JP2011196496A - Inflow air volume-estimating device of belt type continuously variable transmission and control device using this device - Google Patents

Inflow air volume-estimating device of belt type continuously variable transmission and control device using this device Download PDF

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JP2011196496A
JP2011196496A JP2010065625A JP2010065625A JP2011196496A JP 2011196496 A JP2011196496 A JP 2011196496A JP 2010065625 A JP2010065625 A JP 2010065625A JP 2010065625 A JP2010065625 A JP 2010065625A JP 2011196496 A JP2011196496 A JP 2011196496A
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pressure
oil chamber
pulley
primary
ratio
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JP5398611B2 (en
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Katsumasa Imai
Yukihide Sawada
Hiroyuki Tanijiri
Tamotsu Tsukamoto
Masayuki Yamamoto
Takenori Yoneda
勝政 今井
保 塚本
真之 山本
幸秀 澤田
雄紀 米田
裕之 谷尻
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Daihatsu Motor Co Ltd
ダイハツ工業株式会社
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Abstract

To provide an inflow air amount estimation device for a belt-type continuously variable transmission capable of estimating an inflow air amount to a primary oil chamber generated during slow deceleration.
A correlation between a duty ratio signal to a downshift solenoid valve and an inflow air amount per unit time to a primary oil chamber is stored in advance, and a correlation is obtained when it is determined that a maximum pulley ratio has been reached. From the relationship, the inflow air amount per unit time corresponding to the command duty ratio signal to the solenoid valve for downshift is obtained, and using this inflow air amount, the cumulative inflow air amount to the primary oil chamber when the maximum pulley ratio is reached is obtained. calculate.
[Selection] Figure 6

Description

The present invention relates to an apparatus for estimating the amount of air flowing into a primary oil chamber in a belt-type continuously variable transmission.

Conventionally, a belt type continuously variable gear that performs a shift control and a belt clamping pressure control by wrapping a belt between a primary pulley and a secondary pulley and controlling a supply oil amount / hydraulic pressure of an oil chamber provided in both pulleys. A transmission is known. When controlling this continuously variable transmission, the pulley ratio is controlled by controlling the amount of hydraulic oil to the primary oil chamber with a ratio control valve (flow rate control valve), and the hydraulic pressure supplied to the secondary oil chamber is pinched. The belt clamping pressure is controlled by controlling with a control valve (pressure control valve). A signal pressure is applied to the ratio control valve from the upshift solenoid valve and the downshift solenoid valve so as to oppose each other, and the amount of oil to the primary oil chamber is controlled by the magnitude relationship between the signal pressures. On the other hand, a signal pressure is input from the linear solenoid valve to the clamping pressure control valve, and the hydraulic pressure of the secondary oil chamber is controlled in proportion to the signal pressure.

Generally, when the vehicle is decelerated from the deceleration state, it is necessary to return the pulley ratio to the maximum gear ratio (maximum Low) state before the vehicle stops, so the hydraulic oil is discharged from the primary oil chamber. However, the detection accuracy of the pulley ratio deteriorates at the vehicle speed immediately before the vehicle stops, and it cannot be detected whether the pulley ratio has returned to the lowest state. Therefore, in order to ensure the lowest state, the primary ratio is expected even after the arrival of the lowest state. The hydraulic oil continues to be discharged from the oil chamber. Until the pulley ratio reaches the lowest level, the hydraulic oil in the primary oil chamber is discharged through the ratio control valve controlled by the downshift solenoid valve. When the pulley ratio reaches the lowest level, the movable sheave of the primary pulley is stopped. Therefore, there is no change in the volume of the primary oil chamber. However, since the primary oil chamber is located higher than the oil level of the oil pan, the hydraulic oil is continuously discharged through the ratio control valve due to the head difference, and air corresponding to the volume of the discharged hydraulic oil instead. Will flow into the primary oil chamber through the seal portion.

When air (air is a compressive fluid) flows into the primary oil chamber in this way, when hydraulic oil is supplied to the primary pulley at the time of re-starting, the rise of hydraulic pressure is delayed, and a feeling of stickiness of belt slipping and starting occurs. There was a possibility.

Therefore, conventionally, in order to prevent unnecessary discharge of hydraulic oil from the primary oil chamber when the vehicle stops in the lowest state, the supply of hydraulic oil to the primary oil chamber is changed from flow control to pressure control (so-called confinement control). What is switched is known. However, the closing control is a control that is performed after the vehicle stops, and the inflow of air into the primary oil chamber occurs even during deceleration. Therefore, when the closing control time is short or when the closing control is not performed, the belt Can not prevent the feeling of slipping and starting.

The above-described problem becomes more significant in an idle stop vehicle equipped with a continuously variable transmission having only an oil pump driven by an engine. That is, if the idle stop is performed with air in the primary oil chamber when the vehicle is stopped, the oil pump is also stopped during the idle stop, so that no hydraulic pressure is supplied to the primary oil chamber even if the closing control is performed. For this reason, the closing control is performed only for a short time from the stop of the vehicle to the start of the idle stop, and a small-diameter orifice is set in the supply oil passage to the primary oil chamber in the closing control. Air does not escape in a short time. Therefore, air remains in the primary oil chamber at the time of idling stop return (engine restart), and the pressure increase of the primary oil pressure is delayed, belt slippage occurs transiently, and the start performance is deteriorated.

