JPS6258053A - Ignition timing control method for vehicle having automatic transmission - Google Patents

Abstract

PURPOSE:To lighten a shock in changing a speed, by retarding an ignition timing of an engine in changing a gear stage of an automatic transmission, and thereby reducing an output of the engine. CONSTITUTION:A power of an engine 8 is transmitted through a belt-type continuous variable transmission 12 and a gear-type auxiliary transmission 14, etc. to driving wheels. In controlling an ignition timing in association with a speed change of the auxiliary transmission 14 by means of an engine control computer 334, it is first determined according to a gear stage signal SGW whether or not a shift change has been carried out. If YES, a retard quantity from a fundamental ignition timing is decided by using a data map or the like according to an actual speed change ratio and a throttle opening degree. Next, after a predetermined delay time has been elapsed, the fundamental ignition timing is so corrected as to effect retarding of an ignition timing corresponding to the above-mentioned retard quantity for a predetermined period of time, thus reducing an output of the engine.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to an ignition timing control method for a vehicle equipped with an automatic transmission, and more particularly to a technique for reducing shift shock of the automatic transmission.

BACKGROUND ART Vehicles are known that are equipped with an automatic transmission that automatically switches to a predetermined gear according to the driving condition. These include vehicles equipped with so-called automatic ink transmissions, and vehicles equipped with a stepped automatic transmission, that is, an auxiliary transmission, after a continuously variable transmission (for example, as described in Japanese Patent Application Laid-Open No. 60-37455). . In such vehicles, when an automatic transmission performs a gear change operation such as an upshift during heavy vehicle acceleration, a sudden acceleration sequence following a temporary deceleration occurs as the gear train that makes up the gear changes. It was easy to get stuck, and drivability sometimes deteriorated.

In particular, in a vehicle equipped with an auxiliary transmission downstream of a continuously variable transmission, as the gear ratio (manpower shaft rotation speed/output shaft rotation speed) of the continuously variable transmission increases, the auxiliary change i! As the torque input to the machine and the inertia torque of the preceding stage increase, the shift shock that accompanies the shift of the sub-transmission tends to increase and fluctuate, and at the same time, the shift shock that is applied to the friction engagement device when the sub-transmission shifts increases. In some cases, the load became so large that durability could not be obtained.

Means for Solving the Problem The present invention has been made against the background of the above circumstances.
The gist of this is an ignition timing control method for a vehicle equipped with an automatic transmission that automatically switches to a predetermined gear according to the driving condition, and when changing the gear of the automatic transmission, The aim is to delay the ignition timing of the engine to reduce the output of the engine.

Operation and Effects of the Invention With this method, when changing gears of an automatic transmission such as upshifting or downshifting while the vehicle is accelerating,
Since the ignition timing of the vehicle's engine is delayed and the engine output is reduced, the torque input to the auxiliary transmission upon gear train switching is reduced, and shift shock is alleviated. In particular, in vehicles equipped with an auxiliary transmission after the continuously variable transmission, the gear ratio of the continuously variable transmission (
As the input shaft rotational speed/output shaft rotational speed) increases, the torque input to the auxiliary transmission and the inertia torque of the preceding stage thereof increase, so there is a tendency for shift shocks associated with shifting of the auxiliary transmission to increase and to fluctuate. However, according to the present invention, this problem can be effectively resolved and drivability can be improved.

In addition, the load applied to the frictional engagement device during shifting of the auxiliary transmission is reduced, and its durability is increased.

Preferably, the vehicle is equipped with a continuously variable transmission that continuously changes the rotation of the engine, and the automatic transmission is not provided downstream of the continuously variable transmission, and the vehicle The gear is automatically switched to a predetermined gear according to the driving condition of the vehicle. Since the above-mentioned problems are conspicuous in such vehicles, the present invention can be applied more effectively.

