JP2006220202A - Controller for work vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a controller for a work vehicle capable of easily controlling load to decelerate to prevent occurrence of engine stalling when load is abruptly changed. <P>SOLUTION: A change gear ratio pattern is set by a change gear ratio setting dial. Amount of stepping-on of an advance (backing) pedal is read. A target change gear ratio value corresponding to the amount of stepping-on of the pedal on the selected change gear ratio pattern is calculated based on the read numerical value. A load control mode change-over switch 222a is turned ON. Load is controlled when an auxiliary transmission is at low speed and rotational speed of an engine is below a predetermined value. When engine load is proper load, the target change gear ratio value corresponding to the amount of stepping-on of the pedal on the selected change gear ratio pattern is maintained. When load is large load, the target change gear ratio value is controlled to reduce it until a reduction rate of engine rotational speed becomes less than a given value C% (5% in the embodiment). When load is small load, the target change gear ratio value is controlled to increase it until a detected value of engine load becomes 100%. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a work vehicle such as a tractor used for agricultural work or a wheel loader used for civil engineering work, and more specifically, can control the output rotation speed from a hydraulic continuously variable transmission mounted on the work vehicle. It relates to control means.

  In recent work vehicles such as tractors or wheel loaders, for example, in Patent Document 1, the rotational power of an engine is driven via a hydrostatic continuously variable transmission (HST) and a gear-type sub-shift to improve work efficiency. A throttle lever configured to transmit and set the output rotational speed of the engine and a shift pedal for shifting the output rotational speed of the hydrostatic continuously variable transmission are provided, and the shift pedal is depressed. In this case, a configuration is disclosed in which the accelerator operating portion is pushed by the shift pedal, and the engine output rotational speed set and held by the throttle lever can be increased.

Further, in order to apply to a relatively narrow work area where the travel speed range (during forward movement) such as a work vehicle such as a tractor or a wheel loader is several Km / h to several tens Km / h, Patent Document 2 discloses the hydrostatic pressure. Instead of a continuously variable transmission, an input shaft that transmits power from the engine and an output shaft that transmits hydraulic shift output to the left and right wheels are concentrically arranged on both sides of the cylinder block constituting the continuously variable transmission. By adopting a so-called inline type continuously variable transmission in which a hydraulic pump part and a hydraulic motor part are arranged, and the input shaft and output shaft are configured as a double shaft, no hydraulic oil leakage occurs and power transmission efficiency It is disclosed that the fuel consumption of the engine can be reduced.
JP 2003-226165 A JP 2002-89655 A

  By the way, in a work vehicle such as a tractor, in addition to towing work equipment, it is suitable for various types of work such as field tillage / plowing work, crushed soil / leveling work, paddy field cultivation, multi-laying work, etc. It is necessary to travel in the range of torque and gear ratio.

  Therefore, the applicant can change and adjust the swash plate angle of the hydraulic pump in the inline type continuously variable transmission according to the depression amount of the shift pedal, and change rate of the transmission ratio corresponding to the depression amount of the transmission pedal. We thought about providing a gear ratio setting device that changes However, for example, in the case of loader work where a high load is suddenly applied during the work, the engine stall (hereinafter simply referred to as the engine stall) does not occur only by operating the gear ratio setting device, and conversely, For example, in the case of paddy field work where there is little fluctuation in load, there is a problem that it is impossible to prevent torque shortage.

    The present invention has been made in order to solve the above-described problems, and while changing the output rotational speed of a hydraulic continuously variable transmission by a simple operation of setting a gear ratio pattern, the load changes suddenly. In this case, it is an object of the present invention to provide a work vehicle control device that can easily perform load control capable of decelerating so that engine stall does not occur.

  In order to achieve the above object, a control device for a work vehicle according to a first aspect of the present invention includes an input shaft to which power from an engine is transmitted in a transmission case mounted on a traveling machine body, and the input shaft A hydraulic continuously variable transmission including a hydraulic pump, a hydraulic motor, and an output shaft arranged on the same axis and having a sub-transmission mechanism is provided to transmit at least traveling driving force from the output shaft via the hydraulic motor. In the work vehicle configured as described above, a gear ratio setting device that sets a gear ratio pattern of a gear ratio of the rotational speed of the output shaft in the hydraulic continuously variable transmission with respect to the rotational speed of the engine, an engine speed sensor, The output shaft rotational speed sensor, an electronic governor controller, and a load control means, wherein the load control means is in an ON state of a load control switch, and the auxiliary transmission mechanism is turned on at a low speed stage. Is, and the rotational speed of the engine is intended to be executed when less than a predetermined value.

  According to a second aspect of the present invention, in the work vehicle control device according to the first aspect, pattern storage means for storing a plurality of speed ratio patterns, a shift sensor for detecting the depression amount of the shift pedal, and the shift sensor And an actuator for adjusting a swash plate angle of the hydraulic pump based on a detected value of the engine, wherein the load control means has a reduction rate of the engine speed of less than a predetermined value A% and an engine load in the electronic governor controller. When the detected value is equal to or greater than a predetermined value B%, control is performed so as to maintain the target gear ratio on the gear ratio pattern set by the gear ratio setting device.

  According to a third aspect of the present invention, in the work vehicle control device according to the second aspect, the load control means is configured such that the rate of decrease in the engine speed is not less than the predetermined value A% and the detected value of the engine load. Is equal to or greater than the predetermined value B%, the target speed ratio is controlled to decrease until the rate of decrease in the engine speed becomes less than the predetermined value C%.

  According to a fourth aspect of the present invention, in the work vehicle control device according to the second or third aspect, the load control means is configured such that when the detected value of the engine load is less than the predetermined value B%, the detected value is Control is performed to increase the target gear ratio until 100% is reached.

  According to the first aspect of the present invention, the load control means is executed when the load control switch is ON, the auxiliary transmission mechanism is turned on at a low speed, and the engine speed is equal to or higher than a predetermined value. Is. In this load control, the traveling speed is automatically changed so that the engine is not overloaded, in other words, the engine stall does not occur, while the traveling speed at the time of the work which does not like the fluctuation of the engine speed is made low. Therefore, the traveling operation of the work vehicle can be made extremely simple, and the effect of reducing fatigue during long-time work can be achieved.

  According to the invention of claim 2, the running load is within an appropriate range, the rate of decrease in the engine speed is less than a predetermined value A%, and the detected value of the engine load in the electronic governor controller is a predetermined value B. When the ratio is greater than or equal to%, once the operator sets the gear ratio pattern of the gear ratio with the gear ratio setting device, the operator only needs to operate the gear pedal to change the actual gear ratio due to changes in the environment or fluctuations in the work vehicle's running load. When the ratio value deviates from the target gear ratio value, it can be automatically controlled so that it automatically approaches the target gear ratio value. There is an effect that can be reduced.

  According to the invention of claim 3, when the traveling load is in a large range, the rate of decrease in the engine speed is not less than the predetermined value A% and the detected value of the engine load is not less than the predetermined value B%. Since the target speed ratio is controlled to decrease until the rate of decrease in the engine speed becomes less than the predetermined value C%, the target speed ratio is automatically decreased when the traveling load is large. , Engine stall can be prevented.

  According to a fourth aspect of the present invention, when the detected value of the engine load is less than the predetermined value B% when the traveling load is in a small range, the target gear ratio is increased until the detected value reaches 100%. Therefore, by automatically increasing the traveling speed (transmission ratio) to a range where engine stall does not occur, it is possible to perform high-speed traveling work, and to increase the efficiency of farm work.

  DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, drawings when an embodiment of the present invention is applied to a farm tractor as a work vehicle will be described. 1 is a side view of a tractor, FIG. 2 is a rear perspective view of the tractor, FIG. 3 is a side view of the tractor, FIG. 4 is a perspective view of a traveling machine body, FIG. 5 is a skeleton diagram of power transmission, and FIG. 24, FIG. 24 is a gear ratio pattern diagram showing the relationship between the depression amount of the gear pedal and the gear ratio, FIG. 25 is a flowchart of gear ratio adaptive control, and FIG. 26 is a flowchart of load control.

  As shown in FIGS. 1 to 4, the tractor 1 as a work vehicle supports a traveling machine body 2 with a pair of left and right rear wheels 4 as well as a pair of left and right front wheels 3, and is mounted on a front portion of the traveling machine body 2. By driving the rear wheel 4 and the front wheel 3 at 5, the vehicle is configured to travel forward and backward. The engine 5 is covered with a bonnet 6. Further, a cabin 7 is installed on the upper surface of the traveling machine body 2. Inside the cabin 7, there is a steering seat 8 and a steering handle (round handle) 9 that moves the front wheel 3 left and right by steering. And are installed. A step 10 on which the operator gets on and off the cabin 7 is provided, and a fuel tank 11 for supplying fuel to the engine 5 is provided on the inner side of the cabin 10 and below the bottom of the cabin 7. .

  The traveling machine body 2 includes an engine frame 14 having a front bumper 12 and a front axle case 13, and left and right machine body frames 16 that are detachably fixed to the rear portion of the engine frame 14 with bolts 15. A mission case 17 is connected to the rear portion of the body frame 16 for transmitting the rotation of the engine 5 to the rear wheels 4 and the front wheels 3 with appropriate speed change. In this case, the rear wheel 4 is rearwardly attached to the transmission case 17 so as to protrude outward from the outer surface of the transmission case 17 and to the outer end of the rear axle case 18. It is attached via a gear case 19 that is mounted so as to extend.

