JPS6374736A - Speed change ratio control method for belt type continuously variable transmission for automobile - Google Patents

Abstract

PURPOSE:To prevent application of excessive tension to a transmission belt at the time of restarting of a car after quick branking by controlling a speed change direction changeover valve, when the car is going to stop, so that the speed change ratio varies quickly toward the maximum. CONSTITUTION:This belt-type stepless speed changer 12 includes a speed change direction changeover valve device, not illustrated, to switch over the speed change direction through control of supply or exhaust of the working oil to/from a hydraulic cylinder 50, which is to alter the V-groove width of a variable pulley 36 mounted on the input side. A speed change velocity changeover valve device is also included, not illustrated, which is to switch over the speed change velocity through control of the rate of supply or exhaust flow of the working oil to/from said hydraulic cylinder 50. In the state the car is going to stop, the speed change velocity changeover valve is so controlled that the speed change ratio varies quickly toward the maximum. If here the car speed is lower than the specified value, the speed change velocity changeover valve is switched over so that supply/exhaust of the working oil to/from the hydraulic cylinder 50 is suppressed.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of the Invention The present invention relates to a method for controlling a gear ratio of a continuously variable transmission for a vehicle.

Conventional technology An input shaft and an output shaft, an input-side variable pulley and an output-side variable pulley with variable effective diameters provided on the input shaft and output shaft, respectively, and V-grooves of the input-side variable pulley and output-side variable pulley. The gear shifting direction is controlled by supplying hydraulic oil to the wrapped transmission belt and a hydraulic cylinder that changes the V-groove width of the input variable pulley to change the gear ratio, or by discharging the hydraulic oil from the hydraulic cylinder. 2. Description of the Related Art A belt-type continuously variable transmission for a vehicle is known, which includes a shift direction switching valve device for switching, and a shift speed switching valve device for switching the transmission speed by switching the supply flow rate or discharge flow rate of hydraulic oil from the hydraulic cylinder. In such a vehicle belt-type continuously variable transmission, as described in Japanese Patent Application No. 60-70614, the required output is determined so as to occur on the minimum fuel consumption rate curve of the engine. The speed change direction switching valve device is switched so that the target rotational speed of the continuously variable transmission input shaft matches the actual rotational speed, thereby increasing the fuel consumption efficiency of the vehicle. When the vehicle comes to a stop, the speed change direction switching valve device is controlled so that the groove width of the variable pulley on the input side is large and the width of the groove on the variable pulley on the output shaft side is small. Generally, the gear ratio is changed to its maximum value in preparation for restarting the vehicle.

Problems to be Solved by the Invention However, in such conventional control methods, when the vehicle is suddenly braked, the rotation of the continuously variable transmission may stop before the gear ratio of the continuously variable transmission reaches its maximum value. Then, when the vehicle restarts after that, since driving force is required, the continuously variable transmission is rotated again and the gear ratio of the continuously variable transmission is changed toward its maximum value. In such a case, there is an inconvenience that the transmission belt slips. In other words, when a vehicle starts, a relatively large amount of power is transmitted through the transmission belt, and in such a case, the tension of the transmission belt corresponding to the transmitted torque is required, but the continuously variable transmission is not effective. Since the width v1 of the input variable pulley is expanded in order to maximize the speed ratio, the tension in the transmission belt becomes insufficient and slippage occurs.

Means for Solving the Problems The present invention has been made against the background of the above circumstances.
The gist is the input shaft and output shaft, the input-side variable pulley and output-side variable pulley with variable effective diameters provided on the input shaft and output shaft, respectively, and the input-side variable pulley and output-side variable pulley. Supply hydraulic oil to the transmission belt wrapped around the groove of the pulley and the hydraulic cylinder that changes the groove width of the input variable pulley to change the gear ratio, or discharge hydraulic oil from the hydraulic cylinder. In a belt-type continuously variable transmission for a vehicle, the belt-type continuously variable transmission has a shift direction switching valve device that switches the shift direction by changing the shift direction, and a shift speed switching valve device that switches the shift speed by switching the supply flow rate or discharge flow rate of hydraulic oil of the hydraulic cylinder.
A transmission ratio control method in which the transmission direction switching valve device is controlled so that the transmission ratio quickly changes toward a maximum value when the vehicle stops, and when the vehicle speed is lower than a predetermined value. Another object of the present invention is to switch the transmission speed switching valve so as to suppress the discharge of hydraulic oil from the hydraulic cylinder for changing the transmission ratio or the supply of hydraulic oil to the hydraulic cylinder.

