passage 27, and the tank 22. The tilting position of the swash plate of the hydraulic pump 12 is driven by the servo mechanism 28.

  The discharge oil passage 25 is provided with a pressure sensor 26 that detects the discharge pressure of the hydraulic pump 12. A rotation speed sensor 29 that detects the rotation speed of the hydraulic pump 12 is provided on the second output shaft 10 d or the drive shaft 12 a of the hydraulic pump 12.

  The first output shaft 10 c or the drive shaft 11 a of the power transmission mechanism 11 is provided with a torque sensor 30 that detects drive torque. The traveling body 13 is provided with a vehicle speed sensor 127 that detects the vehicle speed.

  The accelerator pedal 23 and the work machine operation lever 24 are provided, for example, in the cab. In the wheel loader 1, the direction of the vehicle body is changed by a steering operation, but this is omitted in this embodiment for convenience of explanation.

  A signal indicating the operation amount of the pressure sensor 26, the rotation speed sensor 29, the torque sensor 30, the vehicle speed sensor 127, the accelerator pedal 23, and the work machine operation lever 24 is input to the controller 20.

  The inverters 3 and 4 are driven and controlled by a control signal output from the electric motor controller 21. The electric motor controller 21 is driven and controlled by a control signal input from the controller 20.

  In the controller 20, a control signal for driving and controlling the servo mechanism 28 of the hydraulic pump 12, a control signal for driving and controlling the valve position of the directional flow control valve 15, and the first and second electric motors 5 and 6 are driven. Control signals for control are generated and output to the servo mechanism 28, the directional flow control valve 15, and the electric motor controller 21, respectively.

  When the controller 20 outputs a control signal for generating a positive (+) polarity torque in the first electric motor 5, the inverter 3 controls the first electric motor 5 to operate as an electric motor. . In this case, the DC power stored in the battery 2 is converted into AC power by the inverter 3 and supplied to the first electric motor 5, and the output shaft 5 a of the first electric motor 5 is rotationally driven. As a result, torque is generated in the first electric motor 5, and this torque is transmitted to the first input shaft 10 a of the planetary gear mechanism 10 via the output shaft 5 a of the first electric motor 5.

  When the controller 20 outputs a control signal for generating negative (−) polarity torque in the first electric motor 5, the inverter 3 controls the first electric motor 5 to operate as a generator. To do. In this case, the torque of the output shaft 5a of the first electric motor 5 is absorbed by the first electric motor 5 to generate power. The AC power generated by the first electric motor 5 is converted into DC power by the inverter 3, supplied to the battery 2 through the DC power supply line, and stored as DC power.

  When the controller 20 outputs a control signal for generating a positive (+) polarity torque in the second electric motor 6, the inverter 4 controls the second electric motor 6 to operate as an electric motor. . In this case, the DC power stored in the battery 2 is converted into AC power by the inverter 4 and supplied to the second electric motor 6, and the output shaft 6 a of the second electric motor 6 is rotationally driven. Thereby, torque is generated in the second electric motor 6, and this torque is transmitted to the second input shaft 10 b of the planetary gear mechanism 10 via the output shaft 6 a of the second electric motor 6.

  When the controller 20 outputs a control signal for generating a negative (−) polarity torque in the second electric motor 6, the inverter 4 controls the second electric motor 6 to operate as a generator. To do. In this case, the torque of the output shaft 6a of the second electric motor 6 is absorbed by the second electric motor 6 to generate power. The AC power generated by the second electric motor 6 is converted into DC power by the inverter 4, supplied to the battery 2 through the DC power supply line, and stored as DC power.

  The drive torque generated by the first electric motor 5 is transmitted to the traveling body 13 via the first input shaft 10a, the sun gear 10S, the planetarium gear 10P, the planetarium carrier 10PC, and the first output shaft 10c of the planetary gear mechanism 10. It is transmitted to the drive shaft 11a. On the other hand, the drive torque generated by the second electric motor 6 is supplied to the drive shaft 12a of the working machines 18 and 19, that is, the hydraulic pump, via the second input shaft 10b and the second output shaft 10d of the planetary gear mechanism 10. 12 is transmitted to the drive shaft 12a of the planetary gear mechanism 10 and the traveling body via the second input shaft 10b of the planetary gear mechanism 10, the gear 10e, the ring gear 10R, the planetarium gear 10P, the planetarium carrier 10PC, and the first output shaft 10c. 13 drive shafts 11a.

  When the battery 2 shown in FIG. 1A is replaced with the engine 8 and the generator 9 shown in FIG. 1B, the output torque of the engine 8 is absorbed by the generator 9 to generate power. Electric power is supplied to the first and second electric motors 5 and 6, and the first and second electric motors 5 and 6 operate as electric motors (positive (+) polarity torque is generated). When the first and second electric motors 5 and 6 operate as generators (negative (−) polarity torque is generated), the AC power generated by the first and second electric motors 5 and 6 is generated. Inverters 3 and 4 are converted to DC power, supplied to the generator 9, and the generator 9 operates as an electric motor. The torque generated by the generator 9 is transmitted to the output shaft of the engine 8, and the engine 8 Is added to the torque generated in

  FIGS. 3A and 3B are main components among the components shown in FIGS. 1A and 2, that is, the battery 2, the electric motors 5 and 6, the planetary gear mechanism 10, and the vehicle body of the hydraulic pump 12. An arrangement example is shown. FIG. 3A is a top view of the wheel loader 1, and FIG. 3B is a side view of the wheel loader 1.

  FIG. 6 is a block diagram showing the control system of the first and second electric motors 5, 6, that is, the configuration of the controller 20.

  As shown in FIG. 6, a signal indicating the operation amount (pedal depression amount) α of the accelerator pedal 23 is input to the deviation calculation unit 30. A signal indicating the current vehicle speed (current rotation speed of the travel output shaft 10 c) β detected by the vehicle speed sensor 127 is input to the deviation calculating unit 30. The deviation calculation unit 30 calculates a deviation α-β between the pedal depression amount α and the current vehicle speed β, and a control command corresponding to the deviation α-β is given to the first electric motor 5. In the torque sensor 30, the current torque of the travel output shaft 10c is detected, and this current torque is given to the first electric motor 5 as a feedback amount. The first electric motor 5 has its rotation speed and torque adjusted according to the deviation α−β.