  A hydraulic working machine lifting mechanism 20 for lifting and lowering a working machine (not shown) such as a tiller is detachably attached to the upper surface of the rear portion of the mission case 17. A working machine such as a tiller is connected to the rear part of the mission case 17 via a lower link 21 and a top link 22. Further, a PTO shaft 23 for the working machine such as the tiller is provided on the rear side surface of the mission case 17 so as to protrude rearward.

  FIG. 15 shows a hydraulic circuit 200 of the tractor 1 of the present embodiment, and includes a working machine hydraulic pump 94 and a traveling hydraulic pump 95 that are operated by the rotational force of the engine 5 as will be described later. The traveling hydraulic pump 95 is connected to a hydraulic cylinder 93 for power steering by the round handle 9 via a control valve 202 for power steering, while switching valves for operating the brake cylinders 68 for the left and right auto brakes 65 respectively. Are connected to left and right autobrake solenoid valves 67a and 67b. Further, the traveling hydraulic pump 95 is provided for the PTO clutch hydraulic solenoid valve (proportional control valve) 104 for the PTO clutch 100 and the transmission parts of the transmission case 17, that is, the hydraulic continuously variable transmission 29 for main transmission described later. Proportional control valve 203, switching valve 204 actuated thereby, shift shift electromagnetic valve 666 of traveling auxiliary shifting hydraulic cylinder 55, forward clutch for actuating hydraulic clutches 40, 42 for switching forward and reverse traveling To switch the solenoid valve 46, the reverse clutch solenoid valve 48, the four-wheel drive hydraulic solenoid valve 80 for the four-wheel drive hydraulic clutch 74 for simultaneously driving the front wheel 3 and the rear wheel 4, and the front wheel 3 to double speed drive. The double speed hydraulic solenoid valve 82 for the double speed hydraulic clutch 76 is connected. The work machine hydraulic pump 94 is connected to a control electromagnetic valve 201 for supplying hydraulic oil to the single-acting hydraulic cylinder 205 in the work machine lifting mechanism 20. As shown in FIG. 15, the hydraulic circuit 200 includes a relief valve, a flow rate adjusting valve, a check valve, an oil cooler, an oil filter, and the like.

  5 to 7 show the mission case 17. The mission case 17 is partitioned inside and behind by a partition wall 31. A front side wall member 32 and a rear side wall member 33 are detachably fixed to the front side and the rear side of the mission case 17 with bolts, respectively. The mission case 17 is formed in a box shape, and a front chamber 34 and a rear chamber 35 are formed inside the mission case 17. The front chamber 34 and the rear chamber 35 are in communication with each other so that the hydraulic oil (lubricating oil) inside these chambers moves relative to each other.

  As shown in FIG. 5, the front side wall member 32 is provided with a front wheel drive case 69 described later. In the front chamber 34, a travel auxiliary transmission gear mechanism 30 and a PTO transmission gear mechanism 96, which will be described later, are arranged. In the rear chamber 35, a hydraulic continuously variable transmission 29, which is a traveling main transmission mechanism described later, and a differential gear mechanism 58 are arranged.

  An engine output shaft 24 is provided on the rear side surface of the engine 5 so as to protrude rearward. A flywheel 25 is attached to the engine output shaft 24 so as to be directly connected. Between a main drive shaft 26 projecting rearward from the flywheel 25 and a main transmission input shaft 27 projecting forward from the front surface of the transmission case 17 via an extendable power transmission shaft 28 having a universal joint at both ends. Connect. The rotation of the engine 5 is transmitted to the main transmission input shaft 27 in the transmission case 17, and then appropriately shifted by the hydraulic continuously variable transmission 29 and the traveling auxiliary transmission gear mechanism 30, via the differential gear mechanism 58. The driving force is transmitted to the rear wheel 4. Further, the rotation of the engine 5 that has been appropriately shifted by the traveling auxiliary transmission gear mechanism 30 is transmitted to the front wheels 3 via the front wheel drive case 69 and the differential gear mechanism 86 of the front axle case 13. Yes.

  Next, FIG. 8 shows an inline hydraulic continuously variable transmission 29 in which a main transmission output shaft 36 is concentrically disposed on the main transmission input shaft 27. A hydraulic continuously variable transmission 29 is installed in the rear chamber 35 via a main transmission input shaft 27. The rear end side of the main transmission input shaft 27 opposite to the input side (front end side) of the main transmission input shaft 27 is rotatably supported by the rear side wall member 33 by a ball bearing 504.

  A cylindrical main transmission output shaft 36 is fitted on the front side of the hydraulic continuously variable transmission 29, that is, on the input side of the main transmission input shaft 27. A main transmission output gear 37 for taking out the main transmission output from the hydraulic continuously variable transmission 29 is provided on the main transmission output shaft 36. The middle of the main transmission output shaft 36 is penetrated by the partition wall 31, and the front end and the rear end protrude into the front chamber 34 and the rear chamber 35, respectively. The middle of the main transmission output shaft 36 is rotatably supported on the partition wall 31 by two sets of ball bearings 502. A main transmission output gear 37 is provided at the front end portion of the main transmission output shaft 36. The input side of the main transmission input shaft 27 protrudes forward from the front end of the main transmission output shaft 36 such that the input side of the main transmission input shaft 27 projects forward from the front end of the main transmission output shaft 36 via the roller bearing 503. (See FIG. 8).

  As will be described below, the hydraulic continuously variable transmission 29 includes a variable displacement type shifting hydraulic pump unit 500 and a constant displacement type shifting hydraulic pressure that operates with high-pressure hydraulic fluid discharged from the hydraulic pump unit 500. A motor unit 501. A cylinder block 505 for the hydraulic pump unit 500 and the hydraulic motor unit 501 is fitted on the main transmission input shaft 27 substantially in the middle between the partition wall 31 and the rear side wall member 33. The main transmission input shaft 27 and the cylinder block 505 are connected by a spline 525. The hydraulic pump unit 500 is disposed on one side of the main transmission input shaft 27 opposite to the input side with the cylinder block 505 interposed therebetween. A hydraulic motor unit 501 is disposed on the other side of the cylinder block 505 that is the input side of the main transmission input shaft 27.

  The hydraulic pump unit 500 includes a first holder 510 that is fixed to the inner surface of the transmission case 17 so as to face the side surface of the cylinder block 505, and a tilt angle that can be changed with respect to the axis of the main transmission input shaft 27. A pump swash plate 509 disposed in one holder 510, a shoe 508 slidably provided on the pump swash plate 509, a pump plunger 506 connected to the shoe 508 via a spherical universal joint, and the pump plunger 506 in a cylinder block 505 is provided with a first plunger hole 507 disposed so as to freely enter and exit. One end of the pump plunger 506 protrudes from the side surface of the cylinder block 505 toward the pump swash plate 509 (right side in FIG. 8). The hydraulic pump unit 500 includes a cylinder block 505, a pump plunger 506, a shoe 508, a pump swash plate 509, and a first holder 510.

  A sleeve 511 fitted to the main transmission input shaft 27, a roller bearing 512, and radial and thrust load roller bearings 513 are interposed between the main transmission input shaft 27 and the first holder 510. A nut 514 is provided behind the main transmission input shaft 27 to prevent the roller bearing 513 from coming out.

  The cylinder block 505 is provided with the same number of first spool valves 536 as the pump plungers 506. In addition, a first radial bearing 537 is disposed in the first holder 510. The first radial bearing 537 is provided in the first holder 510 so as to be inclined at a constant inclination angle with respect to the axis of the main transmission input shaft 27. In FIG. 8, the position rotated about 90 degrees with respect to the pump swash plate 509 (the front side of the drawing of FIG. 8) is separated from the side surface of the cylinder block 505 by about 180 degrees (the back side of the drawing of FIG. 8). The first radial bearing 537 is inclined and supported such that the first radial bearing 537 is close to the side surface of the cylinder block 505.

  On the other hand, the hydraulic motor unit 501 includes a second holder 519 arranged to face the side surface of the cylinder block 505, and a second holder 519 so as to maintain a constant inclination angle with respect to the axis of the main transmission input shaft 27. A motor swash plate 518 to be fixed, a shoe 517 slidably provided on the motor swash plate 518, a motor plunger 515 connected to the shoe 517 via a spherical universal joint, and the motor plunger 515 to be able to enter and exit the cylinder block 505 A second plunger hole 516 is provided. One end of the motor plunger 515 protrudes from the side surface of the cylinder block 505 in the direction of the motor swash plate 518 (left side in FIG. 8).

  A joint member 526 is fixed to the second holder 519 with a bolt 527. The output shaft 36 and the joint member 526 are connected by a spline 528.

  Between the main transmission input shaft 27 and the second holder 519, radial load roller bearings 520 and 521, a sleeve 522 fitted to the main transmission input shaft 27, a radial and thrust load roller bearing 523, Intervenes. A nut 524 is provided to prevent the roller bearing 523 from coming out of the main transmission input shaft 27.

