Work vehicle

Abstract

When switching control is performed, in which a power transmission path of a transmission is switched from an HMT to a direct connection mechanism, a shift control device of a work vehicle performs pump volume reduction control for reducing the pump volume of an HST hydraulic pump of the HMT from maximum volume to zero while a lock-up clutch of the direct connection mechanism is kept in a pressed state. Pump volume reduction control reduces the pump volume of the HST hydraulic pump in at least two stages and is performed such that the rate of reduction in pump volume in a second stage of pump volume reduction control is lower than the rate of reduction in pump volume in a first stage, the second stage following the first stage.

Description

Work vehicles

[0001] The present invention relates to a work vehicle equipped with a transmission for traveling.

[0002] Work vehicles such as wheel loaders and wheel excavators are configured to perform tasks such as excavating, transporting, and loading earth and sand using a work implement, while also propelling themselves using wheels. Since wheel propulsion allows for high-speed travel, the range of vehicle speeds in use is wide. Therefore, work vehicles must be equipped with a travel transmission with a wide speed range. Furthermore, from the perspective of efficiency, the travel transmission must also have high transmission efficiency. One example of a transmission that meets the above requirements while also taking cost into consideration is a continuously variable transmission (CVT). A CVT can operate an engine, which serves as a power source for traveling, at the most fuel-efficient operating point depending on the load conditions.

[0003] One type of CVT is the hydrostatic transmission (HST). The HST is a transmission configured with a closed hydraulic circuit that combines a variable-displacement hydraulic pump and a hydraulic motor. In an HST, continuously varying the displacement of the hydraulic pump or hydraulic motor achieves infinite speed change, and the theoretical speed change range is unlimited. However, because power is transmitted via fluid, the efficiency is relatively low among CVTs, and the efficiency drops even further near both ends of the speed change range. Therefore, the speed change range that can actually be applied to an HST is limited.

[0004] One type of CVT is the hydraulic mechanical transmission (HMT), which is different from the HST in that all power is transmitted via fluid. The HMT is configured so that a portion of the power is transmitted via fluid by the HST, while the remaining power is transmitted mechanically without the use of fluid. The HMT has the function of continuously variable transmission, but by also using mechanical power transmission, it can achieve higher transmission efficiency than a configuration with only an HST. However, the speed range is narrower than a configuration with only an HST, so it is necessary to use a means to expand the speed range.

[0005] Known means for expanding the speed range include a configuration in which an auxiliary transmission is connected in series to the HMT, and a configuration in which the HMT itself has multiple power transmission paths and switches to one of the paths. An example of the latter means for expanding the speed range is the technology described in Patent Document 1. In the transmission structure described in Patent Document 1, an HMT has two power transmission paths and is equipped with a clutch for each path. In this transmission structure, the rotation speeds of both power transmission paths are matched, and the clutches of both systems are simultaneously engaged with slippage, thereby switching from one power transmission path to the other.

[0006] JP 2023-14983 A

[0007] The technology described in Patent Document 1 aims to expand the gear range by making the power transmission path of the HMT itself switchable between two systems. However, with this configuration, because both power transmission paths include the HST path, which transmits power via fluid, power transmission efficiency tends to be low even over the expanded gear range.

[0008] Therefore, a possible solution to the technology described in Patent Document 1 is to change the power transmission system to two systems: one power transmission path via the HMT and one mechanical power transmission path that bypasses the HMT. In this configuration, a mechanical power transmission mechanism with a fixed gear ratio can be used to expand the gear range and improve efficiency. However, when switching the power transmission path from the HMT to the mechanical power transmission mechanism, if the rotation speeds of the two systems do not match, a gear shift shock occurs. In some cases, this can lead to a stall of the drive source (prime mover).

[0009] The present invention has been made to solve the above problems, and its object is to provide a work vehicle that can suppress gear change shock and avoid stalling of the prime mover when switching the power transmission path of the transmission from HMT to a mechanical power transmission mechanism.

[0010] The present application includes a plurality of means for solving the above-mentioned problems. One example thereof is a transmission including a prime mover as a power source, a hydromechanical first power transmission device that continuously changes the speed of the rotational power of the prime mover and transmits it, and a mechanical second power transmission device that has a gear ratio higher than the highest gear ratio of the first power transmission device and transmits the rotational power of the prime mover without passing through the first power transmission device, and a control device that controls the transmission, wherein the first power transmission device includes a mechanical power transmission mechanism that divides the rotational power input from the prime mover into two and outputs them, and a hydrostatic power transmission mechanism that has a variable displacement hydraulic pump and a variable displacement hydraulic motor fluidly connected to each other, and one of the rotational powers output from the mechanical power transmission mechanism is input to the hydraulic pump and output from the hydraulic motor, and the second power transmission device has an engaged state that brings the power transmission path of the second power transmission device into a connected state and an engaged state that brings the power transmission path of the second power transmission device into a disengaged state. In a work vehicle configured to include a clutch that can be switched between a released state and a disengaged state, the control device performs switching control to switch the power transmission path of the transmission from the first power transmission device to the second power transmission device when a predetermined condition is satisfied, the switching control changes the clutch from the released state to a pressed state that allows it to be switched to the engaged state, and performs pump volume reduction control to reduce the pump volume of the hydraulic pump of the hydraulic power transmission mechanism from the state immediately before the switching control to a predetermined volume while keeping the clutch in the pressed state, the pump volume reduction control reduces the pump volume of the hydraulic pump in at least two stages, and is performed so that the pump volume reduction rate in a second stage following the first stage is slower than the pump volume reduction rate in an initial first stage of the pump volume reduction control.

[0011] According to one example of the solution of the present application, when the power transmission path of the transmission is switched from a hydromechanical first power transmission device (HMT) to a mechanical second power transmission device, it is possible to suppress gear shift shock and avoid stalling of the prime mover. Other problems, configurations, and effects will become clear from the description of the following embodiments.

[0012] 1 is a side view showing an overview of a wheel loader as a work vehicle according to a first embodiment of the present invention. FIG. 1 is a schematic diagram showing the configuration of a traveling power transmission system provided in the work vehicle according to the first embodiment shown in FIG. 1. FIG. 2 is a schematic diagram showing the configuration of a transmission in the traveling power transmission system of the work vehicle according to the first embodiment shown in FIG. 3. FIG. 4 is an explanatory diagram showing a power transmission path when the gear ratio of the HMT in the transmission of the traveling power transmission system of the work vehicle according to the first embodiment shown in FIG. 3 is at the lowest speed. FIG. 5 is an explanatory diagram showing a power transmission path when the gear ratio of the HMT in the transmission of the traveling power transmission system of the work vehicle according to the first embodiment shown in FIG. 3 is at the highest speed. FIG. 6 is an explanatory diagram showing a power transmission path of a direct-coupled mechanism in the transmission of the traveling power transmission system of the work vehicle according to the first embodiment shown in FIG. 3 (highest gear ratio of the transmission). FIG. 7 is a time chart showing an example of transitions in the states of the lock-up clutch, the HST hydraulic pump, the engine, and vehicle traveling when switching power transmission paths using a switching method of a comparative example for the transmission shown in FIG. 3. 9 is a time chart showing another example of transitions in the states of the lockup clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path using a switching method of a comparative example for the transmission shown in FIG. 3 (an example of the occurrence of an engine stall). FIG. 10 is a block diagram showing the configuration of hardware and functional units of a gear change control device for a work vehicle according to the first embodiment shown in FIG. 3. FIG. 11 is a flowchart showing an example of a processing procedure for controlling the transmission (switching control of the power transmission path from the HMT to the direct drive mechanism) using the gear change control device for a work vehicle according to the first embodiment shown in FIG. 10. FIG. 11 is a time chart showing an example of transitions in the states of the lockup clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path of the transmission (switching from the HMT to the direct drive mechanism) in the work vehicle according to the first embodiment. FIG. 12 is a block diagram showing the configuration of hardware and functional units of a gear change control device for a work vehicle according to a second embodiment of the present invention.14 is a time chart showing an example of transitions in the states of the lock-up clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path of the transmission in the work vehicle according to the second embodiment shown in FIG. 13 (switching from HMT to direct drive mechanism). FIG. 15 is a block diagram showing an example of the configuration of hardware and functional parts of a gear change control device for a work vehicle according to a third embodiment of the present invention. FIG. 16 is a flowchart showing an example of a processing procedure for control of the transmission (switching control of the power transmission path from HMT to direct drive mechanism) by the gear change control device for a work vehicle according to the third embodiment shown in FIG. 15 is a time chart showing an example of transitions in the states of the lock-up clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path of the transmission in the work vehicle according to the third embodiment shown in FIG. 15.

[0013] Hereinafter, an embodiment of a work vehicle of the present invention will be described with reference to the drawings. In this embodiment, a wheel loader will be described as an example of a work vehicle. Note that the front, rear, left, and right directions described in this specification refer to directions as seen by an operator aboard the work vehicle.

[0014] [First embodiment] First, the general configuration of a wheel loader as a work vehicle according to a first embodiment will be described using Figures 1 and 2. Figure 1 is a side view showing an outline of a wheel loader as a work vehicle according to a first embodiment of the present invention. Figure 2 is a schematic diagram showing the configuration of a traveling power transmission system provided in the work vehicle according to the first embodiment shown in Figure 1.

[0015] In Figure 1, the wheel loader 1 is a work vehicle capable of traveling on wheels and performing work using a work implement. The wheel loader 1 is, for example, an articulated type in which the vehicle is steered by bending near the center, and includes a front frame 2 that forms the front portion of the vehicle body, a rear frame 3 that forms the rear portion of the vehicle body, a pair of left and right front wheels 4 (see also Figure 2) rotatably mounted on the front frame 2, and a pair of left and right rear wheels 5 (see also Figure 2) rotatably mounted on the rear frame 3. In the wheel loader 1, the front frame 2 and the rear frame 3 are connected via a center pin (not shown) so as to be rotatable in the left-right direction, and the front frame 2 is configured to bend in the left-right direction relative to the rear frame 3. The rear frame 3 includes a machine room 3a that houses various devices. Examples of the devices housed in the machine room 3a will be described later.

[0016] A hydraulically driven working device 6 is attached to the front of the front frame 2. The working device 6 is composed of, for example, a lift arm 7 and a bucket 8 serving as a working implement. The base end of the lift arm 7 is attached to the front of the front frame 2 so as to be rotatable in the vertical direction. The base end of the bucket 8 is attached to the tip of the lift arm 7 so as to be rotatable in the vertical direction. The lift arm 7 and the bucket 8 are each driven by a hydraulic cylinder (not shown) that expands and contracts with a supply of pressure oil. The lift arm 7 is configured to raise or lower the bucket 8, for example, by extending or contracting the hydraulic cylinder. The bucket 8 is configured to tilt (rotate upward relative to the lift arm 7) or dump (rotate downward relative to the lift arm 7), for example, by extending or contracting the hydraulic cylinder.

[0017] A cab 9 in which an operator sits is installed in the front of the rear frame 3. Inside the cab 9, operating devices (not shown) are arranged for the operator to operate the wheel loader 1. Examples of operating devices include an accelerator pedal, a forward / reverse command switch, a brake pedal, and a steering wheel.

[0018] 2, the wheel loader 1 is equipped with a traveling power transmission system 20 that transmits rotational power output by a prime mover 11 as a power source to the front wheels 4 and rear wheels 5 to cause the vehicle to travel. The traveling power transmission system 20 is, for example, a four-wheel drive system that drives both the front wheels 4 and the rear wheels 5.

[0019] Specifically, the traveling power transmission system 20 includes a transmission 30 that changes the speed of the rotational power output from the prime mover 11 and transmits it, a propeller shaft 22 that transmits the power output from the transmission 30, a front wheel differential 23 that distributes the power transmitted by the propeller shaft 22 to the left and right front wheels 4, a rear wheel differential 24 that distributes the power transmitted by the propeller shaft 22 to the left and right rear wheels 5, a front wheel drive shaft 25 that transmits the power distributed by the front wheel differential 23 to the front wheels 4, and a rear wheel drive shaft 26 that transmits the power distributed by the rear wheel differential 24 to the rear wheels 5. The prime mover 11 is, for example, composed of an engine alone. The transmission 30 has an input shaft 31 mechanically connected to the prime mover 11 and an output shaft 32 mechanically connected to the propeller shaft 22. 1 and 2, the propeller shaft 22 is configured to be able to bend in the left-right direction by connecting the front and rear propeller shafts 22a, 22b via a propeller shaft joint 22c. The prime mover 11 and the transmission 30 are housed in the engine room 3a, as shown in FIG.

[0020] As shown in Figure 2, a cargo handling hydraulic pump 13, which constitutes one element of a hydraulic system for operating the working implement 6, is mechanically connected to the prime mover 11. The cargo handling hydraulic pump 13 is a hydraulic source for the hydraulic system, and supplies pressure oil to hydraulic cylinders that drive the lift arm 7 and bucket 8 that constitute the working implement 6. As shown in Figure 1, the cargo handling hydraulic pump 13 is housed in the machine room 3a.

