JP3937123B2 - Shift control device for automatic transmission - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a shift control device for an automatic transmission, and more particularly to a shift control device for suitably performing a shift performed by fastening a friction element by increasing an operating hydraulic pressure.
[0002]
[Prior art]
The automatic transmission determines the power transmission path (gear stage) of the gear transmission system by selectively hydraulically operating (engaging) friction elements such as a plurality of clutches and brakes, and switching the friction elements to be operated. It is configured to perform a shift to the next gear stage.
[0003]
Since the automatic transmission has such a configuration, when switching from a certain shift speed to another shift speed, in some cases, a frictional element causes a rotational change in a direction opposite to the engine rotational change caused by the torque of the engine itself. The shift may not proceed or end unless it is caused by the fastening operation of the clutch, and the present invention relates to a shift performed by fastening the friction element by increasing the hydraulic pressure.
[0004]
At the time of the speed change, the effective gear ratio represented by the transmission input / output rotation speed ratio is changed by increasing the engagement capacity of the friction element to be engaged (engagement side friction element) by increasing the operation oil pressure (engagement side operation oil pressure) While the inertia phase changing from the front gear ratio to the post-shift gear ratio is proceeding, it is a common practice to set a shelf pressure corresponding to the transmission input torque in order to prevent a shift shock in the engagement side hydraulic pressure.
[0005]
The shelf pressure of the engagement-side operating hydraulic pressure during the inertia phase has been conventionally set as described in, for example, Japanese Patent Laid-Open No. 7-83321.
That is, as shown in FIG. 8, during the shelf pressure setting period C6, the transmission input torque is obtained by multiplying the engine torque obtained from the internal signal of the engine controller and the torque ratio of the torque converter. Torque sharing pressure P of the friction element on the fastening side that gives a fastening capacity that can be transmitted_B And the return spring pressure P of the fastening side friction element that balances with the return spring force of the fastening side friction element_RTNIn order to advance the inertia phase, the rotational speed of the engagement side friction element necessary to reduce the rotational speed of the engine and the torque converter to the rotational speed after shifting in a predetermined time (during upshift) Inertia induced pressure P_INA And the working side hydraulic pressure P_C With this, the working side hydraulic pressure P_C The shelf pressure as illustrated in FIG. 8 is set.
[0006]
[Problems to be solved by the invention]
By the way, the turbine torque T of the torque converter that is the transmission input torque._t Is the turbine speed (transmission input speed) N_t For example, the engine speed changes with the tendency as shown in FIG._t Turbine torque (transmission input torque) T_t Will rise.
For this reason, the inertia phase proceeds and the turbine speed N_t Is the rotation speed N at the start of the inertia phase_tSTo N at the end of the inertia phase_teTurbine torque (transmission input torque) T as it decreases (during upshift)_t Will increase.
[0007]
Therefore, the torque sharing pressure P of the engagement-side friction element determined according to the transmission input torque_B 8 also increases as the inertia phase progresses as shown in FIG. 8, and as a result, the engagement-side operating oil pressure P in the shelf pressure setting period C6._C Is set to a shelf pressure having a gradient as exemplified in the figure, and the shelf pressure gradient is uniquely determined by the progress speed of the inertia phase (shift), which causes the problem described below.
[0008]
When the calculation error of the transmission input torque becomes larger than the actual value and the shelf pressure varies in the excessive direction, the engagement side friction element is fastened and the turbine speed N in the initial phase of the inertia phase is increased._t As shown in FIG. 9, the turbine torque (transmission input torque) T is large._t Increase rapidly.
As a result, the above torque sharing pressure P_B Suddenly increases, and the engagement side hydraulic pressure P shown in FIG._C Ascending slope becomes steep and turbine speed N_t The divergence cycle that causes a further sudden decrease in the operating pressure P_C There is a concern that a large shift shock may occur at the end of the inertia phase due to the rapid increase of the acceleration.
[0009]
On the contrary, when the calculation error of the transmission input torque becomes smaller than the actual value and the shelf pressure varies in an excessively small direction, the engagement of the engagement side friction element is delayed and the turbine rotational speed N at the initial stage of the inertia phase is delayed._t As shown in FIG. 9, the turbine torque (transmission input torque) T is small._t Increases slowly.
