JPH0543526B2 - - Google Patents

Description

[Detailed description of the invention]

[Industrial Application Field] The present invention relates to a four-wheel drive system configured to directly transmit driving force from a belt-type continuously variable transmission to the front wheels and to the rear wheels via a hydraulic clutch of a transfer device. The present invention relates to a hydraulic clutch control device for a four-wheel drive device that controls the clutch torque of the hydraulic clutch. [Prior art] The four-wheel drive device in the drive system of a four-wheel drive vehicle is
Generally front engine/front drive (FF) or rear engine/rear drive (RR)
Based on this, the vehicle is designed to run in four-wheel drive by transmitting power to the other front and rear wheels via a transfer clutch installed in the transfer device. In this case, it has been proposed to use a hydraulic clutch as the transfer clutch (Japanese Patent Application Laid-Open No. 58-56921, Japanese Patent Application Laid-Open No. 61-1999).
(See Publication No. 157438). Here, in a four-wheel drive system that does not have a center differential and uses four-wheel drive by engagement of a transfer clutch, when the four-wheel drive makes a large turn on a road surface such as a paved road with a large tire grip force, The difference in rotation between the front and rear wheels creates excessive torsional torque in the drive system, which is known to cause problems such as tight corner braking and engine stalling, as well as increased steering force. . Therefore, in the prior art described in Japanese Patent Application Laid-open No. 58-56921, a control circuit adjusts the hydraulic pressure according to the amount of depression of the accelerator pedal, the presence or absence of slip in the front and rear wheels, and the on/off state of the manual changeover switch for two-wheel drive. The clutch torque of the clutch is controlled in four stages. For example, when a slip occurs, the clutch torque is controlled to the maximum to maximize the four-wheel drive function, and when the vehicle is making a large turn, the clutch torque is controlled to the maximum. Torque is controlled to a small level to avoid tight corner braking. In addition, we focused on the fact that the tight corner braking phenomenon occurs when the steering angle is large when driving in four-wheel drive.
Prior art is also known in which the tight corner braking phenomenon is avoided by switching to wheel drive (see Japanese Patent Laid-Open No. 58-61025 and Japanese Patent Laid-Open No. 59-206228). [Problem to be Solved by the Invention] By the way, in the prior art described above, which automatically switches to two-wheel drive when the steering angle becomes large, the tight corner braking phenomenon can be avoided; In this case, the performance as a four-wheel drive will be completely lost. Therefore, there is a problem in that the performance peculiar to a four-wheel drive vehicle cannot be sufficiently exhibited particularly in driving conditions such as on a road surface where slips are likely to occur or when climbing a slope. In this regard, in the prior art described in Japanese Patent Application Laid-open No. 58-56921, when the amount of depression of the accelerator pedal is small and the engine torque is small, the clutch torque is controlled to the minimum, thereby enabling the vehicle to make large turns. At times, the clutch torque is controlled to be small in response to a decrease in engine torque, and the hydraulic clutch is slipped to avoid tight corner braking, so that the four-wheel drive performance can also be maintained. However, when the clutch torque is simply controlled to be small when the engine torque is small, the following problems arise. In other words, the vehicle is equipped with a belt-type continuously variable transmission that automatically shifts up as the vehicle speed increases to the rotational speed of the wheels, and the driving force from the belt-type continuously variable transmission is not transferred to the front wheels. In a four-wheel drive system that uses direct transmission, if the front wheels lift up and spin under special driving conditions such as rough roads, the belt-type continuously variable transmission will automatically shift up and reduce engine torque. The clutch torque of the hydraulic clutch of the transfer device decreases, resulting in insufficient torque to be transmitted to the rear wheels, and it may not be possible to escape from such a situation where the front wheels are stuck, in which case the four-wheel drive function can be fully utilized. The present invention was made in view of the problems of the prior art, and it transmits the driving force from the belt type continuously variable transmission directly to the front wheels, and transmits it to the rear wheels via a hydraulic clutch of a transfer device. 4 configurations to communicate
