JP4152565B2 - Hydraulic control device for automatic transmission for vehicle - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a hydraulic control device for a vehicle automatic transmission, and more particularly to precharge control when a frictional engagement element such as a clutch is fastened.
[0002]
[Prior art]
Conventionally, in a vehicle automatic transmission, when a frictional engagement element such as a clutch is fastened from a released state, oil is rapidly filled into the frictional engagement element and a pipe that supplies oil to the frictional engagement element. A configuration is known in which the pre-charging is performed and the hydraulic pressure of the frictional engagement element to be fastened is quickly increased to the initial pressure of the fastening control.
[0003]
As a technique for optimizing the precharge control, a configuration in which the precharge pressure and the precharge time are changed according to the throttle opening, the oil temperature, the vehicle speed, and the like is disclosed in JP-A-7-027217 and JP-A-6-235451. No. gazette and the like.
[0004]
Japanese Patent Laid-Open No. 5-106722 discloses a configuration in which the precharge pressure is learned and controlled so that the pull-in torque generated at the time of shifting by changing the friction engagement elements becomes a predetermined value.
[0005]
Furthermore, Japanese Patent Laid-Open No. 5-312258 discloses a configuration in which the precharge time is learned and controlled by the rotational behavior (pre-firing speed) after precharge.
Japanese Patent Laid-Open No. 7-174217 discloses a configuration in which the time from the start of shifting to the inertia phase (start of rotation fluctuation) is measured, and the precharge time is changed based on the difference between the measured time and the target time. Has been.
[0006]
[Problems to be solved by the invention]
By the way, in the precharge, in order to complete the oil filling within the target time, it is necessary to fill the oil at a predetermined flow rate. The oil flow rate depends on the discharge amount of the oil pump or the line adjustment pressure (original pressure). When the oil discharged from the same oil pump is supplied to the torque converter and used for lubrication, the oil that can be supplied to the friction engagement element is limited. In some cases, a part of the discharged oil leaks, and it is not always possible to fill the friction engagement element with the oil at a desired flow rate.
[0007]
Also, even if the opening of the valve that controls the supply of oil to the friction engagement element is controlled so that the flow rate becomes the desired flow rate, the actual flow rate varies depending on the oil viscosity, and the desired flow rate In some cases, the oil could not be filled.
[0008]
If the desired flow rate for the frictional engagement element cannot be secured in the precharge control, the time required for filling becomes longer and the shift time becomes longer, and conversely, a large pressure fluctuation will occur if charging is performed for a single time with an excessive flow rate. The problem of letting it occur.
[0009]
In the case where the precharge pressure and precharge time are changed according to the throttle opening, oil temperature, vehicle speed, etc. as in the prior art, it is possible to respond to these changes in conditions, but quantitatively measure the flow rate fluctuations as described above. Since it is not the structure to judge, there existed a problem that a desired flow volume was not obtained accurately.
[0010]
In addition, if the configuration is such that the pre-charge pull-in torque, rotational behavior, and the like are determined, it will be evaluated whether or not the pre-charge control is appropriate for various fluctuation factors including flow rate fluctuations. Since the result is reflected in the shift after the next time, there is a problem that it is not possible to cope with a change in conditions for each shift (every precharge), and it is not always possible to make an optimal correction.
[0011]
The present invention has been made in view of the above problems, and in a precharge control when a friction engagement element is fastened, a hydraulic control device for an automatic transmission for a vehicle capable of accurately filling oil at a desired flow rate. The purpose is to provide.
[0012]
[Means for Solving the Problems]
Therefore, according to the first aspect of the present invention, the engagement / release of the frictional engagement element is controlled by the hydraulic pressure to change the speed, and the opening area of the solenoid that controls the supply of the hydraulic pressure to the frictional engagement element is set to the indicated pressure. In a hydraulic control device for an automatic transmission for a vehicle that increases and changes in response to an increase, a target flow rate when oil is charged into a frictional engagement element to be fastened, a limit flow rate that can be supplied to the frictional engagement element to be fastened, Based on the viscosity, the indicated pressure during precharging is set.
[0013]
According to this configuration, in the precharge in which the friction engagement element to be fastened is quickly charged with oil, the instruction pressure of the friction engagement element is set based on the target flow rate, the limit flow rate indicating the actual supply capacity, and the viscosity of the oil. .
[0014]
In the second aspect of the present invention, the target flow rate, and configured to set larger the limit flow rate is low Tokiho etc. finger manometric.
