Patent Application
20100228456
The present invention relates to a control method for closing a clutch in an automatic manual transmission.
Automatic manual transmissions (commonly named “AMT”) are increasingly widespread, which transmissions are structurally similar to manual gearbox of the traditional type except for the clutch pedal and the gear selection lever operated by the driver being replaced by corresponding electric or hydraulic servo controls.
An automatic manual transmission is provided with a transmission control unit, which while shifting gears, drives the servo controls associated with clutch and gearbox to disengage the current gear and engage the successive gear; furthermore, while shifting gear, the transmission control unit interfaces with the engine control unit to conveniently adjust the torque delivered by the engine so as to adapt the engine operation to the transmission ratio change and to avoid an undesired temporary increase/decrease of engine rpm when the clutch is open or in all cases either while opening or closing. Such a control mode is applied both to traditional single clutch gearboxes and to double clutch gearboxes; the only difference between the two types of gearboxes resides in that in a single clutch gearbox there is a central interval of time in which the clutch is completely open (i.e. the engine idly turns without load), while in a double clutch gearbox the two clutches mutually cross over and therefore a clutch opening/closing situation exists during the whole shifting operation.
A clutch has an engine side angularly integral with the drive shaft of the engine (possibly by interposing a transmission shaft) and a gearbox side, integral with a primary shaft of the gearbox, which is in turn angularly integral with the driving wheels through the currently engaged gear; the rotation speed of the two sides of the clutch is adapted when closing the clutch, because when the clutch is completely closed, the two sides of the clutch are angularly integral with each other, and thus rotate at the same rotation speed. The rotation speed of the driving wheels is imposed by the forward traveling speed of the vehicle, and varies with relatively long times because of the high total inertia of the vehicle itself; instead, the rotation speed of the drive shaft of the engine may be quickly varied in a controlled manner when the drive shaft is idling (i.e. when it is separated from the driving wheels) due to the low total inertia of the engine. Therefore, in the known, currently marketed automatic manual transmissions, when closing a clutch, the rpm of the drive shaft (i.e. of the engine side of the clutch) is controlled according to the rpm of the gearbox side of the clutch (thus according to the rotation speed of the driving wheels, multiplied by the transmission ratio of the currently engaged gear), so as to make a connection between the rotation speed of the engine side of the clutch and the rotation speed of the gearbox side of the clutch as “smooth” as possible. Indeed, if the connection between the rotation speed of the engine side of the clutch and the rotation speed of the gearbox side of the clutch is “violent”, oscillations are triggered in the transmission, which cause longitudinal vibrations on the vehicle, and are thus perceived by the driver and by possible passengers.
However, in the known, currently marketed automatic manual transmissions, the clutch is closed very slowly in order to obtain a “smooth” connection between the rotation speed of the engine side of the clutch and the rotation speed of the gearbox side of the clutch; such a mode is effective in avoiding the onset of oscillations in the transmission, but on the other hand increases the time needed to shift gears, and thus penalizes performance. The problem is particularly evident in double clutch gearboxes, in which the clutches are in an oil bath because, while in a dry clutch used in a conventional single clutch gearbox the transmitted torque remains virtually constant despite slip variations (i.e. differences of rpm between the two sides of the clutch), in an oil bath clutch used in a double clutch gearbox, the torque transmitted by the clutch clearly drops to small slip values, and such a drop of the transmitted torque has a high dispersion from clutch to clutch. Therefore, in an oil bath clutch, controlling the torque transmitted by the clutch at small slip values (i.e. when connecting the rotation speed of the engine side of the clutch to the rotation speed of the gearbox side of the clutch is required) is very complicated.
US2008207393A1 describes a clutch control method which determines a torque target which is to be transmitted by the clutch during a gear shift to reduce the total time required for the gear shift itself. A clutch controller controls an opening degree of the clutch by driving an clutch actuator according to the difference between the torque actually transmitted by the clutch and a target of the torque transmitted by the clutch; moreover, the clutch controller determines whether a speed difference between the two sides of the clutch is reduced with an adequate acceleration and, according to such a determination, the clutch controller corrects the target of the torque transmitted by the clutch.
It is the object of the present invention to provide a control method for closing a clutch in an automatic manual transmission, which method is free from the above-described drawbacks while being easy and cost-effective to be implemented.
