Steering-wheel revolution number correction system of all-wheel-drive vehicle

Turning radius is calculated from steering angles of steering-wheels and articulation angles of respective frames, and the steering-wheels are automatically rotated faster than rear-wheels based on the turning radius, the revolution number of the rear-wheels, engine speed of an engine and a speed range of hydraulic motors, so that conventional trouble for manipulating operation lever etc. while turning a motor grader and skill-requiring work for adjusting the revolution number of the steering-wheels in accordance with the revolution number of the rear-wheels can be eliminated, thereby easily and accurately controlling the revolution number of the steering-wheels and securely preventing tight-corner braking phenomenon.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a steering-wheel revolution number correction system of all-wheel-drive vehicle. More specifically, it relates to a steering-wheel revolution number correction system of a construction equipment provided with an all-wheel-drive, such as a motor grader and a wheel loader.

2. Description of Related Art

Recently, all-wheel-drive for driving all of front and rear wheels is often provided to a motor grader, etc.

The all-wheel drive drives the rear-wheel by an output of an engine transmitted through a transmission and the output of the engine is transmitted to a hydraulic pump, where right and left front-wheels are driven by a pair of hydraulic motors rotated by delivery oil from the hydraulic pump.

Incidentally, when the vehicle turns, because the turning radius of the front-wheel becomes greater than the turning radius of the rear-wheel on account of inner race difference, the front-wheel has to be rotated faster than the rear-wheel. In a rear-wheel-drive vehicle solely driving the rear wheel thereof, the front-wheel freely rotates in accordance with turning radius. However, in a vehicle having an all-wheel-drive where the front-wheel rotates in synchronization with the rotation of the rear-wheel, the rotation difference of the front and rear wheels caused by the inner race difference during turning movement cannot be completely absorbed, so that the vehicle turns with the front-wheel being braked. Accordingly, unless the front-wheel is rotated faster than the rear-wheel, a so-called “tight-corner” braking phenomenon occurs during turning movement and the vehicle does not turn smoothly.

In order to rotate the hydraulic-motor-driven front-wheel faster than the engine-driven rear-wheel, U.S. Pat. No. 4,186,816 discloses switching a control mode of a hydraulic pump to rotate the front-wheel faster than the rear-wheel by a predetermined ratio (several percent).

Further, U.S. Pat. No. 5,147,010 discloses monitoring and controlling the hydraulic pressure applied to the hydraulic motor to change the revolution number of the hydraulic motor in accordance with slippage of the rear-wheel, thereby rotating the front-wheel faster than the rear-wheel.

However, according to the disclosure of the U.S. Pat. No. 4,186,816, because the front-wheel always rotates faster than the rear-wheel by a predetermined ratio, a problem occurs when the vehicle turns at the same speed (the same revolution number of the rear-wheel) and with different turning radius. Specifically, though the front-wheel has to be rotated faster as the turning radius becomes smaller, since the revolution number of the front-wheel is determined irrespective of the turning radius, the front-wheel does not rotate at a revolution number corresponding to the turning radius, so that the tight-corner braking phenomenon cannot be securely prevented.

Further, the rotation of the front-wheel faster than the rear-wheel by the several percent cannot rotate the front-wheel sufficiently fast when the vehicle turns with a small turning radius.

Because the slippage of the rear-wheel is required in the technique shown in U.S. Pat. No. 5,147,010, the front-wheel does not rotate fast when the vehicle is turned and the rear-wheel does not always slip, so that the tight-corner braking phenomenon cannot be securely prevented.

Further, because only the hydraulic pressure to the hydraulic motor is monitored in this technique, when the ground pressure of the front-wheel is decreased and is likely to be slipped when, for instance, the work is conducted while keeping the blade in contact with the ground, the hydraulic motor may be rotated at a high speed without sufficient hydraulic pressure and the front-wheel may keep on slipping. The energy is wasted in the above condition.

On the other hand, an auxiliary lever for manually controlling the rotation of the hydraulic motor may be provided and operated while turning the vehicle for rotating the front-wheel faster than the rear-wheel. In this arrangement, the front-wheel can be rotated faster than the rear-wheel by a desired ratio during turning movement, so that the vehicle can be rotated smoothly.

However, it is so troublesome to operate the auxiliary lever throughout turning the vehicle in addition to steering operation. Further, it takes considerable skill and is not easy to properly adjust the revolution number of the front-wheel in accordance with the speed of the vehicle and the turning radius.

On the other hand, it is possible to calculate the turning radius by detecting the steering angle of the front-wheel and to rotate the front-wheel faster than the rear-wheel based on the turning radius.

However, in an articulating motor grader having the front-wheel and the rear-wheel attached to independent frames, the respective frames being angle-adjustably connected, because the turning radius on the front-wheel side greatly changes relative to the rear-wheel side not only by the steering operation but by articulating the front frame on the front-wheel side relative to the rear frame on the rear-wheel side, accurate turning radius cannot be obtained only by detecting the steering angle.

SUMMARY OF THE INVENTION

In the present invention, following arrangement is used for a steering-wheel revolution number correction system of an all-wheel-drive vehicle for easily and accurately controlling the revolution number of the front-wheel while turning the vehicle and securely preventing the tight-corner braking phenomenon even when the vehicle has an articulating type frame.

In the present invention, the following arrangement is used for an all-wheel-drive vehicle having an articulating frame.

A steering-wheel revolution number correction system of all-wheel-drive vehicle according to an aspect of the present invention is used for an all-wheel-drive vehicle having an all-wheel-drive for driving a rear-wheel by an output transmitted through a transmission, for transmitting the output of the engine to a hydraulic motor and for driving a front-wheel by a hydraulic motor rotated by a delivery oil from the hydraulic pump, a front frame provided with the front-wheel and a rear frame provided with the rear-wheel, the front frame and the rear frame being angle-adjustably connected, the system including: a front-wheel steering angle sensor for detecting a steering angle of the front-wheel; an articulation angle sensor for detecting a connection angle of the front frame and the rear frame; a turning radius operation for calculating a turning radius of the vehicle based on the steering angle of the front-wheel and the connection angle of the frames; a rear-wheel revolution number sensor for detecting a revolution number of the rear-wheel; an engine speed sensor for detecting engine speed of the engine; and a steering-wheel revolution number controller for rotating the front-wheel faster than the rear-wheel based on the turning radius and the revolution number of the rear-wheel.

