JPH0577532B2 - - Google Patents

Description

[Detailed description of the invention] [Industrial application field]

The present invention relates to a hydraulic control device for a four-wheel drive system for a vehicle, and more particularly, when a wet multi-plate hydraulic clutch is used in a four-wheel drive transfer system, clutch torque control is performed at each shift position of the hydraulic clutch. Regarding.

[Conventional technology]

The four-wheel drive device of the transmission system in a four-wheel drive vehicle is
It is based on a front engine/front drive (FF) or a rear engine/rear drive (RR), and is structured to further transmit power to the other front and rear wheels using a transfer device during four-wheel drive. Some transfer clutches use hydraulic clutches. By the way, in a direct-coupled four-wheel drive system that does not have a center differential and drives the four wheels through the engagement of a transfer clutch, the front and rear wheels are almost directly coupled to drive the vehicle on a road surface with a large tire grip force, such as a paved road. In this case, if a large rotational difference occurs between the front and rear wheels when turning, excessive torsional torque will be generated in the drive system. This is known to cause problems such as braking and stalling of the engine, and increased steering force. Conventionally, in a four-wheel drive system equipped with the above-mentioned hydraulic transfer device, hydraulically controlled systems for the purpose of avoiding braking phenomena have been disclosed, for example, in JP-A-56-138020 and JP-A-57-1.
There is a prior art in Publication No. 84227. The former is based on 4-wheel drive in driving conditions below the set vehicle speed, including starting, to improve starting acceleration, and when turning, switches to 2-wheel drive to avoid braking phenomena. It's getting old. The latter type switches to two-wheel drive in the same way as described above when torsional torque is generated in the drive system due to turning or the like during four-wheel drive.

[Problems to be solved by the invention]

By the way, in the case of the prior art described above, the tight corner braking phenomenon can be avoided by switching to two-wheel drive when turning, but in this case, the performance as four-wheel drive is completely lost. Become. Therefore, road surfaces that are particularly susceptible to slips,
There is a problem in that the unique performance of a four-wheel drive vehicle cannot be fully demonstrated in driving conditions such as when riding on a hill. Therefore, it is desired as an optimal control to avoid the braking phenomenon while maintaining the four-wheel drive state even when turning. The braking phenomenon mainly occurs when the vehicle makes a large turn at low speeds, including when starting. Under these driving conditions, the engine-side clutch is often in a half-clutched state, and the magnitude of steering is performed depending on the degree of half-clutching. Therefore, if the transfer device is equipped with a hydraulic clutch and the transmission torque of the hydraulic clutch is controlled in relation to the half-clutch state of the engine side clutch, the clutch torque can be reduced to avoid the braking phenomenon. This makes it possible to simultaneously satisfy two conflicting demands: increasing the clutch torque for four-wheel drive and increasing the clutch torque. In addition, under driving conditions where the clutch on the engine side is directly connected, if the engine torque is used to control the torque of the transfer clutch according to the load, it will be possible to avoid braking phenomena when turning at low loads, and to avoid braking when turning with high loads. It becomes possible to demonstrate the performance of the drive together. Furthermore, in general vehicles, in addition to forward and reverse gears,
In the forward gear, there are also a plurality of shift positions between the low gear and the high gear, and in the case of an automatic transmission, there are respective ranges that limit the shift range. Here, especially in the case of an automatic transmission, the range in which the shifting range is limited is intended to apply engine braking, improve hill climbing, and improve drivability on rough roads, and the reverse gear is also used for escaping from rough roads, etc. In such driving conditions, there is little possibility of large steering changes. Therefore, by correcting the torque capacity using signals that detect driving conditions such as the shift position, accelerator opening, and vehicle speed, it is possible to realize the duality of avoiding braking phenomena and strengthening four-wheel drive. The present invention has been made in view of the above points, and is a four-wheel drive system that is capable of optimally controlling the torque capacity of the hydraulic clutch of the transfer device depending on each shift position throughout the entire range from the start of the vehicle to the time the vehicle is running. The purpose is to provide a hydraulic control device for the device.

[Means to solve the problem]

In order to achieve the above object, the present invention provides a four-wheel system in which power is directly transmitted to one of the front and rear wheels through an engine side clutch and a transmission, and power is transmitted to the other wheel through a transfer device having a hydraulic clutch. In the drive device, there is a means for detecting whether the engine-side clutch is in a half-clutch state and a direct connection state, and under low-speed driving conditions including starting with the half-clutch clutch, the accelerator opening, vehicle speed, etc. are determined based on the torque capacity of the engine-side clutch. means for setting the torque capacity of the hydraulic clutch in response to a signal for detecting a running condition;
At least under the above-mentioned low-speed driving condition with the half-clutch engaged, the hydraulic clutch is characterized by comprising means for correcting the torque capacity of the hydraulic clutch based on a detection signal of the shift position and a signal detecting driving conditions such as accelerator opening and vehicle speed. do.

