DE10308690A1 - Gear changing strategy for automobile dual-clutch gearbox has input shaft of inactive parallel drive train increased during gear pre-selection - Google Patents

Abstract

The gear changing strategy provides preselection of a target gear for the inactive parallel drive train of the dual-clutch gearbox, with the revs of the input shaft of the inactive parallel drive train increased during the gear pre-selection, by closing the clutch of the inactive parallel drive train during the neutral phase. An Independent claim for a dual-clutch gearbox for an automobile is also included.

Description

The present invention relates to a transmission and a shift strategy for a transmission, in particular for a dual clutch transmission, one Vehicle in which a target gear on a non-torque transmitting Transmission input shaft is selected.

Gearboxes and their shift strategies are known from vehicle technology. If z. B. a dual clutch transmission or a parallel transmission in a gear with a constant state, one with the not activated or non-torque transmitting transmission input shaft Gear can be switched. This additional gear change can also, for. B. at an automated manual transmission. This additional Gear changes can be compared in particular with the dual clutch transmission with the automated manual transmission, a higher wear at the Effect synchronization.

Accordingly, the object of the present invention is a transmission and a shift strategy for a transmission, especially for a Dual clutch transmission to propose a vehicle through which in particular the wear during synchronization is reduced and also the Time to synchronize the selected gear is reduced.

This task is procedurally achieved through a shift strategy for a transmission, in particular for a dual clutch transmission, a vehicle in which a Target gear selected on a transmission input shaft that does not transmit torque is solved in which the speed of the non-torque transmitting Transmission input shaft is increased during the gear preselection.

Accordingly, the switching strategy according to the invention provides support synchronization during gear preselection for a transmission, realized in particular in a parallel transmission or the like. By the increase in the speed of the non-torque transmitting Transmission input shaft can be the differential speed during synchronization be significantly reduced.

According to a further development, the proposed switching strategy can provide that the clutch of the transmission input shaft not transmitting torque is preferably closed during the neutral phase in the area code in this way the speed of the non-torque transmitting Increase transmission input shaft.

Particularly advantageous for the later synchronization of the preselected gear it is when the speed of the shift strategy according to the invention is not torque transmitting or not activated transmission input shaft to the Engine speed or the like is adjusted. This is particularly the case if the engine speed between the speed of the non-torque transmitting Transmission input shaft lies before and after the preselection of the gear. This means that the engine speed between the later target speed and the before the preselected speed of the no torque transmitting Transmission input shaft should lie.

The switching strategy according to the invention can further provide that according to increasing the speed of the non-torque transmitting Gearbox input shaft of the preselected gear is synchronized. Through the lower speed difference can synchronize faster and also be carried out with less wear on the transmission.

In an advantageous embodiment of the present invention, the the clutch assigned to the non-torque transmitting transmission input shaft preferably over a predetermined ramp curve of the clutch torque or a defined ramp function to a defined, predetermined value be closed, preferably after reaching the engine speed The clutch is opened again and the preselected gear is synchronized.

The use of the switching strategy according to the invention is particularly advantageous with a downshift. It is also preferred to use one Parallel shift transmission with two transmission input shafts and two assigned Couplings provided. However, it is also possible that the proposed Shift strategy is used in other transmission systems.

Furthermore, the object underlying the invention can be device-wise by a transmission, in particular a dual clutch transmission, for a vehicle with at least two transmission input shafts, each with a clutch is assigned, in particular to carry out the proposed Switching strategy can be solved in which at least one control device or the like for increasing the speed of the non-torque transmitting Transmission input shaft is provided during the gear preselection of a gear.

By increasing the speed of the non-torque transmitting Transmission input shaft and the associated reduction in the synchronizing speed difference, the gear change can be faster and be carried out with less wear in the transmission according to the invention.

Furthermore, the invention relates to a method, a device and their Use for operating a motor vehicle, in particular for control a parallel gearbox (PSG), with a drive motor, a clutch and / or a transmission is 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.

An automatic transmission can also be used as the USG, whereby an automatic transmission a transmission essentially without interruption of tractive power in the switching operations and this is usually through planetary gear stages is constructed.

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 a shows that a throttle valve sensor 15 , an engine speed sensor 16 and a speedometer sensor 17 can be used and transmit measured values or information to the control unit 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 .

The Fig. 1a shows in addition to the load lever 30 and the sensors associated therewith a brake actuator 40 to actuate the service brake or the parking brake, such. 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 is the total ratio of the higher gear. 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. 2a.

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

FIG. 2a shows the course of acceleration during and after an upshift for a vehicle with a dual clutch transmission. It is clear from FIG. 2a 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. 2a.

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 Circuit type, e.g. B. upshift or downshift, the moment in Drivetrain through coordinated actions of the two clutches and the Motor is controlled such that preferably drivetrain vibrations be suppressed and 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, the each continuously connected to the crankshaft of the engine by a friction clutch could be. At one input shaft z. B. gears 1, 3 and 5 lie, while on the other input shaft the gear steps 2, 4 and 6 are. Simple upshifts and downshifts can e.g. B. be carried out in such a way so that the target gear is engaged on the disengaged input shaft (shaft 2) the input shaft of the actual gear (shaft 1) is disengaged while shaft 2 is engaged and the engine speed is the speed of shaft 2 is adjusted.

Such a transition from wave 1 to wave 2 results in a Interruption of traction avoided. It is important to ensure that neither by blending the couplings through the subsequent ones Synchronization of the engine speed a noticeable jerk is stimulated, which namely the driving comfort deteriorates.

In the strategy according to the invention there is a smooth translation change provided with a suitable synchronization. For example, the Strategy according to the invention can also be used for pilot control and then 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, there can be three cases in particular differentiate:

1. Clutch 1 sticks, clutch 2 slips

The moment balance provides the following two equations:

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 adhesion slip or Slip-hold transitions should have to avoid unwanted jerky movements to avoid. Assuming that before and after Gear ratio change, the clutch transmitted should be liable, the Clutch 1 has at least one slip-slip transition and clutch 2 experience at least one slip-detention transition. These transitions can be designed smoothly.

• 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
  
  
  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:
  
  
  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
  
  
  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:
  
  
  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 is not more about the driver's desired torque can be increased. In this case build up a hatch reserve by using 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 ₄ 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, it can be simulated using simulation. It is e.g. B. possible that the equations of motion 5a, 7a and 9a are integrated. Before and after the shift, ie before phase 1 and after phase 4, the transmissible torque of the 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
Tr: -5 Nm - 10 Nmh / km wheel 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 from this behavior can be easily quantified (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 noticeable jerk after phase 4 arises because the Synchronization in phase 4 has not been completed. With this deviation from the ideal strategy, which is artificial in the simulation was also to be expected for 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 shift. 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 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 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 to make the switch is not only carried out without reducing the pull force, but - if Engine torque and clutch torque can be controlled with sufficient accuracy - too without noticeable jerk. If the exact controllability of the moments is not the strategy for precontrol can be used a regulation of the clutch torques is superimposed. Overall, with the strategy indicated using circuits in dual clutch transmissions a minimum of loss of comfort and thus the quality of a conventional step machine.

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 clutch 1 is fully open and that the angular velocity ω _{2 of} the transmission input shaft is 12 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 the context of an advantageous embodiment of the present invention four different switching types can be defined, namely pull Upshift, overrun upshift, train downshift and overrun Downshift. The described lack of comfort can be solved by the Fix the following fade strategy, where the initial gear speed is as Speed of the transmission input shaft of the initial gear (initial gear is inserted) and the target gear speed as the speed of the transmission input shaft of the target gait (target gait 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 can be suitably combined with these.