[0021] In the traveling power transmission system 20 configured in this manner, as shown in Fig. 2, a portion of the rotational power of the prime mover 11 is input to the cargo handling hydraulic pump 13, and the remainder is input to the transmission 30 via the input shaft 31. The rotational power, which has been changed to the desired torque and rotation speed by the transmission 30, is transmitted from the output shaft 32 of the transmission 30 via the propeller shaft 22 to the front wheel differential 23 and the rear wheel differential 24. The power distributed by the front wheel differential 23 is transmitted to the front wheels 4 via the front wheel drive shaft 25, and the power distributed by the rear wheel differential 24 is transmitted to the rear wheels 5 via the rear wheel drive shaft 26. As a result, a driving force Dt that counteracts the traveling load Lt acting on the front wheels 4 and rear wheels 5 is generated at the front wheels 4 and rear wheels 5, causing the wheel loader 1 to travel.

[0022] Next, the details of the configuration of the transmission in the traveling power transmission system for the work vehicle according to the first embodiment will be explained using Fig. 3. Fig. 3 is a schematic diagram showing the configuration of the transmission in the traveling power transmission system for the work vehicle according to the first embodiment shown in Fig. 2.

[0023] 3, the transmission 30 of the traveling power transmission system 20 is configured to be switchable between power transmission by a continuously variable transmission mechanism 35 and power transmission by a direct-coupled mechanism 36 that does not pass through the continuously variable transmission mechanism 35. The continuously variable transmission mechanism 35 is configured by a hydromechanical continuously variable transmission (hereinafter referred to as HMT). The direct-coupled mechanism 36 transmits the rotational power of the prime mover 11 by a mechanical mechanism without passing through the continuously variable transmission mechanism 35.

[0024] Specifically, the transmission 30 includes an input shaft 31 to which the rotational power of the prime mover 11 is input, an output shaft 32 that outputs the rotational power to the front wheels 4 and rear wheels 5 that generate the running load Lt, an HMT 35 provided between the input shaft 31 and the output shaft 32, and a direct-coupled mechanism 36 provided between the input shaft 31 and the output shaft 32. The power transmission path of the HMT 35 and the power transmission path of the direct-coupled mechanism 36 are in a parallel relationship with each other with respect to the input shaft 31 and the output shaft 32.

[0025] The input shaft 31 is mechanically connected to the prime mover 11 (drive shaft) and rotates when driven by the prime mover 11. A first input gear 61 (described later) of the direct-coupled mechanism 36 is provided on the input shaft 31. The output shaft 32 is mechanically connected to the propeller shaft 22 of the traveling power transmission system 20 and outputs power to the propeller shaft 22 to rotate and drive the front wheels 4 and rear wheels 5 that generate a traveling load Lt. The output shaft 32 is provided with an output gear 33 to which output from the HMT 35 is input.

[0026] The HMT 35 transmits a portion of the input rotational power to a mechanical mechanism, and transmits the remaining rotational power via a hydrostatic continuously variable transmission (hereinafter referred to as HST) rather than a mechanical mechanism, via a fluid. The HMT 35 includes, for example, a planetary gear mechanism 40 as a mechanical mechanism and an HST 50 that transmits power hydraulically. The HMT 35 of this embodiment is configured to divide the rotational power input from the prime mover 11 into two parts by the planetary gear mechanism 40, and to continuously switch the ratio of power transmission by the mechanical gears of the planetary gear mechanism 40 to power transmission by the HST 50 (hydraulic). The planetary gear mechanism 40 has an input side element (described in detail below) mechanically connected to the input shaft 31, and two output side elements (described in detail below) mechanically connected to the output shaft 32 and the input side elements (described in detail below) of the HST 50, respectively. The HST 50 has an input-side element (described in detail below) mechanically connected to one output-side element (described in detail below) of the planetary gear mechanism 40, and an output-side element (described in detail below) mechanically connected to the output shaft 32. In the HMT 35, the input rotational power is split by the planetary gear mechanism 40 into a path that goes to the HST 50 and a path that goes directly to the output shaft 32 without passing through the HST 50, and ultimately the power output from the planetary gear mechanism 40 and the power output from the HST 50 are combined (synthesized) at the output shaft 32.

[0027] The planetary gear mechanism 40 includes, for example, a first sun gear 41 and a second sun gear 42, a plurality of first planetary gears 43 that rotate while revolving around the central axis of the first sun gear 41, a plurality of second planetary gears 44 that rotate while revolving around the central axis of the second sun gear 42, and a carrier 45 that rotatably supports the first planetary gears 43 and the second planetary gears 44 and rotates around the central axes of the first sun gear 41 and the second sun gear 42. The carrier 45 is mechanically connected to the input shaft 31. The first sun gear 41 is mechanically connected to a pump input shaft 47 and is connected to an HST hydraulic pump 51 (described later) of the HST 50 via the pump input shaft 47. The second sun gear 42 is mechanically connected to a planetary output gear 48 and is connected to the output shaft 32 via the planetary output gear 48 and the output gear 33. The first planetary gear 43 and the second planetary gear 44 mesh with each other. That is, the first sun gear 41 and the second sun gear 42 mesh with each other via the first planetary gear 43 and the second planetary gear 44. The planetary gear mechanism 40 of the present embodiment divides and transmits the rotational power of the prime mover 11 input to the carrier 45 into two paths: one path is output to the HST 50 via the first sun gear 41 and the pump input shaft 47, and the other path is output to the output shaft 32 via the second sun gear 42 and the planetary output gear 48. The planetary gear mechanism 40 has a function of passively dividing the input rotational speed to the carrier 45 into two output rotational speeds: one path is output via the first sun gear 41, and the other path is output via the second sun gear 42.

[0028] The HST 50 includes an HST hydraulic pump 51 and an HST hydraulic motor 52, which are fluidly connected to each other. The HST hydraulic pump 51 and the HST hydraulic motor 52 are connected via a pair of pipes to form a hydraulic closed circuit. The HST hydraulic pump 51 is mechanically connected to the pump input shaft 47 of the planetary gear mechanism 40 and is rotationally driven by power output from the planetary gear mechanism 40 (first sun gear 41). The HST hydraulic pump 51 is a variable displacement hydraulic pump that can change its displacement volume (discharge volume per rotation, hereinafter referred to as pump volume) and has a regulator 51a that adjusts the pump volume. The HST hydraulic motor 52 is rotationally driven by pressure oil supplied from the HST hydraulic pump 51. The HST hydraulic motor 52 is mechanically connected to a motor output shaft 53 and outputs power hydraulically transmitted from the HST hydraulic pump 51 via the motor output shaft 53 to the output shaft 32. The HST hydraulic motor 52 is configured as a variable displacement hydraulic motor that can change the displacement volume (discharge volume per rotation, hereinafter referred to as motor volume). The HST hydraulic motor 52 has a regulator 52a that adjusts the motor volume.

[0029] The regulator 51 a of the HST hydraulic pump 51 and the regulator 52 a of the HST hydraulic motor 52 are configured to adjust the pump volume and motor volume in response to commands from the transmission control device 80, and are configured to be able to continuously change the pump volume and motor volume within ranges from zero to maximum volume. The gear ratio of the HST 50 is determined according to the ratio between the pump volume of the HST hydraulic pump 51 and the motor volume of the HST hydraulic motor 52. When the pump volume is larger than the motor volume, the gear ratio of the HST 50 is in the direction of increasing speed (a value larger than 1). On the other hand, when the pump volume is smaller than the motor volume, the gear ratio of the HST 50 is in the direction of decreasing speed (a value smaller than 1). By continuously adjusting the pump volume and motor volume, the HST 50 outputs the power input from the planetary gear mechanism 40 to the HST hydraulic pump 51 from the HST hydraulic motor 52 to the output shaft 32 in a manner that allows the speed to be continuously changed.

[0030] The hydraulic closed circuit of the HST 50 includes a relief circuit and a charge circuit. The relief circuit prevents abnormal pressure increases in the hydraulic closed circuit and includes a pair of main relief valves 56 for releasing pressurized oil in the hydraulic closed circuit to a tank 55. The main relief valves 56 are configured to open when the pressure in the hydraulic closed circuit exceeds a set value. The charge circuit replenishes pressurized oil that leaks from the hydraulic closed circuit. The charge circuit includes, for example, a charge pump 57 that supplies pressurized oil to the hydraulic closed circuit, a pair of check valves 58 that prevent flow from the hydraulic closed circuit to the charge pump 57, and a charge relief valve 59 that releases pressurized oil discharged by the charge pump 57 to the tank 55.

[0031] The direct-coupled mechanism 36 includes a first input gear 61 provided on the input shaft 31, a second input gear 62 meshing with the first input gear 61, a direct-coupled shaft 63 connected to the second input gear 62, and a lock-up clutch 64 provided between the direct-coupled shaft 63 and the output shaft 32. The lock-up clutch 64 is configured to be switchable between an engaged state in which the direct-coupled shaft 63 and the output shaft 32 are connected and a disengaged state in which the direct-coupled shaft 63 and the output shaft 32 are disengaged. That is, when the lock-up clutch 64 is in the engaged state, the power transmission path of the direct-coupled mechanism 36 is in a connected state, and the rotational power transmitted to the direct-coupled shaft 63 via the first input gear 61 and the second input gear 62 is transmitted to the output shaft 32. On the other hand, when the lock-up clutch 64 is in the disengaged state, the power transmission path of the direct-coupled mechanism 36 is in a disconnected state, and the rotational power transmitted to the direct-coupled shaft 63 is not transmitted to the output shaft 32. The lock-up clutch 64 is configured so that the pressing pressure that enables switching between the engaged state and the released state is controlled by a command from the transmission control device 80 .

[0032] In the power transmission path of the direct-coupled mechanism 36, which is a mechanical mechanism, the gear ratio (the ratio of the output rotation speed of the direct-coupled shaft 63 to the input rotation speed of the first input gear 61) is fixed. The power transmission path of the direct-coupled mechanism 36 bypasses the power transmission path of the HMT 35, and the gear ratio of the direct-coupled mechanism 36 can be set regardless of the speed range of the HMT 35. Therefore, in order to expand the speed range of the transmission 30, the gear ratio of the direct-coupled mechanism 36 is set to a gear ratio higher than the maximum speed gear ratio of the HMT 35.

[0033] A first speed sensor 38 is provided on the input shaft 31 of the transmission 30 to detect the rotation speed of the input shaft 31. The rotation speed of the input shaft 31, which is the detected value of the first speed sensor 38, corresponds to the rotation speed of the engine serving as the prime mover 11. In other words, the first speed sensor 38 functions as a speed sensor that detects the rotation speed of the engine 11. The first speed sensor 38 outputs a detection signal corresponding to the detected value (the rotation speed of the engine 11) to the engine control device 100.

[0034] The engine control device 100 controls the operation of the engine 11 and is configured to receive detection information from the first speed sensor 38. For example, the engine control device 100 controls the operation of the engine 11 based on the detection information from the first speed sensor 38 to maintain a constant engine speed depending on the driving conditions. The engine control device 100 can estimate engine torque based on the engine speed (detected value of the first speed sensor 38) and the fuel injection amount. When the engine control device 100 detects a decrease in engine speed from the detection information from the first speed sensor 38, it controls the engine 11 to increase the fuel injection amount to maintain the engine speed (prevent stall). The engine control device 100 outputs the detection information from the first speed sensor 38 and control information for the engine 11 to the gear change control device 80 (described below), and is configured to input commands to the engine 11 from the gear change control device 80.

[0035] A second speed sensor 39 is installed on the output shaft 32 of the transmission 30 to detect the rotation speed of the output shaft 32. The rotation speed of the output shaft 32, which is the detected value of the second speed sensor 39, is used to calculate the vehicle speed of the wheel loader 1. In other words, the second speed sensor 39 functions as a vehicle speed sensor that detects information related to the vehicle speed of the wheel loader 1. The second speed sensor 39 outputs a detection signal corresponding to the detected value (the rotation speed of the output shaft 32) to the transmission control device 80.

[0036] Next, the power transmission path in the transmission of the work vehicle according to the first embodiment will be described. First, the power transmission path of the HMT in the transmission will be described using Figures 3 to 5. Figure 4 is an explanatory diagram showing the power transmission path when the gear ratio of the HMT in the transmission of the traveling power transmission system of the work vehicle according to the first embodiment shown in Figure 3 is at the lowest speed. Figure 5 is an explanatory diagram showing the power transmission path when the gear ratio of the HMT in the transmission of the traveling power transmission system of the work vehicle according to the first embodiment shown in Figure 3 is at the highest speed.

[0037] 3, the planetary gear mechanism 40 has a function of passively dividing one input rotation speed into two output rotation speeds, and divides the power transmission path into one that goes to the HST 50 (HST hydraulic pump 51) via the first sun gear 41 and one that goes to the output shaft 32 via the second sun gear 42 without passing through the HST 50. Therefore, when the lock-up clutch 64 is in a disengaged state (when the power transmission path of the direct drive mechanism 36 is in a disconnected state), the gear ratio of the transmission 30 changes continuously in accordance with the gear ratio of the HST 50, which is continuously variable.

[0038] When the reduction ratio of the HST 50 is set to infinity, in other words, when the speed increase ratio of the HST 50 is set to zero, the output rotation speed of the HST 50, i.e., the rotation speed of the HST hydraulic motor 52, must be zero regardless of the input rotation speed to the HST 50, i.e., the rotation speed of the HST hydraulic pump 51. In this case, the speed change ratio of the HMT 35 (transmission 30) is the lowest speed, and the vehicle is in a stopped state. When the vehicle is completely stopped, no power is transmitted, but when the speed increase ratio of the HST 50 is a very small value other than zero, that is, when the reduction ratio of the HMT 35 (transmission 30) is an extremely large finite value, as shown in FIG. 4 , the power transmission path is such that almost all of the power input to the HMT 35 (transmission 30) passes through the HST 50.