[Effect]

Based on the above structure, the present invention divides the engine-side clutch into each region directly connected to the half-clutch, and based on the torque capacity of the power transmitted in each region,
Hydraulic control is performed to determine the torque capacity of the hydraulic clutch of the transfer device, taking each running state into consideration. Then, at least under the above-mentioned low-speed running conditions with the half-clutch engaged, the torque capacity of the hydraulic clutch is corrected based on the shift position of each range of D, R, _{and Ds} , and the running state of each load, thereby reducing the braking phenomenon that accompanies turning. Of course, it is important to avoid
In other cases, it is possible to make full use of the four-wheel drive performance and improve performance on the hill, on rough roads, etc.

【Example】

Embodiments of the present invention will be specifically described below based on the drawings. Referring to FIG. 1, as an example of a four-wheel drive system to which the present invention is applied, a FF-based transverse transaxle type vehicle in which an electromagnetic powder clutch is combined with a belt-type continuously variable transmission will be described. Reference numeral 1 is an electromagnetic powder clutch, 2 is a forward/reverse switching device, 3 is a continuously variable transmission, 4 is a front differential device, 5
is a hydraulic transfer device. The electromagnetic powder clutch 1 is housed in one side of the clutch housing 6, the main case 7 is connected to the other side of the clutch housing 6, and the side case is connected to the side of the main case 7 opposite to the clutch housing 6. 8, a forward/reverse switching device 2,
A continuously variable transmission 3, a front differential device 4, and a transfer device 5 are accommodated, and a clutch housing 6
An extension case 9 is joined to the rear part of the housing. The electromagnetic powder clutch 1 includes a ring-shaped drive member 12 that is integrally connected to a crankshaft 10 from the engine via a drive plate 11, and a disk-shaped driven member that is integrally connected to a transmission input shaft 13 by splines in the rotational direction. It has 14. A coil 15 is placed on the outer peripheral side of the driven member 14.
is built in, and a gap 16 is formed along the circumference between both members 12 and 14.
has electromagnetic powder. Additionally, a slip ring 18 at the hub portion of the driven member 14 equipped with the coil 15
A power feeding brush 19 is slid thereon, and is connected from the slip ring 18 through the inside of the driven member 14 to the coil 15 to form a clutch current circuit. In this way, when a clutch current is passed through the coil 15, the magnetic lines of force generated between the drive and driven members 12 and 14 via the gap 16 cause
Electromagnetic powder is bonded and accumulated in the gap 16 in a chain, and the resulting bonding force causes the driven member 14 to slide and be integrally bonded to the drive member 12, resulting in a clutch connected state. On the other hand, when the clutch current is cut off, the drive by the electromagnetic powder and the coupling force between the driven members 12 and 14 disappear, resulting in the clutch being in a disengaged state. If the clutch current in this case is controlled in conjunction with the operation of the forward/reverse switching device 2, it is possible to switch from the P (parking) or N (neutral) range to the forward D (drive) or D _s (sport) range. Drive) or Reverse R
When switching to the (reverse) range, clutch 1 is automatically connected and disconnected, eliminating the need for clutch pedal operation. Next, the forward/reverse switching device 2 is provided between the input shaft 13 from the clutch 1 and a main shaft 20 disposed coaxially therewith. That is, the input shaft 13 is formed with a reverse drive gear 21 that also serves as a forward engaged side, and the main shaft 20 is formed with a reverse drive gear 21 that also serves as a forward engaged side.
2 are rotatably fitted, and these gears 2
1 and 22 are counter gears 2 supported by a shaft 23
4. They are meshed through an idler gear 26 supported by a shaft 25. And the main shaft 20 and gear 21
A switching mechanism 27 is provided between and 22 . The gears 21 and 2 that are always in mesh here
4, 26, and 22 are connected to the driven member 14 having the coil 15 of the clutch 1, and corresponding to the fact that the inertia mass of this part is relatively large when the clutch is disengaged, the switching mechanism 27 is connected to the hub of the main shaft 20. 28
A sleeve 29 is spline-fitted to the gears 21 and 22 through synchronizing mechanisms 30 and 31. As a result, in the neutral position of P or N range,
The sleeve 29 of the switching mechanism 27 is fitted only with the hub 28, and the main shaft 20 is separated from the input shaft 13.
Next, when the sleeve 29 is meshed with the gear 21 side via the synchronizing mechanism 30, the main shaft 20 is directly connected to the input shaft 13, resulting in a forward movement state in the D or _Ds range. On the other hand, when the sleeve 29 is meshed with the gear 22 side via the synchro mechanism 31, the input shaft 13 is connected to the main shaft 20 via the gears 21, 24, 26, 22, and the engine power is decelerated and reversed. The vehicle will be in reverse mode in R range. The continuously variable transmission 3 has a subshaft 35 with respect to the main shaft 20.
are arranged in parallel, a main pulley 36 and a sub pulley 37 are provided on both shafts 20 and 35, respectively, and an endless drive belt 34 is stretched between both pulleys 36 and 37. The pulleys 36 and 37 are both divided into two parts, one half of the pulley 3