  A signal indicating the operation amount ε of the operation lever 24 is input to the directional flow control valve 15. In the directional flow control valve 15, the valve position is changed according to the operation amount ε, the opening area of the spool is changed, the flow rate of the pressure oil supplied to the hydraulic cylinders 16, 17 is changed, and the working machines 18, 19 are changed. The operating speed is changed. In accordance with the current rotational speed γ of the work machine output shaft 10d detected by the rotational speed sensor 29 so that the hydraulic oil of the desired discharge flow rate (pump discharge amount per unit time) is discharged from the hydraulic pump 12, The tilt position of the swash plate of the hydraulic pump 12 (capacity; pump discharge amount per pump rotation) is adjusted. The current capacity of the hydraulic pump 12 and the current pressure oil pressure detected by the pressure sensor 26 are provided as feedback amounts to the second electric motor 6, and the second electric motor 6 responds to the deviation α−β. The rotation speed is adjusted, and the torque is adjusted according to the current rotation speed γ of the work machine output shaft 10d.

  FIGS. 7 and 8 are flowcharts showing the operation of the control system shown in FIG.

  First, the depression amount α of the accelerator pedal 23 is input (step 101), and it is determined whether or not the deviation α−β between the pedal depression amount α and the current vehicle speed β is not zero (step 102).

  When the deviation α−β is not 0, the torque increase / decrease value of the first electric motor 5 and the increase / decrease values of the rotation speeds of the first and second electric motors 5 and 6 are determined according to the value of the deviation α−β. Is calculated. Specifically, an increase / decrease value of the rotational speed of the travel output shaft 10c is first obtained. Then, the increase / decrease value of the rotation speed of the first input shafts 10a, 10b of the planetary gear mechanism 10 corresponding to the rotation speed increase / decrease value, that is, the increase / decrease value of the rotation speed of the first and second electric motors 5, 6 is obtained. It is done. The rotation speed and torque of the first electric motor 5 are adjusted so that the torque and the rotation speed increase and decrease by the torque increase and decrease value and the rotation speed increase and decrease value corresponding to the deviation α−β, and the second electric motor 6 is adjusted (step 103).

  Next, the current rotational speed β of the travel output shaft 10d is detected (step 104), and the processing from step 101 onward is repeatedly executed in the same manner.

  When the rotation speed of the second electric motor 6 is adjusted in step 103, the rotation speed γ of the work machine output shaft 10d is changed accordingly, and the discharge flow rate (pump discharge amount per unit time) of the hydraulic pump 12 is changed. Changed.

  Therefore, the current rotational speed γ of the work machine output shaft 10d is detected by the rotational speed sensor 29 (step 105), and the hydraulic oil having a desired discharge flow rate (pump discharge amount per unit time) is discharged from the hydraulic pump 12. As described above, the tilt position (capacity; pump discharge per pump rotation) of the swash plate of the hydraulic pump 12 is adjusted according to the current rotational speed γ of the work implement output shaft 10d (step 106).

  The operation amount ε of the operation lever 24 is input (step 107), the valve position of the directional flow control valve 15 is changed according to the operation amount ε, the opening area of the spool is changed, and supplied to the hydraulic cylinders 16 and 17. The flow rate of the pressure oil is changed, and the operating speed of the work machines 18 and 19 is changed (step 108).

  It is determined whether or not the rotational speed γ of the work implement output shaft 10d has changed (step 109). If the rotational speed γ of the work implement output shaft 10d has changed, a torque increment corresponding to the change is calculated. The torque of the second electric motor 6 is adjusted so that the torque increases by the calculated torque increment (step 110).

  Next, it is determined whether or not the torque of the second electric motor 6 is within the rated torque range (step 111). If the torque is within the rated torque range, the processing after step 107 is repeatedly executed in the same manner. However, if it is not within the rated torque range, the tilt position of the swash plate of the hydraulic pump 12 is adjusted to fall within the rated torque range to reduce the capacity of the hydraulic pump 12 (step 112). After the process of step 112, the procedure moves to step 109 again.

  Next, as illustrated in FIG. 4A, description will be made assuming a case where a V-shape operation, which is a typical operation of the wheel loader 1, is performed.

  As shown in FIG. 4 (a), the V shape operation means excavating the earth and sand 40 in which the wheel loader 1 is piled up, loading the earth and sand into the bucket, loading the loaded earth and sand into the dump truck 50, and waiting for the original standby. This is work (operation) in which a series of operations until returning to the position is one cycle.

  V-shape operation includes forward movement (1) in an unloaded state, excavation (2), backward movement in a loaded state (3), forward movement in a loaded state (4), loading (5), backward movement in an unloaded state ( 6) The six modes are configured. In the wheel loader 1, there is a frequent change in work ratio between excavation and traveling.

In other words, when the wheel loader 1 is in an empty state, the wheel loader 1 travels forward to the vicinity of the piled earth and sand 40 (empty load advance (1)) and reaches the vicinity of the piled earth and sand 40. 40, the working machines (boom, bucket) 18 and 19 are lifted and loaded with earth and sand in the bucket (excavation (2)). When the earth and sand are loaded in the bucket, the wheel loader 1 travels backward to the original standby point while moving the working machines 18 and 19 upward (backward loading (3)). When the original standby point is reached, the wheel loader 1 moves forward to the location of the dump truck 50 with the earth and sand being loaded in the bucket while changing the traveling direction (load advance (4)). Then, when the wheel loader 1 reaches the location of the dump truck 50, the traveling is stopped, and the working machines 18 and 19 are lowered to load earth and sand into the dump truck 50 (loading (5)). When all the earth and sand in the bucket is loaded on the dump truck 50, the wheel loader 1 travels backward to the original standby point in an empty state. When returning to the standby point, the work machines 18 and 19 are returned to their original positions and postures (empty reverse (6)).
FIG. 4B is a table showing the magnitude of the load applied to the traveling output shaft (first output shaft) 10c and the work implement output shaft (second output shaft) 10d for each mode of the V shape operation. ing. ◎ and ○ shown in FIG. 4B indicate the magnitude of the load, and ◎ indicates that the load is larger than ○.

  FIG. 5 is a graph showing the time change of the horsepower L3 consumed by the traveling output shaft 10c, the horsepower L2 consumed by the work implement output shaft 10d, and the total horsepower L1 combining both of these during V-shape operation. . Here, it is assumed that the work machine horsepower L2 includes horsepower required for the steering operation. In FIG. 5, the broken line L4 indicates the rated output per one of the first and second electric motors 5 and 6.

  Next, the operation during the V-shape operation will be described in association with the processes of FIGS. 7 and 8 by representing the unloading forward mode (1) and the excavation mode (2).

