Method and apparatus for gear shifting control with improved shift quality during gear shifting

An apparatus for gear shifting control with improved shift quality during gear shifting includes a gearbox unit containing at least one gear branch for power transfer via a hydrodynamic clutch. The hydrodynamic clutch can be coupled in series with a mechanical gearing part containing a device for realizing at least two speeds. In the course of power transfer via the hydrodynamic clutch during a gear shift operation between two mechanical speeds, the fill factor of the hydrodynamic clutch is so altered that a variable which at least indirectly characterizes the rotational speed of the primary bucket wheel of the hydrodynamic clutch is held substantially constant over a certain range of the gear shift operation.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to a method and apparatus for gear shifting 
control with improved shift quality during changing from a first gear to a 
second gear, especially when upshifting. 
2. Description of the Related Art 
Hydrodynamic-mechanical gear box units are known in a multitude of designs 
in the form of automatic transmissions for use in motor vehicles. They 
contain at least one gear branch, which includes a hydrodynamic clutch and 
a mechanical transmission section, installed in tandem with the clutch. 
The mechanical transmission section includes rotational speed/torque 
converters with an appropriate gear ratio, which can be operated through 
appropriate control elements. Various speeds are achieved by appropriate 
operation and/or releasing of control elements in the mechanical 
transmission section. The individual rotational speed/torque converters 
are preferably in the form of spur gear or planetary gear sets. To shift 
gears, it is necessary to actuate or release the appropriate control 
elements of the mechanical transmission section. When shifting up in an 
automatic transmission, for example, two phases occur, namely the 
so-called torque phase and an inertia phase. Before the upward shift is 
effected, the torque input at the transmission clutches includes only the 
torque which is produced by the combustion engine and which runs through a 
torque converter. During the torque transfer phase, pressure is exerted on 
the clutch part that is to be engaged. This torque is divided between the 
releasing clutch and the engaging clutch. At the end of the torque phase, 
the torque that is supported by the releasing clutch drops to zero, and 
the entire torque is transferred to the engaging clutch. After completion 
of the torque transfer, the phase of transferring the gear ratio, the 
so-called inertia phase, begins. During this phase, the rotational speed 
of the combustion engine is quickly reduced to the value of the new gear 
ratio. This occurs to the same degree in which the pressure of the 
engaging clutch increases. This produces a large moment of inertia which 
must be absorbed by the clutch part in addition to the torque that is 
produced by the combustion engine. This inertia torque produces a torque 
thrust which is transmitted to the interior of, for example, a vehicle. 
A multitude of options are known for the reduction of the torque thrust. 
One of these options is described as an example in German patent document 
no. DE 691 13 193 T2. According to the disclosure, control of the shifting 
procedure is provided by continuously monitoring the input speed at the 
transmission input shaft and by comparing a change at the end of the 
torque phase with a stored value. Preferably, the control is improved by 
delaying the motor's timing of ignition by a certain time. An adaptive 
control system is provided to adjust the pressure in a device intended for 
the generation of torque by use of friction for a drive train, which 
includes a combustion motor and a step-up gear. The system further 
includes a shaft for input of the torque between the combustion engine and 
the gear box, an initial friction mechanism for at least partial provision 
of a first transfer path for the torque between the drive shaft and 
primary parts of the transmission for the input of torque, and a second 
friction mechanism which serves to transfer the torque between the drive 
shaft and secondary parts of the transmission. The system further includes 
a primary and a secondary pressure operated servo unit which, when under 
pressure, will operate the first and the second friction mechanisms 
respectively, whereby the creation of the second transfer path for torque 
is accompanied by a reduced motor speed. Also included is a device for 
constantly monitoring the speed at the torque input shaft, that is, at the 
transmission input shaft, and a device for increasing the pressure of the 
second servo unit as a response to a command to change the transmission 
ratio, which requires release of the first friction mechanism and 
actuation of the second friction mechanism. The system further includes a 
device for reduction of pressure in the servo unit, as a response to a 
predetermined measured change in the rotational speed of the torque input 
shaft, and a device to increase pressure in the servo unit as a result of 
the reduction in the clutch ability of the first friction mechanism during 
the shifting procedure. The disadvantage of such a control system is 
essentially that the necessary adaptation of the pressure progression must 
occur very sensitively and is therefore very expensive. Adaptation of the 
pressure value is furthermore always opposite the actual factors at the 
transmission input shaft, particularly depending on the rotational speed 
at the transmission input shaft. Changes in the speed and the mass moment 
of inertia of the combustion engine influence the necessary adaptation of 
the pressure progression. 