FIG. 11 is an example of conventional control when the vehicle stops from a deceleration state and the engine automatically stops after the idle stop determination in the idle stop vehicle. Each time change of vehicle speed, ratio (pulley ratio), engine speed, operation of solenoid valves DS1 and DS2 for shift control, and primary hydraulic pressure is shown. DS1 is an upshift solenoid valve, and DS2 is a downshift solenoid valve. In the primary hydraulic pressure, a solid line indicates a change in hydraulic pressure when air does not flow in, and a broken line indicates a change in hydraulic pressure when air flows in.

t1 is a point in time when the vehicle is decelerated below the predetermined vehicle speed and the pulley ratio reaches the lowest level. Eventually, the vehicle stops, and at time t2, both solenoid valves DS1 and DS2 are turned OFF, and the closing control is started. Subsequently, at time t3, it is determined that idle stop is performed, that is, a predetermined engine stop condition is satisfied. Thereafter, the engine is stopped at time t4, and if a predetermined engine return condition is satisfied at time t5, the engine is restarted. Since the movable sheave of the primary pulley comes into contact with the stopper when reaching the lowest (t1), air may start to flow into the primary oil chamber. In particular, when the period (t1 to t2) from reaching the lowest level to the start of the closing control is long, such as when low-speed operation is continued, the time for the air to flow into the primary oil chamber 13 becomes long. The amount of air flowing into the air increases. When air is not flowing in, as shown by the solid line, the primary hydraulic pressure is quickly increased with the start of the closing control, but when air is flowing in, the pressure increase is delayed as indicated by the broken line. In addition, since the orifice is set in the supply oil passage, the hydraulic oil may not be filled and the inflowing air may remain. Eventually, when the engine is stopped at time t4, the oil pump is also stopped, so that the filling of hydraulic oil is stopped. When the engine is restarted by return to idle stop at time t5, air remains in the primary oil chamber, and there is a possibility of belt slip because there is a delay in boosting the primary oil pressure. Furthermore, since the transmission of the driving force is delayed at the time of idling stop return, the starting performance at the time of restart is deteriorated.

In Patent Document 1, in a continuously variable transmission that performs closing control in a low vehicle speed state equal to or lower than a predetermined vehicle speed, a gear ratio changes due to oil oozing out from a gap in a valve of a hydraulic circuit over time. In order to suppress this, there has been disclosed one that repeatedly executes intermediate pressure control and closing control by setting a duty ratio of a solenoid valve.

In Patent Document 2, the total oil loss amount of hydraulic oil that escapes from the CVT during the closing control is calculated, and when the clutch is connected and the engine speed exceeds the reference speed, the calculation is performed from the mechanical oil pump. A control method for replenishing CVT with hydraulic oil based on the total oil loss amount is disclosed.

In Patent Document 1, in order to suppress oil leakage from the clearance of the spool of the flow rate control valve in a low vehicle speed state equal to or lower than a predetermined vehicle speed, attention is focused only on suppressing the change of the gear ratio from the maximum gear ratio. Therefore, no consideration is given to the point where air flows into the primary oil chamber. Further, even if the oil pressure in the primary oil chamber is controlled to be an intermediate pressure between the pressure at the time when the vehicle speed reaches a predetermined vehicle speed and the pressure in the drain state, the amount of inflow air cannot always be reduced.

In Patent Document 2, there is a problem of hydraulic oil falling out during the closing control. However, since the primary oil chamber is pressure-controlled during the closing control, there is no fear of air flowing in. In Patent Document 2, no consideration is given to the air inflow after the movable sheave of the primary pulley comes into contact with the stopper (after the lowest level).

JP 2008-309271 A JP 2006-183830 A

An object of the present invention is to provide an inflow air amount estimation device for a belt-type continuously variable transmission that can estimate an inflow air amount to a primary oil chamber and can easily determine measures such as belt slip prevention, and a control device using this device. It is to provide.

In order to achieve the above object, the present invention has a primary pulley and a secondary pulley around which a belt is wound, and both pulleys are provided with oil chambers for operating movable sheaves, respectively. The pulley ratio is controlled by controlling the amount of hydraulic oil to the oil flow chamber of the secondary pulley by controlling the flow rate control valve to which the signal pressure is input from the upshift solenoid valve and the downshift solenoid valve. In a belt-type continuously variable transmission configured to control a belt clamping pressure by controlling a supply hydraulic pressure with a clamping pressure control valve and configured to stop the movable sheave of the primary pulley against a stopper in a maximum pulley ratio state. , Means for determining a pulley ratio of the continuously variable transmission, and a duty ratio to the solenoid valve for downshift Means for preliminarily storing the correlation between the number and the amount of air flowing into the oil chamber of the primary pulley per unit time, means for detecting a duty ratio signal commanded to the solenoid valve for downshift, and the pulley When the ratio determining means determines that the maximum pulley ratio has been reached, an inflow air amount per unit time corresponding to a command duty ratio signal to the downshift solenoid valve is obtained from the correlation, and this inflow air amount is calculated. And a means for calculating an integrated inflow air amount into the oil chamber of the primary pulley from when the maximum pulley ratio is reached, and an inflow air amount estimation device for a belt-type continuously variable transmission, characterized in that .

When reaching the lowest (maximum pulley ratio), the movable sheave of the primary pulley comes into contact with the stopper, and the hydraulic oil in the primary oil chamber is continuously discharged via the flow control valve (ratio control valve) due to the hydraulic head difference, Air corresponding to the discharge volume flows into the primary oil chamber through the seal portion. However, since there is no means for directly grasping the amount of air flowing into the primary oil chamber, there is a drawback that the filling of the hydraulic oil into the primary oil chamber at the time of re-starting is delayed, causing belt slip and deterioration of start performance.

Therefore, the present inventors conducted various experiments, and as a result of repeated studies on a method for estimating the amount of air flowing into the primary oil chamber, flow rate control for controlling the amount of oil supplied to the primary oil chamber after reaching the lowest level. It was discovered that there is a correlation between the input signal (duty signal) to the downshift solenoid valve for operating the valve and the amount of air flowing into the primary oil chamber per unit time.

When a normally closed downshift solenoid valve is used, a relatively high duty ratio signal is input in order to maintain the lowest level in the deceleration state at a low vehicle speed. The amount of air flowing into the oil chamber per unit time suddenly increases. As described above, if the correlation between the duty ratio signal and the inflow air amount per unit time is known, the inflow air amount per unit time in the command duty ratio signal can be determined from the command duty ratio signal for the downshift solenoid valve. The integrated inflow air amount to the primary oil chamber can be calculated by time integration of the inflow air amount from the lowest arrival time. If the integrated inflow air amount can be estimated in this way, measures against belt slipping and start performance degradation to be taken next can be easily determined.