(1) At the time of empty load advancement The wheel loader 1 advances in an empty load state. Since it is in an empty state, the operation lever 24 is almost in the neutral position, and the directional flow control valve 15 is almost in the neutral position. Further, since the vehicle is traveling forward, the accelerator pedal 23 is depressed, and the rotation speed and torque of the first electric motor 5 are adjusted according to the depression amount α, and the rotation of the second electric motor 6 is performed. The number is adjusted (step 103 in FIG. 7). As shown in FIGS. 4B and 5, when the vehicle is moving forward with an empty load, a load is applied to the travel output shaft 10 c and horsepower is consumed by the travel output shaft 10 c, but there is almost no load on the work implement output shaft 10 d. The horsepower is almost not consumed.

(2) During excavation During excavation, the wheel loader 1 adjusts the position and posture of the bucket, which is one of the work machines 18 and 19, to the position of the excavation target earth and sand 40, while pushing the bucket with the earth and sand 40. Move forward slightly. When the wheel loader 1 penetrates deeply into the earth and sand 40, the traveling inertia energy of the vehicle body acts effectively, and the vehicle speed rapidly decreases. Therefore, the deviation α-β changes abruptly, the torque of the first electric motor 5 increases (step 103 in FIG. 7), and the torque of the second electric motor 6 increases (step 110 in FIG. 8). . However, the torques of the first and second electric motors 5 and 6 are limited by the current that can be output from the inverters 3 and 4.

  At the time of excavation, the accelerator pedal 23 is depressed and the wheel loader 1 not only keeps pressing the earth and sand 40 by traveling forward, but also operates the operation lever 24 to gradually operate the work machines 18 and 19 (boom and bucket) upward. Efficiently load earth and sand into the bucket. For this reason, the flow rate of the pressure oil supplied to the work machines 18 and 19 changes greatly (step 108 in FIG. 8).

  Thus, during excavation, a large load is applied to both the travel output shaft 10c and the work implement output shaft 10d.

  However, during excavation, the traveling output shaft 10c consumes a large horsepower that is greater than the horsepower consumed by the work implement output shaft 10d. Therefore, the torque of the second electric motor 6 has priority over the work implement output shaft 10d. It operates so as to be distributed to the shaft 10c.

  Further, as described above, since the deviation α-β changes rapidly, the rotation speeds of the first and second electric motors 5 and 6 change dynamically.

  For this reason, the second electric motor 6 changes greatly in the number of rotations according to the traveling state, and the number of rotations γ of the work implement output shaft 10d changes greatly accordingly.

  For this reason, the discharge flow rate (pump discharge amount per unit time) of the hydraulic pump 12 changes greatly, and the operating speed of the working machines 18 and 19 is adjusted in proportion to the operation amount ε of the operation lever 24 (step 107 in FIG. 8). 108) becomes unstable.

  Therefore, in order to maintain the discharge flow rate (pump discharge amount per unit time) of the hydraulic pump 12 at a constant desired flow rate, the swash plate of the hydraulic pump 12 is controlled according to the current rotational speed γ of the work machine output shaft 10d. The tilt position (capacity; pump discharge per pump rotation) is adjusted.

  For example, if the rotation speed of the second electric motor 6 is decreased by depressing the accelerator pedal 23, the rotation speed of the work machine output shaft 10d is decreased and the discharge flow rate of the hydraulic pump 12 (pump discharge per unit time). Amount) is reduced. In this case, in order to suppress a decrease in the discharge flow rate (pump discharge amount per unit time) of the hydraulic pump 12 and maintain a constant desired flow rate, the capacity of the hydraulic pump 12 is adjusted to be large ( Step 106 in FIG.

  As described above, during excavation, a large load is applied to both the travel output shaft 10c and the travel output shaft 10c as shown in FIG. 4B, and the work output shaft 10d is connected to the travel output shaft 10c as shown in FIG. A large horsepower more than the consumed horsepower is consumed, and the combined total horsepower is the largest among the modes during the V-shape operation.

  Note that the following operation is performed during reverse travel and loading.

  During reverse travel, the rotational directions of the planetarium gear 10P (travel output shaft 10c), ring gear 10R, and sun gear 10S of the planetary gear mechanism 10 during forward travel are reversed. For this reason, the 1st and 2nd electric motors 5 and 6 are rotationally driven in the direction opposite to the time of advance. As the second electric motor 6 is reversed, the rotation direction of the work machine output shaft 10d is also reversed, and the hydraulic pump 12 rotates in the reverse direction, so that the hydraulic oil is supplied from the hydraulic pump 12 to the directional flow control valve 15. In addition, the direction switching valve 14 is switched.

  At the time of loading, the rotation output of the traveling output shaft 10c is stopped, and only the work machine output shaft 10d is driven. Since the work implement output shaft 10d is connected to the ring gear 10R, the first electric motor 5 is rotated so that the sun gear 10S is reversed with respect to the ring gear 10R in order to stop the rotational drive of the travel output shaft 10c. Driven. However, since traveling power is not required, the first electric motor 5 may be rotated in an unloaded state.

  FIG. 9 shows the relationship between the rotational speed of the first and second electric motors 5 and 6 and the driving torque, and the relationship between the rotational speed of the travel output shaft 10c and the driving torque.

  In the excavation mode, the rotation speed and drive torque of the first electric motor 5 change as indicated by L11, and the rotation speed and drive torque of the second electric motor 6 use the point L12 and The rotation speed and driving torque of one output shaft (running output shaft) 10c change as indicated by L13. That is, as indicated by L13, the travel output shaft 10c outputs a maximum torque (approximately 320 Nm) from 0 rpm to around 1000 rpm, and is driven at a constant low torque value. At this time, the first electric motor 5 is driven between −1000 rpm and 1000 rpm at a constant torque of about 120 Nm, which is equal to or less than half of the torque of the travel output shaft 10c, as indicated by L11. When the traveling output shaft 10c is 0 rpm, the first electric motor 5 is driven at -1000 rpm.

  When the first electric motor 5 rotates in the negative direction and generates a negative torque, power is generated and electric power is stored in the battery 2 via the inverter 3. 1B, the torque generated by the first electric motor 5 is applied to the engine 8 to assist the engine 8.

  As shown by L12, the second electric motor 6 outputs a torque of 190 Nm at about 500 rpm. The maximum output of the second electric motor 6 is designed based on this operating point.

  The second electric motor 5 is also coupled to the work machine output shaft 10d. For this reason, when excavating a heavy load on the traveling output shaft 10c, if the work machines 18 and 19 are also operated simultaneously, the total horsepower may exceed the motor capacity. It operates to reduce the torque and distribute the torque to the work implement side. That is, the first and second electric motors 5 and 6 are operated with a limit at the maximum torque point. When the work machines 18 and 19 are operated in this state, the second electric motor 6 is The torque of the traveling output shaft 10c is automatically reduced by receiving the torque from the reaction from 19, and a part of the driving torque of the second electric motor 6 is distributed to the work implement output shaft 10d.

