DE10308748A1 - Process and device for driving a motor vehicle controls a parallel switching gear with a double coupling in the drive train - Google Patents

Abstract

A gear for a motor vehicle (1) has an automatically operated friction coupling with a gear shaft coupled to a combustion engine (2) and a drive wheel (6) connected to the gear output. A conversion step between the gear (4) input and output shafts is switchable by an actuator (13b) controlled by a control unit (13). An Independent claim is also included for a process to operate a gear as above.

Description

The invention relates to a method, a device and the use thereof for Operation of a motor vehicle, in particular for controlling a Parallel gearbox (PSG), with a drive motor, a clutch and / or a Gearboxes are provided in the drive train.

Referring to FIG. 1, a vehicle 1, a drive unit 2 such as a motor or an internal combustion engine, on. Furthermore, a torque transmission system 3 and a transmission 4 are arranged in the drive train of the vehicle 1 . In this embodiment, the torque transmission system 3 is arranged in the power flow between the engine and the transmission, with a drive torque of the engine via the torque transmission system 3 to the transmission 4 and from the transmission 4 on the output side to an output shaft 5 and to a downstream axis 6 and to the wheels 6 a is transmitted.

The torque transmission system 3 is a clutch, such as. B. designed as a friction clutch, multi-plate clutch, magnetic powder clutch or torque converter clutch, the clutch can be a self-adjusting or a wear-compensating clutch. The transmission 4 is an uninterruptible manual transmission (USG). According to the idea of the invention, the transmission can also be an automated manual transmission (ASG), which can be shifted automatically by means of at least one actuator. In the following, an automated manual transmission is to be understood as an automated transmission which is shifted with an interruption in tractive force and in which the shifting operation of the transmission ratio is carried out in a controlled manner by means of at least one actuator.

Furthermore, an automatic transmission can also be used as the USG, with a Automatic transmission is a transmission with essentially no interruption in tractive power Switching operations is and usually built up by planetary gear stages is.

Furthermore, a continuously variable transmission, such as a conical pulley belt transmission, can be used. The automatic transmission can also be designed with a torque transmission system 3 arranged on the output side, such as a clutch or a friction clutch. The torque transmission system 3 can furthermore be designed as a starting clutch and / or a reversing set clutch for reversing the direction of rotation and / or a safety clutch with a selectively controllable, transferable torque. The torque transmission system 3 can be a dry friction clutch or a wet friction clutch that runs, for example, in a fluid. It can also be a torque converter.

The torque transmission system 3 has a drive side 7 and an output side 8 , wherein a torque is transmitted from the drive side 7 to the output side 8 by z. B. the clutch disc 3 a by means of the pressure plate 3 b, the plate spring 3 c and the release bearing 3 e and the flywheel 3 d is subjected to force. For this application, the release lever 20 is actuated by means of an actuator, for. B. an actuator operated.

The torque transmission system 3 is controlled by means of a control unit 13 , such as, for. B. a control unit, which can include the control electronics 13 a and the actuator 13 b. In another advantageous embodiment, the actuator 13 b and the control electronics 13 a can also be in two different units, such as. B. housings.

The control unit 13 can contain the control and power electronics for controlling the drive motor 12 of the actuator 13 b. This can for example advantageously achieved that the system as a single space b the installation space for the actuator 13 required with electronics. The actuator 13 b consists of the drive motor 12 , such as. B. an electric motor, the electric motor 12 via a gear such. B. a worm gear, a spur gear, a crank gear or a threaded spindle gear acts on a master cylinder 11 . This effect on the master cylinder 11 can take place directly or via a linkage.

The movement of the output part of the actuator 13 b, such as. B. the master cylinder piston 11 a is detected with a clutch travel sensor 14 which detects the position or position or the speed or the acceleration of a variable which is proportional to the position or engagement position or the speed or acceleration of the clutch. The master cylinder 11 is via a pressure medium line 9 , such as. B. a hydraulic line, connected to the slave cylinder 10 . The output element 10 a of the slave cylinder is with the disengaging means 20 , for. B. a release lever, operatively connected so that a movement of the output part 10 a of the slave cylinder 10 causes the disengaging means 20 is also moved or tilted to control the torque transmitted by the clutch 3.

The actuator 13 b for controlling the transmissible torque of the torque transmission system 3 can be actuatable by pressure medium, ie it can have a pressure medium transmitter and slave cylinder. The pressure medium can be, for example, a hydraulic fluid or a pneumatic medium. The actuation of the pressure medium transmitter cylinder can take place by means of an electric motor, wherein the electric motor provided as the drive element 12 can be controlled electronically. The drive element 12 of the actuator 13 b, besides an electromotive drive element, another, for example, be pressure-actuated drive member. Magnetic actuators can also be used to adjust a position of an element.

In the case of a friction clutch, the transferable torque is controlled in that the friction linings of the clutch disc are pressed in a targeted manner between the flywheel 3 d and the pressure plate 3 b. About the position of the disengagement means 20 , such as. B. a release fork or a central release, the application of force to the pressure plate 3 b or the friction linings can be controlled in a targeted manner, the pressure plate 3 b being moved between two end positions and being able to be set and fixed as desired. One end position corresponds to a fully engaged clutch position and the other end position corresponds to a fully disengaged clutch position. To control a transmissible torque, which is, for example, less than the engine torque currently present, a position of the pressure plate 3 b can be controlled, for example, which lies in an intermediate region between the two end positions. The clutch can be fixed in this position by means of the targeted actuation of the disengaging means 20 . However, it is also possible to control transmissible clutch torques that are defined above the engine torques currently pending. In such a case, the currently occurring engine torques can be transmitted, the torque irregularities in the drive train being damped and / or isolated in the form of, for example, torque peaks.

To control the torque transmission system 3 , sensors are also used which at least temporarily monitor the relevant variables of the entire system and deliver the state variables, signals and measured values necessary for control, which are processed by the control unit, with a signal connection to other electronic units, such as, for example Engine electronics or an electronics of an anti-lock braking system (ABS) or an anti-slip control (ASR) can be provided and can exist. For example, the sensors detect speeds such as wheel speeds, engine speeds, the position of the load lever, the throttle valve position, the gear position of the transmission, an intention to shift and other vehicle-specific parameters.

Fig. 1 shows that a throttle sensor 15, an engine speed sensor 16 and a speedometer sensor may find use 17 and transmit measured values or information to the controller 13. The electronics unit, such as. B. a computer unit, the control electronics 13 a processes the system input variables and transmits control signals to the actuator 13 b.

The gear is as z. B. stage change gear designed, the gear ratios are changed by means of a shift lever 18 or the transmission is operated or operated by means of this shift lever 18 . Furthermore, at least one sensor 19 b is arranged on the shift lever 18 of the manual transmission, which detects the intention to shift and / or the gear position and forwards it to the control unit 13 . The sensor 19 a is articulated on the transmission and detects the current gear position and / or an intention to shift. The intention to shift is detected using at least one of the two sensors 19 a, 19 b in that the sensor is a force sensor which detects the force acting on the shift lever 18 . Furthermore, the sensor can also be designed as a displacement or position sensor, the control unit recognizing an intention to switch from the change in the position signal over time.

The control unit 13 is at least temporarily in signal connection with all sensors and evaluates the sensor signals and system input variables in such a way that, depending on the current operating point, the control unit issues control or regulation commands to the at least one actuator 13 b. The drive motor 12 of the actuator 13 b, z. B. an electric motor receives from the control unit, which controls the clutch actuation, a manipulated variable as a function of measured values and / or system input variables and / or signals of the connected sensors. For this purpose, a control program is implemented in the control unit 13 as hardware and / or as software, which evaluates the incoming signals and calculates or determines the output variables on the basis of comparisons and / or functions and / or characteristic maps.

The control unit 13 has advantageously implemented a torque determination unit, a gear position determination unit, a slip determination unit and / or an operating state determination unit or is in signal connection with at least one of these units. These units can be implemented by control programs as hardware and / or as software, so that by means of the incoming sensor signals, the torque of the drive unit 2 of the vehicle 1 , the gear position of the transmission 4 and the slip that prevails in the area of the torque transmission system 3 and the current operating state of the vehicle 1 can be determined. The gear position determination unit uses the signals from the sensors 19 a and 19 b to determine the currently engaged gear. The sensors 19 a, 19 b are articulated on the shift lever and / or on gear-internal actuating means, such as a central shift shaft or shift rod, and detect them, for example the position and / or the speed of these components. Furthermore, a load lever sensor 31 on the load lever 30 , such as. B. on an accelerator pedal, which detects the load lever position. Another sensor 32 can act as an idle switch, ie when the load lever 30 or accelerator pedal is actuated, this idle switch 32 is switched on and when the load lever 30 is not actuated, it is switched off, so that digital information can be used to identify whether the load lever 30 is actuated. The load lever sensor 31 detects the degree of actuation of the load lever 30 .

Fig. 1 shows in addition to the load lever 30 and the sensors associated therewith a brake actuating element 40 for actuating the service brake or the parking brake, such as. B. a brake pedal, a hand brake lever or a hand or foot operated actuator of the parking brake. At least one sensor 41 is arranged on the actuating element 40 and monitors its actuation. The sensor 41 is, for example, a digital sensor, such as. B. designed as a switch, which detects that the brake actuator 40 is operated or not. With the sensor 41 , a signal device such. B. a brake light, are in signal connection, which signals that the brake is applied. This can be done for both the service brake and the parking brake. However, the sensor 41 can also be designed as an analog sensor, such a sensor, such as a potentiometer, determining the degree of actuation of the brake actuating element 41 . This sensor can also be in signal connection with a signal device.

The following is a possible embodiment of the invention presented here described in which a general strategy for controlling the clutches in a parallel transmission, in particular a double clutch transmission, is specified.

It has been shown that when the torque is transferred, in particular during upshifts, the transmitted torque T _{clB of} the clutch of the higher gear increases from 0 to the expected slip limit. In the meantime, the transmitted torque T _{clA of} the still adhering clutch of the lower gear is maintained until this clutch is torque-free. At this moment the lower gear can be removed.

A mathematical analysis of this strategy shows that the time t ₁ , at which the transmitted torque of clutch A is just 0, lies before the time t ₂ , at which clutch B has reached its expected slip limit. The acceleration a of the vehicle at time t _{1 is given} by the following formula:

In contrast, the acceleration a of the vehicle is given just after the time t ₂ by the following formula:

Where r is the radius of the wheels, T _vehicle running resistance, T _closely the engine torque, J _vehicle inertia moment of the vehicle J _closely the moment of inertia of the engine, i _A, the total ratio of the lower gear, and eat the total ratio of the higher gear is. So i _A > i _B. Assuming that the numerator of the two formulas given above is positive, it can be concluded that a (t ₁ ) <a (t ₂ ), that is to say that the vehicle acceleration during the cross-fading of the clutches falls below the value which it shows after the Crossfade. This is schematically sketched as an example in FIG. 2.

Such a decrease in acceleration can result in a loss of comfort. It would therefore be desirable for a linear transition of the acceleration to be provided without a minimum, as is also indicated in FIG. 2 by the dashed line.

In FIG. 2, the acceleration curve during and after a high-tension circuit is shown for a vehicle having a dual clutch transmission. It is clear from FIG. 2 that when the clutch A is stuck, the acceleration drops below the level of the acceleration in the new gear.

Another possible strategy can provide that T _{clA is reduced} linearly to the value 0, while T _{clB is} simultaneously ramped up linearly. The clutch is liable at the beginning of the cross-fading and the lower gear is not disengaged if clutch A is torque-free. Since the gear on clutch A is not disengaged, clutch A transmits a torque even at time t ₁ . But then no longer from the engine to the transmission, but the other way round. As a result, the acceleration continues to decrease until clutch A finally comes into the slip phase. This procedure is also indicated in FIG. 2.

According to the invention it can also be provided that, depending on the Condition in the gearbox, e.g. B. tensile or shear load, and the type of circuit, for. B. Upshift or downshift, the moment in the drivetrain through coordinated Actions of the two clutches and the engine is controlled so that preferably drive train vibrations are suppressed and a maximum Driving comfort is made possible.

The control strategy according to the invention relates in particular to Dual clutch transmission, which has two transmission input shafts, each be connected to the crankshaft of the engine by a friction clutch can. At one input shaft, for. B. gear steps 1, 3 and 5 are while gear stages 2, 4 and 6 are on the other input shaft. easy Upshifts and downshifts can e.g. B. be carried out so that the Target gear is engaged on the disengaged input shaft (shaft 2) Input shaft of the actual gear (shaft 1) is disengaged while shaft 2 is engaged and the engine speed is adjusted to the speed of shaft 2.

Such a transition from wave 1 to wave 2 results in a Interruption of traction avoided. It should be ensured that neither through the Crossfading of the clutches by the subsequent synchronization of the Engine speed a noticeable jerk is excited, which is namely the driving comfort deteriorated.

In the strategy according to the invention, a jerk-free change of translation is included a suitable synchronization is provided. For example, the Strategy according to the invention can also be used for pilot control and then can be combined with a corresponding regulation.

In order to appropriately illustrate the proposed control strategy, a corresponding model of the drive train with corresponding equations of motion is introduced. The drive train consists of an internal combustion engine with the moment of inertia J _e . Its torque T _e is controllable. Furthermore, two parallel friction clutches are provided which can be connected to the internal combustion engine and whose transmissible torques T _cl1 and T _cl2 can be controlled independently of one another. Furthermore, two partial transmissions are provided, which can be coupled to the internal combustion engine via the friction clutches. The gear ratios of one sub-transmission is i ₁ and that of the other is i ₂ . Both sub-transmissions have a common output shaft. Furthermore, a vehicle with an moment of inertia J _{v is} taken into account, which is connected to the output shaft of the transmission. A torque T _r acts on the vehicle, which is the sum of all driving resistances.

In Fig. 3 is a sketch of this model is shown. A schematic representation of a drive train with a double clutch is shown. T _e is the torque of the internal combustion engine, J _e is its moment of inertia. T _cl1 and T _cl2 are the transmissible torques from clutch 1 and clutch 2. With i ₁ and i ₂ , the ratios of the gears that are engaged on the transmission shafts 1 and 2 are designated. J _v stands for the moment of inertia of the vehicle and T _r stands for the sum of all moments acting from the output side.

Based on this model, three cases in particular can differ to let:

1. Clutch 1 sticks, clutch 2 slips

The moment balance provides the following two equations:

J _e ≙ _e = T _e - T * | cl1 + sign (i ₂ ω _v - ω _e ) .T _cl2 (1a)

J _v ≙ω _v = T _r + i ₁ .T * | cl1 - i ₂ .sign (i ₂ ω _v - w _e ) .T _cl2 (2a)

T * _cl1 is the torque transmitted by clutch 1. Since the engine and the vehicle are rigidly connected to one another via the coupling 1; the following applies:

If equation (3a) is _substituted into equations (1a) and (2a) and equation (4a) into equation (2a), T * _{cl1 is} eliminated and the following equation of motion is obtained:

The condition for the clutch 1 sticking is determined by | T * _cl1 | <T _cl1 T * _cl1 determined. This condition is determined by _substituting equation (5a) into equation (1a) and _solving for T _cl1 . The following equation then results for this holding condition of clutch 1:

2. Clutch slips, clutch 2 is stuck

Analogously to case 1, the equation is:


with the condition of detention

3. Slip clutch 1 and clutch 2

The moment balance provides the following two equations:

It has been shown that the clutches have as little stick slip or Slip-hold transitions should have to prevent unwanted jerky movements avoid. Assuming that before and after the translation change the each transferred clutch should adhere, the clutch 1 can have at least one Stick-slip transition and the clutch 2 at least one slip-stick transition Experienced. These transitions can be made smooth.

• 1. Dabei kann vorgesehen sein, dass die noch übertragene Kupplung 1 an ihre Schlupfgrenze herangefahren wird. Da die Kupplung 2 noch kein Moment überträgt, liegt nach Gleichung 6a die Schlupfgrenze von der Kupplung 1 bei
  
  
  Bei dieser Gleichung sind auf der rechten Seite bis auf T_r sämtliche Größen bekannt. Aus der Gleichung 5a, welche nach T_r aufgelöst wird, ergibt sich für T_r folgendes:
  
  
  Wenn die Gleichung 11a in die Gleichung 10a eingesetzt wird ergibt sich die folgende Gleichung
  
  T_cl1 = |T_e - J_e. ≙_e| (12a)
  
  wobei ≙_e durch fortlaufende Messung und numerische Differentiation der Motordrehzahl bestimmt werden kann. Auf diese Weise wird schließlich T_r ermittelt.
• 2. Ferner wird durch kurzzeitige Erhöhung des Motormoments die Motordrehzahl ebenfalls erhöht, um dadurch eine Schlupfreserve aufzubauen. Diese Schlupfreserve soll gewährleisten, dass beim Überblenden der Kupplung 1 auf die Kupplung 2 die Kupplung 2 nicht wieder anfängt zu haften.
• 3. Des weiteren kann die Kupplung 1 stetig auf den Wert 0 Nm zurückgefahren werden. Vorzugsweise zeitgleich kann die Kupplung 2 stetig von dem Wert 0 Nm an ihre zu erwartende Schlupfgrenze gefahren werden. Dies ist vorzugsweise jenes Kupplungsmoment, bei dem die Kupplung gerade noch schlupfen kann, vorausgesetzt, dass die Motordrehzahl schon synchronisiert ist. Analog zu der Gleichung 10a liegt diese Schlupfgrenze der zweiten Kupplung bei:
  
  
  Dabei ist der Fahrtwiderstand T_r, welcher in die rechte Seite der Gleichung 13a eingeht, durch Auflösen der zweiten Gleichung des Systems nämlich Gleichung 9a bestimmbar; nämlich durch folgende Gleichung:
  
  T_r = J_vn ≙_v + i₁.sign(i₁ω_v - ω_e).T_cl1 + i₂.sign(i₂ω_v - ω_e).T_cl2 (14a)
  
  wobei ≙_v hinreichend genau durch fortlaufende Messung und numerische Differentiation eine der beiden Getriebeeingangsdrehzahlen bestimmt werden kann.
• 4. Schließlich kann die Motordrehzahl durch eine kontrollierte, kurze Reduzierung des Motormoments mit der Drehzahl der zweiten Getriebeeingangswelle synchronisiert werden. Eine mögliche Strategie für einen derartigen Motoreingriff wird nachfolgend dargestellt.

A strategy described below provides for a smooth gear change to a higher gear without a reduction in tractive power with subsequent synchronization of the engine speed.

• 1. It can be provided that the still transmitted clutch 1 is moved up to its slip limit. Since clutch 2 does not yet transmit any torque, the slip limit of clutch 1 is included according to equation 6a
  
  
  In this equation, all quantities except T _r are known on the right. From Equation 5a, which is solved for T _r, T _r is obtained for the following:
  
  
  When equation 11a is substituted into equation 10a, the following equation results
  
  T _cl1 = | T _e - J _e . ≙ _e | (12a)
  
  where ≙ _{e can be} determined by continuous measurement and numerical differentiation of the engine speed. In this way, T _r is finally determined.
• 2. Furthermore, the engine speed is also increased by briefly increasing the engine torque, in order to build up a slip reserve. This slip reserve is intended to ensure that when the clutch 1 is blended onto the clutch 2, the clutch 2 does not start to stick again.
• 3. Furthermore, clutch 1 can be continuously reduced to the value 0 Nm. The clutch 2 can preferably be moved continuously from the value 0 Nm to its expected slip limit at the same time. This is preferably the clutch torque at which the clutch can just slip, provided that the engine speed is already synchronized. Analogous to equation 10a, this slip limit of the second clutch is:
  
  
  The driving resistance T _r , which goes into the right-hand side of equation 13a, can be determined by solving the second equation of the system, namely equation 9a; namely by the following equation:
  
  T _r = J _v n ≙ _v + i ₁ .sign (i ₁ ω _v - ω _e ) .T _cl1 + i ₂ .sign (i ₂ ω _v - ω _e ) .T _cl2 (14a)
  
  where ≙ _v one of the two transmission input speeds can be determined with sufficient accuracy by continuous measurement and numerical differentiation.
• 4. Finally, the engine speed can be synchronized with the speed of the second transmission input shaft by a controlled, brief reduction in the engine torque. A possible strategy for such an engine intervention is shown below.

For example, it is possible that the translation change at full load is carried out and the engine torque in the aforementioned phase 2 no longer can be increased via the driver's desired torque. In this case, a Build up the slip reserve by in the aforementioned phase 1 the moment of Clutch 1 is driven slightly below the slip limit.

• - Die Phase 2 wird solange aufrecht erhalten, bis der Motor und die Getriebeeingangswelle des neuen Ganges synchron laufen. Die Durchführung der Phase 4 kann dann entfallen.
• - Falls steuerungstechnisch eine exakte Synchronisation von ω_e und ω₂ nicht gelingt, könnte vorzugsweise die Phase 2 solange aufrecht erhalten werden, bis ω_e > ω₂ ist. Die Synchronisation kann dann in Phase 4 nachgeholt werden. Sie gelingt dann eher, da die Kupplung 2 in dieser Phase das gesamte Moment überträgt.

