JPH10103473A - Control device for automatic transmission - Google Patents

Abstract

(57) [Problem] To provide an automatic transmission that can set an appropriate precharge period without using an erroneous line pressure different from the current state when a precharge period is set using a line pressure. It is an object to provide a control device. A maximum value of a currently feasible line pressure is obtained from an engine speed detected by an engine speed sensor and a temperature of hydraulic oil detected by an oil temperature sensor. Compare the target line pressure set in the above, select the line pressure of the smaller value, this value and the capacity of the hydraulic chamber of the friction element to be precharged or the oil passage etc. communicating with the hydraulic chamber And a controller 300 for setting a precharge period from the above.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for an automatic transmission mounted on an automobile.

[0002]

2. Description of the Related Art Generally, an automatic transmission mounted on an automobile combines a torque converter and a transmission gear mechanism,
The power transmission path of the transmission gear mechanism is switched by a selective operation of a plurality of friction elements such as a clutch and a brake to automatically shift to a predetermined gear.
A hydraulic control circuit that supplies and discharges the working pressure to and from the hydraulic chamber of each friction element is provided.In this type of automatic transmission,
For example, at the start of a shift, first, there is a case where precharge control is performed to quickly fill an oil passage leading to a fastening chamber or a release chamber of a friction element to be engaged or released by the shift and these hydraulic chambers with hydraulic oil. .

[0003] This precharge control is performed in order to avoid a delay in the operation of engaging or disengaging the friction element. Before shifting, hydraulic oil exists in the oil passage and the hydraulic chamber leading to the friction element. Therefore, if the rise of the operating pressure in the hydraulic chamber is delayed in time, for example, a hydraulic control valve such as a duty solenoid valve for controlling the supply of the operating pressure to the friction element is set at the beginning of the shift. In a fully opened state for a predetermined time, the oil path and the hydraulic chamber leading to the friction element are quickly filled with hydraulic oil, thereby promoting the stroke of the friction element in the fastening direction or the release direction. .

In this case, it is important to determine how long the hydraulic oil is precharged.
If the precharge end timing is too early, in other words, if the precharge time is short, the supply amount of hydraulic oil by the precharge is insufficient, and it still takes time until the friction element is engaged or released. Therefore, the effect of the precharge control is reduced, and the response during shifting cannot be sufficiently improved.

On the other hand, for example, Japanese Patent Application Laid-Open No. 7-27217
Japanese Patent Application Laid-Open Publication No. H11-157686 discloses that a precharge time is determined based on a line pressure. That is, the hydraulic control circuit in this type of automatic transmission includes a regulator valve that adjusts the pressure of the hydraulic oil discharged from the oil pump to a line pressure having an appropriate value according to various situations at that time, The line pressure generated by the regulator valve is a so-called source pressure of the operating pressure supplied to each friction element at the time of gear shifting, steady running, or at the time of engagement operation such as ND or NR. Therefore, by detecting the line pressure at the time of the precharge control, the base flow rate per unit time of the hydraulic oil when the hydraulic control valve is fully opened or the like is determined, and the oil flow through the friction element to be precharged is determined. The time to be precharged is determined from the capacity of the road or the hydraulic chamber.

[0006]

Generally, the line pressure is set to a target value appropriate for each situation at the time of shifting as described above, during steady running, or during an engagement operation. The spool position of the regulator valve is adjusted via, for example, a linear solenoid valve provided in the hydraulic control circuit so that the target line pressure is achieved, which is set based on the throttle opening, the engine speed, and the like ( For example, Japanese Patent Application Laid-Open No. 62-12434
No. 3). Therefore, it is conceivable to use the target line pressure set in this way as the value of the line pressure at the present time for determining the precharge time.

However, in that case, the following problem occurs. That is, for example, at idle
Considering the case where engagement operations such as -D and NR are performed, during this engagement operation, the forward clutch or the reverse clutch is engaged, so that these clutches do not slip and are securely engaged. A target line pressure having such a magnitude is set.

On the other hand, the precharge control is performed immediately after the start of the engagement operation. At that time, the engine speed is still as low as the idle speed. In general, an oil pump is directly or indirectly driven by an engine, and the rotation speed (discharge amount of hydraulic oil) of the pump is proportional to the engine rotation speed. At this time, the rotation speed of the oil pump is also low, and therefore, the discharge amount of hydraulic oil from the pump is also small. Therefore, so that the target line pressure is realized,
Even if the pressure of the hydraulic oil discharged from the oil pump is adjusted, the target line pressure cannot be obtained. In some cases, the maximum achievable line pressure at that time, that is, the adjustment rate by the regulator valve is set to 0. It is conceivable that the discharge pressure itself from the oil pump at this time becomes lower than the target line pressure. In such a case, when the precharge time is determined with the target line pressure as the current line pressure, the base flow rate of the hydraulic oil per unit time at the time of the precharge is determined to be a value larger than the amount that can actually flow. As a result, the precharge time is set shorter than the actually required precharge time, and as described above, the supply amount of the hydraulic oil due to the precharge is insufficient,
As a result, the effect of the precharge control is reduced, and the responsiveness during shifting cannot be sufficiently improved.

Further, when the temperature of the hydraulic oil increases, the fluidity of the hydraulic oil increases, and the amount of leakage from the gaps of the oil pump and the valve body increases. The discharge amount or discharge pressure becomes small, and the same problem as described above may occur.

The present invention addresses the above-described situation when performing precharge control, and determines the precharge time more accurately so that the effect of the precharge control can be sufficiently exerted. An object of the present invention is to provide a control device for an automatic transmission which can always reliably prevent a response delay of the automatic transmission.

[0011]

In order to solve the above problems, the present invention uses the following means.