This four-wheel drive system is based on a wheel drive system, and can avoid tight corner braking while maintaining four-wheel drive performance when turning, and can drive the rear wheels with maximum clutch torque when the front wheels are spinning. The object of the present invention is to provide a hydraulic clutch control device. [Means for Solving the Problems] In order to achieve the above object, the present invention directly transmits the driving force from a belt type continuously variable transmission to the front wheels, and transmits the driving force to the rear wheels via a hydraulic clutch of a transfer device. In a four-wheel drive device configured to transmit data, a control pressure generated by exhaust pressure control of a duty solenoid valve acts on the hydraulic circuit of the hydraulic clutch, and the hydraulic pressure is supplied from the hydraulic source to the hydraulic clutch in accordance with the control pressure. The control unit is equipped with a transfer control valve that continuously increases and decreases the clutch oil pressure that is applied to the clutch, and is a control unit that drives the duty solenoid valve based on a duty ratio signal, so that the clutch oil pressure is continuously increased or decreased in accordance with increases and decreases in engine torque. The control unit is provided with a control unit that sets a duty ratio signal to increase and decrease the control pressure, and to control the control pressure so that the clutch oil pressure reaches a maximum when the front wheels are idling. [Operation] In the present invention employing such means,
When the vehicle makes a large turn, the engine torque decreases, and the control unit outputs a predetermined duty ratio signal to the duty solenoid valve to drive the clutch hydraulic pressure to decrease in accordance with the engine torque. do. Therefore, the control pressure applied to the transfer control valve is continuously controlled to a predetermined value by the duty solenoid valve, so that the clutch oil pressure supplied to the hydraulic clutch is continuously reduced in accordance with the engine torque. The clutch torque of the hydraulic clutch gradually decreases accordingly. Therefore, when the vehicle makes a large turning turn, the hydraulic clutch slips, and the tight corner braking phenomenon is avoided while maintaining the four-wheel drive performance. Additionally, when the front wheels spin when the vehicle is driving on rough roads, the belt-type continuously variable transmission automatically shifts up and engine torque decreases, but in this case, the control unit adjusts the clutch pressure to maximize. A predetermined duty ratio signal is output to the duty solenoid valve to drive it. Therefore, the control pressure acting on the transfer control valve is continuously controlled to a predetermined value by the duty solenoid valve, thereby increasing the clutch oil pressure supplied to the hydraulic clutch to the maximum value. Accordingly, the clutch torque of the hydraulic clutch increases to its maximum value. Therefore, when the front wheels are idling, the rear wheels are driven with the maximum clutch torque, and the four-wheel drive function is fully utilized. Here, since the clutch torque of the hydraulic clutch is continuously controlled to increase or decrease, the driving force can be distributed to the rear wheels without any shock. [Example] Hereinafter, an example of the present invention will be specifically described 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 supply brush 19 is in sliding contact with the clutch current circuit, which is connected from the slip ring 18 to the coil 15 through the inside of the driven member 14. 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 switches the clutch 1
The input shaft 13 is provided between the input shaft 13 and the main shaft 20 coaxially arranged therewith. That is, the input shaft 13
The reverse drive gear 21 also serves as the forward engaged side.
is formed, and a gear 22 on the engaged side for reverse movement is rotatably fitted to the main shaft 20, 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, 24, which are always in mesh here,
26 and 22 are connected to the driven member 14 having the coil 15 of the clutch 1, and the switching mechanism 27 is connected to the hub 28 of the main shaft 20 in response to the fact that the inertia mass of this part is relatively large when the clutch is disengaged. A spline-fitting sleeve 29 is configured to mesh with each gear 21, 22 via synchronizer mechanisms 30, 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 the forward movement state of the D or D _S 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 the main shaft 20 and sub shaft 35, respectively, and an endless drive belt 34 is stretched between both pulleys 36, 37. Both the main pulley 36 and the sub pulley 37 are 2
One pulley half 36a, 37
In contrast to a, the other pulley halves 36b, 37b are 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. Furthermore, the movable pulley half 37b of the sub pulley 37
, spring 4 is installed in the direction to narrow the pulley spacing.
0 is energized. 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 main pulley of the drive belt 34 36, the pulley ratio in the sub-pulley 37 is continuously changed, and the continuously variable power is output to the sub-shaft 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 rotational speed of the subshaft 35 is high, the front differential device 4 has one set of pulleys for the subshaft 35. 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. It communicates with the oil sump side of oil passages 74 and 75. 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 pitot pressure of the oil passage 77 is introduced to the port 81a of the spool 82 on the opposite side of the spring 84, and the pump oil pressure of the oil passage 71 is introduced to the port 81b. The port 81c also has an oil passage 71.