With a modified strategy, downshifts can also be carried out similarly to train shifts. There may be a difference in phase 2. In order to obtain a gear change as smoothly as possible, it may not be sufficient to raise the engine speed ω _e above the speed ω _{1 of} the transmission input shaft that is still being transmitted. If ω _{e is} still below the speed ω _{2 of} the new transmission input shaft during the cross-fading of the two clutches in phase 3, the engine via clutch 2 reduces the vehicle acceleration until ω _e and ω ₂ assume approximately the same values. Therefore, the following modification of the aforementioned phase 2 is also useful for downshifts.

• - Phase 2 is maintained until the engine and the transmission input shaft of the new gear run synchronously. Phase 4 can then be omitted.
• - If an exact synchronization of ω _e and ω ₂ does not succeed in terms of control technology, phase 2 could preferably be maintained until ω _e > ω ₂ . The synchronization can then be made up in phase 4. It is more likely to succeed because clutch 2 transmits the entire torque in this phase.

In particular, to check the effectiveness of the strategy presented these are simulated using simulation. It is e.g. B. possible that the Equations of motion 5a, 7a and 9a are integrated. Before and after Circuit, d. H. before phase 1 and after phase 4, the transferable moment of transmitted clutch can be kept 20% above the transmitted torque.

The following vehicle data result as an example:
J _e : 0.25 kg.m ²
J _v : 70 kg.m ²
T _r : -5 Nm-10 Nmh / km
Radius: 0.3 m
Gear ratio 1:14
Gear ratio 2: 10

With the proposed simulation it can be assumed that the Engine torque can be set instantaneously and precisely. Possible deviations this behavior can be easily quantifiable (e.g. dead time element and / or PT1 behavior).

In FIG. 4, the time course of various sizes is shown during a ratio change from gear 1 to gear 2 according to the above strategy. The speed of the transmission input shaft 1, the speed of the transmission input shaft 2 are indicated in the upper diagram as a solid line, and the motor speed is represented by an x curve. The stepped course shows which clutch is slipping. In the middle diagram, the transmissible torque of clutch 1 and the transmissible torque of clutch 2 are each indicated by solid lines, and the engine torque desired by the driver is indicated by cross-shaped markers. The step-by-step process shows the number of the tax phase. The vehicle speed and the vehicle acceleration over time are shown in the diagram below.

Particular attention should be paid to the acceleration signal in the diagram below. During the transition of the couplings in phase 3 the Acceleration linear and without jerk. The little one from the driver probably hardly perceptible jerk after phase 4 arises because the synchronization in the Phase 4 has not been completed. With this deviation from the The ideal strategy, which was artificially represented in the simulation, is also part of the vehicle.

In FIG. 5, the timing of the same sizes as shown in FIG's. 4 shown over time, whereby a 1-2 shift without the phases 1 and 2 is used here as well. In the upper diagram, the speed of the transmission input shaft 1 and the speed of the transmission input shaft 2 are each shown with a solid line and the motor speed, which is marked with cross-shaped markers. The stepped course shows which clutch is slipping. In the middle diagram, the transmissible torque of clutch 1 and the transmissible torque of clutch 2 are each indicated by a solid line, the engine torque desired by the driver additionally being marked with cross-shaped markers. Furthermore, the current engine torque is indicated by the step-like course. Another step-like display shows the number of the tax phases. In the lower diagram, the vehicle speed and the vehicle acceleration are each indicated by a solid line. The discontinuity of the acceleration curve is clearly recognizable at about 3.75 seconds. This jerk is provided at the end of phase 3, that is to say when the clutches crossfade.

The case of a downshift from 2nd gear to 1st gear is illustrated in FIG. 6. In FIG. 6 the curve is represented by a number of variables during a 2-1- circuit. The upper diagram shows the speed of the transmission input shaft 1 and the speed of the transmission input shaft 2 by means of a solid line. The engine speed is identified by cross-shaped markers. The step-wise course shows which clutch is slipping. In the middle diagram, the transmissible torque of clutch 1 and the transmissible torque of clutch 2 are each represented by a solid line. The engine torque desired by the driver is in turn indicated by cross-shaped markers, the current engine torque likewise having a step-shaped course. Furthermore, the number of the tax phase is indicated by a course. In the lower diagram, the vehicle speed and the vehicle acceleration are each indicated by a solid line.

Phase 2 lasts a little longer than upshifting because the engine speed is to be brought up to the speed of the shaft by clutch 2. If this is not completely successful, a small jerk occurs in phase 3, which is visible in the acceleration signal. Phase 4 is not carried out, since in phase 3 the equality of ω _e and ω ₂ has already been achieved.

Overall, a control strategy for the crossfading of a Transmission shaft to the other transmission shaft in a double clutch transmission can be specified. The tax strategy is designed so that the change is not only is carried out without reducing the tractive force, but - provided the engine torque and Clutch torque can be controlled with sufficient accuracy - even without a noticeable jerk. If the exact controllability of the moments is not guaranteed, the Pre-control strategy used to regulate the Coupling torques is superimposed. Overall, with the specified Strategy gearshifts in dual clutch transmissions with a minimum of Loss of comfort and thus the quality of a conventional one Step machines are outbid.

A further development of the present invention can provide that the engine speed is synchronized after the upshift. With an upshift, at the end of the torque transfer (phase 3), the engine rotates higher than the transmitting transmission input shaft. This slip can be reduced by reducing the engine torque T _e . It is conceivable that e.g. B. T _{e is} reduced to 20 Nm. The engine torque can typically follow the requested torque after a dead time of approximately 100 ms. The torque specification for the synchronization should therefore not only be ended when the slip has been completely eliminated, but preferably earlier. To do this, it may be necessary to estimate the remaining duration of the slip reduction. It should be borne in mind that the clutch 1 is fully open and that the angular velocity ω _{2 of} the transmission input shaft is i ₂ times ω _v . In this way, the equation given under 9a can be considerably simplified and the following equation follows:

Equation 15a can be easily integrated over a time range τ which is chosen to be so short that T _e , T _r and T _cl2 practically do not change, and it follows:

At the end of the synchronization, approximately ω _e = ω ₂ applies. By equating the right sides of the system specified with equations 16a and then resolving them after τ, the remaining duration of the synchronization phase can be estimated and it follows:


where the current speed difference ω (0) - ω ₂ (0) is abbreviated by Δω. During the synchronization phase, the control code can continuously evaluate from equation 17a as soon as τ falls below the dead time of the engine, the requested torque reduction should be canceled and the driver's desired torque should be brought to the foreground again. Downshifts can be synchronized before crossfading, for example. To do this, the engine torque can be increased instead of reduced. Equation 17a can, however, also be used in this case to estimate the synchronization duration.

In an advantageous embodiment of the present invention, four Different shift types are defined, namely train upshift, push Upshift, train downshift and push downshift. Here, the correct the described lack of comfort with the following fade strategy, the initial gear speed as the speed of the transmission input shaft Starting gear (starting gear is engaged) and the target gear speed as speed the gearbox input shaft of the target gear (target gear is engaged) is defined.

This further control strategy preferably comprises six below described phases, which possibly also with the already specified Phases of the aforementioned strategies agree and / or also suitably with these can be combined.

In phase 1, the state of sticking to slip exists. By linear The clutch torque of the transmitted clutch can be reduced in the slip will be brought. At the end of phase 1, i.e. immediately before the start phase 2, it is decided whether the drive train is in the train or Overrun condition. When the slip on the transmitted clutch is positive (Motor speed> speed of the input shaft), can be concluded from this, that there is a train condition. With a negative slip (engine speed <speed the input shaft) is a thrust condition. As soon as the clutch slips proceeded to phase 2.

In phase 2, a so-called slip reserve is built up. This can be done Torque of the transmitting clutch kept constant (or at slip limit) become. By a suitable increase or decrease in the engine torque with respect to of the driver's desired torque, the engine speed is brought to the target speed. If the engine torque is insufficient to reach the target speed within an acceptable time can achieve, for example, the torque of the transmitting clutch be reduced. The target speed when the cable is switched is the maximum of Starting gear speed and the target gear speed plus a slip reserve. at Thrust circuits is the target speed defined by the minimum of Starting gear speed and the target gear speed reduced by the slip reserve.

In phase 3, cross-fading is provided. The moment of the transferred Coupling can preferably with a constant ramp z. B. linearly to the value 0 be reduced while at the same time clutching the target gear is driven to the slip limit.

In the tax strategy proposed here, phase 4 does not continue considered, however, it is mentioned here because it is part of the total Dual clutch strategy is.

Phase 5 is called motor synchronization. With a suitable increase or lowering the engine torque with respect to the driver's desired torque the engine speed can be brought to a desired target speed. If that Engine torque is insufficiently large or small to within the target speed to reach an acceptable time can e.g. B. additionally the torque of the clutch Target gear can be increased. The target speed for train circuits can change by the speed of the transmission input shaft and the added slip reserve result and with overrun circuits the target speed is the speed of the Gearbox input shaft of the target gear reduced by a slip reserve.

Phase 6 is described as slipping to hold. By appropriate increase or decrease in engine torque with respect to Driver's desired torque and / or by closing the clutch of the target gear, the engine speed is brought up to the speed of the clutch of the target gear until the clutch of the target gear passes to stick. The slip transfer is supposed to be done with a smoothing to make a smooth transition and avoid jerky movements.

The proposed control strategy is illustrated in FIG. 7 on the basis of various shift types, namely the train upshift, the push upshift, the train downshift and the push downshift. For each switching type, an upper diagram, a middle diagram and a lower diagram are shown in a column.

The diagram shows the course of different sizes during the respective Circuit shown over time or over the switching phases. In the top Diagrams is the speed of the transmission input shaft with A, the speed of the Gearbox input shaft marked with B and the engine speed with C. In the middle diagrams is the current engine torque with solid Line and the driver's desired torque indicated with a dashed line. In the The diagrams below are the moments transmitted by the couplings Coupling A and coupling B specified.

Overall, knowledge of the slip limits for the clutch is crucial. In phase 2, the torque of the clutch that is still transmitting should be kept at the slip limit. In phase 3, the torque of the clutch of the target gear is driven to the expected slip limit and in phase 5 it is kept at the slip limit. These relationships are particularly clear from FIG. 7.

In this control strategy as well, the slip limits of the clutches can be calculated using the following formula according to a simple model of the drive train, which results from FIG. 8:

• 1. wenn die Kupplung A haftet und die Kupplung B schlupft nach folgender Gleichung:
  
  
  
• 2. Wenn bei gerade schlupfender Kupplung A (Kupplung B ist offen) kann dieser durch Auflösen der oben angegebenen Gleichung 1b nach T_r ermittelt werden:
  
  
  wobei ω_eng die Motorgeschwindigkeit, ω_vehicle die Fahrzeuggeschwindigkeit und ≙_eng die Motorbeschleunigung ist.

Formula 1b above assumes that the slip limit of clutch A is calculated and that clutch B slips or is open. Knowledge of the driving resistance T _{vehicle is} required to evaluate this formula. This can preferably be determined in two ways:

• 1. when clutch A is stuck and clutch B slips according to the following equation:
  
  
  
• 2. If clutch A is just slipping (clutch B is open), this can be determined by solving equation 1b according to T _r given above:
  
  
  where ω _{narrow is} the engine speed, ω _{vehicle is} the vehicle speed and ≙ _{narrow is} the engine acceleration.

It has been found that it is beneficial T _vehicle according to the formula 3b to calculate, as the motor acceleration ≙ is subjected _closely strong fluctuations. In addition, the subsequent calculations of clutch slip limits are consistent with the calculation of T _vehicle because they are based on the same equation. It is particularly advantageous in the proposed control strategy that a reduction or increase in the vehicle acceleration below or above the acceleration in the target gear is avoided. In this way, comfort impairment can also be almost completely prevented.

In Fig. 9 a corresponding flow diagram of the proposed switching strategy is shown in dual clutch transmissions.

A further possible embodiment of the presented here is given below Invention described in which a possible strategy for controlling the Clutches in a dual clutch transmission during a shift between two gears on the same input shaft is proposed.

It was found that with a dual clutch transmission without Interruption of the traction can be switched as long as the target gear and the current gear is engaged with various input shafts. A Dual clutch control strategy is particularly necessary in the event that when shifting a gear or several gears should be skipped, so that the target gear and the starting gear are on the same input shaft, d. H. that the Target gear is not the next higher or lower gear. This circuit can cannot be done directly through a normal dual clutch shift because With this switching method, no change of the transmitted input shaft is necessary.

Accordingly, one task is to find an appropriate one To develop dual clutch control strategy for this aforementioned case at a switching as comfortable as possible is made possible.

In order to carry out this circuit, it can be provided, for example, that preferably an intermediate course is used, in particular on the other Transmission input shaft is. This intermediate course can then take a moment Output transmitted while on the other transmission input shaft Starting gear is switched to the target gear. This way the circuit performed without interrupting the tractive force.

In a double clutch transmission, four different shift types, as already mentioned, can be defined, namely train upshifts, push upshifts, train downshifts and push downshifts. A circuit in which the target gear and the starting gear lie on the same input shaft can be defined by using an intermediate gear by means of the following cross-fading strategy, which is indicated in particular in FIGS. 10 to 13.

The target gear can be assigned, for example, to input shaft B, in which case instead of switching directly to the target gear, a gear of the input shaft A as Intermediate gear is used. The gear ratio of the intermediate course can be larger or smaller than the gear ratio of the initial gear and also larger or smaller than the gear ratio of the target gear. With regard to the Definitions of the terms initial speed and target speed are based on the referenced definitions.

In addition, there is the definition of the term intermediate gear speed. This will defined as the speed of the transmission input shaft of the intermediate gear when the Intermediate gear is engaged.

In the proposed dual clutch control strategy, there are preferably 6 Phases are provided, but phases are also omitted or added can be. It is also possible that these phases in particular with others already mentioned phases can be combined as desired.

Phase 1 and phase 2 correspond to phases 1 and 2 respectively aforementioned dual clutch control strategy, however, in phase 2 one Difference is provided, namely that phase 2 now described twice is carried out, namely when switching from the initial gear to Intermediate gear and then when switching from the intermediate gear to the target gear.

• - Bei Zugbelastung wird in den Zwischengang geschaltet, wobei die Solldrehzahl das Maximum der Anfangsgang-Drehzahl und der Zwischengang-Drehzahl plus der Schlupfreserve ist.
• - Bei Zugbeanspruchung wird in den Zielgang geschaltet, wobei die Solldrehzahl das Maximum der Zwischengang-Drehzahl und der Zielgang- Drehzahl plus der Schlupfreserve ist.
• - Bei Schubbeanspruchung wird in den Zwischengang geschaltet, wobei die Solldrehzahl das Minimum der Anfangsgang-Drehzahl und der Zwischengang- Drehzahl vermindert um die Schlupfreserve ist.
• - Bei Schubbeanspruchungen wird in den Zielgang geschaltet, wobei die Solldrehzahl das Minimum der Zwischengang-Drehzahl und der Zielgang- Drehzahl vermindert um die Schlupfreserve ist.

With this dual clutch control strategy, the target speed can be defined as follows:

• - If there is a tensile load, the gear is shifted into the intermediate gear, the target speed being the maximum of the starting gear speed and the intermediate gear speed plus the slip reserve.
• - In the event of a tensile load, the gear is switched to the target gear, the target speed being the maximum of the intermediate gear speed and the target gear speed plus the slip reserve.
• - In the event of thrust, the gear is shifted into the intermediate gear, the target speed being the minimum of the starting gear speed and the intermediate gear speed reduced by the slip reserve.
• - In the event of thrust loads, the gear is switched to the target gear, the target speed being the minimum of the intermediate gear speed and the target gear speed reduced by the slip reserve.

In phase 3, the state is "crossfade". The moment of transmitting clutch can preferably be with a constant ramp for example linearly reduced to the value 0, while at the same time the Torque of the clutch of the other input shaft has reached the slip limit becomes. This phase can be the one presented here Dual clutch transmission control strategy preferably performed twice become. First when switching from the starting gear to the intermediate gear and then when switching from the intermediate gear to the target gear.

In the 4th phase, the "gear change" state exists. The starting gear can be removed and the target gear inserted.

• - Bei Zug-Schaltungen ist diese die Summe aus Zielgang-Drehzahl und Schlupfreserve.
• - Bei Schubschaltungen ist die Solldrehzahl die Differenz aus der Zielgang- Drehzahl und der Schlupfreserve.

Phase 5 is identical to phase 5 of the dual clutch transmission control strategy previously described, with one difference in the definition of the target speed. In the present dual clutch transmission control strategy, the target speed is defined as follows:

• - In the case of train gearshifts, this is the sum of the target gear speed and the slip reserve.
• - In the case of overrun circuits, the target speed is the difference between the target gear speed and the slip reserve.

Phase 6 is identical to phase 6 of that previously described Dual clutch transmission control strategy. Therefore here is a new one Description omitted.

A flow diagram of the strategy presented here for controlling the clutches in a dual clutch transmission is shown schematically in FIG. 14. The strategy presented is preferably used during a shift between two gears on the same input shaft.

In Figs. 10 to 13 12 are diagrams respectively shown in three columns, the diagrams are indicated for another switching process in each column. The first row or line shows the engine speed with the solid line I, the speed of the transmission input shaft A with the solid line II and the speed of the transmission input shaft B with the dashed line for different gears over time. In the middle upper line, the current engine torque is shown with a solid line and the driver's desired torque with a dashed line over time. In the middle bottom line, the torque transmitted by clutch A is shown with a solid line and the torque transmitted by clutch B with a dashed line over time. The lower line shows the engaged gears of the input shaft over time or the switching phases, the solid line indicating the engaged gears of input shaft A and the broken line indicating the engaged gears of input shaft B. These aforementioned designations apply to all of FIGS. 10 to 13.

In Fig. 10 the aforementioned quantities are shown during a train upshift over time. The left column of FIG. 10 shows the diagrams which result from a 2- (1) -4 train upshift. The gear ratio of the intermediate gear is larger than the gear ratio of the initial gear. The middle column in FIG. 10 shows a 2- (3) -4 train upshift, the gear ratio of the intermediate gear being between the gear ratio of the starting gear and the target gear. The right column in FIG. 10 shows a 2- (5) -4 train upshift, the gear ratio of the intermediate gear being smaller than the gear ratio of the target gear.

FIG. 11 shows a possible course of the variables mentioned during an overrun upshift. The left column of FIG. 11 shows a 2- (1) -4 thrust upshift, in which the gear ratio of the intermediate gear is greater than the gear ratio of the initial gear.

In the middle column of FIG. 11, a 2- (3) -4 thrust upshift is shown, in which the gear ratio of the intermediate gear lies between the gear ratio of the starting gear and the target gear. The right column of FIG. 11 shows a 2- (5) -4 thrust upshift in which the gear ratio of the intermediate gear is smaller than the gear ratio of the target gear.

In Fig. 12 possible variations of the above-mentioned quantities are shown during a tensile downshift. A 4- (1) -2 train downshift is shown in the left column, the gear ratio of the intermediate gear being greater than the gear ratio of the target gear. In the middle column, a 4- (3) -2 train downshift is indicated, whereby this gear ratio of the intermediate gear lies between the gear ratio of the starting gear and the target gear. The right column shows a 4- (5) -2 train downshift, the gear ratio of the intermediate gear being smaller than the gear ratio of the starting gear.

In Fig. 13 possible variations of the mentioned quantities are shown during a push downshift. The left column of FIG. 13 shows a 4- (1) -2 thrust downshift, the gear ratio of the intermediate gear being greater than the gear ratio of the target gear. A 4- (3) -2 thrust downshift is shown in the middle column, with the gear ratio of the intermediate gear lying between the gear ratio of the starting gear and the target gear. The right column of FIG. 13 shows a 4- (5) -2 thrust downshift, the gear ratio of the intermediate gear being smaller than the gear ratio of the initial gear.

A model can also be created for this control strategy for the dual clutch transmission, as already shown in FIG. 8. The formulas relating to FIG. 8 can also be used in this strategy. It is particularly advantageous with this control strategy that an interruption of the tractive force during the shift is avoided. This prevents disadvantageous impairments in comfort.

A further embodiment of the invention presented here follows described in which another control strategy for a dual clutch transmission is proposed.

It should be a simple and accurate method for determining the external Vehicle torque can be specified for the dual clutch transmission.

To implement this strategy, it is particularly necessary to determine the external vehicle _torque T _vehicle in the dual clutch transmission. In particular, the determination of the sum of the external moments that influence the vehicle, such as. B. the air resistance, the frictional resistance between the road and the wheels, the slope of the road (gravity of the vehicle during a downhill / downhill) and / or the application of the vehicle brake, required.

A simplified model of the double clutch transmission, which is already shown in FIG. 8, can be used for this purpose.

