JP5569320B2 - Turbine speed estimation device for torque converter - Google Patents

Description

  The present invention relates to a turbine speed estimating device for a torque converter provided between an engine and an automatic transmission.

The rotation of an engine mounted on an automobile or the like is shifted through a transmission and output to driving wheels. When the transmission is an automatic transmission (a stepped automatic transmission such as a stepped automatic transmission or CVT), a torque converter (hereinafter also referred to as a torque converter) is often provided between the engine and the automatic transmission. ing.
In an automobile having such a torque converter, there is an opportunity to switch between a P or N range, which is in a neutral state, and a travel range such as a D range or an R range by a driver's shift lever operation. In such a situation, since the engine load (hereinafter referred to as torque converter load) due to the torque converter reaction force changes transiently, there is a risk of fluctuations in the engine speed.

  In order to prevent this, it is necessary to estimate the torque converter load and suppress fluctuations in the engine rotational speed. To this end, it is necessary to recognize the input / output rotational speed of the torque converter, Although the engine speed sensor can be used for the speed, the output speed of the torque converter, that is, the turbine speed is required. If a turbine rotational speed sensor is equipped, the turbine rotational speed can be obtained directly, but in order to reduce the cost of the product, it is desired to omit the turbine rotational speed sensor.

In this regard, in Patent Document 1, in an A / T vehicle and a CVT vehicle having a torque converter, the rotational speed is calculated from the accelerator opening, the engine rotational speed Ne, the vehicle speed, and the torque converter characteristics to estimate the input rotational speed. A technique has been proposed in which a sensor for detecting the turbine rotational speed Nt can be omitted.
Patent Document 2 uses the fact that the turbine shaft and the primary shaft can be regarded as the same rotational speed during traveling, and calculates from the primary rotational speed Np or the secondary rotational speed Ns and the primary-secondary reduction ratio. Therefore, a technique has been proposed in which a sensor for detecting the turbine rotation speed Nt can be omitted.

Japanese Patent No. 3104503 JP-A-8-35437

  By the way, in an automatic transmission, the turbine rotational speed Nt may change depending on the operation state of a clutch or brake (hereinafter also simply referred to as an ND clutch) of a forward / reverse switching mechanism in the transmission. It is necessary to grasp the turbine rotation speed Nt and to enable the engine to be operated according to the load torque received by the engine from the torque converter, and to suppress fluctuations in the engine rotation speed due to the change in the load torque.

  Specifically, in the operation state of the ND clutch in this case, the shift range of the transmission is switched from the stop range (P range or N range) to the travel range (R or D range, etc.) by operating the shift lever. A section in which the turbine shaft and the primary shaft are engaged and the turbine rotational speed Nt is decreasing, or conversely, the transmission shift range is changed from the travel range (R or D range, etc.) to the stop range (P range). Alternatively, there is a section where the turbine shaft and the primary shaft are opened and the turbine rotational speed Nt is increased.

In such a section, the accurate turbine rotation speed cannot be obtained by the estimation methods according to Patent Documents 1 and 2, and the engine cannot be operated according to the load torque received from the torque converter. As a result, there is a problem that the drivability of the automobile deteriorates due to fluctuations in the engine rotational speed Ne.
The present invention was devised in view of such a problem, and even if a change occurs in the turbine rotational speed Nt due to the operating state of the forward / reverse switching mechanism in the automatic transmission, the turbine does not depend on the turbine rotational speed sensor. It is possible to estimate the rotation speed and accurately calculate the load torque that the engine receives from the torque converter, so that the drivability improvement of the vehicle, such as prevention of engine rotation fluctuation, can be realized. An object of the present invention is to provide a turbine rotational speed estimation device.

In order to achieve the above object, a torque converter turbine speed estimation device according to the present invention includes a vehicle equipped with an engine and a hydraulic friction engagement element connected to and disconnected from the vehicle, and the transmission / non-transmission of power. The torque in a drive system comprising: an automatic transmission having a forward / reverse switching mechanism for switching transmission; and a torque converter interposed between the engine and the forward / reverse switching mechanism of the automatic transmission. An apparatus for estimating the turbine speed of the converter, the engine speed detecting means for detecting the engine speed Ne, and the oil pressure estimation for estimating the hydraulic pressure P of the hydraulic oil when the hydraulic friction engagement element is connected / disconnected. And an engagement torque estimation means for estimating an engagement torque TCLT of the hydraulic friction engagement element based on the hydraulic pressure P of the hydraulic oil estimated by the hydraulic pressure estimation means The engagement torque TCLT of the hydraulic friction engagement element that is estimated by the engagement torque estimating means, on the basis of the rotational speed Ne of the engine detected by the engine rotational speed detecting means, the hydraulic friction Turbine rotational speed calculating means for calculating a turbine rotational speed Nt of the torque converter when the engaging element is connected / disconnected .

The turbine rotation speed calculation means, based on the rotation speed Ne of the engine, and calculates the transmission torque TauCNe ² inputted from the engine to the automatic transmission, and the engagement torque T _CLT and the transmission torque TauCNe ² It is preferable to calculate the turbine rotation speed Nt based on the difference.
The hydraulic friction engagement element is connected / disconnected by controlling the hydraulic pressure by an electromagnetic valve, and the hydraulic pressure estimation means is operated by passing through the electromagnetic valve based on a drive current value I of a solenoid of the electromagnetic valve. A volume flow rate Qin of oil is calculated, a gap h between elements of the hydraulic friction engagement element is calculated based on the calculated volume flow rate Q _in , and the hydraulic pressure of the hydraulic oil is calculated from the calculated gap h between the elements. It is preferable to estimate P.

