JPH0754111B2 - Ignition timing control method for vehicle equipped with automatic transmission - Google Patents

Description

TECHNICAL FIELD The present invention relates to an ignition timing control method for a vehicle provided with an automatic transmission, and more particularly to a technique for reducing shift shock of the automatic transmission.

Conventional technology and its problems In a vehicle equipped with an automatic transmission, a shift control method is proposed in which the engine output torque is temporarily reduced by retarding the ignition timing of the engine during the shift to suppress shift shock. Has been done. For example, JP-A-59-1260
The shift control method described in Japanese Patent Publication No. 72 is that.

However, having a main transmission and an auxiliary transmission in series,
BACKGROUND ART There is known a vehicle including an automatic transmission whose sub-transmission is automatically switched to a predetermined gear according to the running state of the vehicle. For example, a vehicle equipped with a so-called automatic transmission composed of three speed change parts and an OD part that is selectively switched between a high speed position and a low speed position, and a stepped automatic transmission or auxiliary transmission after the continuously variable transmission. When the above shift control method is applied to a vehicle (described in Japanese Patent Laid-Open No. 60-37455), the ignition timing retard amount during shift is determined without considering the gear ratio of the main transmission. Therefore, when the torque input to the sub transmission and the inertia torque of the preceding stage change as the gear ratio of the main transmission changes, there is a drawback that the retard amount does not reach an optimum value. If the retard amount is too small, the shift shock accompanying the shift of the auxiliary transmission cannot be sufficiently suppressed, and if the retard amount is too large, there is a disadvantage that the suspension of acceleration feeling becomes large.

Means for Solving the Problems The present invention has been made in the background of the above circumstances,
The main point of this is to ignite a vehicle equipped with an automatic transmission that has a main transmission and an auxiliary transmission in series and that automatically switches the auxiliary transmission to a predetermined gear according to the running state of the vehicle. A timing control method, wherein a retard amount of an ignition timing of an engine of the vehicle at the time of shifting of the sub transmission is calculated based on a gear ratio of the main transmission from a relationship stored in advance, and the sub transmission is used. When the gear position is switched, the ignition timing of the engine of the vehicle is delayed by the retard amount to reduce the output of the engine.

With this configuration, when shifting the auxiliary transmission in the automatic transmission, when the ignition timing of the engine of the vehicle is retarded and the output torque of the engine during shifting is temporarily reduced,
Since the ignition timing is retarded by the retard amount calculated based on the gear ratio of the main transmission from the relationship stored in advance and the output torque of the engine is reduced, the gear ratio of the main transmission is different. Since the engine output torque can be appropriately reduced according to the accompanying change in the input torque of the auxiliary transmission and the inertia torque of the preceding stage, the amount of decrease in the output torque of the engine due to the retard angle can be adjusted appropriately regardless of the gear ratio of the main transmission. can get. Therefore, the shift shock and the interruption of the feeling of acceleration associated with the shift of the auxiliary transmission can be suitably eliminated.

Here, preferably, the retard amount is calculated to be a larger value as the gear ratio of the main transmission is larger. Further, preferably, the retard amount is calculated based on the load of the engine, that is, the throttle valve opening degree or the accelerator pedal operation amount. With this configuration, the amount of decrease in the engine output during shifting is controlled more appropriately.

Further, preferably, the main transmission is a continuously variable transmission that continuously changes the rotation of the engine, and the auxiliary transmission is provided at a rear stage of the continuously variable transmission to drive the vehicle. It is a stepped transmission that is automatically switched to a predetermined gear according to the state.

Examples Hereinafter, one application example of the present invention will be described in detail with reference to the drawings.

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

The fluid coupling 10 is fixed to the pump 24 connected to the crankshaft 22 of the engine 8 and the input shaft 26 of the CVT 12, and is connected to the pump 2
A turbine 28 that is rotated by oil from 4 and a lockup clutch 32 that is fixed to the input shaft 26 via a damper 30 are provided. The lockup clutch 32 may be, for example, a vehicle speed or an engine speed or a turbine 28.
When the rotation speed of the crankshaft reaches a predetermined value or higher, the crankshaft 22 and the input shaft 26 are directly connected to each other.