  The cylinder block 505 is provided with the same number of second spool valves 540 as the motor plungers 515. In addition, a second radial bearing 541 is disposed in the second holder 519. The second radial bearing 541 is provided in the second holder 519 so as to be inclined at a constant inclination angle with respect to the axis of the main transmission input shaft 27. In FIG. 8, the cylinder block 505 is on the opposite side (back side in FIG. 8) about 180 degrees so that the position rotated about 90 degrees with respect to the motor swash plate 518 (front side in FIG. 8) is close to the side surface of the cylinder block 505. The second radial bearing 541 is configured to be tilted and supported so as to be separated from the side surface of the first radial bearing. The pump plungers 506 and the same number of motor plungers 515 as the pump plungers 506 are alternately arranged on the same circumference of the rotation center of the cylinder block 505.

  Further, a ring groove-shaped first oil chamber 530 and a ring groove-shaped second oil chamber 531 are formed in the shaft hole of the cylinder block 505 into which the main transmission input shaft 27 is inserted. The cylinder block 505 is formed with a first valve hole 532 and a second valve hole 533 that are arranged at substantially equal intervals on the same circumference of the rotation center. The first valve hole 532 and the second valve hole 533 communicate with the first oil chamber 530 and the second oil chamber 531, respectively. The first plunger hole 507 communicates with the first valve hole 532 via the first oil passage 534, and the second plunger hole 516 communicates with the second valve hole 533 via the second oil chamber 531.

  A first spool valve 536 is inserted into the first valve hole 532. The first spool valves 536 are arranged at substantially equal intervals on the same circumference of the rotation center of the cylinder block 505. The tip of the first spool valve 536 protrudes in the direction of the first holder 510 from the first valve hole 532 with the back pressure spring force, and the tip of the first spool valve 536 abuts the side surface of the outer ring 538 of the first radial bearing 537. Be touched. Then, the first spool valve 536 reciprocates once in one rotation of the cylinder block 505, and the first plunger hole 507 is connected to the first oil chamber 530 or the second oil via the first valve hole 532 and the first oil passage 534. The chamber 531 is configured to communicate with each other alternately.

  A second spool valve 540 is inserted into the second valve hole 533. The second spool valves 540 are arranged at substantially equal intervals on the same circumference of the rotation center of the cylinder block 505. The tip of the second spool valve 540 protrudes in the direction of the second holder 519 from the second valve hole 533 due to the back pressure spring force, and the tip of the second spool valve 540 contacts the side surface of the outer ring 542 of the second radial bearing 541. Be touched. Then, the second spool valve 540 reciprocates once by one rotation of the cylinder block 505, and the second plunger hole 516 passes through the second valve hole 533 and the second oil passage 535, and passes through the first oil chamber 530 or the second oil. The chamber 531 is configured to communicate with each other alternately.

  Further, a hydraulic oil supply oil passage 543 is formed in the central portion of the main transmission input shaft 27 in the axial direction. The supply oil passage 543 is opened at the rear end face of the main transmission input shaft 27 and communicates with the discharge port of the traveling hydraulic pump 95 described above. Further, the first charge valve 544 for supplying hydraulic oil in the hydraulic oil supply oil passage 543 to the first oil chamber 530 and the second charge valve 545 for supplying hydraulic oil in the hydraulic oil supply oil passage 543 to the second oil chamber 531. And are provided.

  And with respect to the hydraulic closed circuit formed between the first and second plunger holes 507 and 516 and the first and second oil chambers 530 and 531, the first and second charge valves 544 and 545 are passed through, The hydraulic oil is supplied from the hydraulic oil supply oil passage 543. It should be noted that hydraulic oil is supplied from the hydraulic oil supply oil passage 543 to the rotating portions of the hydraulic pump unit 500 and the motor unit 501 as lubricating oil via check valves.

  Further, the pump swash plate 509 is disposed on the outer periphery of the small diameter portion of the first holder 510 via an inclination angle adjustment fulcrum 555 as described later (see FIG. 13). The pump swash plate 509 is provided such that its inclination angle is adjustable with respect to the axis of the main transmission input shaft 27. A main transmission hydraulic cylinder 556 for main transmission operation, which is a transmission actuator for changing the inclination angle of the pump swash plate 509 with respect to the axis of the main transmission input shaft 27, is provided (see FIG. 13). The main transmission hydraulic cylinder 556 is configured to change the inclination angle of the pump swash plate 509 so that the main transmission operation of the continuously variable transmission 29 is performed. The first holder 510 is coupled to the rear side wall member 33 that is a non-rotating portion of the transmission case 17 via the holder coupling member 690 so that the pump swash plate 509 does not rotate with respect to the main transmission input shaft 27. (See FIG. 13).

  The main transmission operation of the inline hydraulic continuously variable transmission 29 will be described below. A switching valve 204 is operated by hydraulic oil from a proportional control solenoid valve 203 that operates in proportion to the amount of depression of a forward pedal 232 or a reverse pedal 233, which will be described later, and the hydraulic cylinder 556 is controlled. The inclination angle of the pump swash plate 509 is changed with respect to the 27 axis.

  First, even if the cylinder block 505 is rotated when the tilt angle of the pump swash plate 509 is kept substantially zero so that the pump swash plate 509 is substantially orthogonal to the axis of the main transmission input shaft 27, the first plunger hole The pump plunger 506 is supported in a substantially constant posture at 507 so that the pump plunger 506 does not move forward and backward, and the hydraulic oil in the first plunger hole 507 is not discharged from the first oil passage 534 toward the first valve hole 532 in the discharge stroke of the pump plunger 506. The hydraulic oil is not supplied from the first plunger hole 507 to the second plunger hole 516, and the motor plunger 515 does not advance. Further, since the hydraulic oil is not drawn into the first plunger hole 507 even during the suction stroke of the pump plunger 506, the hydraulic oil is not discharged from the second plunger hole 516 into the first plunger hole 507, and the motor plunger 515 does not retract.

  That is, when the inclination angle of the pump swash plate 509 is substantially zero, the hydraulic motor unit 501 is not driven by the hydraulic pump unit 500. Therefore, the motor swash plate 518 is fixed to the cylinder block 505 via the motor plunger 515, and the cylinder block 505 and the motor swash plate 518 rotate at the same rotational speed in the same direction. The main transmission output shaft 36 is rotated at substantially the same rotational speed, and the rotational speed of the main transmission input shaft 27 is transmitted to the main transmission output gear 37 without being changed.

  Next, as shown in FIG. 9, when the pump swash plate 509 is inclined with respect to the axis of the main transmission input shaft 27, the first radial bearing is rotated by the rotation of the cylinder block 505 that rotates integrally with the main transmission input shaft 27. The first spool valve 536 slides back and forth at the outer ring 538 of 537, and the first oil chamber 530 or the second oil chamber 531 communicates with the first plunger hole 507 alternately every half rotation of the cylinder block 505. Further, the second spool valve 540 slides back and forth at the outer ring 542 of the second radial bearing 541, and the first oil chamber 530 or the second oil chamber 531 alternates with the second plunger hole 5016 every half rotation of the cylinder block 505. Communicated with A closed hydraulic circuit is formed between the first plunger hole 507 and the second plunger hole 516, and hydraulic oil is pumped from the first plunger hole 507 to the second plunger hole 516 in the discharge stroke of the pump plunger 506, The hydraulic fluid is returned from the second plunger hole 516 to the first plunger hole 507 during the suction stroke of the pump plunger 506, and the operation of the axial piston pump and the motor is performed.

  When the pump swash plate 509 is tilted in one direction (positive tilt angle) with respect to the axis of the main transmission input shaft 27, the motor swash plate 518 is rotated in the same direction as the cylinder block 505, and the hydraulic motor unit The main transmission output shaft 36 is rotated at a higher rotational speed than the main transmission input shaft 27, and the rotational speed of the main transmission input shaft 27 is increased and transmitted to the main transmission output gear 37. It is done. That is, the rotational speed of the hydraulic motor unit 501 driven by the hydraulic pump unit 500 is added to the rotational speed of the main transmission input shaft 27 and transmitted to the main transmission output gear 37. Therefore, the shift output (travel speed) from the main shift output gear 37 is proportional to the inclination (positive inclination angle) of the pump swash plate 509 within the range of the rotation speed higher than the rotation speed of the main transmission input shaft 27. As a result, the maximum traveling speed is reached at the maximum inclination (positive inclination angle) of the pump swash plate 509.

  Further, when the pump swash plate 509 is inclined in the other direction (negative inclination angle) with respect to the axis line of the main transmission input shaft 27, the motor swash plate 518 is rotated in the direction opposite to the cylinder block 505, and the hydraulic motor The part 501 is decelerated (reversely rotated), the main transmission output shaft 36 is rotated at a lower rotational speed than the main transmission input shaft 27, and the rotational speed of the main transmission input shaft 27 is decelerated and transmitted to the main transmission output gear 37.

  That is, the rotational speed of the hydraulic motor unit 501 driven by the hydraulic pump unit 500 is subtracted from the rotational speed of the main transmission input shaft 27 and transmitted to the main transmission output gear 37. Therefore, the shift output (travel speed) from the main shift output gear 37 is proportional to the inclination (negative inclination angle) of the pump swash plate 509 within the range of the rotation speed lower than the rotation speed of the main transmission input shaft 27. As a result, the minimum traveling speed is reached at the maximum inclination (negative inclination angle) of the pump swash plate 509. In the embodiment, when the negative inclination angle of the pump swash plate 509 is approximately 11 degrees, the gear ratio is zero. Although slightly different depending on the gear ratio pattern described later, the gear ratio is set to be maximum when the positive inclination angle is approximately 11 degrees.

