JPS6331835A - Hydraulic control method for belt type continuously variable transmission - Google Patents

Abstract

PURPOSE:To eliminate hunting and degradation of operationability due to the hunting, by enabling proper switching between intermediate speed shift and low speed shift with reference to the rotary speed of an input shaft when the actual rotary speed of the input shaft follows to a target rotary speed. CONSTITUTION:In a belt type continuously variable transmission 10, the widths of V-grooves in primary and secondary variable pulleys 24, 26 are varied by respective hydraulic cylinders 44, 50, and the shifting direction is switched by controlling supply/discharge of working oil with respect to the input side hydraulic cylinder 44 by means of a speed change direction switching valve 58. The shifting speed is varied by controlling the flow of working oil being supplied/ discharged with respect to the input side hydraulic cylinder 44 by means of a speed change control valve 60. Here, the speed change direction switching valve 58 is controlled according to a deviation between an actual rotary speed and a target rotary speed of an input shaft 32 such that the speed change control valve 60 is switched to a flow suppressing condition when said deviation is within a referential range variable with the rotary speed of the input shaft.

Description

[Detailed Description of the Invention] Technical Field The present invention relates to a hydraulic control method for a belt-type continuously variable transmission.
In particular, the present invention relates to a technique for improving the control characteristics of the gear ratio of a continuously variable transmission.

A type of conventional belt-type continuously variable transmission has an input shaft, an output shaft,
A pair of input-side variable pulleys and output-side variable pulleys provided on the input shaft and output shaft, a transmission belt wound between the input-side variable pulleys and the output-side variable pulleys, and the input-side variable pulleys and the output side variable pulleys. ■Input-side hydraulic cylinder and output-side hydraulic cylinder that change the groove width of the side variable pulley, and a speed change that switches between a state in which hydraulic oil flows into the input-side hydraulic cylinder and a state in which hydraulic oil flows out from the input-side hydraulic cylinder. A directional switching valve and a variable speed control valve that can be selectively operated into two states: a state in which the flow rate of hydraulic oil flowing into an input hydraulic cylinder or a flow rate of hydraulic oil flowing out from the input hydraulic cylinder is restricted and a state in which it is not restricted. There is a belt-type continuously variable transmission equipped with In such a belt type continuously variable transmission, the shift direction switching valve is controlled so that the actual rotational speed of the input shaft matches the target rotational speed, while the deviation between the actual rotational speed and the target rotational speed is When the deviation is within a predetermined first judgment reference value, the shift speed control valve is switched to the flow rate suppression state, but until the deviation exceeds the first judgment reference value and reaches the second judgment reference value, the shift speed is changed. A hydraulic control method is employed in which a control valve is periodically operated repeatedly between the two states to control the flow rate at an intermediate speed. For example, JP-A-60-9526
This is what is described in Publication No. 2.

According to this hydraulic control method, when the deviation exceeds the first judgment reference value, the shift speed control valve is periodically operated repeatedly between the two states, so that the minimum and maximum speeds corresponding to the two states are The gear ratio of the belt type continuously variable transmission is changed at an intermediate speed, and the control characteristics when making the actual input rotational speed follow the target rotational speed are improved.

In other words, the actual input shaft rotational speed or gear ratio sufficiently follows changes in the target input shaft rotational speed or target gear ratio, which are determined in order to achieve the driving conditions requested by the driver at the optimum fuel efficiency rate. This eliminates the hunting phenomenon, particularly in transient conditions, and prevents engine vibrations and deterioration of drivability caused by it.

Problems to be Solved by the Invention However, subsequent research has revealed that there are still problems with this conventional control method.

In other words, the amount of change in the hydraulic oil capacity in the input variable pulley,
Even if the control operation amount, such as the V-groove width change stroke of the input side variable pulley, is the same, the corresponding amount of change in input shaft rotational speed or change in gear ratio will differ in relation to the actual gear ratio. However, in conventional control methods, a constant value is used as the neutral judgment reference value regardless of the actual gear ratio of the belt type continuously variable transmission, so good control characteristics may not be obtained over the entire range of gear ratio changes. There was. For example, when the throttle valve opening is increased to a nearly full open state in response to a vehicle starting operation, the target input shaft rotation speed is suddenly increased, so the gear ratio is rapidly increased in order to make the actual input shaft rotation speed follow it. It is changed by downshifting, but even if the deviation is close to the target input shaft rotational speed and the downshift is switched to a slower one, it will overshoot,
Therefore, even if the amplifier is switched to a slower amplifier shift in an attempt to follow the target value, the amount of increase in the rotation speed due to the engine output torque is greater than the amount of decrease in the input shaft rotation speed due to this, and the target input shaft rotation speed is The deviation may further increase and exceed the first criterion value for switching between a slow upshift and an intermediate speed upshift.