First, an invention described in claim 1 of the present application (hereinafter referred to as "first invention") is provided with an engine output detecting means for detecting a value relating to an output torque of an engine, and driven by the engine. An oil pump for generating a hydraulic pressure by the hydraulic pump, hydraulic pressure adjusting means for adjusting the hydraulic pressure generated by the oil pump to a predetermined target hydraulic pressure determined based on the detection result of the engine output detecting means, Operating pressure supply means for controlling the hydraulic pressure adjusted to the hydraulic pressure to generate a working pressure and supplying the same to the friction element, and the working pressure supply means is adjusted to the target hydraulic pressure at a predetermined time. A control device for an automatic transmission configured to perform a precharge of supplying a hydraulic pressure to a friction element for a predetermined time as an operating pressure, wherein the control device relates to a rotation speed of the oil pump. Oil pump rotational speed detecting means for detecting a value, an upper limit hydraulic pressure determining means for obtaining a maximum value of the hydraulic pressure which can be generated by the oil pump based on a detection result of the detecting means, and a precharge time by the operating pressure supplying means. At least a precharge time setting means for setting using a value of the operating pressure supplied to the friction element at the time of precharging, wherein the precharge time setting means sets a value of the operating pressure used for setting the precharge time. In this case, a smaller one of the target oil pressure and the upper limit oil pressure is used.

According to a second aspect of the present invention (hereinafter referred to as a "second aspect"), in the first aspect, the operating pressure supply means is engaged when switching from the non-traveling speed range to the traveling speed range. The friction element is precharged at the beginning of the switching, and the target oil pressure is set based on the engine speed and the throttle opening.

According to a third aspect of the invention (hereinafter referred to as a "third invention"), an oil temperature detecting means for detecting the temperature of hydraulic oil is provided in the first or second invention. The oil pressure detecting means sets the maximum value of the oil pressure that can be generated by the oil pressure generating means to a smaller value when the oil temperature is higher than when the oil temperature is lower.

By using the above means, each invention of the present application operates as follows.

According to the first aspect of the present invention, when setting the precharge time, the maximum value of the hydraulic pressure that can be generated by the oil pump is detected, and the achievable upper limit hydraulic pressure and the predetermined hydraulic pressure to be adjusted are determined. Since the smaller value of the target oil pressure and the precharge time are set, the discharge amount from the oil pump is sufficiently large, and when the discharge pressure can be adjusted to the target oil pressure, the target oil pressure is set. On the other hand, when the discharge amount from the oil pump is insufficient and the discharge pressure itself does not reach the target hydraulic pressure, the precharge time is set based on the discharge pressure. Become. As a result, the base flow rate of hydraulic oil per unit time at the time of precharge always matches the base flow rate that can actually flow, and as a result, the precharge time is appropriately set, and the effect of the precharge control is ensured. Will be demonstrated.

The oil pump according to the first aspect of the present invention may be either an oil pump driven directly by an engine or an oil pump driven indirectly.

According to the second aspect of the invention, the target oil pressure is set to a value necessary for engaging the friction element based on the engine speed and the throttle opening, and the discharge pressure itself from the oil pump is adjusted. At the start of the engagement operation at the time of idling where the situation where the target oil pressure does not reach the target oil pressure is likely to occur, the precharge time is appropriately set, and the effect of the precharge control is reliably exhibited.

According to the third aspect of the present invention, the maximum achievable hydraulic pressure is higher at high oil temperatures where the amount of hydraulic oil leaking from gaps in the oil pump, valve body and the like increases than at low oil temperatures. Since the actual hydraulic pressure is set to a small value, the temperature of the actual oil pressure is corrected in an appropriate direction, and a more accurate precharge time is set.

[0020]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The embodiments of the present invention will be described below in terms of a mechanical configuration, a hydraulic control circuit, and control operations such as gear shifting.

Mechanical Configuration First, the overall mechanical schematic configuration of the automatic transmission 10 according to the present embodiment will be described with reference to the skeleton diagram of FIG.

The automatic transmission 10 has, as main components, a torque converter 20 and a transmission gear mechanism driven by the output of the converter 20 (hereinafter, the engine side is referred to as front and the non-engine side as rear). And a plurality of friction elements 51 to 40 such as clutches and brakes for switching a power transmission path formed by the planetary gear mechanisms 30 and 40.
55 and a one-way clutch 56, whereby the first to fourth speeds in the D range and the first to third speeds in the S range are provided.
The first and second speeds in the speed range and the L range and the reverse speed in the R range are obtained.

The torque converter 20 includes a pump 2 fixed in a case 21 connected to the engine output shaft 1.
2 and the pump 22
And a stator 25 interposed between the pump 22 and the turbine 23 and supported by the transmission case 11 via a one-way clutch 24 to increase the torque. And a lock-up clutch 26 provided between the case 21 and the turbine 23 and directly connecting the engine output shaft 1 and the turbine 23 via the case 21. And
The rotation of the turbine 23 is output to the planetary gear mechanisms 30 and 40 via a turbine shaft 27.

Here, behind the torque converter 20, an oil pump 12 driven by the engine output shaft 1 via a case 21 of the torque converter 20 is arranged.

On the other hand, the first and second planetary gear mechanisms 30,
Reference numeral 40 denotes sun gears 31 and 41, and a plurality of pinions 32 ... 32 meshed with the sun gears 31 and 41.
42 ... 42 and these pinions 32 ... 32, 42 ... 4
2 and pinion carriers 33, 43, and ring gears 34 meshed with the pinions 32,.
44.

Then, the turbine shaft 27 and the first
A forward clutch 51 is provided between the planetary gear mechanism 30 and the sun gear 31, a reverse clutch 52 is provided between the turbine shaft 27 and the sun gear 41 of the second planetary gear mechanism 40, and a turbine shaft 27 is provided with the second planetary gear mechanism. A 3-4 clutch 53 is interposed between the pinion carrier 43 and the pinion carrier 40, and a 2-4 brake 54 for fixing the sun gear 41 of the second planetary gear mechanism 40 is provided.