[0015]
According to such a configuration, the target flow rate, by actually larger as command pressure when the supply can limit the flow rate is small, Assurance of target flow rate. According to a third aspect of the present invention, an instruction pressure at the time of precharging is set based on the deviation between the target flow rate and the limit flow rate, and the viscosity of the oil, and the limit flow rate is increased as the oil viscosity increases. On the other hand, the instruction pressure is set to be larger as the oil viscosity is higher than the deviation between the target flow rate and the limit flow rate.
[0016]
According to this configuration, when the viscosity of the oil is high, the limit flow rate is estimated to be higher, while the command pressure is increased to estimate the limit flow rate according to the leakage flow rate, and the oil is filled at the desired flow rate. Let The oil viscosity can be estimated from the oil temperature.
[0017]
In the invention according to claim 4, the target flow rate is calculated based on the target filling time and the filling volume.
According to such a configuration, as a flow rate required to fill the filling volume with oil in the target time from the target value of the time from the start of filling to completion and the filling volume of oil including the friction engagement element and the pipe, A target flow rate is set.
[0018]
In the invention described in claim 5, the limit flow rate is calculated based on the dischargeable flow rate, the flow rate required for other elements, and the leakage flow rate. According to such a configuration, the amount of fluid that is discharged from the oil pump (dischargeable flow rate) minus the oil supplied to other elements such as the torque converter and the lubrication path, and the flow rate that leaks out in the middle is friction-engaged. The critical flow rate is required as the oil that can be supplied to the element .
[0019]
In the invention described in claim 6, the dischargeable flow rate is calculated based on the original pressure of the oil and the rotational speed of the engine combined with the automatic transmission.
According to this configuration, the dischargeable flow rate is obtained from the original pressure (line pressure) of the oil and the rotational speed of the engine that is proportional to the rotational speed of the oil pump that is rotationally driven by the engine.
[0020]
According to the seventh aspect of the present invention, the flow rate required for the other elements is calculated based on the rotational speed of the engine combined with the automatic transmission.
According to this configuration, the flow rate of oil supplied to other elements such as a torque converter and a lubrication path is determined according to the engine speed.
[0021]
The invention according to claim 8 is configured to calculate the leakage flow rate based on the temperature of the oil.
According to this configuration, since the amount of oil leakage varies depending on the viscosity of the oil, the leakage flow rate is calculated based on the temperature of the oil that correlates with the viscosity.
[0022]
【The invention's effect】
According to the first aspect of the invention, even if there is a change in the oil supply capacity and a change in the viscosity of the oil, there is an effect that the friction engagement element can be filled with the oil at the target flow rate.
[0023]
According to the second aspect of the present invention, there is an effect that the target flow rate can be secured by increasing the command pressure (opening area of the solenoid) against a decrease in the oil supply capability. According to the third aspect of the present invention, the critical flow rate can be estimated in response to the decrease in the viscosity of the oil and the increase in the leakage flow rate of the oil, and the viscosity with respect to the deviation between the same target flow rate and the critical flow rate. The higher the value, the higher the command pressure, and the more effective the target flow rate can be secured.
[0024]
According to the fourth aspect of the invention, since the target flow rate is corrected to complete the filling in the target time, the oil filling can be completed in a certain time, and the pressure control accuracy after the filling period is increased. There is an effect that the shift performance can be improved.
[0025]
According to the fifth aspect of the present invention, it is possible to accurately estimate the oil supply capacity to the friction engagement element in consideration of the oil supply state, the flow rate supplied to other elements such as a torque converter, and the amount of leakage. effective.
[0026]
According to the sixth aspect of the present invention, the oil supply state can be accurately estimated from the rotation speed of the oil pump and the adjustment pressure of the oil discharged from the oil pump, so that the oil to the friction engagement element can be estimated. There is an effect that the supply capacity can be accurately estimated.
[0027]
According to the seventh aspect of the present invention, it is possible to accurately estimate the flow rate of the oil that is supplied to the torque converter and the lubrication path and is not supplied to the friction engagement element, and thus the oil supply capacity to the friction engagement element is increased. There is an effect that it can be estimated with high accuracy.
[0028]
According to the eighth aspect of the present invention, it is possible to accurately estimate the change in the leakage flow rate due to the change in the viscosity of the oil, and thus it is possible to accurately estimate the oil supply capacity to the friction engagement element.
[0029]
DETAILED DESCRIPTION OF THE INVENTION
Embodiments of the present invention will be described below.
FIG. 1 shows a transmission mechanism of a vehicle automatic transmission according to an embodiment, and an output of an engine (not shown) is transmitted to a transmission mechanism 2 via a torque converter 1.
[0030]
The transmission mechanism 2 includes two sets of planetary gears G1, G2, three sets of multi-plate clutches H / C, R / C, L / C, one set of brake bands 2 & 4 / B, and one set of multi-plate brakes L & R /. B, one set of one-way clutch L / OWC.