According to the above aspect of the present invention, after detecting the steering angle of the front-wheel by the front-wheel steering angle sensor and detecting the connection angle of the respective frames by the articulation angle sensor, the turning radius on the front-wheel side is calculated by the turning radius operation based on the steering angle and the connection angle. Then, the steering-wheel revolution number controller determines how many times faster the front-wheel has to be rotated than the rear-wheel based on the turning radius, and calculates the necessary revolution number of the front-wheel, i.e., the required revolution number of the hydraulic motor, from the ratio and the rear-wheel revolution number detected by the rear-wheel revolution number sensor. Further, the steering-wheel revolution number controller determines the optimum delivery oil flow rate from the hydraulic pump in accordance with the engine speed detected by the engine speed sensor and feeds the delivery oil to the hydraulic motor to rotate the hydraulic motor with the necessary revolution number, thereby automatically rotating the front-wheel faster than the rear-wheel. Accordingly, the trouble for manipulating the auxiliary lever during turning movement and uncertainty caused in adjusting the revolution number can be eliminated and accurate turning radius can be obtained by additionally considering the connection angle of the respective frames as well as the steering angle of the front-wheel.

On the other hand, according to another aspect of the present invention capable of being applied to an all-wheel-drive vehicle having no articulating frame, following arrangement is adopted.

A steering-wheel revolution number correction system of all-wheel-drive vehicle according to another aspect of the present invention is used for an all-wheel-drive vehicle having an all-wheel-drive for driving a rear-wheel by an output transmitted through a transmission, for transmitting the output of the engine to a hydraulic motor and for driving a front-wheel by a hydraulic motor rotated by a delivery oil from the hydraulic pump, the system including: a steering-wheel revolution number sensor for detecting respective revolution numbers of the right and left front-wheels; a turning radius operation for calculating a turning radius on the front-wheel side based on the revolution numbers of the right and the left front-wheels; a rear-wheel revolution number sensor for detecting a revolution number of the rear-wheel; an engine speed sensor for detecting engine speed of the engine; and a steering-wheel revolution number controller for rotating the front-wheel faster than the rear-wheel based on the turning radius and the revolution number of the rear-wheel.

According to the above aspect of the present invention, after the revolution numbers of the right and left front-wheels are detected by the steering-wheel revolution number sensor, the turning number on the front-wheel side is calculated based on the revolution numbers. Then, the steering-wheel revolution number controller initially determines how many times faster the front-wheel has to be rotated than the rear-wheel based on the turning radius and, subsequently, the necessary revolution number, i.e. required revolution number of the hydraulic motor is calculated by the ratio and the rear-wheel revolution number detected by the rear-wheel revolution number sensor. Further, the optimum delivery oil flow rate from the hydraulic pump in accordance with the engine speed detected by the engine speed sensor is determined by the steering-wheel revolution number controller and feeds the delivery oil to the hydraulic motor to rotate the hydraulic motor with the necessary revolution number, thereby automatically rotating the front-wheel faster than the rear-wheel. Accordingly, the trouble for manipulating the auxiliary lever during turning movement and uncertainty caused in adjusting the revolution number can be eliminated to achieve the above object.

In the above, the system may further include: a front frame provided with the front-wheel; a rear frame provided with the rear-wheel, the front frame and the rear frame being angle-adjustably connected; and an articulation angle sensor for detecting a connection angle of the front frame and the rear frame, where the turning radius operation determines the turning radius based on the revolution number of the right and left front-wheels and the connection angle of the respective frames.

Such an arrangement is suitably used for an articulating type motor grader including a front frame and a rear frame, so that the turning radius can be more accurately obtained.

In the above, a torque converter may preferably be provided between the engine and the transmission, and the rear-wheel revolution number sensor may preferably include a torque converter output side revolution number sensor for detecting a revolution number on an output side of the torque converter and a speed range sensor for detecting a speed range of the transmission connected to the output side of the torque converter.

Some of the all-wheel-drive vehicles transmit the engine output to the transmission via a torque converter. In an such arrangement, the engine speed does not directly become the transmission input side revolution number on account of torque transfer loss in the torque converter.

Accordingly, in the present invention, the revolution number of the output side of the torque converter, i.e., the revolution number inputted to the transmission is detected by the torque converter output side revolution number sensor. Accordingly, the revolution number of the rear-wheel of the all-wheel-drive vehicle provided with the torque converter can be accurately detected by detecting the revolution number inputted to the transmission and the speed range of the transmission.

In the above, the steering-wheel revolution number controller may preferably control the flow rate of the delivery oil from the hydraulic pump to be supplied to the hydraulic motor based on the turning radius and the revolution number of the rear-wheel.

According to the above arrangement, the flow rate of the delivery oil is controlled by arranging the hydraulic pump as variable capacity type, the structure of the hydraulic motor side can be simplified and size thereof can be reduced, so that the hydraulic motor can be suitably accommodated in a narrow disposition space on the front-wheel side.

In the above arrangement, a plurality of speed ranges may preferably be set in the hydraulic motor, and the steering-wheel revolution number controller may preferably control the flow rate of the delivery oil from the hydraulic pump based on the speed range of the hydraulic motor as well as the turning radius and the revolution number of the rear-wheel.

According to the above arrangement, the respective motors are arranged as variable capacity type, the revolution number of the right and left front-wheels can be varied by providing the hydraulic motors, respectively, on the right and left front-wheels. In other words, it is advantageous to rotate the outer wheel side of the right and left front-wheels faster than the inner wheel side, so that the turning movement can be conducted further smoothly by controlling the respective front-wheels in such a manner.

Further, when the engine output is transmitted to the transmission through the torque converter, if the engine speed is rapidly dropped from the high-speed side, though only for a short time, the rear-speed continues to rotate at a high-speed on account of the characteristics of the torque converter. At this time, the rotation difference between the engine speed (hydraulic pump) and the rear-wheel is increased and, even when the flow rate of the delivery oil from the hydraulic pump is set at a maximum, the front-wheel cannot be rotated faster than the rear-wheel, thereby causing tight-corner braking phenomenon during the time period.

However, according to the present invention, because the revolution number of the hydraulic motor for driving the front-wheel is directly controlled, low-torque and high-speed drive can be achieved with small flow rate of delivery oil even when the flow rate of the delivery oil of the hydraulic pump is small, so that the revolution number of the hydraulic motor can be maintained at a high-speed rotation and the tight-corner braking phenomenon can be restrained.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENT(S)

FIGS. 1to5show a motor grader1as an all-wheel-drive vehicle according to first embodiment of the present invention applied with the steering-wheel revolution number correction system of the present invention.