In this way, when the deviation exceeds the first judgment reference value, an upshift of the intermediate speed is started, and the input shaft rotational speed and engine output torque rapidly decrease, and then cross the increasing target rotational speed and undershoot. do. This repetition causes fluctuations in the input shaft rotational speed and transmission torque, which sometimes impairs drivability.

Means for Solving the Problems The present invention has been made against the background of the above circumstances.
The gist of this is that an input shaft and an output shaft, a pair of input-side variable pulleys and output-side variable pulleys provided on the input and output shafts, and windings between the input-side variable pulley and output-side variable pulley. The transmission belt that is hung, the input side hydraulic cylinder and the output side hydraulic cylinder that change the groove width of the input side variable pulley and the output side variable pulley, respectively, and the state in which hydraulic oil is allowed to flow into the input side hydraulic cylinder. A speed change direction switching valve that switches to a state in which hydraulic oil flows out from an input hydraulic cylinder, and a flow rate suppression state that limits the flow rate of hydraulic oil flowing into the input hydraulic cylinder or flowing out from the input hydraulic cylinder. In a belt-type continuously variable transmission equipped with a variable speed control valve that can be selectively switched between two positions: a non-limiting flow rate non-suppression state and a non-restriction state, While controlling the speed change direction switching valve, when the deviation between the actual rotation speed and the target rotation speed is within a predetermined first judgment reference value, the speed change speed control valve is switched to the flow rate suppression state. A hydraulic control method of a type in which the shift speed control valve is periodically switched between the two states until the first judgment standard value is exceeded and the second judgment standard value is reached, the first judgment standard being The purpose is to change the value based on the rotational speed of the input shaft from a predetermined relationship.

Effects of the Invention With this method, the first judgment reference value can be changed based on the rotational speed of the input shaft from a predetermined relationship, so that when making the actual input shaft rotational speed follow the target rotational speed, The intermediate speed shift and slow shift are appropriately switched in relation to the rotational speed of the input shaft, thereby eliminating hunting and the deterioration in drivability caused by the hunting.

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

In FIG. 2, reference numeral 10 is connected to the engine 8, and changes its rotation at a predetermined gear ratio.
It is a belt-type continuously variable transmission that transmits the output of
Hydraulic control device 1 for changing (shifting) the gear ratio
2. On the other hand, the gear ratio controller 14 has
Throttle sensor 16, -next side variable pulley rotation sensor 1
8. From the secondary side variable pulley rotation sensor 20, the transaxle oil temperature sensor 22, etc., a throttle signal ST indicating the throttle operation amount as the required load amount of the engine, and the primary side variable pulley 24 of the belt type continuously variable transmission 10. Rotational speed(
A rotation signal RI representing the actual engine rotation speed), a rotation signal RO representing the rotation speed of the secondary side variable pulley 26, and a temperature signal TH representing the oil temperature of the transaxle are supplied. The gear ratio controller 14 grasps the operating state based on these signals, determines the optimum target engine speed or target gear ratio to obtain the desired operating state mainly at the minimum fuel efficiency, and adjusts the actual engine rotation. Drive signals SD1 and SD2 are supplied to electromagnetic valves 28 and 30 in hydraulic control device 12, respectively, so that the speed or gear ratio matches the target engine speed or target gear ratio. Here, if the rotational speed of the primary side (input side) rotating shaft 32 of the belt type continuously variable transmission 10 is No, and the rotational speed of the secondary side (output side) rotating shaft 34 is N 6 ut, then the gear ratio γ is determined by the following formula, and since the rotational speed N0□ of the secondary rotating shaft 34, which is related to the actual vehicle speed, is common, the target rotational speed N r n and the target gear ratio γ8 are unique. There is a relationship. Therefore, changing the gear ratio γ in order to make the actual rotation speed N in match the target rotation speed Nun'' and making the actual gear ratio γ match the target gear ratio γ1 are substantially the same. .

