SE512198C2 - Clutch torque transfer system control method in e.g. motor vehicle - Google Patents

Abstract

The method involves controlling a torque transfer system with or without load distribution. The clutch torque which is transferable from a drive side to an output side is used as the control value. The control value is calculated in dependence on a drive torque. A sensor system detects measured values and a central control or computer unit determines or calculates the torque in dependence on the relevant operation state of the system. The sensor and controller are connected together. The torque capacity is used as a control value to give pre-determined tolerance about a certain slip limit. The torque transfer system includes a clutch, gear box and setting member. The transferable torque is adapted and controlled. Deviation from the ideal state are compensated long term through correction.

Description

i.;.; ..; tat: ... ia-l-uallt-í-ih fi llli l i -.-.

HI ll 'ztïlläl 1 10 15 20 25 30 512 198 2 for a bridging coupling a special friction device. This means that between the input part and the output part of the torsional mandrel pair only kra fl magasiri are arranged, which can be set in a relative rotation between these parts.

By the determination according to the invention of the torque transmission capacity in the torsion damper, the oscillations occurring in the partial load area, i.e. in the area with drive torques in the order of between 10 and 60% or between 25 and 50% of the nominal torque, can very well be damped.

It may be particularly suitable if the damper allows an angle of rotation which is relatively small in relation to the hitherto known angles of rotation of dampers for transducer bridging couplings. The angle of rotation can be in the order of: t2 to = t8 °, preferably in the order of 13 to ió ”. The total angle of rotation of the damper, thus the total angle of rotation for both directions of rotation, thus amounts to 4 to 16 °, preferably 6 to 12 °. Due to the relatively small angle of rotation of a torsion damper for bridging couplings designed in accordance with the invention, it can be guaranteed that during load change, i.e. at transition fi from push drive d to traction drive and vice versa, the deflections in the damper are kept low, whereby a swiveling of the driveline can be limited. Advantageously, torque shocks or torque proportions in these shocks, which are above the torque of the torsion damper, can be damped or filtered through slippage or sliding of the bridging clutch, so that they are at least substantially kept away from the driveline or gearbox.

For most cases of use, it may be suitable if the pair has a torsional rigidity which is in the order of magnitude between 7 and 30 Nm / °, preferably between 8 and 15 Nm / °. For many applications, however, this torsional rigidity can also be chosen smaller or larger. For most applications, the bridging coupling or torsion damper can be designed in such a way that it has a stop torque in the order of between 30 and 90 Nm, preferably in the order of between 40 and 70 Nm. For slightly motorized vehicles, however, the impact torque can also be designed lower. It may also be necessary to set the impact torque higher in the case of heavily motorized vehicles with a relatively high weight.

The drive system according to the invention can advantageously be used in connection with a method for controlling a slip-coupled bridging clutch depending on the current torque, which ensures a control oriented according to energy-technical and power considerations at least in all forward gears in a gearbox.

However, the drive system according to the invention may also find use in gearbox adjustments or regulations which should be fully open in the first and / or second. / Users \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 512 198 3 or the second off-gear stage or stages of the bridging coupling.

With the aid of the accompanying drawings, the following is an understanding of and details of the control method for use in internal combustion engine motor vehicles and a torque transmission system with a torque converter and a parallel bridging coupling as well as advantages obtained by this steering method. . There fi g shows. 1 is a schematic view of a torque transmission system with a current converter and a friction coupling arranged parallel thereto and the converter bridging, fi g. lb an illustration of the motor torque as a function of the motor speed, fi g. lc the losses in the converter in iron motion, fi g. Id loss effect as a function of vehicle speed, fi g. le the influence of the transducer design on the traction, fi g. 2 a half-sectional view of it against the schematic view in fi g. la corresponding torque transmission system with a converter and a lock-up coupling as well as with a diagram for the associated pressure part control, fi g. 3 shows in a diagram the division of the motor torque into one of the torque transducers and a torque which can be exerted by the bridging clutch in dependence on the slimming occurring in the transducer and the bridging friction clutch, fi g. 4 the engine speed and the difference speed in the converter depending on the time when accelerating a motor vehicle with a changeover process at according to the invention morning-controlled converter bridging, fi g. 5 corresponding to fi g. 4 the driven torque over time when accelerating a vehicle with a shifting process during torque-controlled transducer bridging, fi g. 6 in a diagram as in fi g. 4 the speed ratio at acceleration and at slip-controlled cross-bridging, fi g. 7 corresponding to fi g. 6 in a diagram similar to fi g. 5 the driven torque over time at the acceleration at slimming-regulated transducer bridging, fi g. 8 in a diagram similar to that in fi g. 4 and 6 the speed ratio at the acceleration with transducer bridging opened during a shifting process and closed again after the shifting process, fi g. 9 corresponding to fi g. 8 in a diagram similar to that in fi g. Figs. , fi g. lla a torque transmission system with a hydrodynamic converter bridging fi coupling, fi g. 11b and c a temperature drop in converter bridging connections, över g. lld and lle the maximum surface pressure at Oz \ users \ ya \ dok \ word-dok \ l03289001 .doc U .q i .. “i: c iriiiintzlr: l 10 15 20 25 30 512 198 4 converter bridging couplings, fi g. 12 in a diagram the division of the motor torque into one of the torque converter and one torque transferable by the bridging clutch in dependence on the sliming occurring in the transducer and in the bridging friction clutch, fi g. 13 in a primary characteristic field for a "hard" designed converter the pump torque over the pump speed with the speed ratio turbine / purnp as parameter, fi g. 14 in a secondary characteristic field the turbine torque for the "hard" designed converter over the turbine speed, fi g. The driven characteristic field of a converter commonly designed "hard", fi g. 16 in a diagram similar to fi g. Fig. 13 shows the primary characteristic field for a "softly designed" converter with the pump torque over the pump speed and the speed ratio turbine / purnp as parameter, Fig. 17 in a section characteristic field for the one according to fi g. 16 "softly" designed the transducer turbine torque over the turbine speed, fi g. 18 by means of superimposed secondary characteristic fields according to fi g. 16 and 17 the usable converter area which can be added in the "soft" design of the converter, fi g. 19 in a diagram similar to fi g. The driven characteristic field of the one according to fi g. 18 softly invented converter, fi g. 20 in fl the surface of the transducer invention on the losses, fi g. 21 torque turns of engine and gearbox, fi g. 22 a load shift ratio with conventional fi damper, fi g. 23 effect of slippage on the oscillation ratio, fi g. 24 load change ratio with and without sliming, fi g. 25 the required slimming without and with mini-torsion door pairs, fi g. 26 a "bubble scheme" for the interaction of slippage, mini-torsion columns, cone design and adaptive control, fi g. Fig. 28 shows by means of various diagrams examples of lcriteria for the determination of the kme-falctom or km., - characteristic field, Fig. 28 a fl fate diagram for a method for torque control with adaptation of a transducer bridging coupling, fi g. 29 different diagrams for explaining the additive and multiplicative correction factors, which are useful for the torque control of a clutch, fi g. A circuit diagram of a torque control for a clutch, fi g. 31 a torque transmission system with a hydrodynamic transducer bridging free coupling, fi g. 32 and 33 details of the torsional vibration damper according to fi g. 24, fi g. 34 a possible torsional characteristic line for the damper in a lock-up coupling, fi g. In a diagram the driven characteristic field of a "soft" shaped converter, fi g. 36 the sum frequency of the slimming, fi g. 37 the losses in a mountain run, fi g. 38 a schematic illustration of the control procedure, and fi g. 39 a traction diagram.

That i fi g. 1 and 2, the torque transmission system 10 includes a torque converter 11 and a flow pressure means operable bridging coupling 12, which is connected in parallel with the torque converter. The torque transmission system is operatively connected to an internal combustion engine (not shown) via a shaft only indicated by a shaft 13 and is in turn on the drive side via a driven shaft in turn: 0 15 \ users \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 512 198 14 in drive connection with an automatic transmission connected to the driveline, which is also not shown.

As the schematic half-sectional view of the torque transmission system 10 shows in fi g. 2 in connection with the torque control diagram, the torque converter 11 is a standard disturbance converter. This flow converter consists of a converter cover 16 connected to the driven side of an internal combustion engine, a pump wheel 17 forming together with the converter cover, a turbine wheel 18 connected in turn via a drive shaft to the automatic gearbox (not shown) as well as a guide wheel arranged between pump and turbine wheels. 19. The transducer bridging clutch coupling 12 is disposed between the turbine wheel 18 and the transducer cover 16 and has a coupling disc 20 rotatably connected to the transducer of the transducer wheel, the friction lining 21 cooperating with a counter surface 22 of the transducer cover 16. The friction clutch has a further and a front chamber 25 facing the transducer cover 16.