InFIGS. 1 and 2, the motor grader1is a vehicle having total six wheels, i.e., a pair of front wheels of a left front-wheel2and right front-wheel3, and rear-wheels which are provided in pairs on both sides, the rear-wheels including left rear-front wheel and left rear-rear wheel (not shown) and a right rear-front wheel4and a right rear-rear wheel5, where a blade50provided between the front-wheels2and3and the rear-wheels4and5levels the ground, removes snow, cuts with low load and mixes material, etc.

The front-wheels2and3as well as the blade50are mounted on a front frame51and the rear-wheels4and5are mounted on the rear frame52. As shown inFIG. 2, the front frame51is rotatably connected to the rear frame52by a vertical center pin53at approximately beneath a cockpit. The front frame51is turned by stretching and contracting an articulate cylinder54connected between the frames51and52by operating a lever from the cockpit. The turning radius while turning the motor grader1can be made smaller by turning (articulating) the front frame51relative to the rear frame52.

As represented by the right front-wheel3shown inFIG. 3, the front-wheels2and3are connected to the front frame51(FIG. 1) through a knuckle70, a knuckle support71and a front axle72.

The knuckle70is connected to a knuckle support71through a king pin73and is turned around the king pin73. The turning movement is conducted with a steering cylinder74having both ends connected between the knuckle70and the front axle72by steering operation in the cockpit. The knuckle arms70of the front-wheels2and3are connected by a tie rod75.

The front axle72is composed of a first cross member76supporting lower front portion of the knuckle support71and a second cross member77supporting lower rear portion of the knuckle support71.

A (rod of) leaning cylinder78for slanting (leaning) the right front-wheel3in right and left direction is connected to the upper front of the knuckle support71and a leaning rod79for transmitting the leaning movement to simultaneously lean the left front-wheel2(FIG. 1) is connected to the upper rear of the knuckle support. The leaning movement is effective for further lessening the turning radius in turning the motor grader1.

Since the knuckles70on both right and left sides are connected by the tie rod75as described above, the knuckle on left side (not shown) can be turned by turning the right knuckle70by the single steering cylinder74. However, the steering cylinder74may preferably be provided on the left side and the respective knuckles may be turned by respective steering cylinders74. This is because, when the turning movement is conducted solely by the single steering cylinder74on the right front-wheel3, steering operation can become different between left turn for advancing the rod and right turn for retracting the rod on account of speed difference between advancement and retraction of the rod.

As shown inFIG. 4, the front-wheels2and3are driven by a hydraulic system7connected to an output of the engine6. The rear-wheels4and5are driven through a torque converter8, a transmission9, a final drive10and a tandem device11connected to the other output of the engine6. In other words, the motor grader1is an all-wheel-drive vehicle where the front and rear wheels2to5are driven by the power-generating and power-transmitting devices6to11, the devices6to11constituting an all-wheel drive12.

Almost all of the components of the all-wheel drive12(the engine6, a part of the hydraulic system7, the torque converter8, the transmission9, and the final drive10) are supported by the rear frame52.

The hydraulic system7of the all-wheel drive12includes a hydraulic pump13directly driven by the output of the engine6, and a left hydraulic motor14and a right hydraulic motor15rotated by the hydraulic oil discharged from the hydraulic pump13to drive the front-wheels2and3.

The hydraulic pump13of the present embodiment is a swash plate type axial plunger pump. In the hydraulic pump13, the discharge amount of the hydraulic oil is controlled by changing the angle of the swash plate type (variable capacity type) by a control signal CS outputted from a controller100functioning as a computer. The discharge direction of the hydraulic oil of the hydraulic pump13is switchable between channels A and B by a direction switching signal DS outputted from the controller100, so that the motor grader1is advanced and retracted by changing the discharge direction to change the rotary direction of the respective hydraulic motors14and15.

However, the hydraulic pump to be used is not restricted to swash plate type but may be an angled axis type.

The hydraulic motors14and15are radial piston type in the present embodiment, which is integrated in the knuckle60shown in FIG.3. As shown in FIGS.5(A) and5(B) enlarging cross section, the hydraulic motors14and15include a rotary valve16at the center thereof, a fixed core17with the rotary valve16being inserted, and a rotatable cam ring18disposed on the outer circumference of the fixed core17, the cam ring18being connected to the front-wheels2and3.

A plurality of cylinders19are radially formed on the fixed core17, respective cylinders19advancebly and retractably accommodating a piston21having a roller20at a distal end thereof.

As shown in FIG.5(A), the respective pistons21sequentially advance along circumferential direction by the hydraulic oil fed from a supply port22of the rotating rotary valve16into the cylinder19and retracts in sequence of advancement by the discharge of the hydraulic oil through a delivery port23and a biasing force of biasing member (not shown). During the time, the roller20of the advancing piston21contacts a cam surface24inside the cam ring18to rotate the cam ring18.

On the other hand, FIG.5(B) shows a condition in which all of the pistons21are restracted. Such a condition is achieved by shutting the supply of the hydraulic oil toward the hydraulic motors14and15. The supply of the hydraulic oil is shut by actuating a free wheel valve25by a drive type switching signal KS from the controller100. In this condition, the all-wheel driving of the motor grader1is released to be a rear-wheel driving.

Incidentally, for instance, two-stage speed ranges are set in the hydraulic motors14and15(stationary two stages, and not variable capacity type).

In the first stage, as shown in FIG.5(A), the hydraulic oil is supplied from the supply port22of the rotary valve16to all of the pistons21.

In the second stage, the hydraulic oil is supplied only to four pistons21among the ten pistons21disposed in X-shape (two pairs disposed on intersecting line) and not adjacent in circumferential direction, and the hydraulic oil is not supplied to the remaining six pistons21. In this case, the other six pistons21are completely retracted and are not in contact with the cam ring18as shown in FIG.5(B). In this speed range, supposing the delivery oil flow rate is the same as in the first stage, since the number of the piston21to which the hydraulic oil is supplied decreases to four, the applied hydraulic pressure becomes high and the four pistons21are advanced and retracted at a high speed and, consequently, rotate the cam ring18at a higher speed than in the first stage. On the other hand, if the delivery oil flow rate is less than the first stage, approximately the same revolution number as in the first stage can be obtained.

The first and the second speed ranges are automatically switched in accordance with speed ranges of the transmission9by a solenoid valve26actuated by a speed range switching signal MS from the controller100.

Back to the hydraulic system7shown inFIG. 7, hydraulic sensors27and28are respectively provided on the channel A and B from the hydraulic pump13to judge whether the hydraulic pressure in the hydraulic circuit is proper or not based on a hydraulic pressure detecting signal OS outputted from the hydraulic sensors27and28to the controller100.

An intercommunicating flow path C for intercommunicating the channels A and B is provided on the side of the hydraulic pump13relative to the free wheel valve25, and an inching valve29is provided in the intercommunicating flow path C.