The ignition timing is preferably determined from a predetermined relationship,
ii+? +F2 It is delayed in accordance with the delay 'J'l Eit determined based on the accelerator operation amount or throttle valve opening of the vehicle and the gear ratio of the gear transmission. In this way, the engine power output during gear shifting can be adjusted to the optimum value according to the driving condition of the vehicle expressed by the accelerator operation amount or throttle valve opening of the vehicle and the gear ratio of the continuously variable transmission. It has the advantage of being controlled. In addition, the basic ignition time 1s1 determined by the normal ignition timing controller
may be corrected so that it is delayed by the above-mentioned retard amount.

Further, the amount of retardation of the ignition timing is preferably determined to increase as the gear ratio of the continuously variable transmission increases,
In addition, 1. The larger the throttle valve opening or accelerator operation amount of the vehicle is, the larger the retard amount is. , there is an advantage that the amount of retardation is necessary and sufficiently determined.

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

In FIG. 2, the power of an engine (not shown) is transmitted through a fluid coupling 10. Belt type continuously variable transmission (hereinafter referred to as CVT) 12
．． Sub-transmission 14. Intermediate gear device 16 and differential device 1
8 and is transmitted to a drive wheel (not shown) connected to the drive shaft 2o.

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

The CVT 12 includes variable pulleys 36 and 38 provided on the input shaft 26 and output shaft 34, respectively, and a conduction heald 40 wound around the variable pulleys 1136 and 38.
It is equipped with The variable pulleys 36 and 38 are fixed rotating bodies 42 and 44 fixed to the input shaft 26 and the output shaft 34, respectively, and are provided on the input shaft 26 and the output shaft 34 so as to be movable in the axial direction but not relatively rotatable around the axes. When the movable rotors 46 and 48 are moved by the hydraulic cylinders 5o and 52, the groove width, that is, the hanging diameter (effective diameter) of the transmission belt 40 is changed, and the CVT 12 The gear ratio T (-rotational speed N8 of the human power shaft 26, /rotational speed N.ut of the output shaft 34) is changed.

The hydraulic cylinder 50 is operated exclusively to change the gear ratio γ, and the hydraulic cylinder 52 is operated exclusively to obtain the minimum clamping force within a range where the transmission heald 40 does not slip. The oil pump 54 constitutes a hydraulic pressure source of a hydraulic control device, which will be described later, and is connected to the crankshaft 22 via a not-illustrated connecting shaft passing vertically through the input shaft 26, and is constantly rotationally driven by the engine 8.

The auxiliary transmission 14 is a stepped automatic transmission that is connected in series with the CVT 12 at a later stage and automatically switches between a high gear and a low gear according to the driving conditions of the vehicle. It is provided coaxially with the shaft 34 and includes a Ravigneau type compound planetary gear set. f1 star gear 60 and this first planetary gear 60
and a second planetary gear 62 that meshes with the second sun gear 58, a ring gear 64 that meshes with the first planetary gear 60, and a first
A carrier 66 rotatably supports the planetary gear 60 and the second planetary gear 62. The second sun gear 58 is fixed to a shaft 68 that is integrally connected to the output shaft 34, and the carrier 66 is fixed to an output gear 70. The high speed clutch 72 controls engagement between the lever 68 and the first sun gear 56, and the low speed brake 74 controls engagement between the first sun gear 56 and the first sun gear 56.
The reverse brake 76 controls the engagement of the ring gear 64 with the housing. FIG. 3 shows the operating state of each frictional engagement element of the sub-transmission 14 and the reduction ratio in each range. In the figure, ◯ indicates a locked state, X indicates a released state, and ρ1 and ρ2 are gear ratios defined by the following equation.

ρ1 = Zs+/z.

ρ2−Z5□/Z.

However, ZsI is the number of teeth of the first sun gear 56, and Zs2 is the number of teeth of the second sun gear 56.
The number of teeth Zl of the sun gear 58 is one of the ring gear G4.