Further, the ring gear 34 of the first planetary gear mechanism 30 and the pinion carrier 43 of the second planetary gear mechanism 40
The low reverse brake 55 and the one-way clutch 56 are arranged in parallel between the transmission case 11 and these components, and the pinion carrier 33 of the first planetary gear mechanism 30 and the second planetary gear mechanism Forty ring gears 44 are connected, and the output gear 13 is connected to them.

The output gear 13 is meshed with a first intermediate gear 62 on an idle shaft 61 constituting an intermediate transmission mechanism 60, and is connected to a second intermediate gear 63 on the idle shaft 61 and a differential gear. The input gear 71 of the differential gear 70 meshes with the rotation of the output gear 13 and is input to the differential case 72 of the differential 70.
The left and right axles 73, 74 are driven via the zero.

Here, the relationship between the operating state of the friction elements 51 to 55 such as the clutches and brakes and the one-way clutch 56 and the shift speed is summarized in Table 1 below.

[0030]

[Table 1] The transmission gear mechanism portion of the automatic transmission 10 shown in the above skeleton diagram is specifically configured as shown in FIG. 2, but as shown in FIG. A turbine rotation sensor 305 used for control described later is attached.

The sensor 305 includes a forward clutch 51 whose tip rotates integrally with the turbine shaft 27.
The rotation speed of the turbine shaft 27 is detected by detecting a periodic change in a magnetic field generated by a spline provided on the outer peripheral surface of the drum 51a. I have.

Hydraulic Control Circuit Next, a hydraulic control circuit for supplying and discharging the working pressure to and from the hydraulic chambers provided in the friction elements 51 to 55 shown in FIGS. 1 and 2 will be described with reference to FIG.

Of the above frictional elements, the 2-4 brake 54 composed of a band brake has a fastening chamber 54a and a release chamber 54b as hydraulic chambers to which the operating pressure is supplied, and operates only in the fastening chamber 54a. When pressure is supplied,
-4 The rake 54 is fastened and the operating pressure is supplied only to the release chamber 54b, the operating pressure is not supplied to both the chambers 54a and 54b, and both the chambers 54a and 54b
When the operating pressure is supplied, the 2-4 brake 5
4 is released.

Further, other friction elements 51 to 53, 55
Has a single hydraulic chamber, and the friction element is fastened when operating pressure is supplied to the hydraulic chamber.

(1) Overall Configuration As shown in FIG. 3, the hydraulic control circuit 100 includes, as main components, a regulator valve 101 for generating a line pressure and a manual valve for manually switching a range. 102, a low reverse valve 103, a bypass valve 104, a 3-4 shift valve 105, and a lock-up control valve 106, which operate at the time of gear shifting to switch oil passages leading to the friction elements 51 to 55, and these valves 103 to 106 First and second ON-OFF solenoid valves (hereinafter, referred to as “first and second SVs”) 111 and 112 for operation, and a solenoid relay valve (hereinafter referred to as “relay”) for switching the supply destination of the operating pressure from the first SV 111 "Valve") 1
07, and first to third duty solenoid valves (hereinafter, referred to as “first to third DSVs”) that control generation, adjustment, discharge, and the like of the operating pressure supplied to the hydraulic chambers of the friction elements 51 to 55.
121, 122, 123, etc. are provided.

Here, the first and second SVs 111, 11
Each of the second and first to third DSVs 121 to 123 is a three-way valve, and can obtain a state in which the upper and downstream oil paths are communicated and a state in which the downstream oil path is drained. ing. In the latter case, the oil passage on the upstream side is shut off, so that the operating oil from the upstream side is not drained out in a drain state, and the drive loss of the oil pump 12 is reduced.

Note that the first and second SVs 111 and 112 are O
At the time of N, the upper and downstream oil passages are communicated. Also, the first
When the third DSVs 121 to 123 are OFF, that is, when the duty ratio (the ratio of the ON time in one ON-OFF cycle) is 0%, the third DSVs 121 to 123 are fully opened to completely communicate the upper and downstream oil passages, and When the duty ratio is 10
At 0%, the oil path on the upstream side is shut off to cause the oil path on the downstream side to be in a drain state. At an intermediate duty ratio, the hydraulic pressure on the upstream side is used as the original pressure, and the duty ratio on the downstream side is reduced. An oil pressure adjusted to a corresponding value is generated.

The regulator valve 101 adjusts the pressure of the hydraulic oil discharged from the oil pump 12 to a predetermined line pressure. This line pressure is supplied to the manual valve 102 via the main line 200, and is supplied to a solenoid reducing valve (hereinafter, referred to as a solenoid reducing valve).
(Reducing valve) 108) and the 3-4 shift valve 105.

The line pressure supplied to the reducing valve 108 is reduced by the valve 108 to a constant pressure, and then supplied to the first and second SVs 111 and 112 via the lines 201 and 202. .

This constant pressure is supplied to the relay valve 107 via the line 203 when the first SV 111 is ON, and when the spool of the relay valve 107 is located on the right side in the drawing (the same applies hereinafter). Is
Further, a pilot pressure is supplied to a control port at one end of the bypass valve 104 via the line 204 to urge the spool of the bypass valve 104 to the left. When the spool of the relay valve 107 is located on the left side, the 3-4 shift valve 105
Is supplied as pilot pressure to the control port at one end of
The spool of the 3-4 shift valve 105 is biased to the right.

When the second SV 112 is ON,
The constant pressure from the reducing valve 108 is supplied to the bypass valve 104 via the line 106. When the spool of the bypass valve 104 is located on the right side, the lock-up control valve is further connected via the line 207. The control valve 106 is supplied as pilot pressure to a control port at one end of the control valve 106.
Bias the spool to the left. Also, bypass valve 1
When the spool No. 04 is located on the left side, the line 208
Is supplied as pilot pressure to the control port at one end of the low reverse valve 103 to urge the spool of the low reverse valve 103 to the left.