The converter is a flow gear with pump 17, turbine 18 and articulated wheels 19. Without slippage, it does not transmit any torque. At constant input speed, the following applies: The higher the slip, the higher the torque. Fig. 1b shows this connection for a braked drive, the solid line in fi g. 1b illustrates a conventional transducer and the dashed line represents a soft transducer.

Soft is called a converter, which at the same moment has a higher slip, ie. that it transmits lower torques at the same slip. The softer converter thus puts less resistance to the motor. If the driver requires a higher torque, the softer converter builds up higher speed differences.

The higher speed differences provide the so-called the rubber band effect, the vehicle reacts delayed to gas, ie. it does not react directly to gas.

However, most urn tests begin with a cold phase. If the engine more easily reaches higher speeds in this phase, it will heat up faster and the emissions will be significantly better.

At a given motor speed, the soft converter opposes the motor with a lower torque. If the vehicle is stationary at engine speed, the engine must overcome the transducer torque, and thus the losses at the vehicle standstill will be according to fi g. lc with soft converter lower. I fi g. 1c shows, by way of example, the dry losses at vehicle standstill shown for a soft converter by 0.95, the losses for a conventional converter being indicated by 1.6. 0: \ users \ ya \ dok \ word-dok \ 103289001 .doc 1.1. ,. .. ll ..

U lll »l: j läal: iii-iii 10 15 20 25 30 -512 198 6 With extended torque, e.g. at a given vehicle speed at a given incline, the slip at soft transducer becomes larger, such as fi g. lb shows, and thus the losses also become larger (fi g. ld).

In contrast to the clutch, a torque converter can increase the torque. This torque conversion is at the same diameter with a softly designed transducer higher. If you increase the conversion by the same diameter, the converter becomes softer. The higher conversion leads to an increase in the traction crane (and thus the acceleration capacity) (ñg. Le).

The transducer 11 is supplied in a known manner via a impeller side of the conduit 30 opening into the transducer housing from a source of pressure medium (not shown) with flow pressure part, the pressure control taking place via a control valve 31, which in turn is controlled by a control element 32.

The flow pressure medium, on the other hand, is diverted via a line (not shown) to a cooler 33 only indicated.

In addition to actuating the turbine wheel 18, the pressure of the flow pressure coil on the current side of the impeller 17 also acts in the rear chamber 24 of the friction clutch 12, acts on the pressure plate 20 and presses it against the counter surface 22 of the transducer cover 16 cooperating with its friction lining. Slimming in all drive areas takes place by the more or less large gap between the friction lining 21 of the coupling disc 20 and the cooperating counter surface 22 of the transducer cover 16 due to the slippage, a throttled interference pressure component effect of the tension between the coupling disc 20 and the transducer cover 16 extends. the inner chamber 25 is so controllable by means of a valve connected to this chamber via a line 34 that an adjustable differential pressure effective between the rear and front chambers determines the torque transmitted by the friction clutch 12.

With regard to the parallel arrangement of the transducer 11 and the later bridging friction clutch 12, the motor torque is equal to the sum of the transducer or impeller and the torques transmitted by the clutch, thus Mmom = Mxopplmg + Mrunlphjul- The gear torque, as far as the sum system is disregarded the torques transmitted by the converter, or turbine wheel, thus Mvaxentaaa = Mkopplng + Mrmbinhiul or Mlwvvlins + (MPwnvhlul X Conversion) - The division of the motor torque into one of the transducer and one of the bridging Eriktionskopplirigen transferable. 3 depending on the slimming. It appears that with increasing sliming the proportion of the motor torque transmitted by the transducer increases and consequently the torque transmitted by the clutch decreases. 0: \ users \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 512 198 7 In the control procedure according to the invention, the slip is admittedly not regulated, but depending on the operating state of the motor, the proportion of motor torque to be transmitted by the friction clutch and of a computer unit, for example a microprocessor, the differential pressure necessary for the transmission of the predetermined torque is set in the friction clutch. The grinding is then obtained automatically.

I fi g. 4 illustrates over time the engine speed 40 and the differential speed 41 of the converter when accelerating and when shifting from, for example, the second gear to the third gear. As a result of the acceleration, the engine speed in the second gear first increases up to the triggering of the shifting process and decreases during the 42-shifting shifting torque.

The difference speed of the converter, on the other hand, first remains constant, then increases sharply during the changeover process. If the shift from the second to the third gear drops at 43 the engine speed and the differential speed in the converter, namely the latter after a slight shift to a constant value at a higher level than before the shift torque. This is illustrated in fi g. 4. On the other hand, the engine speed increases slightly with respect to the assumed acceleration in the third gear. It is obvious that at no time does the converter cross-bridge fi -connectively fasten. Instead, run in all drift rotors with slippage.

Of particular interest is the corresponding to fi g. 4 i fi g. Over time illustrated the driving torque 44, which drops sharply at the beginning of the shifting process, then during the phase with great slippage in view of the torque increase thus obtained rises steeply and falls back at the end of the shifting process without appreciable and otherwise immediately decaying afterward oscillation 46 in driveline oscillation. value corresponding to the third gear.

Figs. 6 and 7 show that in the case of slip-regulated transducer bridging, the conditions during the changeover process are completely different. Also fi g. 6 and 7 refer to the switching from the second to the third gear in an accelerating vehicle.

Such as fi g. 6 shows, in the second gear, the engine speed 40 'increases until the triggering of the shifting process at 42', while the differential speed 41 'in the converter and thus the occurring slip remains constant. At the beginning of the shift ear race at 42 ', the motor speed decreases, while the differential speeds in the converter increase. After switching to the third gear, the motor speed and the differential speed in the converter decrease again.

As with slip-regulated transducer bridging, the aim is to keep the differential speed in the transducer layer constant even during the shifting process, the shifting process takes longer than with torque-controlled transducer bridging, because the transducer's turbine cannot yield. At the end of the shifting process, the friction clutch occurs at 47 the addition of the bridging O: \ users \ ya \ dok \ word-dok \ 10328900 l .doc tili i »ii ml ii 10 15 20 25 30 1512 198 s, as the slip control can only work when a deviation has been set, and even then only with a speed limited by the adjusting elements and the control stability. Finally set, such as fi g. 6 shows, after the time-long changeover process, the slip 41 'again at the level prevailing before the changeover process. Even with slip-regulated transducer bridging, the driving torque 44 'drops sharply at the beginning of the shifting process, only to rise steeply in the same way as with torque-regulated transducer bridging and at the end of the actual shifting process with noticeable and only e and er and after decaying oscillations. third gear corresponding value.

It can be seen that with wear-controlled transducer bridging, the speed gradient and the speed difference at the end of the weighing are very large. This is the reason why at the end of the shift dry run the friction clutch is at and with regard to the then completely bridged converter, the above-mentioned e-oscillations in the driveline are set. Also fi g. 8 and 9 show analogously with fi g. 4 and 5 show the acceleration of a vehicle with a shifting process, whereby the transducer bridge opens during the shifting process, but which after the shifting to a higher gear is closed again.

Av fi g. 8 shows that until the triggering of a gear change at 42 "the motor speed 40" rises, while on the other hand the differential speed 41 "in the converter decreases slightly. During the actual gear cycle the motor speed corresponding to the gear shifting to a higher gear decreases. The wavelength variation 41 in the converter increases. of the changeover process, and then decreases again at the end of the changeover and after the expiration of a predetermined time due to the closing of the transducer bridge at 48 it goes towards zero. At the driven torque the conditions are at first quite similar to those of the inventive torque control of the transducer bridging, but towards the rapidly decaying overshoot 46 "directly at the end of the shifting process it comes at the lowering of the differential speed to zero, i.e. at the complete closing of the transducer. to significant shift impacts with only slowly decaying oscillations 49 in the driveline.

Such as the comparison of the control concept according to the invention with the aid of fi g. 4 and 5 with the slip-regulated transducer bridge according to fi g. 6 and 7 and the control concept with during the changeover process opened, e fi is the changeover again closed converter bridging according to fi g. 8 and 9 have shown, significantly less shift shocks occur in the torque-controlled transducer bridging according to the invention than in the other control concepts. This is due to the fact that during the changeover the converter bridge, which is still driven with a predetermined slip, gives way and the diner speed can rise to a corresponding degree.