  In the traveling mode, for example, in the empty forward mode, the rotational speed and driving torque of the first electric motor 5 change as L14, and the rotational speed and driving torque of the second electric motor 6 change as L15. The rotational speed and driving torque of the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10 change as indicated by L16. That is, as indicated by L16, the travel output shaft 10c is driven along a constant horsepower line, that is, along a curve in which the torque is reduced as the vehicle speed increases. For example, as shown in FIG. 1B, when the engine 8 is used as a drive source, the traveling output shaft 10c is driven along a 27.2 kW constant horsepower curve that is the rated output of the engine 8. At this time, the first and second electric motors 5 and 6 are driven along a constant 13.6 kW horsepower curve corresponding to half of the rated output of the engine 8, as indicated by L14 and L15, respectively. In FIG. 9, the arrow indicates a state of changing in the high speed direction, that is, in the speed increasing direction.

Thus, the drive torque of the first electric motor 5 and the drive torque of the second electric motor 6 are added and transmitted to the travel output shaft 10c.

  The higher the vehicle speed (the higher the rotational speed of the traveling output shaft 10c at L16), the higher the rotational speed of the first and second electric motors 5 and 6 (the rotational speed increases at L14 and L15). As described above, the torque and rotation speed of the first and second electric motors 5 and 6 are controlled. However, the maximum vehicle speed is determined according to the gear ratio of the planetary gear 110 (the gear ratio of the sun gear 10S, the ring gear 10R, and the planetarium gear 10P).

  FIG. 10 shows a comparative example of the first embodiment of FIG.

  As shown in FIG. 10, in the comparative example, the traveling output shaft 10c and the work machine output shaft 10d are individually driven by the first electric motor 5 and the second electric motor 6, respectively, and Such torque distribution is not performed.

  FIG. 11 compares the output (kW), maximum torque (Nm), and maximum rotation speed (rpm) of the first and second electric motors 5 used in the first embodiment and the comparative example, respectively. .

  That is, about the output of the 1st electric motor 5 for driving power, in the case of a 1st Example, it may be 1/2 of the case of a comparative example, and the 1st and 2nd electric motors 5 and 6 are the whole. Even if it sees, it is 2/3, and it turns out that a motor capacity can be lowered significantly. Further, the maximum torque of the first electric motor 5 requires a large torque of 320 Nm in the case of the comparative example, but in the case of the first embodiment, it is less than half of that, and the second torque is the second torque. The maximum torque of the electric motor 6 is less than 2/3 of the comparative example, and it can be seen that the motor capacity can be significantly reduced. Further, the maximum rotational speeds of the first and second electric motors 5 and 6 are lower in the case of the first embodiment than in the case of the comparative example, whereby the motor can be reduced in size. I understand. As described above, according to the first embodiment, the capacities of the first and second electric motors 5 and 6 can be reduced and the size can be reduced, so that the cost can be suppressed and the wheel loader 1 can be accommodated in a small space. It becomes possible to do. As a result, the capacity of the inverter and the power supply line can be reduced, and the durability of the power element can be improved. Therefore, not only cost reduction but also reliability can be improved.

  Next, a second embodiment will be described.

  FIG. 13 is a diagram corresponding to FIG. 1A and shows the configuration of the second embodiment. FIG. 14 is a diagram corresponding to FIG. 9, and shows the relationship between the rotational speed and driving torque of the first and second electric motors 5 and 6 and the rotational speed and driving of the travel output shaft 10 c in the second embodiment. The relationship with torque is shown. In the following description, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted as appropriate.

  In the second embodiment, similarly to the first embodiment, the planetary gear mechanism 10 inputs the driving torques of the first and second electric motors 5 and 6 to input the first and second electric motors 5 and 6. The drive torque is distributed and transmitted to the drive shaft 11a of the traveling body 13 and the drive shaft 12a of the work implement. In the planetary gear mechanism 10, all of the driving torque of the first electric motor 5 is transmitted to the driving shaft 11 a of the traveling body 13, and all or part of the driving torque of the second electric motor 6 is transmitted to the traveling body 13. It is configured to be transmitted to the drive shaft 11a.

  That is, as shown in FIG. 13, the output shaft 5 a of the first electric motor 5 is connected to the gear 10 g via the first input shaft 10 a of the planetary gear mechanism 10. The gear 10g meshes with the ring gear 10R of the planetary gear 110.

The output shaft 6 a of the second electric motor 6 is connected to the gear 10 e via the second input shaft 10 b of the planetary gear mechanism 10. The gear 10e meshes with the gear 10f. The gear 10f is connected to the sun gear 10S of the planetary gear 110.
The planetarium gear 10P of the planetary gear 110 is connected to the planetarium carrier 10PC, and the planetarium carrier 10PC is connected to the drive shaft 11a of the traveling body 13 via the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10. It is connected.

  The gear 10 e of the planetary gear mechanism 10 is connected to the drive shafts 12 a of the working machines 18 and 19 via the second output shaft (working machine output shaft) 10 d of the planetary gear mechanism 10.

  According to the configuration of the second embodiment, as shown in FIG. 14, in the excavation mode, the rotational speed and driving torque of the first electric motor 5 change as indicated by L21, and the rotational speed of the second electric motor 6 is changed. The drive torque changes as indicated by L22, and the rotational speed and drive torque of the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10 change as indicated by L23. That is, as indicated by L23, the travel output shaft 10c outputs a maximum torque (approximately 320 Nm) from 0 rpm to around 1000 rpm, and is driven at a constant value of low speed and high torque.

  When the second electric motor 6 rotates in the negative direction and generates a negative torque, the torque generated by the second electric motor 6 is applied to the hydraulic pump 12 and the hydraulic pump 12 is driven to rotate. Is done. For this reason, when excavating while traveling mainly at low speed, the working machines 18 and 19 are operated with a large torque proportional to the traveling power, and a large excavating force is obtained. However, when the excavation work is simply a pressing work and a large excavation force is not required, the inverter 4 is configured to cause the second electric motor 6 to generate electric power to accumulate electric power in the battery 2 or assist the engine 8. May be controlled. In this case, it is desirable to control so that the generated electric power of the second electric motor 6 is increased by relieving the pressure oil discharged from the hydraulic pump 12 at a low pressure and returning it to the tank 22.