In the conventional continuously variable transmission, when the vehicle stops from the deceleration state, the closing control is performed in order to maintain the lowest state. The confinement control is a control for switching the hydraulic oil supply to the primary oil chamber from the flow rate control until then to the pressure control. A pressure control valve (ratio check valve) is provided to carry out the closing control, and the supply oil passage to the primary oil chamber is switched from the flow rate control valve to the pressure control valve. This pressure control valve, for example, inputs a secondary hydraulic pressure to one end side, inputs a primary hydraulic pressure to the other end side, controls the primary hydraulic pressure according to the hydraulic pressure balance, and maintains a pulley ratio. If the start timing of the closing control is set when the integrated inflowing air amount calculated by the above-described inflowing air amount estimating device reaches the reference value, the closing control is started earlier than when the closing control is started after the vehicle is stopped. It can be started, and the air pool in the primary oil chamber can be eliminated early. The reference value may be determined based on, for example, the amount of air that does not cause belt slip at the time of restart, or the amount of air that causes the start delay to be less than a predetermined time.

As a method of suppressing or eliminating the air entering the primary oil chamber, various methods can be used in addition to the early start of the closing control as described above. For example, when the reference value of the air amount is changed by deceleration or road surface gradient and it is determined that the inflow air amount has reached the reference value, the closing control is started or the upshift solenoid valve is temporarily turned on. (Or a high duty ratio), it is also possible to quickly fill the primary oil chamber with the working oil and forcibly discharge the air that has flowed in.

Further, in an idle stop vehicle, when it is determined that air of a reference amount or more has flowed in, it is possible to delay engine torque suppression control or start clutch engagement control when the engine is restarted. In the worst case, it is possible to prohibit idle stop even if other idle stop conditions are satisfied. That is, as an idle stop prohibition condition, it may be added whether or not the integrated air amount to the primary oil chamber is greater than or equal to a reference value.

As described above, according to the present invention, when it is determined that the pulley ratio of the continuously variable transmission is in the maximum pulley ratio state, the duty ratio signal to the solenoid valve for downshift and the primary oil chamber obtained in advance are determined. Based on the correlation with the inflow air amount per unit time, the inflow air amount per unit time corresponding to the command duty ratio signal to the downshift solenoid valve is obtained, and this inflow air amount to the primary oil chamber is obtained. Since the integrated inflow air amount is calculated, the amount of air entering the primary oil chamber can be accurately grasped. For this reason, it is possible to take appropriate measures to suppress belt slippage and start-up performance deterioration according to the amount of air.

It is a skeleton figure which shows the structure of the vehicle which concerns on this invention. It is a detailed sectional view of a primary pulley and a secondary pulley. FIG. 2 is a hydraulic circuit diagram of the hydraulic control device for the continuously variable transmission shown in FIG. 1. It is a hydraulic circuit diagram of the principal part of FIG. It is a figure which shows each characteristic of line pressure P L with respect to solenoid pressure Psls, clutch modulator pressure Pcm, clutch control pressure, and secondary pressure. It is a figure which shows the correlation with the duty ratio of the solenoid valve for downshifts concerning this invention, and the inflow air quantity to a primary oil chamber. It is a figure which shows the time change of the duty ratio and integrated air amount which were calculated using the correlation shown in FIG. It is a figure which shows each time change of the vehicle speed, pulley ratio, engine speed, solenoid valve, and primary oil pressure at the time of idling stop which concerns on this invention. It is a flowchart figure of an example of the control method of the continuously variable transmission which concerns on this invention. It is a flowchart figure of the other example of the control method of the continuously variable transmission which concerns on this invention. It is a figure which shows each time change of the vehicle speed at the time of the conventional idle stop, a pulley ratio, an engine speed, a solenoid valve, and a primary hydraulic pressure.

FIG. 1 shows an example of the configuration of a vehicle equipped with a belt type continuously variable transmission according to the present invention. An output shaft 1 a of the engine 1 is connected to a drive shaft (output shaft) 32 via a continuously variable transmission 2. The continuously variable transmission 2 is provided with a torque converter 3, a transmission 4, a hydraulic control device 7, an oil pump 6 driven by the engine 1, and the like.

The continuously variable transmission 2 includes a forward / reverse switching device 8, a primary pulley 11, a secondary pulley 21, and a V that is wound between the pulleys and transmits the rotation to the primary shaft 10 by switching the rotation of the turbine shaft 5 of the torque converter 3 between forward and reverse. The transmission 4 includes a belt 15, a differential device 30 that transmits the power of the secondary shaft 20 to the drive shaft 32, and the like. The turbine shaft 5 and the primary shaft 10 are arranged on the same axis, and the secondary shaft 20 and the drive shaft 32 are arranged parallel to the turbine shaft 5 and non-coaxially. Therefore, the continuously variable transmission 2 has a three-axis configuration as a whole. The V belt 15 is not limited to a known metal belt composed of a continuous tension band and a large number of blocks supported by the tension band, and other belts such as a chain belt may be used.

The forward / reverse switching device 8 includes a planetary gear mechanism 80, a reverse brake B1, and a direct coupling clutch C1. The reverse brake B1 and the direct coupling clutch C1 are wet multi-plate brakes and clutches, respectively. A sun gear 81 of the planetary gear mechanism 80 is connected to the turbine shaft 5 as an input member, and a ring gear 82 is connected to the primary shaft 10 as an output member. The planetary gear mechanism 80 is a single pinion system, the reverse brake B1 is provided between the carrier 84 supporting the pinion gear 83 and the transmission case, and the direct coupling clutch C1 is provided between the carrier 84 and the sun gear 81. When the direct clutch C1 is released and the reverse brake B1 is engaged, the rotation of the turbine shaft 5 is reversed, decelerated and transmitted to the primary shaft 10, and the drive shaft 32 passes through the secondary shaft 20 in the same direction as the engine rotation direction. Since it rotates, it will be in a forward running state. Conversely, when the reverse brake B1 is released and the direct clutch C1 is engaged, the carrier 84 and the sun gear 81 rotate together, so that the turbine shaft 5 and the primary shaft 10 are directly connected, and the drive shaft 32 passes through the secondary shaft 20. Rotates in the direction opposite to the engine rotation direction, so that the vehicle travels backward.