6a, 37a, the other pulley half 36b,
37b is movable to make the pulley interval variable, and the movable pulley halves 36b, 37b are equipped with hydraulic servo devices 38, 39 which also serve as pistons.
Further, a spring 40 is applied to the movable pulley half 37b of the sub pulley 37 in a direction to narrow the pulley interval. Further, as a hydraulic control system, an oil pump 41 as an operating source is installed next to the main pulley 36. This oil pump 41 is a high pressure gear pump,
The pump drive shaft 42 passes through the main pulley 36, the main shaft 20, and the input shaft 13 to connect to the crankshaft 10.
The oil pressure is directly connected to the engine, and hydraulic pressure is constantly generated while the engine is running. Then, the oil pressure of the oil pump 41 is controlled to supply and drain oil to each hydraulic servo device 38 and 39, and the pulley spacing between the main pulley 36 and the sub pulley 37 is changed to an inverse relationship, so that the pulley 36 of the drive belt 34 , 37 are converted steplessly, and steplessly variable power is output to the subshaft 35. The front differential device 4 has one set of pulleys for the countershaft 35 in view of the fact that the minimum pulley ratio on the high speed side of the continuously variable transmission 3 is very small, for example 0.5, and therefore the rotation speed of the countershaft 35 is high. An output shaft 44 is connected via an intermediate reduction gear 43 . A final gear 46 is connected to the drive gear 45 of this output shaft 44.
mesh, and the final gear 46 to differential mechanism 47
Power is transmitted to the axles 48 and 49 of the left and right front wheels via the axles 48 and 49 of the left and right front wheels. Further, in the transfer device 5, a transfer gear 50 that meshes with the final gear 46 is rotatably fitted to a transfer shaft 51 installed laterally to the vehicle body, and the transfer gear 50 and the shaft are connected to each other. 51, a wet multi-plate hydraulic clutch 52 for four-wheel drive is provided. The direction of the transfer shaft 51 is changed by a pair of bevel gears 53 and 54, and the transfer shaft 51 is connected to a rear drive shaft 55.
Transmission is further configured from the rear drive shaft 55 to the rear wheel side. The hydraulic clutch 52 has a hub 56 that is integral with the transfer gear 50 and a drum 57 that is integral with the transfer shaft 51. Between the hub 56 and the drum 57, there is a plate 59 that is pressed by a piston 58. It will be installed in a multi-panel format. A return spring 60 is applied to the piston 58, and a piston chamber 61 is provided on the opposite side of the plate 59. Further, in the main case 7, a cover 62 is attached on an extension line of the transfer shaft 51, a valve body 63 is attached to the main case 7 inside this cover 62, and a valve body 63 is attached to the main case 7.
3 is equipped with a solenoid means 64. In this way, the clutch hydraulic pressure from the valve body 63 is transmitted to the piston chamber 6 through the oil passage 65 of the transfer shaft 51, etc.
1 to control clutch torque. In FIG. 2, to explain the hydraulic control system of the continuously variable transmission 3, the main pulley side hydraulic servo device 3
8, a cylinder 38a integral with the main shaft 20;
The movable pulley half 36b is fitted into the cylinder 3.
It has a main pulley servo chamber 38b into which line pressure is introduced. Also, in the sub-pulley side hydraulic servo device 39, the cylinder 3 integrated with the sub-shaft 35
The movable pulley half 37b is fitted into the cylinder 39a, and has an auxiliary pulley servo chamber 39b into which line pressure is introduced into the cylinder 39a. The area receiving pressure is increasing. The oil pumped up from the oil reservoir 70 by the oil pump 41 is guided to the pressure regulating valve 80 via the oil passage 71, and a line pressure oil passage 72 branching from the oil passage 71 is always connected to the sub-pulley servo chamber 39b. Communicate to introduce pressure. The oil passage 71 further communicates with a gear ratio control valve 90, and an oil passage 73 for supplying and discharging line pressure between this gear ratio control valve 90 and the main pulley servo chamber 38b communicates, and drains each valve 80, 90. Oil passages 74 and 75 communicate with the oil reservoir side. In addition, a rotation sensor 76 is installed at the cylinder 38a on the main pulley side to take out a control signal pressure of the pitot pressure according to the engine speed in the shift control after the clutch is engaged. is guided to each valve 80, 90 via an oil passage 77. Furthermore, in contrast to the D range, which performs shift control over a wide range including low engine speeds, the D _s range performs shift control only in a high engine speed range, and applies engine braking when the accelerator is released. As a hydraulic system, a relief valve 78 is provided in a drain oil path 74 from a pressure regulating valve 80, and an oil path 79 of a lubrication hydraulic circuit that branches from the upstream side of this relief valve 78 communicates with a select position detection valve 110. An oil passage 88 further branches from the oil passage 79,
It communicates with the actuator 120 of the gear ratio control valve 90. The pressure regulating valve 80 includes a valve body 81 and a spool 8.
2. A sensor shoe 85, which has a spring 84 biased between it and one bush 83 of the spool 82 and engages with the movable pulley half 36b of the main pulley to detect the actual gear ratio, connects the lubrication passage. It is movably supported by a shaft tube 86 that also serves as a shaft tube and communicates with a bush 83. In the valve body 81, the oil passage 7 is connected to the port 81a of the spool 82 on the opposite side from the spring 84.