O: \ users \ ya \ dok \ word-dok \ 103289001 .døc 10 15 20 25 30 512 198 9 In the diagram according to fi g. 10, the curve 50 shows the course of the pressure difference Ap in the lock-up coupling acting as a function of time. Starting from the initial differential pressure Apsm, the pressure difference first rises steeply over time, as the nearest Apgm key 51 shows, to then gradually weaken in pitch and finally asyrnptotically approach a set pressure difference indicated by the dashed line 52. This is done by stepwise approximation, by starting from a differential pressure Apn at a time tn according to the equation given in claim 29 after a sensing interval vidt at a time tm determines the differential pressure denpm, the gradient required for the time difference eftert is calculated and this gradient is set by the hydraulic system and finally this step sequence is repeated continuously until the setpoint for the pressure difference indicated by the dashed line 52 is reached.

At that as an embodiment in fi g. The illustrated torque transmission system 60 is a hydrodynamic torque converter 61 with a bridging coupling 62 and a damper unit 63 acting between the torque converter and the bridging coupling.

The torque converter 61 comprises a pump wheel 65 with a combustion engine (not shown) in a rotationally fixed drive connection, a turbine wheel 67 operatively connected to a drive-side hub 66, a guide wheel 68 connected to the turbine wheel and a rotating wheel attached to the turbine wheel and fitted to the turbine wheel.

The transducer cover 70 is rotatably connected to the impeller 65 and by means of its drive dry connection to the internal combustion engine via drive pins 71, 72 projecting on the side facing away from the impeller, on which a pivot wheel (not shown) of the internal combustion engine is accommodated.

Arranged between the turbine wheel 67 and the transducer cover 70 is a ring piston 74 centrally arranged with the axis of rotation of the transducer, in which case it is a plate-shaped part. This annular piston is radially internally with a sealing hub 75 received on a receiving hub 76 continuing from the hub part 66 connected to the turbine wheel in a rotationally fixed manner and radially externally designed as a coupling friction disc 78 with a conical friction surface 79.

The tapered conical surface 79 of the friction plate 78 on the ring piston 74 cooperates with a corresponding conically shaped counter-friction surface 80 of the transducer cover 70 rotatably connected to the impeller 65. The cones of the cooperating friction surfaces open according to the the side facing away from the turbine wheel 65. With respect to this design, there is formed between the peripheral portion of the turbine wheel 67 and the conically shaped clutch friction disc 78 of the design. the annular piston radially outwardly from the transducer cover 70 enclosed wedge-shaped annular space.

The advantages of the cone design 79 are mainly obtained by the stiffer construction and the larger friction surface: * The divertable loss effect is noticeably greater, and with equal loss effect the maximum oil temperature is lower. It helps to solve problem 5 (control speed), problem 2 (engine pressure) and the rest of problem 3 (control parameter problems).

* The coating load is reduced by the more uniform surface pressing.

* The transferable torque becomes higher. Many single-disc bridge couplings today are already at the limit of their capacity. Due to the flow effects through the cooling oil stream, the transferable torque drops again.

* The weight and mass moment of inertia are lower, as due to the stiffer construction the plate can be chosen thinner.

* The converter size becomes smaller and the controllability thus better.

For the life of the oil, the locally occurring maximum temperature plays an important role. The temperature is kept low by a coating cooling, see fi g. l lb, 1 1c.

Due to the coating cooling, the bearing surface becomes smaller. However, this is not a problem with the cone as the coating pressure is uniform (fi g. 11d and fi g. Ile). In addition, the transferable torque drops through the oil stream, here the reinforcing effect of the cone 79 helps.

In this wedge-shaped annular space, the derailleur unit 63 is accommodated with annularly shaped damper-spring elements 82, which in the circumferential direction on one side support against damper drive parts 83 rotatably connected to the ring piston 74 and support with their other ends against rotor-driven damper-driven parts 84 connected to the turbine wheel 67.

The damper drive parts 83 are leaf-spring-like-shaped, and are arranged on the side of the ring piston 74 facing the turbine wheel 67 and are rotatably connected thereto in the area between the ring-piston timing hub 75 and the clutch friction disc 78 by means of rivets. On the side facing away from the friction surface 79 of the clutch friction disc 78, the following therethrough drive parts 83 extend outwardly from the contouring of the ring piston 74 and dampen the spring elements 86, 87 surrounding the spring elements 82 as well as a spring element on a gable end supporting carrier 88, 89.

The damper-driven parts 84 are ring segments welded together with the peripheral portion of the turbine wheel 67, from which in the direction of the coupling friction disc 78 the coupling friction plate 78 carries står grooves 90 which protrude, which mediate the support of the damper spring elements 82 towards their other ends. The spring elements are thus accommodated between the carriers 88, 89 of the pair of spring drive parts 83 and the projecting carriers 90 on the damper-driven parts 84.

The damper unit 43 of the converter 41 is preferably adapted for the main driving range, as in fi g. 18 and 19 are indicated in the form of shaded surfaces. Such a damper design, which is shown with consideration of the transducer bridging, and guarantees a significantly better damping with respect to that in this main driving range in future complete torque turns, than would be possible with a damper design designed for a larger driving range. In addition, a particularly compact conversion lamp construction is obtained.

The lock-up coupling shown in the drawing as an exemplary embodiment and explained above has a one-sided pressure chamber 92 between the ring piston 74 and the turbine wheel 67 and a rear pressure chamber 93 between the ring piston and the transducer cover 70. The clutch disc 78 is actuated in its counter-friction surface 80 with the transducer lid 70. of the action of the front pressure chamber 92 with flow medium pressure and the setting of the torque transferable by the sk coupling takes place in dependence on the differential pressure acting between the front pressure cores 93.

The flywheel shown above a flywheel, which by means of the drive pins 71, 72 projecting from the transducer cover 70 on the side facing away from the torque converter, is pivotally connected to the transducer cover, initiates the input torque when the lock-up coupling 62 is opened directly on the impeller. considering the hydraulic fluid flow thereby caused via the turbine wheel 67 to the drive hub 66.

When, on the other hand, the lock-up coupling is completely closed and thus the friction disc 78 of the ring piston 74 cooperates non-slip with the friction surface 80 of the transducer cover 70, a direct mechanical transmission of the input torque initiated to the transducer cover 82 to the turbine wheel 67 via the damper cover drive hub 66 to a driveline operatively connected to a gearbox operatively connected.

If, depending on a differential pressure acting between the front and rear pressure cores 92, 93 of the lock-up coupling, the lock-up coupling operates with slimming, the input torque initiated via the transducer cover 70 is divided depending on the slimming into one of the lock-ups clutch 62 and torque transmitted by the converter 61, such as fi g. 12 schematically shows.

The torque transfer from the lock-up clutch 62 to the rotary wheel 67 and the drive-side hub 66 connected thereto guarantee an effective equalization of non-conformities of the initiated torque. Considering the arrangement of the damping spring elements O: \ users \ ya \ dok \ word-dok \ 103289001 .doc I »-r rniivl r l- fi lllhl .runt r 10 15 20 25 30 512 198 12 82 in the peripheral region between the ring piston 74 fi disc 78 and the turbine wheel 67, the mastery of relatively large fi distances is guaranteed.

According to the invention, the friction clutch 12, 42 can be controlled in such a way that in all forward gears it is at least occasionally at least partially closed. In other words, a slip control of the clutch is also provided in the first or from the first gear, whereby a complete closing can also take place.

However, the conical friction surface of the transducer cover 70 and the friction plate 78 may also, as shown by the closely spaced design line shown and indicated at 70a and 78a, be shaped as cones inclined toward the turbine wheel. However, the damping springs 82 may be located radially further inwards, e.g. over the hub 66.

In commonly designed torque transmission systems, the lock-up clutch, which is fully open in the lower gears, is engaged in the upper gears. In the interest of a good overall efficiency and for the limitation of the heat generated, the converter is "hard" designed.

Fig. 13 shows the primary characteristic field of a "hard" shaped converter with the pump torque over the pump speed and the turbine / pump speed ratio as paired.

I fi g. 13 shows a characteristic field of a drive motor with the motor drive torque plotted over the motor speed corresponding to the turbine speed.

Finally is in fi g. 13 also shows the main driving range, which roughly covers the speed range between 750 to 2000 rpm.