As a result, the above torque sharing pressure P_B The engagement side hydraulic pressure P shown in FIG._C Ascending slope of the turbine becomes gentle and the turbine speed N_t The divergence cycle in which the lowering of the pressure is further slowed is repeated, and the engagement side hydraulic pressure P_C As a result, the inertia phase does not end easily, and a feeling of extension of the speed change occurs. In the worst case, the inertia phase (shift) does not end completely.
[0010]
Further, in the above-described conventional shift control, the engagement side operating oil pressure P_C The shelf pressure gradient is uniquely determined by the speed of the inertia phase (shift), so that the shelf pressure gradient is manipulated to tune the transmission quality, and the strength against variation and the reduction of the shift shock are balanced at a high level. There is also a concern that the degree of freedom cannot be given.
[0011]
  In the first aspect of the present invention, the first two problems are that the transmission input torque used for setting the shelf pressure is the transmission input torque every moment during the shelf pressure control period. Based on the fact that it is due to the fact that the effective shelf pressure gradient cannot be arbitrarily set for
  The transmission input torque used for setting the shelf pressure is the transmission input torque that should be generated at the transmission input rotation speed at the end of the inertia phase, and there is an error in the estimation of the transmission input torque or the target fastening capacity. Even if there is an error in the actual engagement capacity, the end of the inertia phase (shift) is compensated for without the feeling of extension during the shift and a large shift shock, so as to solve the former two problemsmainObjective.
[0012]
  The first invention furtherIn addition to the above-described effects, the shelf pressure gradient can be arbitrarily set, whereby the shelf pressure gradient can be manipulated to arbitrarily tune the transmission quality, and the strength against variation and the reduction of the transmission shock can be reduced. To be able to balance in a high dimension, and to solve the latter problem regarding the degree of freedom.SecondObjective.
[0013]
  Claim2No. described in2It is an object of the present invention to easily obtain the transmission input torque at the transmission input rotational speed at the end of the inertia phase.
[0014]
  Claim3No. described in3It is an object of the present invention to more easily obtain the transmission input torque at the transmission input rotational speed at the end of the inertia phase.
[0015]
[Means for Solving the Problems]
  For these purposes, first, a shift control device for an automatic transmission according to the first invention is:The configuration is as follows.
  First of all, to explain the prerequisite automatic transmission,The shelf pressure of the working hydraulic pressure related to the friction element to be set in order to advance the inertia phase of the shift performed by fastening the friction element by increasing the working hydraulic pressure is determined at least according to the transmission input torque.Is.
[0016]
  First1A shift control device for an automatic transmission according to the invention comprises:Such an automatic transmissionIn
  The transmission input torque used for determining the shelf pressure is a transmission input torque that should be generated at the transmission input rotation speed at the end of the inertia phase,
  Furthermore,The shelf pressure determined according to the transmission input torque at the transmission input rotation speed at the end of the inertia phase is set as the initial shelf pressure, and the initial shelf pressure is added with a gradient setting pressure that increases with the passage of time. Is set.
[0017]
  First2The shift control device for an automatic transmission according to the invention isClearlyLeave
  The transmission input torque at the transmission input rotation speed at the end of the inertia phase is based on a map showing the relationship between the transmission input rotation speed and the throttle opening and the transmission input torque, and the transmission input rotation at the end of the inertia phase. It is characterized in that it is configured to search and obtain from a number.
[0018]
  First3A shift control device for an automatic transmission according to the invention is2In the invention,
  The transmission input rotation speed at the end of the inertia phase is calculated from the transmission output rotation speed and the post-shift gear ratio, and is obtained.
[0019]
【The invention's effect】
In the first aspect of the present invention, during the shift performed by engaging the friction element by increasing the operating hydraulic pressure, the operating hydraulic pressure related to the friction element is determined according to at least the transmission input torque in order to advance the inertia phase of the shift. Set the shelf pressure.
In the first invention, the transmission input torque used for determining the shelf pressure is particularly the transmission input torque that should be generated at the transmission input rotation speed at the end of the inertia phase.
[0020]
  Of the first inventionThus, when determining the shelf pressure for the inertia phase progression according to the transmission input torque at the transmission input rotation speed at the end of the inertia phase,
  The transmission input torque at the transmission input speed at the end of the inertia phase does not change during the shelf pressure control period when the throttle opening is constant, and the shift is completed when the throttle opening is constant. Because it is the input torque at the time,
  If there is an error in the estimation of the input torque, or if there is an error in the actual engagement capacity of the friction element, the end of the inertia phase (shift) is compensated without repeating the diverging cycle described above. At the same time, it is possible to realize the compensation without the feeling of extension during the shift and a large shift shock, and it is possible to solve the above-mentioned problems of the conventional apparatus.