In phase 1, the state of sticking to slip exists. By linear The clutch torque of the transmitted clutch can be reduced these are brought into slip. At the end of phase 1, i.e. immediately Before phase 2 begins, a decision is made as to whether the drive train is in the train or overrun condition. If the slip on the transmitted clutch is positive (motor speed> speed of the input shaft), it can be concluded that there is a train condition. With a negative slip (Engine speed <speed of the input shaft) there is an overrun condition. As soon as the clutch slips, the process moves on to phase 2.

In phase 2, a so-called slip reserve is built up. This can the torque of the transmitting clutch constant (or at slip limit) being held. By a suitable increase or decrease in Engine torque with respect to the driver's desired torque becomes the engine speed brought to target speed. If the engine torque is insufficient to adjust the Reaching the target speed within an acceptable time can, for example additionally the torque of the transmitting clutch can be reduced. It is the target speed when the cable is switched the maximum of the starting gear speed and the target gear speed plus a slip reserve. With overruns the target speed is defined by the minimum of the starting gear speed and the target gear speed is reduced by the slip reserve.

In phase 3, cross-fading is provided. The moment of transmitted clutch can preferably with a constant ramp z. B. linearly reduced to the value 0, while at the same time the moment of Clutch of 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. By suitable Increase or decrease in engine torque with respect to Driver's desired torque can set the engine speed to a desired one Target speed are brought. If the engine torque is insufficient or is small in order to reach the target speed within an acceptable time, z. B. additionally the torque of the clutch of the target gear can be increased. there can the target speed in train circuits by the speed of the Transmission input shaft and the added slip reserve result and at Thrust circuits is the target speed the speed of the transmission input shaft of the finish line reduced by a hatch 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 based on the speed of the clutch Bring the target gear until the clutch of the target gear changes to stick. The slip-to-custody transition should be carried out with a smoothing order to achieve a smooth transition and thus jerky movements avoid.

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 various sizes during the each circuit shown over time or over the switching phases. In The upper diagrams shows the speed of the transmission input shaft with A, the Speed of the gearbox input shaft with B and the engine speed with C characterized. The current diagrams are shown in the middle diagrams Motor torque with a solid line and the driver's desired torque with dashed line indicated. In the diagrams below are those of Clutches transmitted moments of clutch A and clutch 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.

It was found that with a dual clutch transmission without Interruption of 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 Shifting cannot be done directly through a normal double clutch shift be carried out because there is no change in the transmitted input shaft this switching method is required.

Accordingly, one task is to find an appropriate one To develop a dual clutch control strategy for the aforementioned case is made possible with the most comfortable switching 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 another transmission input shaft. This intermediate course can then be a Moment transferred to the downforce while on the other Gearbox input shaft the starting gear is switched to the target gear. On in this way, the shift is carried out without interrupting the tensile 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. 10a, 11a to 13.

The target gear can be assigned to the input shaft B, for example then instead of switching directly to the target gear, a gear of the input shaft A as Intermediate gear is used. The translation stage 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 gear Speed is referred to the above 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, if 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 other phases already mentioned can be combined as desired.

Phase 1 and phase 2 correspond to phases 1 and 2 respectively aforementioned dual clutch control strategy, but in phase 2 a difference is provided, namely that phase 2 now described twice, namely first when shifting from the starting gear in the 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 case 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 coupling of the other input shaft to the slip limit is driven. 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. 10a, 11a 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. 10a, 11a 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, with the gear ratio of the intermediate gear 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.

In Fig. 11 a possible course of the above dimensions is shown upshift thrust during a. In the left column of Fig. 11, a 2- (1) -4 boost upshift is shown 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 downshift train. In the left column, a 4- (1) -2 train downshift is shown, 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 the tractive force is avoided during the shift. Thus become disadvantageous Impairment of comfort prevented.

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

It should be a simple and precise method for determining the external vehicle torque specified in the dual clutch transmission become.

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, which affect the vehicle such. B. the air resistance Frictional resistance between the road and the wheels, the slope of the Road (gravity of the vehicle during an uphill / downhill descent) and / or the actuation of the vehicle brake.

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-Ubergang 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:
  
  
  

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 slip-slip transition of clutch A. 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:
  
  
  

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, i _B the Total gear ratio 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, it is particularly advantageous that the calculations of the slip limits of the clutches are possible and thus the implementation of a clutch control strategy m is possible in the simplest possible way.

Furthermore, another possible embodiment of the presented here Invention described. Within the scope of this configuration, a procedure is used proposed with which the determination of the power flow in a transmission, in particular in a double clutch transmission is possible to preferably the implementation of the aforementioned dual clutch control strategies improve.

With the method according to the invention, a simple and possible exact method for determining the power flow in the Proposed dual clutch transmission, which can also be used to determine whether there is a pull or a push condition.

Three situations within a dual clutch circuit are preferred demonstrated on the basis of which the determination of the power flow in the drive train is possible. The pull-push estimate can preferably be made immediately before Shifting, i.e. before the start of the clutch change, be carried out, so that the result obtained is advantageously as current as possible. It is this estimate is also possible at another suitable time perform.

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 _en g ≥ ≙ _eng .J _eng
  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 the appropriate shift strategy (train upshift, push upshift, train Downshift or thrust downshift). In this way unnecessary changes from the train state to the thrust state or vice versa and thus avoiding running through the gear play, which could have a negative impact on driver comfort.

Thus, the comfort impairment is presented by the Dual clutch strategy minimized.

In Fig. 16 is a flow chart is shown, which thrust train estimate illustrates 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 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 To achieve electric motor torque. The strategy according to the invention differs in particular from the aforementioned strategies in that an electric motor is included in the model. Furthermore, upshifting and downshift separated into train upshift and push upshift, d. H. four different switching types instead of two different types, are used. Furthermore, active control of the Output torque with the electric motor or one of the clutches, by 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:
  
  
  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:
  
  
  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: To enable convenient switching from one shaft to another. The Shifting comfort is achieved through the even Vehicle acceleration during the shift. I.e. a bad one Shifting comfort is characterized by significant and sudden There are deviations from the driver's desired acceleration (wheel torque) and especially 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, only two slip detentions Transitions may be provided. For example, a transition can be provided while the disengaged clutch is open and a transition can take place be provided while the new clutch is closed. The Engine torque can be used to increase or decrease actively ensure smooth slip-stick transitions so that the Engine speed can be controlled to prevent slippage during shifting to achieve and the engine synchronization compared to the speed of the new Increase input shaft.

Furthermore, the torque control is provided at the output torque in order to to achieve the driver's desired vehicle acceleration (wheel torque). Primary can achieve this goal by actively using the electric motor torque during the Circuit can be achieved. Can also be secondary if the electric motor is not is available or the desired torque is above the maximum achievable e- Engine torque is one of the clutches to be controlled in such a way Get driver's vehicle acceleration.

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

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, be provided in the clutch control that before switching Torque with the old clutch is reduced to the slip limit. The moment when switching the new shaft by reducing the clutch torque changed on the old shaft and the clutch torque on the new shaft will be raised. 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 positivem Moment von dem Motor zum Abtriebsmoment übertragen wird. Folglich zieht der Motor das Fahrzeug (normale Art der Hochschaltung).
• - Schubhochschaltung: Hochschaltung mit negativem 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 a push situation. This should be done before Circuit recording can be provided (e.g. if the old coupling up to Slip limit is opened). Various estimates can be made be used.

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:

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:

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 procedure the old clutch after reaching the Slip limit is opened slightly. In this situation, the clutch begins to close Hatch. By observing the direction of hatching, namely by Differences in speed on the input and output side of the Coupling, can be determined if there is a pulling or pushing situation. A train situation is characterized by positive slip, i.e. H. the The speed on the inside is higher than on the outside of the clutch. It should be noted that the clutch and engine dynamics (dead time) are one May be subject to query. The estimate should therefore preferably be with one of the other estimates can be combined.