That i fi g. 14 illustrated the secondary characteristic field shows the turbine torque over the turbine speed with indication of the efficiency in the most different power ranges for the hard-formed converter according to the characteristic feature in fi g. 15.

That i fi g. The drive characteristic field shown in Fig. 15, in which the turbine torque of the converter is plotted above the turbine speed, illustrates the converter tube area, in which with increasing speed the turbine torque decreases sharply, as well as the coupling area connecting to the converter area.

Furthermore, again the main driving area shown as a narrowly sloping surface is plotted in the drive characteristic field.

With ordinary torque transmission systems with an interest in a good overall efficiency and for limiting the heat generation in the event of a “hard” converter design, the torque increase decreases sharply with increasing speed. In the middle speed range there is therefore only a small and in the upper speed range finally no torque increase at all.

O: \ users \ ya \ dok \ word ~ dok \ 103289001 .doc 10 15 20 25 512 198, 13 In the primary characteristic field according to fi g. 16, the pump torque is shown above the pump speed with the turbine / purnp speed ratio as a pair for a "soft" designed converter.

The characteristic lines of the "soft" designed converter have for the same parameter as in fi g. 13 a significantly flatter course. The conversion range extends across the center to into the upper speed range. j This leads, such as fi g. l7gvisar, to one in relation to that in fi g. 14 illustrated the secondary field of a hard-formed converter to a sharply widened secondary field. Consequently, in the case of soft conversion lamination, significantly larger acceleration reserves are available, which in many cases makes a shift more than necessary when accelerating.

These acceleration reserves are shown especially in fi g. 18, in which the sectarian field belonging to a hard converter according to fi g. 14 is superimposed on the secondary field of the soft-formed transducer according to fi g. 17. In soft transducer design, the dashed area between the two fi load lines for the two transducers is gained for a torque increase. This also shows it analogously with fi g. 15 i fi g. 19 illustrated the digital field of torque transmission system with soft transducer design. The usable converter area is in relation to the drive characteristic field according to fi g. 15 made larger with the serpentine lying over the dashed line. Incidentally, even in this characteristic field, the main driving area and the area with minimal slip are shown as a narrow sloping surface.

Furthermore, i fi g. 18 plotted driving points 1, 2 and 3. In the case of realized converter designs with a "hard" and "soft" design, the following slip values and efficiencies could be measured in these three points: "hard" converter "soft" converter Sliming Sl (s) Efficiency s (%) Efficiency Tl 'fl Point l 65 0,547 75 0,388 Point 2 40 0,789 60 0,669 Point 3 2 0,980 2 0,980 O: \ uscrs \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 -512 198 14 It can be seen that in the lower and middle speed ranges with "soft" converter design the efficiency compared to the efficiency of a "hard" designed converter is admittedly lower, but there is a noticeable increase in slip and thus an improved torque increase. At operating point 3 in fi g. 18, on the other hand, with hard and soft converter design, slip and efficiency are equal.

Due to the dynamic relationship of hydraulic and mechanical systems, in the case of too rapid an increase in the amount of a torque transmitted by the torque transmission system between transducers and friction clutch influencing parameters, it can produce oscillations of different frequencies by an excessive amount of the jerk or an extension of the iction clutch.

In order to avoid such oscillation generations, a suitable further embodiment of the invention proposes that the setting of an amount calculated from the previously deviating new for a division of the torque to be transmitted between transducer and friction clutch influencing parameter, preferably the diurnal pressure, is delayed according to a function according to a function. .

However, the setting of a new calculated amount from the fi further deviating for a division of the torque to be transmitted between transducer and fi action-coupled actuating parameters can also be delayed according to a function depending on the ditference speed between the driving and the driven side of the torque transmission. an amount calculated from the previously deviated new for a division of the torque to be transmitted between transducer and friction clutch affecting parameters possibly delayed according to a function in dependence on the motor speed gradient.

The loss effect at a given driving condition (weight, rise) decreases with increasing conversion in comparison with a soft converter without higher conversion, but it is generally higher than with a rigid converter, fi g. 20. At high slip, the losses despite at the same time softer converter are not higher than with a stiffer converter, as the higher torque conversion improves efficiency (area A in verk g. 20) Without a bridging coupling, a compromise must be made between standstill losses and acceleration capability (the converter must therefore rather be soft) and between losses in the driving area (here the converter should rather be stiff). Here, the limits are predetermined by the converter physics by O: \ users \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 198 15 512. These have been commissioned in recent years. The transducers so designed are rather rigid.

The significance of the loss effect can be limited by a standard bridging coupling with torsion damper and without slippage. However, there are restrictions. For reasons of comfort (humming, rattling and load shifting), these bridging couplings can only be used at the upper gears and only at higher speeds. Despite this, there are still some comfoitin restrictions.

For the illustration of the problem of humming and rattling is in fi g. 21 shows the oscillation amplitude over the engine speed - generated by a motor arrangement. Depending on the engine speed and hum sensitivity of the vehicle, it is only possible to bridge from a higher speed.

As is known, the motor is driven most of the time at a relatively low speed. Therefore, consumption savings are limited.

An additional problem is the load change ratio and the ratio at in- resp. disconnection of the bridge, fi g. 22. If the driver gives gas in the bridged state, he first receives a jerk swing instead of the desired traction increase. Then the lock-up connection is opened, which in unfavorable cases can even first lead to a short reduction of torque. Only when fi er does the driver receive the desired traction increase. When the lock-up connection is closed, a driveline effect can occur again. Even during the changeover, comfort problems can occur, and therefore the bridge is usually opened before a changeover. In the lower gears, these meanings are strongest, and therefore in ordinary bridging systems are only bridged in the fourth and fifth gears.

The loss effect, which occurs in the first gear during rock clearing, is therefore not reduced by a bridging. At the given cooling effect, this loss effect also limits the permissible softness of the converter.

The "rubber band effect" can not be prevented in many areas and also does not allow the choice of a softer converter.

The transducer bridging system according to the invention concept consists of a sliding fi in bridging coupling in cone shape with mini torque damper, an adaptive control and a soft transducer.

It is a system of slipping. What are the advantages of slipping compared to complete bridging? * Brurnmandet is reduced.

* Load shift ratio is improved.

* The gear quality is significantly improved (completely bridged but not acceptable).

O: \ users \ ya \ dok \ word-dok \ 103289001 .doc .. .__ r_; l _... iii; ill: i 10 15 20 25 30 512 198 16 * The connection quality of the bridging is improved (connection here means that with an open converter the bridging connection is operated, but not unconditionally to slip zero).

By reducing the noise generation, the bridging can be started earlier than with the usual system, see fi g. 23.

The load change ratio and the provision of the traction crane are significantly improved in iron construction with the standard system, see frg. 24. Gives the driver gas, then no jerky oscillation occurs, as the bridging slides through. Through this sliding, the slippage and thus the transducer torque is built up. In addition, there is no torque limitation. The torque increases continuously, through the increasing conversion over the motor torque. It can be bridged earlier, even at the lower gears and at lower speeds.

When there are so many benefits, why are slippery bridges not used everywhere today? Of course, there are also some possible problems with slippage (for explanation fi g. 25): 1. At lower speeds, the slippage necessary for hum avoidance is relatively large, and thereby the loss effect is also large. If the slippage decreases, short-term adhesion occurs, which in many cases causes humming (Ani). 2. Many motors must not be overloaded at high loads.

If it is not permitted to press the motor at high load, there are two possibilities. You increase the engine speed by completely opening the bridge or by letting the bridge slip more strongly. If you let it slip more strongly, higher loss effects occur in the bridging (Ana). 3. Regulating minor slippage is difficult. With "sharp" control parameters, control problems often occur, and if you loosen the parameters, the slimming can decrease noticeably. In many cases, a control has advantages, but even here a swing of the slimming is hardly inevitable. This can lead to persistence (humming risk) or excessive slimming (major losses). 4. The steering is not arbitrarily accurate. The smaller the torque to be set, the more difficult an exact regulation or control becomes. 5. The steering is not arbitrarily fast. In the case of non-stationary processes, the control section needs a set-up time. In these phases, the slimming deviates. Thus, in order to avoid humming, a certain slippage must be reserved. As a result, higher slip values (Ang, Ang) are obtained again. 6. Losses occur on the bridge. A significant problem with slippery couplings is the service life problem. Most of the time, they have been able to cope with the declining loss effects for some time. However, a few thousand kilometers occur, for example. twitching problems. These jerking problems have their cause not least in the damage of the oil - the friction coating is generally still in order. The additives are damaged by local overheating and over time it affects the entire oil. Even at lower dry loss power, the coating must be very well cooled. The safety against local overheating should be as great as possible! To the loss effect by slipping for oscillating fi year coupling and through points 2 and 3 also comes the dry loss effect when coupling and uncoupling the bridging. The lower the speeds and the higher the loads at which the bridging is switched on, the greater the loss effect - especially when a comfortable bridging change is taken into account.