The hydraulic estimating means, in the calculated hydraulic oil volume flow Q _{in (n),} the calculated volumetric flow rate Q _{in (n)} the hydraulic friction engagement element that passes through the solenoid valve at predetermined intervals Based on the corrected volume flow rate corrected by the outflow flow rate _Qout, the hydraulic fluid volume Q (n) in the hydraulic chamber of the hydraulic friction engagement element is calculated, and the calculated hydraulic fluid volume Q (n) in the hydraulic chamber is calculated. It is preferable to calculate the gap h (n) based on

  The hydraulic friction engagement element is connected / disconnected by a piston that moves under controlled hydraulic pressure by the solenoid valve, and the hydraulic pressure estimation means includes a previously obtained hydraulic fluid volume in the hydraulic chamber, a piston position, and the like. And the correspondence relationship between the piston position and the gap between the elements obtained in advance are used to calculate the gap h (n) based on the hydraulic oil volume Q (n) in the hydraulic chamber. It is preferable.

The engaging torque estimating means, as the sum of the contact friction torque T _{FR C} corresponding to the contact friction between the viscous resistance torque T _DRG according to the viscous resistance of the hydraulic fluid the hydraulic friction engagement element, the engaging It is preferable to calculate the torque _TCLT .
Transmission input rotation speed detection means for detecting a transmission input rotation speed Np input from the forward / reverse switching mechanism to the transmission mechanism of the automatic transmission is provided, and the viscous resistance torque _TDRG is obtained from the turbine rotation obtained last time. The difference between the speed Nt (n−1) and the transmission input rotational speed Np detected by the transmission input rotational speed detection means, the number of engagement surfaces and the effective radius (r _{1) of the} hydraulic friction engagement element , R ₂ ) and the viscous resistance coefficient μ _ATF of the hydraulic oil, the contact friction torque T _{FR C} is determined by the hydraulic pressure P (n) of the hydraulic oil and the hydraulic pressure P (n). It is preferable to calculate based on the effective area nS, effective radius [(r ₁ + r ₂ ) / 2] and friction coefficient μ _FRC of the engagement surface of the hydraulic friction engagement element.

  According to the present invention, the turbine rotational speed can be estimated without relying on the rotational speed sensor, for example, when switching between a traveling range and a non-traveling range during idling, or regardless of the driver's shift lever operation when the vehicle is stationary. When performing control that automatically shifts to a neutral state (so-called idle neutral control), the torque converter load can be accurately calculated, and the engine speed is stabilized while monitoring the torque converter load. It is also possible to control to show the behavior.

  The estimated turbine rotation speed is not limited to such control, and can be applied to various other controls.

1 is a configuration diagram illustrating an engine, a transmission, a torque converter, and a turbine rotation speed estimation device according to an embodiment of the present invention. It is principal part sectional drawing which shows the structure of ND clutch with which the transmission concerning one Embodiment of this invention was equipped. It is a figure explaining the characteristic of the capacity coefficient used for estimation of the turbine speed estimating device concerning one embodiment of the present invention. It is a figure explaining the characteristic of the torque ratio used for estimation of the turbine speed estimation apparatus concerning one Embodiment of this invention. It is a figure which shows the operation | movement model of ND clutch for demonstrating simulation of the piston action hydraulic pressure P of ND clutch used for estimation of the turbine speed estimation apparatus concerning one Embodiment of this invention. ND for explaining solenoid valve passage volume flow rate Qin (n) per unit time and volume Q (n) existing in the hydraulic cylinder of the ND clutch used for estimation of the turbine speed estimation device according to the embodiment of the present invention. It is a figure which shows the action | operation model of a clutch. It is a schematic diagram explaining a mode that a piston engages a clutch board in ND clutch with which the transmission concerning one Embodiment of this invention was equipped. It is a figure explaining the characteristic of simulation oil pressure P (n) which simulates the engagement oil pressure of the ND clutch with which the transmission concerning one embodiment of the present invention is equipped corresponding to the clutch board gap of the ND clutch. It is a main flowchart explaining the procedure of estimation by the turbine speed estimation apparatus concerning one Embodiment of this invention. It is a subroutine flowchart explaining the procedure of estimation by the turbine speed estimation apparatus concerning one Embodiment of this invention.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
1 to 10 illustrate an embodiment of the present invention. FIG. 1 is a configuration diagram showing an engine, a transmission, a torque converter, and a turbine rotational speed estimation device, and FIG. 2 is a cross-sectional view of a main part showing a configuration of an ND clutch. FIG. 3 is a characteristic diagram of the capacity coefficient, FIG. 4 is a characteristic diagram of the torque ratio, FIGS. 5 and 6 are operation model diagrams of the ND clutch, FIG. 7 is an explanatory diagram of piston operation of the ND clutch, and FIG. FIG. 9 and FIG. 10 are flowcharts for explaining the estimation procedure.