[0021]
  First1In the inventionMoreThe shelf pressure determined according to the transmission input torque at the transmission input rotation speed at the end of the inertia phase is used as the initial shelf pressure, and the initial shelf pressure is added with the gradient setting pressure that increases with the passage of time. Because
  The shelf pressure gradient can be set arbitrarily depending on how the gradient setting pressure is applied, and this can be achieved by manipulating the shelf pressure gradient to tune the transmission quality arbitrarily or by making the shelf pressure gradient steep. Therefore, it is possible to balance the strength against the variation and the reduction of the shift shock achieved by grading the shelf pressure gradient at a high level, and it is possible to solve the above-mentioned problems of the conventional apparatus regarding the degree of freedom.
[0022]
  First2In the invention, when the transmission input torque at the transmission input rotation speed at the end of the inertia phase is obtained, the inertia phase is completed based on the map showing the relationship between the transmission input rotation speed, the throttle opening, and the transmission input torque. To search and obtain from the transmission input speed of
  The transmission input torque at the transmission input rotational speed at the end of the inertia phase can be easily obtained, and the above-described effects can be easily realized.
[0023]
  First3In the invention,2In order to obtain and calculate the transmission input rotational speed at the end of the inertia phase in the invention from the transmission output rotational speed and the gear ratio after shifting,
  The transmission input torque at the transmission input rotation speed at the end of the inertia phase can be obtained more easily, and the above-described effects can be realized more easily.
[0024]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.
FIG. 1 shows a shift control apparatus for an automatic transmission according to an embodiment of the present invention, wherein 1 is an engine and 2 is an automatic transmission.
The output of the engine 1 is adjusted by a throttle valve that increases in opening degree from fully closed to fully open as the accelerator pedal is operated by the driver, and the output rotation of the engine 1 is transmitted through the torque converter 3 to the automatic transmission. 2 is input to the input shaft 4.
[0025]
The automatic transmission 2 includes a front planetary gear set 6 and a rear planetary gear set 7 which are sequentially placed from the engine 1 side on the input / output shafts 4 and 5 arranged in a coaxial butting relationship, and these are provided as planets in the automatic transmission 2. The main component of the gear transmission mechanism.
The front planetary gear set 6 close to the engine 1 has a front sun gear S._F , Front ring gear R_F , Front pinion P meshing with these_F , And a front carrier C that rotatably supports the front pinion_F A simple planetary gear set consisting of
The rear planetary gear set 7 far from the engine 1 is also_R , Rear ring gear R_R , Rear pinion P meshing with these_R , And a rear carrier C that rotatably supports the rear pinion_R A simple planetary gear set consisting of
[0026]
Friction elements that determine the transmission path (speed stage) of the planetary gear transmission mechanism include low clutch L / C, 2nd and 4th brake 2-4 / B, high clutch H / C, low reverse brake LR / B, and low one. The way clutch L / OWC and the reverse clutch R / C are provided in correlation with the components of the planetary gear sets 6 and 7 as follows.
That is, the front sun gear S_F Can be appropriately connected to the input shaft 4 by the reverse clutch R / C and can be appropriately fixed by the second-speed / four-speed brake 2-4 / B.
[0027]
Front carrier C_F Can be appropriately coupled to the input shaft 4 by the high clutch H / C.
Front carrier C_F Further, the low one-way clutch L / OWC prevents rotation in the direction opposite to the engine rotation, and can be appropriately fixed by the low reverse brake LR / B.
And front carrier C_F And rear ring gear R_R Can be appropriately coupled by a low clutch L / C.
Front ring gear R_F And rear carrier C_R The front ring gear R_F And rear carrier C_R Is coupled to the output shaft 6 and the rear sun gear S_R Is coupled to the input shaft 4.
[0028]
The power transmission train of the planetary gear speed change mechanism includes a selective hydraulic operation (fastening) indicated by solid circles in FIG. 2 of the friction elements L / C, 2-4 / B, H / C, LR / B, and R / C. ) And the self-engagement of the low one-way clutch L / OWC indicated by a solid line ◯ in the same figure, forward first speed (1st), forward second speed (2nd), forward third speed (3rd), forward The fourth forward speed (4th) forward speed and the reverse speed (Rev) can be obtained.
Note that the hydraulic operation (fastening) indicated by the dotted circles in FIG. 2 is a friction element to be operated when engine braking is necessary.