  Next, as shown in FIGS. 5 and 6, the front chamber 34 of the mission case 17 has a forward gear 41 and a reverse gear 43 for switching between forward and reverse, and a traveling auxiliary gear for switching between low speed and high speed. A transmission gear mechanism 30 is disposed.

  A description will be given of switching between forward and reverse movements through the forward gear 41 and the reverse gear 43. As shown in FIG. 6, a travel counter shaft 38 and a reverse rotation shaft 39 are disposed in the front chamber 34 where the main transmission output gear 37 is disposed. The travel counter shaft 38 is fitted with a forward gear 41 connected by a forward wet multi-plate hydraulic clutch 40 and a reverse gear 43 connected by a reverse wet multi-plate hydraulic clutch 42. Is done. The forward gear 41 is meshed with the main transmission output gear 37. The main transmission output gear 37 is engaged with a reverse gear 44 provided on the reverse shaft 39. The reverse gear 43 meshes with a reverse output gear 45 provided on the reverse shaft 39.

  Then, when the forward pedal 232 described later is depressed, the clutch cylinder 47 is operated by the forward clutch solenoid valve 46 to continue the forward hydraulic clutch 40, and the main transmission output gear 37 and the travel counter shaft 38 are connected to the forward gear 41. (See FIGS. 5 and 6).

  On the other hand, when the reverse pedal 233, which will be described later, is depressed, the clutch cylinder 49 is operated by the reverse clutch solenoid valve 48 to continue the reverse hydraulic clutch 42, and the main transmission output gear 37 and the travel counter shaft 38 are connected to the reverse gear 43. (See FIGS. 5 and 6).

  In the neutral position where neither the forward pedal 232 nor the reverse pedal 233 is depressed, both the forward and reverse wet multi-plate hydraulic clutches 40 and 42 are both disconnected, and the front wheel 3 and the rear The travel driving force from the main transmission output gear 37 output to the wheels 4 is configured to be substantially zero (main clutch disengaged state).

  Next, switching between the low speed and the high speed performed through the traveling auxiliary transmission gear mechanism 30 will be described. As shown in FIGS. 5 and 6, a traveling auxiliary transmission gear mechanism 30 and an auxiliary transmission shaft 50 are disposed in the front chamber 34 of the transmission case 17. Between the traveling counter shaft 38 and the auxiliary transmission shaft 50, low-speed gears 51 and 52 for auxiliary transmission and high-speed gears 53 and 54 for auxiliary transmission are provided. Further, a low speed clutch 56 and a high speed clutch 57 which are continued or disconnected by the auxiliary transmission hydraulic cylinder 55 are provided. The low speed clutch 56 or the high speed clutch 57 is continued in the auxiliary speed change hydraulic cylinder 55 by operating the auxiliary speed change lever (not shown) or detecting the rotational speed of the engine 5. The high-speed gear 54 is connected, and the driving force is output from the auxiliary transmission shaft 50 to the front wheels 3 and the rear wheels 4.

  The rear end of the auxiliary transmission shaft 50 extends through the partition wall 31 and extends into the rear chamber 35 of the transmission case 17 (see FIG. 5). A pinion 59 is provided at the rear end portion of the auxiliary transmission shaft 50. In addition, a differential gear mechanism 58 that transmits traveling driving force to the left and right rear wheels 4 is disposed in the rear chamber 35. The differential gear mechanism 58 includes a ring gear 60 that meshes with a pinion 59 at the rear end of the auxiliary transmission shaft 50, a differential gear case 61 provided in the ring gear 60, and left and right differential output shafts 62. The differential output shaft 62 is connected to the rear axle 64 by a final gear 63 or the like, and is configured to drive the rear wheel 4 provided on the rear axle 64 (see FIG. 5).

  Further, a brake 65 is installed on the differential output shaft 62, and the brake 65 is configured to be braked by operating the left and right brake pedals 66. On the other hand, when the steering angle of the handle 9 is detected, the brake cylinder 68 is operated by the left and right autobrake solenoid valves 67a and 67b, and the brake 65 (see FIG. 5, but only one is shown) is automatically braked. It is configured to perform a turning run such as a U-turn.

  Next, switching between the 2WD and 4WD of the front and rear wheels 3 and 4 will be described. As shown in FIGS. 5 and 6, a front wheel drive case 69 is provided on the front side wall member 32 of the mission case 17. The front wheel drive case 69 is provided with a front wheel input shaft 72 and a front wheel output shaft 73. The front wheel input shaft 72 is connected to the auxiliary transmission shaft 50 by gears 70 and 71. The front wheel output shaft 73 is fitted with a four-wheel drive gear 75 connected by a four-wheel drive hydraulic clutch 74 and a double speed gear 77 connected by a double speed hydraulic clutch 76. The four-wheel drive gear 75 and the double speed gear 77 are connected to the front wheel input shaft 72 by gears 78 and 79, respectively.

  Then, by the four-wheel drive operation of the switching lever (not shown) of the two-wheel drive and the four-wheel drive, the clutch cylinder 81 is operated by the four-wheel hydraulic solenoid valve 80 and the four-wheel drive hydraulic clutch 74 is continued, and the front wheel input shaft 72 and the front wheel output shaft 73 are connected by a four-wheel drive gear 75, and the front wheel 3 is driven together with the rear wheel 4.

  Next, switching of the double speed drive of the front wheel 3 will be described. As shown in FIGS. 5 and 6, the clutch cylinder 83 is actuated by the double-speed hydraulic solenoid valve 82 upon detection of a U-turn (direction change at the headland in the field) operation of the steering handle 9, and the double-speed hydraulic pressure is actuated. The clutch 76 is continued, the front wheel input shaft 72 and the front wheel output shaft 73 are connected by the double speed gear 77, and about twice as fast as the speed when the front wheel 3 is driven by the four-wheel drive gear 75. The front wheel 3 is configured to be driven at a high speed.

  As shown in FIG. 5, the front wheel 3 is powered between a front wheel input shaft 84 projecting rearward from the front axle case 13 and a front wheel output shaft 73 projecting forward from the front surface of the transmission case 17. Are connected via a front wheel drive shaft 85 that transmits In addition, a differential gear mechanism 86 that transmits traveling driving force to the left and right front wheels 3 is disposed inside the front axle case 13.

  The differential gear mechanism 86 includes a ring gear 88 that meshes with a pinion 87 at the front end of the front wheel input shaft 84, a differential gear case 89 provided in the ring gear 88, and left and right differential output shafts 90. A front axle 92 is connected to the differential output shaft 90 by a final gear 91 or the like, and a front wheel 3 provided on the front axle 92 is driven. Further, on the outer surface of the front axle case 13, a hydraulic cylinder 93 for power steering that changes the traveling direction of the front wheel to the right and left by the steering operation of the steering handle 9 is disposed.

  As shown in FIGS. 5 and 7, on the front side of the front side wall member 32 of the mission case 17, a work machine hydraulic pump 94 for supplying hydraulic fluid to the work machine lifting mechanism 20, and the mission case 17. And a traveling hydraulic pump 95 for supplying hydraulic oil to the hydraulic cylinder 93 for power steering. A transmission case 17 is used in combination as an oil tank so that hydraulic oil in the case 17 is supplied to the hydraulic pumps 94 and 95.

  Next, switching of the driving speed of the PTO shaft 23 (four forward rotations and one reverse rotation) will be described with reference to FIGS. The front chamber 34 of the transmission case 17 is provided with a PTO transmission gear mechanism 96 that transmits power from the engine 5 to the PTO shaft 23 and a pump drive shaft 97 that transmits power from the engine 5 to the hydraulic pumps 94 and 95 ( (See FIG. 7).

  As shown in FIG. 7, the PTO transmission gear mechanism 96 described in detail later includes a PTO counter shaft 98 and a PTO transmission output shaft 99. The PTO input gear 101 connected by the PTO hydraulic clutch 100 is fitted on the PTO counter shaft 98. An input side gear 102 provided on the main transmission input shaft 27 and an output side gear 103 of the pump drive shaft 97 are meshed with the PTO input gear 101, and the pump drive shaft 97 is connected to the main transmission input shaft 27.

  Then, by continuing the operation of the PTO clutch lever (not shown), the clutch cylinder 105 is operated by the PTO clutch hydraulic solenoid valve 104 (see FIG. 5), and the PTO hydraulic clutch 100 is continued. The PTO counter shaft 98 is connected to the PTO input gear 101.

  Further, a first speed gear 106, a second speed gear 107, a third speed gear 108, a fourth speed gear 109, and a reverse gear 110 are fitted on the PTO speed change output shaft 99 for PTO output (FIG. 5, see FIG.

  A shift shifter 111 is pivotally supported on the PTO shift output shaft 99 by a spline. The gears 106, 107, 108, 109, 110 are configured to be alternatively connected to the PTO shift output shaft 99 by a shift shifter 111. A shift arm 112 connected to a PTO shift lever (not shown) is engaged with the shift shifter 111. Then, a shift operation of the PTO shift lever (not shown) causes the shift shifter 111 to move linearly along the axis of the PTO shift output shaft 99 by the shift arm 112, and each gear 106, 107, 108, 109, Any one of 110 is selected and connected to the PTO shift output shaft 99 (see FIGS. 5 and 7). Therefore, the first, second, third, fourth, and reverse PTO shift outputs are transmitted from the PTO shift output shaft 99 to the PTO shaft 23 via the gears 113 and 114.