• 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. ≙T_emotor 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:
  
  
  (17d) 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:
  
  
  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. ≙T _emotor is an estimate of the external torque T _vehicle , which acts on the vehicle, whereby this can be determined since J _vehicle constant and ≙ _{vehicle are} known. This results in the following equation:
  
  
  (17 d) If the equation is substituted into Equation 16d 17d, results in the following equation i:
  
  
  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:
  
  
  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, 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). It is not to be expected that the comfort of the gearshift will depend on the precisely estimated values of T _vehicle and T _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. In the context of a further embodiment of the present invention, train upshifting with the support of the electric motor is considered when the torque is sufficient.

The tensile torque should be from the old wave to the new wave. This can be done by increasing the Engine speed can be reached via the speed of the old shaft. In this Situation is the old wave to reduce vehicle acceleration opened and the new shaft to increase vehicle acceleration closed. These reactions can cancel each other out. The E- Engine torque is used to control vehicle acceleration or adjust.

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 clutch slips.

Furthermore, the engine torque can be controlled via the driver's desired engine torque to ensure that the engine speed is greater than the speed of the engine old wave is held. The transmitted torque of the old clutch can preferably reduced to 0 with a constant ramp, while the transmitted torque in the new clutch is preferred with the same ramp to the slip limit or slightly above, where controlled the torque of the electric motor according to equations 23d or 25d to achieve the driver's desired acceleration of the vehicle.

Furthermore, the engine torque is reduced to a minimum torque in order to accelerate the engine synchronization when the clutch torque of the old clutch is 0 (clutch disengaged). The new clutch will be easy controlled over the slip limit, 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 engine new input shaft is synchronized. Then the vehicle acceleration further controlled with the electric motor.

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 moment should be on Output during the entire torque transmission from the old shaft to new wave will be preserved. This is usually done by reducing the Motor speed below the speed of the new shaft before the Torque transmission reached. Otherwise, opening the old clutch will result and closing the new clutch to increase the Vehicle acceleration. However, with the support of the electric motor torque can be counteracted, whereby the moment can be transmitted, before the engine speed is below the speed of the new shaft. In 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 clutch slips.

It is also possible to reduce the engine torque to a minimum. The The transmitted torque of the old clutch is increased via a constant ramp decreases the value 0 while the transmitted torque of the new clutch is increased to the slip limit or slightly above with the same ramp, where the torque of the electric motor is controlled according to equation 23d or 25d to achieve the 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 continue to 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 engine new input shaft is synchronized. Vehicle acceleration can then be controlled further with the electric motor.

In Fig. 19 is a simulation of a thrust upshift with the support of an electric motor is shown. 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 with the support of the electric motor if sufficient provided at this moment. The pulling moment should be during the total torque transmission from the old shaft to the new shaft on the Downforce remains. This is usually done by increasing the Motor speed over the speed of the new shaft before the Torque transmission reached. Otherwise, opening the old clutch will result and closing the new clutch to reduce the Vehicle acceleration. However, with the support of the e- Engine torque to be counteracted, the torque being transmitted can be before the engine speed over the speed of the new shaft increases. In this case, the new clutch can synchronize the Engine speed support and therefore the shift significantly accelerate.

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 clutch slips.

It is also possible that the engine torque on the Driver's desired engine torque is controlled. The transferable moment of old coupling can preferably with a constant ramp down to the value 0 be reduced while the transmissible torque of the new clutch preferably with the same ramp to the slip limit or slightly above 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.

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 changed until the motor synchronization.

It is also possible to keep the engine torque just below the Increase desired driving motor torque in order to ensure continuous slip Achieve cohesion when the engine speed matches the speed of the engine new input shaft is synchronized. The vehicle acceleration will then further regulated with the electric motor.

In Fig. 20, the simulation of a traction downshift with the support of the E-motor is illustrated. 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 moment should be on the Output during the entire torque transmission from the old shaft to new wave will be preserved. This can be done by reducing the engine speed below the speed of the old shaft. In this situation opening the old clutch increases vehicle acceleration and reduced by closing the new clutch. These reactions can cancel each other out. The electric motor torque is used for control and Vehicle acceleration setting used.

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

Furthermore, the engine torque can be reduced to a minimum torque, to make sure the engine speed is below the speed of the old one Wave remains. The transferable torque of the old clutch can be achieved with a constant ramp can be reduced to the value 0 while the transferable moment of the new coupling with the same ramp up to the Slip limit or slightly increased, taking the moment of the electric motor is driven according to equation 23d or 25d to the To achieve driver's desired vehicle acceleration.

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 up to or slightly above the slip limit controlled and the vehicle acceleration continues with the electric motor regulated.

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 both simulations.

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 do not adequately support vehicle acceleration. In this case the strategy is supported as much as possible by the electric motor; either with maximum positive or maximum negative available moment, and it can be one of the clutches to control the torque for the output be used 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 shaft. In this situation, opening the old one Clutch reduces vehicle acceleration and by closing the new clutch increases vehicle acceleration. With suitable and parallel coupling operations, these reactions can mutually cancel.

It is possible that the moment of the old clutch up to the slip limit is reduced, then the engine torque until the old hatches Clutch is increased.

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

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. Vehicle acceleration is then controlled via the new clutch.

The engine torque is just below the Driver's desired engine torque increased to an even slip Achieve cohesion when the engine speed matches the speed of the engine new input shaft is synchronized. The vehicle acceleration will further 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 Receive torque transmission from the old shaft to the new shaft stay. This is done by reducing the engine speed to below the Speed of the new shaft reached before torque transmission. In this The situation is solved by opening the old clutch Vehicle acceleration increased and by closing the new clutch Vehicle acceleration reduced. Through suitable and parallel Coupling operations can cancel these reactions against each other become.

It is possible that the moment of the old clutch up to the slip limit is reduced, in which case the engine torque is reduced until the old one 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 that Driver's desired engine torque is increased to ensure even slip Ensure transfer of liability if the engine speed is below the Speed of the new input shaft. The moment of the old clutch is reduced to 0 with a constant ramp, while the Torque of the new clutch according to equations 30d / 32d or 31d / 33d is controlled in order to achieve the 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 Traction torque should be on the output throughout Torque transmission from the old shaft to the new shaft is retained. This can be achieved by changing the engine speed over the speed of the engine new shaft is increased before the torque transmission. In this situation the vehicle acceleration is reduced by opening the old clutch and by closing the new clutch, vehicle acceleration elevated. This can be done by suitable and parallel coupling operations Reactions against each other are canceled.

It is possible that the moment of the old clutch up to the slip limit is reduced and then the engine torque until the old hatches Clutch is increased.

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

The engine torque can easily go below the driver's desired engine torque be reduced to ensure an even slip-to-slip transition to be delivered if the engine speed exceeds the speed of the new input shaft lies. The torque of the old clutch can preferably be kept constant Ramp can be reduced to 0 during the moment of today The clutch is regulated according to equations 30d / 32d or 31d / 33d Driver's vehicle acceleration to achieve until the new clutch liable.

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 the support of an electric motor. The pulling moment should be on the Output during the entire torque transmission from the old shaft to new wave will be preserved. This is done by increasing the engine speed reached via the speed of the old shaft. In this situation, by opening the old clutch reduces the vehicle acceleration and by Closing the new clutch increases 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 up to the slip limit is reduced, in which case the engine torque is reduced until the old one Clutch slips.