[0039] Conversely, when the reduction ratio of the HST 50 is set to zero, in other words, when the speed increase ratio of the HST 50 is set to infinity, the rotational speed of the HST hydraulic pump 51, which is the input rotational speed to the HST 50, must be zero, regardless of the rotational speed of the HST hydraulic motor 52, which is the output rotational speed of the HST 50. When the rotational speed of the HST hydraulic pump 51 is zero, the HST 50 does not receive power from the planetary gear mechanism 40, as shown in FIG. 5 , so the only power transmission path is one in which all of the power input to the HMT 35 (planetary gear mechanism 40) goes directly to the output shaft 32 via the second sun gear 42 without passing through the HST 50. In this case, because the speed increase ratio of the HST 50 is infinity, the speed ratio of the HMT 35 (transmission 30) becomes the highest speed within the range of continuously variable transmission.

[0040] For these reasons, when the pump displacement of the HST hydraulic pump 51 in the HST 50 is set to zero and the motor displacement of the HST hydraulic motor 52 is set to the maximum, the gear ratio of the HMT 35 (transmission 30) becomes the minimum speed. On the other hand, when the pump displacement of the HST hydraulic pump 51 is set to the maximum and the motor displacement of the HST hydraulic motor 52 is set to zero, the gear ratio of the HMT 35 becomes the maximum speed within the range of continuously variable transmission.

[0041] Secondly, the power transmission path of the direct drive mechanism in the transmission will be described using Figures 5 and 6. Figure 6 is an explanatory diagram showing the power transmission path of the direct drive mechanism in the transmission of the traveling power transmission system for the work vehicle according to the first embodiment shown in Figure 3 (highest gear ratio of the transmission).

[0042] In the transmission 30, the pump displacement of the HST hydraulic pump 51 and the motor displacement of the HST hydraulic motor 52 of the HST 50 are both set to zero to interrupt power transmission via the HST 50 and engage the lock-up clutch 64. As a result, the power transmission path of the transmission 30 becomes the power transmission path of the direct-coupled mechanism 36, as shown in FIG. 6 , in which power input to the transmission 30 passes from the first input gear 61 and the second input gear 62 to the output shaft 32 via the direct-coupled shaft 63 and the engaged lock-up clutch 64 without passing through the HMT 35 (both the planetary gear mechanism 40 and the HST 50). The speed ratio of the direct-coupled mechanism 36 is a fixed value corresponding to the gear ratio of the first input gear 61 and the second input gear 62, and is set to a speed ratio higher than the maximum speed within the continuously variable transmission range of the HMT 35.

[0043] As described above, in the transmission 30, the motor displacement of the HST hydraulic motor 52 is set to zero whether the power transmission path is one in which the HMT 35 is at the highest continuously variable speed ratio (see FIG. 5 ) or one in which the direct-coupled mechanism 36 has a speed ratio even higher than the maximum speed of the HMT 35 (see FIG. 6 ). Meanwhile, the pump displacement of the HST hydraulic pump 51 is maximum in the former case and zero in the latter case. To switch the power transmission path between the HMT 35 power transmission path and the direct-coupled mechanism 36 power transmission path when the HMT 35 has the highest continuously variable speed ratio, the pump displacement of the HST hydraulic pump 51 must be changed between maximum displacement and zero. Changing the pump displacement between maximum displacement and zero takes a time period of less than one second to several seconds. Therefore, when the pump displacement is changed between maximum displacement and zero, a transitional state of dual power transmission occurs in which power is transmitted by the HMT 35 and the direct-coupled mechanism 36 simultaneously. When the power transmission of the transmission 30 is in a transient state between the HMT 35 and the direct drive mechanism 36, a gear shift shock occurs, which may lead to engine stall in some cases. Therefore, it is necessary to control the timing for increasing or decreasing the pump volume of the HST hydraulic pump 51 in conjunction with the timing for engaging or disengaging the lock-up clutch 64 of the direct drive mechanism 36.

[0044] When the power transmission path of the transmission 30 is switched from the power transmission path of the direct drive mechanism 36 to the power transmission path of the HMT 35 with the highest continuously variable speed ratio, i.e., when the engaged lock-up clutch 64 is disengaged, the gear ratio of the transmission 30 is switched to the lower speed side. At this time, a load torque acts in a direction that increases the rotation speed of the engine 11, so engine stall does not occur, and engine braking becomes a cause of gear shift shock. In this case, gear shift shock can be suppressed by the engine control device 100 increasing the fuel injection amount of the engine 11 to an extent that offsets the engine brake torque. In other words, when the power transmission path of the transmission 30 is switched from the direct drive mechanism 36 to the HMT 35, gear shift shock can be suppressed simply by controlling the drive of the engine 11.

[0045] On the other hand, when switching from the power transmission path in the maximum speed state of the continuously variable transmission in the HMT 35 to the power transmission path of the direct drive mechanism 36, i.e., when engaging the disengaged lock-up clutch 64, the gear ratio of the transmission 30 switches to the high-speed side. At this time, a load torque acts in a direction that reduces the engine speed. In this case, the above-mentioned problem cannot be solved by controlling the engine 11 alone. Therefore, this embodiment focuses on control of the latter switching in the transmission 30.

[0046] Next, problems that arise when switching the power transmission path from HMT (highest speed state of continuously variable transmission) to the direct drive mechanism in a transmission will be described using a switching method of a comparative example.

[0047] Figure 7 is a time chart showing an example of the transitions in the state of the lock-up clutch 64, the HST hydraulic pump 51, the engine 11, and the vehicle running state when the power transmission path is switched using a switching method of the comparative example for the transmission 30 shown in Figure 3. Figure 7 shows an example in which a gear shift shock occurs using the switching method of the comparative example. In Figure 7, the top row shows the transitions in the pressing pressure of the lock-up clutch 64 of the direct-coupled mechanism 36, the second row shows the transitions in the pump volume of the HST hydraulic pump 51 of the HMT 35, the third row shows the transitions in the rotation speed of the engine 11, the fourth row shows the transitions in the vehicle speed of the wheel loader 1, and the bottom row shows the transitions in the output torque of the engine 11.

[0048] In FIG. 7 , time region A corresponds to the time when the power transmission path of the transmission 30 is the power transmission path of the HMT 35, which is the continuously variable maximum speed power transmission path (see FIG. 5 ). Meanwhile, time region C corresponds to the time when the power transmission path of the transmission 30 is the power transmission path of the direct drive mechanism 36 (see FIG. 6 ). Time region B, which is between time region A and time region C, corresponds to a transient state in which the power transmission path transitions from the continuously variable maximum speed power transmission path of the HMT 35 to the power transmission path of the direct drive mechanism 36. In the switching method of the comparative example, as shown in the third row of FIG. 7 , it is assumed that control is executed to maintain the rotation speed of the engine 11 constant. Note that, when switching the power transmission path from the HMT 35 to the direct drive mechanism 36, the motor displacement of the HST hydraulic motor 52 is zero in all of time regions A, B, and C, and therefore the motor displacement of the HST hydraulic motor 52 is not shown in FIG. 7 .

[0049] In time region A, the power transmission path of the transmission 30 is the highest speed power transmission path of the continuously variable transmission in the HMT 35, so the lock-up clutch 64 is in a released state with zero pressing pressure, and the pump volume of the HST hydraulic pump 51 is at its maximum. Because the engine speed is constant and the gear ratio of the transmission 30 is also constant, the vehicle speed of the wheel loader 1 is also constant.

[0050] In time region C, the power transmission path of the transmission 30 has been completely switched to the power transmission path of the direct drive mechanism 36, so the pump volume of the HST hydraulic pump 51 is zero and the lock-up clutch 64 is in an engaged state with maximum pressing pressure. When the pump volume of the HST hydraulic pump 51 is zero, the HMT 35 no longer transmits power via either the HST 50 or the planetary gear mechanism 40.

[0051] In order to prevent the power transmission of the transmission 30 from being interrupted during the transition of the power transmission path of the HMT 35 shown in FIG. 5 to the power transmission path of the direct-coupled mechanism 36 shown in FIG. 6 (when the pump displacement of the HST hydraulic pump 51 is changed from the maximum displacement to zero), the lock-up clutch 64 must be pressed before the pump displacement is set to zero. Therefore, in time region B during the transition from the power transmission path of the HMT 35 shown in FIG. 5 to the power transmission path of the direct-coupled mechanism 36 shown in FIG. 6, the lock-up clutch 64 is first pressed, as shown in the top row of FIG. 7. Then, as shown in the second row of FIG. 7, the pump displacement of the HST hydraulic pump 51 begins to decrease from the maximum displacement and is ultimately set to zero. Even if the pump displacement starts to decrease before the lock-up clutch 64 is pressed, the power transmission of the transmission 30 will not be completely interrupted unless the pump displacement is set to zero before the lock-up clutch 64 is pressed. However, when the pump displacement of the HST hydraulic pump 51 is reduced, the discharge pressure of the HST hydraulic pump 51 increases relative to the input torque. For this reason, even when a relatively small torque is input to the HST hydraulic pump 51, the main relief valve 56 of the HST 50 opens and the pressure in the hydraulic closed circuit of the HST 50 escapes. In this case, the torque capacity of the HMT 35 decreases, and sufficient power transmission may not necessarily be possible. Therefore, in order to transmit the required power even when the load is heavy, it is necessary to press the lock-up clutch 64 before starting to reduce the pump displacement of the HST hydraulic pump 51.

[0052] However, even if the pressing pressure of the lock-up clutch 64 is increased to its maximum when the pump volume of the HST hydraulic pump 51 is at its maximum volume, the torque capacity of the HST 50 is greater than the torque capacity of the lock-up clutch 64, so the lock-up clutch 64 will be in a slipping state. For this reason, the power transmission path of the transmission 30 will be maintained as the highest speed power transmission path for the continuously variable speed in the HMT 35. As a result, as shown in the fourth row of Fig. 7, the vehicle speed of the wheel loader 1 does not change. At this time, as shown in the bottom row of Fig. 7, the engine torque increases by the amount that the lock-up clutch 64 slips.

[0053] Here, if the pump displacement of the HST hydraulic pump 51 is reduced from maximum while the lockup clutch 64 remains in a pressing state (maximum pressing pressure), the torque capacity of the HST 50 decreases. As a result, slippage of the lockup clutch 64 decreases, and the torque transmission path also shifts to the lockup clutch 64 side of the direct-coupled mechanism 36. If the pump displacement is reduced to zero, the lockup clutch 64 enters an engaged state where it no longer slips, and the torque transmission path shifts completely from the HMT 35 to the lockup clutch 64 side, with the direct-coupled mechanism 36 taking over power transmission. As a result, the gear ratio of the transmission 30 switches to the gear ratio of the direct-coupled mechanism 36, which is higher than the maximum speed of the continuously variable transmission of the HMT 35. When the gear ratio of the transmission 30 switches to the higher speed side, a driving force Dt (see FIG. 2 ) acts in a direction that accelerates the vehicle speed of the wheel loader 1.

[0054] When the torque capacity of the HST 50 falls below the torque capacity of the lock-up clutch 64, the power transmission path of the transmission 30 suddenly shifts from the HMT 35 to the direct-coupled mechanism 36 (the lock-up clutch 64 side). The engine output torque, which has increased due to slippage of the lock-up clutch 64, is further increased when the running load Lt of the front wheels 4 and rear wheels 5 due to vehicle acceleration is suddenly applied to the engine 11. At this time, as shown in the bottom row of FIG. 7, if the engine torque has not reached the upper limit torque for that rotation speed, the engine will not stall. However, as shown in the fourth row of FIG. 7, the vehicle 1 will suddenly accelerate. This acceleration can cause gear shift shock.

[0055] Furthermore, Figure 8 is a time chart showing another example of transitions in the states of the lock-up clutch, HST hydraulic pump, engine, and vehicle running when switching the power transmission path using a switching method of a comparative example for the transmission shown in Figure 3. In Figure 8, the top, second, third, and fourth rows show transitions of the same items as in Figure 7. Figure 8 shows an example in which an engine stall occurs when switching the transmission using the switching method of the comparative example with the engine speed set lower than in the case of Figure 7.

[0056] Time domains A, B, and C in Figure 8 are also similar to those shown in Figure 7. That is, in time domain A, the lock-up clutch 64 of the direct drive mechanism 36 is disengaged, and the power transmission path of the transmission 30 is in a state of the highest-speed power transmission path of the continuously variable transmission in the HMT 35 (see Figure 5). In time domain C, the pump displacement of the HST hydraulic pump 51 is zero, and the power transmission path of the transmission 30 is completely switched to the power transmission path of the direct drive mechanism 36. In time domain B, the power transmission path of the transmission 30 is in a transient state (dual power transmission state) shifting from the HMT 35 to the direct drive mechanism 36. In Figure 8 as well, the motor displacement of the HST hydraulic motor 52 is zero in all of time domains A, B, and C, and therefore the motor displacement of the HST hydraulic motor 52 is not shown.