  In FIG. 6, a rotation detection gear 115 provided on the reverse rotation shaft 39 and an electromagnetic pickup type main transmission output portion rotation sensor 116 for detecting the rotation of the main transmission output gear 37 are installed to face each other, and the main transmission mechanism The output speed of 29 is detected by the main transmission output unit rotation sensor 116. Further, an electromagnetic pickup type vehicle speed sensor 117 for detecting the rotation of the gear 78 of the front wheel input shaft 72 is installed, and the traveling speed (vehicle speed) is determined based on the rotation of the front wheel input shaft 72 and the auxiliary transmission shaft 50. It is comprised so that it may be detected by.

  As is apparent from the above description and FIG. 8 and the like, a transmission case 17 is provided that transmits power from the engine 5, and a hydraulic shift output is provided to the input shaft 27 that transmits power from the engine 5 and the left and right wheels 3 and 4. An inline-type continuously variable transmission 29 arranged on the same axis line as the output shaft 36 to be transmitted is disposed in the transmission case 17, and hydraulic pressure is applied to one side across the cylinder block 505 constituting the continuously variable transmission 29. In a work vehicle in which the hydraulic motor unit 501 is disposed on the other side of the pump unit 500 and the output shaft 36 is fitted on the input shaft 27 to form a double shaft configuration, the input side of the input shaft 27 and the cylinder block 505 are arranged. The hydraulic motor 501 is disposed between the input side of the input shaft 27 and the output side of the output shaft 36 on the same side. Therefore, for example, even if the traveling auxiliary transmission gear, the differential gear, the PTO transmission gear, and the like are installed inside the transmission case 17 as in the transmission structure of the tractor 1, a continuously variable transmission is provided at the rear of the transmission case 17. The installation space for the machine 29 can be easily secured. An installation space such as a PTO transmission gear or a traveling auxiliary transmission gear is secured in the front portion of the transmission case 17 on the input side of the input shaft 27. For example, the transmission case 17 of the tractor 1 can be reduced in size or weight, and the manufacturing cost can be reduced. Can be reduced.

  The differential gear mechanism 58 includes a differential lock mechanism (not shown) that stops the differential operation (the left and right differential output shafts 62 are always driven at a constant speed). When the lock pin provided in the differential gear case so as to freely enter and exit is engaged with the differential gear by operating a differential lock lever (or pedal) (not shown), the differential gear is fixed to the differential gear case. The differential function is stopped, and the left and right differential output shafts 62 are driven at a constant speed.

  Next, the structure of the main transmission hydraulic cylinder 556 for shifting the continuously variable transmission 29 will be described in detail with reference to FIGS. 9, 11, 13, and 14. A cylinder chamber 691 of the main transmission hydraulic cylinder 556 is formed in the rear side wall member 33. The piston 557 of the main transmission hydraulic cylinder 556 is disposed in the cylinder chamber 691 of the main transmission hydraulic cylinder 556 so as to be slidable in the vertical direction. A square columnar base end pin 693 is engaged with a recess 692 formed in the outer periphery of the piston 557 in the middle. A square columnar base end pin 693 is rotatably disposed on one end side of the main transmission arm 558. The middle of the main transmission arm 558 is rotatably supported on the holder connecting member 690 via the arm shaft 694. A square columnar tip pin 696 is slidably engaged with an arm groove 695 on the other end side of the main transmission arm 558. A quadrangular prism-shaped tip pin 696 is rotatably supported on a semicircular tilt angle adjustment fulcrum 555 of the pump swash plate 509. A rotation guide 697 for supporting the tilt angle adjustment fulcrum 555 is disposed in the first holder 510. The rotation guide 697 forms a semi-cylindrical surface concentric with the rotation center of the pump swash plate 509. The tilt angle of the pump swash plate 509 is changed by the guidance of the rotation guide 697.

  As shown in FIGS. 11 and 12, the PTO shaft 23 is arranged in the center of the left and right width of the mission case 17 (one tractor body). A differential gear mechanism 58 is arranged on the right side of the PTO shaft 23 in the traveling direction (FIG. 10). A continuously variable transmission 29 is arranged obliquely above the left side of the PTO shaft 23 in the traveling direction. A piston 557 is arranged on the left side of the continuously variable transmission 29 in the traveling direction. A main transmission hydraulic cylinder 556 is disposed at the corner of the rear side wall member 33 obliquely above the left side in the traveling direction.

  As shown in FIG. 13, the main transmission arm 558 and the arm shaft 694 are disposed at substantially the same height as the axis of the continuously variable transmission 29. As shown in FIG. 14, the main transmission arm 558 is disposed on the left side of the continuously variable transmission 29 in the traveling direction and substantially parallel to this axis. The piston 557 is installed substantially vertically in the rear side wall member 33 so as to slide in the vertical direction. The piston 557 can be installed only by forming the main transmission hydraulic cylinder 556 formation portion of the rear side wall member 33 slightly larger than the diameter of the piston 557.

  The speed change operation of the main speed change hydraulic cylinder 556 will be described. When the corresponding forward clutch solenoid valve 46 or reverse clutch solenoid valve 48 (see FIGS. 5, 6 and 23) is switched by a stepping operation of a forward pedal 232 or a reverse pedal 233, which will be described later, the main transmission hydraulic cylinder 556 is activated. . When the piston 557 moves up or down, the main transmission arm 558 rotates around the arm shaft 694, the tilt angle adjustment fulcrum 555 and the rotation guide 697 rotate and guide the pump swash plate 509, and the pump swash plate The configuration is such that the main transmission operation of the continuously variable transmission 29 is performed by changing the inclination angle of 509. Pump swash plate 509 and first holder 510 are connected to main transmission input shaft 27 so that pump swash plate 509 does not rotate, and first holder 510 and holder connecting member 690 are connected.

  Next, referring to FIGS. 5, 10, 11, and 12, the forward clutch solenoid valve 46, the reverse clutch solenoid valve 48, the left and right autobrake solenoid valves 67a and 67b, the four-wheel drive hydraulic solenoid valve 80, and the double speed The mounting structure of the hydraulic solenoid valve 82 and the PTO clutch hydraulic solenoid valve 104 will be described in detail.

  As shown in FIGS. 10 to 12, a base member 650 is disposed on the rear side surface of the front side wall member 32 at a position lower than the auxiliary transmission shaft 50, the differential output shaft 62, and the hydraulic oil 566, and the bolt It is detachably fixed via. The electromagnetic valves 46, 48, 67 a, 67 b, 80, 82, 104 are installed on the rear surface of the base member 650 so as to protrude rearward. A flat plate cover 651 covers the rear surface of each solenoid valve 46, 48, 67 a, 67 b, 80, 82, 104. The flat cover 651 is detachably fixed to the base member 650 with bolts.

  As shown in FIGS. 10 and 11, the oil filter 652 for filtering the hydraulic oil has a mission behind it with a flat plate lid 651 sandwiched between the solenoid valves 46, 48, 67 a, 67 b, 80, 82, 104. Arranged in the case 17. The oil filter 652 is detachably fixed to the filter lid 653. The filter lid 653 is formed integrally with the fastening member 654. A fastening member 654 is detachably fixed to the outer surface of the mission case 17 via a bolt 655. The oil supply pipe 656 of the working machine hydraulic pump 94 and the traveling hydraulic pump 95 is communicated with the filter lid 653 via the oil passage pipe 657.

  In each clutch cylinder 47, 49, 81, 83, 105 shown in FIG. 5, each solenoid valve 46, 48, 80, 82, 104 is a perforated oil passage formed in the front wall member 32 and the base member 650 ( (Not shown). When the respective solenoid valves 46, 48, 80, 82, 104 are controlled by appropriate means, the respective clutch cylinders 47, 49, 81, 83, 105 are operated, and the respective clutches 40, 42, 74 shown in FIG. , 76, 100 are respectively switched.

  Next, the speed change structure of the auxiliary transmission gear mechanism 30 will be described in detail with reference to FIGS. As shown in FIG. 9, the auxiliary transmission hydraulic cylinder 55 is configured in a double-acting structure in which a piston rod 661 is provided on one side of the piston 660. The auxiliary transmission hydraulic cylinder 55 is formed with a first cylinder chamber 662 in which a piston rod 661 is provided and the other second cylinder chamber 663. An auxiliary transmission shifter 665 is connected to the tip of the piston rod 661 via a shift arm 664. The low speed clutch 56 or the high speed clutch 57 is continued by the auxiliary transmission shifter so that the auxiliary transmission shaft 50 is driven at low speed or high speed.

  The first cylinder chamber 662 is in direct communication with the discharge side of the traveling hydraulic pump 95. The second cylinder chamber 663 communicates with the discharge side of the traveling hydraulic pump 95 via a two-position, three-port shift shift switching valve 666. The shift shift switching valve 666 includes a shift solenoid 667. When the shift shift switching valve 666 is switched by the shift solenoid 667 and the second cylinder chamber 663 is communicated to the discharge side of the traveling hydraulic pump 95 via the shift shift switching valve 666, the difference in pressure receiving pressure on both sides of the piston 660 Thus, the piston 660 moves in the direction in which the piston rod 661 protrudes, and the high speed clutch 57 is continued to drive the auxiliary transmission shaft 50 at high speed (see FIG. 9).