T==NI, l/Nout The belt type continuously variable transmission 10 and the hydraulic control device 12 are constructed as shown in FIG. 3. That is, the belt type continuously variable transmission 10 includes a pair of primary side (input side) variable pulleys 24 and a secondary side (output side) variable pulley provided on the downstream side rotating shaft 32 and the secondary side rotating shaft 34. A pulley 26 and a transmission belt 36 wound between the variable pulleys 24 and 26 are provided, and the rotational force transmitted from the engine to the primary rotating shaft 32 is transferred to the secondary side via the transmission belt 36. The power is transmitted to the rotating shaft 34, and is further transmitted to the output shaft 40 connected to wheels etc. via a differential device (not shown) via a gear device 38 that includes a gear train such as a planetary gear device and can change the direction of rotation. It is now possible to do so. - The next variable pulley 24 is a fixed rotating body 4 fixed to the - next rotating shaft 32.
2, and a movable rotary body 46 that is axially movably but non-rotatably fitted to the negative side rotating shaft 32 and is moved axially by a primary side (input side) hydraulic cylinder 44,
The groove width of the groove formed between the fixed rotary body 42 and the movable rotary body 46, which are a pair of pulley components, is determined according to the hydraulic pressure of the next-side hydraulic cylinder 44, that is, the effective width of the next-side variable pulley 24. The diameter (the diameter of the transmission belt 36) is changed. Similarly, the secondary variable pulley 26 has a fixed rotating body 48 fixed to the secondary rotating shaft 34, and is fitted to the secondary rotating shaft 34 so as to be movable in the axial direction but not rotatable. side) A movable rotary body 52 that is moved in the axial direction by a hydraulic cylinder 50, and depending on the hydraulic pressure of the secondary hydraulic cylinder 50, there is a V formed
By changing the groove width of the groove, the effective diameter is changed. Note that the above-mentioned secondary hydraulic cylinder 44 as the first hydraulic cylinder has a double piston structure, and even if the same line hydraulic pressure is supplied, it has a larger output than the secondary hydraulic cylinder 50 as the second hydraulic cylinder. It is now possible to obtain it.

The hydraulic control device 12 includes a hydraulic pressure generating device 56 that generates hydraulic oil (line hydraulic pressure) at a predetermined pressure corresponding to the hydraulic signal representing the gear ratio and the throttle valve opening supplied from the sensing valve 54, and Shift direction switching in which the shift direction (transmission ratio changing direction) is switched by supplying line hydraulic pressure to the cylinder 44 to increase the pressure, or by allowing discharge of hydraulic oil from the primary side hydraulic cylinder 44 and decreasing the pressure. The valve 58 and a transmission speed control valve 60 that controls the flow rate of line oil supplied to the downstream side hydraulic cylinder 44 or the flow rate discharged from the primary side hydraulic cylinder 44 to change the shift speed (speed ratio change speed). It is equipped with
Line hydraulic pressure is always supplied to the secondary side hydraulic cylinder 50 and the sensing valve 54.

To further explain the hydraulic control device 12 based on FIG. 4, the hydraulic pressure generating device 56 includes a pump device 62, a regulator valve 64, a throttle detection valve 66, a lubricating oil cooler 68, a Tara pressure valve 70, etc. The line oil pressure, which changes depending on the speed and the gear ratio, is supplied to the gear change direction switching valve 58, the gear change speed control valve 60, the sensing valve 54, etc. via the oil passage 72.

The sensing valve 54 includes a spool valve 74,
- A sensing piston 76 is provided which moves together with the movable rotating body 46 of the next variable boob IJ 24 and applies a biasing force corresponding to the gear ratio γ of the belt type continuously variable transmission 10 to the spool valve element 74 via a spring. , the flow area between the input port door 8 and the output boat 80 is changed by the spool valve 74 responsive to the gear ratio, so that the gear ratio pressure signal corresponding to the gear ratio γ shown in FIG. The input port 82 of the valve 64 is supplied.

The throttle detection valve 66 is engaged with a spool valve element 84 and a cam 86 that rotates with the throttle operation, and is moved with the throttle operation, so that a spring applies a biasing force to the spool valve element 84 corresponding to the throttle valve opening degree. A piston 90 is provided to adjust the flow area of the input port 92 communicating with the oil passage 72, and a throttle pressure signal representing the throttle valve opening shown in FIG. 6 is provided to the output port 94. to the input port 96 of the regulator valve 64.

The regulator valve 64 includes a spool valve element 98 and a valve plunger 100 that receives the gear ratio pressure signal and the throttle pressure signal to control the spool valve element 98, and a line boat 102 connected to the pump device 62. By adjusting the flow area in communication with the return oil passage 104, the line oil pressure of the oil passage 72 communicating with the line boat 102 is adjusted as shown in FIG. That is, since the hydraulic pressure output from the pump device 62 is generated by the pump 106 driven by the engine, the hydraulic pressure is applied to the transmission belt 36 of the belt type continuously variable transmission 10 so as not to increase power wastage. The pressure is set to the minimum necessary to prevent slippage, thereby lowering the vehicle's fuel efficiency. Note that the pump device 62 is connected to a lubricating oil cooler 68, a sensing valve 54, and a drain pipe (not shown).
The hydraulic oil returned to the tank 108 from the throttle detection valve 66, the gear change direction switching valve 58, the gear change speed control valve 60, etc. is pumped up by the pump 106, and the hydraulic oil is transferred to the relief valve 110.
The oil is supplied to the regulator valve 64 through an oil passage 112 equipped with the oil.