Therefore, in the low gears in L and D ranges, the first T! l Low speed brake 7 as a friction engagement device
4 is activated and the first sun gear 56 is fixed, power is transmitted at the reduction ratio (1+ρ1/ρ2).
In the high gear stages of the L and D ranges, the entire planetary gear system rotates as a unit by the operation of the high speed clutch 72 as a second frictional engagement device, thereby transmitting power at a reduction ratio of 1. Ru. Furthermore, in the R range, the ring gear 64 is fixed to the housing by the operation of the reverse brake 76, so power is transmitted in reverse rotation at the reduction ratio (1-1/ρ2).

The output gear 70 of the sub-transmission 14 is connected to the differential device 18 via the intermediate gear device 16, and after the engine power is distributed to the left and right drive shafts 20 in the differential device, transmitted to the drive wheels.

FIG. 4 shows a hydraulic control circuit for controlling the vehicle power transmission device shown in FIG. The oil pump 54 pumps hydraulic oil and the like returned into an oil tank (not shown) through a strainer 80 to a suction line pressure oil passage 82 . The throttle nottle valve 84 generates a throttle pressure Pth corresponding to the throttle valve opening θ at its output port 86. The spool 88 of the throttle valve opening 84 receives an acting force that increases as the throttle valve opening θ increases from a throttle cam 90 that rotates together with the throttle valve, and a throttle pressure PLh as a feedback pressure from a control boat 92.
are received in opposite directions, and the line pressure oil passage 82 and the output port 8
Controls opening and closing with 6. The manual valve 94 is in the low position of the shift lever.

S (second), D (drive), neutral to N)
, R (reverse), and P (parking) range operation and output from the output port 258 of the 1 kM subprimary valve 254 to the boat 96 when in the R range. to the hydraulic actuator 76° that operates the reverse brake 76 through the boat 98 when in L and S ranges.
When in the D range, it leads to boats 98 and 100, respectively. The relief valve 102 has a function as a safety valve that releases oil in the line oil passage 82 when the first line pressure PNI of the line pressure oil passage 82 exceeds a predetermined value.

The secondary hydraulic oil passage 104 is connected to the line pressure oil passage 82 via an orifice 106 and a boat 110 from which excess oil from the primary regulator valve 108 is discharged.
The control room 1 is connected to the secondary hydraulic fluid line 104 via
16, and controls the connection between the secondary hydraulic oil passage 104 and the port h 120 in relation to the oil pressure in the control chamber 116 and the load of the spring 118 to predetermine the secondary oil pressure Pz of the secondary hydraulic oil passage 104. Maintain value. Lubricant oil passage 122 is connected to secondary hydraulic oil passage 104 via boat 120 and orifice 124 . The lock-up control valve 126 selectively connects the secondary hydraulic oil passage 104 to the engagement side and release side of the lock-up clutch 32 in the fluid coupling 1o. The lock-up solenoid valve 128 controls opening and closing between the control chamber 130 and the drain 132 of the lock-up control valve 126, and when the solenoid valve 128 is off (non-energized state), the lock-up clutch 32 is opened and closed. The secondary hydraulic pressure Pz from the secondary hydraulic oil passage 104 is transmitted, and power is transmitted via the fluid in the fluid coupling IO. However, fluorite 5'1. When the valve 128 is on (energized state), the secondary hydraulic oil passage 104 is connected to the engagement side of the lock-up clutch 32 and the oil cooler 134.
Secondary hydraulic pressure Pz is supplied from the engine, and power is transmitted via the lock-up clutch 32. Cooler bypass valve 1
36 controls the tala pressure.

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

As a result, the gear ratio is switched to a direction in which the gear ratio is decreased by 1 f)5.