Furthermore, the engine torque can be below the Driver's desired engine torque can be controlled to ensure that the Engine speed is kept below the speed of the old shaft. The transmissible moment of the old clutch is e.g. B. with a constant Ramp reduced to 0 during the transferable moment of the new one The clutch is regulated according to equations 30d / 32d or 31d / 33d To achieve driver's desired vehicle acceleration.

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. Vehicle acceleration is further regulated with the new clutch.

The engine torque is just up to the driver's desired engine torque reduced to achieve a continuous slip-to-slip transition when the engine speed is synchronized with the speed of the new input shaft. The new clutch further regulates vehicle acceleration.

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 assumes a maximum before switching. If no engine torque at Increasing the engine speed beyond the speed of the old shaft is available in this case the old clutch can go below the slip limit the torque transmission can be controlled. The acceleration value can about to be maintained.

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 In this case, a type of ASG circuit can be provided. Here is the old one Clutch fully opened and the new clutch closed when the Engine speed is below the speed of the new shaft.

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

In thrust downshifts, the engine torque is always close before the shift a minimum. There is no engine torque to reduce the Motor speed available below the speed of the old shaft. In this case the old clutch must be below the Hatching limit are kept. The acceleration value can be about to be kept.

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", driving state 5 "end upshift", driving state 6 "shaft 2 active", driving state 7 "preparation for downshift", driving state 8 "torque transmission to shaft 1", driving state 9 "end downshift", 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 + air resistance + rolling resistance
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 due to incorrect dynamic equations visible are, as a result of incorrect dynamic equations that occur during a Time step used in the model, especially in the case when the clutch slip is changed from sticking to slipping (Thrust torque is then calculated as the pulling torque). However, as long as the Acceleration value before and after the peak is the same, it is expected that no jerk is perceptible in the vehicle. It should also be noted that the acceleration curves are greatly increased, so that is why in the courses shown with regard to these effects is to be considered.

As a result, an optimal strategy for comfortable shifting, especially with an ESG gearbox, with and without the support of an electronic Motors proposed.

Overall, it becomes a control strategy for translation changes preferably presented in an ESG dual clutch transmission. The goal this strategy is the control of the output torque (Vehicle acceleration) to a comfortable overlap circuit perform. The primary control means can be the electric motor. If the moment of the electric motor is not sufficiently large, the Control of 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. Alles These types of circuits can be operated with minimal jerk and high comfort be carried out with a sufficient moment from the electric motor or from is made available to the internal combustion engine. The Thrust upshifts and train downshifts can occur when using the Electric motors can be carried out significantly faster and more comfortably because of the Internal combustion engine cannot provide sufficient torque, i. H. Upshift at minimum gas or downshifts at maximum gas.

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

A further embodiment of the present invention follows described which another switching strategy especially for Parallel gearbox provides that all synchronizations are also dispensed with enabled for upshifts.

Especially for reasons of cost and competition, it is desirable on To dispense with synchronizations in dual clutch transmissions without 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.

Gear 1 is engaged on gearbox input shaft 1 (GE1) and gear 2 is engaged on gearbox input shaft 2 (GE2). This can be done in standstill without synchronization by slightly engaging clutches K1 and K2, so that gear 1 can be started , After the transition to gear 2, the gearbox input shaft GE1 can be synchronized to gear 3. For this purpose, gear 1 is first taken out and the transmission input shaft GE1 is switched to the neutral state. The gearbox input shaft GE1 now runs freely and, due to the drag losses in the gearbox, can reach the synchronous speed for gear 3 after some time. 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 gear 1, clutch K1 is applied briefly and transmission input shaft GE1 is already considerably delayed with regard to the speed of transmission input shaft GE2, which at this moment, i.e. immediately after the cross-fade shift, approximately in the ratio of the ratio i ₂ / i _{3 is} slightly above the synchronous speed of gear 3. Finally, this synchronous speed is reached as a result of the temperature-dependent drag torque, which can be recognized by the available speed signals.

If no gear is to be selected, the speed of the Gearbox input shaft GE1 falls below the synchronous speed and with This switching command is easily synchronized by applying clutch K1 become. It is also conceivable that other approaches to the switching strategy according to the invention are applied to this Improve 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 bring 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 functions of the engine in Provides controls for gearbox.

For example, systems can be provided that allow access to the Vehicle and its commissioning without conventional keys enable. For example, it can also be provided that a separate Control button is provided for starting the engine. To start the Motors have to be used especially in automatic transmissions or automated transmissions a starter approval can be provided.

With the present invention, the operation of gears and Starting or stopping the engine can be combined. That is why suggested that the operation of the transmission and the motor regarding the functions of starting and switching off. This also enables simple strategies to represent the functions that the parking of a vehicle with a manual gearbox with inlaid Gear corresponds.

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

For the actuation of automatic or automated gearboxes preferably one or more controls, such as. B. selector lever, Shift paddles, buttons on the steering wheel or the like are used. Preferably two types of actuation are possible. For one thing, only impulses passed on a certain action, e.g. B. in the transmission, namely for example, shift up or down a gear, trigger. On the other hand functions are switched on, e.g. 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, there are also other switchable positions 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 Push button switch for changing gear or changing between automatic and manual switching program is provided.

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

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

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 the failure of a fuse or the like, during a Overlap circuit fails and both clutches a moment transfer. If this is also the case in certain driving situations, e.g. B. on smoother Lane or the like, happens and the driver also goes off the gas because he has the If a failure is noticed, the wheels may slip, for example. By a suitable parameterization of the circuits, especially the train Downshifts can be a behavior of the vehicle, as with a Manual gearbox or automatic gearbox (ASG), can be achieved. Appropriate criteria and Find parameterizations. For example, such Parameterization for downshifts in a parallel transmission in Winter program or wheel slip detection can be provided.

According to the invention, the following can be proposed as a preventive measure: that with 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 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 one other suitable switch is activated by the driver, or by that the wheels just before the liability once or several times have lost. This can e.g. B. when spinning during acceleration, by an ASR intervention, by 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. The wheel slip probability is shown there 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, both with engine drag torque control in particular Failure of the transmission control or one or both clutch actuators is proposed.

In the case of a parallel gearbox (PSG) in particular, that in the course of an overlap shift, the transmission control or at least one clutch actuator fails. For example, this failure caused by the destruction of a fuse.

As a result, both clutches can transmit a moment and this Condition can be frozen if it is self-retaining electromotive actuators. It should be considered whether it is thereby blocking a transmission or at least an increased one Drag torque can come on the wheel, so that due to the failure that Vehicle could lose its lateral stability and break out.

There is a particular risk of breaking out if both Clutches permanently transmit moments and the driver of the gas and the Engine goes into overrun as well as the vehicle is on smooth for example, icy ground.

Basically there is in all types of gearbox, in particular also in Manual transmissions, the risk that the vehicle due to the Load changes from pull to push operation break out. At a Parallel gearbox (PSG) with frozen clutches, so in rare cases of emergency running, the limit shifts 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:

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 to influence in such a way that the aforementioned situations are avoided.

This can be achieved in particular by the engine control by influencing the engine drag torque in particular one Prevent detention from the outset or upon detection of a detachment intervenes accordingly regardless of driver reaction.

The engine control can determine whether the transmission control has failed and the two clutches of the parallel gearbox (PSG) permanently Transmitted moment. Accordingly, it is proposed according to the invention that a special engine drag torque control is active in the engine control is, which in the event of a load change from train to push operation of the Internal combustion engine, which basically take place on smooth roads can, and thus the risk of the wheels sticking to the road prevents. If there is a detachment, you can increase it of the engine torque, the wheels stuck to the road again become.

According to a development of the present invention, possibilities proposed to detect the transmission failure. For this, e.g. B. be provided that the transmission control emergency operation information, for. B. a Emergency run bit on the CAN, sends to the engine control system for activation this special engine drag torque control.