<Configuration of main part of drive system>
First, a configuration of a main part of a drive system of an automobile having a torque converter that is an estimation target of the turbine rotational speed estimation device will be described.

As shown in FIG. 1, the drive system includes an engine (internal combustion engine) 2, an automatic transmission 4, and a torque converter (hereinafter also referred to as torque converter) 6 interposed between the engine 2 and the automatic transmission 4. The output shaft of the automatic transmission 4 is connected to driving wheels (not shown), and the rotation of the engine 2 is shifted and transmitted to the driving wheels for driving.
In the torque converter 6, an impeller 64 coupled to the output shaft 22 of the engine 2 and a turbine liner (also simply referred to as a turbine) 66 coupled to the input shaft 42 of the automatic transmission 4 are opposed to each other in the torque converter casing 62. A viscous fluid (not shown) is interposed between the impeller 64 and the turbine liner 66, and power is transmitted between the impeller 64 and the turbine liner 66 through the viscous fluid.

  In the case of this embodiment, a belt type continuously variable transmission (CVT) is applied to the automatic transmission 4. That is, the automatic transmission 4 includes a primary shaft 46A connected to the input shaft 42 via the forward / reverse switching mechanism 44 in the casing 40, a primary pulley 46 mounted on the primary shaft 46A, and a primary shaft 46A. A secondary shaft 48A equipped in parallel, a secondary pulley 48 equipped on the secondary shaft 48A, and a primary pulley 46 and a belt 50 spanned between the secondary pulley 48, and the primary pulley 46 and the secondary pulley 48 are provided. The gear ratio can be continuously changed by moving the movable pulley in the axial direction to change the effective diameter of each of the pulleys 46 and 48.

The secondary shaft 48A is connected to the output shaft 54 via a gear pair 54a. The rotation shifted by the automatic transmission 4 is transmitted to driving wheels (not shown), and the vehicle is driven by the rotation of the driving wheels.
An electronic control unit (ECU) 10 that controls the engine 2 and the automatic transmission 4 is also provided. The ECU 10 includes an input / output device, a storage device (ROM, RAM, nonvolatile RAM, etc.), a central processing unit (CPU), a timer counter, and the like.

  Further, a crank angle sensor (engine rotation speed sensor as engine rotation speed detection means) 20a that detects the rotation angle of the crankshaft of the engine 2 and outputs it as engine rotation speed information, and the rotation speed of the primary shaft 46A of the automatic transmission 4 Primary rotation speed sensor (primary rotation speed detection means) 20b is detected, and detection information of each sensor 20a, 20b is sent to the ECU 10.

<Main part configuration of forward / reverse switching mechanism>
Although not shown in detail in the forward / reverse switching mechanism 44, for example, a sun gear, a ring gear, a first pinion gear that meshes with the sun gear, a second pinion gear that meshes with the ring gear, and the first and second pinions It can be constituted by a double pinion type planetary gear mechanism comprising a carrier that pivotally supports a gear. In this case, the sun gear is connected to the input shaft 42, the carrier is connected to the primary shaft 46A, and is connected to the casing 40 via a pinion gear, a ring gear, and a reverse brake.

  When the vehicle moves forward, the forward clutch is engaged and the reverse brake is released, and when the vehicle moves backward, the reverse brake is engaged and the forward clutch is released. As for switching between transmission and non-transmission of power, when the torque transmission between the input shaft 42 and the primary shaft 46A is cut off, both the forward clutch and the reverse brake are released. When torque is transmitted between the input shaft 42 and the primary shaft 46A, either the forward clutch or the reverse brake is coupled.

  Such clutches and brakes provided in the forward / reverse switching mechanism 44 are generally constituted by a wet multi-plate clutch device driven by a hydraulic method. In addition, since there is switching between the neutral range (N range) and the drive range (D range) as a representative switching of the forward / reverse switching mechanism 44, such a clutch and a brake are also referred to as an ND clutch.

For example, FIG. 2 shows an example of a wet multi-plate clutch device 52 as a clutch that frictionally engages the first rotating element 44a of the forward / reverse switching mechanism 44 with the second rotating element 44b. If either the first rotation element 44a or the second rotation element 44b is not rotating, it is called a brake.
As shown in FIG. 2, the wet multi-plate clutch device 52 includes a cylinder 52a (here, an annular cylinder) 52a provided with a clutch plate holding portion 52d, which is equipped to rotate integrally with the first rotating element 44a. A piston (here, an annular piston) 52b mounted in the cylinder 52a so as to be movable in the axial direction, a hydraulic chamber 52c surrounded by an inner wall of the cylinder 52a and a pressure-receiving surface of the piston 52b, a second rotating element 44b, A clutch disk holding member 52e that rotates integrally is provided, and a hydraulic oil supply port 52f is provided in the hydraulic chamber 52c.

  The clutch plate holding portion 52d of the cylinder 52a is equipped with a plurality of clutch plates 52g that rotate integrally with the cylinder 52a (first rotating element 44a) and move in the axial direction, and the clutch disk holding member 52e includes a clutch plate 52g. A plurality of clutch disks 52h that rotate integrally with the disk holding member 52e (second rotating element 44b) and are movable in the axial direction are arranged between the clutch plates 52g and are mounted in series. Each clutch plate 52g is a metal plate having a rough surface, and a friction material (facing) is attached to the surface of each clutch disk 52h facing the surface of the clutch plate 52g.