[0029]
The engagement logic of the shift control friction elements L / C, 2-4 / B, H / C, LR / B, and R / C shown in FIG. 2 is realized by the control valve body 8 shown in FIG. In addition to a manual valve (not shown), a line pressure solenoid 9, a low clutch solenoid 10, a second / fourth speed brake solenoid 11, a high clutch solenoid 12, a low reverse brake solenoid 13, and the like are inserted in 8.
[0030]
The line pressure solenoid 9 switches the line pressure, which is the original pressure of the shift control, by turning on and off, and a manual valve (not shown) is driven forward by the driver according to the desired travel mode (D) range position, It is assumed that the vehicle is operated to the reverse travel (R) range position or the parking / stop (P, N) range position.
In the D range, the manual valve corresponds to the low clutch L / C by controlling the duty of the low clutch solenoid 10, the second / fourth speed brake solenoid 11, the high clutch solenoid 12, and the low reverse brake solenoid 13 using the above line pressure as a source pressure. Line pressure is supplied to a predetermined circuit so that the hydraulic pressures of the 2nd and 4th brakes 2-4 / B, the high clutch H / C, and the low reverse brake LR / B can be individually controlled, and the duty of each solenoid It is assumed that the first to fourth speed engagement logic shown in FIG. 2 is realized by the control.
However, in the R range, the manual valve directly supplies the line pressure to the reverse clutch R / C without depending on the duty control of each solenoid, and the low reverse brake solenoid 13 for the low reverse brake LR / B. It is assumed that the reverse engagement logic shown in FIG. 2 is realized by supplying the duty control pressure according to, and engaging them.
In the P and N ranges, the manual valve does not supply the line pressure to any circuit, and the automatic transmission is neutralized by releasing all the friction elements.
[0031]
The transmission controller 14 executes ON / OFF control of the line pressure solenoid 9, and the duty control of the low clutch solenoid 10, the second speed / fourth speed brake solenoid 11, the high clutch solenoid 12, and the low reverse brake solenoid 13, respectively.
For this purpose, the transmission controller 14 includes a signal from a throttle opening sensor 15 that detects the throttle opening TVO of the engine 1, and
Turbine rotational speed N which is the output rotational speed (transmission input rotational speed) of torque converter 3_t A signal from the turbine rotation sensor 16 for detecting
The rotational speed N of the output shaft 5 of the automatic transmission 2_O A signal from the output rotation sensor 17 for detecting
A signal from the inhibitor switch 18 for detecting the selected range;
The engagement-side friction element to be engaged at the time of shifting, that is, as is clear from FIG. 2, the 2nd and 4th speed brakes 2-4 / B at the 1 → 2 shift, the high clutch H / C at the 2 → 3 shift, → At the time of four shifts, the signals from the hydraulic switch group 19 arranged in the second and fourth speed brakes 2-4 / B are respectively input.
Here, the hydraulic switch group 19 is turned on when the operating hydraulic pressure of the corresponding friction element reaches a pressure at which the loss stroke of the friction element ends and the engagement capacity starts to be generated.
[0032]
In order to explain the automatic transmission operation in the D range in which the present invention is concerned, the transmission controller 14 executes a control program (not shown), and based on a predetermined shift map, the throttle opening TVO and the transmission output rotational speed. N_O From the (vehicle speed), a suitable gear position required in the current driving state is searched.
Next, the transmission controller 14 determines whether or not the currently selected shift speed is coincident with the preferred shift speed. If they are not coincident, a shift command is issued and the shift to the preferred shift speed is executed, that is, in FIG. Based on the engagement logic table, the hydraulic pressure of the friction element is changed by duty control of the solenoids 10 to 13 so that the engagement and release switching of the friction element for the speed change are performed.
[0033]
Here, as in the shift between the second speed and the third speed and the shift between the third speed and the fourth speed, the first friction element is released by a decrease in the hydraulic pressure while the second friction is In order to explain the change-over shift performed by fastening the elements by increasing the operating hydraulic pressure, this shift should be released when the shift is a drive-up shift accompanying a vehicle speed increase in the normal drive state (drive state opposite to the engine brake), for example. Release side hydraulic pressure command value P which is the command value of the hydraulic pressure of the friction element_O And the engagement side hydraulic pressure command value P which is the command value of the hydraulic pressure of the friction element to be engaged_C Is given as shown in FIG.