Operation and Effect of the Invention By doing this, when the vehicle speed is lower than a predetermined value,
Since the shift speed switching valve is switched to suppress the discharge of hydraulic oil from the hydraulic cylinder for changing the gear ratio or the supply of hydraulic oil to the hydraulic cylinder, when the vehicle speed is low, the gear change of the continuously variable transmission is Ratio changes are suppressed. Therefore, when the vehicle restarts after sudden braking, even if the gear ratio of the continuously variable transmission attempts to change toward its maximum value, that change is suppressed, eliminating the drop in tension in the transmission belt. This effectively prevents the transmission belt from slipping.

EXAMPLE Hereinafter, an application example of the present invention will be explained in detail based on the drawings.

In FIG. 2, the power of the engine 8 is connected to a drive shaft 20 via a fluid coupling 10, a belt type continuously variable transmission (hereinafter referred to as CVT) 12, an auxiliary transmission 14, an intermediate gear device 16, and a differential device 18. The signal is transmitted to the driven wheels (not shown).

The fluid coupling IO connects a pump 24 connected to the crankshaft 22 of the engine 8, a turbine 28 fixed to the input shaft 26 of the CVT 12 and rotated by oil from the pump 24, and a damper 30 connected to the input shaft 26. A fixed lock-up clutch 32 is provided. The lock-up clutch 32 is activated when, for example, the vehicle speed, the engine rotational speed, or the rotational speed of the turbine 28 exceeds a predetermined value, and connects the crankshaft 22 and the input shaft 26 directly.

The CVT 12 includes an input-side variable pulley 36 and an output-side variable pulley 38 provided on the input shaft 26 and the output shaft 34, respectively, and a transmission belt 40 wound between the variable pulleys 36 and 38. The input side variable boob IJ36 and the output side variable pulley 38 are connected to the input shaft 26
and fixed rotating bodies 42 and 4 fixed to the output shaft 34.
4, and movable rotary bodies 46 and 48 provided on the input shaft 26 and the output shaft 34, respectively, so as to be movable in the axial direction but not relatively rotatable around the axes. By moving the transmission belt 40, the width of the groove, that is, the diameter (effective diameter) of the transmission belt 40 is changed, and the gear ratio r (=
Rotational speed N i of input shaft 26 /rotational speed N of output shaft 34 . ut) is now being changed. The hydraulic cylinder 50 is operated exclusively to change the gear ratio γ, and the hydraulic cylinder 52 is operated exclusively to obtain the minimum clamping force within a range where the transmission belt 40 does not slip. The oil pump 54 constitutes a hydraulic pressure source of a hydraulic control device, which will be described later, and is connected to the crankshaft 22 via a not-illustrated connecting shaft passing vertically through the input shaft 26, and is constantly driven to rotate by the engine.

The sub-transmission 14 is provided coaxially with the output shaft 34 of the CVT 12, and includes a Ravigneau type compound planetary gear device. This planetary gear device includes a pair of first sun gear 56 and second sun gear 58, a first planet gear 60 that meshes with the first sun gear 56, and a pair of first and second sun gears.
A second planetary gear 62 that meshes with the sun gear 58 , a ring gear 64 that meshes with the first planetary gear 60 , and a first planetary gear 6
0 and a carrier 66 that rotatably supports the second planetary gear 62. The second sun gear 58 is fixed to a shaft 68 that is integrally connected to the output shaft 34, and the carrier 66 is fixed to an output gear 70. The high speed clutch 72 controls the engagement between the shaft 68 and the first sun gear 56 , the low speed brake 74 controls the engagement of the first sun gear 56 with the housing, and the reverse brake 76 controls the engagement between the shaft 68 and the first sun gear 56 . Controls the attachment of the housing to the housing. Third
The figure shows the operating state of each frictional engagement element of the sub-transmission 14 and the reduction ratio in each range. In Figure 3,
The ○ mark indicates the binding state, the × mark indicates the release state, and ρl and ρ
2 is a gear ratio defined from the following equation.

ρ1=ZsI/Z.

ρ2=Z, □/Z.

However, Z51 is the number of teeth of the first sun gear 56, and Z5□ is the number of teeth of the second sun gear 56.
The number of teeth of sun gear 58 and Zl are the number of teeth of ring gear 64.

Therefore, in the low gears in the L and D ranges, the low gear brake 74 is operated and the first sun gear 5G
is fixed, power is transmitted at the reduction ratio (1+ρ1/ρ2), but in the high gears of L and D ranges, the entire planetary gear system rotates as a unit due to the operation of the high gear clutch 72. As a result, power is transmitted at a reduction ratio of 1. Furthermore, in the R range, the ring gear 64 is fixed to the housing by the operation of the reverse brake 76, so power is transmitted through reverse rotation at the gear ratio (1-1/ρ2).

Returning to FIG. 2, the output gear 70 of the auxiliary transmission 14 is connected to the differential device 18 via the intermediate gear device 16, and the power of the engine 8 is transmitted to the left and right drive shafts 2 through the differential device 18.
0 and then transmitted to the left and right drive wheels.