• 1. Kupplung A haftet und die Kupplung B ist offen oder schlupft. Daraus ergibt sich folgende Gleichung für das externe Fahrzeugmoment T_vehicle:
  
  
  
• 2. Bei dieser Methode haftet die Kupplung A und die Kupplung B ist vorzugsweise offen. Durch lineares Herunterfahren des Kupplungsmomentes der übertragenden Kupplung A wird diese in den Schlupf gebracht. Insbesondere nach dem Haft-Schlupf-Übergang der Kupplung A kann das externe Fahrzeugmoment bestimmt werden. Es ist auch möglich, dass das externe Fahrzeugmoment nach einem Schlupf-Haft-Übergang bestimmt wird. Bei dieser zweiten Methode ergibt sich das externe Fahrzeugmoment nach folgender Gleichung:
  
  
  
• 3. Bei der dritten Methode haftet keine der Kupplungen und das externe Fahrzeugmoment ergibt sich durch folgende Gleichung:
  
  T_vehicle = ≙_vehicle.J_vehicle - sign(ω_eng - i_A.ω_vehicle).i_A.T_clA - sign(ω_eng - i_B.ω_vehicle).i_B.T_clB (3c)
  

It has been shown that at least three methods for determining the vehicle torque are possible.

• 1. Clutch A sticks and clutch B is open or slips. This results in the following equation for the external vehicle _torque T _vehicle :
  
  
  
• 2. With this method, clutch A is liable and clutch B is preferably open. By linearly reducing the clutch torque of the transmitting clutch A, the clutch A is brought into slip. The external vehicle torque can be determined in particular after the clutch A slip-slip transition. It is also possible that the external vehicle torque is determined after a slip-stick transition. With this second method, the external vehicle torque results from the following equation:
  
  
  
• 3. In the third method, none of the clutches is liable and the external vehicle torque results from the following equation:
  
  T _vehicle = ≙ _vehicle .J _vehicle - sign (ω _eng - i _A .ω _vehicle ) .i _A .T _clA - sign (ω _eng - i _B .ω _vehicle ) .i _B .T _clB (3c)
  

In the aforementioned equations, ω _{narrow is} the engine speed, ω _vehicle the vehicle speed,, _narrow the engine acceleration, ≙ _vehicle the vehicle acceleration, T _narrow the engine torque, J _narrow the moment of inertia of the engine, i _A the overall gear ratio of the first input shaft, eats the overall gear ratio of the gear of the second input shaft, T _clA the transmitted clutch _{torque of} the first clutch and T _clB the transmitted clutch _{torque of} the second clutch.

In Fig. 15 is a flowchart of the estimate of the external torque vehicle is shown. It has been shown that it is advantageous if the external vehicle torque T _{vehicle is determined} according to the second method presented, since the engine acceleration ≙ _narrow and the vehicle acceleration ≙ _vehicle are subject to strong fluctuations. In addition, the subsequent calculations of the slip limits of the clutches are then consistent with the calculation of the external vehicle _torque T _vehicle , since these are based on an equation.

In the strategy presented here for estimating the external vehicle torque in a dual clutch transmission is particularly advantageous that the Calculations of the slip limits of the clutches are possible and therefore the Implementation of a clutch control strategy in the simplest way possible.

Furthermore, another possible embodiment of the invention presented here described. As part of this configuration, a method is proposed with which the determination of the power flow in a transmission, in particular in a Dual clutch transmission is possible to preferably implement the to improve the aforementioned dual clutch control strategies.

With the method according to the invention, a simple and as accurate as possible Method for determining the power flow in the double clutch transmission proposed, which can also be used to determine whether a train or a Thrust condition.

Three situations within a dual clutch circuit are preferred shown, on the basis of which the determination of the power flow in the drive train is possible is. The train-push estimate can preferably be carried out immediately before the switching, ie before the start of the clutch change, so that in advantageously the result obtained is as up to date as possible. It is also possible perform this estimate at another suitable time.

This determination of the power flow according to the present invention can be provided on the basis of a model of the double clutch transmission shown as an example in FIG. 8.

• 1. Wenn eine Kupplung offen ist und die andere Kupplung schlupft oder haftet, kann durch lineares Herunterfahren des Kupplungsmomentes der übertagenden Kupplung diese auch in den Schlupf gebracht werden, falls sie sich nicht schon im Schlupf befindet. Zu Beginn der Schlupfphase der Kupplung kann entschieden werden, ob sich der Antriebsstrang unter Zugbelastung oder Schubbelastung befindet. Wenn der Schlupf an der übertagenden Kupplung positiv ist (Motorgeschwindigkeit ω_eng > Geschwindigkeit der aktiven Eingangswelle ω_inpshaft), so liegt eine Zugbelastung vor. Bei einem negativen Schlupf liegt dann eine Schubbelastung vor. Demnach ergeben sich folgende Bedingungen:
  Zug: ω_eng ≥ ω_inpshaft (positiver Schlupf)
  Schubbelastung: ω_eng < ω_inpshaft (negativer Schlupf)
• 2. Hier wird der Zustand einer haftenden Kupplung vorgesehen. Wenn das eingetragene Motormoment T_eng größer ist als das dynamische Motormoment ≙_eng .J_eng befindet sich das Getriebe im Zugzustand, d. h., der Motor würde dann beschleunigt werden, falls die geschlossene Kupplung in diesen Moment geöffnet wird. Im umgekehrten Fall liegt ein Schubzustand vor, wenn nämlich das eingetragene Motormoment T_eng kleiner als das dynamische Motormoment ≙_eng.J_eng ist, wobei ≙_eng die Motorbeschleunigung und J_eng das Trägheitsmoment des Motors ist. Daraus ergeben sich folgende Bedingungen:
  Zugzustand: T_eng ≥ ≙_eng.J_eng
  Schubzustand: T_eng < ≙_eng.J_eng
• 3. Hier wird der Zustand betrachtet, bei dem keiner der Kupplungen haftet. Ein Zug-Zustand kann dadurch definiert werden, dass die Summe der übertragenen Momente der Kupplungen positiv ist. Das übertragene Kupplungsmoment T transferred|clA/B ist gleich dem Minimalwert aus dem eingestellten Kupplungsmoment T_cla/b und der Schlupfgrenze der Kupplung T slip|clA/B. Daraus ergeben sich folgende Bedingungen:
  
  
  

Three possible methods are presented below, which can be implemented by the method according to the invention, whereby other methods are also possible:

• 1. If one clutch is open and the other clutch slips or sticks, by linearly reducing the clutch torque of the transmitting clutch, it can also be brought into slip, if it is not already in slip. At the beginning of the slip phase of the clutch, a decision can be made as to whether the drive train is under tensile or shear loads. If the slip on the overlying coupling is positive (motor speed ω _eng > speed of the active input shaft ω _inpshaft ), there is a tensile load. If there is negative slip, then there is a shear load. This results in the following conditions:
  _Tension : ω _narrow ≥ ω _inpshaft (positive slip)
  Shear load: ω _narrow <ω _inpshaft (negative slip)
• 2. The condition of an adhering clutch is provided here. If the entered engine torque T _{narrow is} greater than the dynamic engine torque ≙ _narrow .J _tight , the gearbox is in the pull state, ie the engine would then be accelerated if the closed clutch was opened at this moment. In the reverse case, a push condition exists, namely, when the registered engine torque T as the dynamic engine torque is ≙ .J _Eng less _tight, wherein ≙ _closely the motor acceleration, and J is _closely the moment of inertia of the engine. This results in the following conditions:
  Tension condition: T _narrow ≥ ≙ _narrow .J _narrow
  Thrust condition: T _narrow <≙ _narrow .J _narrow
• 3. Here the state is considered in which none of the clutches is liable. A tensile state can be defined in that the sum of the transmitted moments of the clutches is positive. The transmitted clutch torque T transferred | clA / B is equal to the minimum value from the set clutch _torque T _{cla / b} and the slip limit of the clutch T slip | clA / B. This results in the following conditions:
  
  
  

Here ω _{vehicle is} the vehicle speed, i _A the overall gear ratio of the first input shaft, i _B the overall gear ratio of the second input shaft, T transferred | clA the transmitted clutch torque of the first clutch and T transferred | clB the transferred clutch torque of the second clutch.

Knowing the slip limits of both clutches is crucial. The slip limits of the clutches can preferably be calculated according to a model of the drive train already specified in the aforementioned double clutch control strategies, knowledge of the driving resistance T _{vehicle being} necessary for the evaluation. As with the previous control strategies, this can also be determined in at least two ways, namely clutch A is liable and clutch B is slipping or clutch A is slipping and clutch B is open.

Through the methods shown to determine the tension / shear condition, it is advantageously possible during a double clutch shift suitable shift strategy (train upshift, push upshift, train Downshift or thrust downshift). That way unnecessary change from the pull state to the push state or vice versa and thus avoiding running through the gear play, which has a negative effect on could affect the driver's sense of comfort. Thus the Impairment of comfort minimized by the presented dual clutch strategy.

FIG. 16 shows a flow chart which illustrates the pull-push estimate according to the invention. It can be seen that the second and third methods can also be used when a clutch is open. However, the first method is preferable because it is more accurate.

A further embodiment of the present invention follows described which a control strategy for translation changes, in particular in an ESG double clutch transmission.

As part of this configuration, a tax strategy is aimed described, a fast and convenient switching by suitable control of the clutches, the engine torque and the electric motor torque. The strategy according to the invention differs in particular from the aforementioned Strategies by including an electric motor in the model. Further are upshifting and downshifting in train upshifting and pushing Upshift separated, d. H. four different switching types instead of two different types are used. Furthermore, an active control the output torque with the electric motor or one of the clutches, to the desired output torque (driver's desired vehicle acceleration) to reach.

• - Gänge an beiden Wellen werden vorgewählt, d. h. keine Gangsynchronisation wird simuliert.
• - Der Motor, der Elektromotor und die Kupplungsdynamik wird als linear angenommen und es liegt keine Totzeit vor der Rückmeldung vor. D. h. die Änderung des Moments beginnt sofort nach der Anforderung und mit einer konstanten definierten Rampe gegenüber dem angeforderten Moment. Das reale dynamische Verhalten des Motors kann einen signifikanten Effekt bei der Schaltstrategie bewirken.
• - Spiel in dem Antriebsstrang ist nicht in dem Modell berücksichtigt.
• - Dämpfung und Elastizität bei dem Antriebsstrang ist in dem Modell ebenso nicht berücksichtigt.

The drive train is simplified in this shift strategy and illustrated by the following elements shown in FIG. 17. The model consists of an internal combustion engine with a torque that is equal to the moment of inertia J _narrow and a total engine torque T _narrow . The input shaft 1 is connected to the motor via a clutch so that a maximum torque T _cl1 can be transmitted and is connected to the output shaft with a first gear with a ratio i ₁ . The input shaft 2 is also connected to the engine via a clutch so that a maximum torque T _cl2 can be transmitted and is connected to the output shaft via a second gear with the transmission ratio i ₂ . Furthermore, an electric motor with an inertia torque J _emotor and with a motor torque T _{emotor is} provided, which is permanently engaged with the input shaft 2 via a _{gear ratio} i _em . The output shaft or output shaft is connected to the engine by an moment of inertia of the vehicle J _vehicle , which is influenced by a driving resistance T _vehicle . The following simplifications are provided in the model:

• - Gears on both shafts are preselected, ie no gear synchronization is simulated.
• - The motor, the electric motor and the clutch dynamics are assumed to be linear and there is no dead time before the feedback. I.e. the change in torque begins immediately after the request and with a constant defined ramp compared to the requested torque. The real dynamic behavior of the engine can have a significant effect on the shift strategy.
• - Play in the drive train is not taken into account in the model.
• - Damping and elasticity in the drive train are also not taken into account in the model.

• 1. Der Zustand, dass die Kupplung 1 haftet und die Kupplung 2 schlupft, wird betrachtet. Dabei wird eine Momentenbilanz durch folgende Gleichung gegeben:
  
  J_eng. ≙_eng = T_eng - T *|cl₁ - sign(ω_eng - i₂.ω_vehicle).T_{cl 2} (1d)
  
  J_vehicle. ≙_vehicle = T_vehicle + i₁.T *|cl₁ + sign(ω_eng - i₂.ω_vehicle).i₂.T_{cl 2} + i₂.i_em.(T_emotor - J_emotor. ≙_emotor) (2d)
  
  T*_cl1 ist das von der Kupplung 1 übertragene Drehmoment und nicht das Moment, welches die Kupplung 1 in der Lage ist zu übertragen. Da die Kupplung 2 schlupft ist T_cl2 das steuerbare Moment der Kupplung 2. Mit der haftenden Kupplung 1 können folgende zusätzliche Gleichungen gegeben werden:
  
  
  Durch Eliminieren von T *|cl1 und durch Umformen der vorgenannten Gleichungen (1d-4d) wird folgende Gleichung gegeben:
  
  
  Folglich ist die Fahrzeugbeschleunigung durch folgende Formel gegeben
  
  
  Durch Einsetzen der Gleichung 5d in Gleichung 1d kann gezeigt werden, dass die Schlupfphase für Kupplung 1 durch folgende Gleichung gegeben ist
  
  
  
• 2. Es wird der Zustand betrachtet, bei dem die Kupplung 2 haftet und die Kupplung 1 schlupft. Analog zu dem vorherigen Fall können folgende Gleichungen angegeben werden:
  
  
  Die Schlupfbedingung für die Kupplung 2 ist:
  
  
  
• 3. Es wird der Zustand betrachtet, bei dem beide Kupplungen schlupfen. Es folgt folgende Gleichung:
  
  
  

All combinations give realistic results with the model. There are preferably three different cases that represent the dynamic behavior of the ESG dual clutch transmission:

• 1. The state that the clutch 1 sticks and the clutch 2 slips is considered. A torque balance is given by the following equation:
  
  J _eng . ≙ _eng = T _eng - T * | cl₁ - sign (ω _eng - i ₂ .ω _vehicle ) .T _{cl 2} (1d)
  
  J _vehicle . ≙ _vehicle = T _vehicle + i ₁ .T * | cl₁ + sign (ω _eng - i ₂ .ω _vehicle ) .i ₂ .T _{cl 2} + i ₂ .i _em . (T _emotor - J _emotor . ≙ _emotor ) ( 2d)
  
  T * _cl1 is the torque transmitted by clutch 1 and not the moment that clutch 1 is able to transmit. Since the clutch 2 slips, T _{cl2 is} the controllable torque of the clutch 2. The following additional equations can be given with the clutch 1 adhering:
  
  
  By eliminating T * | cl1 and transforming the aforementioned equations (1d-4d) the following equation is given:
  
  
  As a result, vehicle acceleration is given by the following formula
  
  
  By inserting equation 5d into equation 1d, it can be shown that the slip phase for clutch 1 is given by the following equation
  
  
  
• 2. The state is considered in which the clutch 2 adheres and the clutch 1 slips. Analogous to the previous case, the following equations can be given:
  
  
  The slip condition for clutch 2 is:
  
  
  
• 3. Consider the condition where both clutches slip. The following equation follows:
  
  
  

The purpose of this switching strategy according to the invention is to be comfortable Enable switching from one wave to another. The shift comfort is thereby achieved by the smooth vehicle acceleration during the shift. I.e. poor shifting comfort is characterized in that significant and sudden deviations from the driver's desired acceleration (wheel torque) are present and in particular peaks. It should be noted that an increase or a decrease in vehicle acceleration has no impact in connection with switching from one wave to another wave if these waves have different gear ratios.

In the proposed tax strategy, there should only be two slip-to-prison transfers be provided. For example, a transition can be provided during the disengaged clutch is open and a transition can be provided while the new clutch is being closed. The engine torque can be Increase or decrease can be used actively to even slip slip Ensure transitions so that the engine speed can be controlled to Slip during shifting to achieve and engine synchronization compared to the speed of the new input shaft.

Furthermore, the torque control is provided at the output torque in order to To achieve driver's desired vehicle acceleration (wheel torque). Primarily can this goal by actively using the electric motor torque during the shift can be achieved. Secondary can also if the electric motor is not available or that desired torque is above the maximum achievable electric motor torque, one of the Couplings can be controlled to the Get driver's vehicle acceleration.

The driver's desired vehicle acceleration can be determined by the acceleration which is achieved when the vehicle with shaft 1 (acceleration shaft 1) or shaft 2 (acceleration of shaft 2) is driven, when driving with an adhesive clutch, as well as when with Driver's desired engine torque and with the clutch completely disengaged another wave is driven.

The driver's desired acceleration is then changed from the old one when switching Wave acceleration transferred to the new wave acceleration, because the Engine speed is synchronized with the new shaft. Furthermore, at the clutch control can be provided that before switching the moment the old clutch is reduced to the slip limit. The moment is at Switching the new shaft by reducing the clutch torque on the old one Shaft changes and the clutch torque on the new shaft is increased. The Circuit is finished when the engine speed changes to the new one Input shaft speed is synchronized and when the new shaft sticks.

• - Zughochschaltungen, wobei eine Hochschaltung mit positiven Moment von dem Motor zum Abtriebsmoment übertragen wird. Folglich zieht der Motor das Fahrzeug (normale Art der Hochschaltung).
• - Schubhochschaltung: Hochschaltung mit negativen Moment, welches von dem Motor zum Abtriebsmoment übertragen wird. Folglich schiebt der Motor das Fahrzeug (z. B. Hochschaltung während einer Abfahrt nach einer Beschleunigung).
• - Zugrückschaltung: Rückschaltung mit positivem Moment, welches von dem Motor zum Abtrieb übertragen wird. Folglich zieht der Motor das Fahrzeug (z. B. Kick-down).
• - Schubrückschaltung: Rückschaltung mit negativem Moment, welches von dem Motor zum Abtrieb übertragen wird. Folglich schiebt der Motor das Fahrzeug (normale Art von Rückschaltung).

In the control strategy presented here, four main switching situations are considered in particular:

• - Train upshifts, whereby an upshift with positive torque is transmitted from the engine to the output torque. As a result, the engine pulls the vehicle (normal type of upshift).
• - Thrust upshift: upshift with negative torque, which is transmitted from the engine to the output torque. As a result, the engine pushes the vehicle (e.g. upshift during a departure after acceleration).
• - Train downshift: downshift with positive torque, which is transmitted from the engine to the output. As a result, the engine pulls the vehicle (e.g. kick-down).
• - Thrust downshift: downshift with negative torque, which is transmitted from the engine to the output. As a result, the engine pushes the vehicle (normal type of downshift).

For the control strategy according to the invention, it is important to identify precisely whether there is a train or push situation. This should be done before Circuit recording can be provided (e.g. if the old coupling up to Slip limit is opened). Different estimates can be used become.

One possibility is to use the estimation according to Eq. 13d only to be carried out if a clutch is liable. A pull condition exists when there is excess torque on the motor side of the clutches. That is, the engine is accelerated when the clutches are opened. The following equation follows:

T _engine ≥ ≙ _engine .J _engine (13d)

Another possibility can be to use the estimation according to Eq. 14d to be used if T _vehicle and J _vehicle can be estimated with certainty. A train condition exists when the external torque is less than the desired torque in order to achieve the current vehicle acceleration. That is, the vehicle is decelerated when both clutches are opened. The following equation results:

T _vehicle + i ₂ .i _em .T _emotor ≤ ≙ _vehicle. (J _vehicle + (i ₂ .i _em ) ² .J _emotor ) (14d)

Furthermore, the estimate according to Eq. 15d only if both clutches slip. A tensile state exists when the sum of the transmitted clutch torques is positive. It should be noted that the transmitted clutch torque is determined by the minimum of the controllable clutch torque setting point and the clutch slip limit. The following equation results:

It can also be provided that during a circuit recording process the old clutch is opened slightly after reaching the slip limit. In this Situation, the clutch begins to slip. By watching the direction of slippage, namely by differences in the speed at the input and Output side of the clutch, can be determined when a train or Thrust situation exists. A train situation is due to positive slip marked, d. H. the speed on the inside is higher than on the outside the clutch. It should be noted that the clutch and engine dynamics (Dead time) can be subject to a query. The estimate should therefore preferably combined with one of the other estimates.