  The clutch plate 52g and the clutch disk 52h are engaged (contacted) when pressed by the piston 52b, and their relative rotation is restricted by the frictional force generated according to the pressing force (pressing force) at this time. In this example, when the clutch plate 52g and the clutch disk 52h are engaged (sliding friction state), the differential between the first rotating element 44a and the second rotating element 44b is restricted, and the rotational speed difference between the two decreases. When the clutch plate 52g and the clutch disk 52h are completely engaged (static friction state), both are in a constant speed rotation state.

  The piston 52b moves according to the hydraulic pressure of hydraulic oil (ATF: Automatic Transmission Fluid) supplied to the hydraulic chamber 52c through the hydraulic oil supply port 52f. When the hydraulic pressure increases, the clutch plate 52g and the clutch disk 52h are engaged. When the hydraulic pressure is lowered, the clutch plate 52g and the clutch disc 52h are moved in the direction of releasing the engagement (rightward in FIG. 2). .

  The return spring 52i is interposed between the piston 52b and the sub-piston 52k that forms the centrifugal hydraulic balance chamber 52j, and the hydraulic pressure in the centrifugal hydraulic balance chamber 52j generated according to the rotation speed A stable urging force is exerted on the piston 52b regardless of a change in the rotational speed while canceling out the centrifugal hydraulic pressure generated in the hydraulic chamber 52c in accordance with the rotational speed.

  Therefore, while the return spring 52i contracts according to the operating oil pressure P in the hydraulic chamber 52c, the piston 52b takes a position where the operating oil pressure P and the urging force of the return spring 52i are balanced, and according to the position of the piston 52b. The distance between the clutch plate 52g and the clutch disk 52h is determined. Further, the movement of the piston 52b requires the inflow of hydraulic oil into the hydraulic chamber 52c, and this hydraulic oil inflow amount corresponds to the position of the piston 52b.

  As shown in FIG. 1, the hydraulic oil is supplied from a hydraulic source (hydraulic pump or hydraulic oil tank that stores hydraulic oil from the hydraulic pump) 56, and this supply state is adjusted through the hydraulic oil supply valve 58. . An electrically controllable solenoid valve (electromagnetic valve) is applied to the hydraulic oil supply valve 58, and the supply state of the hydraulic oil can be electrically controlled.

<Turbine speed estimation device>
The present turbine speed estimation device is a device that estimates the rotational speed of the turbine 66 of the torque converter 6 when the forward / reverse switching mechanism 44 is switched between connection and disconnection in a vehicle (automobile) having the drive system as described above. It is configured as a functional element.

The ECU 10, a hydraulic estimation means 12 for estimating the oil pressure of the hydraulic fluid when disconnecting the ND clutch (hydraulic friction engagement element), the engaging torque T _CLT of ND clutch based on hydraulic pressure of estimated hydraulic oil Based on the estimated engagement torque _TCLT of the ND clutch, the engine rotation speed Ne detected by the engine rotation speed sensor 20a, the time of the turbine rotation speed at that time Turbine rotational speed calculation means 16 is provided for calculating the rate of change and integrating the time rate of change to calculate the turbine rotational speed of the torque converter 6 .

In particular, the hydraulic pressure estimation means 12 calculates the volume flow rate Q (n) of the hydraulic oil passing through the solenoid based on the drive current value I (n) of the solenoid of the solenoid valve 58, and calculates the calculated volume flow rate Q (n). Based on the calculated clearance h (n) between the elements of the ND clutch, the hydraulic pressure P (n) of the hydraulic oil is estimated from the calculated clearance h (n) between the elements.
Hereinafter, estimation of the turbine rotation speed by the hydraulic pressure estimation means 12, the engagement torque estimation means 14, and the turbine rotation speed calculation means 16 will be described.

<Estimation of turbine rotation speed>
The estimation of the rotational speed (turbine rotational speed) Nt of the turbine 66 of the torque converter 6 during the engagement of the ND clutch will be specifically described.

As shown in FIG. 1, when the equation of motion about the turbine shaft (input shaft 42 of the automatic transmission 4) is established, the transmission torque τCNe ² from the engine ² is input to the turbine shaft 42, and the ND clutch Since the engagement torque _TCLT is output, the equation of motion of the following equation (1) is established around the turbine shaft 42.
The turbine rotation speed Nt follows Formula (1) until the ND clutch is engaged, and can be regarded as Nt = Np regardless of Formula (1) after the ND clutch is engaged.

In addition, hereinafter, parameters relating to each formula are appended to the bottom of each formula.

The time change rate dωt / dt of the turbine angular velocity and the time change rate dNt / dt of the turbine rotational speed have a relationship of dωt / dt = (2π / 60) · (dNt / dt). When (1) is arranged with respect to dNt / dt, the following equation (1 ′) is obtained.

By integrating the equation (1 ′) for each calculation time, a turbine rotation speed estimated value Nt (n) of the current calculation cycle is obtained as in the following equation (2).