To combat these problems, a very simple torsion damper (which can also only be adapted for partial loads), a cone shape and an adaptive control are useful. The interaction between these system components is shown in the "bubble diagram" (fi g. 26). The bubbles surrounded by thick lines then constitute customer requirements and the dashed bubbles the components in the system according to the invention.

Mirtitorion damper: Advantages of this very simple torsion door damper (see also fi g. 25): * Problem 1 (browning) is thus solvable. The impulses which occur during short-term adhesion are thereby filtered and humming does not occur.

* Problem 3 (control parameter problems) is thus partially solvable. Here, too, short-term adhesion no longer has a negative effect.

* Problem 4 steering accuracy at low torque) is thus solvable. At low engine torque, the bridging can be closed with a higher horizontal torque, as the torsion damper takes over the oscillation disconnect.

The slimming can be chosen smaller. In the resonant area of the damper, slippage prevents production. Therefore no fi-element elements are necessary in the damper. The mini torsion damper is lighter and cheaper than a regular torsion damper.

In the following, they should be in the bubble diagram in fi g. 26 shown, strongly bridged connections are once explained with key words: -> softer converter => lower idle losses but => larger losses when driving O: \ uscrs \ ya \ dok \ word-dok \ 103289001 .doc å 20 '512 198 18 goal: reduce consumption => bridge at all gears + increase conversion => improve driving characteristics goal: reduce humming => steering accuracy => on average higher sliming necessary -> miniature damper + adaptive steering => average slip can be lowered + load change, shift, load and connection 10 and impact ratio can be optimized but not stationary and at low engine speeds => greater sliming necessary => heat dissipated => lower local oil temperature 15 -> conical lock-up => large surface => good coating utilization possible => higher cooling oil flow achievable but -> transferable torque is lowered => reinforcement effect of the cone => torque transmission is guaranteed => rigid shape => uniform pressing ln ii. => ideal effect of the coating => constant fi- ict value => good local temperature distribution 25 => increase of the service life 'i! ~ => lower moment of inertia => further improvement of the driving characteristics I fi g. 27 is shown by means of diagrams 1-5 for the criteria for determining the kmffactor or the km characteristic field, which may, for example, be stored in a central processing unit (CPU), where + means a good quality and - a poor quality.

The abscissa in this diagram shows the kmc-falctor and the ordinate the tendency effect of the criteria with regard to the size of the km-factor. As can be seen from a comparison between the idealized characteristic lines plotted in the diagrams, the different criteria are: l nu mi i riš- i: w - »- wt- O: \ users \ ya \ dok \ word-dok \ 103289 (X) l .doc 10 15 20 25 30 '512 198 19 partially opposite, i.e. they are, viewed over the kme factor, opposite. For this reason, taking into account several of these criteria, they must be weighted according to the case of use or in accordance with the desired vehicle condition in relation to their priority or importance. As can be seen from diagram 1, the acoustics or noise ratio cannot be arbitrarily improved by selecting a very small km factor, because otherwise due to the high slip in the bridging coupling an impermissibly high thermal load on it or on the transducer can occur. The negative influence of an excessive slimming in the bridge coupling is shown in diagram 2 in fi g. 27. There are thus boundary conditions, which should not be underestimated or exceeded. However, between the still acceptable boundary conditions, which should not be underestimated or exceeded, a change of the km-factor is possible. As already mentioned, the kmc falctom can be varied depending on the current operating conditions, whereby this variation can take place stepwise or continuously between certain limit values. The km-falctom can advantageously be variable depending on the condition variables of the motor vehicle. These state variables for the motor vehicle or the drive can be registered by a processor, so that the kme factor associated with these state variables can be set or registered. This km factor can, for example, be read from a stored characteristic field.

In many operating states of the internal combustion engine, it can be advantageous if the bridging clutch is actuated in such a way that it can transmit the entire net torque delivered by the internal combustion engine and occurring at any given time. Such an effect on the bridging clutch can be particularly advantageous in the lower operating range of an internal combustion engine, it being then particularly suitable if the bridging clutch has a damper which is designed for this partial load area. Such a damper thus has a bridging or abutment torque which is less than the maximum or nominal torque delivered by the internal combustion engine. This torque can be in the order of between 30 and 60% of the nominal torque of the internal combustion engine. The effect of such a damper is shown in diagram 1 in fi g. 27. The use of such an oscillating torch makes it possible to at least partially combat the acoustic problems prevailing in the lower drive range of an internal combustion engine in connection with a relatively large kme factor.

I fi g. 28 shows a block diagram or a circuit diagram for a torque control with adaptation, which is explained in more detail below. The actuator of the transducer bridging coupling can then take place via an electro-hydraulic I according to fi g. 28, at 1 of different input quantities, the driving torque of the drive unit is first calculated, such as in particular an internal combustion engine. The quantities extracted for this purpose comprise at least two of the following quantities, namely the speed of the drive unit, Oz \ users \ ya \ dok \ word-dok \ l032 89001 .doc,. trim _ -_n_.l * at small in 10 15 20 25 30 512 198 20 load arm position or accelerator pedal position for the fuel supply, negative pressure in the intake system, injection time, consumption, etc. At 2, the connection 1 takes place, which causes a correction of the driving torque. This correction is made by means of correction folders, which are emitted by the system adaptation denoted by 12. These correction factors can smooth out the anomalies occurring in the system in relation to the desired state, namely by equalizing these anomalies by additive, multiplicative and / or non-linear proportions.

At 3, the correct temperature error for the current operating condition is determined or measured.

This factor constitutes the torque ratio Mcoppiing to Md fi dning comma to be set by the control as a -for each operating point according to the type of a characteristic field of the selected weighting of the predetermined value given in Fig. 27. In this case, the design of any damper present in the bridging coupling is of particular importance, since in the presence of such a core at least over a relatively large section of the drive area of the internal combustion engine or the hydrodynamic torque converter can be kept constant.

At 4, the calculation of the set-torque torque takes place by means of the current torque and the corrected drive torque of the drive unit. At 5, a further correction of the set-torque can take place with the additive, multiplicative and / or non-linear proportions resulting from the system adaptation 12. Thus, the interconnection 2 can be achieved. For many use cases, it is sufficient if only one of the two connections 1, 2 is present, whereby preferably the connection 1 is to be maintained.

At 6, the calculation of the setpoints takes place through the corrected setpoint torque and the inverse transfer function for the section, which represents the bridging coupling. At 7, the control output quantities can be calculated on the basis of the set quantities obtained at 6 and the inverse transfer function of the set links. The actuator can be formed in a particularly advantageous manner by means of an electro-hydraulic In a advantageous manner, a proportional valve can be used or even a pulse width modulated valve. At 8, a feedback of the setpoints can take place before a regulation or adaptation. However, this feedback can also be omitted. At 9 a measurement of the actual torque can take place, e.g. by means of a torque shield or strain gauge (DMS). Instead of the measurement of the actual torque at 9 o'clock, a calculation of this torque can also take place through the state variables as well as of the vehicle and converter physics. For this purpose, for example, the motor characteristic field and / or the converter characteristic characteristic field or these characteristic fields representing quantities can be stored in a processor or in a central processor unit. Furthermore, for this purpose, the torque transmission capacity of the transducer bridging coupling or the representing quantities representing the transducer bridging coupling may be stored for this purpose.

If both a measurement of the torque according to point 9 and point 10 takes place, a comparison of the measured torque can be made with the torque calculated from the model. The comparison can then take place as a logical connection-minimum-maximum-formation or as a probability comparison.

I fi g. 28, system adaptation denoted by 12 can be made by the following comparison and the corresponding corrections by following. a) Ironing of corrected set-torque and is-torque, whereby this comparison can also take place long-term, ie. by observing the deviations over a concurrent time window.

Comparison of corrected torque and recalculated torque, whereby this comparison can also take place in the long term, Lex. by observing deviations over a concomitant time window.