[0031]
The two sets of planetary gears G1 and G2 are simple planetary gears composed of sun gears S1 and S2, ring gears r1 and r2, and carriers c1 and c2, respectively.
The sun gear S1 of the planetary gear set G1 is configured to be connectable to the input shaft IN by a reverse clutch R / C, and is configured to be fixed by a brake band 2 & 4 / B.
[0032]
The sun gear S2 of the planetary gear set G2 is directly connected to the input shaft IN.
A carrier c1 of the planetary gear set G1 is configured to be connectable to the input shaft I by a high clutch H / C, while a ring gear r2 of the planetary gear set G2 is a carrier of the planetary gear set G1 by a low clutch L / C. The carrier c1 of the planetary gear set G1 can be fixed by a low & reverse brake L & R / B.
[0033]
A ring gear r1 of the planetary gear set G1 and a carrier c2 of the planetary gear set G2 are directly and integrally connected to the output shaft OUT.
In the speed change mechanism 2 configured as described above, the first to fourth speeds and the reverse are realized by a combination of engagement states of the respective clutches and brakes as shown in FIG.
[0034]
In FIG. 2, the circles indicate the engaged state, and the parts not marked with the symbol indicate the released state. In particular, the engaged state indicated by the black circle of the low & reverse brake L & R / B at the first speed. Indicates fastening in only one range.
[0035]
According to the combination of engagement states of the clutches and brakes shown in FIG. 2, for example, at the time of downshift from the 4th speed to the 3rd speed, the brake band 2 & 4 / B is released and the low clutch L / C is engaged. When downshifting from 3rd to 2nd, the high clutch H / C is released and the brake band 2 & 4 / B is engaged. When upshifting from 2nd to 3rd, the brake band 2 & 4 / B And the high clutch H / C are engaged, and at the time of upshift from the third speed to the fourth speed, the low clutch L / C is released and the brake band 2 & 4 / B is engaged.
[0036]
As described above, a shift in which the friction engagement element is switched by simultaneously controlling the engagement and release of the clutch / brake (friction engagement element) is referred to as a switching shift.
[0037]
The engagement / release operation of each clutch / brake (friction engagement element) is controlled by hydraulic pressure, and the hydraulic pressure supplied to each clutch / brake is adjusted by a solenoid valve, as shown in FIG. By such a mechanism, supply of oil (ATF: automatic transmission fluid) to each clutch and brake is controlled.
[0038]
In FIG. 3, the oil discharged from the oil pump 21 that is rotationally driven by the engine is regulated to a predetermined line pressure by the pressure regulating mechanism 22. The oil adjusted to the line pressure is supplied to each friction engagement element 23 via a solenoid valve 24 provided for each friction engagement element 23, and is also supplied to the torque converter 1 and the lubrication path. .
[0039]
The solenoid valve 24 is duty-controlled by a control unit 25. The control unit 25 includes an oil temperature sensor 26 for detecting the oil temperature and an accelerator operated by the driver. An accelerator opening sensor 27 that detects the opening, a vehicle speed sensor 28 that detects the traveling speed of the vehicle, a turbine rotation sensor 29 that detects the turbine rotation speed of the torque converter 1, an engine rotation sensor 30 that detects the engine rotation speed, and the like. Detection signals are input, and by controlling each solenoid valve 24 based on these detection results, the engagement hydraulic pressure of each friction engagement element 23 is controlled.
[0040]
As shown in FIG. 4, the solenoid valve 24 includes a valve body 31, a spool valve 32 that is slidably inserted in the valve body 31 in the axial direction, and a solenoid that displaces the spool valve 32 in the axial direction. 33.
[0041]
The valve body 31 is connected to a hydraulic passage 34, a drain passage 35 from the pressure regulating mechanism 22, and a supply passage 36 for the friction engagement element 23, and the spool valve 32 selects the hydraulic passage 34 and the drain passage 35. Thus, the operation for putting oil into the friction engagement element 23 and the operation for removing oil are controlled.
[0042]
In addition, the pressure in the supply passage 36 acts on the spool valve 32 via the feedback passage 38 provided with the orifice 37 in the direction in which the hydraulic passage 34 is closed and the drain passage 35 is opened (the leftward direction in the figure). It is configured as follows.
[0043]
Further, the spring 39 is provided to urge the spool valve 32 to the right in the drawing.
Accordingly, the electromagnetic force of the solenoid 33 acts against the biasing force of the spring 39, so that the spool valve 32 is displaced in the left direction in the figure. The electromagnetic force of the solenoid 33 is increased, so that the spool The valve 32 is displaced to the left in the figure, and the drain is increased.
