METHOD FOR CONTROLLING A SHIFTING PROCESS IN A POWERTRAIN OF A VEHICLE

Abstract

A method is provided for controlling a shifting process in a powertrain of a vehicle, the powertrain having a first and a second drive machine, a transmission connecting the drive machines to a transmission output and at least one coupling which can be shifted, wherein during a shifting process an offgoing coupling is disengaged and/or an oncoming coupling is engaged.

Claims

1. A method for controlling a shifting process in a powertrain of a vehicle having a first drive machine and a second drive machine, a transmission connecting the drive machines to a transmission output and at least one shiftable clutch, the method including the steps of: disengaging, during a shifting process, an outgoing clutch and/or engaging an oncoming clutch, operating the transmission in a fixed transmission ratio in a first transmission mode and in a variable transmission ratio in a second transmission mode, during a relief phase, the outgoing clutch is relieved before disengagement, and/or in a synchronization phase a differential speed between a drive side and a driven side of the oncoming clutch is adjusted to zero before engagement, and simultaneously with the relief phase and/or the synchronization phase, a wheel drive torque continues to be adjusted to a target drive torque.

2. The method according to claim 1, characterized in that the relief of the outgoing clutch is carried out by adjusting a mating torque of the outgoing clutch to zero.

3. The method according to claim 1, characterized in that the relief of the outgoing clutch is carried out by dividing the wheel drive torque between the first and second drive machine in such a way that a mating torque of the outgoing clutch is zero.

4. The method according to claim 3, characterized in that a torque distribution between the first and second drive machine is carried out by means of model-based pilot control.

5. The method according to claim 1, characterized in that the adjustment of the differential speed between the drive side and the driven side of the oncoming clutch is carried out before engagement to zero by means of model-based pilot control.

6. The method according to claim 1, characterized in that the regulation of a mating torque of the outgoing clutch and the regulation of the differential speed between the drive side and the driven side of the oncoming clutch during the relief phase or the synchronization phase of the shifting process are carried out immediately one after the other.

7. The method according to claim 1, characterized in that during the relief phase of the outgoing clutch the transmission is operated with a fixed transmission ratio.

8. The method according to claim 1, characterized in that during the synchronization phase of the oncoming clutch the transmission is operated with variable transmission ratio.

9. The method according to claim 1, characterized in that before the relief phase of the outgoing clutch and/or after the synchronization phase of the oncoming clutch the transmission is operated with a fixed transmission ratio.

10. The method of claim 2, characterized in that the relief of the outgoing clutch is carried out by dividing the wheel drive torque between the first and second drive machine in such a way that the mating torque of the outgoing clutch is zero.

11. The method of claim 4, characterized in that the adjustment of the differential speed between the drive side and the driven side of the oncoming clutch is carried out before engagement to zero by means of model-based pilot control.

12. The method of claim 5, characterized in that the regulation of a mating torque of the outgoing clutch and the regulation of the differential speed between the drive side and the driven side of the oncoming clutch during the relief phase or the synchronization phase of the shifting process are carried out immediately one after the other.

13. The method of claim 6, characterized in that during the relief phase of the outgoing clutch the transmission is operated with a fixed transmission ratio.

14. The method of claim 7, characterized in that during the synchronization phase of the oncoming clutch the transmission is operated with variable transmission ratio.

15. The method of claim 8, characterized in that before the relief phase of the outgoing clutch, and/or after the synchronization phase of the oncoming clutch, the transmission is operated with a fixed transmission ratio.

16. The method of claim 4, wherein a torque distribution between the first and second drive machine is carried out by means of at least one trajectory.

17. The method of claim 5, wherein the adjustment of the differential speed between the drive side and the driven side of the oncoming clutch is carried out before engagement to zero by means of at least one trajectory.

18. The method of claim 1, wherein during a relief phase the outgoing clutch is completely relieved before disengagement.

19. The method of claim 1, wherein the at least one shiftable clutch is two shiftable clutches.

Description

[0030] The invention is explained in more detail below by reference to the non-restrictive figures, wherein:

[0031] FIG. 1 schematically shows a hybrid powertrain for carrying out the method according to the invention;

[0032] FIG. 2 shows a curve of the vehicle speed and the vehicle acceleration during a shifting process;

[0033] FIG. 3 shows a torque curve during a shifting process;

[0034] FIG. 4 shows a clutch torque curve during a shifting process; and

[0035] FIG. 5 shows a speed curve of the drive machines during a shifting process.