[0057] In the gear shift switching method of this comparative example, the control of the rotation speed of the engine 11 is different from that shown in Fig. 7. Specifically, as shown in the third row of Fig. 8, it is assumed that the rotation speed of the engine 11 is controlled to be kept constant and lower than that in Fig. 7 within a range that does not cause stalling in order to save fuel.

[0058] In time region B, the switching control of the power transmission path of the transmission 30 is performed in the same manner as in the case shown in Fig. 7. That is, as shown in the top diagram of Fig. 8, first, the lock-up clutch 64 of the direct-coupled mechanism 36 is pressed with maximum pressing pressure. Then, as shown in the second diagram of Fig. 8, the pump displacement of the HST hydraulic pump 51 begins to decrease from maximum displacement while the lock-up clutch 64 remains pressed. At this time, as shown in the bottom diagram of Fig. 8, the engine output torque increases due to slippage of the lock-up clutch 64.

[0059] Furthermore, the pump displacement of the HST hydraulic pump 51 is rapidly reduced to zero. This reduces the torque capacity of the HST 50, preventing the lock-up clutch 64 from slipping. As a result, the torque transmission path shifts from the HMT 35 to the lock-up clutch 64 side, and the direct-coupled mechanism 36 takes over power transmission, switching the gear ratio of the transmission 30 to that of the direct-coupled mechanism 36. When the gear ratio of the transmission 30 switches to the high-speed side, a driving force Dt acts in a direction that accelerates the vehicle speed of the wheel loader 1.

[0060] At this time, the load torque for accelerating the vehicle is superimposed on the load torque caused by the running load and the load torque caused by slippage of the lock-up clutch 64, in a direction that reduces the rotational speed of the engine 11. For this reason, if the engine rotational speed is maintained at a low speed close to the limit corresponding to the running load in order to save fuel, the output torque of the engine will reach the upper limit of the rotational speed, causing the engine to stall, as shown in the bottom part of Figure 8.

[0061] To prevent this engine stall, it is possible to automatically execute control to increase the engine speed immediately before switching to the power transmission path of the direct drive mechanism 36. However, control that automatically increases the engine speed may result in unintended sudden acceleration. In other words, there is a concern that increasing the engine speed itself may result in gear shift shock.

[0062] It is possible to prevent engine stall by always setting the engine speed at a relatively high speed as shown in Figure 7, and to suppress gear shift shock by lowering the engine speed that has been set at a high speed. However, a general characteristic of the engine 11 is that the lower the speed in the low speed range, the better the fuel economy, so it is not desirable to increase the engine speed and increase fuel consumption just to switch to the power transmission path of the direct drive mechanism 36.

[0063] Therefore, the speed change control device 80 (see FIG. 3) of the wheel loader 1 according to this embodiment appropriately controls the reduction in pump volume of the HST hydraulic pump 51 while maintaining the rotational speed of the engine 11 at a low rotational speed, thereby avoiding the occurrence of engine stall when switching the power transmission path from the HMT 35 to the direct drive mechanism 36 in the transmission 30 and suppressing speed change shock.

[0064] Next, the hardware and functions of the gear change control device for the work vehicle according to the first embodiment will be described with reference to Fig. 9. Fig. 9 is a block diagram showing the configuration of the hardware and functional parts of the gear change control device for the work vehicle according to the first embodiment shown in Fig. 3.

[0065] 9, the gear change control device 80 is configured to perform control to change the gear ratio of the transmission 30. The gear change control device 80 has, as its hardware configuration, a storage device 81 made up of, for example, RAM, ROM, etc., and a processing device 82 made up of a CPU, MPU, etc. The storage device 81 pre-stores programs and various information required to perform control to change the gear ratio of the transmission 30. The processing device 82 reads the programs and various information from the storage device 81 as appropriate, and performs processing in accordance with the programs to realize various functions.

[0066] The shift control device 80 according to this embodiment is characterized by a control method used when switching the power transmission path of the transmission 30 from the HMT 35 (the highest speed state of the continuously variable transmission) to the direct-coupled mechanism 36. When switching the power transmission path of the transmission 30 from the HMT 35 to the direct-coupled mechanism 36, the shift control device 80 controls the pump displacement of the HST hydraulic pump 51 while keeping the lock-up clutch 64 pressed, as in the case of the switching control method of the comparative example described above. However, the manner in which the pump displacement is reduced differs from the switching control method of the comparative example. In summary, the shift control device 80 temporarily stops the reduction of the pump displacement and maintains the pump displacement constant. Thereafter, when a predetermined condition is satisfied (when the vehicle speed is equal to or greater than a predetermined value), the shift control device 80 resumes the reduction of the pump displacement, reducing it until the pump displacement reaches zero. This reduces shift shock and prevents engine stalls when the power transmission path transitions from the HMT 35 to the direct-coupled mechanism 36.

[0067] Note that the transmission control device 80 controls the motor displacement of the HST hydraulic motor 52 to maintain zero, as in the case of the switching control method of the comparative example. Also, in this embodiment, it is assumed that the engine control device 110 executes control to maintain the rotation speed of the engine 11 at a predetermined constant when the power transmission path of the transmission 30 is switched. However, the setting of the rotation speed of the engine 11 can be changed depending on the traveling conditions of the wheel loader 1. In other words, if the traveling load increases during switching control of the transmission 30 and the engine control device 100 detects a decrease in engine rotation speed, it controls the engine to increase the fuel injection amount to prevent engine stall.

[0068] The gear change control device 80 has, for example, as functional units that perform this switching control of the power transmission path of the transmission 30, a torque margin determination unit 91 that determines the engine torque margin until the occurrence of an engine stall, a restart timing determination unit 92 that determines the timing to restart reducing the pump volume of the HST hydraulic pump 51, and a transmission control unit 93 that controls the transmission 30. The transmission control unit 93 has a pump volume control unit 95 that controls the pump volume of the HST hydraulic pump 51, a motor volume control unit 96 that controls the motor volume of the HST hydraulic motor 52, and a clutch control unit 97 that controls the lock-up clutch 64.

[0069] The torque margin determination unit 91 is configured to determine, for example, based on an estimated value Te of the output torque of the engine 11 acquired from the engine control device 100, whether or not there is a margin of engine torque relative to the upper limit torque (engine specifications) that can be output by the engine 11 corresponding to the engine speed at the time of determination (when the transmission 30 is being switched). Specifically, the torque margin determination unit 91 determines whether or not the difference between the upper limit torque corresponding to the engine speed at the time of determination and the estimated value Te of the engine torque from the engine control device 100 exceeds a predetermined first threshold. If the difference exceeds the first threshold, it is determined that there is a margin of output torque of the engine 11 relative to the upper limit torque at which engine stall may occur. On the other hand, if the difference is equal to or less than the first threshold, it is determined that there is no margin of output torque of the engine 11 relative to the upper limit torque at which engine stall may occur. The torque margin determination unit 91 outputs the determination result of whether or not there is a margin of engine torque relative to the occurrence of engine stall to the transmission control unit 93.

[0070] The first threshold value is set to a value that provides a margin such that the engine output torque does not exceed the upper limit torque of the engine 11, even when the road load increases due to climbing or external disturbances. If the road load increases during a transient state in which the power transmission path of the transmission 30 transitions from the HMT 35 to the direct drive mechanism 36, the engine control device 100 detects a decrease in engine speed and increases the fuel injection amount to prevent engine stall. Therefore, the first threshold value is set so that a margin is maintained up to the upper limit torque (engine stall) so that the engine control device 100 can respond in time to the expected rate of increase in the road load. The first threshold value is pre-stored in, for example, the storage device 81.

[0071] In this embodiment, the torque margin determination unit 91 makes a determination based on an estimated value of engine torque acquired from the engine control device 100. However, the torque margin determination unit 91 may also make a determination based on a value detected by a torque sensor capable of detecting engine torque. The engine control device 100 can estimate the output torque of the engine 11 based on the actual engine speed of the engine 11 detected by the first speed sensor 38 and the amount of fuel injected into the engine 11.

[0072] The restart timing determination unit 92 determines whether or not it is appropriate to resume the reduction of the pump displacement of the HST hydraulic pump 51 after the reduction has been temporarily stopped. The restart timing determination unit 92 determines the timing to resume the reduction of the pump displacement after the reduction has been temporarily stopped, for example, by determining whether the actual vehicle speed of the wheel loader 1 has reached a predetermined value that is slightly lower than the vehicle speed corresponding to the gear ratio when the power transmission path is the direct drive mechanism 36. The actual vehicle speed of the wheel loader 1 can be calculated based on the actual rotation speed of the output shaft 32 of the transmission 30 detected by the second speed sensor 39. If the actual vehicle speed of the wheel loader 1 is equal to or greater than the predetermined value, it is determined that the transition of the power transmission path of the transmission 30 to the direct drive mechanism 36 has progressed sufficiently, and that the reduction of the pump displacement can be resumed. On the other hand, if the actual vehicle speed of the wheel loader 1 is less than the predetermined value, it is determined that the transition of the power transmission path of the transmission 30 to the direct drive mechanism 36 has not been sufficient, and that the reduction of the pump displacement is not yet ready to be resumed. The restart timing determination unit 92 outputs the determination result as to whether or not it is appropriate to restart the reduction of the pump displacement to the transmission control unit 93 .

[0073] When the power transmission path of the transmission 30 is switched from the HMT 35 to the direct drive mechanism 36, the pump volume control unit 95 of the transmission control unit 93 controls the pump volume of the HST hydraulic pump 51 to decrease from a state immediately before the control to a predetermined volume (zero). Specifically, the pump volume control unit 95 temporarily stops the decrease in the pump volume midway based on the determination result of the torque margin determination unit 91, and maintains the pump volume constant. Thereafter, the pump volume control unit 95 resumes the decrease in the pump volume based on the determination result of the restart timing determination unit 92, and decreases the pump volume until it becomes zero. Note that when the power transmission path of the transmission 30 is switched from the highest-speed power transmission path of the continuously variable transmission in the HMT 35 to the power transmission path of the direct drive mechanism 36, the pump volume of the HST hydraulic pump 51 begins to decrease from its maximum volume. Furthermore, when the torque surplus determination unit 91 determines that there is no margin in the engine output torque relative to the upper limit torque of the engine 11 corresponding to the engine speed during the control (when the difference between the upper limit torque and the output torque is equal to or less than the first threshold value), the pump volume control unit 95 controls the pump volume so that it is maintained constant without decreasing it or is increased.

[0074] The motor volume control unit 96 of the transmission control unit 93 controls the motor volume of the HST hydraulic motor 52 to be maintained at zero when the power transmission path of the transmission 30 is switched from the HMT 35 to the direct drive mechanism 36. Furthermore, when the torque surplus determination unit 91 determines that there is no margin in the output torque relative to the upper limit torque of the engine 11 (the difference between the upper limit torque and the output torque is equal to or less than the first threshold value), the motor volume control unit 96 can also perform control to increase the motor volume.

[0075] The clutch control unit 97 of the transmission control unit 93 controls the pressing pressure of the lock-up clutch 64. Specifically, the clutch control unit 97 is configured to increase the pressing pressure of the lock-up clutch 64 to a maximum value when the power transmission path of the transmission 30 is switched to the direct drive mechanism 36. Furthermore, when the power transmission path of the transmission 30 is switched to the direct drive mechanism 36, if the torque surplus determination unit 91 determines that there is no margin in the output torque relative to the upper limit torque of the engine (the difference between the upper limit torque and the output torque is equal to or less than a first threshold), the clutch control unit 97 is configured to reduce the pressing pressure of the lock-up clutch 64 to a minimum. Note that the clutch control unit 97 of this embodiment does not need to precisely control the pressing pressure of the lock-up clutch 64, and therefore can be configured to perform two-value control of a pressurized state and a depressurized state.

[0076] Next, the processing procedure for controlling the switching of the power transmission path of the transmission from the HMT to the direct drive mechanism by the control device for the work vehicle according to the first embodiment, and the effects of this switching control, will be described with reference to FIGS. 3 and 10 to 12. FIG. 10 is a flowchart showing an example of the processing procedure for controlling the transmission (controlling the switching of the power transmission path from the HMT to the direct drive mechanism) by the gear change control device for the work vehicle according to the first embodiment shown in FIG. 9. FIG. 11 is a time chart showing an example of transitions in the states of the lock-up clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path of the transmission (switching from the HMT to the direct drive mechanism) in the work vehicle according to the first embodiment. FIG. 12 is a time chart showing another example of transitions in the states of the lock-up clutch, the HST hydraulic pump, the engine, and vehicle running when switching the power transmission path of the transmission (switching from the HMT to the direct drive mechanism) in the work vehicle according to the first embodiment (an example in which the running load increases midway).

[0077] In summary, the control of switching the power transmission path of the transmission 30 by the shift control device 80 shown in FIG. 10 involves pressing the lock-up clutch 64 of the direct-coupled mechanism 36 before starting to reduce the pump displacement of the HST hydraulic pump 51 (step S10). Thereafter, the pump displacement of the HST hydraulic pump 51 is reduced (steps S30 and S80). However, the reduction in pump displacement is temporarily stopped midway (step S50), and when the vehicle speed of the wheel loader 1 reaches a predetermined value (step S70), the reduction in pump displacement is resumed (step S80). However, if the rotational speed of the engine 11 decreases and is not maintained constant due to an overload during this control, the pump displacement of the HST hydraulic pump 51 is increased (step S90). Even in the processing of step S90, if there is insufficient margin for output torque relative to the upper limit torque of the engine 11 at which engine stall is a concern, the lock-up clutch 64 is released (step S100).

