Method for operating a dual-clutch transmission for resolving respective tooth-on-tooth positions

Abstract

A method for operating a dual-clutch transmission includes, to resolve a respective tooth-on-tooth position between a toothing of a sliding sleeve and a toothing of a gear wheel during an engaging operation for the respective gear, a relative rotation between the toothing of the sliding sleeve and the toothing of the gear wheel is effected by a twisting torque which acts on the respective gear where the twisting torque is a same size for all of the plurality of gears.

Claims

1. A method for operating a dual-clutch transmission (10) which includes: two sub-transmissions (12, 14) each with an input shaft (16, 18) which is connectable via a friction clutch (24, 26) to an output shaft (22) of a drive engine (20); and a plurality of gears (1, 2, 3, 4, 5, 6, 7, R) with respective gear transmission ratios, wherein each of the plurality of gears are shiftable using a respective friction-synchronized gear clutch, wherein the gear clutches have a respective toothing of a sliding sleeve (42, 44, 46, 48) and a respective toothing of a respective gear wheel (Z1, Z3, Z5, Z7, Z9, Z11, Z12, Z13), wherein respective contact surfaces of the respective toothing of the sliding sleeve (42, 44, 46, 48) and the respective toothing of the respective gear wheel are flat or slightly rounded; and comprising the steps of: wherein, to resolve a respective tooth-on-tooth position between a toothing of a sliding sleeve (42, 44, 46, 48) and a toothing of a gear wheel (Z1, Z3, Z5, Z7, Z9, Z11, Z12, Z13)) during an engaging operation for the respective gear, a relative rotation between the toothing of the sliding sleeve (42, 44, 46, 48) and the toothing of the gear wheel (Z1, Z3, Z5, Z7, Z9, Z11, Z12, Z13) is effected by a twisting torque which acts on the respective gear (1, 2, 3, 4, 5, 6, 7, R), wherein the twisting torque is a same size for all of the plurality of gears (1, 2, 3, 4, 5, 6, 7, R), and wherein a respective clutch torque for the respective friction clutch (24, 26) required to effect the twisting torque is calculated based on a respective gear transmission ratio, which clutch torque is set at the respective friction clutch (24, 26).

2. The method according to claim 1, wherein in the engaging operation, the sliding sleeve (42, 44, 46, 48) is displaced by a gear selector.

3. The method according to claim 2, wherein the twisting torque is solicited by an electronic control unit of the dual-clutch transmission (10) when a position of the gear selector exceeds a predefined threshold value by a predefined amount.

4. The method according to claim 1, wherein the twisting torque is solicited by the electronic control unit when a rotational speed difference between the sliding sleeve (42, 44, 46, 48) and the gear wheel (Z1, Z3, Z5, Z7, Z9, Z11, Z12, Z13) has reduced to a value which is in a range from 10 revolutions per minute inclusive to 40 revolutions per minute inclusive.

5. The method according to claim 2, wherein the twisting torque is solicited by an electronic control unit of the dual-clutch transmission (10) when it is determined that a speed of the gear selector is greater than 20 to 60 millimeters per second.

6. The method according to claim 2, wherein in the engaging operation, the gear selector has a constant hydraulic pressure applied to it from a start of a soliciting of the twisting torque to an end of the engaging operation.

7. The method according to claim 1, wherein the twisting torque is in a range from five newton-meters inclusive to 15 newton-meters inclusive.

8. A method for operating a dual-clutch transmission, comprising: resolving a tooth-on-tooth position between a toothing of a sliding sleeve and a toothing of a gear wheel of a gear during an engaging operation of the gear by effecting a relative rotation between the toothing of the sliding sleeve and the toothing of the gear wheel by a twisting torque which acts on the gear.

Description

BRIEF DESCRIPTION OF THE DRAWING

(1) FIG. 1 shows a schematic depiction of a dual-clutch transmission which is operated or can be operated by means of a method according to the invention.

DETAILED DESCRIPTION OF THE DRAWING

(2) FIG. 1 shows a schematic depiction of a dual-clutch transmission 10 for a motor vehicle, in particular for a car, for instance a passenger car. The dual-clutch transmission 10 has a first sub-transmission 12 and a second sub-transmission 14. The sub-transmission 12 has a first input shaft 16 and the sub-transmission 14 has a second input shaft 18. In the exemplary embodiment shown in FIG. 1, the input shaft 18 is a hollow shaft, which is completely penetrated along the axial direction thereof by the input shaft 16.