During the period when the first solenoid valve 144 is on, the hydraulic pressure in the chamber 150 is discharged through the drain 164 of the first solenoid valve 144.
The spool of the first spool valve 142 is pushed toward the chamber 150 by a spring 152, and the boat 156 receives line pressure Pj.
21 does not occur and the boat 160 is connected to the drain 162. As a result, the gear ratio is switched in the increasing direction.

During the period when the second solenoid valve 148 is off, the second spool valve 1
The spool 46 is activated by the spring 1 due to the secondary hydraulic pressure P2 in the chamber 166.
68, the connection between boat 158 and boat 170 is severed, and boat 172 is connected to boat 174. The boats 170 and 172 are connected to the input hydraulic cylinder 50 of the CVT 12. During the period when the second solenoid valve 148 is on, the hydraulic pressure in the chamber 166 is discharged from the drain 176 of the second solenoid valve 148, and the spool of the second spool valve 146 is pressed toward the chamber 166 by the spring 168.
Boat 158 is connected to boat 170 and boat 172
The connection between the boat 174 and the boat 174 is severed. Boat 174 is connected to boat 160 via oil line 180. The orifice 182 is connected to the boat 15 when the second solenoid valve 148 is turned off.
8 leads a small amount of oil to the boat 170. Therefore, during the period when the first solenoid valve 144 is off and the second solenoid valve 148 is on, the input hydraulic cylinder 50 of the CVT 12
Oil is quickly supplied to the engine, and the gear ratio T quickly decreases. The first solenoid valve 144 is off and the second solenoid valve 148
During the period when is off, c v 'F12 input (jl, oil is supplied to the hydraulic cylinder 50 through the orifice 182
The gear ratio γ of CV'T'12 gradually decreases. When the first solenoid valve 144 is on and the second solenoid valve 148 is on, oil is not supplied to or discharged from the input hydraulic cylinder 50 of the CVT 12.
The gear ratio T of 2 gradually increases according to leakage from the hydraulic cylinder 50 and the like. When the first solenoid valve 144 is on and the second
During the period when the solenoid valve 148 is off, the input hydraulic cylinder 5
0 oil is drained from train 162, so CV
The gear ratio γ of T12 increases rapidly.

The gear ratio detection valve 184 includes a rod 194 that is in sliding contact with the movable rotary body 46 on the input side, and the rod 194 is moved in the axial direction by an amount of displacement equal to the displacement amount of the movable rotary body 46 in the axial direction. . The gear ratio detection valve 184 is the CVT 12
As the amount of displacement of the movable rotary body 46 relative to the fixed rotary body 42 on the input side increases, the discharge flow rate of the oil supplied through the orifice 218 increases. It decreases as it increases. The gear ratio pressure Pr is generated by controlling the discharge amount of the hydraulic medium supplied to the output port 216.

The cut-off valve 226 is the lock-up control valve 126.
The chamber 2 communicates with the control chamber 130 via an oil passage 228.
30. and a spool 234 that moves in relation to the oil pressure in its chamber 230 and the spring force of a spring 232 when the solenoid valve 128 is off, i.e. when the lock-up clutch 32 is in the released state (auxiliary gearshift When changing gears in the machine 14, the lock-up clutch 32 is released to absorb shocks in the power transmission system), and the lock-up clutch 32 is released to absorb the shock of the power transmission system, and the lock-up clutch 32 is closed to transmit the gear ratio pressure Pr to the primary regulator valve 108. prevent.

The primary regulator valve 108 as a first line pressure generating means includes a port 1-236 to which a throttle pressure Ptl is supplied, and a boat 2 to which a gear ratio pressure Pr is supplied.
38, boat 240 connected to line pressure oil line 82
．． Boat 2 connected to the suction side of the oil pump 54
42. and boat 246 ., which is supplied with first line pressure FAI through orifice 244 . A spool 248 that moves axially to control the connection between boats 240 and 242. The spool 24 receives the throttle pressure P.
Spool 250 biasing 8 toward boat 238 and spring 2 biasing spool 248 toward boat 238
It is equipped with 52. The pressure receiving areas of the two lands from the bottom of the spool 248 are AI, A2, and throttle pressure PL, respectively.
The pressure receiving area of the land of the spool 250 that receives h is A3,
And the acting force of the spring 252 is Wl, then the following equation (1)
And (2) holds true.