The CVT 12 includes variable pulleys 36 and 38 provided on the input shaft 26 and the output shaft 34, respectively, and the variable pulleys 36 and 38.
And a conduction belt 40 wound around 38. The variable pulleys 36 and 38 are provided on the fixed rotating bodies 42 and 44 fixed to the input shaft 26 and the output shaft 34, respectively, and on the input shaft 26 and the output shaft 34 so as to be movable in the axial direction and non-rotatable around the shaft. The movable rotors 46 and 48 are moved by the hydraulic cylinders 50 and 52 to change the width of the V groove, that is, the hanging diameter (effective diameter) of the conduction belt 40. Gear ratio γ (= rotation speed N _in of input shaft 26 / rotation speed N of output shaft 34
_out ) is supposed to be changed. Hydraulic cylinder 50
Is operated exclusively for changing the gear ratio γ, and the hydraulic cylinder 52 is operated exclusively so as to obtain the minimum clamping pressure within a range where the slippage of the transmission belt 40 does not occur. In addition,
The oil pump 54 constitutes a hydraulic pressure source of a hydraulic control device which will be described later, and is connected to the crankshaft 22 by a connecting shaft (not shown) longitudinally passing through the input shaft 26 and is constantly driven by the engine 8 to rotate.

The sub-transmission 14 is a stepped automatic transmission that is connected in series with the CVT 12 in a subsequent stage and is automatically switched to a high-speed gear stage and a low-speed gear stage according to the running conditions of the vehicle. It is installed coaxially with 34 and contains a Ravigneaux type compound planetary gear unit. This planetary gear device includes a pair of first sun gear 56 and second sun gear 58, and a first planetary gear 60 that meshes with the first sun gear 56.
A second planetary gear 62 meshing with the first planetary gear 60 and the second sun gear 58, a ring gear 64 meshing with the first planetary gear 60, and the first planetary gear 60 and the second planetary gear 62 rotatably supported. It has a carrier 66 and a carrier. The second sun gear 58 is fixed to a shaft 68 integrally connected to the output shaft 34, and the carrier 66 is fixed to an output gear 70. The high speed clutch 72 controls engagement between the shaft 68 and the first sun gear 56, the low speed brake 74 controls engagement of the first sun gear 56 with the housing, and the reverse brake 76 controls the ring gear 64. Control the engagement of the with the housing. FIG. 3 shows the operating state of each friction engagement element of the auxiliary transmission 14 and the reduction ratio in each range. In the figure, ◯ indicates an engaged state, x indicates an open state, and ρ1 and ρ2 are gear ratios defined by the following equation.

ρ1 = Z _s1 / Z _r ρ2 = Z _s2 / Z _r where Z _s1 is the number of teeth of the first sun gear 56 and Z _s2 is the second sun gear.
58 is the number of teeth, and Z _r is the number of teeth of the ring gear 64.

Therefore, at the low speed gear stages in the L and D ranges, the low speed stage brake 74 as the first friction engagement device is actuated and the first sun gear 56 is fixed, so that the power is reduced at the reduction ratio (1 + ρ1 / ρ2). Is transmitted, but L and D
In the high-speed gear stage of the range, the entire planetary gear device rotates integrally as a result of the operation of the high-speed stage clutch 72 serving as the second friction engagement device, whereby power is transmitted at a reduction ratio of 1. Further, in the R range, the ring gear 64 is fixed to the housing by the operation of the reverse brake 76,
Power is transmitted by reverse rotation of the reduction ratio (1-1 / ρ2).

The output gear 70 of the auxiliary transmission 14 is connected to the differential gear 18 via the intermediate gear device 16, and the power of the engine is the differential gear 18.
In FIG. 1, after being distributed to the left and right drive shafts 20, respectively, they are transmitted to the left and right drive wheels.