  Next, traveling control (shift control) of the work vehicle (traveling vehicle) of the present embodiment will be described. FIG. 23 is a functional block diagram of the travel control means. A travel controller 210 such as a microcomputer having a ROM storing a control program and a RAM storing various data is connected to a battery via a power application key switch 211. (Not shown). The key switch 211 is connected to a starter 212 for starting the engine 5.

  The travel controller 210 is connected to an electronic governor controller 213 that controls the rotation of the engine 5. The electronic governor controller 213 is connected to a governor 214 that adjusts the fuel of the engine 5 and an engine rotation sensor 215 that detects the rotational speed of the engine 5. When a fuel adjustment rack (not shown) provided in the governor 214 detects the rotational position of the manually operated throttle lever 206 with the throttle potentiometer 217, and the rotational speed of the engine 5 is set based on the detected value The fuel adjustment rack is automatically operated via an electromagnetic solenoid (not shown) for driving the fuel adjustment rack so that the set rotation speed of the throttle lever 206 and the rotation speed of the engine 5 coincide with each other by a signal from the electronic governor controller 213. The position is adjusted. It should be noted that the rotation of the engine 5 can be prevented from changing due to a load change or the like, in other words, the engine 5 can be controlled to maintain a substantially constant rotation regardless of the load change.

  Furthermore, as shown in FIG. 23, the travel controller 210 includes various sensors and switches of the input system, that is, a steering potentiometer 218 that detects the rotation amount (steering angle) of the round handle (steering handle) 9, and an operator Forward potentiometer 219 and reverse potentiometer 220 as shift potentiometers for shifting the traveling speed, a gear ratio setting dial 221, a changeover switch 222 for switching the auxiliary transmission gear mechanism 30 between high speed and low speed, and main shift output Main shift output unit rotation sensor 116 for detecting the number of rotations of the unit, a PTO cutoff switch 223 for blocking output to the PTO shaft, and a brake pedal switch 231 for detecting when the brake pedal 230 is depressed. And are connected. The forward potentiometer 219 is a pedal depression position sensor that converts and detects the depression amount of the forward pedal 232 as a shift pedal, and the reverse potentiometer 220 is a pedal that converts and detects the depression amount of the reverse pedal 233 as a transmission pedal. It is a depression position sensor.

  As shown in FIG. 23, the travel controller 210 includes a forced operation button 226 for forcibly turning off various output solenoid valves, that is, the forward clutch solenoid valve 46 and the reverse clutch solenoid valve 48 of the main transmission mechanism. And a high-speed clutch solenoid valve 666 for switching the sub-shift between high speed and low speed, and a proportional control solenoid valve that operates the main transmission hydraulic cylinder 556 in proportion to the amount of depression of a shift pedal (forward pedal 232 and reverse pedal 233) described later. 203 and left and right brake electromagnetics 67a and 67b are connected.

  In the present embodiment, a round handle (steering handle) 9 is disposed on the steering column 234 protruding from the floor plate 235 in front of the steering seat 8 in the driving section (cabin) 7 shown in FIG. A brake pedal 230 is disposed. A throttle lever 206 is arranged on the right side of the steering column 234. A forward pedal 232 and a reverse pedal 233 are arranged in parallel on the right side of the steering column 234. A vehicle speed setting dial 211, a sub shift changeover switch 222, and a PTO cutoff switch 223 are disposed on the right column of the control seat 8, and a PTO shift lever 224 is disposed on the left column of the control seat 8. A differential lock pedal 225 is disposed in front of the left column of the control seat 8. In addition, the floor board 235 forms the substantially whole upper surface in a flat surface.

  With reference to FIGS. 17 to 22, the mounting structure of the forward pedal 232 and the reverse pedal 233 will be described.

  As shown in FIGS. 19 and 21, the forward pedal 232 and the reverse pedal 233 are rotatably fitted to the brake pedal shaft 255 with the pivot support shaft portions 237 a and 237 b at the base ends of the pedal arms 232 a and 233 a. . The tread plates 236a and 236b (or the pedal arms 232a and 233a) of the forward pedal 232 and the reverse pedal 233 rotate obliquely downward from the initial (neutral) position on the upper surface of the floor plate 235 around the pivot support shaft portions 237a and 237b. It is installed as possible. A transmission link mechanism 275 is provided that transmits the pedal depression amounts of the forward pedal and the reverse pedal to the shift potentiometer 220 that is a shift sensor.

  As shown in FIGS. 18 to 21, the transmission link mechanism 275 includes a pair of check links 238 a and 238 b that respectively connect the forward pedal 232 and the reverse pedal 233 to a cam plate 258 described later, and the forward pedal 232 and the reverse pedal 233. The pedal depression force is increased when the pedal depression amount (or depression angle θ) of the stepping plates 236a and 236b becomes equal to or greater than the neutral position restoring means 241 for returning the gear to the neutral position (position where the shift output is substantially zero). Depression resistance changing means 242 is provided. A transmission frame 266 for installing the neutral position restoring means 241 and the stepping resistance changing means 242 is disposed at the attachment portion of the steering column 234.

  As shown in FIGS. 19 and 20, link arms 239a and 239b are respectively installed on the respective rotation support shaft portions 237a and 237b, and one end portion of each check link 238a and 238b is connected to each link via the link shafts 268a and 268b. The arms 239a and 239b are pivotally connected to each other. The other ends of the check links 238a and 238b are rotatably connected to an intermediate portion of a cam plate 258 described later via a support shaft 269. When both the forward and reverse treads 236a and 236b are supported in the initial (neutral) position, the forward check link 238a and the link arm 239a, and the reverse check link 238b and the link arm 239b are provided. The links 238a and 238b and the link arms 239a and 239b are respectively arranged at positions (seesaw structure) that are substantially symmetrical with respect to a straight line connecting the brake pedal shaft 255 and the support shaft 269. The above-described initial (neutral) position is a shift neutral position where the pedal depression amount is substantially zero, that is, a shift neutral position where the shift drive output from the continuously variable transmission 29 is substantially zero.

  Therefore, when the operator steps on one of the stepping plates 236a (236b) of the forward pedal 232 or the reverse pedal 233, the stepping plate 236a (236b) on the stepped-on side moves in the stepping direction (slanting forward and downward), while stepping on. The other tread plate 236b (236a) that is not moved moves in a direction opposite to the stepping direction (slanting front downward) of the stepping plate 236a (236b) on the stepped side 236b (236b).

  On the other hand, when the operator depresses the tread plates 236a and 236b of both the forward and reverse pedals 232 and 233, the depressing operation of the pedals 232 and 233 causes the check links 238a and 238b and the cam plate 258 to be connected. Therefore, the tread plates 236a and 236b of both the plates cannot be moved simultaneously in the stepping direction. In this way, even if the operator steps on both treads 236a and 236b at the same time, both treads 236a and 236b do not move in the treading direction (slanting forward and downward) at the same time, and one of the treads 236a (236b) is not Only when the pedal is depressed, only the pedals 232 and 233 on the side where the pedals 236a and 236b are depressed are operated.

  As shown in FIGS. 19 and 20, the neutral position restoring means 241 is a T-shaped cam having a return spring 256 for returning the footboards 236a, 236b to the initial (neutral) position and a cam groove 257 at the tip. It consists of a plate 258 and a cam roller 265 provided in the cam groove 257 so as to be movable. A base end portion of the cam plate 258 is rotatably connected to one end portion of the transmission frame 266 via the cam shaft 270. The camshaft 270 is disposed on the transmission frame 266. One end of the return spring 256 is locked to the cam shaft 270. The other end side of the return spring 256 is engaged with a roller shaft 267 for fitting the cam roller 265 rotatably. When the cam roller 265 is positioned substantially in the middle of the cam groove 257, the roller shaft 267, the support shaft 269, and the cam shaft 270 are arranged on the same straight line, and the return spring 256 is most contracted. The stepping plates 236a and 236b of the forward pedal 232 and the reverse pedal 233 are configured to be held at initial (neutral) positions, respectively.

  On the other hand, when the operator steps on one of the forward or reverse tread plates 236a, 236b, the cam plate 258 rotates in the forward or reverse direction against the return spring 256 force, and the cam roller 265 moves in the cam groove. 257 is moved in the both end directions from the substantially intermediate portion of 257, and the return spring 256 is expanded in proportion to the moving amount of the cam roller 265. The force for extending the return spring 256 is substantially equal to the pedal depression force in the low speed operation range when the forward or reverse pedals 232 and 233 are depressed to move at low speed.

  As shown in FIGS. 20 and 21, the stepping resistance changing means 242 includes a stepping force increasing spring 260 for increasing the stepping force of the stepping plates 236a and 236b, a pressing force increasing spring 260, and a spring seat 261 and a pulling spring seat 262. A spring cylinder 263 disposed between them, a push-pull rod 264 that connects one end side to the push spring seat 261 and the pull spring seat 262, and a shaft that is rotatable to the other end side of the push-pull rod 264 via a roller shaft 267. The cam roller 265 is supported. The spring cylinder 263 includes a support arm 272. The support arm 272 is rotatably connected to the speed change frame 266 via the arm shaft 273. The spring cylinder 263 is connected to the transmission frame 266. In this case, the operator steps on either one of the step plates 236a, 236b, rotates the cam plate 258, moves the cam roller 265 to the end of the cam groove 257, further steps on the step plates 236a, 236b, When the plate 258 is further continuously rotated in the same direction, the push-pull rod 264 moves in either the push direction or the pull direction, and either the push spring seat 261 or the pull spring seat 262 increases the pedaling force. The spring 260 moves so as to be compressed.