[0078] Time domains A, B, and C in FIG. 11 are the same as those shown in FIG. 7 . That is, in time domain A, the lock-up clutch 64 of the direct-coupled mechanism 36 is disengaged, and the power transmission path of the transmission 30 is in the state of the highest-speed power transmission path of the continuously variable transmission in the HMT 35 (see FIG. 5 ). In time domain C, the pump displacement of the HST hydraulic pump 51 is zero, and the power transmission path of the transmission 30 is completely switched to the direct-coupled mechanism 36. Time domain B is a transient state (dual power transmission state) in which the power transmission path of the transmission 30 shifts from the HMT 35 to the direct-coupled mechanism 36. That is, time domain B shows the changes in the state of the transmission 30 and the engine 11 when the gear change control device 80 goes through steps S10 to S80 (excluding steps S90 and S100) of the processing procedure of the flowchart shown in FIG. 10 . In FIG. 11 as well, the motor displacement of the HST hydraulic motor 52 is zero in all of the time regions A, B, and C, and therefore the motor displacement of the HST hydraulic motor 52 is not shown.

[0079] 12 is the same as the time region A in FIG. 11 . Time region B and time region C are different from the time region B and time region C in FIG. 11 . Time region B shows a case where switching of the power transmission path of the transmission 30 to the direct drive mechanism 36 is canceled due to an increase in the driving load when the power transmission path is switched to the direct drive mechanism 36 side. That is, time region B shows the change in the state of the transmission 30 and the engine 11 when the gear change control device 80 goes through steps S90 and S100 after processing steps S10 to S60 of the processing procedure of the flowchart shown in FIG. 10 . Time region C shows a state where the pump volume of the HST hydraulic pump 51 is at its maximum volume and the pressing pressure of the lock-up clutch 64 is zero, and the power transmission path of the transmission 30 has returned to the original, infinitely variable, maximum speed power transmission path of the HMT 35 (see FIG. 5 ).

[0080] During the switching control of the transmission 30 by the shift control device 80, if the reduction in pump capacity of the HST hydraulic pump 51 is temporarily stopped midway (see step S50 in FIG. 10 and the second row in FIG. 11), the lock-up clutch 64 shown in FIG. 3 will be in a slipping state, but the slippage will gradually decrease and the lock-up clutch 64 will be in a state of transmitting power. As a result, the vehicle speed of the wheel loader 1 will gradually increase (see the fourth row in FIG. 11). By waiting for this gradual increase in vehicle speed, it is possible to avoid a sudden shift to the power transmission path from the HMT 35 to the direct drive mechanism 36 and the accompanying sudden switch to the high-speed gear ratio.

[0081] 7 and 8, the power transmission path of the transmission 30 will suddenly shift from the HMT 35 to the direct drive mechanism 36 when the torque capacity of the HST 50 shown in FIG. 3 falls below the torque capacity of the lock-up clutch 64. Therefore, in order to prevent a sudden shift in the power transmission path of the transmission 30, it is not necessary to take time for the entire process of reducing the pump capacity of the HST hydraulic pump 51 from maximum to zero, but it is necessary to take time to maintain the dual power transmission state within a range that avoids the occurrence of an engine stall.

[0082] However, when the transmission 30 is in a dual power transmission state, energy loss occurs due to slippage of the lock-up clutch 64 and energy loss due to relief of the HST 50 (opening of the main relief valve 56). Therefore, it is desirable to shorten the switching time of the power transmission path of the transmission 30 within a range that allows a sudden shift in the power transmission path of the transmission 30 to be avoided.

[0083] Therefore, the pump displacement of the HST hydraulic pump 51 is quickly reduced until the torque capacity of the HST 50 becomes approximately the same as the torque capacity of the lock-up clutch 64, and then the pump displacement is temporarily stopped. Thereafter, when the lock-up clutch 64 no longer slips, the reduction of the pump displacement is resumed and the pump displacement is quickly reduced to zero. However, even when the reduction of the pump displacement of the HST hydraulic pump 51 is stopped midway, if the pump displacement is reduced too much before the pump is temporarily stopped, there is a concern that the engine may stall, as in the case shown in FIG. 8 . Therefore, the extent to which the reduction of the pump displacement of the HST hydraulic pump 51 is temporarily stopped is controlled to a range in which the occurrence of an engine stall can be avoided (a range in which the engine torque has a margin relative to the upper torque limit of the engine 11) (see the relationship between the second and bottom rows in FIG. 11 ).

[0084] The shift control of the transmission 30 by the shift control device 80 in FIG. 10 can be initiated when a predetermined condition is met. The predetermined condition includes, for example, the vehicle speed of the wheel loader 1 reaching a predetermined value or the operation of a switch installed on the wheel loader 1. Specifically, the processing procedure for the shift control device 80 to control the shift of the transmission 30 is as follows: first, as shown in FIG. 10 , the lockup clutch 64 of the direct-coupled mechanism 36 is pressed (step S10). Specifically, the clutch control unit 97 increases the pressing pressure of the lockup clutch 64 until it reaches its maximum. This causes the pressing pressure of the lockup clutch 64 to change as shown in the top row of FIG. 11 . At this time, the shift control device 80 maintains the pump displacement of the HST hydraulic pump 51 at its maximum displacement, as shown in the second row of FIG. 11 . In other words, the HMT 35 of the transmission 30 is maintained in the state prior to the start of the shift control. These processing procedures are the same as those in the comparative example shown in FIGS. 7 and 8 .

[0085] Next, the gear change control device 80 determines whether there is a margin of engine torque to prevent the engine from stalling (step S20). As described above, when the pump displacement of the HST hydraulic pump 51 is reduced, the engine torque increases due to slippage of the lock-up clutch 64 and the opening of the main relief valve 56 of the HST 50, which raises the concern that the engine may stall (see the bottom row in FIG. 8). Furthermore, if the traveling load increases during this switching control, the engine control device 100 must be able to respond in time with engine control to prevent the engine from stalling. Therefore, before starting to reduce the pump displacement, the gear change control device 80 determines whether there is a margin of engine torque to prevent the engine from stalling.

[0086] Specifically, the torque margin determination unit 91 determines whether the difference between the upper limit torque of the engine 11 corresponding to the engine speed during switching control and the output torque of the engine 11 (an estimated value by the engine control device 100 or a detected value by a sensor) exceeds a first threshold. If the difference exceeds the first threshold, it is determined that there is a margin in the engine torque to prevent the engine from stalling (YES), and the process proceeds to step S30. On the other hand, if the difference is equal to or less than the first threshold, it is determined that there is no margin in the engine torque to prevent the engine from stalling (NO), and the process proceeds to step S100.

[0087] If the answer is YES in step S20 (there is a margin in engine torque), the gear change control device 80 (pump volume control unit 95) reduces the pump volume of the HST hydraulic pump 51 (step S30) and again determines whether there is a margin in engine torque to prevent the occurrence of an engine stall (step S40). If the answer is YES in step S40 (there is a margin in engine torque), the process returns to step S30 again to reduce the pump volume again, and this is repeated until the answer is NO in step S40 (there is no margin in engine torque). The repeated processing of steps S30 to S40 corresponds to controlling the degree of reduction in the pump volume of the HST hydraulic pump 51 within a range that can prevent the occurrence of an engine stall.

[0088] If the determination in step S40 is NO (there is insufficient engine torque margin), the transmission control device 80 (pump volume control unit 95) temporarily stops the reduction in the pump volume of the HST hydraulic pump 51 and maintains the pump volume constant (step S50). At this time, the reduction in pump volume causes energy loss due to relief of the HST 50, reducing the torque capacity of the HST 50 and approaching the torque capacity of the lock-up clutch 64. Therefore, although the lock-up clutch 64 is slipping, the slippage gradually decreases and the lock-up clutch 64 is able to transmit power. In other words, the transmission 30 is in a dual power transmission state in which power transmission by the HMT 35 and power transmission by the direct drive mechanism 36 coexist.

[0089] Next, the transmission control device 80 determines whether there is a margin of engine torque to prevent the occurrence of an engine stall (step S60). If it is determined that there is a margin of engine torque (YES), the process proceeds to step S70, whereas if it is determined that there is no margin of engine torque (NO), the process proceeds to step S90.

[0090] If the answer is YES in step S60 (there is a margin in the engine torque), the transmission control device 80 (restart timing determination unit 92) determines whether the vehicle speed of the wheel loader 1 has reached a predetermined value that is slightly lower than the vehicle speed corresponding to the gear ratio when the transmission 30 is in the power transmission path of the direct drive mechanism 36 (step S70). This determines the timing to resume reducing the pump displacement of the HST hydraulic pump 51 based on the transition state of the power transmission path from the HMT 35 to the direct drive mechanism 36. If the answer is NO in step S70 that the vehicle speed of the wheel loader 1 is less than the predetermined value, the process returns to step S60 again to determine whether there is a margin in the engine torque, and this is repeated until a YES determination is made in step S60 and a YES determination is made in step S70 (the actual vehicle speed of the wheel loader 1 is equal to or greater than the predetermined value). In other words, when the actual vehicle speed of the wheel loader 1 is less than the predetermined value, the pump displacement of the HST hydraulic pump 51 is maintained constant, as shown in the second and fourth rows of FIG. As a result, the vehicle speed is maintained at a gradual increase, as shown in the fourth row of FIG.

[0091] If the determination in step S70 is YES (the actual vehicle speed of the wheel loader 1 is equal to or greater than a predetermined value), the gear change control device 80 (pump volume control section 95) resumes reducing the pump volume of the HST hydraulic pump 51, as shown in the second row of FIG. 11 , and reduces the pump volume to zero (step S80). This completely switches the power transmission path of the transmission 30 from the HMT 35 to the direct drive mechanism 36. At this time, because the vehicle speed of the wheel loader 1 is already approaching the speed corresponding to the gear ratio of the power transmission path of the direct drive mechanism 36, it is possible to prevent a sudden transition of the power transmission path from the HMT 35 to the direct drive mechanism 36 and an accompanying sudden switch to a high-speed gear ratio. After processing step S80, the gear change control device 80 ends control of switching the power transmission path of the transmission 30.

[0092] Furthermore, if step S60 returns NO (no engine torque margin), the gear shift control device 80 (pump displacement control unit 95) increases the pump displacement of the HST hydraulic pump 51 (step S90). For example, while the pump displacement of the HST hydraulic pump 51 is maintained constant (step S50), the driving load may increase due to hill climbing or other factors. In this case, the engine control device 100 detects a decrease in engine speed and increases the fuel injection amount to prevent engine stall. However, because the pump displacement has already been reduced from its maximum displacement, as shown in the bottom row of FIG. 12, the engine torque margin may be relatively small compared to the upper limit torque at which engine stall may occur. In this case, the engine control device 100 may not be able to respond in time to the increase in driving load through engine control, making it impossible to avoid engine stall through engine control alone. Therefore, the gear shift control device 80 increases the pump displacement of the HST hydraulic pump 51 to its maximum displacement, as shown in the second row of FIG. 12.

[0093] After increasing the pump displacement of the HST hydraulic pump 51 in step S90, the transmission control device 80 returns to step S20 and determines whether there is a margin of engine torque to prevent the engine from stalling. In other words, it redoes the switching control for switching the power transmission path of the transmission 30 from the HMT 35 to the direct drive mechanism 36.

[0094] If the determination in step S20 is NO (there is no engine torque margin), the gear change control device 80 (clutch control unit 97) disengages the lock-up clutch 64 (step S100), as shown in the top row of Fig. 12. In other words, the switching control for switching the power transmission path of the transmission 30 from the HMT 35 to the direct drive mechanism 36 is canceled.

[0095] If there is still concern about the occurrence of an engine stall, the transmission control device 80 (motor displacement control section 96) increases the motor displacement of the HST hydraulic motor 52 from zero (processing not shown in FIG. 10 ). As a result, the power transmission of the transmission 30 shifts from power transmission via only the planetary gear mechanism 40 of the HMT 35 to power transmission via the HST 50 as well, so the gear ratio of the transmission becomes lower than the maximum speed of the continuously variable transmission of the HMT 35, making it possible to avoid an engine stall.

[0096] As described above, in the present embodiment, when the power transmission path of the transmission 30 is switched from the HMT 35 to the direct drive mechanism 36, the reduction in the pump volume of the HST hydraulic pump 51 is temporarily stopped midway, and the pump volume is maintained constant. Thereafter, when the vehicle speed of the wheel loader 1 reaches a predetermined value, the reduction in the pump volume is resumed. By temporarily stopping the reduction in the pump volume of the HST hydraulic pump 51, it is possible to gradually transition the power transmission path from the HMT 35 to the direct drive mechanism 36 without stalling the engine 11.