The inching valve29is actuated by a power transfer switching signal BS from the controller100. When the intercommunicating flow path C is blocked by the inching valve29(condition shown in FIG.4), the hydraulic oil discharged from the hydraulic pump13is supplied to the hydraulic motors14and15. When the intercommunicating flow path C is in communication, the hydraulic oil circulates in the channels A to C and is not supplied to the hydraulic motors14and15, thus not transmitting the power. Such inching valve29is used for suspending the drive of the front-wheels2and3by the hydraulic motors14and15for a temporary short time.

The hydraulic motors14and15are connected to the channels A and B in parallel. More specifically, a first left branched flow path D from one side of the left hydraulic motor14and a first right branched flow path E from one side of the right hydraulic motor15are connected to the channel A, and a second left branched flow path F from the other side of the left hydraulic motor14and a second right branched flow path G from the other side of the right hydraulic motor15are connected to the channel B.

The channel A, the first left branched flow path D and the first right branched flow path E are connected by a flow switching valve30.

The flow switching valve30has throttles31and32on the side of the respective branched flow paths D and E. In the flow switching valve30, the hydraulic oil is branched to the respective hydraulic motors14and15with the same flow rate when the hydraulic oil is discharged from the hydraulic pump13to the channel A, for instance, during advancement operation. On the contrary, when the hydraulic oil is discharged toward the channel B, for instance, during retraction movement, the hydraulic oil flowing through the respective branched flow paths D and E is combined at the same flow rate to, consequently, equalize the flow rate branched from the channel B to the second left branched flow path F and the second right branched flow path G.

The first left branched flow path D and the first right branched flow path E are intercommunicated by a bypass flow path H. A throttle33is provided to the bypass flow path H. On account of the bypass flow path H and the throttle33, slightly more hydraulic oil is supplied to the hydraulic motor14(15) to be located outer side during turning movement than the other hydraulic motor15(14) located inner side. Accordingly, one of the front-wheel2(3) as an outer wheel is rotated slightly faster than the other front-wheel3(2) as an inner wheel, to conduct further smooth turning movement.

InFIG. 4, a drive type switching switch34, a front-wheel rotation control lever35, a speed range sensor36as a speed range detecting means, a limit switch37and an angle sensor38as an articulate angle detecting means are connected to the controller100. The drive type switching switch34, the front-wheel rotation control lever35, the speed range sensor36and the limit switch37are respectively provided in the cockpit and the angle sensor38is situated outside the cockpit as shown in FIG.2.

The drive type switching switch34switches the drive type of the motor grader1between all-wheel drive and the rear-wheel drive. When the switch is turned on and off, a switching signal SS is outputted to the controller100and the drive type switching signal KS is outputted to the above-described free wheel valve25by the controller100having received the signal SS to switch the drive type.

The front-wheel rotation control lever35is for manually adjusting a swash plate angle of the hydraulic pump13, which outputs a swash plate angle adjusting signal FS each time the retractable control lever35is operated, and the controller having received the signal FS outputs a control signal CS to the hydraulic pump13to change the angle of the swash plate. Such a front-wheel rotation control lever35is used for reducing torque of the front-wheels2and3when the ground pressure of the front-wheels becomes low during the operation using the blade50, which is effective for preventing slippage of the front-wheels2and3.

The front-wheel rotation control lever35is also used for adjusting target revolution number of the front-wheels2and3in the steering-wheel revolution number correction system, which will be described later.

The speed range sensor36detects a speed range position of a transmission shift lever39, i.e., advancement first to sixth, retraction first to sixth and neutral speed ranges of the lever39, and outputs a speed range detecting signal TS to the controller100. When the controller100judges that the front-wheels2and3has to be driven with low torque and high-speed, the controller100outputs a speed range switching signal MS to the solenoid valve26to switch the speed range of the hydraulic motors14and15to the second step on the higher speed side. When the speed range location is judged neutral by the speed range detecting signal TS, the controller100outputs a power transfer switching signal BS to the inching valve29to stop power transfer toward the hydraulic motors14and15to release drive of the front-wheels2and3.

The speed range detecting signal TS from the speed range sensor36is also used in the steering-wheel revolution number correction system, which will be described later.

The limit switch37is turned on by stepping on the inching pedal40and, while the limit switch37is stepped on, outputs a control device step-on detecting signal PS to the controller100. Because the power is not transmitted to the rear-wheels4and5by a speed switching clutch mechanism and a direction switching clutch mechanism for advancement and retraction (not shown) in the transmission9during the above condition, the controller100outputs the power transfer switching signal BS to the inching valve29to stop power transfer to the hydraulic motors14and15to release drive of the front-wheels2and3.

As shown inFIG. 2, the angle sensor38detects an articulate angle (connection angle) β of the front frame51relative to the rear frame52and outputs an angle detection signal AS to the controller100. The controller100having received the angle detection signal AS outputs a display signal to an indicator (not shown) in the cockpit to display turning condition of the front frame5on the indicator. The angle detection signal AS includes information on articulation direction such as leftward articulation and rightward articulation as well as the articulation angle β.

The angle detecting signal AS from the angle sensor38is also used in the steering-wheel revolution number correction system, which will be described below.

The steering-wheel revolution number correction system60according to the present embodiment will be described below.

InFIG. 4, the steering-wheel revolution number correction system60has a left steering angle sensor61as a front-wheel steering angle detecting means for detecting a steering angle (turning angle) of the left front-wheel2, a right steering angle sensor62as another front-wheel steering angle detecting means for detecting a steering angle (turning angle) of the right front-wheel3, the angle sensor38as an articulation angle detecting means, a turning radius operation63for determining turning radius of the front-wheels2and3based on the steering angle of the front-wheels2and3and the articulation angle β of the respective frames51and52, a rear-wheel revolution number sensor64for detecting revolution number of the rear-wheels4and5, an engine speed sensor65as an engine speed detecting means for detecting engine speed of the engine, and a steering-wheel revolution number controller66for controlling the revolution number of the front wheels2and3based on the turning radius, the revolution number of the rear-wheels4and5and the engine speed of the engine6.

The respective steering angle sensors61and62output a front-wheel steering angle detecting signal VSLand VSRcorresponding to a steering angles αLand αRof the front-wheels2and3relative to the king pin73(FIG. 3) (see FIGS.3and6: αLshows the steering angle of the left front-wheel2).

At this time, the steering angles αLand αRare detected for both of the front-wheels2and3because the steering angles αLand αRdiffer between the left front-wheel2and the right front-wheel3during turning movement. This is on account of the trapezoid link structure formed by the knuckle70, the knuckle support71, the front axle72and the tie rod75.