According to the present invention, a control method for closing a clutch in an automatic manual transmission is provided as claimed in the attached claims.
The present invention will now be described with reference to the accompanying drawings, which illustrate a non-limitative embodiment thereof, in which:
In
Vehicle 1 comprises a control unit 11 of engine 4, which governs the engine 4, a control unit 12 of transmission 6, which governs the transmission 6, and a bus line 13, which is made according to CAN (Car Area Network) protocol, extends throughout the vehicle 1 and allows the control units 11 and 12 to interface with each other. In other words, the control unit 11 of engine 4 and the control unit 12 of transmission 6 are connected to the bus line 13, and may thus communicate with each other by means of messages sent on the bus line 13 itself. Furthermore, the control unit 11 of engine 4 and the control unit 12 of transmission 6 are directly connected to each other by means of a dedicated synchronization wire 14, which is capable of directly transmitting a signal without the delays introduced by the BUS line 13 from the control unit 13 of transmission 6 to the control unit 12 of engine 4.
As shown in
The double clutch gearbox 7 has seven forward speeds indicated by Roman numerals (first gear I, second gear II, third gear III, fourth gear IV, fifth gear V, sixth gear VI and seventh gear VII) and one reverse (indicated by the letter R). The primary shaft 15 and the secondary shaft 17 are mechanically coupled to each other by means of a plurality of gear pairs, each of which defines a respective speed and comprises a primary gear 18 mounted to the primary shaft 15 and a secondary gear 19 mounted to the secondary shaft 17. In order to allow the correct operation of the double clutch gearbox 4, all the odd speeds (first gear I, third gear III, fifth gear V, seventh gear VII) are coupled to a same primary shaft 5a, while all the even speed (second gear II, fourth gear IV, and sixth gear VI) are coupled to the other primary shaft 15.
Each primary gear 18 is keyed onto a respective primary shaft 15 to again integrally rotate with the primary shaft 15 itself, and permanently meshes with the respective secondary gear 19; instead, each secondary gear 19 is idly mounted to the secondary shaft 17. Moreover, the double clutch gearbox 7 comprises four double synchronizers 20, each of which is mounted so as to be coaxial to the secondary shaft 17, is arranged between two secondary gears 19, and is adapted to be actuated to alternatively engage the two respective secondary gears 19 onto the secondary shaft 17 (i.e. to alternatively make the two respective secondary gears 19 angularly integral with the secondary shaft 17). In other words, each synchronizer 20 may be displaced either in one direction to engage a secondary gear 19 onto the secondary shaft 17, or in the other direction to engage the other secondary gear 19 onto the secondary shaft 17.
Each clutch 16 has an engine side 21 which is angularly integral with the drive shaft 5 of engine 4 by interposing a transmission shaft 8, and a gearbox side 22 which is integral with a primary shaft 15 of the gearbox 7, which in turn may be made angularly integral with the secondary shaft 17, and thus with the rear driving wheels 3 by engaging a gear. When each clutch 16 is closed, the two sides 21 and 22 are angularly integral with each other and thus have to rotate at the same rotation speed; instead, when each clutch 16 is at least partially open, the two sides 21 and 22 are free to rotate at two different rotation speeds.
The modes of shifting from a current gear A to a successive gear B are described below. A shift up during the gear sequence will be described for simplicity, and therefore the current gear A has a higher transmission ratio than the successive gear B.
In an initial situation (i.e. before shifting the gear), a clutch 16A is closed to transmit motion to a primary shaft 15A, which in turn transmits motion to the secondary shaft 17 by means of the currently engaged gear A; a clutch 16B is, instead, open and thus isolates a primary shaft 15B from the transmission shaft 8. Before starting to shift up, the successive gear B is engaged to connect the primary shaft 15B to the secondary shaft 17 by means of the gear B itself; once the gear B has been engaged, the shifting starts by opening the clutch 16A to disconnect the primary shaft 15A (thus gear A) from the transmission shaft 8 (i.e. from the drive shaft 5 of engine 4), while closing the clutch 16B to connect the primary shaft 15B (thus gear B) to the transmission shaft 8 (i.e. to the drive shaft 5 of engine 4).