[0034]
First, engagement side hydraulic pressure command value P_C In order to terminate the loss stroke of the engagement side friction element as early as possible during the C1 period from the shift command instant t1, the command value P_C Is a high precharge pressure.
In the subsequent C2 period, if the loss stroke is ended with the precharge pressure, a shock is generated, so the command value P_C Is once lowered below the precharge pressure and gradually increased at a change rate that does not cause a shock.
As a result, the engagement side friction element ends the loss stroke and begins to have the engagement capacity, and the hydraulic switch 19 is turned on at this instant t2.
[0035]
Set time T from instant t2 when hydraulic switch 19 is turned ON_C Command value P during C4 period from when_C A predetermined change rate θ_C To increase the engagement capacity of the engagement side frictional element to the required capacity corresponding to the input torque required for the switching speed.
If the release-side friction element is released at this time, the release-side friction element is switched to the fastening-side friction element, and the transmission input / output rotation ratio N_t / N_O The effective gear ratio i represented by ## EQU3 ## starts to change from the gear ratio before shifting to the gear ratio after shifting, and the inertia phase is started.
[0036]
From the inertia phase start instant t3 to the command value P during the C5 period_C Is further increased at a reduced change rate to reach a predetermined initial shelf pressure corresponding to the transmission input torque, and during the subsequent shelf pressure setting period C6, the command value P_C Is increased at a predetermined change rate that is further reduced to a predetermined gradient θ._T Set the shelf pressure.
Thereafter, in the C7 period until the inertia phase end determination instant t4, the command value P_C In order to prevent a shock at the end of the inertia phase, the pressure is once made smaller than the shelf pressure and then slowly raised, and the command value P is applied during the C8 period from the inertia phase end determination instant t4._C Is rapidly increased to the line pressure, which is the original pressure.
[0037]
Next, release side hydraulic pressure command value P_O In order to secure the release response of the release-side friction element during the O1 period from the shift command instant t1, the command value P_O Is stepped down to a predetermined value, and the command value P is reduced during the subsequent O2 and O3 periods._O Are successively reduced at a small change rate to the release pressure before switching, that is, to the release-side operating hydraulic pressure value at which the release-side friction element does not slip.
Then, the set time T from the instant t2 when the engagement side friction element starts to have the engagement capacity at the end of the loss stroke and the hydraulic switch 19 is turned on._O The command value P during the O4 period until_O Is maintained at the release pressure before the change and the set time T_O The command value P is set so that switching from the release-side friction element to the engagement-side friction element is performed during the O5 period after the elapse of time._O A predetermined change rate θ_O To gradually reduce the engagement capacity of the release side friction element.
[0038]
In the O6 period and the O7 period after detecting that the inertia phase is started by switching from the release side friction element to the engagement side friction element at the instant t3, the command value P_O At a predetermined change rate set in each period to finally reach 0, and in the O8 period, the command value P_O Is kept at 0 and the inertia phase is advanced to its completion.
[0039]
By the way, in this embodiment, the transmission controller 14 applies the engagement-side operating hydraulic pressure command value P in the shelf pressure control period C6 when performing the above-mentioned changing gear shift._C In particular, the control as shown in FIG. 5 is executed by executing the control program shown in FIG. 4, and the shelf pressure shown in FIG.
The shelf pressure control program shown in FIG.
In step 30, the engagement side hydraulic pressure (torque sharing pressure) P is such that the engagement side frictional element can transmit the transmission input torque at the end of the inertia phase._B The routine to determine
In step 40, the return spring pressure P of the fastening side friction element that balances with the return spring force of the fastening side friction element._RTN The routine to determine
In step 50, the inertia induced pressure P of the engagement side friction element necessary to reduce the engine speed and the torque converter speed to the speed after the upshift to advance the inertia phase._INA The routine to determine
In step 60, the engagement side hydraulic pressure command value P_C Shelf pressure gradient θ_T Gradient setting pressure P for setting_SLP A routine for determining is started.
[0040]
Torque sharing pressure (P_B ) In the determination routine, the throttle opening TVO is read in step 31, and the transmission output rotational speed N is read in step 32._O In step 33, the gear ratio i after shifting is read_fin At the end of gear shifting (at the end of the inertia phase) at step 34, the turbine speed N_teN_te= N_O Xi_fin Calculated by
In the next step 35, the throttle opening TVO and the inertial turbine end turbine speed N_teTo the turbine torque T at the end of the inertia phase based on the turbine torque map illustrated in FIG._teSearch for.