4, 5, and 6 show a hydraulic control circuit for controlling the vehicle power transmission device shown in FIG. 2. FIG.

The oil pump 54 pumps the hydraulic oil returned into the oil tank (not shown) to the suction line oil passage 82 through the strainer 80. The throttle valve opening detection pulp 84 generates a throttle pressure P corresponding to the throttle valve opening θ on its output port 86. The spool 88 of the throttle valve 84 receives an acting force from a throttle cam 90 that rotates together with the throttle valve (not shown), which increases as the throttle valve opening θ increases, and a throttle pressure P as a feedback pressure from the control port 92 in opposite directions. The opening/closing of the line oil passage 82 and the output port 86 is controlled accordingly. The manual valve 94 is in the low position of the shift lever)
, D (drive), N neutral), R (reverse)
, and axially positioned in connection with P (parking) range operation, the first line pressure P of the line oil passage 82
Connect IL to port 96 when in R range, to port 98 when in L range, to ports 98 and 100 when in D range,
Lead each. The relief valve 102 has a function as a safety valve that releases oil from the line oil passage 82 when the first line pressure Pj21 of the line oil passage 82 exceeds a predetermined value.

The secondary hydraulic oil passage 104 is connected to the line pressure oil passage 82 via an orifice 106 and a port 110 through which excess oil of the primary regulator valve 108 is discharged.
The control room 1 is connected to the secondary hydraulic fluid line 104 via
16, controls the connection between the secondary hydraulic oil passage 104 and the port 120 in relation to the oil pressure in the control chamber 116 and the load of the spring 118, and controls the secondary oil pressure Pz of the secondary hydraulic oil passage 104 to a predetermined value. to be maintained. Lubricating oil passage 122 is connected to secondary hydraulic oil passage 104 via port 120 or orifice 124. The lock-up control valve 126 selectively connects the secondary hydraulic oil passage 104 to the engagement side and release side of the lock-up clutch 32 within the fluid coupling 10. Solenoid valve 128
is the control chamber 130 of the lock-up control valve 126 and the drain 1
32, and the solenoid valve 128 is turned off (de-energized).
In this case, the secondary hydraulic pressure Pz from the secondary hydraulic oil passage 104 is transmitted to the release side of the lock-up clutch 32, and power is transmitted via the fluid coupling 10. However, solenoid valve 12
If 8 is on (energized), lock-up clutch 3
2 and the secondary hydraulic oil passage 1 to the oil cooler 134
The secondary hydraulic pressure Pz from 04 is supplied, and the power is transmitted via the lock-up clutch 32. Cooler bypass valve 136 controls Tala pressure.

The speed ratio control device includes a speed change direction switching valve device 138 consisting of a first spool valve 142 and a first solenoid valve 144, and a second spool valve 142 and a first solenoid valve 144.
A variable speed switching valve device 140 including a spool valve 146 and a second electromagnetic valve 148 is provided. First solenoid valve 144
is off, the spool of the first spool valve 142 is pressed toward the spring 152 by the secondary hydraulic pressure Pz of the chamber 150, and the first line pressure PlI of the port 154 is
It is sent to the port 158 of the second spool valve 146 via the port 156 of the spool valve 142, and the connection between the port 160 and the drain 162 is cut off.

As a result, the gear ratio T is switched in the decreasing direction.

During the period when the first solenoid valve 144 is on, the hydraulic pressure in the chamber 150 is discharged through the drain 164 of the first solenoid valve 144, and the spool of the first spool valve 142 is held in the chamber by the spring 152.
50 , there is no first line pressure pH at port 156 and port 160 is connected to drain 62 . As a result, the gear ratio is switched in the increasing direction.

During the period when the second solenoid valve 148 is off, the second spool valve 1
The spool 46 is connected to the spring 1 by the secondary hydraulic pressure Pz in the chamber 166.
68, the connection between port 158 and port 170 is broken, and port 172 is connected to port 174. Ports 170 and 172 are connected to C via oil passage 176.
It is connected to the input side hydraulic cylinder 50 of the VT12.

During the period when the second solenoid valve 148 is on, the hydraulic pressure in the chamber 166 is discharged from the drain 178 of the second solenoid valve 148, and the spool of the second spool valve 146 is moved by the spring 168 into the chamber 166.
, port 158 is connected to port 170, and ports 172 and 174 are disconnected.

Port 174 is connected to port 160 via oil passage 180. Orifice 182 directs a small amount of oil from port 158 to port 170 when second solenoid valve 148 is off.