The torque ratio τ of the torque converter 6 and the capacity coefficient C of the torque converter 6 in the formula (1 ′) are design values related to the torque converter 6, and as shown in FIGS. 3 and 4, the turbine rotational speed Nt and the engine rotational speed Ne. The ratio Nt / Ne is given as a map. Here, the previous estimated value is used as the turbine rotation speed Nt.

Next, determine the ND clutch engagement torque _{T CLT} in the formula (1 '). The ND clutch engagement torque _{T CLT,} as shown in equation (3) below, guided by the sum of the viscous resistance torque (viscosity resistance _{section) T DRG} and contact friction torque (contact friction _{term) T FR C.} The viscous resistance term _TDRG is a viscous resistance torque due to ATF that acts in proportion to the relative speed between the friction plates of the ND clutch that is a wet multi-plate type, and the contact friction term T _{FR C} is the ND clutch. This is the contact friction torque generated by the friction plates when the friction plates are engaged.

The viscous resistance term _TDRG of the equation (3) is given by the following equation (4). The previous estimated value Nt (n−1) can be used as the turbine rotational speed Nt in the equation (4), and the difference between the turbine rotational speed Nt (n−1) and the detected transmission input rotational speed Np The viscous resistance term T _DRG can be easily calculated based on the number n of the engagement surfaces of the friction plates of the ND clutch, the effective radii r ₁ and r _2, and the viscous resistance coefficient μ _ATF of the hydraulic oil. .

The contact friction term T _{FR C} in the equation (3) is given by the following equation (5). That is, the contact friction torque T _{FR C} is determined by the hydraulic oil pressure P (n), the effective area nS and the effective radius [(r ₁ + r) of the engagement surface of each friction plate of the ND clutch on which the hydraulic pressure P (n) acts. ₂ ) / 2] and the friction coefficient μ _FRC . In order to calculate this equation (5), the hydraulic oil pressure (ND clutch piston operating oil pressure) P of the ND clutch is required.

Here, estimation of the ND clutch piston working oil pressure P will be described.
In this estimation, the ND clutch piston working oil pressure P is correlated with the ND clutch plate gap h (n), and the ND clutch plate gap h (n) is correlated with the volume Q (n) of the ATF in the cylinder. The volume Q (n) of the ATF in the cylinder can be obtained by integrating the amount of ATF entering / exiting the cylinder, and the amount of ATF entering / exiting the cylinder is the displacement x of the solenoid valve (hydraulic oil supply valve) 58. This is done by paying attention to the point that it can be obtained from (n).

Therefore, the hydraulic pressure estimation means 12 first calculates the volume flow rate Q _in (n) of the ATF that passes through the solenoid valve 58 every calculation cycle, and uses the calculated volume flow rate Q _in (n) for the hydraulic chamber 52c of the ND clutch. is corrected by outflow rate Q _out of, by integrating the corrected volumetric flow rate, it calculates the hydraulic oil volume Q in the hydraulic chamber 52c (n), then, in the hydraulic chamber 52c that are obtained in advance The gap value h (n) based on the hydraulic oil volume Q (n) using the correspondence relation between the hydraulic oil volume and the piston position and the correspondence relation between the previously obtained piston position and the ND clutch plate gap h. And the hydraulic pressure P (n) is calculated from the calculated gap value h (n) using the correspondence between the gap value and the ATF pressure obtained in advance. Details will be described below.

First, a simple ND clutch operation model is considered in order to simulate a piston action hydraulic pressure P that changes in response to a shift range switching command (for example, shift lever operation). As shown in FIG. 5, this operation model is installed in the clutch plate (clutch plate 52g and clutch disk 52h), cylinder 52a, piston 52b, ATF flow path 58a communicating with the cylinder 52a, and ATF flow path 58a. And an electromagnetic solenoid valve 58.

In this operation model, when there is a shift range switching command by a shift lever operation or the like, this becomes a clutch engagement release command depending on the shift range switching. When this clutch engagement release command is issued, the current (coil current) I (n) for driving the solenoid valve 58 flows in response to the primary delay of the command value O (n) with respect to the command value O (n). Thus, a magnetic field (coil magnetic field) H (n) is generated by the current I (n). The clutch engagement command value O (n) is 1 and the clutch release command value O (n) is 0. Next, the magnetic force due to the magnetic field H (n) and the restoring force of the return spring 52 i give a force f (n) to the solenoid valve 58 . The solenoid valve displacement speed v (n) can be obtained by integrating the force f (n) over time, and the solenoid valve displacement x (n) can be obtained by integrating the speed v (n) over time.

Therefore, the solenoid valve displacement x (n) can be derived from the following equations (6.1) to (6.5).
For example, the current I (n) for driving the solenoid valve has a clutch engagement command [O (n) = 1 in addition to the change according to the current response gain k ₁ as shown in the equation (6.1). ] If there is an increase in first-order lag in accordance with the primary filter gain k _2, it decreases clutch opening command [O (n) = 1] is the first order lag corresponding to the primary filter gain k _2. The current magnetic field by I (n) H (n), as shown in equation (6.2) is obtained by processing the current I (n) in the magnetic field response gain _{k 3} for relative current. Restoring force f of the return spring 52i applied to the solenoid valve 58 (n), as shown in equation (6.3), obtained by treating magnetic field H a (n) in response gain _{k 4} force with respect to the magnetic field force The spring force k ₅ x (n−1) obtained by multiplying k ₄ H (n) and the previous value x (n−1) of the displacement (= solenoid valve displacement) of the return spring 52i by the spring constant k ₅ It is obtained as a difference. The solenoid valve displacement speed v (n) is obtained by integrating the force f (n) over time as shown in equation (6.4), and the solenoid valve displacement x (n) is obtained by equation (6.5). As shown in FIG. 5, the speed v (n) can be obtained by integrating the time.