The shift direction switching valve 58 and the shift speed control valve 60 cooperate with each other to constitute a shift valve device. The speed change direction switching valve 58 and the speed change speed control valve 60 each include a first solenoid valve 28 and a second solenoid valve 30 and spool valves 118 and 120, respectively, and the line hydraulic pressure is applied to them through an oil path 72. It is supplied to the first solenoid valve 28 and the second solenoid valve 30, respectively. The first solenoid valve 28 is provided with an orifice 122 in a passage that communicates with the oil passage 72, and when the first solenoid valve 28 is closed (when not energized), the line hydraulic pressure passes through the orifice 122 and is applied to the spool valve. When the spool valve element 118 is actuated on the end surface 125 of the spool valve element 124, the spool valve element 124 is moved against the biasing force of the spring 126. When the action of the line hydraulic pressure on the spool valve element 124 is released by draining oil from the downstream side, the spool valve element 124 is moved according to the biasing force of the spring 126. That is, the spool valve element 124 is positioned at two positions, the supply position and the discharge position, in response to the operation of the first solenoid valve 28, and the spool valve element 124 is positioned at the supply position (the spring 126
side), the oil line 72 and the supply line 128 are connected, while the discharge line 130 and the drain line 131 are cut off, thereby preventing the supply of line hydraulic pressure to the next side hydraulic cylinder 44. An acceptable supply condition is obtained, and in the other position, i.e. the discharge position, the oil line 72 and the supply line 1
28 is cut off, the discharge pipe line 130 is opened to the drain pipe line 131, and the - next side hydraulic cylinder 44 is opened to the drain pipe line 131.
This results in a discharge condition that allows the discharge of hydraulic fluid from.

On the other hand, the second solenoid valve 30 is also provided with an orifice 132, and when it is closed, like the first solenoid valve 28, line hydraulic pressure via the orifice 132 acts on the end surface 135 of the spool valve element 134 of the spool valve 120. As a result, the spool valve element 134 is moved against the biasing force of the spring 136, but when the spool valve element 134 is opened, the action of the line hydraulic pressure on the spool valve element 134 is released and the spool valve element 134 is moved.
4 is moved according to the biasing force of the spring 136. The spool valve 120 is equipped with an output port 138 and an input port 140 that communicate with the downstream hydraulic cylinder 44, and the supply pipe 128 and the discharge pipe 1.
a supply boat 142 and a discharge boat 144 respectively connected to the discharge line 1 through an orifice 146;
30 is provided, and when the spool valve element 134 is positioned on the spring 136 side by the closing operation (when not energized) of the second solenoid valve 30, the input port 140 and the deceleration boat 148 are connected to the input port 140. While the port 148 is cut off, the human-powered boat 140 and the discharge port 144
The output port 138 and the supply port 142 are communicated through an orifice 150 formed in the spool valve element 134. Further, when the spool valve element 134 is moved according to the spring 136 in accordance with the opening operation (at the time of excitation) of the second solenoid valve 30, the connection between the output port 138 and the supply port 142, and between the input port 140 and the deceleration port 148 is The input port 140 and the discharge port 14 are communicated with each other.
4 is cut off.

That is, the spool valve element 134 of the speed change control valve 60 is at a throttle position that suppresses the flow of hydraulic oil from the supply pipe 128 toward the output port 138 based on the operation of the second electric valve 30 that functions as a pilot valve. At the same time, it is placed in one of two positions: the open position and the unrestrained open position,
On the contrary, the two positions in the supply system correspond to the two positions in the discharge system: an open position that does not suppress the flow of hydraulic oil from the input port 140 to the discharge pipe line 130, and a throttle position that suppresses the flow of hydraulic oil from the input port 140 to the discharge pipe line 130. In the applied example, one variable speed side f'[Il valve 60 is common to the supply system and the discharge system. When the speed change direction switching valve 58 is in the supply state, either the throttle state or the open state for the supply is set, and when it is in the discharge state, either the throttle state or the open state for the discharge is the speed change speed. Based on the increase/decrease in the effective diameter of the downstream variable pulley 24, which is selected by the control valve 60, the speed change speed is controlled both when the gear ratio is increased and when the gear ratio is decreased. This speed change control valve 60 is
The second solenoid valve 30 is turned on and off, and the operation of the solenoid valve 30 is controlled by the controller 14.

As mentioned above, this controller 14 is supplied with a throttle signal ST representing the engine accelerator operation amount, that is, the throttle valve opening θ, and based on the throttle operation amount (in other words, the engine load amount), mainly the minimum fuel efficiency is achieved. Primary rotating shaft 3 to obtain the required operating performance
A target rotational speed N i n of No. 2 is determined. For example, as shown in FIG. 8, the controller 14 has a functional formula or a function formula that provides a target rotational speed N in * for operating the engine 8 exclusively at the minimum fuel efficiency rate in response to an arbitrary throttle valve opening θ. A conversion table is stored, and from the relationship determined in advance, the actual throttle signal ST
The target engine rotational speed N,2 is determined based on this. In addition, the controller 14 receives the rotation signal R1 of the negative secondary side variable boolean IJ24, so that the actual rotational speed N8 of the primary side rotating shaft 32. , and the controller 14 compares the actual rotational speed N i n and the target rotational speed N i n ' determined as described above, and calculates the deviation Δni between the two.