[0044]
As shown in FIG. 5, in the switching shift according to the present embodiment, the engagement hydraulic pressure of the friction engagement element to be fastened is gradually increased while the engagement hydraulic pressure of the friction engagement element to be released is gradually decreased, and the release side The torque is switched from the friction engagement element to the fastening side friction engagement element.
[0045]
In addition, when a speed change request that requires the engagement operation of the friction engagement element is generated, first, an instruction pressure higher than the initial pressure of the engagement control is output, whereby oil is rapidly applied to the friction engagement element to be engaged. The pre-charging is performed to fill the friction engagement element with oil, and thereafter, the engagement hydraulic pressure is gradually increased and controlled. This pre-charge will be described in detail below.
[0046]
The flowchart of FIG. 6 shows the main routine of the precharge control.
The precharge control (flow rate control) is started based on the shift determination. As will be described later, the precharge control (flow rate control) is determined to end when the clutch reaction force on the engagement side exceeds a predetermined value, and the control proceeds to pressure control. During (flow rate control), when the turbine rotation speed falls below the first reference speed (when torque is lost) and when the turbine rotation speed exceeds the second reference speed (> first reference speed) (when idling occurs) ) Is forcibly shifted to pressure control.
[0047]
In step S1, a target flow rate of oil supplied to the frictional engagement element to be fastened is calculated.
The calculation of the target flow rate in step S1 is shown in detail in the flowchart of FIG.
[0048]
In step S101, the volume Vc (cc) for filling the oil is set as the sum of the pipe volume and the friction engagement element volume.
The filling volume Vc may be stored in advance for each frictional engagement element to be fastened, and the stored value may be referred to according to the frictional engagement element to be fastened.
[0049]
In step S102, a target time (target filling time) Tgt-TIME (sec) for completing filling is set.
The target filling time Tgt-TIME may be a fixed value, but is preferably changed according to the oil temperature (ATF temperature) as shown in FIG. When the target filling time Tgt-TIME is set according to the oil temperature (ATF temperature), as shown in FIG. 8, the target filling time Tgt-TIME is shortened as the oil temperature at which the oil viscosity decreases is higher. Good.
[0050]
In step S103, the target flow rate Tgt-Q (cc / sec) is calculated from the filling volume Vc and the target filling time Tgt-TIME according to the following equation.
Target flow rate Tgt-Q = [Filling volume Vc] / [Target filling time Tgt-TIME]
In step S2, a dischargeable flow rate Can-Q (cc / sec) is calculated.
[0051]
Specifically, as shown in FIG. 9, a map in which the dischargeable flow rate Can-Q (cc / sec) is stored in advance according to the line pressure PL (Kpa) and the engine rotation speed (rpm) is referred to. The dischargeable flow rate Can-Q (cc / sec) corresponding to the line pressure PL and the engine rotation speed (rpm) is searched.
[0052]
In this embodiment, since the oil pump 21 is driven by the engine, the engine rotation speed (rpm) is used as a value proportional to the rotation speed of the oil pump 21. Therefore, the rotational speed of the oil pump 21 may be obtained and used for calculating the dischargeable flow rate Can-Q.
[0053]
In step S3, the flow rate of oil supplied to the torque converter 1 and the lubrication circuit is calculated as the required flow rate Require-Q (cc / sec) from the dischargeable flow rate Can-Q.
[0054]
Specifically, as shown in FIG. 10, a table in which the required flow rate Require-Q (cc / sec) is stored in advance according to the engine speed (rpm) is referred to, and the engine speed (rpm) at that time is referred to. Search the corresponding required flow rate Require-Q (cc / sec).
[0055]
Since the flow rate of oil supplied to the torque converter 1 and the lubrication path increases as the engine speed (rpm) increases, the required flow rate Require-Q (cc / sec) increases as the engine speed (rpm) increases. It is set to a large value.
[0056]
In step S4, a loss due to leakage from the dischargeable flow rate Can-Q is calculated as a leakage flow rate Leak-Q (cc / sec).
Specifically, as shown in FIG. 11, a leak flow rate corresponding to the oil temperature at that time is referenced by referring to a table in which the leak flow rate Leak-Q (cc / sec) is stored in advance according to the oil temperature (ATF temperature). Search for Leak-Q (cc / sec).
[0057]
When the oil temperature is high, the oil viscosity decreases and the amount of oil leakage increases, so the leakage flow rate Leak-Q (cc / sec) is set to a larger value as the oil temperature increases.