Further, the constant pressure from the reducing valve 108 is also supplied to the control port 101a of the regulator valve 101 via the line 209. In this case, the constant pressure is adjusted by the linear solenoid valve 131 provided on the line 209 according to, for example, the throttle opening of the engine. Therefore, the line pressure is adjusted by the regulator valve 101 according to the throttle opening and the like. Will be adjusted.

The main line 200 led to the 3-4 shift valve 105 is connected to the first line via the line 210 when the spool of the valve 105 is located on the right side.
The accumulator 141 communicates with the accumulator 141.
To introduce line pressure.

On the other hand, the line pressure supplied from the main line 200 to the manual valve 102 is D, S, L
Are introduced into the first output line 211 and the second output line 212 in each forward range, into the first output line 211 and the third output line 213 in the R range, and into the third output line 213 in the N range.

The first output line 211 is connected to the first output line 211.
It is led to the DSV 121 and supplies the first DSV 121 with a line pressure as a control source pressure. The downstream side of the first DSV 121 is connected to a low reverse valve 1 via a line 214.
03, and when the spool of the valve 103 is located on the right side, a further line (servo apply line)
It is guided to the engagement chamber 54a of the 2-4 brake 54 via 215. When the spool of the low reverse valve 103 is located on the left side, the spool is further guided to the hydraulic chamber of the low reverse brake 55 via a line (low reverse brake line) 216.

Here, from the line 214, the line 2
17 is branched and led to the second accumulator 142.

The second output line 212 is connected to the second output line 212.
The line pressure is supplied to the DSV 122 and the third DSV 123, the line pressure is supplied to these DSVs 122 and 123 as the control source pressure, and the line pressure is also supplied to the 3-4 shift valve 105.

The line 212 led to the 3-4 shift valve 105 is led to the lock-up control valve 106 via the line 218 when the spool of the valve 105 is located on the left side. Is located on the left side, and is further led to the hydraulic chamber of the forward clutch 51 via a line (forward clutch line) 219.

Here, the forward clutch line 2
The line 220 branched from 19 is led to the 3-4 shift valve 105. When the spool of the valve 105 is located on the left side, the line 220 communicates with the first accumulator 141 via the aforementioned line 210, and the valve 105 When the spool is located on the right side, it communicates with the release chamber 54b of the 2-4 brake 54 via the line (servo release line) 221.

The downstream side of the second DSV 122 to which the control source pressure is supplied from the second output line 212 is connected to the line 22.
The pilot pressure is supplied to the control port at one end of the relay valve 107 via the control port 2 to supply the pilot pressure to the port, thereby urging the spool of the relay valve 107 to the left. The line 22 branched from the line 222
3 is led to a low reverse valve 103, and the valve 10
When spool 3 is on the right, line 2
Leads to 24.

From this line 224, the orifice 15
1, the line 225 is branched, and the branched line 225 is led to the 3-4 shift valve 105. When the spool of the 3-4 shift valve 105 is located on the left side, the servo Release line 221
Through the release chamber 54b of the 2-4 brake 54.

Also, the orifice 1
A line 226 is further branched from a line 225 branched through the line 51, and the line 226 is further branched.
Is guided to the bypass valve 104, and is guided to the hydraulic chamber of the 3-4 clutch 53 via the line (3-4 clutch line) 227 when the spool of the valve 104 is located on the right side.

Further, the line 224 is directly led to the bypass valve 104, and communicates with the line 225 via the line 226 when the spool of the valve 104 is located on the left side. That is, the line 224 and the line 225
Are connected to bypass the orifice 151.

The downstream side of the third DSV 123 to which the control source pressure is supplied from the second output line 212 is connected to the line 22.
8 and is led to the lock-up control valve 106, and when the spool of the valve 106 is located on the right side, it communicates with the forward clutch line 219.
When the spool of the lock-up control valve 106 is located on the left side, it communicates with the front chamber 26 b of the lock-up clutch 26 via the line 229.

Further, a third output line 213 from the manual valve 102 is led to the low reverse valve 103 to supply a line pressure to the valve 103. When the spool of the valve 103 is located on the left side, it is guided to the hydraulic chamber of the reverse clutch 52 via a line (reverse clutch line) 230.

The line 231 branched from the third output line 213 is led to the bypass valve 104, and when the spool of the valve 104 is located on the right side, the control of the low reverse valve 103 via the line 208 is performed. The line pressure is supplied to the port as the pilot pressure, and the spool of the low reverse valve 103 is urged to the left.

In addition to the above configuration, the hydraulic control circuit 1
00 is provided with a converter relief valve 109. This valve 109 is a regulator valve 10
After the working pressure supplied from 1 through the line 232 is regulated to a constant pressure, this constant pressure is supplied to the lock-up control valve 106 via the line 233. This constant pressure is applied to the lock-up control valve 1
When the spool of the valve 106 is located on the right side, it is supplied to the front chamber 26b of the lock-up clutch 26 via the aforementioned line 229. When the spool of the valve 106 is located on the left side, the rear chamber is provided via the line 234. 26a.

Here, the lock-up clutch 26 is released when the above-mentioned constant pressure is supplied to the front chamber 26b, but the spool of the lock-up control valve 106 is located on the left side, and 3DSV1
When the operating pressure generated at 23 is supplied to the front chamber 26b, a slip state is set, and the slip amount is controlled according to the operating pressure.