  In the traveling mode, for example, in the empty forward mode, the rotational speed and driving torque of the first electric motor 5 change as L24, and the rotational speed and driving torque of the second electric motor 6 change as L25. The rotational speed and driving torque of the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10 change as indicated by L26. That is, as indicated by L26, the travel output shaft 10c is driven along a constant horsepower line, that is, along a curve in which the torque is reduced as the vehicle speed increases. For example, as shown in FIG. 1B, when the engine 8 is used as a drive source, the traveling output shaft 10c is driven along a 27.2 kW constant horsepower curve that is the rated output of the engine 8. At this time, the first and second electric motors 5 and 6 are driven along a constant 13.6 kW horsepower curve corresponding to half of the rated output of the engine 8, as indicated by L24 and L25, respectively. Thus, the drive torque of the first electric motor 5 and the drive torque of the second electric motor 6 are added and transmitted to the travel output shaft 10c.

  The higher the vehicle speed (the higher the rotational speed of the travel output shaft 10c at L26), the higher the rotational speed of the first and second electric motors 5 and 6 (the rotational speed increases at L24 and L25). As described above, the torque and rotation speed of the first and second electric motors 5 and 6 are controlled. However, the maximum vehicle speed is determined according to the gear ratio of the planetary gear 110 (the gear ratio of the sun gear 10S, the ring gear 10R, and the planetarium gear 10P).

  Next, a third embodiment will be described.

  FIG. 15 is a diagram corresponding to FIG. 1A and shows the configuration of the third embodiment. FIGS. 16 to 18 are diagrams corresponding to FIG. 9, and the relationship between the rotational speed and the driving torque of the first and second electric motors 5 and 6 and the traveling output shaft 10 c or work in the third embodiment. The relationship between the rotation speed of the machine output shaft 10d and the drive torque is shown. In the following description, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted as appropriate.

  In the third embodiment, as in the first embodiment, the planetary gear mechanism 10 inputs the driving torques of the first and second electric motors 5 and 6 and inputs the first and second electric motors 5 and 6. The drive torque is distributed and transmitted to the drive shaft 11a of the traveling body 13 and the drive shaft 12a of the work implement. In the planetary gear mechanism 10, all or part of the driving torque of the first electric motor 5 is transmitted to the driving shaft 11 a of the traveling body 13, and all or part of the driving torque of the second electric motor 6 travels. It is configured to be transmitted to the drive shaft 11 a of the body 13.

  In the first embodiment and the second embodiment, the planetary gear mechanism 10 is configured by one planetary gear 110, whereas in the third embodiment, the planetary gear mechanism 10 includes two planetary gears, that is, The first planetary gear 110 and the second planetary gear 210 are configured.

  That is, as shown in FIG. 15, the output shaft 5 a of the first electric motor 5 is connected to the gear 10 h via the first input shaft 10 a of the planetary gear mechanism 10. The gear 10h meshes with the gear 10i. The gear 10 i is connected to the sun gear 110 </ b> S of the first planetary gear 110. Further, the gear 10 h meshes with the ring gear 210 </ b> R of the second planetary gear 210. Thus, when the output shaft 5a of the first electric motor 5 is rotationally driven, the sun gear 110S of the first planetary gear 110 is rotationally driven and the ring gear 210R of the second planetary gear 210 is rotationally driven. Is done.

  The output shaft 6 a of the second electric motor 6 is connected to the gear 10 j via the second input shaft 10 b of the planetary gear mechanism 10. The gear 10j meshes with the gear 10k, the gear 10k meshes with the gear 10l, and the gear 10l meshes with the gear 10m. The gear 10 m is connected to the sun gear 210 </ b> S of the second planetary gear 110. Further, the gear 10j meshes with the ring gear 110R of the first planetary gear 110. Thus, when the output shaft 6a of the second electric motor 6 is rotationally driven, the sun gear 210S of the second planetary gear 210 is rotationally driven and the ring gear 110R of the first planetary gear 110 is rotationally driven. Is done.

  The planetarium gear 110P of the first planetary gear 110 is connected to the planetarium carrier 110PC, and the planetarium carrier 110PC drives the traveling body 13 via the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10. It is connected to the shaft 11a.

  The planetarium gear 210P of the second planetary gear 210 is connected to the planetarium carrier 210PC, and the planetarium carrier 210PC is connected to the working machine 18 via the second output shaft (working machine output shaft) 10d of the planetary gear mechanism 10. It is connected to 19 drive shafts 12a.

  According to the configuration of the third embodiment, as shown in FIG. 16, in the traveling mode, for example, in the empty forward mode, the rotational speed and driving torque of the first electric motor 5 change as indicated by L31, and the second The rotational speed and driving torque of the electric motor 6 change as indicated by L32, and the rotational speed and driving torque of the first output shaft (traveling output shaft) 10c of the planetary gear mechanism 10 change as indicated by L33. That is, as indicated by L33, the travel output shaft 10c is driven along a constant horsepower line, that is, along a curve in which the torque is reduced as the vehicle speed increases. For example, as shown in FIG. 1B, when the engine 8 is used as a drive source, the traveling output shaft 10c is driven along a 27.2 kW constant horsepower curve that is the rated output of the engine 8. At this time, the first and second electric motors 5 and 6 are driven along a constant 13.6 kW horsepower curve corresponding to half of the rated output of the engine 8, as indicated by L31 and L32. Thus, the drive torque of the first electric motor 5 and the drive torque of the second electric motor 6 are added and transmitted to the travel output shaft 10c.

  The higher the vehicle speed (the higher the rotational speed of the travel output shaft 10c at L33), the higher the rotational speed of the first and second electric motors 5 and 6 (the rotational speed increases at L31 and L32). As described above, the torque and rotation speed of the first and second electric motors 5 and 6 are controlled.

  Although FIG. 16 shows a relationship L33 between the rotational speed and torque of the travel output shaft 10c, the work implement output shaft 10d is driven in conjunction with the travel output shaft 10c. That is, the work implement output shaft 10d is driven along a curve on which the torque is reduced as the vehicle speed increases, as with the curve L33.

  During the excavation mode and the loading mode, it is possible to operate in the travel priority mode or the work implement priority mode. The travel priority mode is a mode in which the travel output shaft 10c is prioritized over the work implement output shaft 10d and more torque is distributed to the travel output shaft 10c. The work implement priority mode is the travel output shaft 10c. In this mode, the work implement output shaft 10d is given priority over the work implement output shaft 10d.

  FIG. 17 shows a case where excavation or loading is performed in the travel priority mode. As shown in FIG. 17, the rotational speed and driving torque of the first electric motor 5 change as indicated by L41, and the second The rotation speed and drive torque of the electric motor 6 change as indicated by L42, and the rotation speed and drive torque of the second output shaft (working machine output shaft) 10d of the planetary gear mechanism 10 change as indicated by L43. That is, as indicated by L43, the work implement output shaft 10d is driven along a constant horsepower line, that is, along a curve in which the torque is reduced according to the increase in the rotational speed.