7, and the pump oil pressure of the oil passage 71 is introduced to the port 81b. Further, the oil passage 71 and an oil passage 87 to the speed ratio control valve 90 communicate with the port 81c. A port 81f on the spring 84 side of this port 81c, and a port 81e that is provided between ports 81a and 81b to prevent leakage of pump hydraulic pressure from affecting pitot pressure are provided, and leaked oil is drained to the oil reservoir 70. guided by. Also, the port 8 is connected to the channel portion of the land 82a of the spool 82.
1c and 81d are connected to each other for pressure regulation. That is, the pitot pressure and the pump hydraulic pressure act on the spool 82 in the direction of opening the drain port 81d, whereas the load of the spring 84 according to the speed ratio by the sensor shoe 85 acts on the drain port 81d.
Acts in the direction of closing 1d. As a result, for example, in a low speed gear with a large gear ratio, high line pressure is generated in the port 81c to avoid belt slip, and by moving the pulley half 36b to the right in the figure, it is difficult to shift to a high gear gear with a small gear ratio. Accordingly, the sensor shoe 85 moves to the right in the drawing, and controls the line pressure to be reduced by reducing the load on the spring 84, thus maintaining a pulley pressing force that does not cause belt slip. The gear ratio control valve 90 includes a valve body 91 and a spool 9.
2. A spring 94 is biased between the spool 92 and one operating plunger 93, and the pitot pressure of the oil passage 77 is applied to the port 91a at the end of the spool 92 opposite to the spring 94 in the valve body 91. be guided. Also, the oil passage 73 is connected to the intermediate port 91b.
However, the oil passage 87 is connected to the spring side port 91c.
However, the drain oil passage 75 communicates with the opposite port 91d, and the groove 92a of the spool 92 communicates with the port 91b.
and 91c or 91d are connected to supply and drain line pressure to the main pulley servo chamber 38b. An adjustment plunger 95 protrudes from the inside of the spool 92 toward the spring 94 and is movably inserted. An adjustment spring 97 is installed between the retainer 96 at the tip of the protrusion of the plunger 95 and the operating plunger 93. A return spring 98 is biased between the spool 92 and the spool 92 . The line pressure port 91c communicates with the inside of the spool 92 through the small hole 99 of the spool 92, and supplies line pressure to the spool 92 and the plunger 9.
5, the line pressure changes the amount of protrusion of the plunger 95 relative to the spool 92, that is, the load of the adjustment spring 97. Further, the operating plunger 93 receives the rod 1 from the cam 100 which acts as a lift depending on the accelerator opening degree.
It is separated from the rod 101 and connected via a weak spring 102, and has a stopper 103 to move the same stroke as the rod 101. And plunger 93
The inside communicates with the port 91a via the notch 104, the port 93a, the orifice 105, and the oil passage 106, and a spring 107 that adjusts the load of the spring 102 is attached to the valve body 91 at the end of the spool 92.
energized between. In this way, pitot pressure acts on the spool 92 in the direction of shifting up by introducing line pressure into the main pulley servo chamber 38b through communication between the ports 91b and 91c, while adjusting the line pressure with the spring 94 according to the accelerator opening degree. The load of the spring 97 acts in the direction of draining the main pulley servo chamber 38b and downshifting through communication between the ports 91b and 91d, and the gear ratio is determined by the balanced relationship between the two. Here, when the line pressure before the start of the shift is at its maximum, the adjustment plunger 95 is retracted the most and the spring 97
The load of spring 97 is zero, and from this, spring 97
The shift start point is determined in equilibrium with no shift, and after this shift start point, the load on the spring 97 is increased based on the decrease in line pressure, and the engine rotation is increased as the gear ratio is shifted to a high speed gear. Rise in number. Furthermore, the pitot pressure that is balanced in the above relationship is
It acts on the operating plunger 93 through the oil passage 106 and the like, and cancels out the force exerted on the plunger 93 due to the pitot pressure. In the select position detection valve 110, a valve body 113 having a drain hole 112 is inserted into a valve body 111.
A cam 115 that rotates in response to a selection operation is in contact with the valve body 113. Here, in the cam 115, the range positions of D, N, and R are in the convex part 115a, and the range positions of P and _Ds at both ends are in the concave part 115.
b, and the drain hole 112 is closed in each of the D, N, and R ranges to generate operating oil pressure. Further, an orifice 116 is provided on the upstream side of the branch of the oil passage 88 in the oil passage 79 to prevent the oil pressure in the oil passage 74 from decreasing when the drain hole 112 is opened in the P and D _s ranges. . In the actuator 120, a piston 122 is inserted into a cylinder 121, and a return spring 123 is biased on one side of the piston 122.