In step S5, the required flow rate Require-Q and the leakage flow rate Leak-Q are subtracted from the dischargeable flow rate Can-Q, and the result is the limit flow rate Limit- that can actually be supplied to the engagement side frictional engagement element. Q (cc / sec).
[0058]
Limit-Q (cc / sec) = [Can-Q]-([Require-Q] + [Leak-Q])
In step S6, based on the deviation between the target flow rate Tgt-Q and the limit flow rate Limit-Q and the oil temperature correlated with the oil viscosity, the precharge hydraulic pressure (target clutch pressure in precharge) P-PRI (Kpa) Set.
[0059]
Specifically, as shown in FIG. 12, a map storing precharge hydraulic pressure P-PRI in advance according to the deviation between the target flow rate Tgt-Q and the limit flow rate Limit-Q and the oil temperature is referred to. The precharge hydraulic pressure P-PRI corresponding to the time deviation and oil temperature is searched.
[0060]
Here, the target flow rate Tgt-Q, is set as the pre-charge oil pressure P-PRI (command pressure) is high when there is little limit flow Limit-Q, and a target flow rate Tgt-Q and limit flow rate Limit-Q For this deviation, the precharge hydraulic pressure P-PRI (indicated pressure) is set higher as the oil temperature is lower and the viscosity is higher.
[0061]
The precharge hydraulic pressure P-PRI setting causes changes in oil supply capacity (dischargeable flow rate Can-Q, required flow rate Require-Q, leakage flow rate Leak-Q) to the friction engagement elements, and variations in flow rate. Even if there is a change in the viscosity of the oil to be used, the oil can be filled at the desired flow rate, preventing the filling time from being prolonged and preventing large pressure fluctuations from filling due to excessive flow rate. it can.
[0062]
The precharge hydraulic pressure P-PRI is continuously output until completion of precharge (flow rate control) is determined, and then the target clutch pressure (indicated pressure) is reduced to the initial pressure for pressure control, The clutch pressure (indicated pressure) is gradually increased with a predetermined ramp gradient to shift to pressure control for fastening the friction engagement element (see FIG. 5).
[0063]
In this embodiment, when the end of precharge (flow rate control) is determined, the precharge hydraulic pressure P-PRI is gradually changed from the initial pressure of pressure control at a predetermined time TIMER1.
[0064]
Specifically, when the initial pressure of pressure control is P-RTN-α, the elapsed time from the end of precharge (flow control) is t, and the gain is α, the indicated pressure Pc0 within the predetermined time TIMER1 is ,
Pc0 = P-PRI × (1−α × t ^1/2 )
Asking.
[0065]
The gain α is a value set such that the command pressure Pc0 = the initial pressure P-RTN-α when the elapsed time t is a predetermined time TIMER1.
The control so far is shown in the control block diagram of FIG.
[0066]
That is, the target flow rate is obtained from the target filling time and the volume (volume), the dischargeable flow rate is obtained from the line pressure and the engine rotational speed, the necessary flow rate to be supplied to the torque converter is obtained from the engine rotational speed, and the oil temperature To obtain the leakage flow rate.
[0067]
Then, the result of subtracting the required flow rate and the leakage flow rate from the dischargeable flow rate is defined as a limit flow rate that can be supplied to the friction engagement element, and according to the deviation between the target flow rate and the limit flow rate and the oil temperature (oil viscosity) By setting the precharge hydraulic pressure (indicated pressure at the time of precharge), the target flow rate is ensured, that is, the filling is completed in the target filling time.
[0068]
In the flowchart of FIG. 6, in step S7 and subsequent steps, processing for determining whether to shift to the pressure control is performed.
In step S7, an inflow flow rate Qs for the solenoid valve 24 is calculated.
[0069]
The solenoid inflow flow rate Qs is defined as follows: C is the oil flow coefficient, A is the opening area of the hydraulic passage 34 controlled by the solenoid valve 24, PL is the line pressure, Real-Pc is the clutch hydraulic pressure, and ρ is the oil density.
Qs = C · A · {(PL-Real-Pc) / ρ} ^1/2 (1)
Is calculated as
[0070]
Therefore, in step S7, the solenoid inflow rate Qs is calculated as shown in the flowchart of FIG.
Hereinafter, the calculation of the solenoid inflow flow rate Qs will be described according to the flowchart of FIG. 14 with reference to the control block diagram of FIG.
[0071]
In step S701, in order to obtain the opening area A, first, a solenoid displacement amount X (cm) is calculated.
In the present embodiment, a target clutch oil pressure (indicated pressure) is determined, and the solenoid valve 24 is driven with a duty corresponding to the target clutch oil pressure (indicated pressure). Therefore, the drive duty DUTY (%) of the solenoid is obtained from the target clutch oil pressure (indicated pressure) at that time with reference to a table as shown in FIG.