(3) The dual-clutch transmission 10 is a component of a drive train for driving the motor vehicle. The drive train can have a drive engine 20, also simply referred to as engine. The drive engine 20 is for example an internal combustion engine. Alternatively or additionally, the drive engine 20 may be a reciprocating engine. The drive engine 20 has a drive shaft 22, which is for example designed as a crankshaft. The drive engine 20 can provide torques for driving the motor vehicle via the drive shaft 22, with the torques for driving the motor vehicle also being referred to as driving moments or drive torques.

(4) For producing hybrid vehicles, the dual clutch transmission 10 may additionally have not only an electric machine incorporated into the dual-clutch transmission 10 as shown in FIG. 1 but an electric machine may also be provided incorporated into the drive train on the input side of the first and second friction clutches 24, 26. The electric machine can also be integrated into the dual-clutch transmission at other points or be connected to the crankshaft of the drive engine.

(5) The sub-transmission 12 has a first friction clutch 24 which is also simply referred to as first clutch. The sub-transmission 14 has a second friction clutch 26 which is also simply referred to as second clutch. The clutches are also referred to as starting clutches and are for example designed as multi-disc clutches. It can be seen from FIG. 1 that the input shaft 16 can be connected to the drive shaft 22 via the friction clutch 24. The input shaft 18 can be connected to the drive shaft 22 via the second friction clutch 26. As a result, the driving moments can be transmitted from the drive shaft 22 to the respective input shafts 16 or 18 via the respective clutch.

(6) The dual-clutch transmission 10 has forward travel gears 1, 2, 3, 4, 5, 6 and 7, by means of which a respective forward travel of the motor vehicle can be effected. The dual-clutch transmission 10 also has a reverse gear R, by means of which a reverse travel of the motor vehicle can be effected. Where reference is made hereinafter to a gear, the gear, gears or the gears of the dual-clutch transmission 10, this can be understood to mean the respective forward travel gear 1, 2, 3, 4, 5, 6 and 7 and/or the respective forward gears 1, 2, 3, 4, 5, 6, and 7 and/or the reverse gear R. It can be seen from FIG. 1 that the first sub-transmission 12 comprises the forward travel gears 1, 3, 5 and 7, with the sub-transmission 14 comprising the forward travel gears 2, 4 and 6 and the reverse gear R. The forward travel gears 1, 2, 3, 4, 5, 6 and 7 are also referred to as forward gears. Accordingly, the reverse gear R is also referred to as reverse travel gear.

(7) Moreover, the dual-clutch transmission 10 can comprise an electric machine 28, which for example can be operated in motor mode and thus as an electric motor. In motor mode, the electric machine 28 can provide torques, which can be introduced into the sub-transmission 12. For example, the motor vehicle can be driven by torques provided by the electric machine 28, in particular via the sub-transmission 12. The dual-clutch transmission 10 also has a parking brake 30.

(8) The dual-clutch transmission 10 has a first output shaft 32 and a second output shaft 32. A toothed wheel 36 or 38 is connected non-rotatably to the respective output shaft 32 or 34, with the toothed wheels 36 and 38 simultaneously meshing with a further toothed wheel 40. The toothed wheels 36 and 40 or 38 and 40 form for example a final drive ratio, also referred to as final drive, other than 1. It can furthermore be seen from FIG. 1 that the gears comprise toothed wheels Z1-13. Here, for example, the forward travel gear 1 comprises the toothed wheels Z1 and Z2, the forward travel gear 2 comprises the toothed wheels Z3 and Z4, the forward travel gear 3 comprises the toothed wheels Z5 and Z6, the forward travel gear 4 comprises the toothed wheels Z7 and Z8, the forward travel gear 5 comprises the toothed wheels Z9 and Z10, the forward travel gear 6 comprises the toothed wheels Z11 and Z8, the forward travel gear 7 comprises the toothed wheels Z12 and Z10, and the reverse gear R comprises for example the toothed wheel Z13 and for example the toothed wheel Z3 and for example the toothed wheel Z4. The toothed wheels Z1, Z3, Z7 and Z9 are gear wheels in the form of idler wheels which are rotatably mounted on the output shaft 34. The toothed wheels Z5, Z11, Z12 and Z13 are gear wheels in the form of idler wheels which are rotatably mounted on the output shaft 32.