Another possibility is that the engine control system independently recognizes that the transmission control has failed completely and is activated then independently this special engine drag torque control. On Indication of the transmission failure can e.g. B. the missing update of the CAN Messages from the transmission control unit (values do not change, message Counter frozen). There are other ways to identify the Possible transmission failure.

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. There the special engine drag torque control with low drag torque is shown, whereby the drag torque is regulated at the trap limit.

In the aforementioned cases, the driver has the option of driving the vehicle to actively decelerate by pressing the brake pedal, then the others Systems to ensure driving stability, such as. B. ABS, ESP or the like.

the engine torque is shown in FIGS. 40 to 42 in each case plotted over time, wherein the solid line in each case shows the torque curve, which-drag torque control motor is influenced by the respective. 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 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, in can advantageously be provided, for example, that the load precontrol automatically to the conditions of the respective coupling used is adjusted. A major difference between the different Couplings is that the distance between the points of use of the Disengagement or engagement force for the starting clutch and the powershift clutch are different.

Because the disengagement force of the starting clutch is used for the position reference an additional adaptation can advantageously be dispensed with become.

However, shifting the working range of the powershift clutch requires an adjustment in particular because the load is significantly different Actuator position depends. This can be done if the Load feedforward cause the feedforward to be either too small values or suggests values that are too large. If the pilot control is too low, this leads to sluggish control behavior, since waiting for the integral part to build up must be to overcome the burden. If the pilot control is too large this leads to an override of the control, which considerably affects the actuator can strain.

A procedure is described below for additional information to get over the load curve, for example, the increase in Detect engagement force of the powershift clutch. That way certain position can then be integrated into the load feedforward control.

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 sufficient to measure a position here. For more complicated variations of the However, load profiles may also be required, more complex measurements with which the slope of the load can also be determined. It it is also possible to provide other adjustments to the load precontrol in order to to further improve the process proposed here.

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 the 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, especially disc spring, is placed. Here are apart from the incremental travel sensor, no additional sensors or mechanical parts required. However, this method should preferably be used only for actuators without compensation spring and with mechanical Release system can be used, such as. B. with an uninterrupted 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.


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:

• - 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.

In a test setup, the measurements were carried out with a modified clutch actuator (ASG) on a special test setup. 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 basic structure made of aluminum profile rails which the control unit is attached to the clutch actuator. On the Different load modules can be positioned in aluminum profiles.

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 friction torque is for movements against 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 due of the actuator gearbox is only approximately linear. It also turns Sign around.

The weak spring begins with the measurements initially provided here from a position of 0 increments (with increasing force with negative position) and the strong spring starts at -100 increments. It other conditions are possible.

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

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. An externally attached measurement of the motor current is also clear inaccurate.

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 (Δv/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 (Δv / 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) = v (x) - v (x + 40) (for movements with negative speed).

Since the speed is not measured at every position due to the time sampling of the signals, the speed curve should be partially interpolated here.

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 always look at the speed during a measurement same phase position of the motor revolution (Pos mod 40 = const.), this results in a signal removed from the modulation by motor rotation. Depending on the The phase position chosen shows a slightly different one Speed course Aside from some noise, differ the speed profiles, 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 (| v _i - v _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 enables the kink position of the load characteristic to be found with an accuracy of ± 40 increments.

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 v _i and vi-1 limit the first interval on the hard spring (the position interval found above). The measurements of v _i-2 and v _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 for each individual phase position very accurate. (At best 20 increments), but it is possible Use procedure for several phase positions of a series of measurements. A simple averaging of all possible positions that can be determined in this way is sufficient but not yet to improve the accuracy significantly.

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, who sit on the flanks of the periodic disturbance due to the Interpolation, is subject to very large fluctuations. For this reason care is therefore taken at the beginning of the measurement, these phase positions not to look at. This is done by evaluating the acceleration (△ v / At). Phase positions that are too high in the first motor revolution Limiting acceleration are not 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 is in each case a range of ± 3 to ± 5.5 increments.

The individual measurements lasted between 35 (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). However, these can possibly be reduced even further.

It can be assumed that the hardness of the spring is searched for its starting position is included in the accuracy of the position determination. With a harder one Spring can enable a more accurate determination of the reference position.

The two algorithms described above can preferably use the following algorithm, 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 should accelerate the clutch opening 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 contributes the dynamic behavior of the clutch actuator, since the gear only can be removed when the clutch is open. At the ASG The gear change takes place in such a way that the clutch torque is reduced to 0 Nm is dismantled either ramp-like or parabolic. Only in Coupling state 4, the coupling is fully opened on the HUB. In the time in who only transmits the clutch a few Nm, the gear 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 in state 4 to HUB to drive.

To open the clutch quicker and take the gear out earlier as well as being able to synchronize the new gear earlier, according to one Further development of the invention can be provided just before the gear is removed, the clutch position control on a controller is switched and the adjustment motor in the clutch actuator with max.

Voltage is operated. The position control of the clutch actuator can be such operate to start with a control deviation of approx. <4 mm, to be adjusted so that there are no overshoots if possible. This leads to, that the actuator movement becomes slower from approx. 4-5 Nm. By creating the Max. Voltage to the electric motor is the max. Actuator movement reached. From the point at which the clutch is then opened securely or wide enough (e.g. from the separation point), can switch from the voltage control to the Position control can be switched so that the way up to the position HUB (in state 4, neutral gear, the clutch is opened up to the stroke). The switch to position control must be made again in good time so that the The actuator does not drive against the stop. Synchronizing the next one Ganges only begins when the clutch position reaches a certain value has exceeded so that the clutch is safely opened. If that Clutch opening gets faster, you can move to the next gear synchronize and therefore reduce the switching time. Especially when using Active interlock makes sense and may even be necessary to use the clutch another strategy to open faster, because with Active Interlock still with closed clutch the alley can be approached and switching is carried out much faster.

According to another development of the invention, the same effect as Both of the aforementioned training can be achieved if instead of one Switch voltage control, the control deviation increases. you So instead of driving on HUB, the target path could be higher start so that the position control due to the large control deviation Max. Voltage is applied to the motor. If the coupling is wide enough The target path is opened, namely when the standby point is reached set back to HUB so that the actuator does not drive against the stop can and can be adjusted to HUB as usual with neutral gear.

As part of a further embodiment, a further improvement can be made can be achieved if you also prefer a pilot control for the Clutch actuator inserts. You know the temporal behavior of the Clutch actuator very precisely. As a result, the behavior can be very good. B. with a PT2 link or the like. With this knowledge, the Open the clutch by default, if there is still a moment to be transferred. Due to the actuator delay, the actual Actual torque on the clutch only changed after approx. 30 ms. It takes so long the actuator when energized until it moves noticeably. This time piloted, a quicker reaction of the clutch actuator to a desired target position request can be realized.

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 Path measurement to correct the transmission behavior of couplings in Parallel gearbox (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 Kupplungsaktor 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 actuator and clutch release (e.g. lever mechanism, mechanical shaft, hydraulic line, etc.)

The clutch can be actuated by moving to a specific position the drive unit in the clutch actuator. Through the transmission system the clutch release is actuated and the clutch is either closed, open or set a specific moment. The position of the The drive unit can preferably be monitored with absolute displacement sensors. The Conversion to a transmitted clutch torque is in the Control software calculated using the clutch characteristic curve in which the Transfer function "torque via master cylinder travel" is stored.