FIG. 4 shows a hydraulic control circuit for controlling the vehicle power transmission device shown in FIG. The oil pump 54 is a strainer for the hydraulic oil returned to an oil tank (not shown).
It is pressure-fed to the suction line pressure oil passage 82 via 80. The throttle valve 84 generates a throttle pressure P _th corresponding to the throttle valve opening θ at its output port. The spool 88 of the throttle valve 84 receives an acting force that increases as the throttle valve opening θ increases from the throttle cam 90 that rotates together with the throttle valve and a throttle pressure P _th as a feedback pressure from the control port 92 in the opposite direction. Controls opening and closing of the line pressure oil passage 82 and the output port 86. The manual valve 94 is axially positioned in relation to shift lever L (low), S (second), D (drive), N (neutral), R (reverse), and P (parking) range operations, Sub-primary valve 25 described later
The second line pressure Pl2 output from the output port 258 of 4
In the R range, the reverse brake 76 through port 96
To the hydraulic actuator 76 'for operating the valve, to the port 98 for the L range and the S range, and to the ports 98 and 100 for the D range. The relief valve 102 has a function as a safety valve that releases the oil in the line oil passage 82 when the first line pressure Pl1 in the line pressure oil passage 82 becomes equal to or higher than a predetermined value.

The secondary hydraulic oil passage 104 is connected to the line pressure oil passage 82 via the orifice 106 and the port 110 for discharging the excess oil of the primary regulator valve 108, and the secondary regulator valve 112 is connected to the secondary hydraulic oil passage 112 via the orifice 114. Oil passage
The control chamber 116 is connected to 104, and the connection between the secondary hydraulic fluid passage 104 and the port 120 is controlled in relation to the hydraulic pressure of the control chamber 116 and the load of the spring 118 to control the secondary hydraulic fluid passage 104. Secondary hydraulic Pz
Is maintained at a predetermined value. The lubricating oil passage 122 is connected to the secondary hydraulic oil passage 104 via the port 120 and the orifice 124. The lockup control valve 126 selectively connects the secondary hydraulic oil passage 104 to the engagement side and the release side of the lockup clutch 32 in the fluid coupling 10. The lock-up solenoid valve 128 controls opening / closing of the lock-up control valve 126 between the control chamber 130 and the drain 132, and when the solenoid valve 128 is off (non-excited state), the release side of the lock-up clutch 32. The secondary hydraulic pressure Pz from the secondary hydraulic oil passage 104 is transmitted to the power of the fluid coupling 1
It is transmitted through the fluid in 0. However, when the solenoid valve 128 is on (excited state), the secondary hydraulic pressure Pz from the secondary hydraulic oil passage 104 is supplied to the engagement side of the lockup clutch 32 and the oil cooler 134, and the power is supplied to the lockup clutch 32.
Is transmitted through. The cooler bypass valve 136 controls the cooler pressure.

The gear ratio control device includes a shift direction switching valve device 138 including a first spool valve 142 and a first electromagnetic valve 144, and a speed change speed switching valve device including a second spool valve 146 and a second electromagnetic valve 148.
It has 140. While the first solenoid valve 144 is off, the spool of the first spool valve 142 is pressed toward the spring 152 by the secondary hydraulic pressure Pz of the chamber 150, and the first line pressure Pl1 of the port 154 is the first spool valve. It is sent to the port 158 of the second spool valve 146 via the port 156 of 142, and the connection between the port 160 and the drain 162 is broken. As a result, the gear ratio γ is switched in the decreasing direction. While the first solenoid valve 144 is on, the hydraulic pressure in the chamber 150 is discharged through the drain 164 of the first solenoid valve 144, and the spool of the first spool valve 142 is spring 1
Pressed towards chamber 150 by 52, line pressure on port 156
Pl1 does not occur and port 160 is connected to drain 162.
As a result, the gear ratio is switched in the increasing direction.