Here, since ATF flows in accordance with the solenoid valve displacement x (n), the solenoid valve passage volume flow rate Q _in (n) per unit time and the volume Q (n) existing in the cylinder (hydraulic chamber) 52c are The following equations (7.1) and (7.2) are obtained. Here, for simplification, it is assumed that a constant volume flow Q _out flows _out from the cylinder 52a of the ND clutch , and the ATF pressure difference before and after the solenoid valve 58 is constant at P0.

Further, as shown in FIG. 6, the solenoid passing ATF time flow rate is Qin, the solenoid flow path effective width is w, the solenoid inlet ATF hydraulic pressure is P0, the ATF density is ρ, the cylinder outflow ATF time flow rate is Qout, and the in-cylinder ATF volume is Q.

Here, in the operation model as shown in FIG. 7A, as shown in FIGS. 7B to 7D, the distance between the clutch plates is enlarged for the sake of convenience and changes in the distance between the clutch plates are shown. As shown in FIGS. 7B to 7D, when the state in which the piston engages the clutch plate is considered in a simplified manner, the piston becomes Q (n) / Q by the ATF volume Q (n) existing in the cylinder. Displace by Sf. Accordingly, the ND clutch plate gap h (n) is expressed by the following equation.

On the other hand, if the behavior of this hydraulic pressure is expressed by the clutch plate gap h (n) based on the general change characteristics of the actual ND clutch engagement hydraulic pressure, as shown in the following equation (9) and FIG. The simulated hydraulic pressure P (n) is defined to be “inversely proportional to h (n) and P (n) = 0 when h (n) = h0 and P (n) = P0 when h (n) = hm”. be able to.

Since the simulated hydraulic pressure P (n) can be obtained in this way, the contact friction term T _{FR C} can be obtained by applying the simulated hydraulic pressure P (n) to the ND clutch piston working hydraulic pressure of the above formula (5). Can be calculated. The viscous resistance term _TDRG in the equation (3) can be calculated by the equation (4). If the contact friction term T _{FR C} is obtained, the ND clutch engagement torque T _CLT is obtained by the equation (3). Can be requested. Further, by providing the ND clutch engagement torque T _CLT to equation (1 '), it is possible to determine the time rate of change dωt / dt of the turbine angular velocity by the equation (2), the turbine rotational speed of the current calculation cycle An estimated value Nt (n) is obtained.
<Operation (flow chart)>
Since the turbine speed estimation device for a torque converter according to an embodiment of the present invention is configured as described above, the turbine rotation speed Nt can be estimated, for example, as shown in the flowcharts of FIGS. . The flowcharts of FIGS. 9 and 10 are executed at a predetermined calculation cycle set in advance from when the engine is stopped, such as before the engine 2 is started. In addition, initial values of parameters for estimation calculation are set in advance.

That is, as shown in FIG. 9, first, it is determined whether the calculation cycle is the first time (step a10). If the calculation cycle is the first time, the initial values of the parameters related to the estimation calculation are read (step a20). The initial value of each parameter is a clutch engagement command value O (n), a solenoid valve drive current I (n), a magnetic field H (n) by this current I (n), a magnetic field H (n), and a return spring 52i are solenoids. force applied to the valve 58 f (n), the solenoid valve displacement speed v (n), the solenoid valve displacement x (n), ATF volume Q present in the cylinder (n), the simulated pressure P (n) are both 0 The clutch plate gap h (n) is a value h0 when the clutch is released.

  If the calculation cycle is the first time, the process goes through step a20. If the calculation cycle is not the first time, the process goes to the next step without going through step a20, and the engine rotation speed Ne is detected by the crank angle sensor 20a. The primary rotational speed Np is detected by 20b (step a30). Then, it is determined whether the engine is stopped (Ne = 0?) From the engine speed Ne (step a40). If the engine 2 is stopped, the estimated turbine speed Nt (n) for the current calculation cycle. Is set to 0 (step a100), and this calculation cycle ends.

If the engine 2 is not stopped, it is determined whether or not the engagement of the ND clutch is completed based on whether or not the simulated hydraulic pressure P (n−1) in the previous calculation cycle is equal to or greater than the determination value (step a50). Since ND clutch piston acting hydraulic when engagement of ND clutch is completed can grasped in advance, it is possible to set the determination value from the N D hydraulic pressure P at the completion engagement of the clutch. If the engagement of the ND clutch has been completed, the turbine rotation speed estimated value Nt (n) of the current calculation cycle is set to the primary rotation speed Np obtained by the primary rotation speed sensor 20b (step a90). End the cycle.