The rear-wheel revolution number sensor64is composed of a converter output-side revolution sensor67as a converter output revolution number detecting means for detecting a revolution number of the output side of the torque converter8and the speed range detecting sensor36for detecting speed range position of the transmission shift lever39.

The converter output-side revolution sensor67outputs a converter revolution number detecting signal QS to the controller100.

The engine speed sensor65outputs an engine speed detecting signal ES to the controller100.

The turning radius operation63and the steering-wheel revolution number controller66are software, which is stored in a memory101constituting the controller100such as ROM and RAM and is called from the memory101when the system60is actuated and is implemented by a CPU102as shown in FIG.4.

The memory101of the controller100stores lookup table (not shown) tabling the graph shown inFIGS. 6to9as well as the above software.

FIG. 6is a graph showing a relationship between a means steering angle αAVcalculated by the respective steering angles αLand αRof the left front-wheel2and the right front wheel3and a turning radius RSTon the side of the front wheels2and3.

FIG. 7is a graph showing a relationship between the articulation angle β and the turning radius RARon the side of the front-wheels2and3.

FIG. 8is a graph showing a relationship between the turning radius R of the front-wheels2and3calculated based on the respective turning radiuses RSTand RARand a ratio n1of the steering-wheel revolution number and a rear-wheel revolution number, from which how many times faster the front-wheels2and3have to be rotated relative to the rear-wheels4and5when turning at a predetermined turning radius R can be read.

FIG. 9is a graph showing an optimum swash plate angle θ of the hydraulic pump13corresponding to a ratio n2of the necessary revolution number of the front wheel2and3(necessary revolution number of the hydraulic motors14and15) and the engine speed of the engine6, from which the optimum swash plate angle θ can be read for each speed range of the hydraulic motors14and15.

Next, the revolution number control of the front wheels2and3by the steering-wheel revolution number correction system60will be specifically described below.

Initially, when the motor grader1is running, the turning radius operation63monitors the front-wheel steering angle detecting signal VSLand VSRoutputted from the respective steering angle sensors61and62and the angle detecting signal AS outputted from the angle sensor38.

When the motor grader1starts turning, the turning radius operation63calculates the mean value of the steering angles αLand αRto calculate the average turning angle αAVbased on the information of the steering angle detecting signals VSLand VSR, determines the steering direction and obtains a predetermined turning radius RSTwith reference to the lookup table in the memory10based on the graph of FIG.6. At this time, the turning direction such as leftward turning and rightward turning, is also determined based on the difference of the steering angles αLand αR.

The turning radius operation63detects the articulation angle (connection angle) β based on the information of the angle detecting signal AS, determines the articulation angle (connection angle) β and obtains the predetermined turning radius RARwith reference to the lookup table in the memory101based on the graph of FIG.7.

Subsequently, the turning radius operation63checks whether the turning direction of the front wheels2and3and the articulation direction of the respective frames51and52are the same. When the directions are equal, the turning radius R is calculated by the following function of (1) and, when the directions are different, the turning radius R is calculated by the following function of (2).
R=f1(RST, RAR)  (1)
R=f2(RST, RAR)  (2)

Next, when the turning radius R is calculated as twenty meters by the turning radius operation63, the steering-wheel revolution number controller66refers to the lookup table based on the graph ofFIG. 8to obtain a revolution number ratio n1when the turning radius R=twenty meters. According to the graph ofFIG. 8, n1=1.2, so that the controller66judges that the front-wheels2and3have to be rotated 1.2 times faster than the rear-wheels4and5.

The steering-wheel revolution number controller66calculates a revolution number Mrpm of the rear-wheels4and5. The revolution number Mrpm of the rear-wheels4and5is calculated based on the converter revolution number detecting signal QS from the converter output-side revolution sensor67and the speed range detecting signal TS from the speed range sensor36. The revolution number Mrpm of the rear-wheels4and5multiplied by 1.2 (1.2 Mrpm) is the necessary revolution number required for the front-wheels2and3.

Further, the steering-wheel revolution number controller66monitors the engine speed detecting signal ES from the engine speed sensor65and monitors output condition of the speed range switching signal MS to the solenoid valve26, i.e., the speed range of the hydraulic motors14and15.

In the below description, it is supposed that the engine speed of the engine6is judged as 1000 rpm based on the engine speed detecting signal ES and the speed range of the hydraulic motors14and15is judge as the first step.

According to the above condition, the necessary revolution number of the front-wheels2and3is 1.2 Mrpm and the engine speed of the engine6is 1000 rpm, so that the revolution number ratio n2is 1.2×10−3Mrpm. Accordingly, the steering-wheel revolution number controller66refers to the lookup table based on the graph ofFIG. 9to obtain the optimum swash plate angle θ1when the speed range of the hydraulic motor14and15is in the first step and adjusts the swash plate angle of the hydraulic pump13to θ1to obtain optimum flow rate of the hydraulic oil to rotate the front-wheels2and3driven by the hydraulic motors14and15at a speed 1.2 times faster than the rear-wheels4and5.

Incidentally, when the speed range of the hydraulic motors14and15is in the second step, as shown inFIG. 9, the steering-wheel revolution number controller66obtains an optimum swash angle θ2. The swash plate angle θ2is smaller than the swash plate angle θ1, so that the front wheels2and3are rotated at the same revolution number as in the swash plate angle θ1with less discharge flow rate of hydraulic oil.

The above is the flow of rotation correction control of the front-wheels2and3while turning the motor grader1.

Incidentally, even when the revolution number of the rear wheels4and5is the same Mrpm, the engine speed of the engine6may become different. In the above example, when the revolution number of the rear-wheels4and5is Mrpm, the engine speed of the engine6is 1000 rpm. However, when the speed range of the transmission9is in a higher position (small gear ratio side), the same revolution number of Mrpm of the rear wheels4and5can be obtained with low engine speed of about 800 rpm.

Because not so much of torque is required in descending a slope, even when the revolution number Mrpm of the rear-wheels4and5can be obtained with low engine speed though the transmission9is in the same speed range.

In this condition, since the engine speed 1000 rpm is lowered to 800 rpm in the graph ofFIG. 9, the ratio of revolution number n2increases to 1.5×10−3M.

However, because the n2becomes too great and, when the speed range of the hydraulic motors14and15is in the first step, the optimum swash plate angle may become θ3to exceed the maximum swash plate angle θMAX(the revolution number of the front-wheels2and3does not catch up with the revolution number of the rear-wheels4and5even when the swash plate angle θ is maximized), the controller100automatically switches the speed range of the motors14and15to the second step in advance. Accordingly, the steering-wheel revolution number controller66obtains an optimum swash plate angle θ4, not the optimum swash plate angle θ3.