(9) In contrast, the toothed wheels Z2, Z6 and Z10 are fixed wheels, which are permanently non-rotatably connected to the input shaft 16. The toothed wheels Z4 and Z8 are fixed wheels, which are permanently non-rotatably connected to the input shaft 18. In order, for example, to be able to selectively connect the toothed wheel Z1 and the toothed wheel Z9 to the output shaft 34, in particular in a positive-locking and non-rotatable manner, a sliding sleeve 42 is provided. The sliding sleeve 42 is therefore assigned to the forward travel gears 1 and 5, or the sliding sleeve 42 can be a component of the forward travel gears 1 and 5. In order to be able to selectively connect the toothed wheels Z3 and Z7 to the output shaft 34, in particular in a non-rotatable and positive-locking manner, a sliding sleeve 44 is provided. The sliding sleeve 44 is therefore assigned to the forward travel gears 2 and 4, or can be a component of the forward travel gears 2 and 4. In order to be able to selectively connect the toothed wheels Z5 and Z12 to the output shaft 32, in particular in a positive-locking and non-rotatable manner, a sliding sleeve 46 is provided. The sliding sleeve 46 is therefore assigned to the forward travel gears 3 and 7, or is a component of the forward travel gears 3 and 7. In order finally to be able to connect the toothed wheel Z11 to the output shaft 32, in particular in a positive-locking and non-rotatable manner, a sliding sleeve 48 is provided. The sliding sleeve 48 is therefore assigned to the forward travel gear 6, or is a component of the forward travel gear 6.

(10) The respective sliding sleeve 42, 44, 46 and 48 can be moved translationally relative to the respective output shaft 32 or 34 in the axial direction of the respective output shaft 32 or 34 and thus be axially displaced. The respective gear can therefore be shifted. In order to engage the respective gear, the respective idler wheel is connected non-rotatably to the respective output shaft 32 or 34 by means of the respective associated sliding sleeve 42, 44, 46 or 48. In order to at least reduce, in particular eliminate, a potential rotational speed difference between the respective output shaft 32 or 34 and the respective idler wheel, and therefore be able to synchronize the respective idler wheel with the respective output shaft 32 or 34, the respective gear has a friction-synchronized gear clutch, by means of which the respective gear can be shifted, in particular engaged.

(11) In the case of a single synchronization, the respective friction-synchronized gear clutch has a respective synchronizer ring, which for example has at least one first friction surface. The method according to the invention is independent of the number of friction surface pairings and can be applied to single or multiple synchronizations. In the case of a single synchronization, the first friction surface is for example conical. Furthermore, the respective gear clutch comprises for example a clutch body of the respective idler wheel, also referred to as gear wheel. The clutch body forms for example a second friction surface, which corresponds in particular to the first friction surface and therefore can also be conical in design. In a sufficiently well-known manner, the respective idler wheel and the respective output shaft 32 or 34 can be synchronized by means of the friction surfaces and therefore by means of the gear clutch which forms a synchronization, in particular a friction synchronization. In this case, the respective sliding sleeves 42, 44, 46 and 48, the respective synchronizer ring and the clutch bodies or the respective idler wheel, also referred to as gear wheel, each have a toothing. The toothing of the respective sliding sleeve 42, 44, 46 and 48 is also referred to as first toothing, and the respective shifting toothing of the idler wheel or gear wheel is also referred to as second toothing. The toothing of the synchronizer ring is also referred to as third toothing.

(12) The respective sliding sleeve 42, 44, 46, or 48 is also referred to as clutch sleeve. In order to connect the respective gear wheel (idler wheel) to the respective output shaft 32 or 34 in a positive-locking and non-rotatable manner by means of the respective clutch sleeve, the respective clutch sleeve is moved translationally relative to the gear wheel in the axial direction of the respective gear wheel and is therefore axially displaced, wherein the respective clutch sleeve is axially displaced on the respective gear wheel, i.e., moved in the direction of the respective gear wheel, in particular such that the first toothing of the respective clutch sleeve engages with the respective second toothing of the respective gear wheel. Moreover, the respective gear has a gear transmission ratio, also referred to as transmission ratio, which is formed by the intermeshing toothed wheels of the respective gear. It can be provided here that the gears differ from one another in terms of their respective gear transmission ratios.