(2) Circuit state at each shift speed On the other hand, as shown in FIG. 4, the automatic transmission 10 has the first and second SVs 111,
112, a controller 300 for controlling the first to third DSVs 121 to 123 and the linear solenoid valve 131, and the controller 300 includes:
A vehicle speed sensor 301 for detecting the vehicle speed of the vehicle, a throttle opening sensor 30 for detecting the throttle opening of the engine
2. Engine rotation sensor 30 for detecting engine speed
3. Shift position sensor 304 for detecting a shift position (range) selected by the driver, torque converter 2
0, a turbine rotation sensor 305 for detecting the rotation speed of the turbine 23, and an oil temperature sensor 3 for detecting the oil temperature of the working oil.
06, etc., and these sensors 301 to
Each of the solenoid valves 111, 11 according to the operation state of the vehicle or the engine indicated by the signal from 306, etc.
2, 121 to 123, 131 are controlled. FIG. 2 shows the turbine rotation sensor 305 in an attached state. Also,
The engine rotation sensor 303 also functions as an oil pump rotation sensor that detects the rotation speed of the oil pump 12.

Next, the first and second SVs 111, 112
The relationship between the operating states of the first to third DSVs 121 to 123 and the supply and discharge states of the operating pressure of the friction elements 51 to 55 to and from the hydraulic chamber will be described for each shift speed.

Here, the combinations (solenoid patterns) of the operating states of the first and second SVs 111 and 112 and the first to third DSVs 121 to 123 for each gear are set as shown in Table 2 below.

In Table 2, (○) indicates the first and second SV1.
ON for 11 and 112, 1st to 3rd DSV 121
Reference numerals 123 to 123 are OFF, and all indicate a state in which the upstream oil passage is communicated with the downstream oil passage and the original pressure is supplied to the downstream as it is. (×) indicates the first and second
OFF for SV111 and 112, 1st to 3rd DS
V121 to V123 are ON.
This shows a state in which the upstream oil passage is shut off and the downstream oil passage is drained.

[0063]

[Table 2] (2-1) 1st gear First, in 1st gear (excluding 1st gear in L range), Table 2
As shown in FIG. 5, only the third DSV 123 operates to generate an operating pressure using the line pressure from the second output line 212 as a source pressure, and this operating pressure is supplied via a line 228 to a lock-up control valve. 106. At this point, the spool of the lock-up control valve 106 is located on the right side,
The operating pressure further increases the forward clutch line 219
Is supplied to the hydraulic chamber of the forward clutch 51 as a forward clutch pressure, whereby the forward clutch 51 is engaged.

Here, the forward clutch line 2
19 is connected to the first accumulator 1 via the 3-4 shift valve 105 and the line 210.
Due to the communication with 41, the supply of the forward clutch pressure is performed gently.

(2-2) Second speed Next, in the second speed state, as shown in Table 2 and FIG. 6, in addition to the first speed state, the first DSV 121 also operates.
An operating pressure is generated using the line pressure from the first output line 211 as a source pressure. This operating pressure is supplied to the low reverse valve 103 via the line 214. At this time, the servo release line 215 is further provided because the spool of the low reverse valve 103 is located on the right side.
And supplied to the fastening chamber 54a of the 2-4 brake 54 as servo apply pressure. Thus, the 2-4 brake 54 is engaged in addition to the forward clutch 51.

Since the line 214 communicates with the second accumulator 142 via the line 217, the supply of the servo apply pressure or the application of the 2-4 brake 54 is performed gently. When the spool of the low reverse valve 103 moves to the left when shifting to the first speed in the L range, which will be described later, the hydraulic oil stored in the accumulator 142 is transmitted from the low reverse brake line 216 to the hydraulic pressure of the low reverse brake 55. The room is precharged.

(2-3) Third speed In the third speed state, as shown in Table 2 and FIG. 7, in addition to the above-mentioned second speed state, the second DSV 122 also operates, and the second output line 212 The operating pressure is generated by using the line pressure from the pressure source as the original pressure. This operating pressure is supplied to the low reverse valve 103 through the line 222 and the line 223. At this point, the spool of the valve 103 is also located on the right side, so that the line 224 is further moved.
Will be introduced.

This operating pressure is introduced from the line 224 to the line 225 via the orifice 151,
At this time, since the spool of the 3-4 shift valve 105 is located on the left side, the servo release line 221 is further moved.
Is supplied as a servo release pressure to the release chamber 54b of the 2-4 brake 54 via the. Thereby, the 2-4 brake 54 is released.

Also, the orifice 1
From line 225 branched through 51, line 22
6 is branched, the operating pressure is increased by the line 226.
At this time, the spool of the bypass valve 104 is located on the right side at this point, and the 3-4 clutch line 227
Is supplied to the hydraulic chamber of the 3-4 clutch 53 as a 3-4 clutch pressure. Therefore, in the third speed, the forward clutch 51 and the 3-4 clutch 53 are engaged, while the 2-4 brake 54 is released.

In the third speed, the second DSV 122 generates an operating pressure as described above,
The spool of the relay valve 107 is moved to the left by being supplied to the control port 107a of the relay valve 107 via 2.

(2-4) Fourth gear Further, in the state of the fourth gear, as shown in Table 2 and FIG.
For the third speed state, the third DSV 123 stops generating the operating pressure, while the first SV 111 operates.

By the operation of the first SV 111, a constant pressure from the line 201 is supplied to the relay valve 107 via the line 203. As described above, the spool of the relay valve 107 moves to the left at the third speed. , The constant pressure is
This is supplied to the control port 105a of the four-shift valve 105, and the spool of the valve 105 moves to the right. Therefore, the servo release line 221 is divided into the line 22 branched from the forward clutch line 219.
0, and the release chamber 54b of the 2-4 brake 54 communicates with the hydraulic chamber of the forward clutch 51.

Then, as described above, the third DSV 123 stops generating the operating pressure and sets the downstream side to the drain state, whereby the servo release pressure in the release chamber 54b of the 2-4 brake 54 and the forward clutch 51 Of the third DSV via the lock-up control valve 106 and the line 228.
It will be drained at 123.
The brake 54 is engaged again, and the forward clutch 51 is released.