• 1. Es wird der Zustand betrachtet, bei dem die erste Kupplung haftet und die zweite Kupplung schlupft. Dabei ergibt sich folgende Gleichung
  
  
  in dieser Gleichung ist ≙ target|vehicle eine Vorhersage der Fahrzeugbeschleunigung, welche der Fahrer wünscht. Dem E-Motor liegt dabei die Aufgabe zu Grunde, das Moment T target|emotor an dem Abtrieb vorzusehen, um die Beschleunigung zu erreichen. ≙_vehicle ist eine Abschätzung des externen Moments T_vehicle, welches an dem Fahrzeug wirkt, wobei dieses bestimmt werden kann, da J_vehicle konstant und ≙_vehicle bekannt ist. Damit ergibt sich folgende Gleichung:
  
  
  Wenn die Gleichung 17d in die Gleichung 16d eingesetzt wird, ergibt sich die folgende Gleichung i:
  
  
  Aus dieser Gleichung kann T_vehicle mit der Gleichung 17d während der Schaltung abgeschätzt werden, so dass das E-Motormoment rückgekoppelt kontrolliert werden kann. Eine Vorwärtsschubstrategie schätzt T_vehicle mit Gleichung 17d sofort vor der Schaltung ab und wird in Gleichung 16d mit dem Wert für die gesamte Schaltung verwendet, da diese eine Konstante ist. Wenn Gleichung 18d verwendet wird, kann diese um einen Integralteil erweitert werden, d. h. ein PI-Regler oder dergleichen. Eine mögliche Abschätzung für ≙_vehicle ist in der folgenden Gleichung 19d vorgesehen, wo M die Masse des Fahrzeuges mit normaler Beladung (d. h. zwei Personen) und R der dynamische Radradius ist. Daraus folgt die Gleichung:
  
  ≙_vehicle = m.r² (19d)
  
  In dem folgenden Beispiel wird die Vorhersage der Fahrzeugbeschleunigung, die der Fahrer wünscht, bestimmt. Dabei wird eine Schaltung von der Welle 1 zur Welle 2 angenommen. Die geschätzte equivalente Fahrzeugbeschleunigung an der Welle 1 wird mit folgender Gleichung unter Annahme T_{cl 2} = 0 berechnet
  
  
  Die geschätzte equivalente Fahrzeugbeschleunigung an der Welle 2 unter Voraussetzung T_{cl 1} = 0 wird durch folgende Gleichung bestimmt
  
  
  Nachfolgend wird ein Gewichtungsfaktor zum Schalten von ≙ shaft1|vehicle zu ≙ shaft2|vehicle berechnet, da die Motordrehzahl gegenüber der Welle 2 synchronisiert wird. Es ergibt sich folgende Gleichung:
  
  
  Dieser Gewichtungsfaktor kann zur Vorhersage der Fahrerwunschbeschleunigung verwendet werden, während die Motordrehzahl zwischen der Drehzahl der Welle 1 und der Welle 2 liegt. Es ergibt sich folgende Gleichung:
  
  
  
• 2. Es wird der Zustand betrachtet, bei dem die Kupplung 2 haftet und die Kupplung 1 schlupft. Dabei ergeben sich die folgenden Gleichungen:
  
  
  
• 3. Es wird der Zustand betrachtet, bei dem beide Kupplungen schlupfen. Es ergeben sich folgende Gleichungen:
  
  
  
  
  

Another development of the invention can provide that in the control strategy the shift is carried out with the aid of the electric motor (sufficiently available torque). At least three different situations can be provided:

• 1. The state is considered in which the first clutch sticks and the second clutch slips. The following equation results
  
  
  In this equation, ≙ target | vehicle is a prediction of the vehicle acceleration that the driver desires. The task of the electric motor is to provide the torque T target | emotor on the output in order to achieve the acceleration. ≙ _vehicle is an estimate of the external torque T _vehicle , which acts on the vehicle, which can be determined since J _{vehicle is} constant and ≙ _{vehicle is} known. This results in the following equation:
  
  
  When equation 17d is substituted into equation 16d, the following equation i results:
  
  
  From this equation, T _vehicle can be estimated with equation 17d during the shift, so that the electric motor torque can be checked in a feedback manner. A forward thrust strategy estimates T _vehicle using equation 17d immediately before the shift and is used in equation 16d with the value for the entire shift as it is a constant. If equation 18d is used, this can be extended by an integral part, ie a PI controller or the like. A possible estimate for ≙ _vehicle is provided in the following equation 19d, where M is the mass of the vehicle with a normal load (ie two people) and R is the dynamic wheel radius. The equation follows from this:
  
  ≙ _vehicle = mr ² (19d)
  
  In the following example, the prediction of the vehicle acceleration that the driver desires is determined. A shift from wave 1 to wave 2 is assumed. The estimated equivalent vehicle acceleration on shaft 1 is calculated using the following equation assuming T _{cl 2} = 0
  
  
  The estimated equivalent vehicle acceleration on shaft 2 assuming T _{cl 1} = 0 is determined by the following equation
  
  
  A weighting factor for switching from ≙ shaft1 | vehicle to ≙ shaft2 | vehicle is calculated below since the engine speed is synchronized with respect to shaft 2. The following equation results:
  
  
  This weighting factor can be used to predict the driver's desired acceleration while the engine speed lies between the speed of shaft 1 and shaft 2. The following equation results:
  
  
  
• 2. The state is considered in which the clutch 2 adheres and the clutch 1 slips. The following equations result:
  
  
  
• 3. Consider the condition where both clutches slip. The following equations result:
  
  
  
  
  

A model for a predictive strategy is implied using equations 16d, 20d, 23d, or the like, while in a feedback control strategy, equations 18d, 22d, 25d, or the like are used. The model for the prediction strategy is preferred, but this can also be replaced by a feedback control strategy, in particular in a case in which the value of T _{vehicle is changed} significantly during the shift or the estimate of T _vehicle , ≙ _vehicle far from the correct value is removed, which z. B. can occur as a result of the loading of the vehicle (large deviation from the expected total weight of the vehicle). It is not to be expected that the comfort of the gearshift will depend on the precisely estimated values of T _vehicle and J _vehicle , but an accurate correct detection of the gearshift type is crucial for optimal shifting comfort, in particular the detection of whether a pull or a push switch , is present.

Within the scope of a further embodiment of the present invention Train upshift with the support of the electric motor if there is enough Considered moment.

The tension torque should be from the old one during the entire torque transmission Wave to the new wave remain. This can be done by increasing the Engine speed can be reached via the speed of the old shaft. In this situation the old wave is opened to reduce vehicle acceleration and the New wave closed to increase vehicle acceleration. This Reactions can cancel each other out. The electric motor torque is used to control or adjust vehicle acceleration.

It can be provided that the torque of the old clutch up to Slip limit is reduced and then the engine torque is increased until the old one Clutch slips.

Furthermore, the engine torque can be controlled via the driver's desired engine torque to ensure that the engine speed is above the speed of the old one Wave is held. The transmitted torque of the old clutch can be preferred be reduced to 0 with a constant ramp while the Transmitted torque for the new clutch preferably with the same ramp up to the slip limit or slightly above, the moment of the electric motor controlled according to equations 23d or 25d to obtain the To achieve driver's desired acceleration of the vehicle.

Furthermore, the engine torque is reduced to a minimum torque in order to Accelerate engine synchronization when the clutch torque of the old Clutch 0 is (clutch disengaged). The new clutch is easily over the Slip limit controlled, the vehicle acceleration with the electric motor is controlled further.

Furthermore, the engine torque is just below Driver's desired engine torque increased to an even slip Achieve cohesion when the engine speed matches the speed of the new one Input shaft is synchronized. Then the vehicle acceleration with the E- Motor further controlled.

This illustrated simulation of a train upshift with the support of the electric motor is shown in FIG . Three diagrams are indicated, the upper diagram showing the speed of the motor, shaft 1 and shaft 2 and the clutch over time. The middle diagram shows the moments of clutch 1, clutch 2, the motor and the electric motor over time. While the speed of the vehicle and the vehicle acceleration over time are shown in the diagram below. It should be noted that the shift can be carried out faster if the torque ramp of the clutches and the motor are increased and the new clutch has a higher torque during phase 3.

In addition, a train upshift with the support of the electric motor is included sufficient present moment is provided. The pulling torque should be on the output during the entire torque transmission from the old wave to the new wave remain. This is usually done by reducing the engine speed reached below the speed of the new shaft before the torque transmission. Otherwise, opening the old clutch and closing the new one Clutch to increase vehicle acceleration. However, with the support of the electric motor torque can be counteracted, whereby the torque is transmitted before the motor speed is below the speed of the new shaft. In this case, the new clutch can synchronize the engine speed support and therefore accelerate the circuit significantly.

It can be provided that the torque of the old clutch up to Slip limit is reduced and then the engine torque is reduced until the old one Clutch slips.

It is also possible to reduce the engine torque to a minimum. The The transmitted torque of the old clutch is transferred to the via a constant ramp Value 0 decreases while the transmitted torque of the new clutch with the same ramp to the slip limit or slightly above, the Torque of the electric motor is controlled according to equation 23d or 25d To achieve driver's desired vehicle acceleration.

Furthermore, the new clutch can reach the slip limit or slightly above it can be controlled, the vehicle acceleration with the electric motor can be regulated until engine synchronization is achieved.

It is also possible to reduce the engine torque to just below the Driver's desired engine torque to increase a continuous slip Achieve cohesion when the engine speed matches the speed of the new one Input shaft is synchronized. The vehicle acceleration can then be Motor can be controlled further.

In Fig. 19 is a simulation of a thrust upshift with the support of an E-motor is illustrated. The speed of the motor, the first shaft, the second shaft and the clutch over time is shown in the upper diagram. The middle diagram shows the moments of the first clutch, the second clutch, the electric motor and the motor itself over time. While the speed of the vehicle and the acceleration of the vehicle over time are shown in the diagram below. It should be noted that the spade in the acceleration process is a numerical effect of the simulation process and is therefore not to be expected in the vehicle.

According to a development of the present invention, a train downshift provided with the support of the electric motor when there is sufficient torque. The tension torque should be from the old one during the entire torque transmission Shaft to the new shaft on the output are retained. Usually this will by increasing the engine speed above the speed of the new shaft before Torque transmission reached. Otherwise, opening the old clutch and closing the new clutch to reduce vehicle acceleration. However, this can be counteracted with the support of the electric motor torque the torque can be transmitted before the engine speed over the speed of the new shaft increases. In this case, the new clutch can Synchronize engine speed and therefore support the circuit accelerate significantly.

It can be provided that the torque of the old clutch up to Slip limit is reduced and then the engine torque is increased until the old one Clutch slips.

It is also possible for the engine torque to be greater than the driver's desired engine torque is controlled. The transferable torque of the old clutch can preferably with a constant ramp can be reduced to 0 while the Transmittable torque of the new clutch preferably with the same ramp up to the slip limit or slightly above, the moment of the electric motor is driven according to equation 23d or 25d to the To achieve driver's desired vehicle acceleration.

Furthermore, the new clutch can reach the slip limit or slightly above it can be controlled, the vehicle acceleration with the electric motor continue to can be changed for motor synchronization.

It is also possible to keep the engine torque just below the Driver's desired engine torque to increase a continuous slip Achieve cohesion when the engine speed matches the speed of the new one Input shaft is synchronized. The vehicle acceleration is then Motor further regulated.

In Fig. 20, the simulation of a traction downshift with the assistance of the electric motor is shown. The top diagram shows the speed of shaft 1, shaft 2, the motor and the clutch over time. The middle diagram shows the moments of clutch 1, clutch 2, the motor and the electric motor over time. The speed of the vehicle and the vehicle acceleration over time are indicated in the lower diagram. It should be noted that the shift can be carried out more quickly if the torque ramp of the clutches and the motor is increased and the new clutch assumes a higher torque during phase 3.

Within the scope of a further development of the present invention, a Thrust downshift with the support of an electric motor with sufficient available moment. The thrust should be on the output during the entire torque transmission from the old wave to the new wave remain. This can be done by reducing the engine speed below the Speed of the old shaft can be reached. In this situation, by opening the old clutch increases vehicle acceleration and by closing the new clutch reduced. These reactions can cancel each other out. The electric motor torque is used to control and adjust the Vehicle acceleration used.

It is possible that the torque of the old clutch is reduced to the slip limit and then the engine torque is increased until the old clutch slips.

Furthermore, the engine torque can be reduced to a minimum torque in order to ensure that the engine speed is below the speed of the old shaft remains. The transmissible torque of the old clutch can be kept constant Ramp can be reduced to the value 0 during the transferable moment of new clutch with the same ramp up to the slip limit or slightly is increased, the torque of the electric motor according to equation 23d or 25d is driven in order to accelerate the driver's desired vehicle to reach.

The engine torque is increased by the driver's desired engine torque Accelerate engine synchronization when the moment of the old clutch (completely disengaged) assumes the value 0.

The new clutch is controlled up to the slip limit or slightly above and the vehicle acceleration is further regulated with the electric motor.

The engine torque is just reduced to the driver's desired engine torque to achieve a continuous slip-to-slip transition when the engine speed is synchronized with the speed of the new input shaft. The Vehicle acceleration is further controlled with the electric motor.

In Fig. 21, the simulation of a push downshift with the aid of an electric motor is shown. The upper diagram shows the speeds of shaft 1, the clutch and the motor as well as shaft 2 over time. The middle diagram shows the moments of clutch 1, clutch 2, the motor and the electric motor as well as the driving state over time. The lower diagram shows the vehicle speed and the vehicle acceleration over time. It should be noted that in this simulation, the electric motor only slightly supports the shifting process in order to achieve the desired vehicle acceleration, which consequently assumes almost zero. The first spade in vehicle acceleration is a numerical effect in simulation.

Within the scope of a further embodiment of the present invention, a Circuit provided without suitable support from the electric motor. The E- For example, engine torque may not be available or the torque may not adequately support vehicle acceleration. In this case the Strategy supported as well as possible by the electric motor; either with maximum positive or maximum negative moment available and it can be one of the Couplings are used to control the torque for the output To achieve vehicle acceleration.

• 1. Der Zustand wird betrachtet, bei dem die Kupplung 1 haftet und die Kupplung 2 schlupft. Daraus ergibt sich folgende Gleichung:
  
  
  Dabei kann ≙_vehicle bevorzugt mit der vorher genannten Gleichung 17d abgeschätzt werden und es ergibt sich daraus
  
  
  
• 2. Es wird der Zustand betrachtet, bei dem die Kupplung 2 haftet und die Kupplung 1 schlupft.
  
  
  Dabei wird ≙_vehicle abgeschätzt durch die vorherige Gleichung 25d und es resultiert daraus
  
  
  
• 3. Es wird der Zustand betrachtet, bei dem beide Kupplungen schlupfen. In diesem Fall kann jede Kupplung angesteuert werden, um die Fahrzeugbeschleunigung zu erreichen. Es ergeben sich folgende Gleichungen
  
  
  Dabei kann ≙_vehicle durch die vorher genannte Gleichung 28d abgeschätzt werden und es resultiert daraus
  
  
  

The following simulations assume that the electric motor is not available. Analogous to the developments and refinements of the invention described above, the following equations can be developed.

• 1. The state is considered in which the clutch 1 sticks and the clutch 2 slips. This results in the following equation:
  
  
  In this case, ≙ _{vehicle can} preferably be estimated using the aforementioned equation 17d and it results from this
  
  
  
• 2. The state is considered in which the clutch 2 adheres and the clutch 1 slips.
  
  
  Here ≙ _{vehicle is} estimated by the previous equation 25d and it results from it
  
  
  
• 3. Consider the condition where both clutches slip. In this case, each clutch can be activated to achieve vehicle acceleration. The following equations result
  
  
  Here ≙ _vehicle can be estimated by the aforementioned equation 28d and it results from it
  
  
  

A model for a predictive strategy is implied using equations 26d, 28d, 30d, 31d or the like, while in a feedback control strategy the equations 27d, 29d, 32d, 33d or the like are used. The model for the prediction strategy is preferred, but this can also be replaced by a feedback control strategy, especially in a case in which the value of T _{vehicle is} significantly changed during the shift or the estimate of J _vehicle , ≙ _vehicle far from the correct value is removed, which z. B. can occur as a result of the loading of the vehicle (large deviation from the expected total weight of the vehicle). The comfort of the gearshift is not expected to depend on the precisely estimated values of T _vehicle and J _vehicle .

The following is a train upshift without the support of the electric motor considered. The pulling moment should be on the output throughout Torque transmission from the old shaft to the new shaft is retained. This can be achieved by increasing the engine speed over the speed of the old one Wave. In this situation, opening the old clutch will cause the Vehicle acceleration is reduced and by closing the new clutch vehicle acceleration increases. With suitable and parallel Coupling operations can cancel each other out.

It is possible that the torque of the old clutch is reduced to the slip limit the engine torque is increased until the old clutch slips becomes.

Furthermore, the engine torque can be controlled via the driver's desired engine torque to ensure that the engine speed is above the speed of the old one Wave is held. The transferable torque of the old clutch is e.g. B. with a constant ramp reduced to 0 while the transferable Torque of the new clutch according to equations 30d / 32d or 31d / 33d be changed in order to achieve the driver's desired vehicle acceleration.

The engine torque is reduced to a minimum torque to the Accelerate engine synchronization when the moment of the old clutch (completely disengaged) assumes the value 0. The vehicle acceleration will then controlled via the new clutch.

The engine torque is just below the driver's desired engine torque increased to achieve an even slip-to-slip transition when the Engine speed is synchronized with the speed of the new input shaft. The Vehicle acceleration continues to be controlled with the new clutch.

In Fig. 22, the simulation of a traction upshift without assistance from an electric motor is shown. The top diagram shows the speeds of the clutch, motor, shaft 1 and shaft 2 over time. The middle diagram shows the moments of clutch 1, clutch 2, the motor and the electric motor as well as the desired motor torque over time. The vehicle speed and the vehicle acceleration over time are shown in the diagram below. It should be noted that the shift can be carried out more quickly if the torque ramps of the clutches and the motor are increased, the new clutch assuming a higher torque during phase 3.

The following is a thrust upshift without the support of the electric motor intended. The thrust should be on the output throughout Torque transmission from the old shaft to the new shaft is retained. This is done by reducing the engine speed to below the speed of the engine new shaft reached before torque transmission. In this situation, through opening the old clutch increases vehicle acceleration and through Closing the new clutch reduces vehicle acceleration. By Appropriate and parallel coupling operations can reduce these reactions be canceled against each other.

It is possible that the torque of the old clutch is reduced to the slip limit the engine torque is then reduced until the old clutch slips.

It is also possible that the engine torque is reduced to a minimum. The The torque of the old clutch is calculated according to equations 30d / 32d or 31d / 33d controlled to accelerate the driver's desired vehicle Achieve engine synchronization.

It is conceivable that the engine torque is slightly higher than the driver's desired engine torque is increased to ensure an even slip-to-slip transition if the engine speed is below the speed of the new input shaft. The torque of the old clutch is ramped down to 0 reduced while the torque of the new clutch according to the equations 30d / 32d or 31d / 33d is controlled to the Achieve driver's desired vehicle acceleration until the new clutch sticks.

In Fig. 23 a simulation of a thrust upshift without assistance from the electric motor is shown. The top diagram shows the speeds of shaft 1, shaft 2, the motor and the clutch over time. In the middle diagram, the moments of clutch 1, clutch 2, the electric motor, the motor and the driving state as well as the desired engine torque are shown over time. The vehicle speed and vehicle acceleration over time are shown in the diagram below. It should be noted that the first spade in the acceleration curve of the vehicle is a numerical effect in the simulation and is therefore not to be expected in the vehicle. It should also be noted that the switching time is very long because only a limited negative engine torque is available for engine synchronization. A faster shift can be achieved, for example, using a standard ASG shift strategy or the like.

A further development of the invention is described below Train downshift provided without the support of the electric motor. The pulling moment should be on the output during the entire torque transmission from the old one Wave to the new wave remain. This can be achieved by using the Motor speed via the speed of the new shaft before torque transmission is increased. In this situation, opening the old clutch will cause the Vehicle acceleration is reduced and by closing the new clutch vehicle acceleration increases. Through suitable and parallel Coupling operations can cancel these reactions against each other become.

It is possible that the torque of the old clutch is reduced to the slip limit and then the engine torque is increased until the old clutch slips.

Furthermore, the engine torque can be regulated via the driver's desired engine torque become. The torque of the old clutch is calculated according to equation 30d / 32d or 31d / 33d changed to achieve the driver's desired vehicle acceleration up to There is motor synchronization.

The engine torque can easily go below the driver's desired engine torque be reduced in order to ensure a smooth slip transfer, when the engine speed is higher than the speed of the new input shaft. The The torque of the old clutch can preferably be ramped up to the constant Value 0 can be reduced while the torque of the new clutch according to the Equations 30d / 32d or 31d / 33d is regulated to the Achieve driver's desired vehicle acceleration until the new clutch sticks.

In Fig. 24 a simulation of a traction downshift without assistance from the electric motor is shown. The top diagram shows the speeds of shaft 1, shaft 2, the motor and the clutch over time. The middle diagram shows the moments of clutch 1, clutch 2, the electric motor and the drive motor as well as the driving condition and the desired engine torque over time. The vehicle acceleration and speed over time are shown in the diagram below. It should be noted that faster switching can be achieved with a standard ASG switching strategy or the like.

A further development of the present invention relates to a thrust downshift without Support from an electric motor. The pulling moment should be on the output during of the entire torque transmission from the old wave to the new wave stay. This is done by increasing the engine speed above the speed of the old one Wave reached. In this situation, opening the old clutch will Vehicle acceleration reduced and by closing the new clutch Vehicle acceleration increased. Through suitable and parallel Coupling operations can cancel these reactions against each other become.