(=Ni, "-Ni,").

The controller 14 supplies a pulsed drive current to the second solenoid valve 3o to drive the second solenoid valve 30 ON/OFF while the deviation Δni is within a predetermined range. The duty ratio of the pulsed drive current is changed according to the magnitude of the second solenoid valve 3.
It includes drive control means for changing the respective time ratios of the zero excitation time and the zero demagnetization time. Based on this, the time ratio between the two states of the throttled state and the open state of the speed change speed control valve 60 is controlled so that the speed ratio change speed of the continuously variable transmission 10 decreases as the deviation Δni decreases. It will be done. In order to change the duty ratio of the pulsed drive current, the pulse width may be changed while keeping a constant frequency, that is, the period is constant, or the pulse width may be constant and the period may be changed.

Further, when the deviation Δni, is within a predetermined first judgment reference value AIL+ or A1°, the controller 14 positions the second solenoid valve 30 on the side that suppresses the internal liquid amount in the two states. to stabilize gear ratio control. At this time, the value Aug of the first judgment reference value at the time of upshifting is changed in relation to the rotational speed N in of the primary rotation shaft 32 .

Next, an application example of the method according to the present invention will be explained with reference to the flowchart shown in FIG. 9, along with the operation of the hydraulic control device configured as described above.

First, in step S1, the controller 14 reads the throttle valve opening θ and the actual rotation speed Ni7 of the primary rotation shaft 32 based on the throttle signal ST and the rotation signal R1 of the primary variable pulley 24. Then, in step S2, the target rotation speed N i h corresponding to the throttle valve opening θ is determined, and in step S
3, the target rotation speed N, fi” and the actual rotation speed N 1
A deviation Δni++ from n is calculated. Then, after the judgment reference value determination routine for determining the first judgment reference value AIU at the time of upshifting is executed at the steering knob S4,
In step S5, it is determined whether the deviation Δnin is larger than the neutral judgment reference value A0 or smaller than the neutral judgment reference value -80. This neutral judgment standard value A. is a small value close to zero, and provides hysteresis between upshift and downshift, but may be a value of 0.

In the judgment reference value determination routine, as shown in FIG. 1, first, in step SS1, it is determined whether the actual rotational speed N is less than or equal to a predetermined judgment reference rotational speed N1□. If this determination is affirmative, step SS2 is executed and the first determination reference value A1u at the time of upshift is determined to be the value AA, but if this determination is negative, step S83 is executed. In this step SS3, it is determined whether the actual rotational speed N in is equal to or less than a predetermined determination reference rotational speed N i n □. If this judgment is affirmed, step SS4 is executed and the first judgment reference value A1l is determined to be the value Am, but if this judgment is negative, step SS5 is executed and the first judgment reference value AID is set to the value Am. It is determined to be Ac.

Here, the values adopted as the first judgment reference value Al at the time of upshifting are in the relationship of A, >A, >AA,
In addition, the above judgment reference rotation speed is NInt > N1fi
There is a relationship of l. That is, the first judgment reference value A1g is a function of the rotational speed N in of the primary rotating shaft 32, and has a predetermined constant relationship as shown in FIG. 10.

Returning to FIG. 9, in step S5, Δn1fi
<The state determined as AO is the actual engine rotation speed, that is, the actual rotation speed N1f of the -next side rotating shaft 32.
It means that i is too high than the target rotational speed Ni2,
It is necessary to reduce the actual rotational speed Ni, by executing an upshift that reduces the gear ratio T. That is, in the case of Δrlin <−A +1, as shown in step S6, the drive signal SD is applied to the first electromagnetic valve 28.
When I is not supplied and is turned off, the spool valve 124 of the spool valve 118 is moved to the supply position as shown in FIG. Supply line 128, spool valve 120
After that, the supply state is reached where it can be supplied to the primary side hydraulic cylinder 44, and an upshift is selected to move the movable rotary body 46 of the -next side variable boob 1124 toward the fixed rotary body 42 and expand its effective diameter. be. On the other hand, Δni, 1
>A6, contrary to the above, a downshift that increases the gear ratio γ is selected. That is, step S7
As shown in FIG. 2, when the drive signal SDI is supplied to the magnetic valve 28 and turned ON, the spool valve 124 is moved to the discharge position and the discharge pipe 130 is moved to the discharge position.
is opened to the drain pipe 131, thereby creating a discharge state that allows discharge of hydraulic oil from the primary hydraulic cylinder 44 side, and the movable rotary body 46 of the downstream variable pulley 24 is separated from the fixed rotary body 42. This results in a downshift that can reduce the effective diameter.