[0072]
Next, the solenoid drive duty DUTY is converted into a solenoid drive current I (A) by a table as shown in FIG.
Further, the solenoid drive current I (A) is converted into a solenoid suction force Fsol (Kgf) by a table as shown in FIG.
[0073]
Here, as shown in FIG. 19, the spool valve 32 is displaced to a position where the load of the spring 39 balances the suction force (electromagnetic force) Fsol (Kgf) of the solenoid and the feedback force via the feedback passage 38.
[0074]
Therefore, if the set load of the spring 39 is Fset (Kgf), the spring constant of the spring 39 is Kx, the clutch hydraulic pressure is Real-Pc, and the area of the spool valve 32 on which the feedback force acts is Afb.
Fset + Kx / X = Fsol + Real-Pc / Afb
From the above equation, the solenoid displacement X (cm) is
X = (Fsol + Real-Pc.Afb-Fset) / Kx
Will be required.
[0075]
The calculation of the clutch hydraulic pressure Real-Pc will be described later.
When the solenoid displacement amount X (cm) is obtained as described above, in the next step S702, the opening area A (opening area of the hydraulic pressure supply port) of the solenoid valve 24 is obtained from the solenoid displacement amount X (cm).
[0076]
Specifically, as shown in FIG. 20, a table indicating the correlation between the solenoid displacement amount X and the opening area A is stored in advance, and the solenoid displacement amount X at that time is converted into the opening area A by the table. .
[0077]
Subsequently, in step S703, the flow coefficient C is calculated.
The calculation of the flow coefficient C is performed as shown in the block diagram of FIG.
First, referring to a table in which the viscosity μ is stored in advance according to the oil temperature, the viscosity μ at the oil temperature at that time is obtained, and the ratio of this viscosity μ to the viscosity μ at the reference oil temperature (for example, 80 ° C.) Is calculated.
[0078]
Then, based on the ratio of the flow coefficient C at the reference oil temperature (for example, 80 ° C.) and the viscosity μ, the flow coefficient C corresponding to the oil temperature at that time is obtained.
In step S704, the solenoid inflow flow rate Qs is based on the opening area A, the flow coefficient C and the density ρ corresponding to the oil temperature obtained as described above, and the clutch hydraulic pressure Real-Pc obtained as described later. Is calculated according to the equation (1).
[0079]
In the flowchart of FIG. 6, when the solenoid inflow rate Qs is calculated in step S7, the clutch inflow rate (solenoid discharge flow rate) Qc is calculated in the next step S8, and the clutch reaction force is calculated in step S9.
[0080]
The clutch inflow flow rate (solenoid discharge flow rate) Qc, the clutch hydraulic pressure Real-Pc, and the clutch reaction force can be calculated by solving simultaneous equations of the following equations (2) to (7).
[0081]
Mc · ΔΔYc + Cc · ΔYc + Kc · (Yc + Yco) = Ac · ΔReal−Pc (2)
Vc = Vo + Ac · Yc (3)
Qs−Qc = Vc / K · ΔReal-Pc (4)
Qc = Ac · ΔYc (5)
Real-Pc = Σ (ΔReal-Pc) (6)
Total-Qc = Σ (Qc) (7)
In the above equation, Yc clutch displacement (cm), Yc clutch displacement of the differential value (cm / 10msec), ΔΔYc the differential value of the differential value of the clutch displacement (cm / 10msec ^2), Ac is the clutch piston pressure receiving Area (cm ² ), Cc is the flow coefficient, Mc is the clutch piston load (Kg), Kc is the clutch piston spring constant (Kg / cm), K is the bulk modulus (Kgf / cm ² ), and Vc is the capacity (cc) , Yco is the clutch piston initial set displacement (cm), Total-Qc is the integrated solenoid discharge flow rate, ΔReal-Pc is the differential value of the clutch hydraulic pressure, and Vo is the initial capacity (cc).
[0082]
The clutch piston pressure receiving area Ac, the initial capacity Vo, the clutch piston load Mc, the clutch piston spring constant Kc, and the clutch piston initial set displacement Yco are fixed values given in advance.
[0083]
The bulk modulus K may be a fixed value or may be calculated according to the following formula.
K = Vo / (Vo-Total-Qc) · ΔReal-Pc
As shown in the control block diagram of FIG. 22, by substituting the solenoid inflow flow rate Qs, the clutch inflow flow rate (solenoid discharge flow rate) Qc, the capacity Vc, and the bulk modulus K into the equation (4) (continuous equation). The differential value ΔReal-Pc of the clutch hydraulic pressure is obtained, and the differential value ΔReal-Pc of the clutch hydraulic pressure is integrated to obtain the clutch hydraulic pressure Real-Pc.