When the turning direction of the front-wheels2and3and the articulation direction of the respective frames51and52differ, the direction of the front-wheels2and3may be aligned to the direction of the rear-wheels4and5to run the motor grader1straight. In this case, the actual turning radius is detected as R=∞ according to the above function (2) and it is determined that the ratio of the revolution number n1=1 in the graph of FIG.8. In other words, the necessary revolution number required for the front-wheels2and3in the graph ofFIG. 9becomes equal to the revolution number of the rear-wheels4and5. The optimum swash plate angle θ is calculated from the necessary revolution number in the same manner as in the turning movement.

When the vehicle is turned while the blade50is in contact with the ground or the ground is slippery, and when the turning radius R is twenty meters as in the above, the front-wheels2and3can be slipped when the front wheels2and3are rotated 1.2 times faster than the rear-wheels4and5as in the above. In this case, it is desired that the revolution number ratio n1(FIG. 8) is set smaller.

Accordingly, in the steering-wheel revolution number correction system60of the present embodiment, the operator having checked the ground condition operates the front-wheel rotation control lever35shown inFIG. 4to shift the relationship curve inFIG. 8leftward to uniformly decrease the revolution number ratio n1(FIG. 8) at the turning radius R (see single-dotted line). Accordingly, the torque in the front-wheels2and3can be restrained, thereby further preventing slippage.

According to the present embodiment, the following effects can be obtained.(1) Because the turning radius R is determined from the steering angle αLand αRof the front-wheels2and3and the articulation angle β of the respective frames51and52and the front-wheels2and3are automatically rotated faster than the rear-wheels4and5based on the turning radius R, the revolution number M of the rear-wheels4and5, the engine speed of the engine6and the speed range of the hydraulic motors14and15, the conventional trouble for operating the correcting lever while turning the motor grader1or the skilled work for adjusting the revolution number of the front-wheels2and3to the revolution number of the rear-wheels4and5can be eliminated, so that the revolution number of the front-wheels2and3can be easily and accurately controlled to securely prevent the tight-corner braking phenomenon.(2) Because the target revolution number of the front wheels2and3is determined based on the turning radius R, the front-wheels2and3can be rapidly rotated at a revolution number corresponding to the turning radius R where, for instance, the target revolution number of the front-wheels2and3can be set higher as turning radius R becomes smaller, so that the tight-corner braking phenomenon can be further securely prevented.

Further, in the present system60, the revolution number of the front-wheels2and3can be sufficiently set high irrespective of slippage of the rear-wheels4and5, so that the front-wheels2and3can be securely rotated at a high speed while turning the vehicle without necessarily accompanying slippage of the rear wheels4and5.(3) Because the revolution number ratio n1in increasing the speed of the front-wheels2and3can be changed by the front-wheel rotation control lever35, the torque of the front-wheels2and3can be lowered by setting the revolution number ratio n1at a lower level when the front-wheels2and3are likely to slip, so that the front-wheels2and3can be securely in contact with the ground to prevent slippage. Accordingly, the vehicle can be turned smoothly when the blade50is in contact with the ground to reduce ground pressure of the front-wheels2and3or when the vehicle is on pressed snow.(4) Because the flow rate of delivery oil from the hydraulic pump13can be controlled by operating the front-wheel rotation control lever35and the revolution number can be adjusted as desired in accordance with slippage of the front-wheels2and3, when the blade50is used during operation, continuous slippage of the front-wheels2and3with decreased ground pressure can be avoided, thus preventing useless energy consumption.(5) Because the rear-wheel revolution number sensor64is composed of the converter output rotation sensor67for detecting revolution number of output side of the torque converter8and the speed range sensor36for detecting the speed range of the transmission9connected to the output side of the torque converter8, the revolution number Mrpm of the rear-wheels4and5can be accurately calculated by detecting the revolution number inputted into the transmission9and the speed range in the transmission9even in the present embodiment where the engine speed does not directly become the input revolution number of the transmission9on account of the torque converter8.(6) Because the steering-wheel revolution number correction system60rotates the hydraulic motors14and15and, as a result, the front-wheels2and3by controlling the flow rate of the delivery oil from the hydraulic pump13, the size of the hydraulic motors14and15can be reduced by simplifying the structure thereof, where, for instance, the hydraulic motors14and15is not required to be variable capacity type. Accordingly, the hydraulic motors14and15can be suitably accommodated in a narrow space such as inside of the knuckle70.(7) Though the two-stage speed range is set in the hydraulic motors14and15, since the steering-wheel revolution number correction system60deals the speed range of the hydraulic motors14and15as a parameter for controlling the flow rate of the delivery oil from the hydraulic pump13, the speed range can be accurately detected, so that the swash plate angle θ of the hydraulic pump13can be accurately adjusted in any arrangement where the speed ranges are set in either the first or the second step, thereby obtaining an optimum delivery oil flow rate to smoothly turn the vehicle.(8) In order to curb the fuel consumption of the engine6, the rear-wheels4and5are sometimes rotated at a high speed by setting the speed range of the transmission9on the side of small gear ratio while maintaining the low engine speed. In this case, since the hydraulic pump13is also driven at a low revolution number, even when the discharge flow rate by the hydraulic pump13is set maximum, the rotation of the hydraulic motors14and15reaches limit thereof, so that the front-wheels2and3may not be able to rotate faster than the rear-wheels4and5.

On the other hand, by using the hydraulic motors14and15of the present embodiment, when the delivery oil flow rate from the hydraulic pump13is maximized and the hydraulic motors14and15cannot be rotated more rapidly at the first speed range, the hydraulic motors14and15can be rotated at a higher speed by switching the speed range to the second step. Further, fuel expenses can be lowered by driving with low torque and high-speed.

The second embodiment includes a steering-wheel revolution number correction system160instead of the steering-wheel revolution number correction system60of the first embodiment. Accordingly, the description for common components will be omitted and the description will be concentrated on the steering-wheel revolution number correction system160as a different component.

InFIG. 10, the steering-wheel revolution number correction system160includes a left pickup sensor161as a steering-wheel revolution number detecting means for detecting the revolution number of the left front-wheel2, a right pickup sensor162as another steering-wheel revolution number detecting means for detecting the revolution number of the right front-wheel3, a turning radius operation163for determining the turning radius of the front-wheels2and3based on the revolution number of the front wheels2and3, a rear-wheel revolution number sensor164for detecting revolution number of the rear-wheels4and5, an engine speed sensor165as an engine speed detecting means for detecting the engine speed, and a steering-wheel revolution number controller166for controlling the revolution number of the front-wheels2and3based on the turning radius, the revolution number of the rear-wheels4and5and the engine speed of the engine6.