The operating oil pressure of the oil passage 88 is guided to the other piston chamber 124. Further, a lever 125 at the tip of the piston 122 can engage with a pin 126 of the rod 101 of the gear ratio control valve 90, and when there is no operating oil pressure in the P and D _s ranges, the piston 122 and the lever 125 close the rod 101. is forcibly pushed to a predetermined stroke to limit the shift range to the high engine speed side. As a result, when the accelerator is released in the D _s range, engine braking will be applied by downshifting. Furthermore, in order to correct the characteristics of the D _s range, a balance-type correction lever 128 is provided between the sensor shaft 85, which changes according to the gear ratio, and the lever 125 at the tip of the piston of the actuator 120, with the intermediate portion supported by a pin 127. It will be done. This correction lever 12
8, one end engages with the piston lever 125 only when the actuator 120 is pushed in, and in this state, the gear shift is shifted to the low gear side with a large gear ratio, and the sensor shoe 85 reaches the position of the predetermined gear ratio. Then, the lever 12 is attached to the sensor shoe 85.
8 is engaged. Thereafter, as the gear ratio increases, the piston 122 of the actuator 120 is pulled back, and the piston 122 is returned to approximately the original standby position at the maximum gear ratio. Next, the hydraulic control system of the hydraulic clutch 52 will be explained. First, an oil passage 130 branches from the oil passage 71 of the line pressure circuit in the hydraulic control system of the continuously variable transmission 3, and this oil passage 130 communicates with a pressure regulating valve 140 that always regulates the reducing pressure to a constant level. 130
An oil passage 131 branching from the transfer control valve 1
Connects to 50. Further, the reducing pressure oil passage 132 from the pressure regulating valve 140 communicates with the duty solenoid valve 133 via the control side of the control valve 150.
A clutch pressure oil passage 134 from the control valve 150 communicates with the piston chamber 61 of the hydraulic clutch 52. Note that the reference numeral 135 is an orifice. The pressure regulating valve 140 includes a valve body 141 and a spool 14.
2. A hydraulic chamber 143 that communicates with one reducing pressure oil passage 132 of the spool 142, an oil passage 144 that guides the reduced pressure to the oil passage 132, and the spool 14.
It consists of a spring 145 biased against the other of the two. And one hydraulic chamber 143 of the spool 142
The spool 142 is moved due to the balanced relationship between the force at
Deriving the line pressure of the oil passage 130 from a or draining it from the drain port 141b to adjust the pressure,
The oil pressure is guided to the reducing pressure oil passage 132 and the hydraulic chamber 143 through the oil passage 141, and thus a constant reducing pressure is always generated in the reducing pressure oil passage 132. That is, hydraulic chamber 1
If the land pressure receiving area at 43 is S, the reducing pressure is P _R , and the spring force is F, then P _R ·S=F, and a constant reducing pressure is always generated by P _R =F/S. The solenoid valve 133 closes the drain port 13 based on the duty signal from the control unit 160.
By opening 3a, the line pressure regulating valve 140 performs exhaust pressure control on the reducing pressure P _R to generate a control pressure P _C , which is applied to the transfer control valve 150. The transfer control valve 150 includes a valve body 151,
Spools 152 with different land pressure receiving areas, a hydraulic chamber 153 into which the control pressure P _C of one of the spools 152 is introduced, and a spring 154 biased toward the other side.
The oil passage 1 is introduced from the port 151a.
31 line pressure is controlled to generate clutch pressure P _T , and this clutch pressure P _T is taken out from port 151b to oil passage 134. That is, the force due to the clutch pressure P _T due to the land pressure receiving area difference of the spool 152,
The force due to the control pressure P _C of the hydraulic chamber 153 acts downward, and the force of the spring 154 acts upward in opposition thereto. And when the control pressure P _C increases,
When the spool 152 is moved downward to close the port 151a and the drain port 151c is opened to lower the clutch pressure P _T and the control pressure P _C is lowered, the spool 152 is moved upward to close the port 151a.
The clutch pressure P _T increases by increasing the opening degree of the clutch. As a result, the following equation holds true between the control pressure P _C , the clutch pressure P _T , the spring force F, the spool large diameter area S ₁ , and the small diameter area S ₂ . P _C・S ₂ + P _T (S ₁ − _{S 2} )=F P _T = (F − P _C・S ₂ )/(S ₁ − _{S 2} ) Here, since S ₁ , S ₂ , and F are constant , clutch pressure P _T is duty-controlled control pressure P _C ,
It will be controlled in an inversely proportional relationship. To explain this with reference to FIG. 4a, when the duty ratio of the solenoid valve 133 is 0%, no pressure is exhausted at all, and the control pressure P _C becomes the highest value, which is equal to the reducing pressure P _R of the pressure regulating valve 140. As the duty ratio gradually increases from this state and the pressure is exhausted, the control pressure P _C decreases, resulting in a characteristic as shown by the broken line. On the other hand, in relation to the control pressure P _C mentioned above, the clutch pressure P _T is zero in a region smaller than a certain duty ratio d ₁ , and increases proportionally after that duty ratio d ₁ , resulting in a characteristic as shown by the solid line. Become. Referring to FIG. 3, an electrical control system including a control unit 160 will be described. First, a 4WD switch 161 that detects switching of four-wheel drive, an accelerator switch 162 that detects opening of the accelerator, an accelerator opening switch 163 that detects engine load, and an engine rotation sensor 16.