[0036] FIG. 1 shows by way of example a simplified mechanical schematic of a topology of a powertrain 1 having a first drive machine E and a second drive machine M of a vehicle, wherein in the embodiment example the first drive machine E is formed by an internal combustion engine and the second drive machine M is formed by an electric machine. However, the first drive machine E can also be an electric machine. The powertrain 1 has a transmission 2, which connects the drive machines E, M to a transmission output 5 and thus to the drive wheels of a motor vehicle which are not shown further. Transmission 2 in the embodiment example has an extended Ravigneaux planetary gear set 3 and a simple planetary gear set 4. The extended Ravigneaux planetary gear set 3 has a first sun gear S.sub.1, a second sun gear S.sub.2, a common planet carrier PT.sub.12 for a set of first planet gears P.sub.1 and a set of second planet gears P.sub.2, a first ring gear R.sub.1 and a second ring gear R.sub.2, wherein the first planet gears P.sub.1 mesh with the second planet gears P.sub.2. The simple planetary gear set 4 has a third sun gear S.sub.3, which engages a third planet gear P.sub.3 of a planet carrier PT.sub.3, and a third ring gear R.sub.3. Furthermore, the transmission 2 has a shiftable first clutch C.sub.0, a shiftable second clutch C.sub.1, a shiftable third clutch C.sub.2 and a shiftable fourth clutch C.sub.3, wherein the shiftable fourth clutch C.sub.3 is designed as a brake. When engaged, the first clutch C.sub.0 establishes a drive connection between the second drive machine E and the third ring gear R.sub.1 When engaged, the second clutch C.sub.1 establishes a drive connection between the third ring gear R.sub.3 and the common planet carrier PT.sub.12. When engaged, the third clutch C.sub.2 connects the first sun gear S.sub.1 to the planet carrier PT.sub.3 of the first planetary gear set 4. The fourth clutch C.sub.3 fixes the second sun gear S.sub.2 in engaged condition.

[0037] In FIG. 2 to FIG. 5, for example, a shifting process from a first gear G1 with a fixed transmission ratio FGR to a second gear G2 with a fixed transmission ratio FGR and simultaneous vehicle acceleration a (train upshift) using the method according to the invention is shown, in which, for example, the third clutch C.sub.3 is disengaged and the first clutch C.sub.1 is engaged. Power distribution occurs between the first drive machine E and the second drive machine M, wherein the first drive machine E is operated stationary. The second drive machine M supports the shifting process, wherein the values for the power distribution are provided stationary in an FGR transmission mode by the electronic hybrid control unit HCU.

[0038] FIG. 2 shows the speed v and the acceleration a of the vehicle over the time t before, during and after a shifting process. As can be seen from FIG. 2, the acceleration a of the vehicle takes place over 10 seconds s, the shifting time t.sub.s of the shifting process is about 0.7 seconds (from 3.8 to 4.5 s). This shifting time t.sub.s can only be seen as an example. In the case of faster actuators, the shifting time t.sub.s can be reduced.

[0039] FIG. 3 shows the course of the clutch torque .sub.C3 of the outgoing clutch C.sub.3, the clutch torque .sub.C1 of the oncoming clutch C.sub.1, the drive torque .sub.M of the first drive machine ICE and the drive torque .sub.E of the second drive machine EM over time t for a shifting process. Clutch torque is understood here as the maximum torque to be transmitted via the clutch plates of the corresponding clutch. It can clearly be seen that the outgoing clutch C.sub.3 is completely disengaged at time t=4 s and the oncoming clutch C.sub.1 is completely engaged at time t=4.3 s.

[0040] FIG. 4 shows the course of the mating torques .sub.SC1/.sub.SC3 of the outgoing clutch C.sub.3 or the oncoming clutch C.sub.1, as well as the speed difference .sub.C1 of the clutch plates of the clutch C.sub.1 to be engaged. The term mating torque here refers to the torque actually applied to the respective clutch and transmitted via the clutch plates.

[0041] FIG. 5 shows the speed curve .sub.E and .sub.M of the first drive machine E or the second drive machine M during a shifting process.