The pickup sensors161and162output a steering-wheel revolution number detecting signals NSLand NSRcorresponding to the revolution number of the front-wheels2and3to the controller100.

The rear-wheel revolution number sensor164is composed of a converter output-side revolution sensor167as a converter output revolution number detecting means for detecting the revolution number of the output side of the torque converter8and the speed range sensor36for detecting the speed range of the transmission shift lever39.

The converter output-side revolution sensor167output a converter revolution number detecting signal QS to the controller100.

The engine speed sensor165outputs an engine revolution number detecting signal ES to the controller100.

The turning radius operation163and the steering-wheel revolution number controller166are software, which is stored in the memory101constituting the controller100such as ROM and RAM and is called from the memory101when the system60is actuated and is implemented by the CPU102as shown in FIG.10.

The memory101of the controller100stores a lookup table (not shown) tabling the graph shown inFIGS. 8,9and11as well as the above software.

FIG. 11is a graph showing turning radius R and turning direction of the front-wheels2and3corresponding to respective revolution number NLand NRof the left front-wheel2and the right front wheel3.

FIGS. 8 and 9are the same as described in the first embodiment.

Next, revolution number control of the front-wheels2and3by the steering-wheel revolution number correction system160will be described below in detail.

Initially, when the motor grader1is running, the turning radius operation163monitors the steering-wheel revolution number detecting signal NSLand NSRoutputted from the respective pickup sensors161and162. When the motor grader1starts turning, for instance, rightward, because the left front-wheel2is located outside and the right-front-wheel3is located inside, the left front-wheel2rotates faster than the right front-wheel3while being slightly braked.

The turning radius operation163judges that the revolution number NLof the left front-wheel2is greater than the revolution number NRof the right front-wheel3and the vehicle is turning rightward based on the information of the steering-wheel revolution number detecting signal NSLand NSR, and obtain the turning radius, for instance, twenty meters, with reference to the lookup table in the memory101based on the graph of FIG.11.

The steering-wheel revolution number controller166calculates a revolution number ratio n1when the turning radius R=twenty meters with reference to the lookup table based on the graph of FIG.8. According to the graph ofFIG. 8, n1=1.2, so that the steering-wheel revolution number controller166determines that the front-wheels2and3has to be rotated 1.2 times faster than the rear-wheels4and5.

The steering-wheel revolution number controller166calculates a revolution number Mrpm of the rear-wheels4and5. The revolution number Mrpm of the rear-wheels4and5is calculated based on the converter revolution number detecting signal QS from the converter output-side revolution sensor167and the speed range detecting signal TS from the speed range sensor36. The revolution number Mrpm of the rear-wheels4and5multiplied by 1.2 (1.2 Mrpm) is the necessary revolution number required for the front-wheels2and3.

Further, the steering-wheel revolution number controller166monitors the engine speed detecting signal ES from the engine speed sensor165and monitors output condition of the speed range switching signal MS to the solenoid valve26, i.e., the speed range of the hydraulic motors14and15.

In the below description, it is supposed that the engine speed of the engine6is judged as 1000 rpm from the engine speed detecting signal ES and the speed range of the hydraulic motors14and15is in the first step.

According to the above condition, the necessary revolution number of the front-wheels2and3is 1.2 Mrpm and the engine speed of the engine6is 1000 rpm, so that the revolution number ratio n2is 1.2×10−3Mrpm. Accordingly, the steering-wheel revolution number controller166refers to the lookup table based on the graph ofFIG. 9to obtain the optimum swash plate angle θ1when the speed range of the hydraulic motor14and15is in the first step and adjusts the swash plate angle of the hydraulic pump13to θ1to obtain optimum flow rate of the hydraulic oil to rotate the front-wheels2and3driven by the hydraulic motors14and15at a speed 1.2 times faster than the rear-wheels4and5.

Incidentally, when the speed range of the hydraulic motors14and15is in the second step, as shown inFIG. 9, the steering-wheel revolution number controller166obtains an optimum swash angle θ2. The swash plate angle θ2is smaller than the swash plate angle θ1, so that the front wheels2and3are rotated at the same revolution number as in the swash plate angle θ1with less discharge flow rate of hydraulic oil.

The above is the flow of rotation correction control of the front-wheels2and3while turning the motor grader1.

Incidentally, even when the revolution number of the rear wheels4and5is the same Mrpm, the engine speed of the engine6may become different. In the above example, when the revolution number of the rear-wheels4and5is Mrpm, the engine speed of the engine6is 1000 rpm. However, when the speed range of the transmission9is in a higher position (small gear ratio side), the same revolution number of Mrpm of the rear wheels4and5can be obtained with low engine speed of about 800 rpm.

Because not so much of torque is required in descending a slope, even when the revolution number Mrpm of the rear-wheels4and5can be obtained with low engine speed though the transmission9is in the same speed range.

In this condition, because the engine speed 1000 rpm is lowered to 800 rpm in the graph ofFIG. 9, the ratio of revolution number n2increases to 1.5×10−3M.

However, because the n2becomes too great and, when the speed range of the hydraulic motors14and15are in the first step, the optimum swash plate angle may become θ3to exceed the maximum swash plate angle θMAX(the revolution number of the front-wheels2and3does not catch up with the revolution number of the rear-wheels4and5even when the swash plate angle θ is maximized), the controller100automatically switches the speed range of the motors14and15to the second step in advance. Accordingly, the steering-wheel revolution number controller166obtains an optimum swash plate angle θ4, not the optimum swash plate angle θ3.

When the motor grader1moves straight, the turning radius is detected as R=∞ since there is no difference between the revolution number NLand NRof the front-wheels2and3and it is determined that the ratio of the revolution number n1=1 in the graph of FIG.8. In other words, the necessary revolution number required for the front-wheels2and3in the graph ofFIG. 9becomes equal to the revolution number of the rear-wheels4and5. The optimum swash plate angle θ is calculated from the necessary revolution number in the same manner as in the turning movement.

When the vehicle is turned while the blade50is in contact with the ground or the ground is slippery, and when the turning radius R is twenty meters as in the above, the front-wheels2and3can be slipped when the front wheels2and3are rotated 1.2 times faster than the rear-wheels4and5as in the above. In this case, it is desired that the revolution number ratio n1(FIG. 8) is set smaller.