Control Operations Next, among the specific control operations performed by the controller 300, the characteristic portions of the present invention will be described mainly regarding the control during manual operation between the respective ranges while the vehicle is stopped.
In particular, the control at the time of the ND engagement operation will be described as an example.

(1) ND operation control When a manual operation is performed from the N range to the D range, the engagement control of the forward clutch 51 is performed.
At this time, in order to reduce the shock at the time of engagement of the forward clutch 51, control is performed to switch to the first speed via the third speed. Therefore, when this operation is performed, supply and discharge control of the 3-4 clutch pressure by the second DSV 122 and supply control of the forward clutch pressure by the third DSV 123 are performed.

(1-1) Control of Third DSV The supply control of the forward clutch pressure by the third DSV 123 is performed according to a program shown in FIG. 9. First, in step S1, a calculated hydraulic pressure Ps is calculated according to a program described later, and then in step S2. Then, it is determined whether or not a precharge flag Fp for which a value is set in precharge control described later is 1 or not.

If the precharge flag Fp has been set to 1, in step S3, the third DSV
If the precharge flag Fp is reset to 0 while the duty ratio of 123 is set to 0 (fully open), that is, if the precharge control is completed, step S4
According to S5, a signal of a duty ratio corresponding to the calculated oil pressure Ps is output to the third DSV 123 until the shift is completed, and a forward clutch pressure corresponding to the calculated oil pressure Ps is supplied.

After that, if the shift is completed, step S
According to 6, S7, the forward clutch pressure is increased to a predetermined value by decreasing the duty ratio to 0% at a constant rate.

(1-2) Calculation of Calculated Oil Pressure Ps The calculation of the calculated oil pressure Ps in step S1 in the above program is performed according to the program shown in FIG.

In this program, first, in step S1
At 1, S12, the engine speed N before the operation to the D range is performed.
e, and a hydraulic pressure Pt also corresponding to the throttle opening θ before the operation to the D range is calculated based on the respective maps. Then, in step S13, the hydraulic pressure Pe and the hydraulic pressure Pe are compared. Calculate the higher oil pressure Ps
Is adopted as the initial value Ps ′ of

Here, as shown in FIGS. 11 and 12, in the maps of the oil pressure Pe corresponding to the engine speed Ne and the oil pressure Pt corresponding to the throttle opening θ, these oil pressures Pe and Pt are expressed as follows. It is set to increase as the engine speed Ne or the throttle opening θ increases.

Then, in step S14, it is determined whether or not a predetermined time T1 has elapsed since the operation to the D range. Until the predetermined time T1 has elapsed, the calculated hydraulic pressure Ps is held at the initial value Ps' in step S15. At the same time, if the predetermined time T1 has elapsed, in step S16, the calculated hydraulic pressure Ps is increased at a constant rate (coefficient C1) in accordance with the elapse of time from the elapse of the time.

As a result, the duty ratio of the third DSV 123 and the forward clutch pressure corresponding thereto are reduced as shown in FIG.
, The turbine rotational speed Nt decreases during the ND operation.

(1-3) Control of the Second DSV On the other hand, the control by the second DSV 122 during the ND operation
-4 The supply control of the clutch pressure (and the servo release pressure) is performed according to the program shown in FIG. 14. First, in step S21, it is determined whether or not the turbine speed Nt has become smaller than a predetermined value C2. Until the step S22, in step S22, a duty ratio signal having a predetermined value C3 for obtaining a predetermined 3-4 clutch pressure is output to the second DSV 122.
, The 3-4 clutch pressure is quickly increased and the state is maintained. This allows
The 3-4 clutch 53 is engaged, and a third speed state is established.

Thereafter, the turbine speed Nt is increased to a predetermined value C2.
If it becomes smaller, the 3-4 clutch 53 is released by increasing the duty ratio up to 100% at a constant rate in steps S23 and S24. Thus, the shift to the first speed is completed.

(2) Line pressure control during ND operation In the automatic transmission 10, the hydraulic control circuit 100 generates operating pressures for engaging and releasing the friction elements.
By selectively supplying this to each friction element according to the gear position, a gear position according to the operating range is achieved, but the hydraulic control circuit 100 is supplied to the friction element. The discharge pressure of the oil pump 12 is adjusted to a predetermined line pressure by a regulator valve 101 as a source pressure of the operating pressure.

In this case, the line pressure is appropriately set in accordance with various situations, and the line pressure control for engaging the forward clutch 51 is performed during the engagement operation from the N range to the D range. Will be This control is performed in accordance with the program of FIG. 15. First, in step S31, it is determined whether or not a predetermined time T2 has elapsed from the time of the above operation. Execute line pressure control during traveling.

On the other hand, after the ND operation, the predetermined time T2
Until elapses, according to steps S32 and S33,
Based on the maps shown in FIGS. 16 and 17, a line pressure P1 according to the engine speed Ne and a line pressure P2 according to the throttle opening θ are obtained.

Here, the line pressure in this case may be basically the former line pressure P1.
When an operation such as ND is performed in a so-called air blowing state, the friction element cannot be securely engaged with the line pressure P1, and a line pressure P2 corresponding to the throttle opening is obtained. In step S34, these line pressures P
Set the higher one of 1, P2 to the target line pressure P _0, it is to carry out the line pressure control.

(3) Precharge Control Next, setting control of the precharge flag Fp whose value is determined in step S2 of the program in FIG. 9, that is, precharge control will be described.