FIG. 2 shows a specific structure of the transmission 4. The primary pulley 11 includes a fixed sheave 11a integrally formed on the primary shaft 10, and a movable sheave 11b supported on the primary shaft 10 so as to be axially movable and integrally rotatable. A cylinder 12 fixed to the primary shaft 10 is provided behind the movable sheave 11 b, and an oil chamber 13 is formed between the movable sheave 11 b and the cylinder 12. Hydraulic oil is supplied to the oil chamber 13 from an oil passage 41 provided in the transmission case 40 through the axial hole 10 a of the primary shaft 10. Shift control is performed by controlling the flow rate of this hydraulic oil with ratio control valves 76 and 77 described later. A seal 42 is provided at a connection portion between the oil passage 41 and the axial hole 10a. Further, a seal 43 is provided on the outer peripheral surface of the movable sheave 11 b that is in sliding contact with the inner peripheral surface of the cylinder 12.

The secondary pulley 21 includes a fixed sheave 21a formed integrally on the secondary shaft 20, and a movable sheave 21b supported on the secondary shaft 20 so as to be axially movable and integrally rotatable. A piston 22 fixed to the secondary shaft 20 is provided behind the movable sheave 21 b, and an oil chamber 23 is formed between the movable sheave 21 b and the piston 22. By controlling the hydraulic pressure (secondary pressure) supplied to the oil chamber 23, a belt clamping pressure necessary for torque transmission is applied. In addition, a bias spring 24 for providing an initial clamping pressure is disposed in the oil chamber 23. In the supply oil passage in the vicinity of the oil chamber 23 of the secondary pulley 21, a hydraulic pressure sensor 108 (see FIG. 3) for detecting the secondary pressure is provided.

One end portion of the secondary shaft 20 extends toward the engine side, and the output gear 27 is fixed to this end portion. The output gear 27 meshes with the ring gear 31 of the differential device 30, and power is transmitted from the differential device 30 to the drive shaft 32 extending left and right to drive the wheels.

The continuously variable transmission 2 is controlled by an electronic control unit 100 (see FIG. 1). The electronic control unit 100 includes an engine speed sensor 101, a secondary pulley speed sensor 102, a throttle opening (or accelerator opening) sensor 103, a shift position sensor 104, a primary pulley speed sensor 105, a brake signal sensor 106, a CVT. Detection signals are input from the hydraulic oil temperature sensor 107 and the hydraulic pressure sensor 108 that detects the secondary pressure. Of course, other signals may be input as the input signal. The pulley ratio can be detected based on detection signals from the primary pulley rotation speed sensor 105 and the vehicle speed sensor 102. Since the secondary pulley rotation speed corresponds to the vehicle speed, the sensor 102 also serves as a vehicle speed sensor. The electronic control unit 100 is linked to an engine control ECU (not shown), and an idle stop execution determination signal is input from the engine control ECU. The idle stop execution determination condition (engine stop condition) includes vehicle speed 0, accelerator off, brake on, and the like, and the engine restart condition (return condition) includes brake off, accelerator on, and the like.

The electronic control device 100 controls a solenoid valve built in the hydraulic control device 7. The hydraulic control device 7 is connected to the oil pump 6, the oil chamber 13 of the primary pulley 11, the oil chamber 23 of the secondary pulley 21, the reverse brake B1, and the direct coupling clutch C1. The electronic control unit 100 determines the target primary rotational speed according to a shift map set in advance according to the vehicle speed and the throttle opening, and controls the solenoid valve in the hydraulic control unit 7 to thereby control the continuously variable transmission 2. Adjust the oil amount / hydraulic pressure of the oil chambers 13 and 23 of the primary pulley 11 and the secondary pulley 21 to control the primary rotational speed to a target value and control the belt clamping pressure to a target value that does not cause belt slip. Yes. The hydraulic control device 7 also has a function of controlling the hydraulic pressure supplied to the reverse brake B1 and the direct coupling clutch C1.

FIG. 3 is a hydraulic circuit diagram of an example of the hydraulic control device 7, and FIG. 4 is a hydraulic circuit diagram of the main part thereof. 3, 71 is a regulator valve, 72 is a clutch modulator valve, 73 is a solenoid modulator valve, 74 is a garage shift valve, 75 is a manual valve, 76 is an upshift ratio control valve, 77 is a downshift ratio control valve, 78 is a ratio check valve and 79 is a clamping pressure control valve. Further, SLS is a linear solenoid valve that outputs a solenoid pressure Psls for performing pressure regulation control of the line pressure, transient control of the reverse brake B1 and direct coupling clutch C1, and pressure control of the oil chamber 23 of the secondary pulley 21. DS1 is an upshift solenoid valve that generates an upshift signal pressure Pds1, and DS2 is a downshift solenoid valve that generates a downshift signal pressure Pds2. The solenoid valves DS1 and DS2 have a function of performing not only shift control but also closing control. In this embodiment, the linear solenoid valve SLS uses a normally open linear solenoid valve, and the solenoid valves DS1 and DS2 both use a normally closed solenoid valve.

Solenoid valves DS1 and DS2 are controlled as follows according to the running state.

In Table 1, ○ indicates an operating state, and × indicates a non-operating state. In addition, (circle) and x contain not only an ON state or an OFF state but a duty control state. The closing control for simultaneously turning off both solenoid valves is performed to keep the maximum pulley ratio while the vehicle is stopped and to prevent belt slippage when the vehicle restarts. On the other hand, the closing control for turning on both solenoid valves is performed during the garage shift.