The above-described shifting modes are diagrammatically shown in
As previously mentioned, the engine sides 21 of both clutches 16 are angularly integral with the drive shaft 5 of engine 4 by interposing the transmission shaft 8 and thus they always rotate at the same rotation speed ωE as the drive shaft 5 of engine 4. The gearbox side 22 of clutch 16A is integral with the primary shaft 15A and until the current gear A is engaged, it rotates at a rotation speed ωA which is equal to the rotation speed of the rear driving wheels 3 multiplied by the transmission ratio of the current gear A; similarly, the gearbox side 22 of the clutch 16B is integral with the primary shaft 15B and until the successive gear B is engaged, it rotates at a rotation speed ωB which is equal to the rotation speed of the rear wheels 3 multiplied by the transmission ratio of the current speed B. By first approximation, when shifting gear, the forward traveling speed of vehicle 1 (thus the rotation speed of the rear driving wheels 3 and the rotation speed of the gearbox sides 22 of both clutches 16) is assumed to remain substantially constant when shifting gear; such an assumption is quite realistic because the duration of the gear shifting is short (indicatively a few tenth of a second) and, under normal conditions, the high total mechanical inertia of vehicle 1 does not allow large speed variations in such a short interval of time.
A model of engine 4 and transmission 6 is determined during a step of designing the vehicle 1, which allows the behavior of engine 4 and transmission 6 to be simulated; in particular, such a model uses the transmissibility functions of the clutches 16, each of which provides the torque which is transmitted by the corresponding clutch 16 according to the opening degree (i.e. to the position when controlling the position or to the hydraulic pressure when controlling the pressure) of the clutch 16 itself.
During the whole gear shift, i.e. from the moment T1 of starting the opening of clutch 16A corresponding to the current gear A to the moment T3 of completely closing the clutch 16B corresponding to the successive gear B, the control unit 12 determines a torque target TC-TAR which is to be transmitted through the clutches 16 according to the required evolution of the gear shift; in other words, the desired evolution of the gear shift is generally described by a speed target during the gear shift and by a longitudinal acceleration target during the gear shift, and according to these speed and acceleration targets, a torque target TC-TAR may be determined which is to be transmitted through the clutches 16. It is worth noting that the torque target TC-TAR which is to be transmitted by the clutches 16 is equal to the sum of a torque target TCA-TAR which is to be transmitted through the clutch 16A and a torque target TCB-TAR which is to be transmitted through the clutch 16B. According to the torque target TC-TAR which is to be transmitted through each clutch 16, the opening degree (i.e. the position when controlling the position or the hydraulic pressure when controlling the pressure) of the clutch 16 itself is determined by using the corresponding transmissibility function; in other words, once the required evolution of the torque target TC-TAR which is to be transmitted through each clutch 16 has been determined, the evolution of the opening degree (position or hydraulic pressure) is imposed by the physical features of the clutch 16 itself and is supplied by the corresponding transmissibility function.
During the whole gear shifting, i.e. from the moment T1 of starting the opening of the clutch 16A corresponding to current gear A to the moment T3 of completely closing the clutch 16B corresponding to the successive gear B, the control unit 12 of transmission 6 determines a target torque TE-TAR of the engine 4 according to the torque target TC-TAR which is to be transmitted through the clutches 16; moreover, the control unit 12 of transmission 6 communicates the target torque TE-TAR of the engine 4 to the control unit 11 of engine 4 by means of the bus line 13 and/or by means of the dedicated synchronization wire 14, so that the control unit 11 of engine 4 drives the engine 4 to pursue the torque target TE-TAR of the engine 4.
As shown in
After moment T2, the rotation speed ωE of the drive shaft 5 (i.e. the rotation speed ωE of the engine side 21 of clutch 16B) is gradually decreased to reach the rotation speed ωB of the primary shaft 15B (i.e. the rotation speed ωB of the gearbox side 22 of clutch 16B) as quickly as possible; when the rotation speed ωE of the engine side 21 of clutch 16B is close to the rotation speed ωB of the gearbox side 22 of clutch 16B, i.e. at an impact interval I in which the rotation speed ωE of the engine side 21 of clutch 16B approaches the rotation speed ωB of the gearbox side 22 of clutch 16B (e.g. when the speed difference Δω is lower than a threshold which may take the value of 250 rpm), a particular control mode is actuated to ensure that the connection between the rotation speed of the engine side 21 of clutch 16B and the rotation speed of the gearbox side 22 of cutch 16B is “smooth” and quick at the same time.