[0097] As described above, the wheel loader 1 (work vehicle) according to the first embodiment is equipped with the engine 11 (prime mover) which is the power source for traveling, the transmission 30 including the HMT 35 (hydromechanical first power transmission device) which transmits the rotational power of the engine 11 (prime mover) by continuously changing the speed, and the direct-coupled mechanism 36 (mechanical second power transmission device) which has a gear ratio higher than the highest gear ratio of the HMT 35 (first power transmission device) and transmits the rotational power of the engine 11 (prime mover) without going through the HMT 35 (first power transmission device), and the transmission control device 80 (control device) which controls the transmission 30. The HMT 35 (first power transmission device) includes a planetary gear mechanism 40 (mechanical power transmission mechanism) that divides the rotational power input from the engine 11 (prime mover) into two and outputs them, and an HST 50 (hydraulic power transmission mechanism) that has a variable displacement HST hydraulic pump 51 (hydraulic pump) and a variable displacement HST hydraulic motor 52 (hydraulic motor) that are fluidly connected to each other, and one of the rotational powers output from the planetary gear mechanism 40 (mechanical power transmission mechanism) is input to the HST hydraulic pump 51 (hydraulic pump) and output from the HST hydraulic motor 52 (hydraulic motor). The direct drive mechanism 36 (second power transmission device) is configured to include a lock-up clutch 64 (clutch) that can be switched between an engaged state that connects the power transmission path of the direct drive mechanism 36 (second power transmission device) and a disengaged state that disconnects the power transmission path of the direct drive mechanism 36 (second power transmission device). The shift control device 80 (control device) performs switching control to switch the power transmission path of the transmission 30 from the HMT 35 (first power transmission device) to the direct drive mechanism 36 (second power transmission device). This switching control includes changing the lockup clutch 64 (clutch) from a released state to a pressing state that enables switching to an engaged state, and performing pump volume reduction control to reduce the pump volume of the HST hydraulic pump 51 (hydraulic pump) of the HST 50 (hydraulic power transmission mechanism) from the state (maximum volume) immediately before the above-mentioned switching control to a predetermined volume (zero) while keeping the lockup clutch 64 (clutch) in the pressing state.The pump volume reduction control reduces the pump volume of the HST hydraulic pump 51 (hydraulic pump) in at least two stages, and is performed so that the rate of reduction of the pump volume in the second stage (T1 to T2 (see the time axis at the bottom of FIG. 11)) following the first stage is slower than the rate of reduction of the pump volume in the initial first stage (Ts to T1 (see the time axis at the bottom of FIG. 11)).

[0098] When the pump volume is reduced in at least two stages as in this configuration, by making the reduction rate in the second stage (T1 to T2) after the first stage (Ts to T1) slower than the reduction rate in the first stage from the start of pump volume reduction control, it is possible to suppress gear shift shock and avoid stalling of the engine 11 (prime mover) when switching the power transmission path of the transmission 30 from the HMT 35 (first power transmission device) to the direct drive mechanism 36 (second power transmission device).

[0099] Furthermore, the transmission control device 80 (control device) according to the present embodiment is preferably configured to control the pump volume reduction control of the HST hydraulic pump 51 (hydraulic pump) so that, within the period (Ts to Te) from the start time Ts (see the time axis at the bottom of FIG. 11) to the end time Te (see the time axis at the bottom of FIG. 11), during a first period (Ts to T1) from the start time Ts to the first time T1 (see the time axis at the bottom of FIG. 11) and a second period (T2 to Te) from the second time T2 (see the time axis at the bottom of FIG. 11) to the end time Te, the pump volume reduction control is controlled so that the rate of reduction in the pump volume of the HST hydraulic pump 51 (hydraulic pump) is smaller than the rate of reduction in the first period (Ts to T1) and the rate of reduction in the second period (T2 to Te) and is smaller than a set value (zero).

[0100] According to this configuration, by setting the pump volume reduction rate of the HST hydraulic pump 51 (hydraulic pump) of the HST 50 (hydraulic power transmission mechanism) in the third period (T1 to T2) of the pump volume reduction control to be smaller than the reduction rate in the first period (Ts to T1) and the second period (T2 to Te) and smaller than a set value (zero), an increase in the output torque of the engine 11 (prime mover) can be avoided or suppressed, and the change in the transition state of the power transmission path between the HMT 35 (first power transmission device) and the direct-coupled mechanism 36 (second power transmission device), which are in a dual power transmission state, can be made more gradual than in the first period (Ts to T1) and the second period (T2 to Te). Therefore, when the power transmission path of the transmission 30 is switched from the HMT 35 (first power transmission device) to the direct-coupled mechanism 36 (second power transmission device), shift shock can be suppressed and a stall of the engine 11 (prime mover) can be avoided.

[0101] Furthermore, the transmission control device 80 (control device) according to this embodiment controls the pump volume of the HST hydraulic pump 51 (hydraulic pump) to be kept constant during the third period (T1 to T2) of the pump volume reduction control.

[0102] With this configuration, by maintaining a constant pump volume, it is possible to maintain a dual power transmission state in which power transmission by the HMT 35 (first power transmission device) and power transmission by the direct drive mechanism 36 (second power transmission device) coexist for a long period of time. This maintains a gradual increase in vehicle speed, thereby reliably suppressing gear shift shock.

[0103] In addition, the transmission control device 80 (control device) according to this embodiment is configured to perform pump volume reduction control at a first time point T1, which is the time point at which the difference between the upper limit torque that can be output by the engine 11 (prime mover), corresponding to the rotation speed of the engine 11 (prime mover) during pump volume reduction control, and the output torque of the engine 11 (prime mover) becomes equal to or less than a predetermined threshold value.

[0104] According to this configuration, once the margin of output torque of the engine 11 (prime mover) against stalling of the engine 11 (prime mover) has decreased to a certain extent, the rate of decrease in pump volume is made smaller (zero) than the rate of decrease in the first period (Ts to T1), thereby making it possible to accelerate the transition of the power transmission path of the transmission 30 to the direct drive mechanism 36 while avoiding stalling of the engine 11 (prime mover).

[0105] Furthermore, the speed change control device 80 (control device) according to this embodiment is configured to perform pump volume reduction control, with the point in time T2 being the point in time when the vehicle speed of the wheel loader 1 (work vehicle) reaches a predetermined value that is set in accordance with the speed ratio of the direct-coupled mechanism 36 (second power transmission device) and the rotation speed of the engine 11 (prime mover) during pump volume reduction control.

[0106] According to this configuration, when the vehicle speed of the wheel loader 1 (work vehicle) reaches a predetermined value corresponding to the gear ratio of the direct-coupled mechanism 36 (second power transmission device), the transition of the power transmission path of the transmission 30 from the HMT 35 (first power transmission device) to the direct-coupled mechanism 36 (second power transmission device) has progressed reliably, and therefore, by making the rate of decrease in pump volume greater than the rate of decrease in the second period (T1 to T2) (by resuming the decrease in pump volume), it is possible to quickly complete the transmission 30 switching control while avoiding gear shift shock.

[0107] In addition, the transmission control device 80 (control device) according to this embodiment is configured to perform pump volume reduction control so that the output torque of the engine 11 (prime mover) does not exceed the upper limit torque that can be output by the engine 11 (prime mover), which corresponds to the rotation speed of the engine 11 (prime mover) during pump volume reduction control.

[0108] According to this configuration, by reducing the pump volume so that the output torque of the engine 11 (prime mover) does not exceed the upper limit torque, it is possible to avoid stalling of the engine 11 (prime mover) during the transition of the power transmission path of the transmission 30 from the HMT 35 (first power transmission device) to the direct drive mechanism 36 (second power transmission device).

[0109] Furthermore, the transmission control device 80 (control device) according to this embodiment is configured to perform control to increase the pump volume of the HST hydraulic pump 51 if the rotation speed of the engine 11 (prime mover) is not maintained constant and drops due to overload during control to reduce the pump volume.

[0110] According to this configuration, by increasing the pump volume of the HST hydraulic pump 51, the load on the engine 11 (prime mover) can be reduced, and therefore stalling of the engine 11 (prime mover) can be avoided.

[0111] Furthermore, the transmission control device 80 (control device) according to this embodiment is configured to alleviate the state in which the lock-up clutch 64 (clutch) is pressed when the pump volume of the HST hydraulic pump 51 is increased during pump volume reduction control, if the difference between the upper limit torque that can be output by the engine 11 (prime mover) at this time and the output torque of the engine 11 (prime mover) is equal to or less than a preset threshold value.

[0112] According to this configuration, in addition to increasing the pump volume, the load on the engine 11 (prime mover) can be further reduced by releasing the pressed state of the lock-up clutch 64 (clutch), thereby avoiding stalling of the engine 11 (prime mover).

[0113] [Second Embodiment] Next, a work vehicle according to a second embodiment of the present invention will be described using Figures 13 and 14. Figure 13 is a block diagram showing the configuration of the hardware and functional parts of a gear change control device for a work vehicle according to the second embodiment of the present invention. Figure 14 is a time chart showing an example of the transition in the state of the lock-up clutch, HST hydraulic pump, engine, and vehicle travel when switching the power transmission path of the transmission in the work vehicle according to the second embodiment shown in Figure 13 (switching from HMT to direct drive mechanism). In Figures 13 and 14, parts with the same reference numerals as those shown in Figures 1 to 12 are similar parts, and detailed description thereof will be omitted.

[0114] The work vehicle according to the second embodiment differs from the first embodiment in the following ways: In the first embodiment, when the power transmission path of the transmission 30 is switched from the HMT 35 to the direct drive mechanism 36, the pump displacement of the HST hydraulic pump 51 is reduced while maintaining the engine speed constant (control of the engine control device 100) (see the second and third rows of FIG. 11 ). In contrast, the gear change control device 80A according to the second embodiment is configured to increase the engine speed while temporarily stopping the reduction in pump displacement of the HST hydraulic pump 51 when the power transmission path of the transmission 30 is switched from the HMT 35 to the direct drive mechanism 36, thereby restarting the reduction in pump displacement when the vehicle speed of the wheel loader 1 is lower than a predetermined value.

[0115] Specifically, the gear change control device 80A according to the second embodiment shown in FIG. 13 includes a torque margin determination unit 91A and a transmission control unit 93 (a pump displacement control unit 95, a motor displacement control unit 96, and a clutch control unit 97) similar to the functional units of the gear change control device 80 according to the first embodiment, and further includes an engine control unit 98 that commands the engine control device 100 about the engine speed. However, the torque margin determination unit 91A not only determines the timing to temporarily suspend the reduction of the pump displacement, but also determines the timing to resume the reduction of the pump displacement, like the restart timing determination unit 92 in the first embodiment. Specifically, the torque margin determination unit 91A determines whether the output torque of the engine 11 has a margin relative to the upper limit torque at which engine stall may occur. As in the first embodiment, for example, when the difference between the upper limit torque that the engine can output at the engine speed during switching control and the engine torque Te (estimated value) from the engine control device 100 becomes equal to or less than a first threshold, the torque margin determination unit 91A outputs a determination result to temporarily stop the reduction of the pump displacement of the HST hydraulic pump 51. Furthermore, when the difference between the upper limit torque of the engine and the engine torque Te (estimated value) from the engine control device 100 becomes equal to or greater than a second threshold value, the torque margin determination unit 91A outputs a determination result that causes the reduction of the pump displacement of the HST hydraulic pump 51 to resume. Note that the transmission control unit 93 of the gear change control device 80A of the present embodiment is similar to the functional unit of the gear change control device 80 according to the first embodiment, and therefore a description thereof will be omitted.

[0116] The engine control unit 98 outputs to the engine control device 100 a command to increase or decrease the rotation speed of the engine 11 in accordance with the process of reducing the pump displacement of the HST hydraulic pump 51. Specifically, the engine control unit 98 outputs a command to gradually increase the rotation speed of the engine 11 from the point in time when the reduction in pump displacement of the HST hydraulic pump 51 is temporarily stopped (determined by the determination result of the torque margin determination unit 91A). Furthermore, from the point in time when the reduction in pump displacement of the HST hydraulic pump 51 is resumed (determined by the determination result of the torque margin determination unit 91A), the engine control unit 98 outputs a command to reduce the rotation speed of the engine 11 within a range that does not cause engine stall. In addition, the engine control unit 98 outputs a command to gradually increase the reduced engine rotation speed to return it to the original rotation speed immediately before the switching control. Note that the point in time when the reduction in pump displacement is temporarily stopped, which is the timing to increase the rotation speed of the engine 11, also includes the vicinity of the point in time of the temporary stop.

[0117] Time domains A, B, and C in Fig. 14 are the same as those in the first embodiment shown in Fig. 11. That is, in time domain A, the power transmission path of the transmission 30 is in a state of the highest speed power transmission path of the continuously variable transmission in the HMT 35 (see Fig. 5). In time domain C, the power transmission path of the transmission 30 is in a state in which it has completely switched to the direct drive mechanism 36. In time domain B, the power transmission path of the transmission 30 is in a transient state (dual power transmission state) in which it is shifting from the HMT 35 to the direct drive mechanism 36.

[0118] In time region B, the switching control of the transmission 30 by the shift control device 80A of this embodiment differs from the switching control of the shift control device 80 according to the first embodiment in the following ways. As shown in the second and third rows of FIG. 14 , the shift control device 80A gradually increases the engine speed via the engine control device 100 from the point at which the reduction in pump displacement of the HST hydraulic pump 51 is temporarily stopped. Thereafter, the reduction in pump displacement of the HST hydraulic pump 51 is resumed before the actual vehicle speed of the wheel loader 1 reaches a predetermined value, that is, when the actual vehicle speed is lower than the predetermined value. This makes it possible to resume the reduction in pump displacement earlier than in the first embodiment. By increasing the engine speed in advance before the reduction in pump displacement resumes, the upper limit torque that can be output by the engine 11 is increased (see the bottom row of FIG. 14 ), and therefore engine stall can be avoided even if the reduction in pump displacement resumes earlier.