Then, the above deviation Δn,n (= N i++”
In order to control the rate of change of the gear ratio γ in relation to the magnitude of the absolute value 1Δnt,,l of N++s), the first judgment reference value A1 is set as shown in FIG.
u and the first judgment reference value Alu are set, and at the time of downshifting, the second judgment reference value A is set as shown in FIG.
2 (1) and the first judgment reference value AID are set and prepared in the controller 14.
Except for standard value A20% AzIll, and A11
) is a constant value, and generally there is a relationship such that A2LI≧A+u and A2°≧A16, and AID≧A, u. Therefore, the deviation Δni is within the range of H, I, or J during upshifting, and within the range of KSLSM during downshifting.

In a state where the shift direction switching valve 58 is maintained in the upshift state (hydraulic oil supply state M) by executing step S6, first step S8 By 1Δn! n
It is determined whether l is larger than a reference value Azu.

If it is determined that 1Δn iJ>A 211, the drive signal SD2 is supplied to the second electromagnetic valve 30 in step S9 to turn it on. This state corresponds to a driving state where the duty ratio is 100%.

When the second solenoid valve 30 is turned ON (excited), the oil passage 7
2 line oil flows directly from the supply line 128 to the output boat 138 via the spool valve 118 in the supply state, and is rapidly supplied to the downstream hydraulic cylinder 44. The amount of oil supplied at this time is Q as shown in Fig. 13,
As a result of this supply, the effective diameter of the primary variable pulley 24 is rapidly increased, resulting in a high shift speed (
transmission ratio change speed) is obtained.

On the other hand, in step S8, lΔn i n l >A
If it is determined that it is not tu (not within the range of J), it is determined in step SIO whether 1Δn, 11 is smaller than AIL+ whose value was determined in step S4, that is, whether it is within the range of H. . lΔntnl≦
If AID is determined, the drive signal SD2 is not supplied to the second solenoid valve 30 in step Sll, and the second solenoid valve 3
0 is the 0FF (demagnetized) state. This OFF state is a driving state with a duty ratio of 0%.

When the second solenoid valve 30 is turned off, the spool valve 134 is in the state shown in FIG.
It will be supplied to the primary side hydraulic cylinder 44 via. In this throttling state, the amount of supplied oil is Q2 as shown in FIG. 13, and therefore the effective diameter of the downstream variable pulley 24 is slowly expanded, resulting in a low shift speed during upshifting.

On the other hand, if it is determined in step 5 that the vehicle is in the downshift state, the shift direction switching valve 58 is brought into the downshift state (discharge state) in step S7. In this case as well, the first reference value AI at the time of upshifting is
It is substantially the same except that L+ and the second judgment reference value A211 are changed to the first reference value AIfl and the first judgment reference value A11 at the time of downshifting. That is, step S12
If it is determined that 1Δnin l >A ll1l, the second solenoid valve 30 is turned off in step 13, and a high shift speed in the downshift direction is obtained. Also, in step S13, 1Δntnl≦A
If it is determined that it is ID, the second solenoid valve 30 is turned on in step 315, and a downshift is performed at a low shift speed. Here, the first
Figure 4 shows the relationship between the operating state of the second solenoid valve 30 and the amount of oil flowing out from the primary hydraulic cylinder 44 in the downshift state as described above, where Q represents the maximum flow rate and Q4 represents the minimum flow rate. It shows.

When the deviation 1Δnz++l is between the first judgment reference value A11 and the first judgment reference value A11 during an upshift, and when the deviation 1Δn7,11 is between the first judgment reference value A11 and the first judgment reference value A11 during a downshift,
When the value is between the judgment reference value A11) and the second judgment reference value A2g, that is, when it is in the area 3 and the L65 area in FIGS. 11 and 12, duty control is executed. In such a case, in order to change the duty ratio according to the change in the deviation ]Δn+nl, first in step S16 it is determined whether the deviation 1Δntnl is decreasing or increasing. That is, it is determined whether the difference between the latest calculated deviation 1Δni, . . . l and the deviation 1 Δn, . If it is determined to be positive, the deviation IΔn1M1 is in the process of decreasing, so
In step S17, an operation is performed to make the duty ratio to be determined smaller than the duty ratio D' of the previous cycle by ΔD, and the duty ratio is determined. On the other hand, if it is determined to be negative, it means that the deviation 1Δnznl tends to increase, and in this case, the step of increasing the duty ratio to be determined by ΔD from the previous duty ratio D 318 is executed to determine the duty ratio.

In order to determine the duty ratio, for example, a basic equation shown in the following equation (1) is stored in advance in the controller 14.

D, = ((1/B)X INmun' Nin1+C
)X100...(1) However, B and C are constants, and B=1000. C=0.3. Azo=300゜A
If tD"50, l Ni, l'-NiJ = 300
At the point in time, the arithmetic expression +11 is executed, and the duty ratio D1 at that time becomes 60%.