When the cut-off valve 226 is open and the gear ratio pressure Pr is coming to the boat 238, Pl-(83・pt+,+wt nb P
r)/ (H2-81) ・ ・ ・ ・ ・ (1) If the cut-off valve 226 is closed and the gear ratio pressure Pr is not coming to the boat 238, +11 - (83 ・ P + Wl) / ( 82-AI
) ・ ・ ・ (2) The sub-primary valve 254 as the second line pressure generating means is connected to the input port 256 . to which the first line pressure Pβ1 is introduced. Output port 258 for generating second line pressure Pβ2. Boat 260 where the gear ratio specific pressure Pr is dedicated. A second line pressure PN2 as a feedhack pressure is introduced through an orifice 262 at the bow 1-264. Input port 1-256 and output port 258
The spool 266° that controls the opening and closing of the boat 268. is subjected to throttle pressure P. The spool 266 is moved to the boat 260 in response to the throttle pressure Pth from the boat 268.
The spool 270. and spool 266
It has a spring 272 that biases the boat 260 toward the boat 260. The pressure receiving area of the two lands from the bottom of the spool 266 is B
1, I32. Spool 270 that receives throttle pressure P
Defining the pressure receiving area of the land as B3 and the elastic force of the spring 272 as W2, the second
line pressure P12 is output.

PA2 = (B3 ・Pth +W2 Bl ・Pr
) / (8281)・・・・・・・(3) The shift valve 274 selects between hydraulic actuators 72° and 74″ that actuate the high speed clutch 72 and the low speed brake 74 of the auxiliary transmission 14. It applies hydraulic pressure to the shift lever, S, L.
Input port 276, output ports 278, 280, orifice 282 to which second line pressure P12 is introduced during range operation
drain oil passage 2 having a drain 284 and terminating in a drain 284;
Boat 288 connected to 86. In the D range, the first line pressure P from the boat 100 of the manual valve 94
61 is supplied, other control ports 302, 304, drain 306, spool 308,
and a spring 310 biasing the spool 308 toward the control port 1-304. control port 302,
A secondary hydraulic pressure Pz is introduced to the control port 304 through an orifice 312, and the hydraulic pressure of the control ports 302 and 304 is controlled by a shift solenoid valve 314. The pressure receiving areas of the two bottom lands of the spool 308 are each Sl.

B2 and St<52. Further, the on/off state of the solenoid valve 314 is controlled in relation to the driving parameters of the vehicle.

When spool 308 is in the spring 310 position, input port 276 is connected to output port 278 and output port 280 is connected to boat 288. Therefore, the second line pressure P/2 is applied to the piston 3 from the output port 278.
18 and the high-speed hydraulic actuator 72°, the pressure inside the low-speed hydraulic actuator 74' is exhausted, and the auxiliary transmission 14 becomes a high gear.

When the spool 308 is in the position on the control boat 304 side, the input port 276 is connected to the output port 280, and the output port 278 is connected to the drain 306. Therefore, the second line pressure Pff2 from the output port 280
is supplied to the low-speed hydraulic actuator 74', and the pressure inside the high-speed hydraulic actuator 72° is exhausted, so that the sub-transmission 14 becomes a low-speed gear.

In the case of L range and S range, the first line pressure P11 is not introduced to the control port 300, so the solenoid valve 3
14 is turned off, the spool 308 is initially operated by the secondary hydraulic pressure Pz acting on the land of the pressure receiving area S2, and then by the secondary hydraulic pressure P2 acting on the land of the pressure receiving area S1.
It moves to the spring 310 side, but when the solenoid valve 314 is turned on, the oil pressure in the control ports 302 and 304 decreases, so
The spool 308 moves toward the control boat 304 according to the biasing force of the spring 310. Therefore, L range and S
In the range, the auxiliary transmission 14 is switched between a high gear and a low gear in response to turning on and off the solenoid valve 314.