Evaluation of additional signals, such as e.g. connection or disconnection of additive saws, such as e.g. clearing system, compressor, etc., changeover process. b) Sensing the system deviations registered under a) in additive, multiplicative and / or non-linear proportions of Mming and Mkoppl-lng and the resulting division into the corresponding adaptation loops 1 and 2 or in the connections 1 and 2.

The sensing or registration of the corresponding proportions of Md fi míng and / or Mkopp fi ng can, for example, take place according to the three diagrams in fi g. 29.

In the diagonal rune according to Example 1 in fi g. 29, an is-torque sequence and a set-torque sequence are plotted over time, the over-moment being exhibiting a sudden change.

This abrupt change can, for example, be traced back to the connection of an attachment, such as e.g. a compressor. The change of the torque caused by this additive, which is available to the converter, can be taken into account by an additive proportion, and by means of this the motor torque can be corrected in a corresponding manner.

In the diagram according to example 2 in fi g. 29 is also drawn an is-torque sequence and a set-torque sequence over time. From this diagram it appears that the ratio between the moment belonging to a certain time becomes substantially constant, although the difference between the moments changes. In this way, varying deviations between the is-moment and the should-moment can take place through a multiplicative proportion. Such a process between the is-moment and the should-moment can e.g. is traced back to the O: \ users \ ya \ dok \ word-dok \ 103289001 .doc haÄllllt _.._ nia.lz ... a .... 10 15 20 25 30 512 198 22 the frictional engagement or course of the friction value process occurring between the friction surfaces of the bridging coupling. This is a multiplicative link share.

In the diagram according to example 3 in fi g. 29 also shows an is-torque sequence and a set-torque sequence over time. As can be seen, the two moments change over time, however, over time the difference between the two moments is at least approximately constant.

Such a deviation between the current moment and the desired set moment can be equalized by means of an additive proportion. Such a course between the is-moment and the should-moment can be traced back to a deviation of the set size atria for the bridging coupling.

I fi g. 30 shows a fl fate diagram for a torque control with a very simple construction adaptation. The bridging coupling is then controlled electrohydraulically via a proportional valve and a pulse width modulated valve. The output signal fi from the control date or control output quantity is an adjusting current, which is set proportionally to that at e.g. the pulse width modulated output of the computer adjacent the sensing ratio. The coupling torque is obtained from the pressure difference controlled in this way by the transducer bridging coupling or between the two pressure cores of the bridging coupling.

The system adaptation is limited to the adaptive correction of the drive torque, the deviation of which results from the difference between the setpoint and the actual torque.

In an embodiment according to fi g. 30 lapses ironed with fi g. 28 the coupling 2 as well as the return of the corrected driving torque (M-an-corr).

I fi g. 30 is shown at 6 DP setpoints, namely as a function of the set-torque torque as the main quantity and also depending on the corrected drive torque (M-tor) and the turbine speed (n-turbine) as a parameter.

Function block 7 according to fi g. 28 is in fi g. Divided into two sub-function blocks, namely in 7a and 7b. The sub-function blocks 7a and 7b are each provided with a feedback Sa and 8b. The input quantities for the inverse transfer function for the set-top box (7 = 7a and 7b) are the setpoint difference (DP setpoint) calculated in block 6. The output quantity is formed by the associated sensing ratio as a control output quantity. The connecting actuator is divided into the electric actuator part, which is formed by an end step and the valve winding, as well as into the hydraulic actuator part, which determines the corresponding pressure effect of the transducer bridging coupling. the input to the electric switch is the sensing ratio. This is converted on the output side into is-current. Depending on this is-current (I-is), the hydraulic actuator part sets a corresponding pressure effect of the transducer bridging coupling. This is done by setting a corresponding pressure difference between the frame, Lex. 24, 25 according to fi g. 2 i O: \ users \ ya \ dok \ word 10 15 20 25 30 '512 198 23 converter bridge coupling. Block 7a constitutes the inverse function of the hydraulic actuator part, in which the associated set-current (I-set) is calculated by the setpoint pressure (DP setpoint). This part of the actuator has a feedback of the measured is-pressure (DP-ers) in the form of a pressure adaptation, which is shown by the block 8a. This pressure adaptation emits it (I-should-correct). The part 7b for the transmission function 7 of the set link constitutes the electrical part which calculates the associated sensing ratio from the corrected brush current. For this, a PID control algorithm was used. In this case, the control deviation I-should-corr = -I-is (I-is measured after the valve winding) is calculated with a PID controller the input quantity I-should-R for the inverse transmission ratio of the electric actuator part.

The i fi g. The selected numbering 1-12 for the individual blocks mainly corresponds to the numbering of the individual blocks in fi g. 28. In this way, the individual function blocks for the special electrohydraulic design according to fi g. According to the general embodiment, the second inversion corrected current current according to fi g is corrected. 28.

De i fi g. The individual designations given have the following meanings: DP setpoint = setpoint pressure difference in the lock-up or converter bridging circuit. Corresponds to the pressure difference between the prevailing pressures in the chambers on both sides of the piston.

DP is = between the two chambers in the converter bridging circuit. is-uyckdifference p-e fi er = pressure after the lock-up or converter bridging coupling, thus the pressure in the chamber 25 or in the return line 34 according to fi g. 2.

I-set = set-current for the electro-hydraulic valve.

A-n = speed difference between impeller and turbine wheel, thus A-n = n-impeller - n-turbular wheel.

Quantities denoted by "* -corr" correspond to quantities corrected by adaptation.

De i fi g. Before the condition denoted by the 10 denoted block, the vehicle quantities include the slip in the bridging coupling or the converter.

As further fi amgår of fi g. For example, the speed difference A-n = n-impellers - n-turbine wheels does not constitute a control variable, as is the case with the known slip controls. Vid O: \ users \ ya \ dok \ word ålar .. r. .,. ii .... i. ~ ail., t ...__ i.iaiišl i. 1-: Ä iiliillll Élln 10 15 20 25 30 512 198 24 torque control according to the invention, the speed An of this speed is used as the state variable for the distance to be controlled to observe any torque deviations, which in turn in the adaptation have a corrective effect on the control through corresponding connections.

In this case, the observed torque values, e.g. according to the type of a running time window is stored over a certain period of time, in order to detect the proportions of deviations of the clutch and motor.

This takes place in the system adaptation specified by 12.

The control according to the invention further has the advantage that the adaptation of the disturbing part of the drive torque can take place even with a fully opened lock-up or transducer bridging coupling, thus at km, = 0. For this, the nominal drive torque (M-an) is compared with the the transducer adjacent the moment, which takes place in the connection 1 according to fi g. 28 or at the procedure step in fi g. 28 and 30. By this adaptation, before a later closing of the bridging coupling, any deviations for the drive torque (M-an) can be taken into account in the open state of the bridging coupling. For this purpose, in the system adaptation 12, the torque abutting the transducer is measured, and preferably for this transducer characteristic field is stored in this system adaptation. Thereby, by measuring the speed difference between turbine wheels and impellers, the abutting moment can be calculated. This converter torque is then ironed with the nominal drive torque (N -an) for the motor or drive unit. This driving torque (M-an) can be obtained from one in block 1 according to fi g. 28 and 30 stored stationary motor characteristics field, namely due to the measured state variables, such as special engine speed, load arrester position, consumption, injection amount or injection time, etc. The speed difference between turbine wheels and impellers can be registered in block 10.

Furthermore, it is possible to register the transducer torque already in the block 10, whereby then the transducer characteristic field is stored in the block 10.

At that as an embodiment in fi g. 31 illustrated the torque transmission system 110 is a hydrodynaric torque converter 111 with a bridging coupling 112 and a damper unit 135 acting between the torque converter and the bridging coupling.

The torque converter 111 comprises a impeller 117 with an internal combustion engine (not shown), a turbine 118 operatively connected to a drive-side hub 114, a guide wheel 119 arranged in the perturbation circuit between the impeller and the turbine and a pivot-mounted connecting wheel with the impeller.

The transducer cover 116 is rotatably connected to the impeller 117 and by means of its drive connection to the internal combustion engine via projecting portions 116a projecting on the side facing the impeller 118 facing the impeller 118. to which a drive disk (not shown) of the internal combustion engine is attachable.

Between the turbine wheel 118 and the radial portion of the transducer cover 116, an annular piston 136 centered coaxially with the axis of rotation of the transducer is arranged, in which case it is a plate-shaped part. This annular piston is radially internally received on a drive hub 114 rotatably connected to the turbine wheel 118 and forms radially outwardly a conical portion which is provided with a suitable coating 121. The annular piston 136 cooperates with a corresponding conically shaped counter-friction surface 122 on the transducer cover 16.