If not, the engagement of the ND clutch is complete, the viscous resistance term _{T DRG} of ND clutch engagement torque _{T CLT,} previously estimated value of the turbine rotation speed using Nt (n-1) of the by equation (4) Calculate (step a60). Moreover, to calculate the contact friction term _{T FR C} of the ND clutch engagement torque _{T CLT} (step a70).
The calculation of the contact friction term T _{FR C} (step a70) is performed as shown in the sub-flowchart of FIG. That is, first, the shift lever position is read (step b10). The clutch engagement release command [command value O (n)] can be grasped based on the shift lever position, that is, in which shift range the shift lever position is. The clutch engagement release command value O (n) obtained from the shift lever position is determined (step b20), and the clutch engagement release command value O (n) is a command to open the travel range, that is, the ND clutch. For example, the clutch engagement release command value O (n) is set to 0 (step b30), and if it is a command to engage the ND clutch, the clutch engagement release command value O (n) is set to 1 (step b40). .

Next, a current (coil current) I (n) for driving the solenoid valve 58 is calculated by the above equation (6.1) (step b50), and a magnetic field (coil magnetic field) H (n) by the current I (n) is calculated. Is calculated by the above equation (6.2) (step b60). The force f (n) acting on the solenoid valve 58 is calculated by the above equation (6.3) (step b70). Further, as shown in the equation (6.4), the solenoid valve displacement speed v (n) is calculated by integrating the force f (n) over time (step b80), and the equation (6.5) is calculated. ), The solenoid valve displacement x (n) is calculated by integrating the speed v (n) over time (step b90).

Since ATF flows in accordance with the solenoid valve displacement x (n), as shown in the above equation (7.1), the solenoid valve passage volume flow rate Qin (n) per unit time is calculated (step b100), As shown in the equation (7.2), the volume Q (n) existing in the cylinder (hydraulic chamber) 52c is calculated (step b110).
Since the piston is displaced by Q (n) / Sf due to the ATF volume Q (n) existing in the cylinder, the ND clutch plate gap h (n) is calculated as shown in the equation (8) (step). b120). Then, the simulated hydraulic pressure P (n) is calculated from the relationship between the clutch plate gap h (n) and the behavior of the hydraulic pressure as shown in the equation (9) (step b130).

By applying this simulated oil pressure P (n) to the ND clutch piston working oil pressure of the above equation (5), the contact friction term T _{FR C} is calculated (step b140).
Referring again to FIG. 9, by such operation, contact friction term of ND clutch engagement torque _{T CLT} viscous drag term _{T DRG} (step a60) and ND clutch engagement torque _{T CLT T FR C} (step a70) When is obtained, it calculates the ND clutch engagement torque _{T CLT} adds the contact friction term _{T FR C} and viscous resistance term _{T DRG} (step a80).

Further, the value Nt / Ne of the ratio between the turbine rotational speed Nt and the engine rotational speed Ne is calculated using the previous estimated value Nt (n-1) as the turbine rotational speed Nt (step a110). Further, the torque ratio τ of the torque converter 6 and the capacity coefficient C of the torque converter 6 are retrieved from the maps (FIGS. 3 and 4) using the ratio value Nt / Ne as arguments (step a120).
The obtained ND clutch engagement torque T _CLT , torque converter torque ratio τ, torque converter capacity coefficient C, and engine rotational speed Ne detected at step a30 are substituted into the above equation (1 ′) to obtain the turbine rotational speed. Is calculated (step a130), and as shown in the above equation (2), the time rate of change dNt / dt of the turbine rotation speed is integrated for each calculation period, thereby obtaining the turbine rotation speed. An estimated value Nt (n) is calculated (step a140).

<Effect>
Thus even when the change in the turbine rotational speed Nt to move between the state ND clutch release and engagement opening occurs with without resorting to a turbine rotational speed sensor turbine rotation speed Nt (n) Can be estimated.

Therefore, for example, when the vehicle is idling, the torque converter load can be accurately calculated using the turbine rotational speed Nt even when switching between the traveling range and the non-traveling range, and the behavior of the engine rotational speed can be stabilized. Become.
Further, the input command is not limited to the driver's shift lever operation. In other words, the ND clutch operates in the same manner as described above even when performing control (so-called idle neutral control) that automatically shifts to a state equivalent to the neutral state regardless of the driver's shift lever operation when stopping. Also in this case, it is possible to estimate the turbine rotational speed Nt (n) and stabilize the behavior of the engine rotational speed using the turbine rotational speed Nt.

Furthermore, the application of the turbine rotational speed Nt obtained by the present invention is not limited to the case where the control for stabilizing the behavior of the engine rotational speed is performed by using the calculation of the torque converter load described above. Is also applicable.
Although the embodiments of the present invention have been described above, the present invention is not limited to such embodiments, and some of the above embodiments may be used or modified as appropriate without departing from the spirit of the present invention. Or can be implemented.