While the second solenoid valve 148 is off, the spool of the second spool valve 146 is pressed toward the spring 168 by the secondary hydraulic pressure Pz of the chamber 166, and the connection between the port 158 and the port 170 is cut off.
172 is connected to port 174. Ports 170 and 172 are CVT
It is connected to 12 input side hydraulic cylinders 50. While the second solenoid valve 148 is on, the hydraulic pressure in the chamber 166 is the second solenoid valve 148.
Of the second spool valve 146 is urged toward the chamber 166 by the spring 168 and the port 158
Is connected to port 170, and the connection between ports 172 and 174 is broken. Port 174 is port 160 through oil passage 180
Connected to. The orifice 182 guides a small amount of oil from the port 158 to the port 170 when the second solenoid valve 148 is off.
Therefore, the first solenoid valve 144 is off and the second solenoid valve 148 is
While is on, oil is rapidly supplied to the input side hydraulic cylinder 50 of the CVT 12, and the gear ratio γ rapidly decreases. While the first solenoid valve 144 is off and the second solenoid valve 148 is off, oil is supplied to the input side hydraulic cylinder 50 of the CVT 12 via the orifice 182, and the gear ratio γ of the CVT 12 is changed.
Becomes gradually smaller. When the first solenoid valve 144 is on and the second solenoid valve 148 is on, oil is not supplied to or discharged from the input side hydraulic cylinder 50 of the CVT 12, and the gear ratio γ of the CVT 12 is the leakage from the hydraulic cylinder 50. It gradually increases in accordance with the above. While the first solenoid valve 144 is on and the second solenoid valve 148 is off, the oil in the input hydraulic cylinder 50 is discharged from the drain 162, so that the gear ratio γ of the CVT 12 increases rapidly.

The gear ratio detection valve 184 is provided with a rod 194 slidably contacting the input side movable rotary body 46, and the rod 194 is moved in the axial direction by a displacement equal to the axial displacement of the movable rotary body 46. . The gear ratio detection valve 184 is a fixed rotating body on the input side of the CVT 12.
Since the discharge flow rate of the oil supplied through the orifice 218 increases as the displacement amount of the movable rotator 46 with respect to 42 increases, the gear ratio specific pressure Pr of the output port 216 becomes equal to the gear ratio γ.
Decreases with increasing. Gear ratio Pr is output port 216
It is generated by controlling the discharge amount of the hydraulic medium supplied to.

The cut-off valve 226 includes a chamber 230 communicating with the control chamber 130 of the lockup control valve 126 via an oil passage 228, and a spool that moves in relation to the hydraulic pressure in the chamber 230 and the spring force of the spring 232. 234 and the solenoid valve 128 is off, that is, when the lock-up clutch 32 is in the released state (when shifting is performed in the auxiliary transmission 14, the lock-up clutch is used to absorb the impact of the power transmission system. 32 is released), and closed to prevent transmission of the transmission gear ratio Pr to the primary regulator valve 108.

The primary regulator valve 108 as the first line pressure generating means has a port 23 to which the throttle pressure P _th is supplied.
6, Port 238 to which the transmission gear ratio Pr is supplied, line pressure oil passage 82
Port 240 connected to the oil pump 54, the port 242 connected to the suction side of the oil pump 54, and the port 246 supplied with the first line pressure Pl1 via the orifice 244, the port 240 being moved in the axial direction. Spool 248 that controls the connection between port 242 and port 242, spool 250 that receives throttle pressure P _th and urges spool 248 toward port 238, and spool 2
A spring 252 is provided which biases 48 toward port 238. The pressure receiving areas of the two lands from the bottom of the spool 248 are A
1, A2, the pressure receiving area of the land of the spool 250 that receives the throttle pressure P _th are A3, and the acting force of the spring 252 is W1, the following equations (1) and (2) are established.