  The force (the pedal depression reaction force) for compressing the pedal force increasing spring 260 is substantially equal to the pedal depression force in the high speed operation range when the forward pedal or the reverse pedal 232, 233 is depressed to move at high speed. Therefore, the stepping resistance force of the forward pedal 232 and the reverse pedal 233 increases abruptly in the middle of their stepping amount. That is, when the stepping plates 236a and 236b are stepped on beyond the stepping amount in the low-speed movement range (stepping amount until the cam roller 265 is moved to the end of the cam groove 257), the stepping reaction force (pedal stepping force) of the pedal 232 (233) ) Is increased stepwise by the stepping resistance changing means 242 so that the operator can easily feel that the acceleration is intended to exceed a predetermined value.

  As shown in FIGS. 19 and 20, between the bracket 240 and the sensor link 271, a linear potentiometer or the like as a depression detection sensor for detecting the depression amount (or depression angle) of the respective depression plates 236a and 236b. A shift potentiometer 220 is provided. The bracket 240 is integrally connected to the transmission frame 266. The sensor link 271 is integrally connected to the cam plate 258. The sensor arm 220 a of the speed change potentiometer 220 is always pressed against the sensor link 271 by the spring force of a spring (not shown) built in the speed change potentiometer 220. The sensor arm 220a rotates in conjunction with the sensor link 271. Both the shift potentiometer 220 and the cam plate 258 are installed on the shift frame 266 so that the relative position between the shift potentiometer 220 and the cam plate 258 can be determined with high accuracy.

  Next, constant speed ratio control (speed ratio adaptive control) will be described. Here, the gear ratio means the ratio of the rotational speed of the output shaft 36 of the hydraulic continuously variable transmission 29 to the engine rotational speed, and the same applies hereinafter.

  An input shaft 27 to which power from the engine 5 is transmitted is disposed in a mission case 17 mounted on the traveling machine body 2, and the input shaft 27, a hydraulic pump unit 500 for shifting, a hydraulic motor unit 501, an output shaft 36, In a work vehicle that includes a hydraulic continuously variable transmission 29 arranged on the same axis and configured to transmit at least traveling driving force from the output shaft 36 via the hydraulic motor unit 501, a preset target The actual (real) gear ratio may not match the gear ratio due to environmental changes and disturbances (mainly loads related to travel). Therefore, the control in which the actual (actual) engine speed and the output shaft 36 are fed back to the electronic governor controller 213 so as to approach or match the target speed ratio is called constant speed ratio control (speed ratio adaptive control). . In other words, while controlling the rotational speed of the engine 5 so as to be maintained at a substantially constant value regardless of the fluctuation of the load, the actual (real) gear ratio is within a predetermined value% of the target gear ratio. The proportional solenoid valve 203 for controlling the gear ratio of the hydraulic continuously variable transmission 29 is controlled. Therefore, the speed ratio pattern of the speed ratio of the rotational speed of the output shaft 36 in the hydraulic continuously variable transmission 29 with respect to the rotational speed of the engine 5 is stored in a RAM (anytime readable / writable memory) as pattern storage means in the travel controller 210.

  This speed ratio pattern is a pattern in which the speed ratio increases in proportion to an increase in the depression amount of the speed change pedal (forward pedal 232 and reverse pedal 233), and the proportional function may be a linear function. It may be a function of a quadratic curve. The pattern storage means stores a plurality of gear ratio patterns in a function table format or a map format (see a gear ratio diagram as shown in FIG. 24). According to the type of farm work and field conditions (soil quality, paddy field, field land, etc.), in the embodiment of FIG. 24, 15 types of speed ratio patterns (speed ratio lines) are prepared and stored in the pattern storage means in advance. I am letting. When the operator selects the scale of the transmission ratio setting device (setting dial) 221, any one pattern can be set (instructed) from among a plurality of types of transmission ratio patterns. In other words, the gear ratio setting dial 221 is for changing (adjusting and setting) the rate of change of the gear ratio corresponding to the amount by which the shift pedal is depressed.

  In the embodiment shown in FIG. 24, the abscissa represents the pedal depression amount (indicated as a percentage of the maximum depression amount, the rightward is for the forward pedal and the leftward is for the reverse pedal), and the vertical axis (see the left vertical axis). Is the speed ratio of the rotational speed of the output shaft 36 in the hydraulic continuously variable transmission 29 with respect to the rotational speed of the engine 5 [(the rotational speed of the output shaft 36 / the rotational speed of the engine 5) = 0-2]. A diagram of the pattern is shown. In FIG. 24, No. 1 to No. 15 are set in order from the lower line, and the forward speed ratio pattern and the reverse speed ratio pattern are set to be the same (symmetric). Further, FIG. 24 shows a diagram (indicated by a broken line) showing a change in the pedal depression force with respect to the pedal depression amount. That is, the right vertical axis represents the pedal effort (indicated as a percentage of the maximum value).

  In the present embodiment, the pedaling force of the forward pedal 232 (or the reverse pedal 233) is suddenly increased by the second spring means 242 having the above-described configuration at the midway portion (for example, a position at about 70% of the total pedaling amount). It is supposed to be. That is, when the pedal depression amount is from 0% to about 70%, the pedal depression force increases linearly and proportionally with a low gradient by resisting the urging force of the first spring means 241. At a position where the pedal depression amount is about 70%, the resistance force by the second spring means 242 is added, so the pedal depression force suddenly changes from about 20% of the maximum value to about 50%, and the pedal depression amount thereafter is about 70%. From 100 to 100%, the pedal effort increases in a substantially linear proportion with a high gradient.

  Each speed ratio pattern (speed ratio line) is also different between a region where the pedal depression amount is less than about 70% and a region where the pedal depression amount is larger than that. In the embodiment of FIG. 24, the No. 1 to No. 11 line (No. 1 to No. 11 in order from the lower line in FIG. 24) has a pedal depression amount less than the position of about 70%. The gradient of the gear ratio line is set to be low in the region, and the gradient of the gear ratio line is set to be high in the region exceeding the position of about 70%. When adopting a gear ratio line, when the operator depresses the pedal beyond the normal pedal depression amount range (within about 70%) with the intention of acceleration, it can be felt by a sudden increase in the pedal depression force. . Also, as will be described later, when the operator wants to increase the speed beyond that set by the gear ratio setting dial in response to environmental changes, disturbances, etc., the operator does not have to bother to change the gear ratio setting dial setting. Thus, the speed can be increased simply by depressing the pedal more than a predetermined amount.

  In the No. 12 line (the twelfth line from the lower line in FIG. 24), the gear ratio line is set substantially in a straight line until the pedal depression amount is 0% to 100%.

  In the No. 13 and No. 14 lines (referred to as No. 13 and No. 14 from the lower line in FIG. 24), the gradient of the transmission ratio line in the region where the pedal depression amount exceeds about 70%. Is set lower than the gradient in an area less than about 70% of the position. The gear ratio line of No. 15 is a straight line that becomes the gear ratio 2 at a position where the pedal depression amount is about 70% (immediately before the pedal contacts the second spring means 242).

  Next, the gear ratio adaptive control mode will be described with reference to the flowchart (FIG. 25). As described above, the proportional control electromagnetic valve 203 is operated in proportion to the amount of depression of the forward pedal 232 (or the reverse pedal 233), and the main transmission hydraulic cylinder 556 is driven by the hydraulic oil from now on, so that the hydraulic pressure that is the main transmission mechanism. The hydraulic oil discharge amount of the hydraulic pump unit 500 of the continuously variable transmission 29 is controlled. In this case, automatic control is performed in which the ultimate target value or the maintenance target of the gear ratio set by the setting dial 211 is performed in accordance with changes in the environment. More specifically, the amount of depression of the shift pedal is self-monitored, and the amount of depression Is automatically controlled so as to follow the target line of the gear ratio set by the gear ratio setting dial 211 in accordance with the change of the gear ratio. Thus, the rotational speed of the main transmission output shaft 36 can be changed and adjusted steplessly. Therefore, in the gear ratio adaptive control, the engine is started (S1), and then the operator selects and determines a desired gear ratio pattern according to the work type or the like with the gear ratio setting dial (S2). A predetermined speed ratio pattern stored in the RAM 210 (a readable / writable memory as needed) is read. Next, when the operator depresses the forward pedal 232 (reverse pedal 233) to move the tractor 1 forward (reverse), the pedal depression amount is read into the travel controller 210 from the forward potentiometer 219 (reverse potentiometer 220) (S3). Based on the read numerical value, the calculation unit of the travel controller 210 calculates a target gear ratio value corresponding to the pedal depression amount on the selected gear ratio pattern (S4).