  Although FIG. 17 shows a relationship L33 between the rotation speed and torque of the work implement output shaft 10d, the traveling output shaft 10c is also driven in conjunction with the work implement output shaft 10d. That is, the travel output shaft 10c is driven along the curve L33 shown in FIG.

  When the torque of the travel output shaft 10c (L33 in FIG. 16) and the torque of the work implement output shaft 10d (L43 in FIG. 17) are compared, the travel output is compared with the torque of the work implement output shaft 10d at the same rotation speed. It can be seen that the torque of the shaft 10c is larger.

  The rotation speed of the work machine output shaft 10d varies according to a change in the rotation speed (vehicle speed) of the travel output shaft 10d and a change in torque. For this reason, the capacity of the hydraulic pump 2 is changed in accordance with the fluctuation of the rotation speed of the work machine output shaft 10d (the rotation speed of the hydraulic pump 2), and the hydraulic oil having a predetermined desired discharge flow rate is discharged from the hydraulic pump 12. Control is performed (step 106 in FIG. 8).

  FIG. 18 shows a case where excavation or loading is performed in the work machine priority mode. As shown in FIG. 18, the rotation speed and driving torque of the first electric motor 5 change as indicated by L51, and the second The rotational speed and driving torque of the electric motor 6 change as indicated by L52, and the rotational speed and driving torque of the second output shaft (working machine output shaft) 10d of the planetary gear mechanism 10 change as indicated by L53. That is, as indicated by L53, the work implement output shaft 10d is driven along a constant horsepower line, that is, along a curve in which the torque is reduced in accordance with the increase in the rotational speed.

  However, in order to always stop the rotation drive of the travel output shaft 10c, the second electric motor 6 follows the curve L52 as the first electric motor 5 shifts from the low rotation to the high rotation along the curve L51. Thus, control is performed so as to shift from low rotation to high rotation.

  Since the work implement output shaft 10d is not interlocked with the rotation of the travel output shaft 10c, the torque of the work implement output shaft 10d can be changed regardless of the vehicle speed. That is, if the work implement output shaft 10d is maintained in a low rotation and high torque state, a larger torque can be output from the work implement output shaft 10d as compared with the travel priority mode of FIG.

  For this reason, the work efficiency at the time of excavation and loading improves.

  FIG. 19 compares the output (kW), maximum torque (Nm), and maximum rotational speed (rpm) of the first and second electric motors 5 used in the third embodiment and the comparative example (FIG. 10), respectively. It shows.

  In other words, the output of the first electric motor 5 is 1/2 of that of the comparative example in the case of the third embodiment, and 2/3 of the first and second electric motors 5 and 6 as a whole. It can be seen that the motor capacity can be greatly reduced. The maximum torque of the first electric motor 5 requires a large torque of 320 Nm in the case of the comparative example, but in the case of the third embodiment, it is less than half that. It can be seen that the capacity can be kept low. Further, the maximum rotational speeds of the first and second electric motors 5 and 6 are higher in the case of the third embodiment than in the case of the comparative example, whereby the motor can be reduced in size. I understand. As described above, according to the third embodiment, since the capacities of the first and second electric motors 5 and 6 can be reduced and the size can be reduced, the cost can be reduced and the wheel loader 1 can be accommodated in a small space. It becomes possible to do. As a result, the capacity of the inverter and the power supply line can be reduced, and the durability of the power element can be improved. Therefore, not only cost reduction but also reliability can be improved.

  The first, second, and third embodiments described above are merely examples, and various modifications are possible.

  In particular, the planetary gear mechanism 10 is not limited to the configurations of the first, second, and third embodiments, and various configuration examples are conceivable.

  20A, 20B, 20C, and 20D show various configuration patterns of the planetary gear mechanism 10. FIG. The description will be made assuming that the reference numerals in FIG. 20 are the same as those used in the first, second, and third embodiments.

  FIG. 20A shows a pattern in which the planetary gear mechanism 10 is constituted by one planetary gear 110, where the first input shaft 10 a is connected to the planetary gear 110 and the second input shaft 10 b is connected to the planetary gear 110. A configuration pattern in which the second input shaft 10b is connected to the second output shaft 10 is shown. The configuration pattern in FIG. 20A has two patterns depending on the connected components. These two patterns correspond to the planetary gear mechanism 10 of the first embodiment and the second embodiment.

  In the case of the first embodiment, the first input shaft 10a is connected to the sun gear 10S of the planetary gear 110, the second input shaft 10b is connected to the ring gear 10R of the planetary gear 110, and the second input shaft 10b. Is connected to the second output shaft 10d.

  In the case of the second embodiment, the first input shaft 10a is connected to the ring gear 10R of the planetary gear 110, the second input shaft 10b is connected to the sun gear 10S of the planetary gear 110, and the second input shaft 10b. Is connected to the second output shaft 10d. In both of these patterns, the planetarium carrier 10PC (planetarium gear 10P) of the planetary gear 110 is connected to the first output shaft 10c.

  In the planetary gear 110, a relationship is established in which the addition of the outputs of the ring gear 10R and the sun gear 10S becomes the output of the planetarium carrier 10PC (planetarium gear 10P). By effectively utilizing this characteristic, in the first and second embodiments, the first output shaft 10c is configured as the traveling output shaft 10c, and the second output shaft 10d is configured as the work implement output shaft 10d. All of the driving torque of the first electric motor 5 is transmitted to the driving shaft 11a of the traveling body 13, and all or part of the driving torque of the second electric motor 6 is transmitted to the driving shaft 11a of the traveling body 13. Thus, a large torque is output from the travel output shaft (first output shaft) 10c during excavation.

  FIG. 20B shows a pattern in which the planetary gear mechanism 10 is constituted by two planetary gears 110 and 210, and the first input shaft 10 a is connected to the first planetary gear 110 and the second planetary gear 210. 2 shows a configuration pattern in which the second input shaft 10b is coupled to the first planetary gear 110 and the second planetary gear 210. In the configuration pattern of FIG. 20B, there are 12 patterns depending on the connected components. These 12 patterns are shown in FIG.

  FIG. 21 shows each component of the first planetary gear 110 and the second planetary gear 210 to which the first input shaft 10a is connected for each pattern (second column from the left in the table), the second Each component (third column from the left in the table) of the first planetary gear 110 and the second planetary gear 210 to which the input shaft 10b is coupled, and each of the first and second output shafts 10c and 10d are coupled. The components of the first and second planetary gears 110 and 210 (the rightmost column in the table) are shown in the table. In FIG. 21, S represents a sun gear, R represents a ring gear, and P represents a planetarium carrier (planetarium gear).