Therefore, during the period when the first solenoid valve 144 is off and the second solenoid valve 148 is on, oil is quickly supplied to the input hydraulic cylinder 50 of the CVT 12, and the gear ratio γ is rapidly reduced. During the period when the first solenoid valve 144 is off and the second solenoid valve 148 is off, oil is supplied to the input hydraulic cylinder 50 of the CVT 12 via the orifice 182, and the gear ratio γ of the CVT 12 gradually decreases. Become. When the first solenoid valve 144 is on and the second solenoid valve 148 is on, oil is not supplied to or discharged from the input hydraulic cylinder 50 of the CVT 12, and the gear ratio γ of the C It will gradually increase depending on the leakage, etc. The first solenoid valve 144 is on and the second solenoid valve 14
8 is off, the oil in the input hydraulic cylinder 50 is discharged from the drain 162, so the gear ratio γ of the CVT 12 increases rapidly.

The gear ratio detection valve 184 is shown in detail in FIG. Sleeves 186, 188 are coaxially disposed within bore 190 and are axially secured by snap rings 192. A rod 194 passes through the end of the sleeve 186 and has a spring seat 196 fixed to the tip. Another rod 198 fixed to one end of the rod 194 is connected to the movable rotating body 4 on the input side.
6 and moves the rod 194 in the axial direction by an amount of displacement equal to the amount of displacement in the axial direction of the movable rotating body 46.

Spool 200 has lands 202, 204 and is axially movably fitted within sleeve 188. Land 202 is space 206 between lands 202 and 204
The land 2 has a passage 210 that communicates with the oil chamber 208.
04 controls the opening area of the boat 212 of the sleeve 188 into the space 206. The boat 212 is connected to the drain 214 through a space around the outer circumference of the sleeve 186. The oil chamber 208 has an output port 216 that generates a gear ratio pressure Pr, and the output port 216 is connected to the line oil passage 82 via an orifice 218. The spring 220 is provided between the spring receiver 196 and the sleeve 188 and biases the rod 194 in the direction of pushing it out of the sleeve 186.
The spring is provided between the spring receiver 196 and the flange 224 of the spool 200 to urge the spool 200 toward the oil chamber 208.

Therefore, as the amount of displacement of the movable rotor 46 relative to the fixed rotor 42 on the input side of the CVT 12 increases, the gear ratio T increases. As the rod 194 is pushed out of the sleeve 186 due to the increase in displacement of the movable rotary body 46, the oil chamber 2
The force applied to the spool 200 by the spring 222 in the 08 direction decreases.

As a result, the spool 200 moves toward the rod 194, and the land 204 increases the opening area of the boat 212 to increase the oil discharge flow rate, so that the gear ratio pressure Pr of the output port 216 decreases. The gear ratio pressure Pr is output port 21
6, the upper limit thereof is defined as the first line pressure Pj21.

The broken lines in FIGS. 8 and 9 indicate the gear ratio pressure Pγ and the gear ratio γ.
This example shows two relationships. As described below, the first
The line pressure pH decreases as the gear ratio T decreases, but
When the gear ratio pressure Pr decreases to the gear ratio T where it becomes equal to the line pressure pH (this gear ratio γ is a function of the throttle pressure Pub and therefore the engine output torque Te), in the gear ratio range below that, Py=P11. . In addition, in FIGS. 8 and 9, the two-dot chain line is the ideal value of the first line pressure P11, and TI>T2.

The cut-off valve 226 is the lock-up control valve 126.
The chamber 2 communicates with the control chamber 130 via an oil passage 228.
30 and a spool 234 that moves in relation to the oil pressure in that 230 and the spring force of a spring 232, the solenoid valve 1
28 is off, that is, when the lock-up clutch 32 is in the released state (when the auxiliary transmission 14 changes gears, the lock-up clutch 32 is released in order to absorb the shock of the power transmission system). ), the gear ratio pressure Pγ is in the closed state and the primary regulator pulp 1
Prevent transmission to 08.

The primary regulator pulp 108 as a first line pressure generating means includes a boat 236 to which throttle pressure P is supplied, a boat 238 to which gear ratio pressure Pr is supplied, a boat 240 connected to the line oil path 82, and a boat 240 connected to the line oil path 82, A boat 242 connected to the suction side of the pump 54 and a first line pressure P , 41 via an orifice 244
a spool 248 that moves axially to control the connection between boats 240 and 242; a spool 250 that receives throttle pressure P and urges spool 248 toward port 238; and a spring 252 that biases the spool 248 toward the port 238. Assuming that the pressure-receiving areas of the two lower lands of the spool 248 are A1 and A2, the pressure-receiving area of the land of the spool 250 that receives the throttle pressure P is A3, and the acting force of the spring 252 is Wl, the following equation (1) and ( 2
) holds true.

When the cut-off valve 226 is open and the gear ratio pressure Pγ is coming to the port 238, P l 1 = (A3・P+Wl −AI・P r
) / (A2-Al) ・ ・ ・ ・ ・ (1) When the cut-off valve 226 is closed and the gear ratio pressure Pγ does not come to the port 238, Pβ1 = (A3・Pth + Wl) / (A2-A
1) ...(2) Note that the line pressure Pi1 of equations (11 and (2)) is shown by a solid line and a dashed-dotted line in FIGS. 8 and 9, respectively.