When the cutoff valve 226 is open and the transmission specific pressure Pr is coming to the port 238, Pl1 = (A3 ・ P _th ＋ W1−A1 ・ Pr) / (A2−A1) (1) Cutoff When the valve 226 is closed and the transmission specific pressure Pr does not come to the port 238, Pl1 = (A3 · P _th + W1) / (A2-A1) (2) Sub-primary as second line pressure generating means valve
254 is an input port 256 through which the first line pressure Pl1 is introduced, an output port 258 that generates a second line pressure Pl2, and a gear ratio specific pressure Pr.
Port 260 through which the second line pressure Pl2 as the feedback pressure is introduced through the orifice 262
4, spool 266 for controlling opening / closing of input port 256 and output port 258, port 268 to which throttle pressure P _th is guided, and spool 266 which receives throttle pressure P _th from port 268
Has a spool 270 that biases the spool toward the port 260, and a spring 272 that biases the spool 266 toward the port 260. The pressure receiving area of the two lands from the bottom of the spool 266 is B1,
When the pressure receiving area of the land of the spool 270 that receives B2 and the throttle pressure _Pth is defined as B3 and the elastic force of the spring 272 is defined as W2, respectively, the second line pressure Pl2 is output according to the following equation (3).

Pl2 = (B3 · P _th ＋ W2−B1 · Pr) / (B2−B1) ··· (3) The shift valve 274 includes a high speed clutch 72 and a low speed brake 74 of the auxiliary transmission 14. Which selectively actuates the hydraulic pressure in the hydraulic actuators 72 'and 74' for operating the input port 276 and the output port to which the second line pressure Pl2 is introduced during the D, S, L ranges of the shift lever. 278,
280, a port 288 having an orifice 282 and connected to a discharge oil passage 286 ending at the drain 284, a control port 300 to which the first line pressure Pl1 is supplied from the port 100 of the manual valve 94 in the D range, etc. Control port
302, 304, a drain 306, a spool 308, and a spring 310 biasing the spool 308 toward the control port 304. The secondary hydraulic pressure Pz is guided to the control ports 302 and 304 via the orifice 312, and the hydraulic pressure of the control ports 302 and 304 is controlled by the solenoid valve 314 for shifting. The pressure receiving areas of the two lands from the bottom of the spool 308 are S1 and S2, respectively.
<S2. Further, the on / off of the solenoid valve 314 is controlled in relation to the driving parameter of the vehicle.

When the spool 308 is in the position of the spring 310 side, the input port
276 is connected to output port 278 and output port 280 is connected to port 288. Therefore, the second line pressure Pl2 from the output port 278 causes the accumulator having the piston 318.
The hydraulic pressure is supplied to the hydraulic actuator 320 for 320 and the high speed gear, and the hydraulic actuator 74 'for the low speed gear is exhausted, so that the auxiliary transmission 14 becomes the high speed gear.

When the spool 308 is located on the control port 304 side, the input port 276 is connected to the output port 280 and the output port 278
Is connected to the drain 306. Therefore, output port 2
The second line pressure Pl2 from 80 is supplied to the hydraulic actuator 74 'for the low speed stage, and the hydraulic actuator 72' for the high speed stage is exhausted, so that the auxiliary transmission 14 becomes the low speed gear stage.

In the L range and the S range, since the first line pressure Pl1 is not guided to the control port 300, when the solenoid valve 314 is turned off, the spool 308 initially operates the secondary hydraulic pressure that acts on the land of the pressure receiving area S2. The secondary hydraulic pressure Pz acting on the land of the pressure receiving area S1 by Pz moves to the spring 310 side, but when the solenoid valve 314 is turned on, the hydraulic pressure of the control ports 302, 304 decreases, so that the spool 308 moves to the spring 310. It moves to the control port 304 side according to the urging force of. Therefore, in the L range and the S range, switching between the high speed gear stage and the low speed gear stage of the auxiliary transmission 14 is performed in response to ON / OFF of the solenoid valve 314.