4. Vehicle speed sensor 165, shift range switch 166 that detects each range of D, D _S , and R, sensor 167 that detects depression of the brake, clutch current sensor 1 that detects the capacity of electromagnetic powder clutch 1.
68 and an oil temperature sensor 169. The signals from each switch and sensor are input through the input interface 170 only when the 4WD switch 161 is on. In the control unit 160, the accelerator opening calculation section 171 calculates the gear ratio (pulley ratio) based on the relationship between the engine rotation speed on the gear input side and the vehicle speed on the transmission output side.
The shift pattern in the continuously variable transmission 3 is determined by the relationship between the engine speed and vehicle speed for each accelerator opening as shown in Fig. 4c, so refer to this shift pattern map. Find the accelerator opening degree. In the electromagnetic powder clutch 1, the clutch current is controlled as shown in FIG. 4b in the half-clutch state in the region D ₁ below the set vehicle speed V ₁ shown in FIG. Find the clutch capacity directly from
Therefore, the engine torque calculation section 173 calculates the engine output torque from the accelerator opening degree and the engine rotational speed by referring to the engine torque map. Oil temperature determining section 174 determines oil temperature. Further, the driving condition determining unit 175 includes a low speed determining unit 175a including starting at a set vehicle speed V ₁ or less, a low speed determining unit 175a including starting at a set vehicle speed V 1 or less;
It has a clutch direct connection determination section 175b of V ₁ or more and a brake operation determination section 175c. Low speed determination section 17
The signal from 5a is input to the area determination section 176,
The accelerator switch 162 releases the accelerator, the accelerator opening switch 163 determines low load and high load regions, and for each load, the shift range switch 166 determines the shift position. There is also a region determination unit 177 for the clutch direct connection determination unit 175b, and the set engine rotation speed shown in FIG.
Based on N ₁ as a reference, low load areas below N 1 and high load areas D ₂ and D ₃ above N 1 are determined. Here, the engine speed _N1 is set lower than the shift start point _N2 when the accelerator is fully open, to prevent the torque of the transfer clutch from fluctuating near the shift start point when the accelerator is fully open. Furthermore, the brake operation determination section 17
The signal 5c is input to the deceleration calculating section 178 to calculate the deceleration. The output signals of each of the calculation sections 172, 173, 178 and the determination sections 176, 177 are input to a hydraulic clutch torque calculation section 180, and a constant setting section 179 is sent to the calculation section 180. A ratio constant is entered,
The optimum hydraulic clutch torque is set in each region. An example of the sharing ratio in this case is shown in the table below. Therefore, under low-speed running conditions with a half-clutch engaged, the hydraulic clutch torque is calculated by using the transmission torque of the electromagnetic powder clutch 1 as a base and multiplying it by each sharing ratio in the table. Under these driving conditions, there is the greatest possibility that the vehicle will undergo a large steering turn, and at the same time, the vehicle will also attempt to escape from a rough road. And the large steering change is caused by the normal driving D under low load driving condition.
Or it is done in R range, and when driving on rough roads, it is under high load.
Performed by D _s or R range. Therefore, when opening the accelerator, the sharing ratio of the D, D _s , and R ranges is fixed at the minimum, and when the opening is small, the sharing ratio of the D and R ranges is as small as 20%, and when the opening is large, The sharing ratio of D _s and R range is
It is set to a large value of 60%, and the torque capacity is corrected accordingly. Furthermore, in the continuously variable transmission 3, a shift pattern is set as shown in _FIG .
The degree to which a low speed gear with a large gear ratio is used is large, and the degree to which a high speed gear is used increases as the vehicle speed increases. Therefore, the engine torque is
It is large at low vehicle speeds and decreases as the vehicle speed increases. From this, in the clutch direct connection region where the set vehicle speed is V ₁ or more, and in the region D ₂ where the set engine speed is less than or equal to N ₁ , find the engine torque for each of the above vehicle speeds, and use that torque as a base and multiply it by the smaller sharing ratio in the table. Then, the hydraulic clutch torque is calculated based on the relationship between vehicle speed and engine speed. On the other hand, in the high load side region _D3 where the set engine speed N1 is higher than the set engine speed _N1 , there is almost no possibility of large steering. Therefore, the hydraulic clutch torque is calculated in relation to the vehicle speed by multiplying the large sharing ratio in the table based on the engine full-open performance torque with the maximum engine torque. Under these driving conditions, most of the gear changes have started in the continuously variable transmission 3, especially in the forward gear D.
Since there is no need to make a difference between the and D _s ranges,
In all ranges, the sharing ratio is set as low as 30% in area _D1 , and as large as 50% in area _D3 . Furthermore, if the deceleration is large during braking, the torque selected from any one of the above based on the driving conditions at the time of braking is used as a base, and the calculation is performed in the same manner as described above.