Accordingly, in the steering-wheel revolution number correction system160of the present embodiment, the operator having checked the ground condition operates the front-wheel rotation control lever35shown inFIG. 10to shift the relationship curve inFIG. 8leftward to uniformly decrease the revolution number ratio n1(FIG. 8) at the turning radius R (see single-dotted line). Accordingly, the torque in the front-wheels2and3can be restrained to prevent slippage.

According to the present embodiment, the following effect can be obtained.(1′) Because the turning radius R is determined from the revolution number NLand NRof the front-wheels2and3, and the front-wheels2and3are automatically rotated faster than the rear-wheels4and5based on the turning radius R, the revolution number M of the rear-wheels4and5, the engine speed of the engine6and the speed range of the hydraulic motors14and15, the conventional trouble for operating the correcting lever while turning the motor grader1or the skilled work for adjusting the revolution number of the front-wheels2and3to the revolution number of the rear-wheels4and5can be eliminated, so that the revolution number of the front-wheels2and3can be easily and accurately controlled to securely prevent the tight-corner braking phenomenon.

Further, the effects (2) to (8) of the above-described first embodiment can be obtained in the second embodiment.

Incidentally, the scope of the present invention is not restricted to the above embodiment but includes other arrangement as long as an object of the present invention can be achieved, which includes following modifications.

For instance, the arrangement shown inFIG. 12may be applied to the first embodiment (FIG. 4) and the arrangement shown inFIG. 13may be applied to the second embodiment (FIG.10). In the arrangement of theFIGS. 12 and 13, in addition to controlling the delivery oil flow rate by setting the hydraulic pump as variable capacity type, hydraulic motors114and115are made variable capacity type such as swash plate type and angled axis type to vary the revolution number at a predetermined supply oil flow rate, where the swash plate angle and angle of the angled axis of the hydraulic motors114and115are independently controlled by control signals CSLand CSRfrom the controller100to rotate the front-wheels at a speed higher than the rear-wheels4and5without employing the solenoid valve26shown inFIG. 4or10.

According to the arrangement using the variable capacity hydraulic motors114and115, the outer wheel of the left and right wheels2and3can be positively moved faster and the inner wheel can be positively moved slower in turning the motor grader1, so that the turning movement can be conducted more smoothly.

Because the hydraulic motors and115can be directly controlled by the control signals CSLand CSRfrom the controller100to change the revolution number, the low-torque high-speed drive can be securely achieved when the delivery oil flow rate from the hydraulic pump13is small. Accordingly, when the engine speed of the engine6is rapidly decreased from the high-speed side on account of the characteristics of the torque converter8, the rotation of output side of the torque converter8can be maintained at a high speed, so that the front-wheels can be continuously rotated at a high speed even when the rear-wheels4and5still are rotated at a high-speed, thereby restraining tight-corner braking phenomenon during the condition.

Incidentally, the hydraulic pump13may be stationary type and the front-wheels2and3may be rotated faster than the rear-wheels4and5only by controlling the hydraulic motors114and115, which is included in the scope of the present invention.

The swash plate angle and angle of the angled axis of the hydraulic motors114and115may not be controlled by independent control signals CSLand CSRbut may be controlled by a single control signal.

Though each of the right and left front wheels2and3are driven by independent hydraulic motors14,15,114and115in the above-described embodiments and modifications shown inFIGS. 12 and 13, both of the front-wheels2and3may be driven by a single hydraulic motor.

Both of the steering angle αLand αRof the front-wheels2and3are detected in the first embodiment because the steering angle αLand αRcaused by steering operation differ in the left front-wheel2and the right front-wheel3on account of link mechanism, and also because the turning direction can be easily identified by the difference between the steering angle αLand αR. In other words, since the above embodiment uses a link mechanism where the inner wheel steering angle is always greater than the steering angle on the outer wheel side during turning movement, when, for instance, the steering angle αLis detected greater than the steering angle αR, it is easily recognized that the vehicle is in leftward turning where the left wheel2is located inner side and the right wheel3is located outside.

The above explanation suggests that the turning radius RSTis not necessarily obtained from both of the steering angles αLand αRif the turning direction of the front-wheels2and3can be identified by the other means such as a sensor. Specifically, if the turning direction of the front-wheels2and3can be identified, one of the front-wheel steering angles relative to the other front-wheel steering angle can be calculated in view of the structure of the link mechanism, and the rotating radius RSTcan also be calculated from the actually detected steering angle and the calculated steering angle.

However, because it is costly to provide other means for identifying the turning radius of vehicle and the calculation of the turning radius can be complicated, it is preferable that both of the steering angle and the turning angle can be detected using the same two steering angle sensors61and62as in the above embodiment.

Though the steering angle sensors61and62for detecting the steering angles αLand αRaround the king pin73are used as the front-wheel steering angle detecting means in the first embodiment, the front-wheel steering angle detecting means according to the present invention is not restricted to such arrangement, but a sensor for detecting advancement and retraction of the right and left steering cylinders74may be used.

Though the angle detection signal AS from the angle sensor38for detecting the articulation angle is used only for displaying rotation of the front frame51on the indicator in the cockpit in the second embodiment, the angle detecting signal AS may be used as a parameter for controlling the delivery oil flow rate from the hydraulic pump13.

When the front frame51is turned relative to the rear frame52, because the difference between the turning radius of the front-wheels2and3and the turning radius of the rear-wheels4and5becomes greater, though the front-wheels2and3are of the same turning radius, there may be difference in easiness for turning the vehicle. Accordingly, when the front frame51is turned, the revolution number ratio n1ofFIG. 8determined when the front frame51is not turned may be properly corrected in accordance with the articulation angle. Accordingly, the revolution number of the front-wheels2and3can be adjusted properly in accordance with the articulation angle, thereby achieving the steering-wheel revolution number correction system more suitable for the motor grader1.

The steering-wheel revolution number correction system according to the present invention is not restricted to be applied to the motor grader1as in the respective embodiments, but may be applied to the other all-wheel-drive vehicle having an all-wheel drive having a rear-wheel driven by an output of the engine transmitted through a transmission, a hydraulic pump receiving the output of the engine and a hydraulic motor rotated by the delivery oil from the hydraulic pump to drive the front-wheel.

For instance, the steering-wheel revolution number correction system60of the first embodiment is not limited to be applied to the motor grader1, but may be used for an articulating type all-wheel-drive vehicle having a front frame provided with a front-wheel and a rear frame provided with the rear-wheel which are angle-adjustably connected such as a construction equipment of a large-size wheel loader having an articulating frame etc.

The steering-wheel revolution number correction system160of the second embodiment is not restricted to be applied to the motor grader1but may be used for the other construction equipment such as a wheel loader and a work vehicle such as non-articulating dump truck.