  The case of the third embodiment corresponds to the pattern 2 in FIG. That is, in the case of the third embodiment, the first input shaft 10a is connected to the sun gear 110S (S in FIG. 21) of the first planetary gear 110 and the ring gear 210R (see FIG. 21) of the second planetary gear 210. 21), the second input shaft 10b is connected to the ring gear 110R (R in FIG. 21) of the first planetary gear 110, and the sun gear 210S (FIG. 21) of the second planetary gear 210. S) and the first and second output shafts 10c and 10d are connected to the planetarium carriers 110PC and 210PC (P and P in FIG. 21) of the first and second planetary gears 110 and 210, respectively. Thus, the planetary gear mechanism 10 is configured.

  FIG. 20C shows a pattern in which the planetary gear mechanism 10 is constituted by two planetary gears 110 and 210. The first input shaft 10a is connected to the first planetary gear 110, and the second input shaft 10b. Shows a configuration pattern in which is connected to the second planetary gear 210, the first output shaft 10 c is connected to the second planetary gear 210, and the second output shaft 10 d is connected to the first planetary gear 110. ing. In the configuration pattern of FIG. 20C, there are approximately 20 patterns depending on the connected components.

  FIG. 20D shows a pattern in which the planetary gear mechanism 10 is constituted by two planetary gears 110 and 210, and the first input shaft 10 a is connected to the first planetary gear 110 and the second planetary gear 210. The configuration pattern in which the second input shaft 10 b is connected to the second planetary gear 210 and the second output shaft 10 d is connected to the first planetary gear 110 is shown. In the configuration pattern of FIG. 20D, there are 30 or more patterns depending on the connected components.

  In view of the fact that the total capacity of the electric motor can be reduced and the energy efficiency is improved, the configuration patterns shown in FIGS. 20A and 20B are particularly effective.

  Further, in each of the above embodiments, the case where the torque distribution transmission mechanism 10 is constituted by a planetary gear mechanism has been described. However, the first and second electric motors 5 and 6 are input to input the driving torque. Any configuration can be used as long as it can distribute and transmit the drive torque of the second electric motors 5 and 6 to the drive shaft 11a of the traveling body 13 and the drive shaft 12a of the work machines 18 and 19.

  For example, if the torque distribution / transmission mechanism 10 is configured by a combination of gears, the present invention is not limited to the planetary gear mechanism, and can be configured in any manner within the scope of the differential gear mechanism.

  22 (a) and 22 (b) show differential gear mechanisms equivalent to each other.

  FIG. 22A shows the planetary gear 110 used in the first to third embodiments. FIG. 22B shows a differential gear 310 used in a differential device for a rear wheel of an automobile.

  Both the planetary gear 110 and the differential gear 310 are equivalent in that torque is input from the first input shaft and the second input shaft, these input torques are added, and the added torque is output from the output shaft. It is. Therefore, for example, the planetary gear 110 may be replaced with the differential gear 310 to constitute the torque distribution / transmission mechanism 10 of the first to third embodiments.

  Next, an embodiment in which the consumption of excess power is suppressed and the energy efficiency can be further increased will be described.

  FIG. 23A shows a fourth embodiment in which a clutch 61, a clutch 62, and a brake 63 are added to the first embodiment.

  That is, in the first embodiment, even when only traveling power is required and work machine power is not required, the work machine output shaft 10d is always driven to rotate, so that the pressure oil discharged from the hydraulic pump 12 is not discharged. In some cases, energy was lost due to wasteful discharge into the tank. Therefore, as shown in FIG. 23 (a), a clutch 61 that engages and disengages the second input shaft 10b and the second output shaft (work machine output shaft) 10d is provided, and only the traveling power is provided. Is required and the work machine power is not required, the clutch 61 is disengaged, and all the drive torques of the first and second electric motors 5 and 6 are supplied to the first output shaft (travel output shaft). It may be controlled so that it is transmitted to 10c and not transmitted to the second output shaft (work machine output shaft) 10d.

  In the first embodiment, the second electric motor 6 always uses not only the work machine output shaft 10d but also the ring gear 10R of the planetary gear 110 even in a situation where only the work machine power is required and traveling power is not required. Due to the rotational drive, energy loss may occur. Therefore, as shown in FIG. 23 (a), when the clutch 62 for connecting and releasing the connection between the second input shaft 10b and the gear 10e is provided, only the work machine power is required and the travel power is not required. The clutch 62 is disengaged, and only the second output shaft (working machine output shaft) 10d is rotated by the second electric motor 6, and the ring gear 10R of the planetary gear 110 is controlled not to be rotated. Also good.

  Further, a brake 63 that brakes the ring gear 10R of the planetary gear 110 may be provided, and the engagement and release of the clutches 61 and 62 and the operation and release of the brake 63 may be controlled according to the work mode. When the brake 63 is operated, the ring gear 10R is fixed at a fixed position, and when the operation of the brake 63 is released, the ring gear 10R is in a free rotating state.

  FIG. 23B shows the work mode, the drive shaft (travel output shaft, work implement output shaft) to which power is distributed, the engagement and release states of the clutches 61 and 62, the operation and release state of the brake 63, and the power. It is the table | surface which showed the relationship with a source.

  In the work mode A, only the second electric motor 6 is rotationally driven as a power source, the clutch 61 is engaged, the clutch 62 is released, and the brake 63 is operated. In this case, the power of the second electric motor 6 is transmitted to the second output shaft (work machine output shaft) 10d.

  In the work mode B, the first and second electric motors 5 and 6 are rotationally driven as power sources, the clutch 61 is released, the clutch 62 is engaged, and the brake 63 is released. The In this case, the powers of the first and second electric motors 5 and 6 are added together and transmitted to the first output shaft (traveling output shaft) 10c.

  In the work mode C, the first electric motor 5 is rotationally driven as a power source, the clutch 61 is released, the clutch 62 is released, and the brake 63 is operated. In this case, the power of the first electric motor 5 is transmitted to the first output shaft (travel output shaft) 10c.

  In the work mode D, the first and second electric motors 5 and 6 are rotationally driven as power sources, the clutch 61 is engaged, the clutch 62 is engaged, and the brake 63 is released. Is done. In this case, the power of the first and second electric motors 5 and 6 is distributed and transmitted to the first output shaft (travel output shaft) 10c and the second output shaft (worker output shaft) 10d. Is done.

  In the work mode E, the first and second electric motors 5 and 6 are rotationally driven as power sources, the clutch 61 is engaged, the clutch 62 is released, and the brake 63 is operated. The In this case, the powers of the first and second electric motors 5 and 6 are independently supplied to the first output shaft (travel output shaft) 10c and the second output shaft (work machine output shaft) 10d, respectively. Communicated.