[Table] The duty ratio setting section 181 is
The duty ratio is set according to the torque calculated at 0, and the signal of this duty ratio is
It is input to the solenoid valve 133 via the drive section 182. Furthermore, when the oil temperature is lower than the set value _t1 , the duty ratio is corrected by a signal from the correction coefficient setting section 183. Next, the operation of the hydraulic control device configured as described above will be explained. When the vehicle starts running, when the engine speed increases due to depressing the accelerator, the electromagnetic powder clutch 1
is engaged by the clutch current. Therefore, by shifting the forward/reverse switching device 2 to the forward gear, engine power is input to the main pulley 36 via the main shaft 20. At the start of running, the pressure in the main pulley servo chamber 38b is exhausted by the gear ratio control valve 90 of the hydraulic control system, and the drive belt 34 is wound furthest around the main pulley 36 to become the lowest gear with the largest gear ratio, after which the engine rotates. As the pitot pressure increases according to the number of pulleys, line pressure is also introduced into the main pulley servo chamber 38b, thereby increasing the belt winding diameter while narrowing the pulley interval of the main pulley 36. In this way, the engine speed is continuously variable to keep the engine speed constant, and this shifting power is taken out from the main pulley 36 to the subshaft 35 via the drive belt 34 and sub pulley 37, and further transmitted to the front differential device 4. . By the way, when the 4WD switch 161 is off, the control unit 160 outputs a signal with a duty ratio of around 0%, the solenoid valve 133 is completely closed, and the control pressure P _C is maximized. Therefore, the clutch pressure P _T becomes zero due to the transfer control valve 150 and the hydraulic clutch 52 is released, thereby transmitting the power of the front differential device 4 only to the front wheels and driving in FF two-wheel drive. becomes. Next, the operation when the 4WD switch 161 is turned on will be explained using the flowchart in FIG. 5 and the characteristic diagrams in FIGS. 6 to 8. First, by turning on the 4WD switch 161,
An input interface 170 allows signals from each switch and sensor to be input to and controlled by the control unit 160. Therefore, under low-speed running conditions including starting at a set vehicle speed _V1 or lower, the clutch capacity of the electromagnetic powder clutch 1 is calculated in the clutch torque calculating section 172 using the clutch current. If there is no braking, the region is determined by the region determining section 176 based on the signal from the low speed determining section 175a of the driving condition determining section 175, and if the accelerator is released, in which the accelerator switch 162 is turned off, the hydraulic clutch torque is applied in all ranges. The calculation unit 180 selects the smallest fixed share ratio of 10%. As a result, the hydraulic clutch torque calculation section 1
At step 80, the torque capacity of the hydraulic clutch 52 is calculated based on the curve _l1 of 10% of the maximum value when the engine is fully opened in FIG. 6, and a duty signal based on this is input to the solenoid valve 133. Therefore, the transfer control valve 150 supplies oil to the hydraulic clutch 52 and slightly engages it, and 10% of the transmitted torque of the electromagnetic powder clutch 1 in the half-clutch state is also transmitted to the rear wheel by the transfer device 5. This results in four-wheel drive driving. In this driving condition, the hydraulic clutch is set to 5 in all ranges.
Since the torque capacity of the hydraulic clutch 52 is very small and it is easy to slip, the torsional torque of the drive system in the case of a large turning is sufficiently absorbed by the slip of the hydraulic clutch 52, and the occurrence of a braking phenomenon is avoided. Next, in the case of a low load where the accelerator switch 162 is turned on but the accelerator opening switch 163 is turned off, for example, if the shift is to the D or R range, the calculation section 180 selects a sharing ratio of 20%. Therefore, the torque capacity of the hydraulic clutch 52 becomes a little larger as shown by curve _l2 in FIG. 6, and gradually increases proportionally at this rate as the capacity of the electromagnetic powder clutch 1 increases. As a result, the performance of the four-wheel drive is improved, and since the vehicle is still susceptible to slipping, it is possible to avoid the braking phenomenon. On the other hand, in the case of the shift position in the D _s range, the sharing ratio is 30% and the torque capacity is larger than the above, as shown by curve l ₂ ' in Figure 6, and the four-wheel drive performance is strengthened, making it easier to climb up and downhill. Improves road running performance. On the other hand, when the accelerator opening switch 163 is also turned on and the load is high and the D range is selected, the calculation section 180 selects a sharing ratio of 50%, and the torque capacity of the hydraulic clutch 52 is set as shown by curve _l3 in FIG. Ru. Therefore, under this condition, the hydraulic clutch 52 directly connects the front and rear wheels so as to equalize their driving forces, and the four-wheel drive, for example, starts and accelerates sufficiently. Also, D _s
Or, if the shift position is in the R range, the sharing ratio
At 60%, the torque capacity is maximized as shown by the curve l ₃ ' in Figure 6, and the maximum amount of power that can be gripped is transmitted to the rear wheels, making it possible to effectively escape from rough roads. It will be done. Under clutch-direct driving conditions where the vehicle speed is higher than the set vehicle speed _V1 , the engine torque calculation section 173 calculates the engine torque for each vehicle speed, as shown by curves _l1 , _l2 , and _l3 in FIG. If there is no braking, the region is determined by the region determining section 177 based on the signal from the clutch direct connection determining section 175b of the driving condition determining section 175, and the region is determined in the region where the set engine rotation speed is N ₁ or less.
In the case of _D2 , a sharing ratio of 30% is selected by the hydraulic clutch torque calculation section 180 at all shift positions in all ranges. As a result, the calculation section 18
0, the clutch capacity of the hydraulic clutch 52 is calculated based on the above engine torque curves l ₁ , l ₂ , l ₃ and the 30% curves l ₁ ′, l ₂ ′, l ₃ ′. The duty signal based on the solenoid valve 133
Enter. Therefore, the hydraulic clutch 52 is supplied with oil by the transfer control valve 150 and engaged in a slippable state, and 30% of the engine torque is also transmitted to the rear wheels by the transfer device 5.
It will be driven by wheel drive. In this running state, since slip is possible in all ranges, the torsional torque of the drive system when steering is absorbed by the slip of the hydraulic clutch 52, thereby avoiding the occurrence of a braking phenomenon. In addition, the torque capacity of the hydraulic clutch 52 is controlled in accordance with the change in engine torque according to the gear ratio at each vehicle speed, so the engine torque is always maintained.
30% of the power is transmitted to the rear wheels to ensure constant four-wheel drive performance. On the other hand, in the case of region D ₃ where the set engine speed is N ₁ or more, the calculation unit 180 similarly selects a sharing ratio of 50% in all ranges, and the curve shown in FIG.
l Set the full engine performance torque of ₄ to the base torque. Therefore, the torque capacity of the hydraulic clutch 52 is set as a 50% curve _l4 ' with respect to this torque, and is directly connected to equalize the driving force of the front and rear wheels, maximizing the performance of the four-wheel drive. Exercising to the limit. Although one embodiment of the present invention has been described above, clutches and transmissions other than the embodiment can be similarly applied to RR base and vertical transaxle types. In addition, the input means of the electric control system, the configuration of the control unit, and the sharing ratio between the front and rear wheels are as follows:
The present invention is not limited to the above embodiment, and can also be applied to a full-time type without a 4WD switch. Furthermore, even under driving conditions where the clutch is directly connected, the torque capacity may be made variable for each shift position.