In the D range, the first line pressure P11 is applied to the control port 300.
is guided, so once the spool 308 is in the position on the spring 310 side, the control port 30 is connected to the land of the pressure receiving area S2.
The first line pressure P61 from 0 is applied, and the spool 30B is held at the position of the spring 310 regardless of whether the solenoid valve 314 is turned on or off thereafter. Therefore, sub-transmission 1
4 is held in high gear.

The shift timing valve 324 has a control port 326 that communicates with the hydraulic actuator 72' for the high speed gear, and a spool 328 whose axial position is controlled by the hydraulic pressure of the control port 326, and the shift timing valve 324 has a spool 328 whose axial position is controlled by the hydraulic pressure of the control port 326. The flow rate of oil supplied to the hydraulic actuator 72' for the high speed gear and the amount of oil discharged from the hydraulic actuator 74° for the low speed gear are controlled during an upshift.

FIG. 5 shows an electronic circuit that controls the operation of the hydraulic control device described above. The CVT control computer 330, which includes a CPtJ, RAM, ROM, etc., receives information about the throttle valve opening θ from a sensor (not shown).

Rotational speed N0ut of the output shaft 34 of the CVT 12 (rotational speed n3° of the input side rotating shaft of the sub-transmission I4), rotational speed N in of the input shaft 26 of the CVT 12 + shift lever
A signal representing each operating position is supplied.

The CPU in the CVT control combiator 330 utilizes the temporary storage function of the RAM and processes input signals according to a program pre-recorded in the ROM, and changes the lock-up clutch 32, the gear ratio of the CVT 12, and the gear position of the auxiliary transmission 814. In order to control the solenoid valves 128, 144, 148
, 314 are outputted via amplifiers 332, respectively.

In the CVT control computer 330, a main routine (not shown) is executed to initialize the electronic control device and read input signals from each sensor. Vehicle speed etc. are calculated, and according to input signal conditions, diagnosis is performed to diagnose whether the engine, CVT 12, etc. are operating normally, and is determined in advance based on vehicle speed V and throttle valve opening θ. Lock-up control for controlling the solenoid valve 128 that operates the lock-up clutch 32 from the relationship, vehicle speed ■.

Shift control that automatically switches the gear of the sub-transmission 14 to either a high gear or a low gear based on the throttle valve opening θ7 gear ratio, etc., and a shift that changes the gear ratio of the CVT 12 to an optimal value. Ratio control etc. are repeatedly executed sequentially or selectively.

Further, the CV T fi control computer 330 outputs a speed ratio signal SSR representing the speed ratio of the CVT 12 and a gear position signal SGW representing the current gear position of the sub-transmission 14 to the engine control computer 334. The engine control computer 334 includes a CPU, RAM, ROM, etc., and detects the throttle valve opening θ from a sensor (not shown).
Signals representing the intake air amount of the engine 8, the rotation angle of the crankshaft, the ignition cylinder, the cooling water temperature of the engine 8, the atmospheric pressure, etc. are supplied. The CPU in the engine control computer 334 utilizes the memory function of the RAM and processes input signals according to a fuel control program (not shown) and an ignition control program (not shown) stored in advance in the ROM, and determines the air-fuel ratio of the air-fuel mixture and the ignition timing. A fuel injection signal S is sent to a fuel injection valve 336 and an ignition device 338 provided in the engine 8 in order to maintain the fuel injection valve optimally depending on the driving condition of the vehicle.
Outputs F F and ignition signal SEE.

Hereinafter, the operation of ignition timing control in accordance with the shift of the sub-transmission 14 in the engine control computer 334 will be explained according to the flowchart shown in FIG.