It is possible that the torque of the old clutch is up to the slip limit is reduced, the engine torque then being reduced until the old clutch slips.

Furthermore, the engine torque can be below the driver's desired engine torque be controlled to ensure that the engine speed is below the Speed of the old shaft is maintained. The transferable moment of the old Clutch is z. B. reduced to 0 with a constant ramp while the transmissible torque of the new clutch according to equations 30d / 32d or 31d / 33d is regulated to increase the driver's desired vehicle acceleration to reach.

The engine torque is increased by the driver's desired engine torque Accelerate engine synchronization when the moment of the old clutch (completely disengaged) assumes the value 0. The vehicle acceleration is with the new clutch further regulated.

The engine torque is just up to the driver's desired engine torque reduced to achieve a continuous slip transfer when the Engine speed is synchronized with the speed of the new input shaft. The Vehicle acceleration is further regulated with the new clutch.

In Fig. 25 a simulation of a push downshift without assistance from the electric motor is shown. The top diagram shows the speeds of shaft 1, shaft 2, the motor and the clutch over time. The middle diagram shows the moments of clutch 1, clutch 2, the electric motor, the drive motor and the desired engine torque as well as the driving state over time. The vehicle speed and vehicle acceleration are shown in the diagram below. It should be noted that the first spade in the acceleration curve of the vehicle is a numerical effect in the simulation and consequently does not occur in the vehicle. It should also be noted that the shift can be carried out more quickly if the torque ramps of the clutches and the motor are increased and the new clutch has a higher torque during phase 3.

Furthermore, according to a development of the present invention it is provided that both the available engine torque and the E- Engine torque is sufficient for the aforementioned control strategies. Through the The following situations can make appropriate changes in strategies be provided.

With upshifts, it can be provided that the engine torque is always on Assumes maximum before switching. If no engine torque to increase the Engine speed over the speed of the old shaft is available in this case the old clutch to below the slip limit during the Torque transmission can be controlled. The acceleration value can be about to be kept.

In the case of thrust upshifts, the engine torque is always close before the shift a minimum. The motor synchronization takes too long. In this case a type of ASG circuit can be provided. The old clutch is full opened and the new clutch closed when the engine speed is below the Speed of the new shaft is.

In downshifts, the engine torque before the shift is always close to a maximum. The motor synchronization takes too long. In this case a kind of ASG circuit can be used. The old clutch is fully open and the new clutch is closed when the engine speed exceeds the speed the new wave.

In thrust downshifts, the engine torque is always close before the shift a minimum. There is no engine torque to reduce the engine speed available below the speed of the old shaft. In this case, the old one The clutch is kept below the slip limit during torque transmission become. The acceleration value can be maintained approximately.

In Figs. 18 to 25 include driving states 1 to 10 referred to, wherein the driving state 1 "start-up", the driving state 2 "wave active", the running condition 3 "Preparation upshift", the driving state 4 "torque transmission to shaft 2 ', the driving state 5 "end upshift", the driving state 6 "shaft 2 active", the driving state 7 "preparation for downshift", the driving state 8 "torque transmission to shaft 1 ', the driving state 9" end downshift ", the driving state 10" idle "and the driving state 11 mean "other driving states".

FIG. 26 shows a flow diagram of a possible shift strategy for train upshifts, overrun upshifts, downshifts and downshifts.

FIG. 27 also shows a flow chart for the control of the electric motor.

The following basic data can preferably be used for the switching strategies which are shown in FIGS. 18 to 25:
J _eng = 0.2 kg.m ²
J _vehicle = 117 kg.m ²
J _emotor = 0.022 kg.m ² i ₁ = 14 (gear ratio on shaft 1)
i ₂ = 10 (gear ratio on shaft 2)
i _em = 1.2 (gear ratio between electric motor and shaft 2)
T _emotor = maximum torque
T _vehicle = brake friction + uphill or downhill runs + Luftwiders tan d + Rollwiders tan d
r = 0.3 m (wheel radius)
m = 1300 kg (mass of the vehicle and the load)

It should be noted that some numerical effects (peaks) in the Acceleration curves are visible as a result of incorrect dynamic equations, as a result of incorrect dynamic equations that occur during a time step be used in the model, especially if the Clutch slip is changed from sticking to slipping (thrust torque is then calculated as tensile moment). However, as long as the acceleration value before and after the peak is the same, it is expected that there will be no jerk in the vehicle is noticeable. It should also be noted that the acceleration curves are greatly enlarged, so that this is the case with the courses shown with regard to these effects.

As a result, an optimal strategy for comfortable shifting, in particular proposed for an ESG transmission, with and without the support of an electric motor.

Overall, a control strategy for translation changes is therefore preferred presented in an ESG dual clutch transmission. The goal of this strategy is Control of the output torque (vehicle acceleration) to make it comfortable Perform overlap switching. The primary control means can be the electric motor. If the moment of the electric motor is not large enough, can to control the output torque (vehicle acceleration) also one of the Couplings or the like can be used.

This control strategy specifically uses four different types of Overlap shifts defined, namely train upshifts, Thrust upshifts, train downshifts and thrust downshifts. All these Circuit types can be done with minimal jerk and high comfort with a sufficient moment from the electric motor or from the Internal combustion engine is provided. The upshifts and the Train downshifts can be significantly faster and faster when using the electric motor be carried out more comfortably because the internal combustion engine is insufficient Can provide moment, d. H. Upshift at minimum gas or Downshifts at maximum gas.

Furthermore, switching strategies with the electric motor are possible, which reduce the energy input reduce the couplings in order to reduce the heat load and wear to minimize.

A further embodiment of the present invention follows described which another switching strategy especially for Parallel gearbox provides for the waiver of all synchronizations for Upshifts enabled.

Especially for reasons of cost and competition, it is desirable on To dispense with synchronizations in double clutch transmissions without the The quality of the circuit suffers.

Accordingly, a shift strategy is proposed in which the aforementioned Disadvantages are avoided.

It is possible to provide shift strategies in which only certain synchronizations of the transmission input shafts are provided. For this purpose, the transmission input shafts are braked to the synchronous speed of the target gear using the synchronizations at the highest gears. The target gear can then be engaged. However, it has been shown that the drag losses in the transmission during the traversing paths at the synchronizers and to the target gear, especially at low temperatures, can result in the synchronous speed already being undercut. This effect can be advantageously used in the control strategy presented here. The control strategy according to the invention is shown schematically in FIG. 28, the speed of the transmission input shaft 1 and the transmission input shaft 2 being shown over time or over the individual shift phases. With this strategy, the upshifts also manage without synchronization. A 1-2-3 upshift is indicated by way of example in FIG. 28.

Here, gear 1 is engaged on the transmission input shaft 1 (GE 1) and gear 2 is engaged on the transmission input shaft 2 (GE 2), this being possible in the stationary state without synchronization by slightly applying the clutches K 1 and K 2, so that the gear is in gear 1 can be approached. After the transition to gear 2, the gearbox input shaft GE 1 can be synchronized to gear 3. To this end, gear 1 is first taken out and the transmission input shaft GE 1 is switched to the neutral state. The gearbox input shaft GE 1 now runs freely and can, after a while, reach the synchronous speed for gear 3 as a result of the drag losses in the gearbox. However, this is not acceptable for reasons of switching time, which is clear from the illustration in FIG. 29. The speed gradient as a result of the drag torque over temperature is shown there.

Therefore, immediately after removing the gear 1, the clutch K1 briefly applied and the GE 1 gearbox input shaft already considerably up the speed of the transmission input shaft GE 2 is delayed, which in this Moment, i.e. immediately after the crossfade switching, approximately in the ratio of Gear ratio 12/13 is slightly above the synchronous speed of gear 3. Eventually this synchronous speed as a result of the temperature-dependent drag torque reached, which can be recognized by the existing speed signals.

If no gear is to be selected, the speed of the Gearbox input shaft GE 1 fall below the synchronous speed and with the present Switching command can be easily synchronized by applying clutch K1. It is the other approaches in the inventive method are also conceivable Shift strategy can be applied to further improve this shift strategy.

The shift strategy presented here is preferably for parallel shift transmissions can be used, but the shift strategy can also be used for other transmissions become.

A flowchart of the aforementioned shift strategy for parallel shift transmissions is indicated schematically in FIG. 30.

It is also conceivable that in the shift strategy for speed adjustment during upshifts, which utilizes the drag torque in the transmission, preferably when the shaft of the outgoing gear is clutched after the shift has taken place, the old gear is removed and the neutral state N is shifted briefly in order to thus bringing the clutch to the speed of the new shaft. The possible remaining speed difference can preferably take place by utilizing the drag torque in the transmission. Without utilizing the drag torque, this variant of the shift strategy is suitable for advantageously reducing the required synchronization power. This is particularly important in the case of larger gears and the associated larger inertias of the masses to be synchronized. A corresponding switching flow diagram is shown in Fig. 30a.

A further embodiment of the present invention follows described the integration of engine functions into controls for transmissions.

For example, systems can be provided that allow access to the vehicle and enable its commissioning without conventional keys. For example, it can also be provided that a separate control button for Starting the engine is provided. To start the engine especially in automatic transmissions or also in automated transmissions Starter approval will be provided.

The purpose of the present invention is to operate gearboxes and start them or stopping the motor can be combined. It is therefore proposed that the Operation of the transmission and the engine regarding the functions of the starting and the shutdown can be summarized. This also enables simple Strategies for representing the functions associated with parking a vehicle manual gearbox with gear engaged.

It is also possible to add additional functions of the gearbox and motor consider.

For the actuation of automatic or automated transmissions can be preferred one or more controls, such as B. selector lever, paddle shifters, buttons on Steering wheel or the like., Are used. Preferably two types of Operations possible. On the one hand, only impulses are passed on, one certain action, e.g. B. in the transmission, namely for example a gear up or switch down, trigger. On the other hand, functions are switched on, such as z. B. a possible automatic mode.

• - P-Parken; das heißt, erster Gang ist eingelegt und die Kupplung ist geschlossen
• - S-Start/Stop; das heißt, bei getretener Bremse ist die Kupplung geöffnet
• - N-Neutral; das heißt, kein Gang ist eingelegt und die Kupplung ist offen
• - A-Automatischer Modus; das heißt, Gangwechsel entsprechend der Schaltpunkte
• - M-Manueller Modus; das heißt, manuelle Gangwechsel sind möglich
• - R-Rückwärtsgang; das heißt, Rückwärtsgang ist eingelegt.

According to the present invention, the following switchable positions can preferably be provided:

• - P parking; that is, first gear is engaged and the clutch is closed
• - S start / stop; this means that the clutch is open when the brake is depressed
• - N-neutral; that is, no gear is engaged and the clutch is open
• - A-automatic mode; that is, gear changes according to the shift points
• - M-manual mode; that is, manual gear changes are possible
• - R reverse gear; that is, reverse gear is engaged.

Of course, other switchable positions are also conceivable that are included in this concept can be incorporated.

The positions A and M can preferably be combined by a new position D, the change being able to take place via corresponding key signals A / M. In addition, there are the button functions for changing gears or for starting and stopping the engine.
+ = Upshift
- = downshift
on = switch on starter
off = stop engine.

FIG. 31 shows a possible arrangement for a selector lever with a corresponding layout, with typing functions being marked by squares and switching functions being indicated by circles.

A major advantage of such a selector lever is that the previous separation of the position P has created possible functions in two different positions for modified operation. As an alternative to a selector lever, separate switches, buttons or the like can also be used. The switch can enable the switching function of FIG. 31. The button functions for starting and stopping the engine can also be enabled by another button switch. This push button switch can possibly also be integrated in a changeover switch. Furthermore, it is conceivable that, for example, the pushbutton switch for the gear change or the change between automatic and manual shift program is provided by double use.

This possible embodiment of the present invention can be used in all automated manual transmissions and also for purely automatic transmissions Find.

A further embodiment of the present invention follows described in which a parameterization in the case of certain driving situations and / or driving programs is provided for downshifts in transmissions.

Preferably with a parallel gearbox (PSG), for example Dry clutches and self-locking electromotive actuators can occur that the transmission control due to a fault, for example due to failure of a fuse or the like, during an overlap circuit fails and both clutches transmit for a moment. If so, too certain driving situations, e.g. B. on a smooth road or the like. Happens and the Drivers also go off the gas because he notices the failure, for example Wheels slip. By appropriate parameterization of the circuits, especially the train downshifts, can behave like the vehicle a manual transmission or an automatic transmission (ASG), can be achieved. Suitable criteria and parameterizations are required to find. For example, such a parameterization in downshifts with a parallel gearbox in the winter program or at Wheel slip detection can be provided.

According to the invention, it can be proposed as a preventive measure that in the event of an excessive wheel slip probability, preferably only Train downshifts are carried out with full traction interruption.

In Fig. 32, a train downshift is shown schematically in a parallel transmission with traction interruption. In the diagram above the speed is shown over time. The middle diagram shows the moment over time, indicating the torque reduction in the old gear and the torque build-up in the new gear. The lower diagram shows the tensile force over time, the interruption of the tensile force being indicated by the course of the tensile force. In these diagrams, the engine speed is indicated by a solid line, the clutch torque of the new clutch by a dashed line, the clutch torque of the old clutch by a dotted line and the tractive force at the output by a dash-dotted line. It becomes clear that the tractive effort during the downshift of the parallel gearbox is carried out like an automatic gearbox.

As part of a further development, it is also possible that, with regard to the automatic converter, which the parallel shift transmission is to replace fully, for example, a residual filling torque is transmitted to the drive train via the clutch of the outgoing gear, while the engine is brought to the target speed by a motor intervention new gear is accelerated. This has the advantage that the drive train remains tensioned and no disruptive game runs occur. This procedure is indicated schematically in FIG. 33, in which a train downshift is indicated in a parallel transmission with tractive force filling. Also in Fig. 33, the engine speed by a solid line, the course of the new clutch by a dashed line and the old clutch by a dotted line and the curve of tensile force to the driven by a chain line is shown.

It is possible that both clutch actuators can fail during the clutch overlap. Then at least one clutch would grind permanently. If the driver were also to accelerate, the wheels could lose grip on slippery roads and with a slightly higher probability than a vehicle with a manual transmission. This situation is indicated in FIG. 34 in the form of a power flow through the parallel shift transmission in the engine overrun mode during an overlap shift (clutches frozen).

In the context of an advantageous further development, it is proposed that, in the event of an excessive wheel slip probability, only downshifts with complete traction power interruption are carried out, as is also shown in FIG. 35 in a power flow diagram. The power flow diagram represents a parameterization of the train downshift with an increased wheel slip probability.

An increased wheel slip probability can e.g. B. present if that Winter program by pressing a winter button switch or another appropriate switch is activated by the driver, or by the fact that the wheels have previously lost liability once or several times. This can z. B. when spinning during acceleration, through an ASR intervention an ABS intervention or the like.

It is conceivable that when determining wheel slip z. B. a suitable wheel slip probability parameter is set high and then reduced again over time, in particular the longer no wheel slip can be observed. This is indicated in Fig. 36. There the wheel slip probability is shown over time.

It is also possible that a stepless reduction in the level of refill, in particular during a train downshift, is preferably determined as a function of the wheel slip probability parameter. This is indicated in Fig. 37. The degree of filling is shown there via the wheel slip probability.

A further embodiment of the present invention follows described in which an engine drag torque control especially in the event of failure the transmission control or one or both clutch actuators proposed becomes.

Especially with a parallel gearbox (PSG) it can happen that in the Course of an overlap the transmission control or at least one Coupling actuator fails. For example, this failure can result from the destruction a backup.

As a result, both clutches can transmit a moment and this Condition can be frozen if it is self-retaining electromotive actuators. It is important to consider whether it will blocking a transmission or at least an increased drag torque can come on the wheel, so that due to the failure the vehicle can be damaged Could lose lateral stability and break out.

There is a particular risk of breakout if both clutches permanently transmitted moments and the driver of the gas and the engine in the Overrun goes as well as the vehicle itself on smooth, for example icy Underground.

Basically there is in all types of gearbox, in particular also in Manual transmissions, the risk that the vehicle due to the load change breaks out from train to push operation. With a parallel gearbox (PSG) with frozen clutches, i.e. in the rare case of emergency operation, the shifts Limit and thus increases the probability of this failure.

Fig. 38 shows an example of the power flow in such a situation. The drag capacities of the engine and the second clutch add up at the output.

The output torque can be calculated using the following formula, for example:

M _ab = M _mot .i ₁ + M _K2-Schlupf .sign (n _Mot - n ₂ ). (I ₂ - i ₁ ) (1e)

The driver can basically move the wheels by pressing the power pedal (affect the engine torque) again stick to the road.

An object of the present invention is to control the engine in this way influence that the aforementioned situations are avoided.

This can be achieved in particular by the engine control system Influencing the engine drag torque, in particular a detachment from prevented in advance or upon detection of a detachment regardless of the Driver reaction intervenes accordingly.

The engine control can determine whether the transmission control has failed and the two clutches of the parallel gearbox (PSG) permanently torque transfer. Accordingly, it is proposed according to the invention that a special Engine drag torque control in the engine control is active at a Load changes from pull to push operation of the internal combustion engine, which is yes can generally take place on slippery roads, and thus the risk of Prevents the wheels from sticking to the road. If there is a detachment has come, the wheels can again by increasing the engine torque Stick to be brought on the road.

According to a development of the present invention, possibilities for Transmission failure detection proposed. For this, e.g. B. be provided that the transmission control provides emergency operation information, e.g. B. an emergency bit on the CAN, to the engine management system, which is used to activate this special engine Drag torque control can lead.

Another possibility is that the engine control system independently recognizes that the transmission control has failed completely and then activates it independently this special engine drag torque control. An indication of the Transmission failure can e.g. B. the missing update of the CAN messages from Transmission control unit (values do not change, message counter frozen). Other options for identifying the transmission failure are also conceivable.

FIG. 39 schematically shows a flow chart for activating the engine drag torque control in the engine control. The engine drag torque control can also be activated when the transmission control is reset.

Within the scope of a further embodiment of the present invention Possibilities described how the engine control to the aforementioned situation can react.

It is conceivable that the engine control no longer allows the maximum drag torque in overrun mode and under no circumstances enables the internal combustion engine to be shut off in overrun. The engine torque can be when the accelerator pedal is not pressed, e.g. B. can be set to 0 Nm or a small negative value, as is also shown in FIG. 40. This figure shows the special engine drag torque control with low drag torque.

It is also possible that the engine control delays the implementation of the driver's request after overrun and builds up the retarding overrun torque only very slowly. This is shown schematically in FIG. 41, which shows the special engine drag torque control with delayed drag torque build-up.

Furthermore, it can be provided that the engine torque z. B. is increased (reduce drag torque) until the wheels stick again if it defies these measures to a detachment of the driven wheels against the road, which the engine control can determine from the speed comparison of the driven and the non-driven wheels. This is shown in Fig. 42. The special engine drag torque control with low drag torque is shown there, whereby the drag torque is regulated at the trap limit.

In the aforementioned cases, the driver has the option of driving through the vehicle Activate the brake pedal actively decelerating, then the other systems for ensure driving stability, such as B. ABS, ESP or the like.

In Figs. 40 to 42, the engine torque is in each case plotted over time, wherein the solid line in each case shows the torque curve defined by the respective motor-braking control is affected. The dash-dotted line shows the normal engine drag torque. FIG. 42 additionally shows a second diagram in which the rotational speed is shown over time, the solid line representing the rotational speed or the adhesive tear for the driven wheels and the dotted line representing the rotational speed curve of the non-driven wheels.

A further embodiment of the present invention follows described, in which an automatic adaptation of a load precontrol to the Load profile of the clutch actuator is proposed.

As part of an optimization of the clutch position controller for the Coupling actuator implemented a load concept. This load pre-control corresponds to a position-dependent portion for the output of the control. The The course of the load precontrol can be related to the load profile of the clutch be adjusted to get good regulation. There are others too Customization options for load pre-control possible.

Due to different constructions of the different couplings however, there are also different load profiles. Thus can be more advantageous Be provided that, for example, the load pre-control automatically the conditions of the respective coupling used are adjusted. On the main difference between the different couplings is that the distance between the points of engagement or disengagement for the Starting clutch and the powershift clutch are different.

Since the clutch release force is used for the position reference, an additional adaptation can advantageously be dispensed with.

The shift of the working range of the powershift clutch, however, requires one Adaptation in particular because the load differs significantly from the actuator position depends. If the load precontrol is fixed, this can lead to: the feedforward either suggests values that are too small or values that are too large. at If the feedforward control is too low, this leads to sluggish control behavior the construction of the integral part must be maintained in order to overcome the load. If the feedforward control is too large, this leads to the control being overridden, which can put a considerable load on the actuator.

A method is described below to obtain additional information about the To get load history, for example, the increase in the engagement force of the Powershift clutch to be recognized. The position determined in this way can then be integrated into the load precontrol.