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 very far from the distance measurement in the clutch actuator removed, in addition to the coupling properties, the influences of the Release system through adaptations or zero point adjustment of the Path measurement are taken into account. So z. B. by Temperature changes the transmission behavior of a lever system or the volume of a line filled with hydraulic fluid can be changed. Due to these changes, a must be made at regular intervals Zero adjustment of the distance measurement can be carried out. This can be done by Approaching 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 carried out during the Driving with the clutch closed and gear engaged or during Stand with gear designed to be carried out (clutch is briefly closed).

The clutch characteristic can also be adjusted using the touch point adaptation be adjusted. The clutch actuator position is determined at which the Clutch begins to transmit a minimal moment. This adaptation can when the vehicle is stationary (foot brake or hand brake is applied) with gear and / or engine are idling. The Clutch is slowly closed until a minimal moment is transmitted becomes. The idle controller of the engine control reacts to the closing of the Coupling with an increase in engine torque by exactly the amount of transmitted clutch torque (3-4 Nm) so that the idling speed remains constant. This can be done based on the reaction of the engine torque 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 However, parallel shift gearboxes (PSG) can redefine these strategies become.

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

In contrast to clutch systems, they contribute Dual clutch transmissions two independent clutch actuators transmitted moments of the clutches A and B. Usually is one Clutch with the partial transmission of the odd gears (1, 3, 5) the other Coupling connected to the partial transmission of the even gears (2, 4, possibly R). in the Driving state, 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 Ä+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. When gears are engaged in the sub-transmissions of clutch A + B, disengage the gears of clutch A + B, carry out the zero point adjustment of clutch Ä + 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. there 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 for example:

• 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 be used with any transmission, especially with automated manual transmissions (ASG), e.g. B. parallel gearboxes (PSG), or in 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 proposed is to the occurrence of switching noises or gear noises, such as. B. the gear rattle, by engaging the non-driven Clutch with a predetermined torque of a selected gear to prevent 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 reach the natural frequency of the transmission.

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 advisable that this is done by engaging the non-driven input shaft coupling (IPSI) takes place while it is in neutral, causing inertia the second input shaft to the total inertia of the transmission come in addition. This in turn becomes the resonance frequency of the transmission reduced.

• - 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. Thereby vibrations are caused in the transmission, which are high Have amplitude and outside the phases of the motor and the rest Vehicle. The noise generated by these vibrations can are perceived by the vehicle occupants of a vehicle and are therefore annoying. As a result, ways must be found to get around this Prevent vibrations without affecting other aspects of the vehicle, such as z. B. affect performance and fuel consumption. This can be done by the strategy described above can be achieved.

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 system can be changed. The system that is a Vehicle powertrain is, the following parameters can be the largest Influence on the natural frequency of the gearbox: Inertia of the Input shaft, inertia of the output shaft, Gear ratio, damping spring constant of the clutch, Spring constant of the input shaft, spring constant of the axis and Mass inertia of the vehicle. It is also possible that there are others appropriate parameters can be changed to match the natural frequency of the transmission to influence.

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

Unfortunately, determining the performance of the vehicle is still pending from other aspects, such as B. Fuel consumption and Acceleration power, so the parameters that are used to prevent the Noises are used, thereby being limited. As a result in known strategies, the use of clutch damping springs provided to reduce the spring rate of the system, one Dual mass flywheel is used, which reduces the inertia of the System is increased. However, with a dual clutch transmission, the increases Possibility of implementing a number of software-controlled Strategies to actively suppress these unwanted sounds.

The differences between a parallel gearbox or one Double clutch transmission and a conventional manual transmission essentially in that in the parallel transmission two clutches and two input shafts are provided for the transmission or for the transmission are. Under normal driving conditions, one of the input waves is the Vehicle driven (IPSA), while the other shaft (IPSI) basically is disengaged, which is selected under driving conditions. For example, one of the input shafts can be gears 1, 3, 5 and Reverse gear and the other shaft 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 and the choice of gear on the non-driven shaft for the driver not apparent. This makes it possible that the inertia of the transmission by engaging and disengaging the coupling on the IPSI shaft can be set appropriately.

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 more advantageous Way 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 through the wheel half-waves shown.

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 the aforementioned equations are preferably used.

When analyzing the model, there is a comparison between the strategy with coupled neutral IPSI wave and the well-known strategy with disengaged IPSI shaft. It was also considered if the IPSI The shaft is disengaged and is in engagement 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. From this it can be seen that the IPSI clutch is engaged with no gear engaged is better than the normal strategy in which the clutch is disengaged and the IPSI wave is not engaged with any gear. The strategy in which the IPSI The shaft engages with a gear is very effective for the transmission in terms of preventing the resonance frequency. However, this is not the case Case in first gear, but wherever the IPSA translation (Translation of the activated wave) is higher than the IPSI translation (Translation of the not activated wave). This means that the IPSI Mass inertia is reduced if both translations are run through.

An embodiment of the present invention is described below, in which a further gear selection strategy, especially with one Parallel gearbox, such as. B. a dual clutch transmission proposed is to the occurrence of switching noises or gear noises, such as. B. the gear rattle, by engaging the non-driven Clutch with a predetermined torque of a selected gear to prevent 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 reach the natural frequency of the transmission.

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 can be achieved by increasing the effective mass inertia of the transmission, so the gear noise excitation frequencies are below normal Driving ranges of the engine are. In this variant, this can be done by Engage the non-driven input shaft coupling (IPSI) with a predetermined torque are made possible while the aforementioned Shaft is engaged with a gear. This is optimal for preventing Gearbox noise, because of the inertia of the second input shaft added to the first driven shaft, so that the entire Mass inertia of the transmission is increased significantly. 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.

The equations 1 g and used in the aforementioned shift selection strategy 2 g can be used in this shift selection strategy. Likewise the Conditions under which the noise excitation frequencies can be influenced, are also applicable 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 2 g and 4 g it can be seen that the Engage the IPSI coupling with the shaft, which engages with one gear is optimized, the transmission rattle suppression, since there is the possibility that Increase the inertia of the transmission and thereby the resonance frequency to reduce the transmission.

With this gear selection strategy, the three-mass model, which in. The effectiveness of the intervention to test the coupling to the IPSI shaft which communicates with a gear engaged, are shown Fig. 58, is used. 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 to influence the 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, be avoided.

During torque transmission in a dry clutch system a certain amount of energy from the clutch in the form of thermal Energy absorbed. This thermal energy can be high Coupling temperatures cause the functionality of the components limited or permanent damage to the clutch; especially when the maximum temperatures at the Friction surface from 350 to 380 ° Celsius. Strategies are becoming Used in automated manual transmissions (ASG) to address these situations to react at high temperatures. With these strategies, the Clutch uses a fan. This creates a convection, so that a hot clutch disc and the clutch itself cooled down can be. There are several options for a dual clutch transmission (DSG) available to respond to these high temperature situations and one implement 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 gearbox, 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 means can cool them down.

This strategy can be used preferentially if the vehicle is with moves at a constant speed and z. B. clutch 1 is engaged is because 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 arises on the hot clutch.

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

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

A fourth strategy is also envisaged, which involves changing the Provides switching strategies for the clutches. This can aim to provide fast torque ramps for the two clutches in order to Reduce switching time. In addition, the engine torque can be faster Be controlled in a way. As a result, driving comfort is reduced, however the thermal energy at the couplings is reduced. This strategy can help every moment during the driving cycle.

A fifth strategy is also proposed, instead of Shift strategy of a parallel gearbox with two activatable in parallel Couplings using a strategy of an automated manual transmission (ASG) Torque interruption is used during the circuits. The Vehicle can be moved as if it had only one clutch. This Strategy can be used at any moment during the driving cycle lake become.