In FIG. 3, only the hydraulic circuit relating to the primary pulley 11, the secondary pulley 21, the reverse brake B1 and the direct coupling clutch C1 is shown, but the hydraulic circuit such as the lockup clutch 3a built in the torque converter 3 is the same as that of the present invention. Omitted because there is no direct relationship.

The regulator valve 71 is a valve that regulates the discharge pressure of the oil pump 6 to a predetermined line pressure P L, and regulates the line pressure P L according to the solenoid pressure Psls input to the signal port 71a.

The clutch modulator valve 72 is a valve that outputs a clutch modulator pressure Pcm that is a source pressure of supply pressures (P C1 , P B1 ) to the direct coupling clutch C1 and the reverse brake B1. The line pressure P L is input to the input port 72a, and the clutch modulator pressure Pcm is output from the output port 72b. The output pressure is fed back to the first signal port 72c so as to face the spring load. Therefore, the clutch modulator pressure Pcm is adjusted to a constant pressure corresponding to the spring load.

The solenoid modulator valve 73 is a valve that regulates the clutch modulator pressure Pcm and generates a constant solenoid modulator pressure Psm corresponding to the spring load. The solenoid modulator pressure Psm is the original pressure of the upshift solenoid valve DS1 and the downshift solenoid valve DS2, and is also supplied to the clamping pressure control valve 79.

The garage shift valve 74 is for switching the oil passage so that the supply pressure to the direct coupling clutch C1 and the reverse brake B1 can be transiently controlled when the shift lever is switched from N → D or N → R (at the time of garage shift). It is a switching valve. The right side of the center line in FIG. 3 is a transient state, and the left side is a holding state. A spool 74b urged in one direction by a spring 74a is provided, and signal ports 74c and 74d to which an upshift signal pressure Pds1 and a downshift signal pressure Pds2 are input in the same direction as the spring load are formed. Yes. A solenoid modulator pressure Psm is input to the counter port 74h in a direction opposite to the spring load. Since the solenoid valves DS1 and DS2 are both turned on during the garage shift, the signal pressures Pds1 and Pds2 inputted to the signal ports 74c and 74d are both turned on, and the spool 74b moves downward against the spring 74a and moves to the right side. It becomes a transient state. The solenoid pressure (transient pressure) Psls input to the port 74e is output from the output port 74f and supplied to the direct coupling clutch C1 or the reverse brake B1 via the manual valve 75. Therefore, engagement can be started gently while avoiding the engagement shock of the direct clutch C1 or the reverse brake B1 by the solenoid pressure Psls. When at least one of the signal pressures Pds1 and Pds2 is turned OFF, the left side is held, and the clutch modulator pressure (holding pressure) Pcm input to the port 74g is output from the output port 74f and is directly coupled via the manual valve 75. C1 or reverse brake B1 is supplied. Therefore, the engagement state of the direct clutch C1 or the reverse brake B1 can be maintained regardless of the operation of the linear solenoid valve SLS.

The manual valve 75 is a manually operated valve mechanically connected to the shift lever. The manual valve 75 is switched to each range of P, R, N, D, S, and B, and the hydraulic pressure supplied from the garage shift valve 74 is directly coupled to the clutch C1. Alternatively, it selectively leads to the reverse brake B1. The input port 75a is supplied with hydraulic pressure from the garage shift valve 74, the output port 75b is connected to the direct clutch C1, and the output ports 75c and 75d are both connected to the reverse brake B1. In the R range, the manual valve 75 supplies the hydraulic pressure to the direct clutch C1 and drains the hydraulic pressure of the reverse brake B1. In the D, S, and B ranges, the manual valve 75 supplies the hydraulic pressure to the reverse brake B1 and drains the hydraulic pressure of the direct clutch C1. . In the P and N ranges, which are non-traveling ranges, the hydraulic pressures of the direct clutch C1 and the reverse brake B1 are drained together.

The upshift ratio control valve 76 and the downshift ratio control valve 77 change the valve opening area according to the relative relationship between the upshift signal pressure Pds1 and the downshift signal pressure Pds2, and return to the oil chamber 13 of the primary pulley 11. This is a flow control valve that adjusts the amount of hydraulic oil. That is, as shown in FIG. 4, the upshift ratio control valve 76 includes a spool 76b biased in one direction by a spring 76a, and a signal pressure Pds2 is applied to a signal port 76c on one end side where the spring 76a is accommodated. Is entered. The signal pressure Pds1 is input to the signal port 76d on the other end side facing the spring load. Line pressure P L is supplied to the intermediate input port 76 e, and the output port 76 f is connected to the oil chamber 13 of the primary pulley 11. A port 76h connected to a port 78h of a ratio check valve 78 described later is formed between the input port 76e and the drain port 76g, and a downshift ratio control is provided between the output port 76f and the signal port 76d. A port 76 i connected to the port 77 f of the valve 77 and the port 78 d of the ratio check valve 78 is formed.

The downshift ratio control valve 77 includes a spool 77b biased in one direction by a spring 77a, and a signal pressure Pds1 is input to a signal port 77c on one end side in which the spring 77a is accommodated. The signal pressure Pds2 is inputted to the signal port 77d on the other end side facing the spring load. In the intermediate portion, a drain port 77e, a port 77f connected to the port 76i of the upshift ratio control valve 76, and a port 77g connected to the port 78f of the ratio check valve 78 are formed in this order.

The ratio check valve 78 switches the hydraulic pressure of the oil chamber 13 of the primary pulley 11 from flow control to pressure control during the closing control, and maintains the primary pressure at a predetermined pressure corresponding to the ratio with the secondary pressure. It is a pressure control valve. The ratio check valve 78 includes a spool 78b biased in one direction by a spring 78a, and the hydraulic pressure of the secondary pulley oil chamber 23 is input to a signal port 78c on one end side in which the spring 78a is accommodated. The oil pressure of the primary oil chamber 13 is input to the signal port 78d on the other end side facing the spring load via the ports 76f and 76i of the upshift ratio control valve 76. Note that the pressure receiving area of the signal port 78d to which the primary pressure is input is larger by α times than the pressure receiving area of the signal port 78c to which the secondary pressure is input. Line pressure P L is supplied to the input port 78e, and the output port 78f is connected to the port 77g of the downshift ratio control valve 77. Further, a port 78h connected to the port 76h of the upshift ratio control valve 76 is formed between the output port 78f and the drain port 78g.