The duty ratio D1 at the time of upshifting is determined in step S16. as described above. Although determined in S17 and S18, during downshift, the second solenoid valve 30
Since the flow rate suppression state and non-suppression state are opposite to those at the time of upshift with respect to ON/OFF of Whether Δn1fi
It is judged by the sign or negative of . If Δnkn 0, it is an upshift state, so the above is output as it is as the duty ratio, but if Δni+s>0, it is a downshift state, so in step S20,
The following formula (2) % formula %) is executed, and the resulting reduced duty ratio D2 is:
It is output as the determined value for downshifting. In addition, if it corresponds to the above (1) 2, the following equation (3), Dt-(1-((1
/B) X I N! TI”-Nifil-i-C)
)X100 ... (3) can be mentioned, but
In any case, if D2 is, for example, 40%, Dl will be 60%.

In this way, the duty ratio D1 or D2 is finally determined in steps S19 and S20, and the pulsed drive current of such duty ratio or D2 is
supplied to the second solenoid valve 30 in step 321;
It is repeatedly turned on and off. Then, in step S22, the duty ratio D+ obtained through the above steps is read as the duty ratio D' of the previous stage for the next duty ratio determination. The above series of steps are then repeatedly executed.

Due to the duty control described above, the actual rotational speed N is controlled both during upshifts and downshifts.
As i fi approaches the target rotational speed N i h , a state is obtained in which it gradually approaches the target rotation speed N i h , so good gear ratio control responsiveness without overshoot or hunting can be obtained. In order to achieve excellent responsiveness, it is not necessary to use a proportional flow control valve such as a servo valve, which also prevents variations in control characteristics caused by pulp sticks and the like.

Further, the gear ratio T of the belt type continuously variable transmission 10 as described above is
In the control, the first judgment reference value A1 at the time of upshifting
g is the rotational speed of the engine 8, that is, the -next rotation shaft 32
Since the rotational speed N1 is changed in relation to the rotational speed N1, the problem when starting the vehicle is solved.

That is, in the conventional case where the first judgment reference value A1g is set to a constant value, as shown in FIG.
At o, when the accelerator pedal is depressed to the substantially fully open position of the throttle valve and the target rotational speed N in '' rises, the first electromagnetic valve 28 is turned on in order to follow this target rotational speed N i n ''. Then, the second solenoid valve 30 is turned off, and a rapid downshift is executed. Note that when the vehicle initially starts, a start control routine (not shown) is executed, so that the second solenoid valve 30 is kept in the ON state until time t1 has elapsed, and a slow downshift is performed. It is designed to prevent slipping. At the time of such a start, the target rotation speed N i n is different from the actual rotation speed N
When in approaches, as shown between time 1t and time t, the deviation 1Δn+nl first becomes less than the first judgment reference value A0, and duty control of the second solenoid valve 30 is started to reduce the intermediate speed. Then, as shown between time t and time t4, when the deviation 1Δn1nl becomes less than the first judgment reference value A1°, the second solenoid valve 30 is turned on and a slow downshift is performed. . At the above time t4, the actual rotational speed N i n is the target rotational speed N i n'
indicates a state in which the value exceeds the neutral judgment reference value A0. When the overshoot amount further increases, both the first solenoid valve 2 and the second solenoid valve 30 are turned off to eliminate the deviation.
The F state is set and a slow upshift is started. However, in such a start, the statorol valve is started in a fully open state, so the primary rotating shaft 3 due to the slow upshift described above
The effect of increasing the rotational speed due to the output torque of the engine 8 is greater than the effect of decreasing the rotational speed of 2, and the amount of overshoot further expands, and the deviation 1Δn...I reaches the first judgment reference value A1U at the time of amplifier shift. do. The times in FIG. 16 also indicate this state. As a result, an upshift of the intermediate speed is started, and the rate of change in the gear ratio γ of the belt type continuously variable transmission 1o suddenly increases, the rotational speed N, 7 suddenly decreases, and the output torque decreases. Also, for this reason, the actual rotational speed N8. , once falls below the target rotational speed N,l, but the same phenomenon as above occurs and the amplifier shift at an intermediate speed is temporarily performed again.

Time t6 in FIG. 16 indicates this state. In this way, when the vehicle starts, fluctuations in output torque occur,
There were cases where the drivability was poor.

However, according to the present embodiment, the first judgment reference value A1g at the time of amplifier shift is changed to a larger value as the rotational speed of the engine 8, that is, the rotational speed N in of the -next input shaft 32 increases. As shown in FIG. 15, time t4
From time 1. Deviation 1Δn even during the overshoot period leading to
anl does not exceed the first judgment reference value AIL+,
This eliminates fluctuations in output torque and the impediment to drivability caused by the fluctuations in the output torque, as in the conventional case.