In Fig. 43, the engagement or disengagement force for the power-shift clutch and the starting clutch is represented on the actuator position. It is clear from this figure that there is a difference in the distance between the application points of the disengagement or engagement force for the starting clutch and the powershift clutch.

The result is a table with fixed values for the load pre-control (LV), which is shown below:

Another table with a table adapted to the measured reference position of the engagement force of the power shift clutch (TP) is shown below.

With this configuration, the simplest possible adaptation of the Load pre-control is provided on the coupling system, since this is a simple one There is a parallel shift in the characteristics of the powershift clutch. Therefore, it is enough to measure it a position here. For more complicated variations of the load profiles it may also be necessary to carry out more complex measurements with which can also determine the slope of the load. It is also possible for others To provide adjustments to the load pre-control to the proposed here To further improve procedures.

In the method described above, a position comparison for a Coupling actuator, in particular with an incremental displacement measurement, by a "Sniffing function" enables, here preferably brushless motors in the Coupling actuators are provided. It is also possible to use other motors To use clutch actuators.

A method is specified with which, in particular, the Actuator speed can be recognized at which position the release system on the spring, in particular disc spring. Here are besides that Incremental displacement sensor no additional sensors or mechanical parts required. However, this method should preferably only be used for actuators without Compensation spring and used with mechanical release system: be like z. B. in an uninterruptible manual transmission (USG) with EZA.

The properties of a permanently excited direct current motor are considered as theoretical bases, in which a speed is set at a predetermined voltage U, which is linearly related to the braking torque.

Here are:
n: engine speed
n ₀ : idling speed
M _R : The frictional torque acting on the motor
M _A (U): Starting torque (torque when n = 0). The following applies: M _A (U) α U.

Therefore, in a system in which the inertia of the motor has a negligible influence (≙J <M _R ), the load can be deduced from the speed of the motor. A position-dependent load thus leads to a position-dependent speed.

With a linear load curve, this also leads to a linear speed curve:

In a test setup, the measurements were modified with a Coupling actuator (ASG) performed on a special test setup.

• - Keine integrierte Kompensationsfeder
• - Inkrementalwegsensor (aus einem Wählmotor) auf der Ankerwelle des Kupplungsaktors mit 40 Inkrementen pro Umdrehung → ca. 40. inkremente pro mm am Geberzylinder → ca. 12.5 µm/INC an der (theoretischen) Tellerfeder.

The variations of the clutch actuator are:

• - No integrated compensation spring
• - Incremental travel sensor (from a selector motor) on the armature shaft of the clutch actuator with 40 increments per revolution → approx. 40 increments per mm on the master cylinder → approx. 12.5 µm / INC on the (theoretical) disc spring.

The test setup enables different loads (springs, detents, Stops or their combinations) on the clutch actuator.

The structure consists of a base frame made of aluminum profile rails on the the control unit is attached to the clutch actuator. On the aluminum profiles different load modules can be positioned.

The tappet of the clutch actuator is firmly connected to a rod which is slidably mounted over the basic structure. The locking profile and the stop washers (also for the springs) can be freely positioned on this rod and can also be detached. This means that the positions of the different loads can be varied and several loads can be used in combination. The test setup is shown schematically in FIG. 44.

• - (ehemalige) Kompensationsfeder mit 5.9 N/mm.
  Sie dient der Simulation einer kleinen Last, wie sie auch im realen System benötigt wird, um ein sicheres Funktionieren des Ausrücksystems zu gewährleisten.
• - Kupplungs-Feder mit 19.9 N/mm
  Sie simuliert die Kraft der Tellerfeder.

Two spring loads were used for the test setup presented here, whereby both springs were loaded in the same direction.

• - (former) compensation spring with 5.9 N / mm.
  It is used to simulate a small load, as is also required in the real system, in order to ensure the safe functioning of the release system.
• - Coupling spring with 19.9 N / mm
  It simulates the force of the disc spring.

A load profile shown in FIG. 45 is thus to be generated, the force profiles over the position for two springs 1, 2 and the total profile thereof being shown.

This position-dependent load results from the (almost) self-locking Actuator gear a position-dependent (and movement-dependent) Frictional torque on the engine. The moment of friction is opposed to movements the force almost proportional to this force.

In the following, increments are always used for the position information. The Conversion between increments and the true position is based on the Actuator gearbox only approximately linear. The sign also turns around.

The weak spring begins with the measurements initially provided here a position of 0 increments (with increasing force with negative Position) and the strong spring starts at -100 increments. There are others too Conditions possible.

The motor voltages are specified via the pulse width ratio. A value of 255 corresponds to approx. 10 V. The speed can be in arbitrary Units can be specified. It is preferably simply those in the control unit given values.

It has been shown that when using an incremental travel sensor, the Actuator speed is the most accurate signal for the state of the actuator. Also An externally attached measurement of the motor current is significantly less precise.

First, 46, the result of a series of measurements is shown in Fig., In which under the above-shown load of the actuator with different voltages (PWM ratios ± 30, ± 35. ± 40, ± 45, ± 50, ± 55 and ± 60) is driven in both directions.

• - Die Auswirkung der Last auf die Geschwindigkeit des Aktors zeigt sich weitaus besser bei Bewegungen entgegen der Last (hier bei negativen Geschwindigkeiten) als mit der Last. Dieses Verhalten lässt sich aus den Eigenschaften des selbsthemmenden Getriebes erklären.
• - Es zeigt sich eine deutliche Störung mit einer Periode von 40 Inkrementen = 1 Motorumdrehung.
• - Beim Laufen auf die Last erkennt man (unter der periodischen Störung) eine konstante Geschwindigkeitsänderung über dem Ort (Δν/Δx) unabhängig von der Größe der Geschwindigkeit.

Basic properties are already evident in this measurement:

• - The effect of the load on the speed of the actuator is shown much better with movements against the load (here with negative speeds) than with the load. This behavior can be explained from the properties of the self-locking gear.
• - There is a clear fault with a period of 40 increments = 1 motor revolution.
• - When running on the load you can see (under the periodic disturbance) a constant change in speed over the location (Δν / Δx) regardless of the size of the speed.

The reproducibility of the measurements is very high.

In the measurement shown in FIG. 47, the actuator is moved back and forth eight times in both directions with the same voltage (PWM ratio = ± 35). The measurement curves agree very well.

Even the small speed fluctuations with a period of 3 to 4 Increments are reproducible. Presumably these are magnetic To lead anchor locking back.

When evaluating the speed measurements, the Change in speed per revolution can be calculated.

The first consideration is to calculate the difference in speed for exactly 40 increments in advance for each position.

B (x) = ν (x) - ν (x + 40) (for movements with negative speed.

Because due to the temporal sampling of the signals, the Speed is measured, should in part the speed curve be interpolated.

If the aforementioned measurement data are evaluated in this way, FIG. 48 results, the speed changes per revolution being indicated there.

In FIG. 48, the movement from right to left. The signal is recognizable by the noise but so bad that it cannot be used for positional adjustment in this form. Averaging reduces noise, but at the same time worsens position resolution.

The calculation of the kink position by linear interpolation of the Speeds can be provided.

If you look at the speed during a measurement always at the same Phase position of the motor revolution (Pos mod 40 = const.), One of the results Modulation signal corrected by motor rotation. Depending on the chosen one Phase position shows a slightly different speed curve. Apart from some noise, the speed profiles differ however only through a speed offset.

This is shown in FIG. 49 for two phase positions. Accordingly, different speed profiles with different phase positions are shown.

A next step for the evaluation can be to recognize from which engine revolution you are on the hard spring. This is possible by simply exceeding the limit value (| ν _i - ν _i-1 |> C). The limit value C depends on the properties of the motor and the hardness of the spring. This is in Fig. 50 with V (Pos); U = -45 shown, whereby a rough position reference results from the speed changes.

This allows the kink position of the load characteristic to be found with an accuracy of ± 40 increments become.

A better position can be obtained by looking at the last one Speed measurements used as support points for straight lines and their Intersection calculated.

The measurements of ν _i and ν _i-1 limit the first interval on the hard spring (the position interval found above). The measurements of ν _i-2 and ν _i-3 limit the last interval before the hard spring. A determination of the exact reference position is shown in FIG. 51.

The position determined in this way is still not very great for each individual phase position I agree. (In the best case, 20 increments), but this procedure is possible to be used for several phases of a series of measurements. An easy one However, averaging all possible positions that can be determined in this way is not sufficient to significantly improve accuracy.

Errors caused by noise can be partially recognized and suppressed. So it can happen that the determined intersection of the two straight lines outside the interval [x _i ; x _i-3 ]. Such a case can be easily recognized and the result for averaging can be suppressed.

It has also been shown that the speed determination for phase positions sit on the flanks of the periodic disturbance due to the interpolation, is subject to very large fluctuations. For this reason, therefore, is about to At the beginning of the measurement, be careful not to consider these phase positions. This is done by evaluating the acceleration (Δν / Δt). Phases that are in the limit the first engine revolution to excessive acceleration, do not turn considered further.

The averaging from the selected values for the The reference point determination then reaches the accuracy of approx. 10 to 12 Increments (or less than 5.5 increments standard deviation).

The quality of the process can be tested by measurements with different starting positions for the hard spring. The specified positions are only approximate because they were set by hand. The reproducibility of the recognition of the same position is important here. Such a check of the determined positions of the reference point is shown in FIG. 52.

Except for motor voltages that are too small (<40) or too large (> 55) Position at which the kink of the force characteristic is found, regardless of the Motor voltage. The scatter of the determined positions lies in one Range from ± 3 to ± 5.5 increments.

The individual measurements lasted between 3 s (for U = -35) and 1.6 s (U = -55). If doing the measurements on the necessary range of motion are limited, times between 1.6 s (U = -35) and 0.6 s (U = -55) are obtained. However, these can possibly be reduced even further.

It can be assumed that the hardness of the spring whose starting position is sought goes into the accuracy of the position determination. A harder spring can a more precise determination of the reference position is made possible.

The following algorithm can preferably be used in the above-described methods, a set of data being required for each phase position in which the reference position is sought.

The process is shown in an overview below.

The start of the measurement and the determination of the phase positions under consideration is shown in FIG. 53. A flowchart for determining the reference points for the individual phase positions is shown in FIG. 54.

A further embodiment of the invention is described below, in which the Clutch opening should be accelerated when changing gears.

So far, there has been the problem that gear changes in an automated Manual transmission (ASG) should be carried out faster. Part of it bears that Dynamic behavior of the clutch actuator as the gear is only removed can be when the clutch is open. The ASG changes gear such that the clutch torque is either ramped or down to 0 Nm is parabolically degraded. The clutch is only fully opened in clutch state 4 HUB open. During the time when the clutch can only transmit a few Nm the aisle can be pulled out. There are other strategies for that Clutch torque reduction, but they all aim to do just that Reduce clutch torque to 0 Nm and then move to HUB in state 4.

To open the clutch faster and take the gear out earlier as well According to further training, being able to synchronize the new gear earlier of the invention that shortly before the aisle is removed, the clutch position control is switched to a controller and the Adjustment motor in the clutch actuator with max. Voltage is operated. The Position control of the clutch actuator can operate such that at a Control deviation of approx. <4 mm is started to regulate so that it is as possible there are no overshoots. This leads to the actuator movement starting at approx. 4-5 Nm slows down. By creating the max. The voltage on the electric motor is the Max. Actuator movement reached. From the point at which the clutch is then safe or is opened wide enough (e.g. from the separating point), can Voltage control can be switched to the position control so that the way up to the HUB position (in state 4, neutral gear, the clutch is up to HUB opened) is reached. The switch to position control must be timely again done so that the actuator does not move against the stop. Syncing the Next gear only begins when the clutch position reaches a certain value has exceeded so that the clutch is safely opened. When opening the clutch is faster, you can synchronize the next gear and therefore the Reduce switching time. It makes sense especially when using Active Interlock possibly even necessary to open the clutch faster with a different strategy, since in Active Interlock the alley was still approached with the clutch closed can be and the switching is carried out much faster.

According to another development of the invention, the same effect as in the aforementioned training can be achieved if instead of a Switch voltage control, the control deviation increases. You could So instead of driving on HUB, first set the target path even higher so that the Position control due to the large control deviation the max. Tension on the Motor engages. Is the clutch then opened wide enough, namely when the If the standby point is reached, the target path is set to HUB again so that the Actuator cannot drive against the stop and as usual with neutral gear HUB can be adjusted.

A further improvement can be achieved within the scope of a further configuration if you prefer additional pilot control for the clutch actuator starts. You know the timing behavior of the clutch actuator very well. As a result, the behavior can be very good. B. with a PT2 link or the like. be replicated. With this knowledge, the coupling can do it by default open if you want to transmit the moment. Through the Actuator deceleration, the actual actual torque on the clutch is only after approx. 30 ms changed. The power supply takes so long until it becomes noticeable emotional. If this time is pre-controlled, the Clutch actuator can be realized on a desired target position request.

These proposed possibilities are explained together in FIG. 55. These strategies can be used particularly in vehicles with automated manual transmissions, electronic clutch management, uninterrupted manual transmissions and with parallel manual transmissions.

A further embodiment of the invention is described below, in which in particular adaptation strategies and zero point adjustments of the distance measurement Correction of the transmission behavior of clutches in parallel gearboxes (PSG) can be proposed.

It has been shown that an adaptation of electromechanically operated Clutches preferably in a parallel gearbox (PSG) is required.

Accordingly, strategies for adapting the clutch characteristics in Parallel gearboxes proposed.

• - Kupplungsaktor, beinhaltet eine von der Steuerungssoftware geregelte Antriebseinheit (z. B. Elektromotor + Verschiebeeinheit)
• - Kupplungsausrücker, der die Ausrückkraft auf die Tellerfeder der Kupplung überträgt.
• - Übertragungssystem, die Verbindungsstrecke zwischen Kupplungsakfor und Kupplungsausrücker (z. B. Hebelmechanismus, mechanische Welle, Hydraulikleitung etc.).

The electromechanical clutch actuator, especially in automated manual transmissions, can preferably be divided into approximately three components:

• - Clutch actuator, contains a drive unit controlled by the control software (e.g. electric motor + displacement unit)
• - Clutch release, which transfers the disengagement force to the clutch disc spring.
• - Transmission system, the link between the clutch adapter and clutch release (e.g. lever mechanism, mechanical shaft, hydraulic line, etc.).

The clutch can be actuated by moving to a specific position Drive unit in the clutch actuator. Through the transmission system Clutch release actuated and the clutch either closed, opened or set a certain moment. The position of the drive unit can be preferred be monitored with absolute displacement sensors. Conversion to a transferred one Coupling torque is in the control software using the clutch characteristic calculated in which the transfer function "torque via master cylinder travel" is deposited.

According to the present invention, a clutch adaptation and a Zero point adjustment of the clutch characteristic curve provided. Because the real one Clutch actuation is very far from the distance measurement in the clutch actuator, in addition to the clutch properties, the influences of the release system by adapting or adjusting the zero point of the distance measurement become. So z. B. by temperature changes the transmission behavior of a lever system or the volume of a filled with hydraulic fluid Line can be changed. Due to these changes must be made regularly Intervals a zero point adjustment of the distance measurement can be carried out. This can be achieved by moving to a certain, excellent position of the clutch be carried out, which can be recognized as such by the clutch actuator. This point could e.g. B. be the pressure point of the plate spring at which the Release force increases suddenly. This zero point adjustment can be done while driving with clutch closed and gear engaged or while standing at gear can be carried out (clutch is briefly closed).

The clutch characteristic can also be adjusted using the touch point adaptation become. The clutch actuator position at which the clutch engages is determined minimal moment begins to transmit. This adaptation can be done while standing Vehicle (foot brake or hand brake is applied) with gear engaged and / or Engine run at idle. The clutch is slowly closed until a minimal moment is transmitted. The idle controller of the engine control reacts to the clutch closing by increasing the engine torque exactly the amount of the transmitted clutch torque (3-4 Nm), so that the Idle speed remains constant. Based on the reaction of the engine torque the transmitted torque of the clutch, and thus the position of the touch point be determined and adapted. Other adaptations are also possible.

For the use of electromechanically operated clutch actuators in Parallel shift transmissions (PSG), these strategies can be redefined.

Furthermore, coupling adaptations and a zero point adjustment in Parallel transmission considered.

In contrast to clutch systems, control in dual clutch transmissions two independent clutch actuators the transmitted moments of the clutches A and B. Usually a clutch with the sub-transmission is the odd gears (1, 3, 5) the other clutch is connected to the partial transmission of the even gears (2, 4, possibly R). When driving, the active clutch A / B is closed and transmits this Engine torque via an engaged gear of the partial transmission to the output. The other, passive clutch B / A is either open when the gear is engaged or closed with gear in gear (gear N).

• 1. Bei ausgelegtem Gang (Gang N) im Teilgetriebe der passiven Kupplung B/A folgt Durchführen des Nullpunktsabgleichs von Kupplung A + B
• 2. Bei eingelegtem Gang im Teilgetriebe der passiven Kupplung B/A folgt Auslegen des Ganges (Gang N), Durchführen des Nullpunktsabgleichs von Kupplung A + B und Wiedereinlegen des gleichen Ganges.

The following strategies are preferably conceivable for carrying out the zero point adjustment of the distance measurement in double clutch transmissions:
When the vehicle is running (active clutch A / B transmits engine torque) the following can be provided:

• 1. When the gear is designed (gear N) in the partial transmission of the passive clutch B / A, the zero point adjustment of clutch A + B follows
• 2. When a gear is engaged in the partial transmission of the passive clutch B / A, the gear is disengaged (gear N), the clutch A + B is zeroed and the same gear is re-engaged.

• 1. Bei ausgelegten Gängen (Gang N) in den Teilgetrieben der Kupplung A + B folgt Durchführen des Nullpunktsabgleichs der Kupplung A + B
• 2. Bei ausgelegtem Gang im Teilgetriebe der Kupplung A/B und eingelegtem Gang im Teilgetriebe B/A, folgt Durchführen des Nullpunktsabgleichs der Kupplung A/B und Auslegen des Ganges der Kupplung B/A, Durchführung des Nullpunktsabgleichs sowie Wiedereinlegen des gleichen Ganges.
• 3. Bei ausgelegtem Gang im Teilgetriebe der Kupplung A/B und eingelegtem Gang im Teilgetriebe B/A folgt Auslegen des Ganges der Kupplung B/A, Durchführen des Nullpunktsabgleichs der Kupplung A + B und Wiedereinlegen des gleichen Ganges
• 4. Bei eingelegten Gängen in den Teilgetrieben der Kupplung A + B folgt Auslegen des Ganges der Kupplung A/B des Nullpunktsabgleichs der Kupplung A/B und Wiedereinlegen des gleichen Ganges sowie Auslegen des Ganges der Kupplung B/A und Durchführen des Nullpunktsabgleichs der Kupplung B/A sowie Wiedereinlegen des gleichen Ganges
• 5. Bei eingelegten Gängen in den Teilgetrieben der Kupplung A + B folgt Auslegen der Gänge der Kupplung A + B, Durchführen des Nullpunktsabgleichs der Kupplung A + B und Wiedereinlegen der gleichen Gänge.

When the vehicle is stationary, the following can be provided:

• 1. With designed gears (gear N) in the partial transmissions of clutch A + B, the zero point adjustment of clutch A + B follows
• 2. With the gear in the sub-transmission of clutch A / B in gear and the gear in sub-transmission B / A engaged, carry out the zero point adjustment of clutch A / B and disengage the gear of clutch B / A, carry out the zero point adjustment and re-engage the same gear.
• 3. With the gear in the sub-transmission of clutch A / B in gear and the gear in sub-transmission B / A engaged, disengage the gear of clutch B / A, carry out the zero point adjustment of clutch A + B and re-engage the same gear
• 4. With gears engaged in the sub-transmissions of clutch A + B, disengaging the gear of clutch A / B follows the zero point adjustment of clutch A / B and re-engaging the same gear as well as disengaging the gear of clutch B / A and performing the zero point adjustment of clutch B / A and re-engaging the same gear
• 5. With gears engaged in the sub-transmissions of clutch A + B, disengage the gears of clutch A + B, carry out the zero point adjustment of clutch A + B and reinsert the same gears.

The zero point adjustment can preferably be carried out with valid ones Entry conditions are repeated at certain time intervals. It can the order of the clutches A / B of the current or future Driving situation to be adjusted. The following strategies can be used for this to be proposed:

• 1. Bei einer Fahrt soll immer die Kupplung zuerst adaptiert werden, deren letzter erfolgreicher Nullpunktsabgleich am längsten zurück liegt.

If the vehicle is running (active clutch A / B transmits engine torque), the following can be provided:

• 1. When driving, the clutch should always be adapted first, the last successful zero point adjustment being the longest ago.