SUMMARY OF THE INVENTION 
The present invention provides a method and apparatus for controlling gear 
shifts on hydrodynamic-mechanical gear boxes such that the mentioned 
disadvantages are avoided. In particular, a high level of shifting comfort 
is achieved with little or no moment distortion during shifting, which is 
caused by rotational speed changes and the mass moment of inertia of the 
combustion engine. Design and control expenditures are kept as low as 
possible. 
The present invention relates to a method and apparatus for improving the 
shifting quality during gear shifting, i.e., shifting between two speeds 
in a transmission. The transmission contains at least one gear branch for 
power transfer via a hydrodynamic clutch, which can be coupled in series 
with a mechanical transmission section in the power transfer branch, 
containing devices for realizing at least two speeds. During the shifting 
procedure between two mechanical speeds in the hydrodynamic-mechanical 
power transfer transmission branch, the fill factor of the hydrodynamic 
clutch is monitored and adjusted in such a manner that a variable, which 
at least indirectly characterizes the rotational speed of the pump wheel 
of the hydrodynamic clutch, is held substantially constant over a certain 
range of the shifting operation. Consequently, at the beginning of the 
shifting operation, control of the rotational speed of the pump wheel 
starts through appropriate filling control at the hydrodynamic clutch, 
whereby the speed level prior to shifting is maintained. A firmly 
definable desired rotational speed, which may be established for every 
shifting operation may also serve as a comparison or desired value. 
In known methods, shifting between individual gears during pull-upshifting, 
that is, when changing from a first lower gear into a second higher gear, 
normally leads to a suppression of the rotational speed. The reduction of 
the rotational speed occurs at least at the primary bucket wheel and the 
components coupled with this, in particular the machine drive shaft. With 
the method of the present invention, however, the rotational speed of the 
pump wheel is controlled to an almost constant: value, and the 
hydrodynamic clutch becomes softer and allows the necessary slippage to 
occur naturally in the hydrodynamic clutch. Because of this, in the 
shifting process during changeover of the power transfer unit and/or the 
ratio of rotational speed/torque in the mechanical driving gear, no 
changes in the rotational speed occur at the transmission input shaft or 
in the drive motor coupled with it. Shifting itself can be accomplished 
smoothly and without great pressure build-up. A reduction of slippage is 
again achieved by the change in the driving speed or the forcible 
hardening of the hydrodynamic clutch, particularly through increase in 
filling, under consideration of the minimal acceptable rotational speed of 
the drive motor. The time span for the control process can be from the 
beginning of the shifting process to the time of reaching the synchronous 
rotational speed. However, the time span can also be chronologically 
displaced from the beginning of the shifting process. Alternatively, the 
time span can be completed when a certain slippage value, that is, a 
certain difference between the rotational speeds of the primary bucket 
wheel and the secondary bucket wheel on the hydrodynamic clutch is 
reached. Preferably, the control process is concluded only then, when the 
slippage has been reduced to a minimum. Normally, this would be two or 
three percent of the slippage level normally prevailing in hydrodynamic 
transfer of power. The drive motor may then speed up further, while the 
hydrodynamic clutch operates again at maximum efficiency. 
The method of the present invention is generally suitable for all gear 
shift operations in transmissions in which the hydrodynamic clutch is not 
equipped with beveled blading and participates in the transfer of power. 
On clutches that are equipped with beveled blading, the method of the 
invention is preferable for improvement of the gear shifting quality in 
pull upshifting. In push-upshifting, pull-reverse shifting and 
push-reverse shifting, such a control system is normally unnecessary. In 
push operation, in which a filled hydrodynamic clutch that is equipped 
with beveled blading is turned on, the drive motor coupled with the 
transmission input shaft and, therefore, the primary bucket wheel of the 
hydrodynamic clutch cannot be dragged. In this scenario, the hydrodynamic 
clutch operates in a centrifugal mode when driving the secondary bucket 
wheel through the transmission output side. The vehicle rolls without 
braking too harshly when the gas supply is interrupted, i.e., there is no 
load change impact and the drive motor runs with slightly increased idling 
speed. The behavior of the hydrodynamic component corresponds to that of a 
hydrodynamic rotational speed/torque converter. Push up-shifting 
situations have substantially no effect on the drive motor coupled to the 
transmission input shaft and are therefore barely noticeable. 
In traction operation, the motor normally supports itself at the secondary 
bucket wheel during transfer of power in the hydrodynamic-mechanical power 
transfer branch. This also applies to reverse shifting, during which the 
rotational speed of the secondary bucket wheel increases. The motor speed 
cannot be increased by the selected gear. The hydrodynamic clutch again 
operates in a centrifugal mode, which is why the motor mass cannot 
negatively influence the shifting process. 