[0119] Furthermore, the gear change control device 80A resumes the reduction of the pump volume of the HST hydraulic pump 51, and at the same time reduces the engine speed via the engine control device 100 to a range that does not cause the engine to stall. This is because, by resuming the reduction of the pump volume more quickly than in the first embodiment, gear change shock may occur unless the engine speed is reduced.

[0120] The transmission control device 80A completely switches the power transmission path of the transmission 30 to the direct drive mechanism 36 by ultimately reducing the pump displacement of the HST hydraulic pump 51 to zero. As a result, the gear ratio of the transmission 30 becomes the gear ratio of the direct drive mechanism 36. However, because the transmission control device 80A reduces the engine speed to avoid gear shift shock when the reduction of the pump displacement resumes, if the engine speed remains reduced, the vehicle speed of the wheel loader 1 will not reach the vehicle speed corresponding to the original engine speed before the switching control. Therefore, the transmission control device 80A gradually increases the engine speed via the engine control device 100 to return it to the original engine speed before the switching control. This makes it possible to gradually accelerate the vehicle to a speed corresponding to the gear ratio of the direct drive mechanism 36, which uses the original engine speed as an input. Therefore, the power transmission path of the transmission 30 can be switched to the power transmission path of the direct drive mechanism 36 without generating gear shift shock.

[0121] In this way, in the present embodiment, by raising the upper limit torque of the engine (engine speed) by an amount necessary to avoid engine stall when the reduction in pump displacement of the HST hydraulic pump 51 is temporarily stopped, the period during which the reduction in pump displacement is temporarily stopped (the period during which the pump displacement is maintained constant) can be ended earlier than in the first embodiment. This shortens the period of dual power transmission in which energy loss occurs due to slippage of the lock-up clutch 64 and relief operation of the HST 50, resulting in fuel savings.

[0122] According to the second embodiment described above, as in the first embodiment, by setting the rate of reduction in the pump volume of the HST hydraulic pump 51 (hydraulic pump) of the HST 50 (hydraulic power transmission mechanism) in the third period of the pump volume reduction control (T1 to T2 on the time axis at the bottom of FIG. 14) to be smaller than the rate of reduction in the first period (Ts to T1 on the time axis at the bottom of FIG. 14) and the second period (T2 to Te on the time axis at the bottom of FIG. 14) and smaller than the set value (zero), it is possible to avoid or suppress an increase in the output torque of the engine 11 (prime mover), and to make the change in the transition state of the power transmission path in the HMT 35 (first power transmission device) and the direct drive mechanism 36 (second power transmission device), which are in a dual power transmission state, more gradual than in the first period (Ts to T1) and the second period (T2 to Te). Therefore, when the power transmission path of the transmission 30 is switched from the HMT 35 (first power transmission device) to the direct drive mechanism 36 (second power transmission device), shift shock can be suppressed and stalling of the engine 11 (prime mover) can be avoided.

[0123] Furthermore, the transmission control device 80A (control device) according to the second embodiment simultaneously performs prime mover rotation speed control to control the rotation speed of the engine 11 (prime mover) when performing switching control to switch the power transmission path of the transmission 30 from the HMT 35 to the direct drive mechanism 36. The prime mover rotation speed control includes gradually increasing the rotation speed of the engine 11 (prime mover) from a start time Ts of the pump volume reduction control, starting to reduce the rotation speed of the engine 11 (prime mover) at a second time T2 of the pump volume reduction control, and temporarily reducing the rotation speed of the engine 11 (prime mover) to a speed lower than the rotation speed of the engine 11 (prime mover) at the start time Ts.

[0124] According to this configuration, the upper limit torque of the engine 11 (prime mover) can be raised by increasing the engine speed of the engine 11 (prime mover) from the stage where the reduction of the pump displacement is temporarily stopped, so the period during which the reduction of the pump displacement is temporarily stopped can be ended earlier than in the first embodiment. In other words, the timing at which the reduction of the pump displacement is resumed can be accelerated. This shortens the period during which the power is dually transmitted between the HMT 35 (first power transmission device) and the direct-coupled mechanism 36 (second power transmission device), which causes energy loss, thereby improving fuel economy.

[0125] Furthermore, the transmission control device 80A (control device) according to the second embodiment performs prime mover drive control to control the drive of the engine 11 (prime mover) when performing switching control to switch the power transmission path of the transmission 30 from the HMT 35 to the direct drive mechanism 36. The prime mover drive control includes gradually increasing the upper limit torque that can be output by the engine 11 (prime mover) from a start time Ts of the pump volume reduction control, starting to reduce the upper limit torque of the engine 11 (prime mover) at a second time T2 of the pump volume reduction control, and temporarily reducing the upper limit torque of the engine 11 (prime mover) to a value lower than the upper limit torque of the engine 11 (prime mover) at the start time Ts.

[0126] According to this configuration, by raising the upper limit torque of the engine from the stage where the reduction of the pump displacement is temporarily stopped, the period during which the reduction of the pump displacement is temporarily stopped can be ended earlier than in the first embodiment. In other words, the timing at which the reduction of the pump displacement is resumed can be accelerated. This shortens the period during which the power is dually transmitted between the HMT 35 (first power transmission device) and the direct-coupled mechanism 36 (second power transmission device), which causes energy loss, thereby improving fuel economy.

[0127] [Third Embodiment] Next, a work vehicle according to a third embodiment of the present invention will be described using Figures 15 to 17. In Figures 15 to 17, parts that are the same as those shown in Figures 1 to 11 have the same reference numerals, and therefore detailed description thereof will be omitted. Figure 15 is a block diagram showing the configuration of the hardware and functional units of a control device for a work vehicle according to the third embodiment of the present invention.

[0128] The work vehicle according to the third embodiment differs from the first embodiment in that the transmission 30 performs power transmission path switching control assuming the case where a load-handling load caused by the operation of the working implement 6 occurs simultaneously during travel. Load fluctuations that occur on the wheel loader 1 during transmission 30 switching control include not only fluctuations in the traveling load but also load fluctuations due to the simultaneous occurrence of a load-handling load for the operation of the working implement 6. In this case, the load fluctuations are particularly large because the load-handling load acts superimposed on the traveling load. As shown in FIG. 3 , the engine 11, which serves as the traveling power source, also drives the load-handling hydraulic pump 13 that operates the working implement 6. Therefore, a portion of the engine 11's power is used for the load-handling hydraulic pump 13, thereby reducing the power available for travel. This means that the occurrence of a load-handling load not only limits traveling performance but also increases the risk of engine stall. For example, if a load-handling load is superimposed on the engine 11 while the engine 11 is operating at the minimum rotational speed necessary to avoid stalling under the current traveling load in consideration of fuel efficiency, engine stall may occur. Therefore, in the switching control of the power transmission path of the transmission 30, it is necessary to also respond to load fluctuations caused by the occurrence of cargo loads.

[0129] 2 and 3, the power of the engine 11 is partially distributed to the load-handling hydraulic pump 13 upstream of the transmission 30 of the traveling power transmission system 20, so when a load is loaded, the load on the engine 11 increases. In this case, in order to avoid an unintended decrease in engine speed or the occurrence of an engine stall, it is possible to ensure the required engine torque by controlling the engine 11 alone to maintain or increase its speed.

[0130] However, when a cargo load is superimposed on the engine 11, in the control of switching the power transmission path of the transmission 30, the timing for temporarily suspending the reduction in pump volume of the HST hydraulic pump 51 must be determined not only within a range that can avoid the occurrence of engine stall as in the first embodiment, but also within a range that prevents excessive torque from being input from the engine 11 to the transmission 30 in order to suppress gear shift shock. That is, in this embodiment, the determination of temporarily suspending the reduction in pump volume is different from that in the first embodiment.

[0131] The first embodiment is directed to control of switching of the power transmission path of the transmission 30 when the wheel loader 1 is traveling only without operating the work implement 6. When there is no cargo load and only a traveling load is generated, only the traveling load fluctuates, and even if the power of the engine 11 fluctuates, it is always equal to the input power to the transmission 30. For this reason, if the engine 11 is controlled to a rotation speed near the limit at which it does not stall given the engine load torque at that time, engine stall will not occur, and the input torque to the transmission 30 will be minimized, so the wheel loader 1 will not suddenly accelerate. In the first embodiment, the output torque of the engine 11 is suppressed because it is assumed that the engine rotation speed will be kept low with an emphasis on fuel efficiency. For this reason, excessive torque will not be input from the engine 11 to the transmission 30, so it was sufficient for the transmission control device 80 according to the first embodiment to control the degree to which the pump displacement of the HST hydraulic pump 51 is reduced (the timing at which the reduction in pump displacement is temporarily stopped) within a range in which engine stall can be avoided.

[0132] On the other hand, when the power of the engine 11 is distributed between the traveling load side and the cargo load side, the engine speed cannot be reduced below the speed required to drive the cargo handling hydraulic pump 13. Therefore, from the perspective of the traveling power transmission system 20, the torque of the engine 11 may not have reached near the upper limit of torque that can be output at that current rotational speed, and the engine 11 may be operating with some reserve power. In this state, if the pump displacement of the HST hydraulic pump 51 is reduced to near the limit of the range in which engine stall can be avoided, as in the first embodiment, the pump displacement will be significantly reduced compared to the first embodiment. This significantly reduces the torque capacity of the HST 50, causing a sudden shift in the power transmission path from the HMT 35 to the direct-coupled mechanism 36, which may cause the wheel loader 1 to suddenly accelerate. In other words, if the pump displacement of the HST hydraulic pump 51 is reduced to the limit of the range in which engine stall can be avoided when the engine 11 has some reserve power, there is a concern that a gear shift shock may occur.

[0133] Therefore, the gear change control device 80B according to this embodiment is configured to perform control to reduce the pump displacement of the HST hydraulic pump 51 so as to avoid engine stall and prevent excessive torque from being input from the engine 11 to the transmission 30. The functional units of the gear change control device 80B according to this embodiment shown in Figure 15 differ from those of the gear change control device 80 according to the first embodiment in that the determination method of the torque margin determination unit 91B is different. The other functional units 92 and 93 of the gear change control device 80B according to this embodiment are the same as the functional units of the gear change control device 80 according to the first embodiment, and therefore a description thereof will be omitted.

[0134] The torque margin determination unit 91B of the shift control device 80B does not determine the engine torque margin for engine stall, but determines the margin for the virtual maximum value of input torque to the transmission 30 to suppress shift shock. The virtual maximum torque is, for example, the upper limit torque that can be output corresponding to the engine 11 rotation speed set in the first embodiment, which does not assume driving of the cargo handling hydraulic pump 13 (load handling load), as the virtual maximum value (input threshold) of input torque to the transmission 30. For example, the torque margin determination unit 91B calculates the input torque to the transmission 30 by subtracting the torque of the cargo handling hydraulic pump 13 estimated from the discharge pressure of the cargo handling hydraulic pump 13, which is the detected value of the pressure sensor 15, from the estimated value of engine torque acquired from the engine control device 100. The torque margin determination unit 91B then determines whether the calculated input torque to the transmission 30 is less than the virtual maximum value. If the input torque (calculated value) to the transmission 30 is less than the virtual maximum value, it is determined that excessive torque is not being input to the transmission 30. On the other hand, if the input torque (calculated value) of the transmission 30 is equal to or greater than the virtual maximum value, it is determined that excessive torque is being input to the transmission 30. The torque margin determination unit 91B outputs the determination result as to whether excessive torque is being input to the transmission 30 to the transmission control unit 93. The virtual maximum value is stored in advance in the storage device 81, for example.

[0135] In this embodiment, the determination by the torque surplus determination unit 91B is based on the torque of the cargo handling hydraulic pump 13 estimated from the estimated value of the engine torque taken from the engine control device 100 and the discharge pressure of the cargo handling hydraulic pump 13, which is the detection value of the pressure sensor 15. However, the determination by the torque surplus determination unit 91B can also be based on the detection value of a torque sensor that can detect the input torque of the transmission 30.

[0136] Next, the processing procedure for switching control of the power transmission path of the transmission by the gear change control device of the work vehicle according to the third embodiment and the effects of this switching control will be described with reference to Figure 3 and Figures 15 to 17. Figure 16 is a flowchart showing an example of the processing procedure for control of the transmission (switching control of the power transmission path from the HMT to the direct drive mechanism) by the gear change control device of the work vehicle according to the third embodiment shown in Figure 15. Figure 17 is a time chart showing an example of the transition in the states of the lock-up clutch, HST hydraulic pump, engine, and vehicle running when switching the power transmission path of the transmission (switching from the HMT to the direct drive mechanism) in the work vehicle according to the third embodiment shown in Figure 15.

[0137] 16, switching control of the power transmission path of the transmission 30 by the gear change control device 80B shown in FIG. 15 can be initiated when a predetermined condition is met. The predetermined condition includes, for example, the vehicle speed of the wheel loader 1 reaching a predetermined value or the operation of a switch installed on the wheel loader 1. In summary, as with the first embodiment, this switching control involves pressing the lock-up clutch 64 of the direct-coupled mechanism 36 before starting to reduce the pump volume of the HST hydraulic pump 51 (step S10). Thereafter, the pump volume of the HST hydraulic pump 51 is reduced (steps S30 and S80). However, the reduction in pump volume is temporarily stopped midway (step S50), and when the vehicle speed of the wheel loader 1 reaches a predetermined value (step S70), the reduction in pump volume is resumed (step S80).