(3-1) Outline of Precharge Control In general, in this type of automatic transmission, the operating pressure generated by the hydraulic control circuit is supplied to the hydraulic chamber of the friction element at the time of shifting or engagement operation. When the friction element is fastened, even if the operating pressure is generated immediately after the start of the operation and supplied to the hydraulic chamber of the friction element, an oil passage from the hydraulic control circuit to the hydraulic chamber of the friction element is initially provided. In addition, since no hydraulic oil is present in the hydraulic chamber, there is a problem that the hydraulic pressure does not immediately rise in the hydraulic chamber and the fastening operation of the friction element is delayed.

Therefore, after the output of the gearshift command or the start of the operation, the hydraulic control valve such as a duty solenoid valve for controlling the supply of the operating pressure to the friction element is fully opened for a predetermined time, and the hydraulic pressure of the friction element is changed. In some cases, control is performed to quickly fill the oil passage to the chamber and the hydraulic chamber with hydraulic oil, and this is called precharge control. By performing such precharge control, the response delay of the engagement operation of the friction element is eliminated.

The precharge control will be generally described as follows in accordance with the program shown in FIG. 18, taking the case of ND operation as an example. This control is performed after the start of the ND operation (time t0 shown in FIG. 13) by the third DS in FIG.
This is executed in parallel with the control program of V123. First, in step S41, the controller 300 initializes the total supply amount Qt of the working oil supplied to the forward clutch 51 by this precharge control to 0 first, and then In step S42, a base flow rate Q passing through the DSV 123 when the third DSV 123 is fully opened (duty ratio 0%) is obtained from the line pressure at that time based on the map set as shown in FIG. In this case, the map is set so that the base flow rate Q increases as the line pressure increases. This is because even when the third DSV 123 is fully opened, the base flow rate Q of the hydraulic oil passing therethrough is set. Is changed by the line pressure at that time, and the base flow rate Q increases as the line pressure increases.

Next, in step S43, the oil temperature correction coefficient K is read from the map set as shown in FIG.
In the oil temperature correction coefficient map, the correction coefficient K is set to be smaller than 1 as the temperature of the hydraulic oil decreases. Then, in step S44, the base flow rate Qx corrected by temperature is obtained by multiplying the base flow rate Q by the correction coefficient K. As a result, even if the fluidity of the hydraulic oil changes depending on the temperature, the base flow rate Qx that always matches the amount that can actually flow can be calculated.

Then, in step S45, the corrected flow rate Qx is integrated, and the total supply amount of hydraulic oil Qt from the start of the control to the present time is calculated.
At 6, it is determined whether or not the total supply amount Qt has exceeded a predetermined value V. Until the total supply amount Qt exceeds the predetermined value V, the process proceeds to step S47, where the precharge flag Fp is set to one.
When t exceeds the predetermined value V, the process proceeds to step S48 to reset the precharge flag Fp to 0.

In this case, the predetermined value V is set to a value corresponding to an oil passage from the valve in the hydraulic control circuit 100 to the hydraulic chamber of the friction element and the capacity of the hydraulic chamber. Therefore, at the time of the ND engagement operation, the predetermined value V is set from the third DSV 123 to a value corresponding to the volume of the line 228, the line 220, and the line 219 or the hydraulic chamber of the forward clutch 51, When the total supply amount Qt exceeds the predetermined value V, the oil passage or the hydraulic chamber is filled with the working oil, and the engagement operation of the forward clutch 51 is temporally promoted.

(3-2) Determination of Line Pressure for Setting Precharge Time When performing the precharge control, the precharge period Tpr (see FIG. 13), that is, the precharge flag Fp is set to 1 It is important to properly set the period set in the setting operation. In particular, if this period is too short, the response delay of the fastening operation will not be sufficiently eliminated. On the other hand, the target line pressure Po set by the line pressure control at the time of the ND operation shown in FIG. 15 is used as the line pressure used to calculate the base flow rate Q in step S42 in FIG. But for example,
When the vehicle is idling, such as when the vehicle is stopped, the engine speed is low, and thus the amount of hydraulic oil discharged from the oil pump 12 is small. As a result, even if the pressure of the hydraulic oil discharged from the oil pump 12 is adjusted by the regulator valve 101 so that the target line pressure Po is realized, the target line pressure Po cannot be obtained. As shown in FIG. 13 as the actual line pressure Pr, the maximum line pressure that can be realized at the time of this ND operation,
That is, it is conceivable that the discharge pressure itself from the oil pump 12 when the adjustment rate by the regulator valve 101 is set to 0 becomes lower than the target line pressure Po.

In such a case, if the precharge time is determined by using the target line pressure Po as the current line pressure, the base flow rate Q of the hydraulic oil per unit time at the time of the precharge is larger than the amount that can actually flow. As a result, the precharge time is set to be shorter than the actually required precharge time, and the amount of hydraulic oil supplied by the precharge control becomes insufficient, and the effect of the precharge control is reduced. In addition, the responsiveness during shifting cannot be sufficiently improved.

When the temperature of the hydraulic oil increases, the fluidity of the hydraulic oil increases, and the amount of leakage from the gaps of the oil pump 12 and the valve body increases. The discharge amount or discharge pressure of the oil becomes small, and the same problem as described above may occur.

Therefore, the controller 300 performs control for appropriately setting the precharge period in FIG.
According to the program shown in (1), the following is performed.

That is, when it is determined in step S51 that the engagement control is being performed, the controller 300 determines the type of the shift, for example, ND or NR, in step S52. Step S5
At 3, the target line pressure Po is determined by performing the line pressure control according to FIG. 15 described above according to the type of the shift.

Next, the controller 300 proceeds to step S
At 54, the current engine speed Ne (oil pump 12
22 and FIG. 23 based on the
From the map set as shown in (1), the current maximum achievable line pressure, that is, the actual line pressure Pr is obtained. At this time, the actual line pressure Pr is set so as to increase as the engine speed Ne (the speed of the oil pump 12) increases or as the oil temperature decreases.