【Effect of the invention】

As described above, according to the present invention, in a hydraulic control system for a four-wheel drive system having a hydraulic clutch in the transfer device, control is divided into low speed ranges including half-clutch start and clutch direct coupling ranges. Therefore, it is possible to accurately deal with braking phenomena that occur when making a large turn in a low speed range and with full-scale four-wheel drive with direct clutch connection. In the low speed range, control is performed separately for each condition: open accelerator, low load, and high load, so optimal control is performed to avoid braking phenomena and obtain four-wheel drive performance as required. In addition, the clutch capacity actually transmitted in the half-clutch state is calculated, and the torque capacity of the hydraulic clutch is set based on this.
Its torque capacity is optimized. Furthermore, since the torque capacity is set low when the accelerator is released and the vehicle is in low load driving conditions when the D and R ranges are in the normal driving position, the braking phenomenon that occurs when making a large steering turn is preferentially avoided. . On the other hand, the torque capacity is set high in the D _s range,
Even in the R range, the shift position is set high under high loads, so the vehicle's performance in this range and on rough roads will be improved by making full use of the performance of the four-wheel drive.

[Brief explanation of the drawing]

FIG. 1 is a sectional view showing an example of a four-wheel drive device to which the present invention is applied, FIG. 2 is a circuit diagram showing an embodiment of a hydraulic control system, FIG. 3 is a circuit diagram of an electric control system, and FIG.
Figure a is a characteristic diagram of a solenoid valve, b is a current characteristic diagram of an electromagnetic powder clutch, c is a speed change pattern of a continuously variable transmission, and Figure 5 is a flow chart explaining the operation.
6 to 8 are torque characteristic diagrams of the transfer clutch according to the present invention. 1... Electromagnetic powder clutch, 2... Forward/forward switching device, 3... Continuously variable transmission, 4... Front differential device, 5... Transfer device, 52... Hydraulic clutch, 133... Solenoid valve, 140...Pressure regulating valve, 150...Transfer control valve, 160...
Control unit, 175... Running condition determination section, 17
5a...Low speed determination section, 175b...Clutch direct connection determination section, 176...Region determination section, 180...Hydraulic clutch torque calculation section.

Claims (1)

[Scope of Claims] 1. A four-wheel drive system in which power is directly transmitted to one of the front and rear wheels via an engine side clutch and a transmission, and power is transmitted to the other wheel through a transfer device having a hydraulic clutch, A means for detecting a state in which the engine-side clutch is in a half-clutch state and a direct connection state;
means for setting the torque capacity of the hydraulic clutch in response to a signal detecting driving conditions such as accelerator opening, vehicle speed, etc.; and means for setting the torque capacity of the hydraulic clutch in response to a signal detecting driving conditions such as accelerator opening, vehicle speed, etc.; A hydraulic control device for a four-wheel drive system, comprising: a signal for detecting a running state of the hydraulic clutch; and means for correcting the torque capacity of the hydraulic clutch.