Since the first line pressure Pl1 is guided to the control port 300 in the D range, once the spool 308 reaches the position on the spring 310 side, the first line pressure Pl1 from the control port 300 acts on the land of the pressure receiving area S2. The spool 308 is held at the position on the spring 310 side regardless of whether the solenoid valve 314 is turned on or off thereafter. Therefore, the sub transmission 14 is held in the high speed gear stage.

The shift timing valve 324 has a control port 326 that communicates with a hydraulic actuator 72 ′ for a high speed stage, and a spool 328 whose axial position is controlled by the hydraulic pressure of the control port 326, and shifts from a low speed gear stage to a high speed gear stage. Control the flow rate of oil supplied to the hydraulic actuator 72 'for the high speed stage and the discharge amount of oil from the hydraulic actuator 74' for the low speed stage at the time of the upshift.

FIG. 5 shows an electronic circuit that controls the operation of the hydraulic control device described above. A CVT control computer 330 including a CPU, a RAM, a ROM, etc. is provided with a throttle valve opening θ, a rotation speed N _out of the output shaft 34 of the CVT 12 (a sub transmission 14
The input speed of the input side rotary shaft n _in ), the rotation speed N _in of the input shaft 26 of the CVT 12 and the shift lever operating position are respectively supplied. The CPU in the CVT control computer 330 is RAM
In order to control the lockup clutch 32, the gear ratio of the CVT 12, and the gear stage of the auxiliary transmission 14, the solenoid valves 128, 144, 148, 314 are used to process the input signal according to a program stored in advance in the ROM while utilizing the temporary storage function of the. A signal for driving is output via the amplifying device 332.

In the CVT control computer 330, by executing a main routine (not shown), the electronic control unit is initialized and the input signals and the like from each sensor are read, based on the read signal. Relationships calculated in advance based on the diagnosis, the vehicle speed V, and the throttle valve opening θ for diagnosing whether the engine, CVT 12, etc. are operating normally according to the vehicle speed V and the input signal conditions. From the lockup control for controlling the solenoid valve 128 that operates the lockup clutch 32 from the vehicle speed V, the throttle valve opening θ, the gear ratio, and the like, the gear position of the auxiliary transmission 14 is set to the high speed gear position and the low speed gear position. Shift control that automatically switches to one of the gears, and
A gear ratio control for changing the gear ratio of the CVT 12 to an optimum value is sequentially or selectively repeated.

Further, the CVT control computer 330 outputs a gear ratio signal SSR representing the gear ratio γ of the CVT 12 and a gear stage signal SGW representing the current gear stage of the auxiliary transmission 14 to the engine control computer 334. Engine control computer 334 is CPU, RA
M, ROM, etc. are provided, and signals indicating throttle valve opening θ, intake air amount of engine 8, crankshaft rotation angle, ignition cylinder, cooling water temperature of engine 8, atmospheric pressure, etc. are supplied from sensors not shown. Has been done. The CPU in the engine control computer 334 processes the input signal in accordance with a fuel control program and an ignition timing control program (not shown) stored in advance in the ROM while utilizing the storage function of the RAM, and determines the air-fuel ratio and the ignition timing of the air-fuel mixture. A fuel injection valve provided in the engine 8 for optimally maintaining the operating condition of the vehicle.
The fuel injection signal SFF and the ignition signal SEE are output to the 336 and the ignition device 338.

Hereinafter, in the engine control computer 334,
The operation of the ignition timing control associated with the shift of the auxiliary transmission 14 will be described with reference to the flowchart of FIG.

First, by executing step S1, input signals such as the gear ratio signal SSR, the gear stage signal SGW, and a signal indicating the throttle valve opening θ are read. Next, at step S2, it is determined whether or not there is a shift change based on the gear stage signal SGW, that is, whether the gear stage of the auxiliary transmission 14 has been switched from the low speed gear stage to the high speed gear stage or from the high speed gear stage to the low speed gear stage. It is determined whether or not. If this judgment is denied, the following steps are skipped, but if affirmed, the following step S3 is executed. In step S3, the content of the timer T is cleared to zero. In the following step S4, the retardation amount δ from the basic ignition timing is determined based on the actual gear ratio γ and the throttle valve opening θ of the CVT 12, which is the main transmission, from the previously determined relationship shown in FIG. 6, for example. To be done. The relationship shown in FIG.
It is an experimentally obtained in advance so as not to cause a shift shock at the time of shift-up of 14 and to prevent the ignition timing of the engine 8 from being unnecessarily retarded, and is a ROM in the form of a data map or a functional formula. Is stored in
The larger the actual gear ratio γ and throttle valve opening θ, the larger the retard amount δ.