(13) It is preferably provided here that teeth of the first toothing are flat or planar or else slightly rounded on the end face thereof which faces the second toothing in the axial direction, while for example teeth of the second toothing are flat or planar or else slightly rounded on the end face thereof which faces the first toothing in the axial direction. As a result, what is referred to as a tooth-on-tooth position can arise between the first toothing and the second toothing. Such a tooth-on-tooth position can prevent the teeth of the first toothing from being able to be inserted in the tooth gaps of the second toothing, and so such a tooth-on-tooth position can prevent the toothings from being brought into engagement with one another.

(14) Therefore, a method for operating the dual-clutch transmission 10 is provided, wherein, in the context of the method, tooth-on-tooth positions between the respective first toothing and the respective second toothing are resolved, i.e., eliminated. To this end, between the first toothing of the respective clutch sleeve and the second toothing of the respective gear wheel, during an engaging operation for engaging the respective gear, a relative rotation between the first toothing and the second toothing, i.e., a relative rotation between the respective clutch sleeve and the respective gear wheel, is effected by means of a twisting torque which acts on the respective gear, in particular on the respective clutch sleeve. The twisting torque is a torque which acts on the respective gear, in particular on the respective clutch sleeve, in order to effect such a relative rotation between the respective clutch sleeve and the respective gear wheel that the respective tooth-on-tooth position is eliminated, i.e., resolved.

(15) In order here to be able to achieve a particularly advantageous and in particular quiet, low-wear and smooth operation of the dual-clutch transmission 10, the twisting torque is the same size for all gears, and a respective clutch torque for the respective friction clutch 24 or 26 required to effect the twisting torque is calculated based on the respective gear transmission ratio. This means that the twisting torque is effected by means of the respective friction clutch 24 or 26 which belongs to the sub-transmission 12 or 14 whose gear is engaged or is to be engaged during the engaging operation, in particular such that the respective friction clutch 24 or 26 is at least partially closed. As a result, the clutch torque set is transmitted from the drive shaft 22 to the respective input shaft 16 or 18 via the respective friction clutch 24 or 26. The twisting torque which acts on the clutch sleeve and rotates same relative to the respective gear wheel results from the clutch torque depending on the gear transmission ratio.

(16) It is conceivable for the method according to the invention to be applied to dual-clutch transmissions with for example electromechanically actuated gear selectors.

(17) In the respective engaging operation, the respective clutch sleeve is displaced by means of a gear selector. The gear selector is for example a gear selector piston which is moved translationally in particular in the axial direction of the respective output shaft 32 or 34 and is therefore axially displaced, in order for example to axially displace a respective selector fork, and via the respective selector fork, the respective clutch sleeve. To this end, a fluid, in particular a liquid, is applied to the gear selector piston, with the fluid having a pressure. This means that for example the pressure is applied to the gear selector. Furthermore, it can be provided that the fluid is applied to the gear selector by means of an electrically actuatable valve, such that the fluid, and therefore the pressure, is applied electrohydraulically to the gear selector and therefore to the respective clutch sleeve. This therefore provides electrohydraulic actuation of the gear selector and therefore the respective clutch sleeve.

(18) The dual-clutch transmission 10 is for example operated by means of an electronic computing unit, which is also referred to as electronic control unit or transmission control unit. The transmission control unit can for example control the respective friction clutch 24 or 26 in order thereby to set or effect the clutch torque and consequently the twisting torque and/or to solicit the twisting torque. The soliciting of the twisting torque is therefore to be understood for example to mean that the transmission control unit, in particular a module of the transmission control unit, solicits effecting or exerting the twisting torque, for example by the respective friction clutch 24 or 26 or by another module of the transmission control unit.

(19) In order to engage the respective gear, the respective clutch sleeve is for example displaced into an end position. It was found to be advantageous to generate an at least substantially constant, gear-dependent twisting torque at an exact point in time of the engaging operation, in order to be able to reach the end position of the respective clutch sleeve or the respective gear. This twisting torque makes it possible to eliminate the tooth-on-tooth position which can arise because of the flat tooth geometry, and so the respective clutch sleeve can then be guided into its end position. The tooth-on-tooth position can occur, for example, between what is referred to as a synchronization position and an engaging position of the respective clutch sleeve. At the point in time of the engaging operation, the exact synchronization position and what is referred to as the “kiss point” of the respective clutch are preferably known. “Kiss point” is to be understood to mean an actuation state in which clutch discs of the respective clutch are just touching. The twisting torque is preferably solicited when a position of the gear selector, also referred to as gear selector position, exceeds a predefined threshold value, in particular by a predefined amount. In particular, the twisting torque is solicited when the gear selector position is greater than the previously learnt synchronization position plus 0.2 millimeter (mm), minus what is referred to as a fork deflection. The fork deflection is to be understood to mean a preferably elastic deformation, in particular deflection, of the selector fork which occurs during the engaging operation and which is subtracted from the synchronization position. In other words, a setpoint position is calculated by 0.2 mm being added to the previously learnt synchronization position, and the fork deflection then being subtracted. If the determined gear selector position exceeds this setpoint position, the twisting torque is solicited.