The lock-up coupling 112 has a rear pressure chamber 124 between the ring piston 136 and the turbine wheel 118 and an upper pressure chamber 125 between the ring piston 136 and the transducer cover 116. The piston 136 is actuated in its coupling position cooperating with the front pressure chamber 122 by actuating the front pressure chamber 125. The magnitude of the torque transferable by the friction clutch 112 depends on the differential pressure set between the pressure core frames 124, 125.

The torsional torque pair 135 is designed in such a way that its bridging torque or impact torque is less than the nominal torque, i.e. the maximum torque of the internal combustion converter 110 driving the internal combustion engine. This means that the crane bearings 137 of the torsion damper 135 are designed in such a way that they can non-resiliently capture the entire torque of the internal combustion engine. The relative distortion between the input part 138 of the torsion damper 135 rotatably connected to the piston 136 and the uniform output part 139 can take place through windings for the springs 137 running on the block or preferably through stops arranged between the input part 138 and the output part 139.

The output part 139 of the pair of mandrels 135 is rotatably connected to the turbine hub 114 in a manner known per se via an axial plug connection formed by teeth.

As fi amgår of fi g. 32, the entrance part 138 cooperating with the collar magazine 137 can be formed by segment-shaped construction parts 140, wherein two such construction parts 140, arranged back to back, are arranged diarnetrally opposite. These pairs of segment-shaped structural members 140 are rotatably connected to the piston 136 via rivet connections 141. I fi g. 33 the output part 139 is shown in side view. The flange 139 has an annular base body 139a as well as two diametrically opposed radial arms 142 with recesses 143 for the power magazines 137. The arms 142 are axially received between the paired structural members 140. These pairwise segmented structural members form - viewed in their circumferential direction 144 receiving holes 145 for the arms 142. In Fig. 33, the abutment contours 140 formed by the segment-shaped 0: \ users \ ya \ dok \ word-dok \ 103289001.doc 'Éfl fi l fi lilllllllllv ll:: il 10 15 20 25 30 512198 26 146 for the arms 142 indicated by dashed lines. The piston 136 has axial depressions distributed over the circumference, which in the direction of the turbine wheel 118 form directed projections 147, against which the fastening portions 144 of the segment-shaped structural parts 140 facing the piston 136 abut. The structural parts 140 also have recesses 148 for the veins 137. These recesses 148 are in the axial direction shown in the axial direction in line with the recesses 143 in the exit part 139. In the exemplary embodiment according to fi g. 31-33, the air gases 137 are play-free received in recesses 143 and 148. However, for many use cases it may also be mandatory if at least one of the springs 137 relative to a recess 143 and / or 148 has play. Likewise, at least one of the springs 137 may be mounted with a predetermined locking ring in a window 143 and / or a window 148.

Because the torsion damper 135 according to the invention is only adapted to a partial load area, this can be designed particularly simply, whereby a cost-effective production is also made possible.

According to an embodiment of the invention, the torsional mandrel pair 135 can be designed in such a way that approximately 40-50% of the maximum, thus nominal torque of the internal combustion engine can be transmitted via the springs 137. The relative angle of rotation covered by the electric magazine 137 between the input part 138 and the output part 139 can, as can be seen from fi g. 34, in the order of 5 °. 1 fi g. 34 shows the relative angle of rotation between the input part 138 and the output part 139 of the damper 135 at the towing area of the motor vehicle. In push mode, this relative angle of rotation can be equal or have a different value. Likewise, the distortion stiffness of the torsion damper 135 can be different in the pulling direction and pushing direction. This is achieved by a corresponding dimensioning of the windows 143 and 148 and the springs 137. Likewise, the torsion damper 135 can have a characteristic line in its steps, whereby the characteristic line area, which corresponds to push-pull and traction drive, can also have different courses.

Av fi g. 34 it appears that the torsional mandrel pairs 135 are bridged at an angle of 5 ° or come to abutment and the torque transmitted by the elasticity or compression of the springs 137 is limited to about 45 Nm. A torque socket 135 thus formed can advantageously be used in connection with hydrodynamic torque converters which have a slip-controlled bridging coupling. The stop torque of 45 Nm is suitable for motors, which have a maximum nominal torque in the order of 80 - 200 Nm. The bridging torque of the damper 135 is suitably determined in such a way that it preferably covers the entire main driving range of a motor vehicle. The main driving area is the area which is most often used over the entire service life of a motor vehicle. This main driving range suitably comprises at least the area of the motor vehicle field which is decisive for the FTP75 cycle and / or for the ECE cycle (city, SEK 90 / hour, SEK 120 / hour).

The main driving area is thus the area in which the vehicle is most often used. Due to the traffic structures occurring in the individual countries, this driving range between the individual countries may be slightly different.

In it i fi g. In the drive characteristic pitch of a torque converter III with a soft converter design, the main driver area is shown as a narrowly hatched surface. Furthermore, i fi g. 35 shows the torque converter area of the torque converter. In this converter tube area, the bridging coupling 112 is open. The main driving range is surrounded by a gear range, in which it is preferably driven with a minimal slip in the bridging clutch 112. The main driving range extends from a lower speed A up to an upper speed B. The lower speed A then corresponds at least substantially to the idle speed, which may be in the order of idle speed. 700 - 800 rpm. The upper speed grinder B can be in a speed range of between 2000 and 3000 rpm, and e.g. exhibit the value 2200 rpm. The range with slip can have an upper speed limit C, which can correspond to the maximum speed for the internal combustion engine, but which can also be below that and e.g. have a value of between 3000 and 4000 rpm.

Due to the design according to the invention of the torsional torque pair 135, the torque converter 111 can be completely bridged in the main driving area, i.e. the bridging coupling 112 is driven without slippage, i.e. kM fold is greater than 1, e.g. 1.1. In this main driving range, the oscillation isolation between the internal combustion engine and the trailed gearbox is practically complete via the torsional vibration damper 135. Only peak torque is captured by the control or bridging clutch 112 in the main driving range so that this is reduced to the maximum torque in torque, which, however, is greater than the adjacent torque fi of the internal combustion engine.

In the area of slippage, the bridging coupling 112 is controlled or regulated in such a way that a certain slippage occurs between the friction surfaces 121, 122 of the bridging coupling 112.

Due to this slip, there is also a relative rotation between impeller 117 and turbine wheel 118.

In the area of slippage (km factor is less than 1, eg 0.9) according to fi g. The disturbing torques in this still occurring torque are attenuated mainly by slippage.

In the main driving area as well as in the area with sliming, for better vibration isolation, if in the driveline conditions can occur with higher oscillation amplitudes, ie. for example at resonance, the load shifting stroke or the like the transferable torque of the bridging clutch 112 is reduced. This can be done by changing the kme field tower.

As shown in fi g. 4, the torsional dampers of the bridging coupling 112 may also be designed in such a way that these connecting to a rotation angle with a relatively small rotation angle stiffness have a relatively small rotation angle, in which the torsional rigidity amounts to fl times that of the first rotation angle. I fi g. 34, this second angle of rotation extends over 2 °.

The torsional rigidity in this second rotational angle can amount to 7 - 15 times the torsional rigidity in the first rotational angle. At that in fi g. 34, the torsional rigidity in the first angle of rotation is in the order of 8 Nm / ° and in the second angle of rotation in the order of 70 Nm / °.

In the main driving area according to fi g. By means of the km-factor, the torque transmitted by the bridging coupling 112 is set to approximately 1.1 - 1.2 times the actual torque torque actually present. The control or control of the torque transferable by the bridging clutch 112 can take place in the main driving area in such a way that the torque transferable by the bridging clutch 112 does not fall below a minimum value. This value must be at least 1% of the nominal torque of the internal combustion engine. The minimum torque transmitted by the bridging coupling 112 in the main driving area can amount to, for example, 5 Nm. However, this lower limit can be shifted downwards or upwards, depending on the case of use. Thus, the minimum torque transmitted in the main driving range of the bridging clutch 112 can also be set to a value which is very close to the maximum motor torque occurring in the main driving range, preferably slightly less than this.

In it i fi g. The area denoted by "slippery area" is set by km, - the factor the torque transmitted by the bridging clutch 112 to 0.8 to 0.95 times the momentarily abutting moment from the combustion engine. The torque transmission capacity of the bridging clutch 112 is also dependent on the current torque from the internal combustion engine, which must be transmitted. In other words, this means that with increasing speed of the internal combustion engine, the torque transmitted by the bridging clutch also increases, and with a decrease in the speed delivered by the internal combustion engine, the torque transmission capacity of the bridging clutch 1 12 also decreases.