During the closing control, both solenoid valves DS1 and DS2 are turned OFF or ON, so that the upshift ratio control valve 76 is in the right position in FIG. 4 and the downshift ratio control valve 77 is in the left position in FIG. When the sum of the load due to the secondary pressure and the spring load is relatively larger than α times the load due to the primary pressure, the ratio check valve 78 is in the left position in FIG. 4 and is connected to the input port 78e of the ratio check valve 78. The supplied line pressure P L is supplied from the output port 78f to the primary oil chamber 13 through the ports 77g and 77f of the downshift ratio control valve 77 and the ports 76i and 76f of the upshift ratio control valve 76. Conversely, when the α times the load due to the primary pressure is relatively larger than the sum of the load due to the secondary pressure and the spring load, the ratio check valve 78 is switched to the right position in FIG. Therefore, the primary pressure is drained from the output port 78f and the port 78h via the ports 76h and 76g of the upshift ratio control valve 76. Actually, the spool 78b of the ratio check valve 78 is balanced at an intermediate position between a position connecting the output port 78f and the input port 78e and a position connecting the output port 78f and the port 78h. Thus, the ratio check valve 78 can control the primary pressure so that the ratio between the primary pressure and the secondary pressure has a predetermined relationship, and can maintain the predetermined gear ratio. The supply oil passage connecting the ratio check valve 78 and the primary oil chamber 13 passes through the upshift ratio control valve 76 and the downshift ratio control valve 77, and has a small diameter in the oil passage between the ports 78f and 77g. A simple orifice 90 is set. Due to the action of these orifices 90, a sudden speed change is prevented when switching to the closing control.

The clamping pressure control valve 79 is a valve for controlling the hydraulic pressure (secondary pressure) of the hydraulic oil chamber 23 of the secondary pulley 21. A spool 79g biased in one direction by a spring 79f is provided, and a constant pressure Psm is supplied from a solenoid modulator valve 73 to a signal port 79a on one end side facing the spring load. Line pressure P L is supplied to the input port 79b, the output port 79c is connected to the hydraulic oil chamber 23 of the secondary pulley 21, and the secondary pressure is fed back to the port 79d. The solenoid pressure Psls is supplied to the signal port 79e on the other end side in which the spring 79f is accommodated. Port 79h is a drain port. Therefore, the hydraulic pressure obtained by amplifying the solenoid pressure Psls input to the signal port 79e with a predetermined amplification degree can be supplied to the hydraulic oil chamber 23 of the secondary pulley 21 as a secondary pressure. The hydraulic pressure (secondary pressure) in the hydraulic oil chamber 23 is detected by the hydraulic pressure sensor 108, and the belt clamping pressure or the belt transmission torque can be obtained based on the detected hydraulic pressure.

FIG. 5 shows characteristics of the line pressure P L , the clutch modulator pressure Pcm, the clutch control pressure, and the secondary pressure with respect to the solenoid pressure Psls. The line pressure P L is adjusted to a hydraulic pressure substantially proportional to the solenoid pressure Psls. The clutch modulator pressure Pcm is the same as the line pressure P L until the solenoid pressure Psls reaches a predetermined value, and is limited to a constant pressure when it exceeds the predetermined value. Further, since the solenoid pressure Psls is directly supplied to the reverse brake B1 or the direct coupling clutch C1 in a transient state, the clutch control pressure becomes the solenoid pressure Psls itself. The secondary pressure is proportional to the solenoid pressure Psls and is adjusted to a hydraulic pressure slightly lower than the line pressure P L. As shown in FIG. 5, although both the clutch control pressure and the secondary pressure are controlled by the solenoid pressure Psls, the secondary pressure is always set to exceed the clutch control pressure.

FIG. 6 shows the duty ratio input to the downshift solenoid valve DS2 and the amount of air flowing into the primary oil chamber 13 per unit time in the lowest state (the state where the movable sheave 11b is in contact with the stopper 12a). Is map data showing the correlation. Here, the inflow air amount is estimated from the discharge amount by measuring the discharge amount per unit time of the hydraulic oil from the primary oil chamber 13 when the duty ratio is changed. This map data is obtained experimentally in each continuously variable transmission and is stored in advance in the memory of the electronic control unit 100. In FIG. 6, the inflowing air amount is 0 when the duty ratio is d1 or less, and the inflowing air amount increases proportionally when the duty ratio is d1 or more, and the inflowing air amount becomes the maximum value when the duty ratio is d2 or more. FIG. 6 merely shows an example of the correlation between the duty ratio and the inflow air amount, and it is clear that the correlation differs from that in FIG. 6 depending on the solenoid valve DS2 and the ratio control valves 76 and 77.

Although FIG. 6 shows only one characteristic of the duty ratio of the downshift solenoid valve DS2 and the amount of inflow air per unit time, for example, a plurality of characteristics may be set using the oil temperature as a parameter. That is, since the viscosity of the hydraulic oil is high at low temperatures and the inflow amount of inflow air with respect to the duty ratio tends to decrease, the inflow air amount with respect to the duty ratio signal may be set low at low temperatures. If the correlation between the duty ratio and the inflow air amount cannot be maintained at low temperatures, the characteristics shown in FIG. 6 may be used only during warm-up.

FIG. 7 is data showing the relationship between the change in the duty ratio and the integrated air amount. The integrated air amount can be easily calculated by obtaining the inflow air amount per unit time corresponding to the duty ratio from FIG. 6 and integrating it over time. As shown in the figure, it can be seen that when the duty ratio input to the downshift solenoid valve DS2 is increased stepwise from d1, the integrated air amount entering the primary oil chamber 13 increases in a quadratic function. FIG. 7 is merely an example of the time change of the duty ratio and the integrated air amount, and it is obvious that the change may differ from FIG. 7 depending on the configuration of the hydraulic circuit and the duty ratio.