In step S5, it is determined whether or not the content of the timer T has reached a predetermined constant value T ₁ , and if not, step S5 is repeatedly executed, but if it is determined that it has been reached. The next step S6 is executed. This value T ₁
Corresponds to a delay time from when the shift-up of the auxiliary transmission 14 is instructed to when the retardation of the ignition timing for alleviating the shift shock is started. In step S6, the basic ignition timing is changed (corrected) so that the ignition timing is retarded corresponding to the retard amount δ determined in step S4. Then, step S7 is executed, whether the content of the timer T has reached a predetermined value T ₂ which is determined in advance is determined. This constant value T ₂ corresponds to the time during which the ignition timing retard control is performed when the auxiliary transmission 14 is shifted up. Then, when the determination in step S7 is denied,
If the affirmative determination is made, the retard control of the ignition timing is stopped in step S8 and then the ignition timing retard control routine for shifting shock mitigation is ended.

As described above, according to the present embodiment, when the auxiliary transmission 14 is automatically shifted according to the running state of the vehicle, the auxiliary transmission 14 is shifted up for a certain period (time corresponding to T ₂ −T ₁ ) of FIG. As shown by the broken line in (b), the ignition timing is retarded. Since the ignition timing of the engine 8 and the output torque have a general relationship shown in FIG. 7, the output torque of the engine 8 is changed as the auxiliary transmission 14 is switched from the low speed gear stage to the high speed gear stage. It is suppressed appropriately. At the time of this switching, the gear train for establishing the high speed gear stage is newly established after the gear train for establishing the low speed gear stage is released, but the output torque of the engine 8 is suppressed as described above. , The input side torque of the auxiliary transmission 14 is made smaller than before, and the output side member (output gear 7
The torque newly transmitted to 0) and the inertia torque of the preceding stage are reduced, and the shift shock is appropriately mitigated.
In particular, in the vehicle having the sub-transmission 14 at the rear stage of the CVT 12 which is the main transmission as in the present embodiment, the input torque of the sub-transmission 14 and the inertia of the preceding stage as the transmission ratio γ of the CVT 12 increases. Since the torque increases, the shift shock accompanying the shift of the auxiliary transmission 14 tends to be large and vary. However, according to the present embodiment, the actual gear ratio of the CVT 12 is changed from the pre-stored relationship shown in FIG. Since the output torque of the engine is appropriately reduced by retarding the ignition timing by the retard amount δ calculated based on γ, the input of the auxiliary transmission accompanying the change in the gear ratio of the main transmission is different. Since the engine output torque is reduced according to the change in the torque and the inertia torque in the preceding stage, the engine output torque reduction amount due to the ignition timing retard is appropriately obtained regardless of the gear ratio γ of the CVT 12. Therefore, the shift shock and the interruption of the feeling of acceleration associated with the shift of the auxiliary transmission 14 can be suitably eliminated. In addition, the load applied to the friction engagement device at the time of shifting the auxiliary transmission is reduced, and its durability is improved. The broken line in FIG. 8 (a) shows the torque at the time of upshifting of the output side member of the auxiliary transmission 14, and this torque change is significantly smaller than that in the conventional case shown by the solid line, and the shift shock does not occur. It shows that it has been relaxed. In the figure, the decrease in output torque before retarding the ignition timing is indicated by the accumulator.
This is because the preset pressure of 320 has been lowered in advance compared to the conventional one. That is, in this embodiment, the auxiliary transmission
Since the output torque of 14 is determined by the torque capacity of the high speed clutch 72, the engine 8 is delayed by the ignition timing so that the speed can be changed smoothly even if the torque capacity is reduced in advance.
The output torque of is reduced.