In FIG. 6, the sub-primary valve 254 as a second line pressure generating means has an input port 256 through which the first line pressure PII is introduced from the port 98 of the manual valve 94 during the LJD range, and a second line pressure PR2. A generated output port 258, a port 260 to which the gear ratio pressure Pγ is introduced, and a second line pressure P as a feedback pressure.
12 through an orifice 262
, a spool 266 that controls opening and closing of the input port 256 and the output port 258, a port 268 to which throttle pressure Pth is introduced, and a spool that urges the spool 266 toward the port 260 in response to the throttle pressure P from the port 268. 270, and spool 266 to port 260
It has a spring 272 that biases it toward. Spool 26
Defining the pressure receiving area of the two lands from the bottom of 6 as B1 and B2, the pressure receiving area of the land of the spool 270 receiving throttle pressure P as B3, and the elastic force of the spring 272 as W2, the following equation (3) is established. do.

P12=(B3・Pth+W2 Bl・Pγ) /(
B2-Bl) (3) FIG. 10 shows the relationship between the second line pressure P12 generated by the sub-primary valve 254 and its ideal value.

Shift valve 274, input port 276 and output port 278 to which second line pressure P12 is introduced when in L range
．． 280, port 288 connected to drain oil passage 286 having orifice 282 and terminating in drain 284, port 1 of manual valve 94 when in D range.
A control boat 300 to which the first line pressure Pβ1 is supplied from 00, other control ports 302 and 304, and a drain 3
06, a spool 308, and a spring 310 that biases the spool 308 toward the control port 304. Secondary hydraulic pressure Pz is guided to the control ports 302 and 304 via an orifice 312. Also, control port 3
The oil pressure at 02.304 is controlled by a solenoid valve 314.

The pressure receiving areas of the two bottom lands of the spool 308 are Sl and B2, respectively, and SL<32. Further, the on/off state of the solenoid valve 314 is controlled in relation to the driving parameters of the vehicle, and when the solenoid valve 314 is on, oil is discharged from the drain 316.

When the spool 308 is in the spring 310 side position, the human powered boat 276 is connected to the output port 278 and the output port 280 is connected to the port 288. Therefore, the second line pressure Pf2 is applied to the piston 3 from the output port 278.
18 and the high speed clutch 72, and the auxiliary transmission 14 becomes high speed.

When the spool 308 is in the position on the control port 304 side, the input port 276 is connected to the output port 280 and the output port 278 is connected to the drain 306. Therefore, the second line pressure P12 from the output port 280 is supplied to the low gear accumulator 322, and the sub-transmission 14
becomes the low gear.

In the case of the L range, the first line pressure Pβ1 is not guided to the control port 300, so when the solenoid valve 314 is turned off, the spool 308 is initially driven by the secondary hydraulic pressure Pz acting on the land of the pressure receiving area S2. Afterwards, it moves toward the spring 310 side by the secondary hydraulic pressure Pz acting on the land of the pressure receiving area S1, but when the solenoid valve 314 is turned on, the control port 30
2. Since the oil pressure at 304 decreases, the spool 308 moves toward the port 304 according to the biasing force of the spring 310.

That is, in the L range, the sub-transmission 142 is switched between a high gear position and a low gear position in conjunction with turning on and off the solenoid valve 314.

In the D range, the first line pressure PNI is applied to the control port 300.
is guided, so once the spool 308 is in the position on the spring 310 side, the control port 30 is connected to the land of the pressure receiving area S2.
A first line pressure pH from 0 is applied, and the spool 308 is held at the position on the spring 310 side regardless of whether the solenoid valve 314 is turned on or off thereafter. Therefore, the sub-transmission 14
is held in the high speed stage.

The shift timing valve 324 is connected to the high speed clutch 7.
2, and a spool 328 whose axial position is controlled by the oil pressure of the control port 326, and is used to supply oil to the high speed clutch 72 during upshift from low speed to high speed. The supply flow rate and the amount of oil discharged from the low speed brake 74 are controlled.

FIG. 11 shows an electronic circuit that controls the operation of the hydraulic control device described above. Electronic control device 330 equipped with a so-called microcomputer consisting of a CPU, RAM, ROM, etc.
A signal representing the throttle valve opening θ is supplied from a throttle valve opening sensor 340 disposed in the intake pipe of the engine 8.