Thereafter, in step 36, the turbine torque T at the end of the inertia phase_teEngagement side hydraulic pressure (torque sharing pressure) P such that the engagement side frictional element can transmit (the transmission input torque at the end of the inertia phase) to the limit._B Turbine torque T at the end of inertia phase_teCalculated by multiplying by the torque / hydraulic conversion factor.
[0041]
Return spring pressure (P_RTN ) In the determination routine, in step 41, the return spring pressure P required to finish the loss stroke of the engagement side friction element in balance with the return spring force so as to crush the return spring of the engagement side friction element._RTN Is read.
[0042]
Inertia induced pressure (P_INA ) In the determination routine, in step 51, the turbine rotational speed difference ΔN before and after the gear shift._t This can be calculated from the turbine rotation read value at the start of the inertia phase and the turbine rotation speed at the end of the inertia phase that can be determined as in step 34.
Next, at step 53, an inertia induced torque necessary for lowering the transmission input rotational speed at a predetermined rate of change of rotation in order to advance the inertia phase, and an engagement necessary for the engagement side friction element to generate this torque. Coefficient A representing the relationship between the rate of change in turbine rotation and the engagement-side hydraulic pressure_INA Ask for.
Thereafter, in step 53, target inertia phase time data T from the start of the inertia phase to the end of the inertia phase._INA In step 54, the inertia induced pressure P of the engagement-side friction element necessary to reduce the engine speed and the torque converter speed to the speed after the upshift to advance the inertia phase._INA P_INA = A_INA × ΔN_t / T_INA Calculated by
[0043]
The gradient setting pressure (P_SLP ) In the determination routine, in step 61, the shelf pressure gradient θ_T The shelf pressure gradient data such as_T Multiplied by the elapsed time from the start of the shelf pressure calculation (elapsed time after entering the shelf pressure control period C6 in FIGS. 3 and 5), as shown in FIGS. 3 and 5, the engagement side hydraulic pressure command value P_C Shelf pressure gradient θ_T Gradient setting pressure P for setting_SLP Ask for.
[0044]
In step 22, the torque sharing pressure P obtained in steps 36, 41, 54, and 62, respectively, is obtained._B Return spring pressure P_RTN , Inertia induced pressure P_INA , Gradient setting pressure P_SLP The shelf pressure is determined by adding together the engagement side hydraulic pressure P_C In the shelf pressure control period C6, the shelf pressure is controlled to become the shelf pressure as exemplified by the solid line in FIGS.
[0045]
By the way, in the present embodiment, as described above, the engagement-side hydraulic pressure P is used to advance the inertia phase during the shift in which the inertia phase proceeds only by engaging the friction element._C When determining the shelf pressure to be set to, the transmission input torque (turbine torque) T at the end of the inertia phase is not used as the transmission input torque without changing the transmission input torque that changes every moment during the shift._teTherefore, the torque sharing pressure P corresponding to the transmission input torque is used._B As is clear from FIG. 5, the torque sharing pressure P is not changed during the shelf pressure control period C6._B Is required after the end of the shift, and the end of the inertia phase (shift) is compensated without the repetition of the diverging cycle that has occurred in the conventional device, and the feeling of extension of the shift and a large shift shock The compensation can be realized without any problem, and the above-mentioned problems of the conventional apparatus can be solved.
[0046]
In addition, in the present embodiment, the transmission input torque (turbine torque) T at the end of the inertia phase_teThe shelf pressure determined according to the initial shelf pressure is set as the initial shelf pressure, and the initial shelf pressure increases with time as shown in FIG._SLP Is added to the shelf pressure gradient θ_T To set the gradient set pressure P_SLP Shelf pressure gradient θ depending on how_T Can be arbitrarily set as shown by a two-dot chain line or a one-dot chain line in FIG._T Can be used to arbitrarily tune the transmission quality, or the shelf pressure gradient θ_T The strength against variation achieved by steepening and the shelf pressure gradient θ_T It is possible to balance the reduction of the shift shock achieved by relaxing the speed at a high level, and to solve the above-mentioned problem of the conventional apparatus related to the degree of freedom.
[0047]
Further, in the present embodiment, the transmission input torque (turbine torque) T at the end of the inertia phase_teIs obtained, the transmission input speed (turbine speed N_t ) And transmission input torque (turbine torque) T_t Transmission input speed at the end of the inertia phase (turbine speed N_te) And input torque (turbine torque) T at the end of the inertia phase_teTransmission input torque (turbine torque) T at the end of the inertia phase_teCan be easily obtained, and the above-described effects can be easily achieved.