An oil passage 87 to the gear ratio control valve 90 is in communication with the gear ratio control valve 90. A port 81f on the spring 84 side of this port 81c, and a port 81e provided between ports 81a and 81b to prevent leakage of pump hydraulic pressure from affecting pitot pressure, leaked oil is drained and oil is removed. It is led to the reservoir 70. Further, the ports 81c and 81d are communicated with each other at the chamfer portion of the land 82a of the spool 82 to regulate the pressure. 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. Therefore, the sensor shoe 85 moves to the right in the drawing, and controls the line pressure to be lowered by lowering the load on the spring 84, thereby always 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 94d 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 95.
The amount of protrusion of the plunger 95 relative to the spool 92, that is, the load of the adjustment spring 97, is changed by the line pressure. Further, the operating plunger 93 receives the rod 1 from the cam 100 which acts as a lift depending on the accelerator opening degree.
The rod 101 is separated from the rod 101 and connected via a weak spring 102, and has a stopper 103 so as to move in 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 in a state where there is 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 D _S at both ends are in the concave part 115a.
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 part 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 whose intermediate portion is supported by a pin 127 is provided between the sensor shaft 85 that changes according to the gear ratio and the lever 125 at the tip of the piston of the actuator 120. . This correction lever 128
has one end engaged with the piston lever 125 only when the actuator 120 is pushed in, and
In this state, when the sensor shoe 85 reaches a position of a predetermined gear ratio by shifting to a lower gear with a larger gear ratio, the other end of the lever 128 engages with the sensor shoe 85. 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 is connected to the transfer control valve 1.
Duty solenoid valve 13 via the control side of 50
3, and a clutch pressure oil passage 134 from the transfer 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 144, 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 duty solenoid valve 133 opens the drain port 133a based on the duty signal from the control unit 160, thereby controlling the line pressure regulating valve 1.
40, the reducing pressure P _R is discharge-controlled to generate a control pressure P _C , which is transferred to the transfer control valve 1.
Acts on 50. 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, the clutch pressure P _T is lowered and the control pressure P _C is lowered.
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 duty solenoid valve 133 is 0%, no pressure is exhausted at all, and the control pressure P _C reaches the highest value, which is equal to the reducing pressure P _R of the pressure regulating valve 140.
From this state, as the duty ratio gradually increases and the pressure is exhausted, the control pressure P _C decreases to a characteristic as shown by the broken line. On the other hand, in relation to the above control pressure P _C , 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 register switch 166 that detects each range of D, D _S , and R, sensor 167 that detects depression of the brake, clutch current sensor 168 that detects the capacity of electromagnetic powder clutch 1, and oil temperature sensor It has 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. For electromagnetic powder clutch 1, set vehicle speed as shown in Fig. 4c.
In the region _D1 below _V1 , the clutch current is controlled as shown in Figure 4b in the half-clutch state, so
The clutch torque calculating section 172 directly calculates the clutch capacity from this clutch current. 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. The driving condition determining section 175 also includes a low speed determining section 175a including a start at a set vehicle speed _V1 or lower, a clutch direct connection determining section 175b at a set vehicle speed _V1 or higher, and a brake operation determining section 175c. The signal from the low speed determining section 175a is input to the region determining section 176, where the accelerator release is performed by the accelerator switch 162, and the low load and high load regions are determined by the accelerator opening switch 163.For each load, the shift register switch 166 further determines the low load and high load regions. The position is determined. There is also a region determining section 177 for the clutch direct connection determining section 175b, which determines low load areas below the set engine rotation speed _N1 shown in FIG. 4c, and high load areas _D2 and _D3 above it. It will be judged. 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. Further, the signal from the brake operation determination section 175c is transmitted to the deceleration calculation section 17.
8 to calculate the deceleration based on the change in vehicle speed detected by the vehicle speed sensor 165. 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. Also, under driving conditions where the clutch is directly connected, the hydraulic clutch torque is calculated in the same way using the engine torque as a base. Furthermore, when the deceleration of braking is large, the torque selected from any one of the above based on the driving conditions at the time of the operation is used as a base, and the calculation is performed in the same manner as described above.