[0084]
On the other hand, the equation of motion shown in equation (2) is
Mc · ΔΔYc = Ac · ΔReal−Pc−Cc · ΔYc−Kc · (Yc + Yco)
If Mc · ΔΔYc is obtained from the above equation, ΔΔYc is obtained because the clutch piston load Mc is a known value.
[0085]
Then, ΔYc is obtained by integrating ΔΔYc, and Yc is obtained by integrating ΔYc.
When ΔYc is obtained, the clutch piston pressure receiving area Ac is a known value, and hence the clutch inflow rate (solenoid discharge flow rate) Qc is obtained from the equation (5).
[0086]
Further, Cc · ΔYc in the equation of motion shown in the equation (2) is obtained from ΔYc.
Furthermore, from Yc, a capacity Vc that changes due to oil filling according to equation (3) and Kc · (Yc + Yco) in the equation of motion shown in equation (2) are obtained.
[0087]
Here, in the released state of the clutch, the solenoid opening area A = 0, the capacity Vc = Vo, the clutch displacement amount Yc = 0, the clutch hydraulic pressure Real-Pc = 0, the solenoid inflow flow rate Qs = 0, the clutch inflow flow rate (solenoid discharge flow rate). ) Since Qc = 0, by repeating the calculation with this state as an initial value, the clutch inflow flow rate (solenoid discharge flow rate) Qc that changes with precharging, the clutch hydraulic pressure Real-Pc, and Kc · (Yc + Yco) is determined.
[0088]
In the block diagram of FIG. 22, the differential value is indicated by a dot on the symbol, and the symbol with two dots indicates that it is a differential value of the differential value.
In step S10 of the flowchart of FIG. 6, it is determined to switch from flow control (precharge control) for filling oil to pressure control for controlling the engagement pressure (transmission torque capacity) of the friction engagement element to the target pressure.
[0089]
Specifically, as shown in the flowchart of FIG. 23, first, in step S1101, the clutch reaction force Kc · (Yc + Yco) is compared with a predetermined value, and if the clutch reaction force Kc · (Yc + Yco) is less than or equal to the predetermined value. In step S1102, the flow control (precharge) is continued and the precharge hydraulic pressure P-PRI is output as the command pressure.
[0090]
On the other hand, when the clutch reaction force Kc · (Yc + Yco) exceeds a predetermined value, it is determined that precharge is completed, and the process proceeds to step S1103 to shift to pressure control for controlling the clutch hydraulic pressure to the target pressure.
[0091]
As described above, if the configuration is such that the shift to the pressure control is determined based on the clutch reaction force Kc · (Yc + Yco), the shift to the pressure control can be performed after the precharge is actually completed. The control accuracy of the clutch hydraulic pressure can be improved.
[0092]
The configuration may be such that the time when the target filling time has elapsed since the start of precharge is determined as the time when flow control (precharge control) is completed.
When the transition from flow rate (pre-charge) control to pressure control is determined, as indicated above, the indicated pressure is gradually reduced to the initial pressure P-RTN-α and then the fastening side command is issued with a predetermined ramp. The friction engagement element is fastened by increasing the pressure.
[0093]
In the ramp control, the target clutch hydraulic pressure (indicated pressure) is determined so that the share of the transmission torque capacity commensurate with the input shaft torque of the transmission is gradually shifted from the release side to the engagement side, and the target clutch hydraulic pressure is controlled by the control duty. And the control duty is output to the solenoid valve 24.
[0094]
Further, since the gain of the actual pressure (clutch hydraulic pressure) with respect to the command pressure changes according to the oil temperature (viscosity), it is preferable that the command pressure is corrected according to the oil temperature in the ramp control.
[0095]
However, the content of the pressure control after the precharge control is not limited to the above, and the above precharge control is applied if the precharge is performed when the friction engagement element is fastened. And thereby achieve a similar effect.
[Brief description of the drawings]
FIG. 1 is a diagram showing a transmission mechanism of an automatic transmission according to an embodiment.
FIG. 2 is a diagram showing a correlation between a combination of engagement states of friction engagement elements in the speed change mechanism and a gear position;
FIG. 3 is a system diagram showing a hydraulic control system of the automatic transmission.
FIG. 4 is a sectional view showing details of a solenoid valve in the hydraulic control system.
FIG. 5 is a time chart showing a state of shifting by changing friction engagement elements in the embodiment.
FIG. 6 is a flowchart showing a main routine of precharge control in the embodiment.