As shown in
In particular, a measuring sensor 23 (or, alternatively, an estimating algorithm) supplies the rotation speed ωB of the gearbox side 22 of clutch 16B (i.e. the rotation speed ωB of the primary shaft 15B) to the control unit 12 of transmission 6, and the control unit 11 of engine 4 supplies the rotation speed ωE of the engine side 21 of clutch 16B (i.e. the rotation speed ωE of the drive shaft 5) to the control unit 12 of transmission 6. The control unit 12 of transmission 6 determines the speed difference Δω between the rotation speed ωE of the engine side 21 of clutch 16B and the rotation speed ωB of the gearbox side 22 of clutch 16B, and thus provides the speed difference Δω both to a control block 24, which determines the acceleration target αTAR according to the speed difference Δω, and to a derivation block 25 which determines the acceleration α deriving the speed difference Δω over time. The control unit 2 of transmission 6 determines an acceleration error εα by comparing the acceleration α with the acceleration target αTAR, and thus determines a first contribution TCB-TAR1 of the torque target TCB-TAR, which is to be transmitted through the clutch 16B according to the acceleration target αTAR by means of a PID regulator 26 PID which attempts to cancel the acceleration error εα.
A control block 27 of the control unit 12 of transmission 6 determines a second contribution TCB-TAR2 of the torque target TCB-TAR which is to be transmitted through the clutch 16B according to the desired evolution of the gear shift. The torque target TCB-TAR which is to be transmitted through the clutch 16B is determined by algebraically adding (i.e. taking the sign into account) the first contribution TCB-TAR1 of the target torque TCB-TAR which is to be transmitted through the clutch 16B to the second contribution TCB-TAR2 of the target torque TCB-TAR which is to be transmitted through the clutch 16B. From the above description, it is apparent that the first contribution TCB-TAR1 of the torque target TCB-TAR which is to be transmitted through the clutch 16B is determined by means of a closed loop control (i.e. feedback control) and serves the function of conferring stability and robustness to the overall control, while the second contribution TCB-TAR2 of the torque target TCB-TAR which is to be transmitted through the clutch 16B is determined by means of an open loop control and serves the function of conferring performance (i.e. rapidity) to the overall control. By varying the weight of the two contributions TCB-TAR1 and TCB-TAR2 of the torque objective TCB-TAR which is to be transmitted through the clutch 16B, robustness (and thus comfort) or performance (thus rapidity) may be favoured.
According to an alternative embodiment (not shown) gearbox 7 is of the single-clutch type and thus comprises only one clutch 16 and only one primary shaft 15; also in this case, the above-described control method is practically applied according to the same modes.
To summarize the above, the control unit 12 of transmission 6 normally determines the torque target TCB-TAR which is to be transmitted through the clutch 16B only according to the desired evolution of the position of clutch 16B; instead, when the rotation speed ωE of the engine side 21 of clutch 16B is close to the rotation speed ωB of the gearbox side 22 of cutch 16B, the torque target TCB-TAR which is to be transmitted through the clutch 16B is also determined according to an acceleration target αTAR which depends on the speed difference Δω between the rotation speed ωE of the engine side 21 of clutch 16B and the rotation speed ωB of the gearbox side 22 of clutch 16B. Thereby, a connection between the rotation speed ωE of the engine side 21 of clutch 16B and the rotation speed ωB of the gearbox side 22 of clutch 16B may be obtained, which is very “smooth” while being also particularly quick. Furthermore, it is very simple to calibrate such a connection to favour either “smoothness” or rapidity according to the driving style chosen by the driver (e.g. sporty or comfortable style) simply by adjusting the parameters of the calculation block 24 which determines the acceleration target αTAR according to the speed difference Δω.
In brief, the above-described control method has many advantages. Firstly, a connection between the rotation speed ωE of the engine side 21 of clutch 16B and the rotation speed ωB of the gearbox side 22 of clutch 16B may be obtained, which is very “smooth” while being also particularly quick. Secondly, the above-described control method is simple and cost-effective to be implemented, because it does not require the installation of additional physical components, and does not results in increasing the power of the control unit 12 of transmission 6 as no high additional calculation power is needed.
| Number | Date | Country | Kind |
|---|---|---|---|
| BO2009A 000127 | Mar 2009 | IT | national |