• 1. Bei stehendem Fahrzeug wird immer bei der Kupplung des wahrscheinlichen Ganges der Anfahrt zuerst der Nullpunktsabgleich durchgeführt.
  Hintergrund dieser Strategie ist das Faktum, dass das Moment dieser Kupplung den Komfort der nachfolgenden Anfahrt bestimmt.
• 2. Es werden immer bei beiden Kupplungen der Nullpunktsabgleich gleichzeitig durchgeführt.
  Hintergrund dieser Strategie ist das Faktum, dass das Auslegen der Gänge im Stand immer mit störenden Geräuschen verbunden ist. Diese werden minimiert, wenn beide Gänge gleichzeitig ausgelegt werden und der Nullpunktsabgleich anschließend durchgeführt wird.

When the vehicle is stationary, the following can be provided:

• 1. When the vehicle is at a standstill, the zero point adjustment is always carried out when the likely gear is approached.
  The background to this strategy is the fact that the moment of this clutch determines the comfort of the following approach.
• 2. The zero point adjustment is always carried out simultaneously for both couplings.
  The background to this strategy is the fact that laying out the aisles in the stand is always associated with disturbing noises. These are minimized if both gears are designed at the same time and the zero point adjustment is then carried out.

• 1. Wenn in beiden Teilgetrieben ein Gang eingelegt ist, kann die Tastpunktadaption nacheinander durchgeführt werden.
• 2. Wenn im Teilgetriebe der Kupplung A/B ein Gang, und im Teilgetriebe B/A kein Gang eingelegt ist, soll der Gang im Teilgetriebe B/A eingelegt werden und eine Tastpunktadaption für Kupplung A + B durchgeführt werden.
• 3. Bei einer Tastpunktadaption soll immer die Kupplung zuerst adaptiert werden, deren letzte erfolgreiche Tastpunktadaption am längsten zurück liegt (vorgenannte Strategie unter Pkt. 1).
• 4. Die Tastpunktadaption wird immer bei der Kupplung des wahrscheinlichen Ganges der Anfahrt zuerst durchgeführt (vorgenannte Strategie unter Pkt. 2).

The following strategies are preferably conceivable for the touch point adaptation, which can only be carried out when the vehicle is stationary:

• 1. If a gear is engaged in both sub-transmissions, the touch point adaptation can be carried out one after the other.
• 2. If a gear is engaged in the sub-transmission of clutch A / B and no gear is selected in sub-transmission B / A, the gear should be engaged in sub-transmission B / A and a touch point adaptation for clutch A + B should be carried out.
• 3. In the case of a tactile point adaptation, the coupling should always be adapted first, the last successful tactile point adaptation was the longest ago (the aforementioned strategy under point 1).
• 4. The tactile point adaptation is always carried out first when the probable gear of the approach is coupled (aforementioned strategy under point 2).

A flow chart for a touch point adaptation is given by way of example in FIG. 56.

The aforementioned measures can also be suitably combined with one another become.

The method described in the aforementioned embodiment of the invention can in every transmission, especially in automated manual transmissions (ASG), e.g. B. Parallel gearbox (PSG), or with an electronic Coupling management can be used.

A further embodiment of the present invention follows described, in which a gear selection strategy in particular with a Parallel gearbox, such as. B. a dual clutch transmission is proposed to the occurrence of switching noises or gear noises, such as. B. that Gear rattle, by engaging the non-driven clutch with a predetermined torque of a selected gear on the input shaft prevent.

It has been shown that gear rattling is preferred at engine speeds should be prevented below the speeds in typical driving cycles. The Gear rattle is generated by the impulses of the engine, which the Reach the natural frequency of the gear.

Accordingly, the goal of the proposed strategy is the so-called NVH Improve characteristics of a vehicle by the presence of Transmission noise or switching noise in normal driving conditions is eliminated. This can be done by increasing the effective inertia of the Transmission can be achieved so that the transmission noise excitation frequencies are below the normal driving ranges of the engine. It is appropriate that this by engaging the non-driven input shaft coupling (IPSI) takes place while it is in the neutral state, which causes the inertia of the second input shaft is added to the total inertia of the transmission. This in turn reduces the resonance frequency of the transmission.

• - höhere Geräuschreduzierung als mit IPSI Kupplung im Neutralzustand und ausgekuppelten Zustand
• - Hydrauliknachfüllung kann simultan mit der Geräuschreduzierung erfolgen
• - Getriebe wird weniger betätigt als bei eingekuppelter IPSI Welle und offener Kupplung.

The advantages of engaging the IPSI coupling with the shaft in neutral can be summarized as follows:

• - higher noise reduction than with IPSI coupling in neutral and disengaged condition
• - Hydraulic refilling can take place simultaneously with the noise reduction
• - Gearbox is actuated less than with an engaged IPSI shaft and an open clutch.

The procedure for this strategy is very simple. The engine speed is read in, in particular by using a signal that is already used in the transmission control software and by comparing the calculated natural frequency for the selected active gear. If the engine speed falls below the limit of the excitation frequency, the IPSI clutch is engaged during the neutral state. As soon as the engine speed crosses the resonance range, the transmission can return to the previous state, in which the noise is suppressed. This proposed strategy is shown in a flow chart shown in FIG. 57.

The gear rattling occurs when impulses are triggered Internal combustion engine excites the natural frequency of a transmission. This will Causes vibrations in the transmission, which have a high amplitude and are outside the phases of the engine and the rest of the vehicle. By These vibrations can generate noise from vehicle occupants Vehicle are perceived and are therefore disruptive. Consequently ways must be found to prevent these vibrations without doing so other aspects of the vehicle, such as B. performance and fuel consumption, to influence. This can be achieved through the strategy described above.

The equation for the natural frequency of an object can be expressed by the following equation:

From this equation it can be seen that there are two ways to change the Natural frequency of an object there. On the one hand, the properties of the Spring constant can be changed or on the other hand, the inertia of the Systems are changed. The system that is a Vehicle powertrain, the following parameters can have the greatest impact on the natural frequency of the gearbox: inertia of the input shaft, Mass inertia of the output shaft, gear ratio, Damping spring constant of the clutch, spring constant of the input shaft, Spring constant of the axle and inertia of the vehicle. It is also possible, that other suitable parameters are changed to the natural frequency to influence the transmission.

When using these parameters, the natural frequency of the Vehicle, so that the engine speed, which is required to the System in an area that is small enough so that a System excitation is not achieved under normal driving conditions.

Unfortunately, determining the performance of the vehicle still depends on other aspects, e.g. B. fuel consumption and acceleration performance, so that the parameters used to prevent noise are thereby be limited. As a result, in known strategies Use of clutch damping springs provided to keep the spring constant of the system using a dual mass flywheel which increases the inertia of the system. However, with one Dual clutch transmission increases the possibility of implementing a number of software-controlled strategies to actively avoid these unwanted sounds suppress.

The differences between a parallel gearbox or one Double clutch transmission and a conventional manual transmission are in the essential in that in the parallel transmission two clutches and two Input shafts are provided for the transmission or for the transmission. Under normal driving conditions, one of the input shafts of the vehicle is driven (IPSA), while the other shaft (IPSI) is basically disengaged, which is selected under driving conditions. For example, one of the Input shafts the gears 1, 3, 5 and reverse gear and the other shaft the Realize gears 2 and 4.

These clutches and gear selection are controlled by two control modules driven. As a result, the engaging and disengaging of the IPSI as well the choice of gear on the non-driven shaft for the driver is not seen. This makes it possible that the inertia of the transmission by Coupling and uncoupling of the coupling on the IPSI shaft set appropriately can be.

In principle, the following equation can be used to calculate the inertia of a conventional manual transmission:

However, by using the coupling on the IPSI shaft to couple this shaft in neutral, the following equation results:

As a result, the inertia of the drive can be increased, whereby the Resonance frequency of the drive is reduced and thus advantageously is outside the normal driving range of the vehicle.

A simple three-mass model can be developed to test the effectiveness of this strategy. The model can be designed as shown in FIG. 58.

The connection between the engine and the transmission is made by the Clutch damper and through the input shaft as well as the connection between the transmission and the vehicle represented by the wheel half-waves.

For example, the following values can be used for the model:
J _motor = 0.1743 kg.m ²
J _ipsi = 0.02748 kg.m ²
J _ipsa = 0.02748 kg.m ²
J _ops = 0.008394 kg.m ² (calculation also for the differential)
J _vehicle = 106.58 kg.m ²
K _{damper clutch 1 and 2} = 7.08 Nm / °
K _axis = 117 Nm / °

The following gear ratios can preferably be used when calculating the inertia of the gear:
1 .: 3,727
2 .: 1.96
3 .: 1.32
4 .: 0.945
5 .: 0.756

When calculating the inertia of the gearbox, the aforementioned Equations are preferred.

The analysis of the model shows a comparison between the strategy with the neutral IPSI wave engaged and the known strategy with the IPSI wave disengaged. It was also considered when the IPSI shaft was disengaged and engaged with the lowest gear of the shaft. The results are shown in Fig. 59, the engine speeds relating to a 4-cylinder engine with two pulses generated per cycle.

A comparison of switching strategies is shown there. It can be seen from this that the Engaging the IPSI clutch with no engaged gear better than that normal strategy is where the clutch is disengaged and the IPSI shaft with no gear is engaged. The strategy in which the IPSI wave with one gear is very effective in preventing the transmission from occurring Resonance frequency. However, this is not the case with the first course, however wherever the IPSA translation (translation of the activated wave) is higher than the IPSI translation (translation of the non-activated wave). This means, that the IPSI inertia is reduced when both translations go through become.

An embodiment of the present invention is described below in another gear selection strategy, especially in a parallel transmission, such as B. a dual clutch transmission, is proposed to prevent the occurrence of Switching noises or gear noises, such as. B. the gear rattle engaging the non-driven clutch with a predetermined one Prevent torque of a selected gear on the input shaft.

It has been shown that gear rattling is preferred at engine speeds should be prevented below the speeds in typical driving cycles. The Gear rattle is generated by the impulses of the engine, which the Reach the natural frequency of the gear.

Accordingly, the goal of this strategy is the so-called NVH characteristics of a To improve the vehicle by the presence of gear noises or Switching noise is eliminated in normal driving conditions. This can be done by Increase the effective inertia of the gearbox can be achieved, so that Transmission noise excitation frequencies below the normal driving ranges of the Motors lie. In this variant, this cannot be done by engaging the driven input shaft coupling (IPSI) with a predetermined Torque are made possible while the aforementioned shaft with a gear in Intervention is. This is optimal for preventing gear noise, since Mass inertia of the second input shaft to the first driven shaft in addition, so that the total mass inertia of the transmission increases significantly becomes. This in turn reduces the natural frequency of the transmission.

• - verbesserte Getriebegeräusch-Reduzierung im 1. Gang verglichen mit anderen Strategien
• - Möglichkeit zur simultanen Durchführung der Tastpunktadaption
• - Möglichkeit simultan einen niedrigen Kupplungskraftverlauf zu erzeugen.

The advantages of engaging the clutch with the non-driven shaft (IPSI), which engages with one gear, can be summarized as follows:

• - Improved gear noise reduction in 1st gear compared to other strategies
• - Possibility to perform the touch point adaptation simultaneously
• - Possibility to simultaneously generate a low clutch force curve.

The procedure for this alternative strategy is very simple. The engine speed is read in, in particular by using a signal that is already used in the transmission control software and by comparing the calculated natural frequency for the selected active gear. When the engine speed falls below the excitation frequency limit, the IPSI is engaged in one gear, optimally increasing the inertia of the transmission, engaging the IPSI clutch at a predetermined torque which is dependent on the desired level of noise prevention , The optimal gear can be determined from a table in the software. When the engine speed has passed the resonance range, the transmission can return to the state it was in before the noise cancellation. This strategy is shown in a flow chart in FIG. 60.

Equations 1g and 2g used in the aforementioned shift selection strategy can be used in this shift selection strategy. Likewise the Conditions under which the noise excitation frequencies influence each other can also be used here.

However, with this variant, the IPSI clutch is engaged and the shaft is engaged with one gear, so that the following equation results:

By comparing the two equations 2g and 4g it can be seen that the Engaging the IPSI coupling with the shaft which is in engagement with a gear, optimized gear rattle suppression, since it is possible to Increase the inertia of the transmission and thereby the resonance frequency of the Reduce transmission.

With this gear selection strategy, too, the three-mass model shown in FIG. 58 can be used to test the effectiveness of the clutch engagement on the IPSI shaft engaged with a gear. This model has already been described in the previously described embodiment, so that a further description is not necessary.

The analysis of the model shows a comparison between the known strategy with disengaged IPSI shaft and the strategy with disengaged IPSI clutch while the shaft is engaged with a gear, the IPSI clutch with about 4 and 10 Nm with the Shaft is engaged. In all examples in which the IPSI shaft is engaged with one gear, the highest possible gear inertia was chosen. The results are shown in Fig. 59, the engine speeds relating to a 4-cylinder engine with two pulses generated per cycle.

Fig. 61 shows a comparison of the switching strategies. From this figure it can be seen that there is a significant advantage of engaging the IPSI clutch with an engaging gear in preventing gear rattling at first gear where both conventional transmission strategies are surpassed in which the IPSI clutch is in neutral from that with no one Gear engaged shaft is disengaged. However, the proposed strategy is not quite as effective for gears 2 to 5. There, by disengaging the IPSI clutch, a slight improvement in noise suppression can be made possible.

In the present strategy, both couplings of the two shafts can also be used are each engaged with a gear, especially around the natural frequencies of the To influence transmission.

A further embodiment of the present invention follows described in which switching strategies are specified, in particular high Avoid clutch temperatures, with these strategies preferred Parallel gearboxes (PSG) or double clutch gearboxes are used, to cool the components of the couplings.

It has been shown that by reducing the heat development on a Coupling and thermal on the associated coupling system Damage to the clutch, especially the friction surfaces, avoided become.

During torque transmission in a dry clutch system, a certain amount of energy from the clutch in the form of thermal energy added. This thermal energy can cause high coupling temperatures cause, which limits the functionality of the components or one causes permanent damage to the clutch; especially when exceeded the maximum temperatures on the friction surface of 350 to 380 ° Celsius. Strategies for automated manual transmissions (ASG) are currently used to respond to these situations at high temperatures. at In these strategies, a fan is used for the clutch. This will make one Convection is generated so that a hot clutch disc and the clutch itself can be cooled down. With a dual clutch transmission (DSG) several options available to respond to these high temperature situations and implement an appropriate cooling strategy.

In Fig. 62, a double clutch system is illustrated schematically, comprising two clutch discs, which each have their own drive input shaft. In this known transmission, the first input shaft can implement gears 1, 3 and 5. The second input shaft can implement gears 2, 4 and 6. In normal driving conditions, one of the clutches transmits the torque to be transmitted (active clutch) while the other clutch is not activated. As a result of this arrangement, additional strategies are required to cool the couplings in high temperature situations.

According to the present invention, temperature-dependent strategies, developed especially for parallel gearboxes (PSG).

A first possible strategy can provide that on the active clutch, which transmits the torque, opposite input shaft minimal gear is engaged and the inactive clutch is opened so that the open clutch can rotate as long as it is not activated, which causes this can cool down.

This strategy can be used preferentially if the vehicle is with moves at a constant speed and z. B. the clutch 1 is engaged, since then u. U. the temperature at the second clutch 2 may be too high.

A second strategy can provide for gears on only one input shaft be switched to avoid that the input shaft with the hot Coupling comes into contact. This eliminates additional heat, which on the hot clutch.

This strategy may be preferred during upshifts or downshifts and are used if the coupling temperature is too high.

A third strategy is also proposed, in which the first gear or the second gear used depending on clutch temperatures that are too high becomes. This strategy can be particularly useful when starting vehicles and when high coupling temperatures are used.

A fourth strategy is also envisaged, which involves changing the Provides switching strategies for the clutches. This can aim to be quick To provide torque ramps for the two clutches to increase the switching time to reduce. In addition, the engine torque can be faster can be controlled. As a result, the driving comfort is reduced, but the thermal energy at the couplings reduced. This strategy can be used by anyone Torque during the driving cycle.

A fifth strategy is also proposed, in which instead of the switching strategy a parallel transmission with two clutches that can be activated in parallel Strategy of an automated manual transmission (ASG) with torque interruption is used during the circuits. The vehicle can be moved as if it has only one clutch. This strategy can be used at any moment during of the driving cycle lake.

Furthermore, a sixth strategy is provided, which it will use during the Starting allows torque to be transmitted in parallel through both clutches. This divides the thermal load between the two clutches, which reduces the thermal energy. The time available in the kept the vehicle in a steady state with respect to the gradient can be increased in an advantageous manner. This strategy can in particular used for multiple starts of the vehicle on an incline become. Furthermore, this strategy can also be provided if the driver uses the Vehicle position due to slip on the clutch and through the applied Moment by pressing the accelerator pedal at a value.

A further embodiment of the present invention follows described which support synchronization during the area code provides a clutch, especially in a parallel transmission (PSG).

The proposed strategy is that during a gear change preferably in a lower gear on the non-activated input shaft (IPS) a parallel gearbox, the clutch of this non-activated input shaft is used to reduce the differential speed for synchronization.

When a parallel gearbox is in a constant state gear the engaged gear can be in the non-activated input shaft (IPS) be switched. This additional gear change can be done with an automated Manual transmission (ASG) can be compared. The additional gear changes can compared to an automated manual transmission, an increased wear of the Effect synchronization.

This wear can be reduced in an advantageous manner by using the clutch of the non-activated input shaft IPS if the engine speed lies between the speed before and after the preselection of the gear change. For this purpose, the clutch is closed during the neutral phase in the preselection. The engine speed increases with the speed of the non-activated input shaft. It should be assumed that a parallel gearbox with two IPS shafts and two clutches is provided, with one clutch for each IPS shaft. FIG. 63 is a flow chart shown this strategy invention.

A simulation without the use of the clutch for the synchronization of the preselected gear is shown in FIG. 64. With this synchronization, the synchronous speed is approx. 1250 U / m.

A simulation using the clutch for the synchronization of the preselected gear is shown in FIG. 65. With this synchronization, the synchronous speed is approximately 500 U / m, whereby the synchronous speed is reduced by more than half compared to the aforementioned synchronization.

In the two simulations shown in FIGS . 64 and 65, NEng_y are the engine speed, NIps_i1y / i2y speed of the IPS shaft 1/2, TrqClAct_i1m the current clutch torque of the IPS1 shaft, cg_so_drvdelay the target gear for driving, gc_bo_act_i1m / i2m the current gear the IPS wave 1/2.

By reducing wear during synchronization during the Preselection for upshifts can estimate the speed of the IPS shaft be provided during the neutral phase. The speed control can then be modulated accordingly. The model is dependent on Factors such as B. the friction (i.e. influence of temperature) or the age of the Transmission. When using a speed sensor on the input shaft these factors can be taken into account appropriately. It is also still possible to provide other modifications to the strategy presented to the to further improve the inventive method.

A further embodiment of the present invention follows described which one strategy for choosing the gear on one is not activated input shaft for dual clutch transmissions in vehicles.

Dual clutch transmissions have two engaged gears at the same time provided, which are each in engagement with an input shaft. however becomes only one shaft for torque transmission in the drive train used (except during torque transfer when switching the Transitions). It is selected which gears on the non-activated Input shaft are engaged. This can vary depending on Parameters such as B. the vehicle speed, the accelerator pedal position, the Engine speed or the like can be provided. There is also a possibility that other parameters are taken into account.

The following are requirements that the Limit gear selection functionality:

Especially in a normal driving situation, a higher gear should always be used be selected. In downshifts, the time to increase the Engine speed to a suitable speed longer than the switching time at the Transmission.

Furthermore, it should be considered that there are no problems with kick-down the increasing time for the engine speed because it is always longer than the switching time. Accordingly, a special treatment of kick-down is in a Gear selection is not required.

Furthermore, uphill and downhill detection is normal automated manual transmissions (ASG) also available with a gear selector.

When descending and braking the engine (overrun), especially when negative drive torque, there is no time expectation possible and therefore gear selection cannot be used in a kick-down.

Accordingly, the following basic strategy strategy can be proposed, in the first version of which only the next gear above the activated gear is selected. The following table results:

A second version is a further improved basic version of the strategy, and a first improvement can be achieved by implementing a consideration of the vehicle speed. This results in the following table:

In this case, the xnA value and the xnB value can be used to determine the best Gear selection can be calibrated and some hysteresis as speed limits exhibit. For safe operation, the gear selector should be in neutral choose when the vehicle is stationary. This creates dangerous situations avoided if a clutch (i.e. with a selected gear) reacts incorrectly and an unwanted coupling takes place. This can also be provided if the clutch is too hot and needs to be cooled. Cooling the Clutch can be done better if the clutch rotates. At a Electronic manual transmission with gear selection should be in the neutral state Battery charging can be enabled.

A further improved basic version of the strategy is proposed below, in which starting processes in second gear are made possible as an improvement, in which the neutral state is preselected on the other shaft. The following table results for this improved version:

It can be provided that the vehicle speed is a predetermined one Speed is such that a predetermined vehicle acceleration is ensured.