Then, in step S55, the controller 300 determines whether the target line pressure Po and the actual line pressure Pr
In step S56 or step S57, the line pressure having the smaller value is determined according to the result.
The line pressure used for calculating the base flow rate Q in step S42 of FIG.

Thus, when setting the precharge time, the hydraulic pressure that can be generated by the oil pump 12 or the maximum achievable line pressure Pr that can be adjusted by the regulator valve 101 and the target line pressure Po, whichever is smaller. Since the precharge time is set by using the value of the above, the discharge amount from the oil pump 12 is sufficiently large, and if the discharge pressure can be adjusted to the target line pressure Po, the value is determined based on the target line pressure Po. The precharge time is set, while the discharge amount from the oil pump 12 becomes insufficient, and the discharge pressure itself becomes the target line pressure Po.
, The precharge time is set based on the actual line pressure Pr. As a result, the base flow rate Q of the hydraulic oil per unit time at the time of precharge always matches the base flow rate that can actually flow, and as a result, the precharge time is appropriately set, and the effect of the precharge control is reduced. It will surely be demonstrated.

In this case, as shown in FIG. 23, in particular, at a high oil temperature at which the amount of leakage of hydraulic oil from the gaps of the oil pump 12 and the valve body increases, the actual line pressure is lower than at a low oil temperature. Since the pressure Pr is set to a small value, the actual line pressure is temperature-corrected in an appropriate direction, and a more accurate precharge time is set.

[0106]

As described above, according to the present invention, the precharge time is prevented from being determined by a parameter different from the actual state, and a proper and sufficient precharge time is always secured. The effect of (1) is sufficiently exhibited, so that it is possible to reliably avoid the response delay of the fastening operation of the friction element.

[Brief description of the drawings]

FIG. 1 is a skeleton diagram showing a mechanical configuration of an automatic transmission according to an embodiment of the present invention.

FIG. 2 is a sectional view showing a configuration of a transmission gear mechanism of the automatic transmission.

FIG. 3 is a circuit diagram of a hydraulic control circuit.

FIG. 4 is a control system diagram of the automatic transmission.

5 is a main part enlarged circuit diagram showing a first speed state of the hydraulic control circuit in FIG. 5;

FIG. 6 is an enlarged circuit diagram of a main part showing a state of the second speed in the same manner.

FIG. 7 is an enlarged circuit diagram of a main part showing a state of the third speed in the same manner.

FIG. 8 is an enlarged main part circuit diagram showing a state of the fourth speed in the same manner.

FIG. 9 is a flowchart showing a control operation of a third duty solenoid valve during an ND engagement operation.

FIG. 10 is a flowchart showing the operation of the subroutine during the control.

FIG. 11 is a conceptual diagram of a map used for the subroutine.

FIG. 12 is a conceptual diagram of a map.

FIG. 13 is a time chart showing changes of respective data during an ND engagement operation.

FIG. 14 is a flowchart illustrating a control operation of a second duty solenoid valve during an ND engagement operation.

FIG. 15 is a flowchart illustrating an operation of line pressure control during an ND engagement operation.

FIG. 16 is a conceptual diagram of a map used for the line pressure control.

FIG. 17 is a conceptual diagram of a map.

FIG. 18 is a flowchart showing an operation of precharge control.

FIG. 19 is a conceptual diagram of a map used for the precharge control.

FIG. 20 is a conceptual diagram of a map.

FIG. 21 is a flowchart showing an operation of a line pressure determination control used in the precharge control.

FIG. 22 is a conceptual diagram of a map used for the control.

FIG. 23 is a conceptual diagram of the map.

[Explanation of symbols]

 Reference Signs List 10 automatic transmission 30 transmission gear mechanism 51 forward clutch 123 third duty solenoid valve 300 controller 303 engine rotation sensor 306 oil temperature sensor

 ────────────────────────────────────────────────── ─── Continued on the front page (72) Inventor Kenji Sawa 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Mazda Co., Ltd.

Claims (3)

[Claims] 1. An engine output detecting means for detecting a value related to an output torque of an engine, an oil pump driven by the engine to generate a hydraulic pressure, and a hydraulic pressure generated by the oil pump being detected by the engine output detecting means. Hydraulic pressure adjusting means for adjusting to a predetermined target hydraulic pressure determined based on the hydraulic pressure, and a hydraulic pressure adjusted to the target hydraulic pressure by the hydraulic pressure adjusting means to generate a working pressure and supply the working pressure to the friction element Means, and the operating pressure supply means is provided with:
A control device for an automatic transmission, configured to perform a precharge operation of supplying a hydraulic pressure adjusted to a target hydraulic pressure as it is as a working pressure to a friction element for a predetermined time at a predetermined time, wherein the rotation speed of the oil pump is Oil pump rotational speed detecting means for detecting a value related to the oil pressure, upper limit hydraulic pressure determining means for obtaining a maximum value of the hydraulic pressure that can be generated by the oil pump based on the detection result of the detecting means, At least using the value of the operating pressure supplied to the friction element at the time of precharging, and the precharge time setting means is configured to set the operating pressure used for setting the precharge time. A control device for an automatic transmission, characterized in that a smaller one of the target oil pressure and the upper limit oil pressure is used as the value. 2. The operating pressure supply means precharges a friction element that is engaged when switching from a non-traveling shift range to a traveling shift range at an early stage of the switching, and sets the target oil pressure to the engine rotation speed. The control device for an automatic transmission according to claim 1, wherein the control device is set based on the number and the throttle opening. 3. An oil temperature detecting means for detecting the temperature of the hydraulic oil, wherein the actual oil pressure detecting means has a maximum value of the hydraulic pressure which can be generated by the hydraulic pressure generating means when the oil temperature is high as compared with when the oil temperature is low. The control device for an automatic transmission according to claim 1 or 2, wherein is set to a small value.