Next, the control method of the continuously variable transmission according to the present invention will be described with reference to the time chart of FIG. FIG. 8 shows an example of control when the vehicle stops from the deceleration state and the engine is automatically stopped after the idle stop determination. The vehicle speed, pulley ratio, engine speed, operation of solenoid valves DS1, DS2, primary hydraulic pressure Each time change is shown. In the primary hydraulic pressure, the solid line represents the hydraulic pressure change of the present invention, and the broken line represents the conventional hydraulic pressure change (see FIG. 10).

In FIG. 8, t1 is the time when the pulley ratio reaches the lowest level, t2 'is the start of the closing control, t3 is the idle stop execution determination, t4 is the engine stop, t5 is the engine restart, and is closed 10 is the same as that of FIG. 10 except for the start of the insertion control (t2 ′). When the vehicle is decelerated from the deceleration state, it is necessary to return the pulley ratio of the continuously variable transmission to the lowest state before the vehicle stops. Therefore, the solenoid valve DS1 is turned off and the solenoid valve DS2 is turned on (for example, duty ratio d1 or more). . Therefore, the hydraulic oil in the primary oil chamber 13 is discharged through the ports 76f, 76i, 77f, and 77e of the ratio control valves 76 and 77, and the discharge of the hydraulic oil is continued even after reaching the lowest state. When reaching the lowest level, the movable sheave 11b of the primary pulley 11 comes into contact with the stopper 12a, so that the volume of the primary oil chamber 13 does not change, and thereafter, the air corresponding to the volume of the discharged hydraulic oil becomes the primary oil chamber 13 Flows in. Specifically, since the hydraulic oil is discharged via the ratio control valves 76 and 77 due to the oil level difference (head H) between the shaft hole 10a and the oil pan even after returning to the lowest level, the seal 42 is Air is connected from the oil passage 41 of the arranged transmission case 40 and the shaft hole 10a of the primary shaft 10 or from the gap between the inner peripheral surface of the cylinder 12 where the seal 43 is disposed and the outer peripheral surface of the movable sheave 11b. Flows in (see FIG. 2).

In the present invention, when it is determined that the pulley ratio has reached the lowest level at time t1, the command duty ratio to the solenoid valve DS2 is detected, the inflow air amount with respect to the command duty ratio is read from FIG. The integrated air amount from time (t1) is calculated. When the integrated air amount exceeds the reference value (t2 ′), the closing control is started. This reference value is determined on the basis of an air amount that does not cause belt slip when the engine is restarted, for example, and is set to about 10 cm 3 , for example. The start (t2 ′) of the closing control is earlier than the conventional starting of the closing control (t2, see FIG. 10). Therefore, the inflow time of air becomes short and the inflow amount of air into the primary oil chamber 13 can be suppressed.

In the closing control, both the solenoid valves DS1 and DS2 are turned off, and the primary oil chamber 13 is supplied with hydraulic pressure from the ratio check valve 78. However, the filling of the hydraulic oil is delayed due to the orifice 90 (see FIG. 4) set in the supply oil passage. In this embodiment, since the closing control is started when the integrated air amount exceeds the reference value, the closing control period (t2 ′ to t4) can be made longer than before, and the engine is stopped. In addition, air accumulation in the primary oil chamber 13 can be reliably eliminated. As is apparent from FIG. 8, the primary oil pressure increase gradient in the closing control is larger than that in the conventional example (FIG. 10). The reason is that the closing control is started from a state where the amount of air entering the primary oil chamber 13 is small, so that the hydraulic oil is quickly charged and the pressure is quickly increased.

When the engine is restarted at time t5, the air in the primary oil chamber 13 has already been discharged, so the oil pressure in the primary oil chamber 13 rises quickly, eliminating the delay in boosting the belt clamping pressure, and thus causing belt slippage. Can be prevented. In addition, since the engine driving force can be transmitted promptly at the time of idling stop return, the starting performance at the time of restart is improved.

FIG. 9 shows an example of a control method for a continuously variable transmission according to the present invention. First, the belt low return determination is performed (step S1). This Low return determination is based on whether or not the vehicle speed is lower than the predetermined vehicle speed V1, the pulley ratio is greater than the predetermined value R1 close to the lowest level, and the duty ratio of the solenoid valve DS2 is greater than the predetermined duty ratio d1 for a predetermined time. judge. If the Low return determination is negative, it is next determined whether or not the vehicle speed is lower than V2 that is lower than V1 (step S2). If the vehicle speed is less than V2 and has not returned to the lowest level, the Low return is given up. Then, the closing control is performed (step S3), and then the idle stop (IDS) is activated (step S4).

Next, the inflow air amount (Q) corresponding to the duty ratio of the solenoid valve DS2 is estimated (step S5), then the integrated air amount is calculated from the inflow air amount Q, and this integrated air amount is compared with the reference value Q1 ( Step S6). The reference value Q1 can be set to about 10 cm 3 , for example. If the accumulated air amount ≦ Q1, the air amount in the primary oil chamber is equal to or less than the allowable amount, and therefore the shift control is continued (step S7). The shift control is duty control of the solenoid valve DS2 for controlling the pulley ratio to the Low side. It is determined whether or not the vehicle has completely stopped (step S8). If the vehicle has completely stopped, an idle stop (IDS) is activated (step S9). On the other hand, if the integrated air amount> Q1, it is considered that an unacceptable amount of air has entered the primary oil chamber, so the closing control is performed immediately (step S10), and then whether or not the vehicle has completely stopped is determined. It judges (step S11), and when it stops completely, idle stop (IDS) is operated (step S12). Since the start of the closing control is advanced, almost no air remains in the primary oil chamber already at the time of idling stop, and a delay in boosting response when the engine is restarted next time can be prevented.