Furthermore, a sixth strategy is provided, which it will use during the Starting allows torque to be parallel through both clutches transfer. This reduces the thermal load between the two clutches divided, which reduces the thermal energy. The time available in which the vehicle is in a steady state with respect to the Gradient is kept, can be increased in an advantageous manner. This Strategy can be particularly useful when the vehicle is started several times an incline can be used. This strategy can also be provided when the driver changes the vehicle position by slipping on the clutch and by the applied moment by pressing the accelerator pedal at one Wants to hold value.

A further embodiment of the present invention follows described which support synchronization during the Preselection of a clutch provides, 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) are switched. This additional gear change can be done with a automated manual transmission (ASG) can be compared. The additional Gear changes can, compared to an automated manual transmission, one cause higher wear of the 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. As a result, the speed of the non-activated input shaft is preferably increased to the engine speed. 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 the engine speed, Nlps_ily / i2y speed of the IPS shaft 1/2, TrqClAct_i1m the current clutch torque of the IPS 1 shaft, cg_so_drvdelay the target gear for driving, gc_bo_act_i1m / i2m the current gear for 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 depend on various parameters such as B. the vehicle speed, the Accelerator pedal position, engine speed or the like. It exists also the possibility that other parameters are taken into account.

The following are requirements that can limit the gear selection functionality:
In a normal driving situation in particular, a higher gear should always be selected. In downshifts, the time to increase the engine speed to a suitable speed is longer than the shift time in the transmission.

Furthermore, it should be considered that there are no problems with a kick down occur with the increasing time for the engine speed, since this always is longer than the switching time. Accordingly, special treatment is one Kickdowns with a gear selection are not required.

Furthermore, uphill and downhill detection is normal automated manual transmissions (ASG) also available with a gear selector. When downhill and engine braking (caster), especially at negative drive torque, there is no time expectation possible and therefore gear selection cannot be used during 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 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 the preset is used.

Furthermore, for handling a tip-in (kick down), the Consider acceleration position or movement and a suitable one Select tip-in gear preselection before the shift illustration shows the same tip-in recognizes.

Furthermore, in the event of full braking (panic braking) to avoid possible problems with full braking that are provided in this Situation is selected as gear selection the neutral state:

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

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 rocking function, namely that Rock the vehicle forward and backward to come out from a trench, provided that z. B. in the winter program or the like. The second gear and the reverse gear is set to the Free rocking function. In this way, the Powertrain between reverse gear and second gear to Free rocking or free swinging can be switched so that the Vehicle out of the obstacle, e.g. B. can be removed from a trench.

Furthermore, it can be provided that the vehicle acceleration in an overview or is stored in a memory, so that when the Vehicle speed rapidly due to braking or acceleration is changed, this is recorded or taken into account when selecting the gear becomes. This can e.g. B. can be realized when the vehicle speed, which is used for the decision of the gear selection, one predetermined estimated speed based on the measured Acceleration, is. In this case it is also possible Detect gear change changes more quickly.

It is also possible to go uphill and at Vehicle speeds that are below a certain limit, to select 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 that the gear preselection for parallel gearboxes is canceled. Further can be used features for an electronic manual transmission such be implemented when using even for a parallel gearbox is advantageous. For example, in the electronic transmission Neutral state on the inactive shaft at standstill can be selected to the Recharge battery.

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

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 it may happen with the aforementioned types of gearbox that the vehicle can roll back when starting on a mountain. This may be because the creep torque of the clutch is only built up when the driver Brake released.

• 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 in Can set movement when the driver is still on the brake pedal a correlation with the falling brake pressure can be particularly advantageous. When the vehicle is stopped on an incline by the driver with the brake , it can be assumed that the brake pressure threshold has been exceeded. This can already be done when the brake is released Creep can be activated as a starting aid. There are other possible ones Dependencies when considering the creep torque build-up 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 even with light braking for example in the plane, do not let it stop. Furthermore, a Creeping against the brake increases the energy input into the clutch. As a result, idle fuel consumption may increase and disruptive vibrations can possibly occur in the vehicle. Furthermore, it should be provided that the known tactile strategy, which works at a low torque level, preferably with the invention Creep strategy can be combined.

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 naturally used and preferred in other driving situations in automated manual transmissions (ASG) and in parallel manual transmissions (PSG) can be used.

A further embodiment of the present invention follows described which a switching strategy, in particular for Parallel shift transmission with upshifts with wrongly selected Finish line, suggests.

It has been shown that it is particularly useful in parallel gearboxes it can happen that at the beginning of the shift the wrong gear on the Target wave has been launched. For a successful overlap circuit the target gear must then be changed first. Especially at Train downshifts can be used to get the engine on that time Accelerate target speed. The vehicle is like this when shifting up accelerates for a long time until the target gear is engaged. Especially at manual activation by the driver can result in a loss of comfort come.

Accordingly, it is proposed in the switching strategy according to the invention that by reducing the engine torque preferred at the beginning of the Upshifting the driver can give the impression of being already a circuit can be carried out. This can be the moment of the engine preferably be lowered to a maximum, so that the driver Acceleration experienced by the driver's desired torque in the target gear would have. The only difference for the driver is that the Synchronization of the engine speed to the target gear later than with one Gear change with 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 at Full load switching the danger in the range of the rev limiter Coming internal combustion engine 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.

With possible strategies for the overlap circuit with one Parallel shift transmission can be provided that the internal combustion engine the vehicle is used to correct inaccuracies in the moments balance components involved. For example, for the Overlapping two slipping clutches can be used. This can in a first phase, the clutch of the old gear slipped become. So far, this has been made possible by the internal combustion engine, which is the Hatches until the end of the overlap.

The problem is that the internal combustion engine is not always as precise and above all can not be controlled so quickly that in a relatively short time Time inaccuracies in the overlap can be compensated for.

Therefore, it is a task of the strategy according to the invention, a possibility to specify where no engine intervention is required, and yet this Strategy is robust against torque errors of the clutches.

In the proposed strategy, the part of the overlap is preferred takes into account up to which the clutch of the target gear takes the moment has completely taken over. Accordingly, the core of the invention Strategy in that the coupling of the target gear with uninfluenced Engine torque the slip on the clutch of the old gear as possible keeps constant. This can preferably be done by controlling the target clutch a slip control during the overlap, for example with a simple PID controller or the like can be superimposed. It is also possible, that the regulation with the opening clutch and / or with both Couplings is realized.

It is particularly advantageous in the strategy according to the invention that the Coupling generally quicker on target requirements than that Internal combustion engine reacts. Furthermore, the couplings are in contrast to Internal combustion engine (full load) not restricted at high torques. Of Furthermore, the strategy according to the invention makes it possible for a stable Regulation never affects the downforce more than the driver it is asked for. Furthermore, there is none in the strategy according to the invention 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 lie significantly 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, sonderel 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, 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 be 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} - f (t)
M _{Kupplung_neu} = f (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 especially with upshifts or the new gear Downshifts are used. The speed change of the Input shafts can be due to the clutch torque applied Counter moments from the output, such as driving resistance, Slope or the like, do not flow back into the controller, however but basically flow in the same direction, d. H. the slip can be reduced if the clutch of the target gear is closed further.

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 in the representations in FIG. 74, in particular the course delta_n_soll. 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 the second clutch can only be closed after the slip reserve has been reached become. The build-up of the hatch reserve can be based on a time-dependent Structure of the target slip are regulated so that the clutch of the Target starts relatively early 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 The strategy according to the invention can provide that in the case of train return or Thrust upshifts continue the engine speed to the target speed of the brought new gear, which corresponds to phase 2, before with the Overlap is started.

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 Can hold 200 revolutions / min. The use of the slip control while driving with a partial transmission can also be used for slip build-up be used. For this purpose, the slip during the Overlap can be controlled with the clutch of the old gear.