  On the other hand, the traveling controller 210 reads the engine speed from the engine rotation sensor 206 at every constant time interval (sampling time interval) during traveling, and similarly, at every constant time interval (sampling time interval). The output part rotation sensor 116 detects and reads the main transmission output part rotation speed (S5). From the ratio of the engine speed (denominator) and the main shift output speed (numerator) at the sampling time (current) as the moving average value of a plurality of (4-5) sampling values, the actual (actual) current A speed ratio value is calculated, and it is determined whether or not the current speed ratio value is substantially equal to the target speed ratio value (S6).

  When the current gear ratio value deviates from the target gear ratio value by a predetermined percentage that is larger or smaller (S6: no), the travel controller 210 corrects the applied voltage value of the proportional control solenoid valve 203, whereby the hydraulic pump unit 500 The main shift operation is executed to change or adjust the swash plate angle and control the amount of hydraulic oil discharged to the hydraulic motor unit 501 to increase or decrease the rotational speed of the main shift output shaft 36 (S7). If the current gear ratio value is substantially equal to the target gear ratio value (if the current gear ratio value is within ± predetermined% with respect to the target gear ratio value) (S6: yes), the rotational speed of the main gear output shaft 36 is maintained. (S8).

  In this way, feedback control is performed so that the current gear ratio value approaches (or matches) the target gear ratio value. In these cases, for example, when the output torque fluctuation of the engine is insufficient due to disturbance such as an increase in driving load such as when a tractor enters a soil area containing a lot of water from a dry soil area Of course, when the engine speed deviates from a predetermined value due to environmental changes, the electronic governor controller 213 is operated to control the engine speed to be maintained at a predetermined value.

  If such gear ratio adaptive control is employed, the operator only needs to operate the forward pedal 232 or the reverse pedal 233, which is a gearshift pedal, after setting the gear ratio pattern of the gear ratio with the gear ratio setting dial 221. When the actual gear ratio value deviates from the target gear ratio value due to environmental changes or fluctuations in the work vehicle's running load, it can be automatically controlled so that it automatically approaches the target gear ratio value. The traveling operation can be made extremely simple by approximating the traveling operation in an automobile with a continuously variable transmission mechanism, and there is an effect that fatigue for a long time can be reduced.

  Further, since the automatic control is executed by correcting the operation of the proportional control solenoid valve 203, there is an effect that the control can be performed finely and quickly. Further, since the electronic governor 214 for controlling the rotational speed of the engine 5 according to the load is provided, it is not necessary for the operator to manually adjust the engine throttle, and the rotational speed of the engine 5 and the hydraulic stepless speed change. The engine speed and the pump swash plate 509 angle can be controlled so that the output of the machine 29 can be kept highly efficient, and the traveling operation of the work vehicle can be changed to the traveling operation in an automobile with a continuously variable transmission mechanism. It can be approximated to make it extremely simple, and there is an effect that fatigue for a long time can be reduced.

  Instead of detecting the rotational speed of the main transmission unit, it is also possible to detect the vehicle speed (the rotational speed of the wheel) and calculate the gear ratio as an equivalent meaning.

  Next, load control will be described. The load control is a control for the purpose of high-efficiency and simple driving operation in rotary tillage work, grass-related harvesting packing work, loader work, etc. that dislikes fluctuations in the engine speed of the tractor. 210.

  An example of this load control will be described with reference to a flowchart (FIG. 26). After engine start (S130), a gear ratio pattern is set with a gear ratio setting dial (S131). Next, the depression amount of the forward (reverse) pedal is read (S132). Based on the read numerical value, the calculation unit of the travel controller 210 calculates a target gear ratio value (same as described above) corresponding to the pedal depression amount on the selected gear ratio pattern (S133). On the other hand, the travel controller 210 reads the engine speed and the speed of the main transmission output shaft 36 by sampling similar to S5 described above (S134). Next, from the ratio of the engine speed (denominator) and the main transmission output speed (numerator) at the sampling time (current) as a moving average value of a plurality (4 to 5) of sampling values, the actual (actual) ) Is calculated (S135).

  Next, the engine load factor (detected value) is read by the electronic governor controller 213 (S136). The engine load factor is the ratio of the current engine output to the engine rated output. Next, the user turns on the load control mode switch 222a (S137). The high-speed / low-speed sub-shift switch 222 is turned on to the low speed side (S138: yes) and the engine speed is equal to or higher than a predetermined value (1200 rpm in the embodiment) (S139: yes). The following load control is executed. When the auxiliary transmission changeover switch 222 is on the high speed side or the engine speed is less than a predetermined value, load control is not executed (S140).

  When the load is an appropriate load, that is, the detected value of the engine load read in S136 is equal to or greater than a predetermined value B% (80% in the embodiment), and the engine speed reduction rate is a predetermined value A% (the embodiment) Is less than 10%) (S141: yes), the target gear ratio value corresponding to the pedal depression amount on the selected gear ratio pattern is maintained (S142).

  When the load is a large load, that is, the engine speed reduction rate is not less than a predetermined value A% (10% in the embodiment) and the detected value of the engine load is not less than a predetermined value B% (80% in the embodiment). Sometimes (S143: yes), the target speed ratio value is controlled to decrease until the engine speed reduction rate becomes less than a predetermined value C% (5% in the embodiment) (S144). As a result, when an excessive load is applied to the engine during farm work, the target speed ratio value is automatically reduced without changing the speed ratio pattern or adjusting the pedal depression amount such as forward movement. The load can be returned to the appropriate value range.

  On the contrary, when the load is small, that is, when the detected value of the engine load is less than a predetermined value B% (80% in the embodiment) (S145: yes), the detected value of the engine load becomes 100%. Until the target gear ratio value is increased (S144). As a result, high-efficiency work can be performed by increasing the traveling speed without overloading the engine.

  In the load control mode, when the load is large, the speed is not increased even if the shift pedal is depressed to the speed increasing side. Further, when the shift pedal is depressed to the deceleration side, control is performed so that the vehicle decelerates when the target gear ratio value becomes smaller than the current gear ratio value.

It is a side view of the tractor for farm work. It is a diagonally rear perspective view of a tractor. It is side surface explanatory drawing of a tractor. It is a perspective view of a tractor airframe. It is a skeleton diagram of power transmission. It is explanatory drawing of the travel transmission part of a mission case. It is explanatory drawing of the PTO transmission part of a mission case. It is explanatory drawing of the transmission without permission of a mission case. It is a hydraulic circuit diagram of the continuously variable transmission of the transmission case. It is bottom explanatory drawing which shows the inside of a mission case. It is a bottom perspective view showing an oil filter and a solenoid valve. It is the bottom perspective view which removed the oil filter. It is a side view which shows the speed change operation part of a continuously variable transmission. It is a perspective view which shows a continuously variable transmission and a rear side wall member. It is a hydraulic circuit diagram in the whole tractor of the present invention. It is a top view which shows the cabin part of a tractor. It is a top view which shows a forward (reverse) pedal part. It is a side view which shows a forward (reverse) pedal part. It is a side view which shows a 1st spring means part. It is a side view which shows a 2nd spring means part. It is a top view which shows a 1st spring means part and a 2nd spring means part. It is a side view which shows a brake pedal part. It is a functional block diagram of the control means of this invention. It is a gear ratio diagram. It is a flowchart of gear ratio adaptive control. It is a flowchart of load control.

Explanation of symbols

5 Engine 17 Mission case 27 Input shaft 29 Continuously variable transmission 36 Output shaft 116 Main transmission output part rotation sensor 232 Forward pedal 233 Reverse pedal 206 Engine rotation sensor 210 Travel controller 213 Electronic governor controller 219 Forward potentiometer 220 Reverse potentiometer 222a Load control mode Changeover switch 500 Hydraulic pump unit 501 Hydraulic motor unit 556 Main transmission hydraulic cylinder (transmission actuator)

Claims (4)

An input shaft for transmitting power from the engine is disposed in a transmission case mounted on the traveling machine body, the input shaft, the hydraulic pump, the hydraulic motor, and the output shaft are disposed on the same axis, and a sub-transmission mechanism In a work vehicle configured to transmit at least traveling driving force from the output shaft via the hydraulic motor,
A speed ratio setting device for setting a speed ratio pattern of a speed ratio of an output shaft speed in the hydraulic continuously variable transmission with respect to the engine speed, an engine speed sensor, the output shaft speed sensor, and an electronic A governor controller and load control means;
The load control means is executed when a load control switch is ON, the auxiliary transmission mechanism is turned on at a low speed, and the engine speed is equal to or higher than a predetermined value. . Pattern storage means for storing a plurality of gear ratio patterns, a shift sensor for detecting the depression amount of the shift pedal, and an actuator for adjusting the swash plate angle of the hydraulic pump based on the detected value of the shift sensor;
The load control means is set by the gear ratio setting device when the rate of decrease in the engine speed is less than a predetermined value A% and the detected value of the engine load in the electronic governor controller is greater than or equal to a predetermined value B%. 2. The control apparatus for a work vehicle according to claim 1, wherein control is performed so as to maintain a target speed ratio on the speed ratio pattern.   The load control means sets the engine speed reduction rate to a predetermined value C% when the engine speed reduction rate is not less than the predetermined value A% and the engine load detection value is not less than the predetermined value B%. The work vehicle control device according to claim 2, wherein the control is performed so that the target gear ratio is decreased until the target gear ratio is less than the value.   The load control means controls the target gear ratio to be increased until the detected value reaches 100% when the detected value of the engine load is less than the predetermined value B%. 4. The work vehicle control device according to 3.

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