(20) It is preferably further provided that the twisting torque is solicited, in particular by the transmission control unit, when a rotational speed difference between the respective clutch sleeve and the respective gear wheel has reduced to a value which is in a range from 10 revolutions per minute inclusive to 40 revolutions per minute inclusive, in particular which is 20 revolutions per minute. The twisting torque is preferably solicited, in particular by the transmission control unit, when it is determined that a speed of the gear selector is greater than 20 to 60 millimeters per second (mm/s), in particular greater than 40 mm/s. The speed is for example the speed at which the gear selector is axially displaced. In other words, the twisting torque is preferably solicited when the gear selector speed is once again greater than 20 to 60 mm, in particular once again greater than 40 mm/s.

(21) In order to exclude distortion of the gear selector position by the fork deflection, dependent for example on the above-mentioned pressure, the fork deflection is subtracted out from the determined gear selector position in a position query during which the gear selector position is determined. To this end, the fork deflection is calculated out of the determined, in particular detected, gear selector position, or is subtracted therefrom. The gear selector position is detected for example by means of a sensor, in particular a Hall sensor.

(22) The twisting torque is preferably 10 newton-meters (Nm). It has proven particularly advantageous if the twisting torque is in the range from 5 Nm inclusive to 15 Nm inclusive. It has been found that the fork deflection is for example approximately 0.3 mm at a pressure of 5 bar. In order for example to generate the same twisting torque of for example 10 Nm for each gear at the respective gear, in particular at a respective synchronization point, in particular at the respective clutch sleeve, of the respective gear, the clutch torque is back-calculated starting from the twisting torque and using the gear transmission ratio. The twisting torque acting on the respective gear is for example the mathematical product of the respective gear transmission ratio and the clutch torque, and so for example in order to determine the clutch torque, the twisting torque is divided by the gear transmission ratio. The result of the mathematical division is the clutch torque, which is for example set at the respective clutch.

(23) For example, in engaging operations as a part of high-load shifting operations, in particular as soon as the twisting torque is solicited, the pressure, also referred to as gear selector pressure, is simultaneously set to a constant level, which is, for example, in the range from 3 bar inclusive to 5 bar inclusive, in particular is 4 bar. The ratio of pressure to twisting torque is for example selected such that the clutch torque can overcome the axial force at the gear selector. If for example the tooth-on-tooth position cannot be resolved, i.e., not eliminated, the clutch torque is increased, with the gear selector pressure then for example remaining unchanged.

(24) In engaging operations in the course of low-load and/or smooth shifting, a constant twisting torque is for example likewise solicited, which is for example the same for each gear, however no constant gear selector pressure is then set, but rather the gear selector pressure is for example lowered to a target value starting from a starting value. The starting value is for example 3.5 bar or is in a range from 3 bar inclusive to 5 bar inclusive, with the target value possibly being for example 2 bar or being in a range from 1.5 bar inclusive to 2.5 bar exclusive. The emphasis here is not on shifting speed but rather on smooth operation, in particular in terms of noise and pressure.

LIST OF REFERENCE CHARACTERS

(25) 1 Forward travel gear 2 Forward travel gear 3 Forward travel gear 4 Forward travel gear 5 Forward travel gear 6 Forward travel gear 7 Forward travel gear 10 Dual-clutch transmission 12 First sub-transmission 14 Second sub-transmission 16 Input shaft 18 Input shaft 20 Drive engine 22 Drive shaft 24 Friction clutch 26 Friction clutch 28 Electric machine 30 Parking brake 32 Output shaft 34 Output shaft 36 Toothed wheel 38 Toothed wheel 40 Toothed wheel 42 Sliding sleeve 44 Sliding sleeve 46 Sliding sleeve 48 Sliding sleeve R Reverse gear Z1-13 Toothed wheel