〔Effect of the invention〕

As explained above, according to the present invention, when the vehicle makes a turn with a large steering turn, the engine torque decreases, and the control unit sends a predetermined duty ratio signal to reduce the clutch oil pressure in accordance with the engine torque. It outputs to the solenoid valve and drives it. Therefore, the control pressure applied to the transfer control valve is continuously controlled to a predetermined value by the duty solenoid valve, so that the clutch oil pressure supplied to the hydraulic clutch is continuously reduced in accordance with the engine torque. The clutch torque of the hydraulic clutch gradually decreases accordingly. Therefore, when the vehicle makes a large turning turn, the hydraulic clutch slips, making it possible to avoid tight corner braking while maintaining four-wheel drive performance. Additionally, when the front wheels spin when the vehicle is driving on rough roads, the belt-type continuously variable transmission automatically shifts up and engine torque decreases. A predetermined duty ratio signal is output to the duty solenoid valve to drive it. Therefore, the control pressure acting on the transfer control valve is continuously controlled to a predetermined value by the duty solenoid valve, thereby increasing the clutch oil pressure supplied to the hydraulic clutch to the maximum value. Accordingly, the clutch torque of the hydraulic clutch increases to its maximum value. Therefore, when the front wheels are idling, the rear wheels can be driven with the maximum clutch torque, and the four-wheel drive function can be fully demonstrated. Here, since the clutch torque of the hydraulic clutch is continuously controlled to increase or decrease, the driving force can be distributed to the rear wheels without any shock.

[Brief explanation of the drawing]

Fig. 1 is a sectional view of a four-wheel drive device to which an embodiment of the present invention is applied, Fig. 2 is a circuit diagram showing a hydraulic control system in an embodiment, and Fig. 3 is a circuit diagram showing an electric control system in an embodiment. Circuit diagram, Fig. 4a is a characteristic diagram of a duty solenoid valve in one embodiment, b is a current characteristic diagram of an electromagnetic powder clutch, c is a diagram showing a speed change pattern of a continuously variable transmission, and Fig. 5 is a diagram showing the characteristics of a duty solenoid valve in one embodiment. FIG. 7 is a hydraulic circuit diagram showing the main parts of another embodiment of the 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... Duty Solenoid valve, 14
0...Pressure regulating valve, 150...Transfer control valve,
160...Control unit, 165...Vehicle speed sensor, 181...Duty ratio setting section, 184...
Propeller shaft rotation sensor, 185...Front wheel slip judgment unit, 186...4WD full lock switch, 19
0... Actuator.

Claims (1)

[Scope of Claims] 1. A four-wheel drive system configured to directly transmit driving force from a belt-type continuously variable transmission to the front wheels and transmit it to the rear wheels via a hydraulic clutch of a transfer device, comprising: A control pressure generated by exhaust pressure control of the duty solenoid valve acts on the hydraulic circuit of the clutch, and there is a transfer valve that continuously increases or decreases the clutch hydraulic pressure supplied from the hydraulic source to the hydraulic clutch according to the control pressure. As a control unit that drives the duty solenoid valve based on a duty ratio signal, the clutch oil pressure continuously increases and decreases in accordance with increases and decreases in engine torque, and when the front wheels are idling, the clutch oil pressure increases and decreases. A hydraulic clutch control device for a four-wheel drive system, characterized in that a control unit is provided for setting a duty ratio signal to control the control pressure to a maximum exhaust pressure. 2. A first range characterized in that the control unit is configured to set a duty ratio signal based on an ON signal of a 4WD full lock switch to perform exhaust pressure control on the control pressure so that the clutch oil pressure is maximized. Hydraulic clutch control device for a four-wheel drive device as described in 2. 3 The above transfer control valve has a transfer control valve that operates in response to the ON operation of the 4WD full lock switch and maximizes the clutch oil pressure regardless of the control pressure controlled by the duty solenoid valve. Claim 1 is characterized in that an actuator for forcibly actuating is attached.
Hydraulic clutch control device for a four-wheel drive device as described in 2.