2 Engine (Internal combustion engine)
4 Automatic transmission 6 Torque converter (torque converter)
10 Electronic control unit (ECU)
DESCRIPTION OF SYMBOLS 12 Oil pressure estimation means 14 Engagement torque estimation means 16 Turbine rotational speed calculation means 20a Crank angle sensor (engine rotational speed detection means)
20b Primary rotational speed sensor (primary rotational speed detection means)
DESCRIPTION OF SYMBOLS 40 Casing of automatic transmission 4 42 Input shaft of automatic transmission 4 44 Forward / reverse switching mechanism 44a, 44b Rotation element of forward / reverse switching mechanism 44 46A Primary shaft 46 Primary pulley 48A Secondary shaft 48 Secondary pulley 50 Belt 52 Wet multi-plate clutch Device (ND clutch, hydraulic friction engagement element)
52a Cylinder 52b Piston 52c Hydraulic chamber 52d Clutch plate holding part 52e Clutch disk holding member 52g Clutch plate 52h Clutch disk 52i Return spring 54 Output shaft 54a Gear pair 56 Hydraulic source (hydraulic pump or hydraulic oil tank)
58 Hydraulic oil supply valve (solenoid valve, solenoid valve)
62 Torcon casing 64 Impeller 66 Turbine liner (turbine)

Claims (7)

An engine mounted on a vehicle; an automatic transmission having a forward / reverse switching mechanism for switching between forward / reverse travel and transmission / non-transmission of power by connecting / disconnecting a hydraulic friction engagement element; and the engine and the automatic A torque converter interposed between the forward / reverse switching mechanism of a transmission, and a device for estimating a turbine speed of the torque converter in a drive system comprising:
Engine rotational speed detecting means for detecting the rotational speed Ne of the engine;
A hydraulic pressure estimation means for estimating a hydraulic pressure P of the hydraulic oil when the hydraulic friction engagement element is connected / disconnected;
_Engagement torque estimation means for estimating an engagement torque _TCLT of the hydraulic friction engagement element based on the hydraulic pressure P of the hydraulic oil estimated by the hydraulic pressure estimation means;
Based on the engagement torque _{TCLT of the} hydraulic friction engagement element estimated by the engagement torque estimation means and the engine rotation speed Ne detected by the engine rotation speed detection means, the hydraulic friction is determined. Turbine rotational speed calculation means for calculating a turbine rotational speed Nt of the torque converter when the engaging element is connected / disconnected is provided. The turbine rotation speed calculation means includes
Based on the rotational speed Ne of the engine, a transmission torque τCNe ² (where τ is a torque ratio of the torque converter and C is a capacity coefficient of the torque converter) input from the engine to the automatic transmission is calculated. on the basis of the difference between the transmission torque TauCNe ² and the engaging torque T _CLT, and calculates the turbine rotation speed Nt, the turbine speed estimation device of a torque converter according to claim 1, wherein. The hydraulic friction engagement element is connected and disconnected by controlling the hydraulic pressure by a solenoid valve,
The hydraulic estimating means, on the basis of the driving current value I of the solenoid of the solenoid valve to calculate the volume flow Q _in the hydraulic oil passing through the solenoid valve, the hydraulic based on the calculated said volume flow rate Q _in The turbine speed of the torque converter according to claim 1, wherein a gap h between the elements of the friction engagement element is calculated, and the hydraulic pressure P of the hydraulic oil is estimated from the calculated gap h between the elements. Estimating device. The hydraulic pressure estimation means includes
The volume flow rate Q _in (n) of the hydraulic oil passing through the solenoid valve is calculated at predetermined intervals, and the calculated volume flow rate Q _in (n) is corrected by the outflow flow rate Q _out in the hydraulic friction engagement element. Based on the corrected volume flow rate, the hydraulic fluid volume Q (n) in the hydraulic chamber of the hydraulic friction engagement element is calculated, and the gap h is calculated based on the calculated hydraulic fluid volume Q (n) in the hydraulic chamber. The turbine speed estimation device for a torque converter according to claim 3, wherein (n) is calculated. The hydraulic friction engagement element is connected / disconnected by a piston that moves by controlling the hydraulic pressure by the electromagnetic valve,
The hydraulic pressure estimation means uses the correspondence relationship between the hydraulic fluid volume in the hydraulic chamber obtained in advance and the piston position, and the correspondence relationship between the piston position and the gap between the elements obtained in advance. The turbine speed estimation device for a torque converter according to claim 4, wherein the gap h (n) is calculated based on a hydraulic oil volume Q (n) in the hydraulic chamber. The engagement torque estimating means includes
The engagement torque T _CLT is calculated as a sum of a viscous resistance torque _TDRG corresponding to the viscous resistance of the hydraulic oil and a contact friction torque T _FRC corresponding to the contact friction of the hydraulic friction engagement element. The torque converter turbine speed estimation device according to any one of claims 1 to 5. A transmission input rotation speed detecting means for detecting a transmission input rotation speed Np input from the forward / reverse switching mechanism to the transmission mechanism of the automatic transmission;
The viscous resistance torque _TDRG is a difference between the previously obtained turbine rotational speed Nt (n-1) and the transmission input rotational speed Np detected by the transmission input rotational speed detecting means, and the hydraulic frictional coefficient. Calculated based on the number of engagement surfaces and effective radii (r ₁ , r ₂ ) of the joint element and the viscous resistance coefficient μ _ATF of the hydraulic oil,
The contact friction torque T _FRC includes the hydraulic pressure P (n) of the hydraulic oil, the effective area nS of the engagement surface of the hydraulic friction engagement element on which the hydraulic pressure P (n) acts, and the effective radius [(r ₁ 7. The turbine speed estimation device for a torque converter according to claim 6, wherein calculation is performed based on + r ₂ ) / 2] and a friction coefficient μ _FRC .

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