In FIG. 9, the method of speeding up the start of the closing control according to the amount of air flowing into the primary oil chamber 13 has been described, but the idle stop execution determination may be changed according to the amount of air flowing in. For example, as shown in FIG. 10, when it is determined that the integrated air amount has exceeded the reference value Q1, the shift control is continued (step S13), and even after it is determined that the vehicle has completely stopped (step S14), Stop (IDS) is prohibited (step S15). That is, even when all other idle stop conditions are satisfied, idle stop is not performed. In this case, the closing control may not be performed, and may be started after the vehicle is stopped as in the conventional case (FIG. 10).

Furthermore, instead of prohibiting the IDS operation of FIG. 10, the IDS may be operated and the belt input torque at the time of engine restart may be limited. Specifically, the belt input torque may be limited by limiting the engine torque at restart or limiting the clutch engagement torque in order to prevent the hydraulic response delay and belt slippage.

In the present invention, since the amount of air flowing into the primary oil chamber 13 can be estimated, belt slippage and start performance degradation such as early start of closing control, prohibition of idle stop, limitation of belt input torque, etc. as described above It is possible to quickly determine measures for suppressing the problem. In the present invention, since the integrated air amount can be estimated using an existing device (including a hydraulic circuit) simply by changing the control software, the cost can be increased without causing an increase in cost.

In the above description, the idle stop vehicle has been described as an object. However, even in a non-idle stop vehicle, when the vehicle travels at a reduced speed with the maximum pulley ratio, air enters the primary oil chamber, and thus the present invention can be similarly applied. .

In the above embodiment, two ratio control valves 76 and 77 for upshifting and downshifting are provided as ratio control valves, and upshift solenoid valve DS1 and downshift solenoid valve DS2 are provided for both control valves. The flow rate of the hydraulic oil in the primary oil chamber 13 is controlled by inputting the signal pressure from the opposite to each other. However, the ratio control valve is constituted by a single flow control valve, and the upshift solenoid valve DS1 is connected to both ends thereof. The flow rate of the hydraulic oil in the primary oil chamber 13 may be controlled by inputting the signal pressure of the downshift solenoid valve DS2 so as to face each other. Further, although the output pressure of the ratio check valve 78 is supplied to the primary oil chamber 13 via the ratio control valves 76 and 77, the present invention is not limited to this and may be supplied via another oil passage. . In this case, it is preferable to set an orifice for preventing a sudden shift in the other oil passage.

1 Engine 2 Continuously Variable Transmission 4 Transmission 6 Oil Pump 7 Hydraulic Control Device 11 Primary Pulley 13 Primary Oil Chamber 21 Secondary Pulley 23 Secondary Oil Chamber 76 Upshift Ratio Control Valve (Flow Control Valve)
77 Ratio control valve for downshift (flow control valve)
78 Ratio check valve (pressure control valve)
79 Nipping control valve 90 Orifice 100 Electronic controller 101 Engine speed sensor 102 Secondary pulley speed sensor 103 Throttle opening sensor 104 Shift position sensor 105 Primary pulley speed sensor 108 Hydraulic sensor SLS Linear solenoid valve DS1 Solenoid valve for upshift DS2 Solenoid valve for downshift

Claims (2)

  1. A primary pulley and a secondary pulley around which a belt is wound;
    Each of the pulleys is provided with an oil chamber for operating a movable sheave,
    By controlling the amount of hydraulic fluid to the oil chamber of the primary pulley with a flow rate control valve in which the signal pressure is input from the upshift solenoid valve and the downshift solenoid valve, the pulley ratio is controlled, The belt clamping pressure is controlled by controlling the hydraulic pressure supplied to the oil chamber of the secondary pulley with a clamping pressure control valve.
    In the belt-type continuously variable transmission configured to stop when the movable sheave of the primary pulley hits the stopper in the maximum pulley ratio state,
    Means for determining a pulley ratio of the continuously variable transmission;
    Means for preliminarily storing a correlation between a duty ratio signal to the downshift solenoid valve and an inflow air amount per unit time to the oil chamber of the primary pulley;
    Means for detecting a duty ratio signal commanded to the downshift solenoid valve;
    When the pulley ratio determination means determines that the maximum pulley ratio has been reached, an inflow air amount per unit time corresponding to a command duty ratio signal to the downshift solenoid valve is obtained from the correlation, and the inflow air Means for calculating an integrated inflow air amount to the oil chamber of the primary pulley from the time when the maximum pulley ratio is reached using an amount;
    An inflow air amount estimation device for a belt-type continuously variable transmission.
  2. An inflow air amount estimation device according to claim 1;
    Instead of the flow rate control valve, a pressure control valve that controls the hydraulic pressure supplied to the oil chamber of the primary pulley to maintain the pulley ratio;
    The calculated integrated inflow air amount is compared with a predetermined reference value, and when the integrated inflow air amount exceeds the reference value, the flow control valve to the pressure control valve to the oil chamber of the primary pulley. A control device for a belt-type continuously variable transmission, comprising: switching means for switching a supply oil path.
JP2010065625A 2010-03-23 2010-03-23 Inflow air amount estimation device for belt type continuously variable transmission and control device using this device Expired - Fee Related JP5398611B2 (en)

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Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH09292009A (en) * 1996-04-26 1997-11-11 Unisia Jecs Corp Control device for continuously variable transmission
JP2003343709A (en) * 2002-05-29 2003-12-03 Toyota Motor Corp Control device for continuously variable transmission
JP2007162932A (en) * 2005-12-10 2007-06-28 Hyundai Motor Co Ltd Hydraulic control system for vehicular continuously variable transmission

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH09292009A (en) * 1996-04-26 1997-11-11 Unisia Jecs Corp Control device for continuously variable transmission
JP2003343709A (en) * 2002-05-29 2003-12-03 Toyota Motor Corp Control device for continuously variable transmission
JP2007162932A (en) * 2005-12-10 2007-06-28 Hyundai Motor Co Ltd Hydraulic control system for vehicular continuously variable transmission

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