  For example, a switch for selecting the operation modes A to E is provided on the operation panel of the wheel loader cab, and according to the operation of the selection switch, as shown in FIG. The release, the operation of the brake 63, and the release can be automatically controlled.

  In many cases, the wheel loader is operated in a situation where greater power is required for the travel output shaft 10c than for the work machine output shaft 10d. If the work mode B is selected in such a situation, all the power obtained by adding up the two electric motors 5 and 6 is transmitted only to the travel output shaft 10c, so that the work efficiency can be further improved.

  Further, when traveling on a general road, if the work mode C is selected, the vehicle is driven by the power of one electric motor 5, thereby improving fuel efficiency.

  Further, by selecting the work mode A at the loading work and selecting the work mode D or the work mode E according to the situation at the time of excavation, the work efficiency is improved throughout the V shape operation. In particular, in the V-shape operation, as described above, excavation, transportation, and loading are repeated in a short time, so that it is difficult to select the above-described work mode on the operation panel. Therefore, it is possible to automatically select a work mode as shown in FIG. That is, assuming V-shape operation basically, when the work implement operating lever 24 is not operated for a certain period of time after selecting the work mode D, the clutch 61 is released and the operation mode B is automatically shifted. Is done. In this case, it is assumed that the load is in the bucket. For this reason, the opening of the spool of the control valve 15 is closed so that the working machine does not fall. Further, when the work machine operation lever 24 is operated during operation in the work mode B, the clutch 61 is engaged and the work mode D is restored. If the accelerator pedal 23 is not depressed for a certain time after selecting the work mode D, the clutch 62 is released, the brake 63 is operated, and the operation mode A is entered. Further, when the accelerator pedal 23 is depressed during operation in the operation mode A, the operation mode D is restored.

  As described above, according to the present embodiment, since the engagement and release of the clutches 61 and 62 and the operation and release of the brake 63 are controlled according to the working state, consumption of excess power is suppressed and energy is reduced. Efficiency can be increased.

  In the first embodiment, the case where the clutch 61, the clutch 62, and the brake 63 are provided has been described. However, the clutches 61, 62, and the brake 63 may be provided similarly in the second and third embodiments. .

1A and 1B are configuration diagrams of the first embodiment. FIG. 2 is a block diagram of the first embodiment. FIGS. 3A and 3B are diagrams showing examples of arrangement of components in the wheel loader. 4A and 4B are diagrams for explaining each mode during the V-shape operation. FIG. 5 is a graph showing the consumed horsepower. FIG. 6 is a block diagram of the control system. FIG. 7 is a flowchart showing the processing contents in the controller. FIG. 8 is a flowchart showing the processing contents in the controller. FIG. 9 is a diagram showing the relationship between the rotational speed and torque in the first embodiment. FIG. 10 is a configuration diagram of a comparative example. FIG. 11 is a diagram for explaining the effect of the first embodiment. FIG. 12 is a front view of the planetary gear. FIG. 13 is a block diagram of the second embodiment. FIG. 14 is a diagram showing the relationship between the rotational speed and torque in the second embodiment. FIG. 15 is a block diagram of the third embodiment. FIG. 16 shows the relationship between the rotational speed and torque in the third embodiment. FIG. 17 is a diagram showing the relationship between the rotational speed and torque in the third embodiment. FIG. 18 is a diagram showing the relationship between the rotational speed and torque in the third embodiment. FIG. 19 is a diagram for explaining the effect of the third embodiment. 20A, 20B, 20C, and 20D are diagrams illustrating configuration patterns of the planetary gear mechanism. FIG. 21 is a table showing various patterns of the configuration pattern shown in FIG. 22A and 22B are diagrams for explaining that each is equivalent as a differential gear. FIG. 23A is a block diagram of the fourth embodiment, and FIG. 23B is a table corresponding to FIG. FIG. 24 is a diagram illustrating processing for changing the work mode.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Wheel loader 5 1st electric motor 6 2nd electric motor 10 Torque distribution transmission mechanism 11a Driving shaft of traveling body 12a Driving shaft of working machine

Claims (8)

In the traveling work machine including the traveling body (13) and the work machines (18, 19),
A first electric motor (5);
A second electric motor (6);
The drive torques of the first and second electric motors (5, 6) are input, and the drive torques of the first and second electric motors (5, 6) work with the drive shaft (11a) of the traveling body (13). A torque distribution transmission mechanism (10) distributed and transmitted to the drive shaft (12a) of the machine (18, 19) ,
The torque distribution transmission mechanism (10) transmits the entire driving torque of the first electric motor (5) to the driving shaft (11a) of the traveling body (13), and the driving torque of the second electric motor (6). A traveling work machine characterized in that all or a part of the traveling machine is transmitted to the drive shaft (11a) of the traveling body (13) . The traveling work machine according to claim 1, wherein the torque distribution transmission mechanism is constituted by two planetary gears. The traveling work machine according to claim 1, further comprising control means for drivingly controlling the first and second electric motors (5, 6). In the traveling work machine including the traveling body (13) and the work machines (18, 19),
A first electric motor (5);
A second electric motor (6);
The drive torques of the first and second electric motors (5, 6) are input, and the drive torques of the first and second electric motors (5, 6) work with the drive shaft (11a) of the traveling body (13). A torque distribution transmission mechanism (10) distributed and transmitted to the drive shaft (12a) of the machine (18, 19);
With
The output shaft (6a) of the second electric motor (6) is connected to the second input shaft (10b) of the torque distribution transmission mechanism (10),
The second output shaft (10d) of the torque distribution transmission mechanism (10) is connected to the drive shaft (12a) of the work machine (18, 19),
A clutch (61) that engages and disengages the second input shaft (10b) and the second output shaft (10d) is provided.
A traveling work machine characterized by that. The torque distribution transmission mechanism (10) is composed of a planetary gear mechanism,
The output shaft (6a) of the second electric motor (6) is connected to the second input shaft (10b) of the torque distribution transmission mechanism (10),
The second output shaft (10d) of the torque distribution transmission mechanism (10) is connected to the drive shaft (12a) of the work machine (18, 19),
A clutch (61) that engages and disengages the second input shaft (10b) and the second output shaft (10d) is provided.
The traveling work machine according to claim 4, further comprising a clutch (62) for connecting and releasing the second input shaft (10b) and the planetary gear (110). The traveling work machine according to claim 5, wherein a brake (63) for braking the ring gear (10R) of the planetary gear (110) is provided. The traveling work machine according to claim 4 or 5, further comprising control means for controlling engagement and release of the clutch (61, 62). The traveling work machine according to claim 6, further comprising control means for controlling engagement and release of the clutch (61, 62) and controlling operation and release of the brake (63).

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