Additionally, a vehicle speed sensor 342 detects the rotational speed of the output shaft 34 of the CVT 12 or the output gear 70 of the sub-transmission 14.
A signal representing the vehicle speed V is sent from the input shaft rotation sensor 344, a signal representing the rotation speed N in of the input shaft 26 of the CVT 12 is sent from the temperature sensor 346, a signal representing the engine coolant temperature T8 is sent from the shift position sensor 348 to the shift lever. A signal representing the operating position P of − is supplied to the electronic control device 330, respectively. The CPU in the electronic control unit 330 processes input signals according to a program stored in advance in the ROM while utilizing the temporary storage function of the RAM, and controls the solenoid valve 12.
A signal for driving 8.144.148.314 is outputted via the amplifier device 332, respectively.

In the electronic control device 330, a control mode is selected by first executing the steps shown in FIG. 12.

That is, when step SMI is executed to initialize the electronic control unit, step SM2 is executed to read human input signals from each sensor, and the dialog in step SM3 is executed according to the read signal conditions. Gnosis, mutual control with the engine computer in step SM4, lock-up control in step SM5, and speed change control in step SM6 are executed sequentially or selectively. The diagnosis in step SM3 is for diagnosing whether the engine 8, CVT 12, etc. are operating normally. The mutual control with the engine computer in step SM4 is to control the mutual relationship with the engine computer that controls the ignition timing, fuel injection amount, etc. of the engine 8. The lock-up control in step SM5 is for controlling the electromagnetic valve 128 that operates the lock-up clutch 32 from a relationship determined in advance based on the vehicle speed (2) and the throttle valve opening θ. Note that the above relationship can also be determined from the rotational speed of the turbine 28 of the fluid coupling 10 and the throttle valve opening θ eye. In addition, the shift control in step SM6 is performed using the shift lever.
Based on the position P, vehicle speed V, and throttle valve opening θ, the switching control of the auxiliary transmission is performed in step SM7, or the control is performed in step S.
This is for determining whether to perform shift control of the M8 CVT. The sub-transmission switching control in step SM7 is for controlling the solenoid valve 314 that selectively switches the sub-transmission 14 to high gear, low gear, or reverse based on the shift position P, vehicle speed, etc. It is something. Then, in step SM9, the control value is output.

In the CVT speed change control in step SM8, steps having the functions shown in the control function block diagram of FIG. 13 are executed. That is, the input shaft rotation sensor 344 that detects the rotation speed of the input shaft 26 supplies the rotation speed N i n of the input shaft 26 to the deviation calculation means 358 . The throttle valve opening sensor 340 detects the throttle valve opening θ and supplies it to the target rotational speed determining means 354 and the switching reference value setting means 356. The target rotational speed determining means 354 determines the throttle valve opening θ, the vehicle speed, and the sub-transmission 14 based on a predetermined relationship such that the required output represented by the throttle valve opening θ is generated on the minimum fuel efficiency curve of the engine. Input shaft 2 of CVT12 based on the range of
6 target rotational speed N i +s.

is determined and supplied to the deviation calculation means 358. The deviation calculating means 358 calculates the deviation Δn1, (=N,...
.

−Ni, ) is calculated and supplied to the electromagnetic valve control means 360. On the other hand, the switching reference value setting means 356 determines reference values n1 to n8 for switching the solenoid valves 144 and 148 based on the throttle valve opening degree θ from a predetermined relationship, and supplies them to the solenoid valve control means 360. do. Here, each reference value is n, >n3 >n2>n, >n4 >n7
>n6 >n, and nI , n= , nz ,
ns is a positive value, and n4, n7, n, , nI are negative values. Note that the switching reference values nl to n8 may be constant values.

The electromagnetic valve control means 360 maintains the actual vehicle speed (■) at a predetermined constant value (■). Determine whether the following is true or not, ■.

If the following, the first solenoid valve 144 and the second solenoid valve 148
is excited (turned on) to suppress the rate of change in the gear ratio γ of the CVT 12, but if it is not less than v0, that is, if the running speed is normal, the switching reference values nl to n8 are compared with the deviation Δni. Solenoid SQL 1 of the first solenoid valve 144 and solenoid 5 of the second solenoid valve 148
A drive signal for exciting OL2 is output. Control operations corresponding to the functions of this electromagnetic valve control means 360 are shown in the flowchart of FIG. That is, in step SSI, it is determined whether the actual vehicle speed V is less than or equal to a predetermined constant value v0, and if this determination is affirmative, step SS2 is executed and the first solenoid valve 144 and the second solenoid valve Both valves 148 are brought into an excited state, but if the answer is NO, the CVT speed change control subroutine of step SS3 is executed. This CVT speed change control subroutine is executed as shown in FIG. 14, and as a result, the speed ratio γ of the CVT 12 is controlled in a direction that eliminates the deviation Δn, and the speed of change of the speed ratio γ is reduced as shown in FIG. Deviation Δ
Controlled in relation to nin. Furthermore, when the vehicle suddenly brakes, a rapid downshift is performed by a gear ratio control routine for sudden braking (not shown), and the gear ratio γ of the CVT 12 is changed to its maximum value Tml1w. Note that FIG. 16 is a diagram showing the contents of each flag F in FIG. 14.