Further improvements can be planned. For example Prevent oscillating gear shifts at fast Speed changes can be provided that a time counter is used, which counts the time to change the gear selector before using the preset becomes.

For the handling of a tip-in (kick-down), the Consider acceleration position or movement and a suitable tip-in Select gear selection before the gearshift image recognizes the same tip-in.

Furthermore, in the event of full braking (panic braking), this can be avoided possible problems with full braking that are provided in this Situation is selected as gear selection of neutral.

Furthermore, in the case of driving mode deviations, it can be provided that the Vehicle speeds and the hysteresis deviations depending on suitably calibrated to the driving mode (normal, sport or winter program or the like) become.

With heavy loads or uphill driving can be for the best possible Gear selection with respect to the drive torque are provided that the Drive torque (active clutch torque) when assessing when the Gear selection is determined, is used.

Furthermore, in particular with a free rocking function, namely free rocking the vehicle forward and backward to come out of a ditch, be provided that, for. B. in the winter program or the like. The second course and the reverse gear is set to the rocking function perform. In this way, in the drive train between the Reverse gear and the second gear for rocking or swinging freely be switched so that the vehicle out of the obstacle, for. B. from one Digging, can be removed.

Furthermore, it can be provided that the vehicle acceleration into a Overview or is recorded in a memory, so that when the Vehicle speed changes rapidly due to braking or acceleration is recorded or taken into account when selecting the gear. This can z. B. can be realized if the vehicle speed, which for the Decision of the gear selector used is a predetermined estimated Speed based on the measured acceleration. In this case it is also possible to recognize gear selection changes more quickly.

It is also possible, when driving uphill and at vehicle speeds, which are below a certain limit, to preselect a lower gear.

It is also possible to make other changes to the to further improve the inventive strategy.

In the case of electronic gearboxes in particular, it can be provided that the Gear preselection for parallel gearboxes is canceled. Can also Features used for an electronic manual transmission implemented in this way if the use is also advantageous for a parallel transmission. For example, the neutral state can occur in the electronic manual transmission the inactive shaft at standstill to charge the battery.

A further embodiment of the present invention follows described, which a hill start aid preferred by creep torque activation provides when releasing the brake.

Automatic gearboxes (ASG) and parallel gearboxes (PSG) are automatic transmissions, which are similarly operated by the driver, as well conventional automatic converter transmissions. However, it has been shown that at the aforementioned transmission types can occur that the vehicle at Starting to roll back on a mountain. This may be because that Creep torque of the clutch is only built up when the driver releases the brake has solved.

• 1. Wenn der Bremsdruck von oben kommend eine definierte Schwelle unterschreitet, kann die normale Kriechstrategie aktiviert werden.
• 2. Wenn der Bremsdruck von oben kommend eine definierte Schwelle unterschreitet, kann in funktionaler Abhängigkeit bevorzugt vom Bremsdruck ein Kriechmoment an der Kupplung aufgebaut werden. Beispielsweise kann mit abnehmendem Bremsdruck ein stetig steigendes Kriechmoment aufgebaut werden, bis das maximale zulässige Kriechmoment an der Kupplung anliegt.
• 3. Bei dem Aufbau des Kriechmomentes können verschiedene funktionale Abhängigkeit berücksichtigt werden. Beispielsweise kann das Kriechmoment stetig linear, progressiv, degressiv oder auch stufenweise mit abnehmendem Bremsdruck erhöht werden.

Accordingly, rolling back, particularly when driving uphill, can be prevented or significantly reduced by an earlier creep torque build-up. For this purpose, it can be provided that, in the case of an ASG as well as a PSG system, it can be recognized, for example on the basis of the brake pressure, that the braking force requested by the driver drops. For example, the creep torque can be built up when the temperature falls below a defined threshold. Various options can be provided for this:

• 1. If the brake pressure coming from above falls below a defined threshold, the normal creep strategy can be activated.
• 2. If the brake pressure coming from above falls below a defined threshold, a creep torque can be built up on the clutch, depending on the function of the brake pressure. For example, a continuously increasing creep torque can be built up with decreasing brake pressure until the maximum permissible creep torque is applied to the clutch.
• 3. Various functional dependencies can be taken into account when building the creep torque. For example, the creep torque can be increased steadily in a linear, progressive, degressive or stepwise manner with decreasing brake pressure.

Since after the activation of the creep torque the vehicle is already moving can put, if the driver is still on the brake pedal, a correlation be particularly advantageous with the falling brake pressure. When the vehicle is on an incline held by the driver with the brake can be assumed that the brake pressure threshold has been exceeded. So already when the brake is released, the crawl can be activated as a starting aid. They are too other possible dependencies when considering the structure of the Creep torque conceivable.

When building the creep torque, however, it should be taken into account that the Creep torque is not generally activated at low brake pressures, otherwise the vehicle could move, for example, in the event of a slight brake application Level, don't let it stop. Furthermore, a constant crawl against the Brake increases the energy input into the clutch. Consequently, the Increase fuel consumption while idling and possibly cause annoying vibrations occur in the vehicle. It can also be provided that the known strategy preferred to determine the tactile point, which works at a low torque level can be combined with the creep strategy according to the invention.

The crawling strategy presented is illustrated by means of a flow chart in FIG. 66.

Further, 67 a possible functional relationship between the braking pressure and the creep torque is shown with respect to a threshold value in Fig.. In Fig. 68, a jump-like Kriechmomentenaufbau is shown when falling below the threshold brake pressure. Finally, a history of the creeping torque, 69 is illustrated in FIG. Via the brake pressure with an increased creep torque is provided when closing a low creep torque and the brake is opened.

It should also be noted that the relationships shown in FIGS. 67 to 69 only apply if the brake pressure threshold has been exceeded. Otherwise, the dotted lines, which show the closing of the brake, also apply to the opening of the brake.

Accordingly, a hill start aid is proposed, in which a Creep torque build-up depending on the brake pressure or one Brake pressure threshold is provided. This hill start aid can of course also used in other driving situations and preferred in automated Manual gearboxes (ASG) and parallel gearboxes (PSG) are used.

A further embodiment of the present invention follows described, which a shift strategy, in particular for parallel gearbox Train upshifts with incorrectly selected target gear, suggests.

It has been shown that it occurs particularly in parallel gearboxes may be that the wrong gear is engaged on the target shaft at the start of the shift has been. The target gear must then be used for a successful overlap shift be changed first. You can do this especially with downshifts Use time to accelerate the engine to the target speed. at Upshifts will continue to accelerate the vehicle until the target gear is inserted. In particular when the driver triggers it manually, it can come to loss of comfort.

Accordingly, it is proposed in the switching strategy according to the invention that by reducing the engine torque, preferably at the beginning of the train upshift the driver can get the impression that there is already a circuit be performed. For this purpose, the torque of the motor can preferably be maximal be lowered so that the driver experiences the acceleration that he has with his Driver's desired moment in the target gear. The only difference is for the Drivers in that the synchronization of the engine speed to the finish gear later than with a circuit with a correctly selected target gear.

A particular advantage of this strategy is that the vehicle is in the first Phase is no longer fully accelerated, and especially with full load circuits Danger in the area of the speed limiter of the internal combustion engine arrive, is reduced.

The following equation can apply to the maximum reduced engine torque:

The lowering of the output torque to the target gear level can be determined, for example, by comfort. If it takes longer to engage the target gear than to shut down the engine on M _Red , the target torque M _Red can be held until the target gear is engaged. If the phase is shorter, you can start overlapping immediately (before reaching M _Red ). In this case, however, there may be a restriction due to the maximum travel speed of the couplings.

In FIG. 70, the torque and speed curves are shown schematically in the case of a train upshift when the target gear is engaged in the first phase. In Fig. 71 a flow diagram is shown for the switching strategy presented here, wherein the first phase of the circuit, so the engagement of the target gear, is taken into account. In Fig. 72, a flowchart of the shifting strategy according to the invention, wherein the second phase of the circuit, so the overlap is considered.

The switching strategy according to the invention presented here can be used in particular with uninterruptible gearboxes (USG) and especially at Parallel gearboxes are used.

A further embodiment of the present invention follows described which overlap circuits with a slip control over the couplings provides.

With possible strategies for the overlap circuit with one Parallel shift transmission can be provided that the internal combustion engine of the Vehicle is used to correct inaccuracies of the moments involved Compensate components. For example, two can overlap slipping clutches are used. In a first phase, the Clutch of the old gear to be put into slip. So far this has been done by the Internal combustion engine allows the slip until the end of the overlap holds.

The problem is that the internal combustion engine is not always as precise and above all, it cannot be controlled so quickly that in a relatively short time Inaccuracies in the overlap can be compensated for. Therefore it is a task of the strategy according to the invention to indicate a possibility in which does not require engine intervention, and yet is robust to this strategy Torque errors of the clutches is.

In the proposed strategy, the part of the overlap is preferred taken into account, up to which the clutch of the target gear completely torque take over. Accordingly, the core of the strategy according to the invention is that the clutch of the target gear starts slipping when the engine torque is unaffected the clutch of the old gear keeps as constant as possible. This can preferably be done slip control to control the target clutch during the overlap, for example with a simple PID controller or the like. It is also possible that the regulation with the opening clutch and / or with two clutches is realized.

It is particularly advantageous in the strategy according to the invention that the clutch in usually faster to target requirements than the internal combustion engine. Furthermore, the clutches are in contrast to the internal combustion engine (full load) high moments not limited. Furthermore, the invention Strategy enables that with stable control never more moment on the downforce acts as desired by the driver. Furthermore, the invention Strategy no engine intervention required.

In the strategy according to the invention, for a sensible regulation, it should be taken into account that the regulation times between the target and the actual specification are clearly below the overlap time. It has been shown that the clutch of the target gear can only moderately follow the target torque, and thus corrections of the torque have little effect. The reason may be that the control behavior is very sluggish (τ _PT1 = 50 ms) and, on the other hand, this may be due to the course of the clutch characteristic curve, with large changes in the release at small moments.

• - Es werden keine Rampen bei der Überschneidung verwendet, sondern parabolisch gestartet und danach in einen linearen Verlauf übergegangen. Dadurch kann die zweite Kupplung langsam in einen Arbeitsbereich kommen, in dem die Regelzeiten ähnlich der ersten Kupplung sind.
• - Die Überschneidungsdauer wird geeignet angepasst. Für das in der Simulation benutzte Model des Positionsreglers der Kupplung (τ_PT1 = 50 ms) ergibt sich eine Mindestzeit von ca. 400 bis 500 ms für die Dauer der Überschneidung.

In the strategy according to the invention, the following can be provided to avoid the aforementioned:

• - No ramps are used for the overlap, but started parabolically and then changed to a linear course. As a result, the second clutch can slowly come into a working range in which the control times are similar to the first clutch.
• - The overlap period is adjusted appropriately. For the model of the position controller of the clutch used in the simulation (τ _PT1 = 50 ms), there is a minimum time of approx. 400 to 500 ms for the duration of the overlap.

There are also other measures or combinations of the above Measures possible. It is also conceivable that the measures mentioned are modified in order to further improve the strategy according to the invention.

Furthermore, according to a development of the invention, it can be provided that the speed controller is implemented. It can be provided during the specification of the engine torque and the torque of the old clutch that the torque of the new clutch is regulated appropriately. The following equations can apply to the moments:
M _Mot = M _{driver request}
M _{clutch_alt} = M _{clutch_alt_start} - ƒ (t)
M _{Kupplung_neu} = ƒ (t) - (M _{Mot_Start} - M _Mot) + M _regulator

The input variable for the PID controller can be the difference from the Engine speed and the input shaft speed of the old gear in particular Upshifts or the new gear can be used in downshifts. The speed change of the input shafts can be due to the Coupling torques or the counter torques from the output, such as Driving resistance, road gradient or the like, not in the controller flow back, but flow in the same direction in principle, d. H. the slip can be reduced if the clutch of the target gear is further closed becomes.

In order to ensure that a sufficient slip reserve is available during the overlap, the desired slip can preferably be built up to the required reserve at the beginning of the overlap. This can also be seen from the representations in FIG. 74, in particular the course delta n. In the representations of Fig. 74, the rotational speed and the torque is shown over time in each case.

Accordingly, the clutch of the old gear can be opened slowly, and only after reaching the slip reserve, the second clutch can be closed. The build-up of the hatch reserve can be done via a time-dependent structure of the Target slip should be regulated in such a way that the clutch of the target gear is relatively early begins to build up the target slip.

It is possible that there is a positive slip in the case of train circuits and at A negative slip is built up. In the case of the invention Strategy can be provided that in the case of train downshifts or thrust upshifts the engine speed is still brought to the target speed of the new gear, which corresponds to phase 2 before the overlap begins.

In Figs. 74 and 75, the strategies are shown in Zughochschaltungen example. A simulation of an overlap during a 1-2 train upshift is indicated in FIG. 75.

It has been shown that the regulation of the slip also errors in the Can eliminate clutch torque and slip between 0 and about 200 Rotation / min. The use of the slip control during the Driving with a partial transmission can also be used for the slip build-up. For this purpose, preferably the slip during the overlap with the clutch of the old aisle.

A flowchart of the strategy according to the invention is shown in FIG. 76, the control of the overlap being realized up to the point at which the old clutch is fully opened.

The strategy according to the invention can preferably be used in parallel gearboxes (PSG), but may also be used in automatic transmissions.

A further embodiment of the present invention follows described, which in particular an emergency operation with opening clutches relates to a parallel transmission.

It has been shown that a general opening of the Couplings is required. Furthermore, there can be at least one possibility for emergency opening of the clutches or for stopping the vehicle. It can also hydraulic clutch actuation and clutches used become. Overall, there should be a possibility that makes it possible to Run open the clutches. For example, this can be done by a Electromotive clutch actuators are provided.

The automated double clutch system consists in particular of a dry double clutch, which can include pressed-in, force-released clutches. A hydrostatic disengagement system with at least one emergency valve is also provided. Furthermore, an electromotive clutch actuator is used to actuate the clutches. Such a double clutch system is shown schematically in FIG. 77.

By opening the emergency valve or by opening several emergency valves, for example, one per clutch, the fluid from the pressure chamber of the hydrostatic release system in a depressurized room, for example in a Storage containers escape, and the clutches can then be opened without force become.

The emergency running behavior according to the invention can be both by the type and decisively influenced by the actuation or control of the emergency valve become.

It is possible for a normally open valve to be supplied by the transmission control unit of the transmission, which is basically actively closed during operation of the control unit, and to be opened automatically in principle when the transmission control unit is switched off or when the power supply to the control unit fails. This is also shown in Fig. 79. In this figure, a possible circuit diagram variant with an emergency valve is shown, which is controlled by the transmission control unit.

Furthermore, a normally open valve can be supplied by the ignition in parallel to the transmission control unit. It is provided that this valve is basically actively closed when the ignition is switched on. When the ignition is switched off, the valve can be opened so that the vehicle can be moved with the engine switched off. This variant is shown in Fig. 78. In this figure, another possible circuit diagram variant with an emergency valve is indicated, which is provided in parallel to the transmission control unit.

Furthermore, a normally open valve can be supplied by the transmission control unit and is normally actively closed when the control unit is in operation. With the detection of defined system errors, such as the failure of a clutch actuator or the processor in the control unit and / or in certain situations, such as protection against stalling for the internal combustion engine and when the transmission control unit is switched off or the power supply to the transmission control unit fails, Emergency valve can be opened. This is shown in particular in FIGS. 77 and 80. A suitable control logic for the emergency running behavior according to the invention is indicated in FIG. 80.

Mechanical valve actuation can also be provided. This can be an exclusive mechanical actuation or an additional one electrical control of the emergency valve are provided so that a manual emergency release of the clutches, for example for towing the Vehicle, is advantageously possible.

The aforementioned variants with regard to the arrangement and the actuation at least of an emergency valve can also be suitably modified and / or arbitrarily can be combined with each other.

The emergency running behavior proposed according to the invention can be used in particular for Parallel gearboxes can be used advantageously because there, unlike one so-called add-on automated manual transmission (ASG), the roll-away protection is not guaranteed by a parking lock, and thus an opening of the clutch in this situation is not safety-critical. It is also possible that Proposed emergency running behavior with an electronic clutch management System (EKM) is provided.

A further embodiment of the present invention follows described, which a parameterization especially in overrun downshifts in a parallel gearbox.

Accordingly, it is proposed that no intermediate gas is provided for increasing the engine speed, particularly in the case of overrun downshifts. This means that the vehicle can be operated in a fuel-efficient manner in overrun mode. When the slip is reduced after a change in gear ratio, the clutch can conclude that the torque that can be transmitted is higher than the engine drag torque, and can therefore increase the engine speed until the clutch of the new lower gear sticks. This is shown in FIG. 81, with the torque curves of the on-going clutch, the coming clutch and the internal combustion engine and the speed of the internal combustion engine being shown. In particular, the curves for a downshift in a parallel gearbox without intermediate gas are indicated.

If, for example, winter operation (winter program) is activated in the transmission control unit, which the driver can request, for example, by means of a winter button or winter switch, or which the transmission control system can automatically know through the accumulation of ABS and / or ESP control processes, an intermediate gas can occur during overrun downshifts to be activated. This means that the clutch of the lower gear does not need to build up any or only a minimal over torque when reducing slip after the gear ratio change, and an increase in thrust torque in wheel-road contact is largely avoided. This is shown in particular in FIG. 82. In this figure, as in FIG. 81, the torque and speed curves are shown in a thrust downshift, but here with intermediate gas.

In the case of a parallel gearbox, the gearbox is to be compared to other gearboxes greatest possible consumption advantage can be achieved. With thrust downshifts this means that there is no need for the intermediate gas to raise the engine. This can result in a consumption advantage of over 1% for certain driving cycles can be achieved.

It should be noted that the receiving clutch is an additional Coupling torque must be transferred to the speed of the internal combustion engine compared to the drag torque of the internal combustion engine. This can but when the road is slippery, especially in winter, the wheels stick open lead on the road and thus cause dangerous driving situations. As a result in the parameterization according to the invention a compromise between the Consumption savings and driving safety realized.

A corresponding flowchart for the parameterization selection is shown in FIG. 83.

A further embodiment of the present invention follows described, in which a reduction in thermal stress in particular the starting clutch of a double clutch is proposed.

It has been shown that, especially with repeated approaches and stops on the mountain in particular the starting clutch is subject to considerable thermal stress.

Accordingly, a possibility is given for the capacity of the second clutch a double clutch transmission to increase the thermal load capacity of the To use the entire system.

According to the invention, it can be provided that at a limit temperature to be specified, such as 200 ° C. in the areas shown in FIGS. 84 and 85, the second clutch of the double clutch for transmitting the torque on repeated starts or when stopping on the mountain to use with the coupling. The second clutch is usually not the starting clutch. By sharing the second clutch, the overall thermal capacity of the clutch can advantageously be increased, and thus the maximum possible duration of stopping on the mountain or the maximum possible number of start-up cycles until the maximum permissible clutch temperature is reached, as is also the case is indicated in FIGS. 84 and 85. The actual clutch temperature can preferably supply the clutch temperature model of the clutch control as an input variable.

The ratio of the transmitted torque components of clutch 1 and clutch 2 The overall moment can be controlled by control and / or regulation strategies be divided that both clutches approximately simultaneously the maximum reach the permissible coupling temperature. For example, after falling below the above-mentioned limit temperature due to a cooling phase original starting strategy can be used. There are others too Modifications to the starting strategy presented here are possible in order to reduce the thermal To further reduce the load on the double clutch.

In Fig. 84, the temperatures of the two clutches 1 and 2 are each shown with a 1-coupling strategy and a 2-coupling strategy, wherein the driving state holding is selected on the mountain.

In Fig. 85, the temperatures of the clutches 1 and 2 are also plotted over time, wherein the 2-coupling strategy are shown in the upper diagram, the 1-coupling strategy and in the lower diagram. For both diagrams, multiple approaches, especially on the mountain, have been selected.

The claim submitted with the application is a proposal for formulation without prejudice for obtaining further patent protection. The applicant reserves the right to do more, so far only in the description and / or drawings claimed combination of features.

The exemplary embodiments are not to be understood as a limitation of the invention. Rather, there are numerous changes within the scope of the present disclosure and modifications possible, in particular such variants, elements and Combinations and / or materials, for example by combination or Modification of individual in connection with that in the general description and embodiments and the claims described and in the Features or elements contained in drawings or process steps for the person skilled in the art with regard to the solution of the problem can be taken from and by Combinable features for a new item or new ones Lead process steps or process step sequences, also insofar as they Test and work procedures concern.

Claims (2)

1. Transmission for a motor vehicle with at least one automated by means of a operated friction clutch can be coupled to an internal combustion engine Transmission input shaft and at least one with at least one drive wheel related transmission output shaft, being between the at least one transmission input shaft and the at least one Transmission output shaft each have at least one gear ratio by means of one Control unit controlled actuator is switchable. 2. A method of operating a transmission according to claim 1.