First, by executing step S1, input signals such as the speed change ratio signal SSR, gear stage signal SGW, and a signal representing the throttle valve opening θ are read. Next, in step S2, it is determined whether or not there has been a shift change based on the gear signal SGW, that is, whether the gear of the sub-transmission I4 has been changed from a low gear to a high gear or from a high gear to a low gear. It is determined whether or not. If this judgment is denied, the following steps are skipped, but if this judgment is affirmed, the following step S3 is executed. In this step S3, the contents of timer 'F are cleared to zero. In the subsequent step S4, a retard angle δ from the basic ignition timing is determined based on the actual gear ratio T and throttle valve opening θ from the predetermined relationship shown in FIG. 6, for example. The relationship shown in FIG. 6 was experimentally determined in advance to prevent shift shock from occurring when the sub-transmission 14 is upshifted, and to prevent gJl from being unnecessarily retarded at the time of ignition of the engine 8. The retardation amount δ is stored in the ROM in the form of a data map or a functional formula, and the larger the actual gear ratio T and throttle valve opening θ, the larger the retardation amount δ.

In step S5, it is determined whether or not the content of the timer T has reached a predetermined constant value T. If the content of the timer T has not reached a predetermined constant value T, step S5 is repeatedly executed. The next step S6 is executed.

This value T corresponds to a delay time from when an upshift of the auxiliary transmission 14 is commanded until the start of retardation of the ignition timing to alleviate the shift shock. In step S6, the basic ignition timing is changed (corrected) so that the ignition timing is retarded by the amount of retardation δ determined in step S4. And step S7
is executed, and it is determined whether the contents of the timer T have reached a predetermined constant value T2. This constant value T
2 corresponds to the time during which the ignition timing retard control is performed when the sub-shift a14 is shifted up. And step S7
If the judgment is negative, steps 86 and subsequent steps are executed again, but if the judgment is affirmative, the step knob S8 is
After the above-mentioned ignition timing retard control is stopped,
The ignition timing retard control routine for alleviating the shift shock is ended.

As described above, according to the present embodiment, when the sub-transmission 14 is shifted up automatically depending on the driving condition of the vehicle, only a certain period of time (time corresponding to Tz-T+) is required.
As shown in FIG. 8 (bl), the ignition timing is retarded. Since the ignition timing of the engine 8 and the output torque are in the general relationship shown in FIG. 14 is switched from a low gear to a high gear, the output torque of the engine 8 is appropriately suppressed.At this time of switching, the gear train that established the low gear is released, and then the high gear is established. However, since the output torque of the engine 8 is suppressed as described above, the torque on the input side of the auxiliary transmission 14 is made smaller than before, and the shift amplifier or downshift At this time, the output side member of the sub-transmission 14 (
The torque newly transmitted to the output gear 70) and the inertia torque of the preceding stage are reduced, and the shift shock is suitably alleviated. In particular, in a vehicle equipped with an auxiliary transmission at the rear stage of the continuously variable transmission as in this embodiment, as the gear ratio T of the continuously variable transmission increases, the torque input to the auxiliary transmission increases and the torque input to the auxiliary transmission increases. Due to the increase in inertia torque, there was a tendency for the shift shock accompanying the shift of the auxiliary transmission to be large and fluctuate, but according to the present invention, this has been effectively eliminated.
Improves drivability. Moreover, the load applied to the friction engagement device during shifting of the sub-transmission is reduced, and its durability is increased. The broken line in FIG. 8(a) shows the torque of the output side member of the sub'lld machine 14 during upshifting, and this torque change is significantly smaller than in the conventional case shown by the solid line, and the shift shock This shows that the situation has been relaxed. In the figure, the decrease in the output torque before the ignition timing is retarded is due to the fact that the set pressure of the accumulator 320 has been lowered in advance than before. That is, in this embodiment, since the output torque of the sub-transmission 14 is determined by the torque capacity of the high-speed clutch 72, the engine is retarded by retarding the ignition timing so that the gear can be smoothly shifted even with the previously reduced torque capacity. 8's output torque is lowered.