In addition, the above-mentioned predetermined constant value ■. is the 14th gear ratio when the vehicle restarts with the gear ratio γ not at its maximum value.
A value at which the transmission belt 40 does not slip even if the speed change control subroutine shown in the figure is executed, for example, the deviation Δn
, , is smaller than n2, resulting in a slow downshift, or the deviation Δn, , is n, or n4
The vehicle speed is selected so that the upshift is smaller and slower. Such a vehicle speed is usually the minimum value that can be collected by the vehicle speed sensor 342, for example, a value of about 3-/h.

In this way, according to this application example, the vehicle speed ■ is the predetermined constant value ■ as described above. When the following happens, both the first solenoid valves 144 and 148 are turned on, and the CVT 12
is considered to be a slow downshift state, so even if the rotation of CVTI 2 is stopped before the gear ratio γ of CVT 12 reaches its maximum value due to a sudden braking operation of the vehicle, the vehicle speed ■ will remain constant when the vehicle restarts. The value of■. Gear ratio γ until
This suppresses the rapid change in the transmission belt 40, thereby eliminating the slippage of the transmission belt 40 and the resulting decrease in durability and noise generation.

That is, as an example, as shown in the timing chart of FIG. 17, when a sudden braking operation of the vehicle is performed, a rapid downshift occurs due to the gear ratio control routine during sudden braking (not shown), and the gear ratio T quickly increases. It will be done. In this state, the CVT 12 is rotating and the hydraulic oil in the input side hydraulic cylinder 50 is flowing to the second spool valve 146 and the first
Despite being discharged to the drain through the spool valve 142, the belt is moving (the movable rotating body 46 is moving), so a hydraulic pressure P+ is generated in the hydraulic cylinder 50 on the input side. That is, the hydraulic pressure Pc in the hydraulic cylinder 50 becomes Pl.

In normal braking, the vehicle speed decreases along the broken line in Figures 18 and 20, and the gear ratio reaches γamX before it stops, but in sudden braking, the vehicle speed decreases along the dashed line, so the speed change The wheels may stop rotating before the ratio γ reaches its maximum value. In this application example, even in such a case, the first
As shown at point A in Figure 7, the vehicle speed V is a predetermined constant value ■ before the wheels stop. Since it is determined that the second electromagnetic valve 148 is below, the second electromagnetic valve 148, which had been in the off state, is switched to the on state to enter a slow downshift state.

Therefore, even when the wheels stop and the rotation of the CVT 12 stops, the discharge of hydraulic oil from the hydraulic cylinder 50 is suppressed, so the oil pressure Pc in the hydraulic cylinder 50 does not drop much.

Therefore, during the subsequent restart operation, the hydraulic cylinder 5
Since the oil pressure Pc is relatively maintained within 0, slippage of the transmission belt 40 is eliminated. When the vehicle restarts, the vehicle speed ■ becomes the constant value ■.

As shown at point B in FIG. 17, the second electromagnetic valve 148 is turned off, resulting in a rapid downshift, and the gear ratio T is increased. As a result, when the gear ratio γ reaches its maximum value, the groove width of the input side variable pulley 36 is maximized, and the working oil pressure in the hydraulic cylinder 50 becomes approximately atmospheric pressure. Section C in FIG. 17 shows this state.

Incidentally, the vehicle speed is ■. Hydraulic cylinder 5 when
In the conventional case of a type in which the hydraulic oil flowing out from the 0 is not suppressed, the result is as shown by the dashed line in FIG. 17. That is,
Since the state in which the first solenoid valve 144 is on and the second solenoid valve 148 is off is maintained during the sudden braking and subsequent restart operations, the transmission ratio T of the CVT 12 has not yet reached its maximum value Twax. When the rotation of the CVT 12 stops in this state, the hydraulic oil is discharged from the input side hydraulic cylinder 50, and the oil pressure Pc in the hydraulic cylinder 50 changes from Pl, which had been maintained until then, to P2, which is close to atmospheric pressure, and the belt tension decreases. do. At this time, since the belt rotation is stopped, the gear ratio γ also hardly increases. Therefore, when the accelerator pedal was operated to restart the vehicle, the 17th
As shown in D in the figure, there is a temporary increase in oil pressure due to the rotation of the transmission belt 40, but overall the oil pressure Pc in the hydraulic cylinder 50 is still close to atmospheric pressure due to the delay in the generation of oil pressure.
2 and the tension of the transmission belt 40 is insufficient, so it is inevitable that the transmission belt 40 will slip due to the transmission of power from the engine 8. Note that rotation of the CVT 12 stops due to sudden braking, which occurs not only when the vehicle stops but also when the wheels lock.