[0042] The shifting process can be divided into three time ranges t.sub.1, t.sub.2, t.sub.3, which can be modified separately:

[0043] Relief phase t.sub.1: Relieving the disengaging (outgoing) clutch C.sub.3 (3.8-4.0 s: duration of t.sub.1; 0.2 seconds)

[0044] In the first time range t.sub.1, the aim is to relieve the load on the clutch C.sub.3 to be disengaged in order to avoid grinding immediately after breakaway. For this purpose, either the mating torque .sub.SC3 applied to the outgoing clutch C.sub.3 can be controlled directly to zero, or the power distribution between the first and second drive machine M, E can be modified by the electronic hybrid control unit HCU to implicitly achieve the identical goal. The course of the relevant torque .sub.SC3 is shown in FIG. 4. As soon as the clutch C.sub.3 is completely relieved (time 4.0 s) the clutch C.sub.3 can be disengaged without loss (see .sub.C3 in FIG. 3) and the synchronization phase t.sub.2 can begin.

[0045] Synchronization phase t.sub.2: Synchronization of the clutch plates of the engaging (oncoming) clutch C.sub.1 (4.0-4.3 s: Duration of t.sub.2: 0.3 seconds) The aim is to synchronize the clutch plate speeds or eliminate the speed difference .sub.C1 of the clutch plates of the clutch C.sub.1 to be engaged (see FIG. 4). As soon as there is no speed difference .sub.C1 between the clutch plates C.sub.1 to be engaged (time 4.3) the respective clutch C.sub.1 can be engaged without loss (see torque .sub.SC1 of the oncoming clutch C.sub.1 in FIG. 4). Once synchronization is complete, the restoration of the power distribution required by the electronic hybrid control unit HCU can begin.

Recovery Phase t.SUB.3.: Restoration of Power Distribution (by HCU)

[0046] The aim of the third phase t.sub.3 is to restore the power distribution required by the HCU between the first drive machine E and the second drive machine M. This transition is also carried out gently in order to avoid unrealistic loads on the actuators (e.g. torque jumps). This process corresponds to a gentle loading of the newly engaged clutch C.sub.1 (see torque .sub.SC1 of the oncoming clutch C.sub.1 in FIG. 4).

[0047] During the illustrated shifting process from the first gear G1 driven in FGR mode, a change into CVT mode takes place in the first time range t.sub.1 and second time range t.sub.2 and then again into FGR mode of the second gear G2.

[0048] Compared to the prior art, the method according to the invention has the following advantages: [0049] The shifting strategy is not dependent on the shifting process (e.g. pull-up shifting). [0050] The shifting process can be carried out without loss, as there are no sliding clutches. For real clutch actuation (finite rise times) only the differential speed of the engaging clutch must be kept at zero for the duration of the clutch actuation or an additional phase after phase 1 must be inserted in which the differential speed of the disengaged clutch is kept at zero for the duration of the clutch actuation. The strategy is therefore also robust against inaccuracies in the clutch actuation. This makes it possible to replace friction clutches in a powertrain topology with much cheaper and more robust claw clutches. [0051] The shifting process can be carried out completely without jerking: There is no effect of the shifting process on the desired vehicle acceleration, not even if the clutches are activated stepwise. [0052] It is not necessary to adapt the shifting process for changes in the nominal drive torque (driver) during the shifting process. [0053] The time sequence of the shifting process does not have to be subdivided into torque transfer and speed phase. [0054] The method can be implemented with a simple adjustment effort: For example, a simple pilot control principle can be used, which is based on fulfilling two wishes simultaneously using two actuators.

[0055] The basic idea of the pilot control is that actuating signals for the shifting process are calculated in such a way that a certain behavior is achieved for degrees of freedom; trajectories (e.g. the vehicle speed) are specified for this purpose. To solve this problem, the system must be inverted. This generally leads to causality problems. This restriction can be removed by assuming that the derivations of the trajectories are known. The actuating signals are then calculated from a filtered linear combination of these derivations, wherein the filter and the linear combination contain the inverse model behavior. The linear combination results from the counter polynomials of the inverse transmission matrix of the system. The transmission matrix defines the model input/output behavior in the frequency domain, in this case the behavior of the wheel drive torque and slip speed of the oncoming clutch (in the case of CVT transmission mode) or the wheel drive torque and mating torque of the outgoing clutch (in the case of FGR transmission mode). The filters result from the denominator polynomials of this inverse transfer matrix. The number of necessary derivations corresponds to the relative degrees of the individual transmission paths. The relative degrees are given by the order differences (degrees of difference) in the transfer matrix.

[0056] The method described is suitable for all powertrain topologies with two drive machines E, M and a transmission 2, which has at least one transmission mode with fixed transmission ratio FGR and at least one transmission mode with variable transmission ratio CVT, with at least two shiftable clutches C.sub.1, C.sub.3. It is not limited to a certain number of gears or transmission type.