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

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

Furthermore, instead of the actual throttle valve opening θ, an amount representing a required output such as an accelerator pedal operation amount or intake pipe negative pressure may be used. In this way, the present invention can be suitably applied even when the vehicle is equipped with a diesel engine that does not use a throttle valve. The ignition timing in this case is adjusted by the fuel injection timing.

Furthermore, in the above embodiment, the CVT 12 and the sub-transmission 1
4, the present invention is also applicable to vehicles equipped with an automatic transmission that automatically changes gears to multiple gears, that is, a so-called automatic transmission.

In addition, in the flowchart of FIG.
Ignition timing retard control is executed to alleviate the shift shock on the condition that the engine is upshifted.
The fact that the acceleration of the vehicle exceeds a predetermined constant value may be added to the conditions for executing the ignition timing retard control. In this way, there is an advantage that the ignition timing retard control can be eliminated in a region where the shift shift 3.7 is inherently small and is not a problem.

Further, in the above embodiment, when the sub-transmission 14 is upshifted, the ignition timing is retarded to suppress the output torque of the engine 8, but at the same time the ignition timing is retarded, the fuel injection amount is The output F and the torque of the engine 8 may be suppressed by also limiting the amount of torque.

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

[Brief explanation of the drawing]

FIG. 1 is a flowchart illustrating the main part of the operation of the apparatus to which the present invention is applied. FIG. 2 is a schematic diagram showing a power transmission device for a vehicle to which the present invention is applied. Figure 3 is the second
FIG. 3 is a diagram showing the relationship between the range of the sub-transmission and the frictional engagement device in the device shown in the figure. FIG. 4 is a detailed circuit diagram of a hydraulic control system for operating the device of FIG. 2. FIG. 5 is a block diagram showing an electrical control circuit provided in the device of FIG. 2. 6th
The figure is a diagram showing the relationships used in the flowchart of FIG. 1. FIG. 7 is a graph 1 showing the general relationship between engine ignition timing and output 1 herk. FIG. 8 is a time chart illustrating the operation of the device shown in FIG. 2. 12: Belt type continuously variable transmission 14: Sub-transmission δ: Retard angle γ: Gear ratio θ: Throttle valve opening Fig. 2 Fig. 7 Ignition timing Fig. 8

Claims (5)

[Claims] (1) An ignition timing control method for a vehicle equipped with an automatic transmission that automatically switches to a predetermined gear according to driving conditions, the method controlling the ignition timing of the engine of the vehicle when changing the gear of the automatic transmission. 1. A method for controlling ignition timing for a vehicle equipped with an automatic transmission, characterized in that the output of the engine is reduced by delaying the ignition timing. (2) The vehicle is equipped with a continuously variable transmission that continuously changes the rotation of the engine, and the automatic transmission is provided downstream of the continuously variable transmission and is configured to adjust the running state of the vehicle. An ignition timing control method for a vehicle equipped with an automatic transmission according to claim 1, wherein the automatic transmission is automatically switched to a predetermined gear according to the following. (3) The ignition timing is delayed in accordance with a retard amount determined based on a predetermined relationship based on the accelerator operation amount or throttle valve opening of the vehicle and the gear ratio of the continuously variable transmission. An ignition timing control method for a vehicle equipped with an automatic transmission according to claim 2. (4) The ignition timing of a vehicle equipped with an automatic transmission according to claim 3, wherein the amount of retardation of the ignition timing is increased as the gear ratio of the continuously variable transmission becomes larger. Control method. (5) Ignition timing control for a vehicle equipped with an automatic transmission according to claim 3, wherein the amount of retardation of the ignition timing is increased as the throttle valve opening of the vehicle increases. Method.