One application example of the present invention has been described above based on the drawings, but
The invention also applies in other aspects.

For example, in the above application example, the vehicle speed ■ is a constant value ■. By determining a slow downshift when the
Although the outflow of the hydraulic oil in the hydraulic cylinder 50 on the input side is suppressed, the outflow of the hydraulic oil may be stopped.

In addition, the input side hydraulic cylinder 50 of the above-mentioned application example has the input side variable boob IJ 36
Although the groove width is expanded, the groove width may be expanded by supplying hydraulic oil. In such a case, the supply of hydraulic oil is suppressed when the vehicle speed V becomes equal to or less than a certain value V0.

In addition, in the above application example, when restarting, the vehicle speed ■ is a constant value ■. The system is configured so that the outflow of the hydraulic oil in the input side hydraulic cylinder 50 is restricted when the value exceeds the constant value ■, as shown in FIG. A different value ■ may be used. At this time, ■. ,■8 is CVT
12 is determined to be the lowest possible vehicle speed immediately before or at the time of stopping rotation. In this way, the gear ratio γ of the CVT 12 can be brought as close to its maximum value T mmx as possible during sudden braking.

Further, in the application example described above, a constant value is used for V, but it may be a function of other variables.

It should be noted that the above description is merely an example of application of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of the drawing]

FIG. 1 is a flowchart illustrating the main part of the operation of the apparatus to which the present invention is applied. FIG. 2 is a schematic diagram showing a power transmission device for a vehicle to which the present invention is applied. Figure 3 is the second
FIG. 3 is a diagram showing the relationship between the range of the sub-transmission and the frictional engagement device in the device shown in the figure. 4 to 6 are circuit diagrams showing in detail a hydraulic control system for operating the apparatus of FIG. 2. FIG. 7 is a sectional view showing the gear ratio detection valve of FIG. 6 in detail. FIGS. 8 and 9 are graphs showing the characteristics of the first line oil pressure shown in FIGS. 4 to 6. FIG. Figure 10 is the second part of Figures 4 to 6.
It is a graph showing the characteristics of line oil pressure. Figure 11 is the second
FIG. 3 is a block diagram showing a circuit provided in the device shown in the figure. FIG. 12 is a flowchart illustrating the operation of the apparatus of FIG. 2. Figure 13 is the CV of the flowchart in Figure 12.
FIG. 2 is a functional block diagram illustrating functions in steps of speed change control of T. FIG. 14 is a diagram showing details of the shift control subroutine of FIG. 1. FIG. 15 is a diagram illustrating the operation of the CVT in relation to deviation. FIG. 16 is a diagram illustrating the operations of the first solenoid valve and the second solenoid valve in the operation shown in FIG. 14 in a side-by-side manner. Figure 17 shows
Vehicle speed V, gear ratio γ, oil pressure in the input hydraulic cylinder, 1st
2 is a time chart showing changes in the electromagnetic valve and the second electromagnetic valve in relation to the operation shown in FIG. 1; FIG. 18 is a graph showing the vehicle speed and CVT input shaft rotational speed during braking of the vehicle in relation to the gear ratio. FIG. 19 is a diagram corresponding to FIG. 1 showing another example of application of the present invention. 20th
The figure is a graph showing the relationship between vehicle speed and gear ratio when the vehicle is in a braking state. 12: CVT (belt type continuously variable transmission) 26: Input shaft 34: Output shaft 36 Two-man power side variable pulley 38: Output side variable pulley 40: Transmission belt 50: Hydraulic cylinder 138: Shift direction switching valve device 140: Shift speed Switching valve device T: Gear ratio ■: Vehicle speed

Claims (1)

[Scope of Claims] An input shaft and an output shaft, an input-side variable pulley and an output-side variable pulley with variable effective diameters provided on the input shaft and the output shaft, respectively, and the input-side variable pulley and the output-side variable pulley. Supply hydraulic oil to the transmission belt wrapped around the V-groove of the transmission belt and the hydraulic cylinder that changes the V-groove width of the input variable pulley to change the gear ratio, or discharge the hydraulic oil from the hydraulic cylinder. In a belt-type continuously variable transmission for a vehicle, the belt type continuously variable transmission has a shift direction switching valve device that switches the shift direction by changing the shift direction, and a shift speed switching valve device that switches the shift speed by switching the supply flow rate or discharge flow rate of hydraulic oil of the hydraulic cylinder. A transmission ratio control method in which the transmission speed changeover valve device is controlled so that the transmission ratio changes rapidly toward a maximum value when the vehicle speed is lower than a predetermined value. A belt-type continuously variable transmission for a vehicle, characterized in that the speed change switching valve is switched so as to suppress the discharge of hydraulic oil from the hydraulic cylinder for changing the gear ratio or the supply of hydraulic oil to the hydraulic cylinder. transmission ratio control method.