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 shift transmissions (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 if the transmission control fails, a general Opening the clutches is required. Furthermore, at least one possibility for emergency opening of the clutches or for stopping the vehicle be provided. It can also be a hydraulic clutch actuation and pressed clutches are used. Overall, there should be a way be given, which makes it possible to open the clutches in emergency operation. For example, this can be done by means of an electromotive clutch actuator be 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 running valves, for example one per clutch, the fluid from the Pressure space of the hydrostatic disengagement system in a pressure-free space, For example, in a storage container, escape, and the couplings can then be opened without force.

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

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 become. This can be an exclusive mechanical actuation or an additional electrical control of the emergency valve can be provided, so that a manual emergency opening of the couplings, for example Towing the vehicle is possible in an advantageous manner.

The aforementioned variants with regard to the arrangement and the actuation at least one emergency valve can also be suitably modified and / or can be combined with one another as required.

The emergency running behavior proposed according to the invention can be used in particular for Parallel gearboxes can be used advantageously because there, unlike a so-called add-on automated manual transmission (ASG) that Roll-away protection is not guaranteed by a parking lock, and thus a Opening the clutch in this situation is not critical to safety. It is also possible that the 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 in particular Thrust downshifts in a parallel transmission provides.

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 but can on a slippery road, especially in winter, to detach the Drive wheels on the road and thus cause dangerous driving situations.

Accordingly, a compromise is made in the parameterization according to the invention between saving fuel consumption and driving safety.

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 proposed in particular for the starting clutch of a double clutch becomes.

It has been shown that, especially with repeated travel and stopping on Berg especially the starting clutch is thermally significantly stressed.

Accordingly, one possibility is given, the capacity of the second clutch in a double clutch transmission to increase the thermal Resilience of the entire system to use.

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 Coupling 2 at the total torque can be controlled and / or Control strategies are divided so that both clutches Reach the maximum permissible coupling temperature almost simultaneously. For example, after falling below the limit temperature mentioned above the original starting strategy was used again due to a cooling phase become. There are other modifications to the one presented here Approach strategy possible to the thermal load of the double clutch further decrease.

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. Further advantages and advantageous configurations result from the subclaims and the drawings described below.

Show it:

Fig. 1 is a flow diagram of a switching strategy according to the invention; and

Fig. 2 different courses of engine and transmission sizes during a gear change according to the proposed shift strategy.

In Fig. 1 is a flow diagram of a possible embodiment of the shifting strategy according to the invention. At the beginning it is checked whether a gear preselection is desired for a gear change during a downshift, however this gear change must not yet be active. If these requirements are not met, the switching strategy according to the invention is ended.

If the above requirements are met, another will Step checked whether the engine speed NEng_y between the target speed of the Transmission input shaft and the speed of the transmission input shaft before the Gear selection. If not, the shift strategy according to the invention completed.

Otherwise, the clutch is not during the neutral phase torque transmitting transmission input shaft via a defined Ramp function closed to a defined value to speed Align Nlps_i1y of the transmission input shaft to the engine speed NEng_y. Only then is the clutch opened and the synchronization of the preselected gear continued and then the inventive Shift strategy ended.

In FIG. 2, curves of various engine and transmission variables during a downshift, are shown in accordance with the proposed shift strategy. With NEng_y (1) the engine speed, with Nlps_i1y (2) or Nlps_i2y (3) the speed of the non-torque-transmitting transmission input shaft or the torque-transmitting transmission input shaft, with TrqClAct_i1m (4) the current clutch torque of the non-torque-transmitting transmission input shaft, with cg_so_drvdelay (cg_so_drvdelay ) the target gear for driving and cg_bo_act_i1m (6) or cg_bo_act_i2m denotes the current engaged gear for the non-torque transmitting transmission input shaft or the torque transmitting transmission input shaft.

By closing the clutch during the neutral phase and with it associated increase in speed Nlps_i1y of the non-torque transmitting Gearbox input shaft up to engine speed NEng_y becomes too synchronizing speed difference significantly reduced. In comparison with known shift strategies can thus the synchronous speed at shift strategy according to the invention by more than half to about 500 rpm (rpm) can be reduced.

The subsequent synchronization, which is indicated by an arrow in FIG. 2, can be carried out more quickly and with less wear due to the lower speed difference.

Reduction of wear during synchronization during gear Preselection for upshifts can be done by estimating the speed of the Transmission input shaft can be realized during the neutral phase. The Speed behavior can then be modulated accordingly. The Speed model can depend on factors such. B. the friction (i.e. influence of Temperature) or the age of the gear unit. When using a speed sensor on the transmission input shaft can these factors be considered appropriately. It is also possible to make other modifications to be provided in the switching strategy presented in order to further improve this.

The claims filed with the application are Proposed wording without prejudice for obtaining further patent protection. The The applicant reserves the right to add more, so far only in the description and / or Drawings combinations of features disclosed.

Relationships used in subclaims point to the others Training the subject of the main claim by the features of respective subclaim; they are not a waiver of attainment an independent, objective protection for the combinations of features to understand the related subclaims.

Since the subjects of the subclaims with regard to the prior art reserves the right to make its own and independent inventions on the priority day Applicant before becoming the subject of independent claims or To make divisional declarations. You can continue to work independently Inventions contain one of the objects of the previous Have independent claims independent design.

The exemplary embodiments are not intended to limit the invention understand. Rather, there are numerous within the scope of the present disclosure Changes and modifications possible, especially 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 as well as the claims described features and elements contained in the drawings or process steps for the person skilled in the art with a view to solving the Task can be removed and combined to create a new one Subject or new process steps or process step sequences lead, also insofar as they relate to manufacturing, testing and working processes.

Claims (10)

1. Shift strategy for a transmission, in particular for a dual clutch transmission, of a vehicle in which a target gear is selected on a transmission input shaft that does not transmit torque, characterized in that the speed (Nlps_i1y) of the transmission input shaft that does not transmit torque is increased during the gear pre-grinding. 2. Shift strategy according to claim 1, characterized in that the clutch the transmission input shaft not transmitting torque during the Neutral phase of both gear preselection is closed. 3. switching strategy according to claim 1 or 2, characterized in that the Speed (Nlps_i1y) of the transmission input shaft that does not transmit torque is adjusted to the engine speed (NEng_y) when the engine speed (NEng_y) between the speed (Nlps_i1y) of the non-torque transmitting Transmission input shaft before and after the gear pre-grinding of the gear change lies. 4. switching strategy according to claim 3, characterized in that the preselected gear after increasing the speed (Nlps_i1y) of not torque-transmitting transmission input shaft is synchronized. 5. Switching strategy according to one of the preceding claims, characterized characterized that it is checked at the beginning whether a prefix for a desired gear change is activated and whether the engine speed (NEng_y) between the speed (Nlps_i1y) of the transmission input shaft before and after the gear change. 6. switching strategy according to claim 5, characterized in that in the Neutral phase the clutch of the non-torque transmitting Transmission input shaft is closed at the speed (Nlps_i1y) of the Align the transmission input shaft to the engine speed (NEng_y). 7. Shift strategy according to claim 6, characterized in that the clutch over a predetermined torque curve to a defined value is closed. 8. switching strategy according to claim 6 or 7, characterized in that after When the engine speed (NEng_y) is reached, the clutch opens and the selected gear is synchronized. 9. Switching strategy according to one of the preceding claims, characterized characterized that it is used in a downshift. 10. Transmission, in particular double clutch transmission, for a vehicle at least two transmission input shafts, each with a clutch is assigned, in particular for performing the switching strategy according to a of claims 1 to 9, characterized in that at least one Control device for increasing the speed of the not torque transmitting transmission input shaft during the preselection of a Ganges is provided.