Although one application example of the present invention has been described above, the present invention is also applied to other aspects.

For example, in the above-described embodiment, the retard amount of the ignition timing is obtained based on the actual throttle valve opening θ and the gear ratio γ from the relationship shown in FIG. 6, but the actual throttle valve opening θ and the gear ratio It may be obtained based on one of γ, or even if the retardation amount δ is a constant value, a temporary effect can be obtained.

Further, in place of the actual throttle valve opening θ, an amount representing the engine load such as the accelerator pedal operation amount or the intake pipe negative pressure may be used. With this configuration, the present invention is preferably applied even when the diesel engine that does not use the throttle valve is mounted on the vehicle. The ignition timing in this case is adjusted by the fuel injection timing.

Further, in the above-described embodiment, the vehicle provided with the CVT 12 and the auxiliary transmission 14 is described, but an automatic transmission that automatically shifts to multiple gears, that is, a vehicle equipped with a so-called automatic transmission. The present invention is also applied to.

Further, in the flowchart of FIG. 1, the ignition timing retard control for alleviating the shift shock is executed under the condition that the auxiliary transmission 14 is upshifted, but the vehicle acceleration is set to a predetermined constant value. May be added to the condition for executing the ignition timing retard control. By doing so, there is an advantage that the ignition timing retard control in the region where the shift shock is originally small and is not a problem can be eliminated.

Further, in the above-described embodiment, when the auxiliary transmission 14 is upshifted, the output torque of the engine 8 is suppressed by retarding the ignition timing. The output torque of the engine 8 may be suppressed by being limited.

It should be noted that the above is merely an application example of the present invention, and the present invention can be variously modified without departing from the spirit thereof.

[Brief description of drawings]

FIG. 1 is a flow chart for explaining an essential part of the operation of the device to which the present invention is applied. FIG. 2 is a skeleton diagram showing a power transmission device of a vehicle to which the present invention is applied. Figure 3 is second
It is a figure which shows the relationship of the range of a sub transmission and the friction engagement device in the apparatus of a figure. FIG. 4 is a circuit diagram showing in detail a hydraulic control device for operating the device of FIG. Fifth
The figure is a block diagram showing the electrical control circuit provided in the device of FIG. FIG. 6 is a diagram showing the relationship used in the flowchart of FIG. FIG. 7 is a diagram showing a general relationship between engine ignition timing and output torque.
FIG. 8 is a time chart explaining the operation of the apparatus shown in FIG. 12: Belt type continuously variable transmission 14: Auxiliary transmission δ: Delay amount γ: Gear ratio θ: Throttle valve opening

Claims (4)

[Claims] 1. An ignition timing control for a vehicle, comprising an automatic transmission that has a main transmission and an auxiliary transmission in series, and the auxiliary transmission is automatically switched to a predetermined gear according to the running state of the vehicle. A method of calculating the retard amount of the ignition timing of the engine of the vehicle at the time of gear shifting of the sub-transmission from a prestored relationship based on the gear ratio of the main transmission, An ignition timing control method for a vehicle having an automatic transmission, wherein the ignition timing of the engine of the vehicle is delayed by the retard amount to reduce the output of the engine when the gears are switched. 2. The ignition timing control method for a vehicle according to claim 1, wherein the retard amount is calculated to have a larger value as the gear ratio of the main transmission increases. 3. The ignition timing control method for a vehicle according to claim 1, wherein the retard amount is calculated based on the load of the engine. 4. The main transmission is a continuously variable transmission that continuously changes the rotation of the engine, and the auxiliary transmission is provided at a rear stage of the continuously variable transmission, and a running state of a vehicle. The ignition timing control method for a vehicle according to any one of claims 1 to 3, which is a stepped transmission that is automatically switched to a predetermined gear according to the above.