[0138] Here, unlike the first embodiment, when the pump displacement of the HST hydraulic pump 51 is reduced (steps S30 and S80) and when the reduction in the pump displacement of the HST hydraulic pump 51 is temporarily stopped and maintained constant (step S50), it is determined whether the input torque to the transmission 30 is less than a virtual maximum value (input threshold value) (steps S20B, 40B, S60B). This is to prevent excessive torque from being input to the transmission 30 during the process of reducing the pump displacement of the HST hydraulic pump 51 from the maximum displacement to zero. This prevents a sudden shift to the power transmission path of the direct-coupled mechanism 36 caused by excessive torque to the transmission 30, thereby suppressing gear shift shock.

[0139] Time domains A, B, and C in FIG. 17 are the same as those in the first embodiment shown in FIG. 11 . That is, in time domain A, the power transmission path of the transmission 30 is in a state of the highest speed power transmission path of the continuously variable transmission in the HMT 35 (see FIG. 5 ). In time domain C, the power transmission path of the transmission 30 is in a state in which it has completely switched to the direct drive mechanism 36. In time domain B, the power transmission path of the transmission 30 is in a transient state (a dual power transmission state) in which it is shifting from the HMT 35 to the direct drive mechanism 36. That is, time domain B shows the changes in the state of the transmission 30 and the engine 11 when the gear change control device 80B goes through steps S10 to S80 (excluding steps S90 and S100) of the processing procedure of the flowchart shown in FIG. 16 .

[0140] In this embodiment, the engine speed is higher than in the first embodiment to drive the cargo handling hydraulic pump 13, as shown in the third row of FIG. 17 . Therefore, as shown in the fourth row of FIG. 17 , the vehicle speed of the wheel loader 1 is higher from time region A to time region C than in the first embodiment. However, the degree of change in vehicle speed over time (slope) is the same as in the first embodiment. This is because, in this embodiment, in which a traveling load and a cargo handling load are simultaneously generated, even if the engine 11 is operating with a reserve of power (a state in which there is a large margin of engine torque against engine stall), when the pump displacement of the HST hydraulic pump 51 is reduced or temporarily stopped, a determination is made (steps S20B, 40B, S60B) to avoid input of excessive torque to the transmission 30, and the input torque of the transmission 30 is limited to the engine torque in the first embodiment (the upper limit torque corresponding to the engine speed when there is no cargo handling load). This makes it possible to prevent a sudden shift in the power transmission path to the direct-coupled mechanism 36. That is, even when the shift control device 80B of this embodiment controls the switching of the power transmission path of the transmission 30, shift shock can be suppressed.

[0141] In this embodiment, the input torque of the transmission 30 is limited to less than the virtual maximum value, so as shown in the fifth row of FIG. 17, the engine torque margin before the engine stalls is larger than in the first embodiment where there is no load on the cargo handling hydraulic pump 13 (see the bottom row of FIG. 11).

[0142] According to the third embodiment described above, as in the first embodiment, by setting the rate of reduction in the pump volume of the HST hydraulic pump 51 (hydraulic pump) of the HST 50 (hydraulic power transmission mechanism) in the third period of the pump volume reduction control (T1 to T2 on the time axis at the bottom of FIG. 17) to be smaller than the rate of reduction in the first period (Ts to T1 on the time axis at the bottom of FIG. 17) and the second period (T2 to Te on the time axis at the bottom of FIG. 17) and smaller than the set value (zero), it is possible to avoid or suppress an increase in the output torque of the engine 11 (prime mover), and to make the change in the transition state of the power transmission path in the HMT 35 (first power transmission device) and the direct drive mechanism 36 (second power transmission device), which are in a dual power transmission state, more gradual than in the first period (Ts to T1) and the second period (T2 to Te). Therefore, when the power transmission path of the transmission 30 is switched from the HMT 35 (first power transmission device) to the direct drive mechanism 36 (second power transmission device), shift shock can be suppressed and stalling of the engine 11 (prime mover) can be avoided.

[0143] The wheel loader 1 (work vehicle) according to the third embodiment described above also includes a cargo handling hydraulic pump 13 (equipment) that is driven by input of a portion of the rotational power of the engine 11 (prime mover). The transmission control device 80B (control device) is configured to perform pump displacement reduction control so that the torque input from the engine 11 (prime mover) to the transmission 30 does not exceed a virtual maximum value (input threshold value).

[0144] According to this configuration, when the cargo handling hydraulic pump 13 (equipment) is driven simultaneously while traveling, the pump volume is reduced within a range that can prevent excessive torque from being input from the engine 11 to the transmission 30, thereby avoiding a sudden shift in the power transmission path from the HMT 35 to the direct-coupled mechanism 36, thereby suppressing gear shift shock and preventing the engine 11 (prime mover) from stalling.

[0145] [Other Embodiments] The present invention is not limited to the first to third embodiments described above, and includes various modifications. The above embodiments have been described in detail to clearly explain the present invention, and are not necessarily limited to those including all of the configurations described. For example, it is possible to replace part of the configuration of one embodiment with the configuration of another embodiment, or it is also possible to add the configuration of another embodiment to the configuration of one embodiment. Furthermore, it is also possible to add, delete, or replace part of the configuration of each embodiment with other configurations.

[0146] For example, in the above-described embodiment, an example was shown in which the prime mover 11, which is the power source for traveling the wheel loader 1, was configured as a single engine. However, the prime mover can also be configured as a combination of an engine and an electric motor, or as a single electric motor.

[0147] In the above-described embodiment, an example has been shown in which planetary gear mechanism 40 has first sun gear 41, second sun gear 42, a plurality of first planetary gears 43, a plurality of second planetary gears 44, and carrier 45, and rotational power input to carrier 45 is divided into two paths and transmitted to be output to HST 50 via first sun gear 41 and to be output to output shaft 32 via second sun gear 42. However, the components of the gear mechanism can be selected arbitrarily as long as the planetary gear mechanism has a configuration that functions to divide and transmit rotational power input from prime mover 11 into two paths and transmit the power to be output to HST 50 and to be output to output shaft 32.

[0148] Furthermore, in the above-described embodiment, an example was shown in which the gear change control device 80, 80A, 80B temporarily suspends the reduction of the pump displacement of the HST hydraulic pump 51 midway and controls the pump displacement to maintain a constant value. However, the gear change control device does not necessarily have to temporarily suspend the reduction of the pump displacement midway and maintain the pump displacement constant. For example, within the period from start time Ts (see the time axis at the bottom of FIG. 10 ) to end time Te (see the time axis at the bottom of FIG. 10 ) of the pump displacement reduction control, the gear change control device can also be configured to control the pump displacement so that the rate of reduction of the pump displacement is smaller than the rate of reduction in the first period (Ts to T1) and the rate of reduction in the second period (T2 to Te) and is smaller than a set value during a third period from first time T1 to second time T2, which is between a first period from start time Ts to first time T1 (see the time axis at the bottom of FIG. 10 ) and a second period from second time T2 (see the time axis at the bottom of FIG. 10 ) to end time Te. By making the rate of decrease of the pump volumetric capacity slower than the rates of decrease in the first period (Ts to T1) and the second period (T2 to Te) and slower than the set value, the time during which the HMT 35 and the direct drive mechanism 36 are in a dual power transmission state can be increased. This has the effect of preventing a sudden shift of the power transmission path of the transmission 30 to the direct drive mechanism 36. The set value of the rate of decrease of the pump volumetric capacity is such that the dual power transmission state of the HMT 35 and the direct drive mechanism 36 can be maintained.

[0149] In the above-described embodiment, an example has been shown in which the gear change control devices 80, 80A, 80B and the engine control device 100 are configured separately. However, a configuration in which the gear change control devices 80, 80A, 80B and the engine control device 100 are mounted in a single control device is also possible.

[0150] In the above-described embodiment, an example of a configuration in which the engine 11 drives the cargo handling hydraulic pump 13 has been shown. However, a configuration in which the engine 11 drives a generator instead of the cargo handling hydraulic pump 13 is also possible.

[0151] REFERENCE SIGNS LIST 1... Wheel loader (work vehicle), 11... Prime mover, 13... Load handling hydraulic pump (equipment), 30... Transmission, 35... HMT (hydromechanical first power transmission device), 36... Direct coupling mechanism (mechanical second power transmission device), 40... Planetary gear mechanism (mechanical power transmission mechanism), 50... HST (hydraulic power transmission mechanism), 51... HST hydraulic pump (hydraulic pump), 52... HST hydraulic motor (hydraulic motor), 64... Lock-up clutch (clutch), 80, 80A, 80B... Transmission control device (control device)

Claims

1. A work vehicle comprising: a prime mover as a power source; a transmission including a hydromechanical first power transmission device that continuously changes the speed and transmits the rotational power of the prime mover; and a mechanical second power transmission device that has a gear ratio higher than the highest gear ratio of the first power transmission device and transmits the rotational power of the prime mover without passing through the first power transmission device; and a control device that controls the transmission, wherein the first power transmission device includes a mechanical power transmission mechanism that divides the rotational power input from the prime mover into two and outputs them, and a hydraulic power transmission mechanism that has a variable displacement hydraulic pump and a variable displacement hydraulic motor fluidly connected to each other and whereby one of the rotational powers output from the mechanical power transmission mechanism is input to the hydraulic pump and output from the hydraulic motor, and wherein the second power transmission device includes a clutch that can be switched between an engaged state that connects the power transmission path of the second power transmission device and a disengaged state that disconnects the power transmission path of the second power transmission device, the control device performs switching control to switch the power transmission path of the transmission from the first power transmission device to the second power transmission device when a predetermined condition is satisfied, the switching control including changing the clutch from the released state to a pressing state which enables switching to the engaged state, and performing pump volume reduction control to reduce the pump volume of the hydraulic pump of the hydraulic power transmission mechanism from a state immediately before the switching control to a predetermined volume while the clutch is kept in the pressing state, the pump volume reduction control reducing the pump volume of the hydraulic pump in at least two stages, and performing the pump volume reduction control so that the pump volume reduction rate in a second stage following the first stage is slower than the pump volume reduction rate in an initial first stage of the pump volume reduction control.

2. A work vehicle as claimed in claim 1, characterized in that the pump volume reduction control is performed by controlling the pump volume of the hydraulic pump so that the rate of reduction of the pump volume of the hydraulic pump is smaller than the rate of reduction in the first period and the rate of reduction in the second period and is smaller than a set value during a third period from the first period to the second period, which is between a first period from the start time to the first period and a second period from the second period to the end time, during a period from the start time to the end time of the pump volume reduction control.

3. A work vehicle according to claim 2, wherein the control device controls the pump volume of the hydraulic pump to be maintained constant during the third period of the pump volume reduction control.

4. A work vehicle according to claim 2, wherein the control device performs the pump volume reduction control at a point in time when the difference between the upper limit torque that can be output by the prime mover, which corresponds to the rotation speed of the prime mover during the pump volume reduction control, and the output torque of the prime mover becomes equal to or less than a preset threshold value, as the first point in time.

5. A work vehicle as described in claim 2, wherein the control device performs the pump volume reduction control at the second point in time when the vehicle speed of the work vehicle reaches a predetermined value set in accordance with the gear ratio of the second power transmission device and the rotation speed of the prime mover during the pump volume reduction control.

6. A work vehicle as described in claim 1, wherein the control device performs the pump volume reduction control so that the output torque of the prime mover does not exceed an upper limit torque that can be output by the prime mover, which corresponds to the rotation speed of the prime mover during the pump volume reduction control.

7. A work vehicle as described in claim 2, wherein the control device simultaneously performs prime mover rotation speed control for controlling the rotation speed of the prime mover when performing the switching control, and the prime mover rotation speed control includes gradually increasing the rotation speed of the prime mover from the start point of the pump volume reduction control, starting to reduce the rotation speed of the prime mover at the second point of the pump volume reduction control, and temporarily reducing the rotation speed of the prime mover to a rotation speed lower than the rotation speed of the prime mover at the start point.

8. A work vehicle as described in claim 2, wherein the control device simultaneously performs prime mover drive control for controlling the drive of the prime mover when performing the switching control, and the prime mover drive control includes gradually increasing the upper limit torque that can be output by the prime mover from the start point of the pump volume reduction control, starting to reduce the upper limit torque of the prime mover at the second point of the pump volume reduction control, and temporarily reducing the upper limit torque of the prime mover to a value lower than the upper limit torque of the prime mover at the start point.

9. A work vehicle as described in claim 1, comprising equipment that is driven by inputting a portion of the rotational power of the prime mover, and wherein the control device performs the pump volume reduction control so that the torque input from the prime mover to the transmission does not exceed an input threshold value.

10. A work vehicle as described in claim 1, wherein the control device controls the hydraulic pump to increase its pump volume if the rotation speed of the prime mover is not maintained constant and decreases due to overload during the pump volume reduction control.

11. A work vehicle as described in claim 10, wherein the control device controls the clutch to relax its pressing state when the pump volume of the hydraulic pump is increased during the pump volume reduction control and the difference between the upper limit torque that can be output by the prime mover and the output torque of the prime mover at this time is equal to or less than a preset threshold value.

12. A work vehicle according to claim 1, wherein the predetermined condition includes the speed of the work vehicle reaching a predetermined value or a switch being operated.

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