Due to the design according to the invention of the transducer bridging coupling as well as its control, an optimal operation of a motor vehicle is possible from an energy point of view.

By operating with the slimming-free bridging coupling in the mainly utilized operating states, compared with those in these operating states not bridged or with O: \ users \ ya \ dok \ word-dok \ 103289001 .doc 10 15 20 25 30 "512 198 29 The main speed range is approximately between 600 and 2200 to 3000 rpm or the average speed of about 1800 rpm. Thus, in the main driving range, the override clutch is substantially closed, so that the prevailing engine torque is transmitted through the engine torque. The oscillation damping takes place in this main driving range by the rotary vibration damper 135 arranged in the force fi or torque fl fate of the transducer bridging coupling 112. The torsion damper 135 is then provided with a relatively small turning angle and the approximate torque of the torque damper approximation of the principal torque. Depending on motorization and vehicle weight, this upper limit torque can amount to 15 to 50% of the maximum engine torque. With a damper constructed in this way, oscillations can be controlled in the driving area with lower drive torques, which produce a disturbing hum. Disturbing load change reactions in the driveline are suppressed or avoided by the relatively small angle of rotation of the torsion damper. Load change shocks are limited by sliding relative to each other when the stop torque or bridging torque of the damper fi surfaces of the bridging coupling slides relative to each other. This limits the torque to be transmitted. The torque peaks are damped by slipping in the bridging clutch. Over the main driving area or in the driving area, in which the adjacent torque is greater than the limit torque transmitted by the pair of bridges, the bridging coupling is controlled in such a way that a slip occurs. Disturbing load change reactions are avoided due to the slimming set in this way. In speed ranges or torque ranges over the main driving range, in which there are no disturbing oscillations, the clutch can also be closed at a torque value which is greater than the adjacent engine torque. For certain speed ranges, in which disturbing objects occur, the bridging coupling can be reconnected to slip. The latter may be particularly suitable for the occurrence of a resonant speed. Even in the main driving area or in the area with relatively small motor torques, it may be obligatory when passing resonances to open the bridging coupling or to considerably reduce the torque transmitted by it. Due to the design according to the invention and the control or regulation of the bridging coupling, especially the so-called hums are eliminated, which are not avoidable by a partially closed, i.e. slippery bridging coupling, namely due to the adhering / sliding conditions between the friction surfaces of this bridging coupling O: \ users \ ya \ dok \ word-dok \ 103289001.doc fi l Supplementary explanations for understanding the invention or the inventions as well as further advantages of the inventions or constructions according to the invention compared to the state of the art will be apparent from the following: 1: 2: - :: - iii: z:: iii 10 15 20 25 30 512 198 adjoining description.

Which control strategy should be selected for the converter bridging circuit system according to the invention? The problems that occur with slippage regulation have already been mentioned above.

The basic problem is that a regulation deviation must first occur before the regulation reacts.

In addition, there are areas in which the predetermination should not be achievable, e.g. no higher slip can be adjusted than can occur with an open converter. In the case of shifts, it has a negative effect if the control works against the function of the gearbox. Held e.g. in the case of upshifts, the slip is too low, at the end of the shift there is an attachment and thus comfort disturbances. Solutions can be imagined against all these regulatory problems - usually this is not the optimal solution after all. The LuK control concept therefore works torque-controlled, and system deviations are evened out through adaptation. The bridging torque is determined from the motor torque: Mömwg fl mg = Mmm, * bridging factor Consequently no is set This is also shown in the sum occurrence diagram in fi g. 36, the solid line indicating the state of the art and the dashed line representing the tanks according to the invention.

Whether the bridge is fully opened or slidably closed is determined only by energy considerations. An example: At an extreme mountain run (3600 kg, 12%) the low bridge cannot be closed completely, as e.g. traction fi reserve is not sufficient or the motor must not be pressed too hard. Then it is constantly compared, if the total losses are smaller, when bridging slippery or the bridging completely opens, see. G. 37.

When the driver wants a traction force, he increases the load arm position. First, the engine torque increases. When this torque is not sufficient, the driver increases the load arm position further, to make his additional acceleration request known. In ordinary systems, they are often downshifted, in order to increase the traction crane through a shorter gear ratio. In the case of WL systems, it is first checked whether, by opening the bridging coupling, a traction control is to be expected. This is the case when the converter should be in the converter area after opening. If this is the case, the bridge is opened, arms are shifted down. This check is done constantly. In order to improve collaboration, it is a learning obligation to also adapt the gearbox transmission lines to this concept. This adaptation is especially effective in combination with a soft transducer (see next chapter). This fi loso fi can be roughly shown in a shift line diagram (Fig. 38).

O: \ users \ ya \ dok \ word-dok \ 10328900l .doc 10 15 20 25 30 512 198 31 Already through these measures (sliming, cone, mini-dampers and adaptive control) a clear improvement of consumption is achieved, as it can be bridged in all gears. The softer converter also provides a further noticeable improvement.

To begin with, this contribution is transferred to the converter design. As a bridging in all areas is not possible with the systems common today, the converter must be correspondingly rigidly designed. With the WL concept, it is possible to take advantage of a soft converter and avoid the disadvantages. The advantages are a significantly better traction and lower downtime losses. The disadvantages - in many areas under load higher losses and "gurnmibandse fl ekten" - are avoided by the constantly useful bridging coupling.

With this design, further prominent advantages are achieved. Driving characteristics are significantly improved and consumption is significantly reduced. In addition, emissions increase disproportionately. The test cycles begin with a cold phase. When the bridge is opened, the motor at a soft converter reaches its operating temperature significantly faster, which has a favorable effect on the emission.

At acceleration 0 to 100 km / h, there are no major differences between standard four-speed gearboxes and five-speed gearboxes, as the gears at the lowest gears are almost equal. With the four-speed gearbox with the LuK-WL system, considerable acceleration is obtained (through the softer convertible lamp). ironed with the usual five-speed gearbox. Significant improvements can also be achieved with regard to consumption, both ironed with ordinary four-speed gearboxes and also with ironwork with ordinary five-speed gearboxes. Significant improvements are also to be expected in the errands.

A cost-effective solution with many advantages: The combination of a converter bridging system with a four-speed gearbox.

The converter bridging system in combination with a four-speed gearbox can achieve similar benefits in terms of driving characteristics and consumption as a five-speed gearbox with a standard bridging clutch and at an overall significantly lower weight and cost than a five-speed gearbox (completely apart from the possible savings in development savings ).

The combination of a soft converter with a wide-laid four-speed gearbox shows in the drawbar fi diagram in most areas even a higher drawbar fi than a five-speed gearbox with a standard converter (fi g. 39). It is also seen that in the gear units with lower load, the five-speed gearbox must be shifted over two gear stages and the four-speed gearbox with the converter bridging system not at all, ie. gearbox shift frequency is reduced. The larger gear jumps are captured by the softer converter.

O: \ users \ ya \ dok \ word-dok \ 103289001 .doc 5 'I' 2 'l 9 8 32 The invention is not limited to the exemplary embodiment shown and described, but in particular also includes variants which can be formed by combinations of in compounds. features or elements described with the present invention. Furthermore, individual features or modes of operation described in connection with the features may constitute an independent invention. 0: \ users \ ya \ dok \ word ~ dok \ 103289001 .doc

Claims (5)

Patent claims 1. Combustion engine drive system and slip-controlled bridging clutch fi is a hydrodynamic torque converter, characterized in that the bridging clutch contains a torsional torque pair, the actuating torque of which is lower than the rated torque of the pre-torque motor. Drive system with combustion engine according to claim 1, characterized in that the actuating torque amounts to between 10 and 60% of the maximum torque of the dry combustion engine, preferably to between 25 and 50%. Drive system with dry combustion engine according to Claim 1 or 2, characterized in that the damper does not have an ection device. 4. A combustion engine drive system according to any one of claims 1-3, characterized in that the damper allows a relatively small turning angle in the order of magnitude of j; 2 - 6 °, preferably in 3 - 8 °. Drive system with internal combustion engine according to one of Claims 1 to 4, characterized in that the damper has a stiffness of 7 to 30 Nm / °. í.-_.....-___._.____-_-._- O: \ users \ MD \ DOK \ WORD-DOK \ P32890SE0l.p.doc