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 first
Although the judgment reference value A1U is determined based on a stepwise relationship as shown in FIG. 10, it may be determined based on a linear relationship as shown in FIG. 17, for example.

Furthermore, in the above application example, the judgment reference value determination routine for determining the first judgment reference value A1u at the time of amplifier shift may be provided between step S8 and step SIO in FIG. .

In addition, in the above application example, the first judgment reference value A Io at the time of upshift and the first judgment reference value A Io at the time of downshift.
Although D and D were considered separately, a common value may be used. In such a case, the above-mentioned judgment reference value determination routine is performed in step S12 and step S in FIG.
It can also be provided between 14 and 14.

In addition, within the scope of the spirit of the present invention,
It goes without saying that the present invention can be applied with various changes, improvements, combinations, etc. based on the knowledge of those skilled in the art.

[Brief explanation of the drawing]

FIG. 1 is a diagram showing a main part of the flowchart in FIG. 9, and is a diagram showing an example of application of the method of the present invention, and FIG. 2 is a diagram showing an example of a belt type continuously variable transmission to which the present invention is applicable. It is a figure showing an outline. FIG. 3 is a diagram showing the configuration of the main part of FIG. 2. FIG. 4 is a sectional view of a main part for explaining in more detail the structure of the hydraulic control device shown in FIGS. 2 and 3. FIG. FIG. 5 is a diagram showing the relationship between the gear ratio pressure signal output from the sensing valve of FIG. 4 and the amount of movement of the movable rotating body. FIG. 6 is a diagram showing the relationship between the throttle pressure signal output from the throttle valve of FIG. 4 and the throttle valve opening degree. FIG. 7 is a diagram showing the relationship between the line oil pressure adjusted by the regulator valve of FIG. 4 and the throttle valve opening. FIG. 8 is an example of a graph showing a target engine rotational speed corresponding to the throttle valve opening, and FIG. 9 is a flowchart for explaining the operation in one application example of the present invention. FIG. 10 is a diagram showing the relationships used in the flowchart of FIG. 9. FIGS. 11 and 12 are explanatory diagrams illustrating patterns for determining the magnitude of the deviation between the target rotational speed and the actual rotational speed in the example of FIG. 9. 13 and 14 are diagrams respectively showing the change characteristics of the supply flow rate and the discharge flow rate with respect to the duty ratio of the signal that drives the electromagnetic valve in the shift speed control valve device of FIG. 4. FIG. 15 is a timing chart showing the operation explained in FIG. 9. FIG. 16 is a diagram corresponding to FIG. 15 showing the conventional case. 1st
FIG. 7 is a diagram corresponding to FIG. 10 in another application example of the present invention. 10: Belt type continuously variable transmission 24 Primary side variable pulley (input side variable pulley) 26:
Secondary variable pulley (output variable pulley) 32 Primary rotating shaft (input shaft) 34: Secondary rotating shaft (output shaft) 36: Transmission belt 44 Primary hydraulic cylinder (input hydraulic cylinder) 50
: Secondary side hydraulic cylinder (output side hydraulic cylinder) 58: Shift direction switching valve 60: Shift speed control valve N direct fi'': Target rotational speed AIU: First judgment reference value

Claims (1)

[Claims] An input shaft and an output shaft, a pair of input-side variable pulleys and output-side variable pulleys provided on the input and output shafts, and a pair of input-side variable pulleys and output-side variable pulleys that are wrapped between the input-side variable pulley and the output-side variable pulley. an input-side hydraulic cylinder and an output-side hydraulic cylinder that change the V-groove widths of the input-side variable pulley and the output-side variable pulley, respectively; a state in which hydraulic oil flows into the input-side hydraulic cylinder; and the input side. A speed change direction switching valve that switches to a state where hydraulic oil flows out from the side hydraulic cylinder, and a flow rate suppression state and restriction that restricts the flow rate of the hydraulic oil flowing into the input side hydraulic cylinder or flowing out from the input side hydraulic cylinder. In the belt-type continuously variable transmission equipped with a speed change control valve that can be selectively switched between two positions, a flow rate non-suppression state and a flow rate non-suppression state, the speed change is performed so that the actual rotational speed of the input shaft matches the target rotational speed. While controlling the directional switching valve, when the deviation between the actual rotational speed and the target rotational speed is within a predetermined first judgment reference value, the speed change control valve is switched to a flow rate suppression state; In the hydraulic control method, the flow rate is controlled at an intermediate speed by periodically switching the variable speed control valve between the two states during a period when the flow rate exceeds the first judgment reference value and reaches the second judgment reference value. A hydraulic control method for a belt-type continuously variable transmission, characterized in that the first judgment reference value is changed based on the rotational speed of the input shaft from a predetermined relationship.