[0048]
Further, in the present embodiment, the transmission input rotational speed (turbine rotational speed N at the end of the inertia phase in the third invention)._te) As shown in steps 32 to 34 of FIG._O And gear ratio i after shifting_fin Therefore, the transmission input torque (turbine torque) T at the end of the inertia phase is calculated._teCan be obtained more easily, and the above-described effects can be realized more easily.
[0049]
Where shelf pressure gradient θ_T The shift characteristics during upshifting in the drive state when the slope is indicated by the solid line in FIG. 5, the slope indicated by the two-dot chain line, or the slope indicated by the one-dot chain line are the shifts when the shelf pressure is high or low. A description will be given below together with the characteristics with reference to FIGS.
The speed change characteristic is discussed by the time series change of the effective gear ratio i. Basically, the height of the shelf pressure determines the change gradient of the effective gear ratio i, and the change gradient of the effective gear ratio i becomes steeper as the shelf pressure increases. The shelf pressure gradient θ_T Determines the increasing rate of the change gradient of the effective gear ratio i, and the shelf pressure gradient θ_T As the speed becomes steeper, the increasing rate of the change gradient of the effective gear ratio i becomes larger.
[0050]
As shown by α in FIG. 6, the shelf pressure is determined to be relatively high as shown by the solid line in FIG. 5, and the shelf pressure gradient is relatively small as shown by the solid line in FIG. The gear ratio i changes as indicated by α ′ in the figure, and the overall change gradient becomes relatively steep, but the increase rate of the change gradient is relatively moderate.
When the shelf pressure is relatively high as shown by β in FIG. 6 and the shelf pressure gradient is relatively large as shown by the two-dot chain line in FIG. 5, the effective gear ratio i is β ′ in FIG. As shown, the overall change gradient becomes relatively steep, and the increase rate of the change gradient becomes relatively large. In this case, the transmission output torque waveform has a large peak torque as indicated by β in FIG. 7 and causes a large shift shock.
[0051]
When the shelf pressure is relatively high as indicated by δ in FIG. 6 but the shelf pressure gradient is as small as shown by the one-dot chain line in FIG. 5, the effective gear ratio i is δ ′ in FIG. The overall change gradient is relatively steep, but the rate of increase of the change gradient is rather gradual. In this case, the transmission output torque waveform does not have a large peak torque as shown by δ in FIG. 7 and is a waveform that terminates the shift, and theoretically becomes an ideal torque waveform in terms of shift shock.
[0052]
As shown by γ in FIG. 6, the shelf pressure gradient is as large as β, but when the shelf pressure is relatively low, the effective gear ratio i has an overall change gradient as shown by γ ′ in FIG. Although it becomes relatively gentle, the rate of increase of the change gradient becomes relatively large. In this case, the transmission output torque waveform has a relatively large peak torque as indicated by γ in FIG. 7, and a slightly large shift shock is generated.
When the shelf pressure is relatively low as indicated by ε in FIG. 6 and the shelf pressure gradient is as small as indicated by δ, the effective gear ratio i is compared with the overall change gradient as indicated by ε ′ in FIG. The rate of increase of the change gradient is considerably reduced. In this case, the transmission output torque waveform does not have a peak torque as shown by ε in FIG. 7, but the shift is not completed easily.
[0053]
From the above, it is common knowledge that the shelf pressure is determined as indicated by δ in FIG. 6 so as to obtain an ideal torque waveform indicated by δ in FIG. 7, and the height of the shelf pressure (initial pressure) is shown in FIG. It is considered that it is optimal to determine the solid pressure gradient 5 as described above and to determine the shelf pressure gradient as indicated by δ in FIG.
However, in this case, the shelf pressure gradient is small and easily affected by variations in pipe resistance and hydraulic fluid viscosity, and the shelf pressure gradient is unchanged when the shelf pressure height (initial pressure) varies downward. Therefore, the shelf pressure becomes as shown by ε in FIG. 6, for example, and the speed change characteristic becomes the worst speed change characteristic extended as shown by the same sign ε in FIG.
On the other hand, when the shelf pressure gradient is set to a relatively steep slope as indicated by β in FIG. 6, although the shock is bad, even if the shelf pressure height (initial pressure) varies downward, γ in FIG. As shown in FIG. 8, it is possible to ensure that the inertia phase is finished within a predetermined time, and the worst speed change characteristic of delaying can be avoided.