FIG. 7 is a flowchart showing a target flow rate calculation routine in the precharge control according to the embodiment;
FIG. 8 is a diagram showing a table of oil temperature → target filling time in the embodiment.
FIG. 9 is a diagram showing a map of line pressure and engine rotational speed → dischargeable flow rate in the embodiment.
FIG. 10 is a diagram showing a table of oil temperature → required flow rate in the embodiment.
FIG. 11 is a diagram showing a table of oil temperature → leakage flow rate in the embodiment.
FIG. 12 is a diagram showing a map of (target flow rate−limit flow rate) and oil temperature → precharge oil pressure in the embodiment;
FIG. 13 is a block diagram showing precharge control according to the deviation between the target flow rate and the limit flow rate in the embodiment.
FIG. 14 is a flowchart showing a routine for calculating a solenoid inflow rate in the precharge control according to the embodiment.
FIG. 15 is a block diagram showing calculation control of the solenoid inflow flow rate in the precharge control according to the embodiment.
FIG. 16 is a diagram showing a table of target clutch pressure → solenoid drive duty in the embodiment;
FIG. 17 is a diagram showing a table of solenoid drive duty → solenoid drive current in the embodiment;
FIG. 18 is a diagram showing a table of solenoid drive current → solenoid attraction force in the embodiment;
FIG. 19 is a state diagram showing a load balance state of the solenoid valve.
FIG. 20 is a diagram showing a table of solenoid displacement → opening area in the embodiment;
FIG. 21 is a block diagram showing calculation control of a flow coefficient in the embodiment.
FIG. 22 is a block diagram showing calculation control of the clutch inflow rate, clutch hydraulic pressure, and clutch reaction force in the embodiment.
FIG. 23 is a flowchart showing switching determination from flow rate control to pressure control in the embodiment;
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 ... Torque converter 2 ... Transmission mechanism 21 ... Oil pump 22 ... Pressure regulation mechanism 23 ... Friction engagement element 24 ... Solenoid valve 25 ... Control unit 26 ... Oil temperature sensor 27 ... Accelerator opening sensor 28 ... Vehicle speed sensor 29 ... Turbine rotation Sensor 30 ... Engine rotation sensor 31 ... Valve body 32 ... Spool valve 33 ... Solenoid 34 ... Hydraulic passage 35 ... Drain passage 36 ... Supply passage 37 ... Orifice 38 ... Feedback passage 39 ... Spring G1, G2 ... Planetary gear H / C ... High Clutch R / C ... Reverse clutch L / C ... Low clutch 2 & 4 / B ... 2-speed / 4-speed band brake L & R / B ... Low & reverse brake

Claims (8)

For vehicles in which the engagement / release of the frictional engagement element is controlled by hydraulic pressure to change the speed, and the opening area of the solenoid that controls the supply of hydraulic pressure to the frictional engagement element is increased in response to an increase in the indicated pressure . In the automatic transmission hydraulic control device,
Based on the target flow rate when the friction engagement element to be fastened is filled with oil, the limit flow rate that can be supplied to the friction engagement element to be fastened, and the viscosity of the oil, setting the command pressure at the time of precharging A hydraulic control device for an automatic transmission for a vehicle. Wherein the target flow rate, the limit flow rate is low Tokiho etc. before SL and sets larger command pressure claim 1 hydraulic control device for a vehicular automatic transmission according. It is configured to set the indicated pressure at the time of precharging based on the deviation between the target flow rate and the limit flow rate, and the viscosity of the oil, and
The higher the oil viscosity, the higher the limit flow rate is estimated, while the higher the oil viscosity is, the higher the indicated pressure is set for the deviation between the target flow rate and the limit flow rate. The hydraulic control apparatus of the automatic transmission for vehicles as described. The hydraulic control device for an automatic transmission for a vehicle according to any one of claims 1 to 3, wherein the target flow rate is calculated based on a target filling time and a filling volume. The vehicle automatic transmission according to any one of claims 1 to 4, wherein the limit flow rate is calculated based on a dischargeable flow rate, a flow rate required for other elements, and a leakage flow rate. Hydraulic control device for the machine. 6. The hydraulic control device for an automatic transmission for a vehicle according to claim 5, wherein the dischargeable flow rate is calculated based on an original pressure of oil and a rotational speed of an engine combined with the automatic transmission. The hydraulic control apparatus for an automatic transmission for a vehicle according to claim 5 or 6, wherein a flow rate required for the other elements is calculated based on a rotational speed of an engine combined with the automatic transmission. The hydraulic control device for an automatic transmission for a vehicle according to any one of claims 5 to 7, wherein the leakage flow rate is calculated based on an oil temperature.

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