Therefore, in order to avoid the worst speed change characteristic of extending even if the height of the shelf pressure (initial pressure) varies downward, the shelf pressure gradient indicated by δ in FIG. It is preferable to set the shelf pressure so that the shelf pressure gradient as shown by α intermediate between the shelf pressure gradient shown by β in FIG. 6 tends to increase.
[0054]
In the present embodiment, the shelf pressure gradient θ as illustrated by the solid line, the two-dot chain line, and the one-dot chain line in FIG._T Can be set arbitrarily, so this shelf pressure gradient θ_T Can be used to arbitrarily tune the transmission quality, and as described above, the shelf pressure gradient θ_T The strength against variation achieved by steepening and the shelf pressure gradient θ_T It is possible to achieve a high level balance between the reduction of the shift shock that is achieved by reducing the speed.
[0055]
In each of the above embodiments, when the automatic transmission is a direct acting valve type in which the hydraulic pressure of the friction element is directly controlled by individual solenoids, the automatic transmission is driven. Although the case where the up-shifting shift is performed has been described, the present invention is not limited to the types of these automatic transmissions and the types of the shifting shifts, but is applicable to other types of automatic transmissions and other types of shifting shifts. It is of course possible to apply the same concept, and in these cases, the same effect can be obtained.
[Brief description of the drawings]
FIG. 1 is a schematic system diagram showing a transmission train of an automatic transmission including a shift control device according to an embodiment of the present invention, and a shift control system thereof.
FIG. 2 is a view showing a relationship between a selected shift stage of the automatic transmission and a frictional logic of a friction element.
FIG. 3 is a time chart showing time-series changes in the engagement-side hydraulic pressure command value and the release-side hydraulic pressure command value when the shift control device according to the embodiment performs a drive-up switching shift.
FIG. 4 is a flowchart showing a shelf pressure calculation program for an engagement-side hydraulic pressure command value that is executed by the transmission controller at the time of drive-up change gear shifting in the shift control device according to the embodiment;
FIG. 5 is a time chart showing a time-series change in shelf pressure determined by the program of FIG. 4;
FIG. 6 is a time chart showing a relationship between a set shelf pressure and a time-series change in effective gear ratio.
7 is a time chart showing time-series changes in transmission output torque when several kinds of shelf pressures shown in FIG. 6 are used.
FIG. 8 is a time chart showing a time-series change in shelf pressure set by a conventional apparatus.
FIG. 9 is a characteristic diagram showing the relationship between turbine rotational speed and turbine torque for each throttle opening.
[Explanation of symbols]
1 engine
2 Automatic transmission
3 Torque converter
4 Input shaft
5 Output shaft
6 Front planetary gear set
7 Rear planetary gear set
8 Control valve
9 Line pressure solenoid
10 Low clutch solenoid
11 2-speed and 4-speed brake solenoid
12 High clutch solenoid
13 Low reverse brake solenoid
14 Transmission controller
15 Throttle opening sensor
16 Turbine rotation sensor
17 Output rotation sensor
18 Inhibitor switch
19 Hydraulic switch
L / C low clutch
2-4 / B 2-speed / 4-speed brake
H / C high clutch
LR / B low reverse brake
R / C reverse clutch
L / OWC Rowan Way Clutch

Claims (3)

Automatic shift in which the shelf pressure of the working hydraulic pressure related to the friction element to be set in order to advance the inertia phase of the shift performed by fastening the friction element by increasing the working hydraulic pressure is determined at least according to the transmission input torque In the machine
The transmission input torque used for determining the shelf pressure is a transmission input torque that should be generated at the transmission input rotation speed at the end of the inertia phase ,
The shelf pressure determined according to the transmission input torque at the transmission input rotation speed at the end of the inertia phase is set as the initial shelf pressure, and the initial shelf pressure is added with a gradient setting pressure that increases with the passage of time. shift control apparatus for an automatic transmission is characterized in that setting the. Oite to claim 1, the transmission input torque at the transmission input speed during the inertia phase ends, transmission input speed and the inertia phase based on a map showing the relationship between the throttle opening and the transmission input torque A shift control device for an automatic transmission, characterized in that it is configured to search and obtain from a transmission input rotation speed at the end. 3. The shift control apparatus for an automatic transmission according to claim 2 , wherein the transmission input rotation speed at the end of the inertia phase is calculated from the transmission output rotation speed and the gear ratio after shifting.

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