With regard to push-reverse shifting with transmission arrangements having 
a hydrodynamic clutch equipped with beveled blading, the same comments 
apply as for push-upshifting. Here too, the motor speed is not affected by 
the gear shifting operations. 
The method of the invention provides the advantage that lower drive motor 
speeds are attained by matching the output of the drive motor with the 
desired driving performance. Overall, a higher level of efficiency is 
achieved under driving conditions. Because of the displacement of slippage 
or compensation for slippage during gear changes into the hydrodynamic 
clutch, the individual shifting components necessary for achieving the 
individual speeds are far less stressed. Since during the gear shifting 
process no moment distortions occur which would be caused by changes in 
the rotational speeds and by mass moments of inertia, the gear shifting 
process proceeds almost smoothly which manifests itself in increased gear 
shifting comfort. Very low shifting speeds may be selected, since the 
rotational speed of the drive motor does not change during shifting. The 
progression of rotational speed of the drive motor during gear shifting 
with hydrodynamic clutch is constant. 
The device includes a gear box unit having a first hydrodynamic gear 
component with at least a primary and a secondary bucket wheel, which 
together form a toroidal operating chamber and function as a hydrodynamic 
clutch. The gear box unit also has a second mechanical gear component with 
at least a device for achieving two different speeds, and control or 
regulating equipment. The control or regulating unit includes at least two 
inputs and one output in order to achieve operation consistent with the 
invention. A first input is coupled with a device to measure or establish 
a signal for the presence of a shifting process. A second input is coupled 
with a device capable of measuring at least one value which at least 
indirectly characterizes the average rotational speed of the primary 
bucket wheel of the hydrodynamic clutch. The value at least indirectly 
characterizing the average rotational speed of the hydrodynamic clutch of 
the primary bucket wheel is compared with a desired value which is 
preferably consistent with the value of the rotational speed measurable at 
the beginning of a gear change, or with the indirectly characterized 
rotational speed value of the primary bucket wheel. Depending upon the 
deviation between the actual and the desired value, a variable Y is formed 
which serves to trigger a device to change the degree of fill of the 
hydrodynamic clutch. Changing the degree of fill can be accomplished 
through a control or regulating unit. The control or regulating unit is 
secondary to the rotational speed control for the primary bucket wheel. 
The following options exist for regulating the degree of fill: 
a) Controlling the degree of fill within the scope of the rotational speed 
control; or 
b) Controlling the degree of fill within the scope of degree of fill 
control which is secondary to a rotational speed control. 
For details on the execution of rotational speed control, we refer to the 
comments in "Voith--Hydrodynamic in drive technology", Krauskopf 
Engineering Digest 1987, Chapter "Control and adjustments", which is 
incorporated herein by reference. The adjustment of the degree of fill or 
realization of rotational speed adjustments is not limited by a concrete 
implementational example. It is, however, important that a value which 
represents at least an indirectly characterized rotational speed of the 
pump wheel is set during the initial speed levels, and that this value is 
held constant. 
The transmission of the present invention displays good efficiency 
throughout the entire range and no disadvantages are noted regarding 
shifting comfort and driving characteristics when compared with currently 
known transmissions. During the operational condition of "driving", the 
hydrodynamic clutch component is preferably operated with maximum fill 
since there can be only minimal slippage. The clutch therefore provides 
excellent efficiency which, depending upon load and rotational speed, may 
be between 90 and 99%. These values are not achieved by any hydrodynamic 
converter. The hydrodynamic component which operates as a clutch during 
the start up process shifts hard. In spite of this, it remains a hydraulic 
transfer component with all its advantages, for example damping of 
vibration and noise reduction. In all subsequent shifting sequences, the 
slippage of the hydrodynamic component is increased, and shifting is 
softer and achieves values which are consistent at least with the slippage 
of a hydrodynamic converter. Shifting of individual gears in the first 
operational condition may be rapid. The slippage can be displaced into the 
hydrodynamic component and may then be reduced again by increasing the 
degree of fill. Thereby, comfortable transitions are achieved for 
realization of the individual speeds, while at the same time reducing the 
load on the shifting elements. 
Even with suddenly occurring load changes, the slippage of the hydrodynamic 
component may be increased for a short period, so that a soft hydraulic 
connection results. Also, the support in push operation is very low. The 
dreaded load change impact is then a non-event. In addition, the 
characteristic curve of the utilized flexible damper displays better 
characteristics during load changes. The hydrodynamic component is then 
bridged in the operational condition of "driving", when shifting between 
the individual gears no longer causes loss of shifting comfort. This 
function is ensured by appropriate control of the filling and emptying 
processes. 
In the braking mode, the hydrodynamic clutch can be utilized as a retarder.

Corresponding reference characters indicate corresponding parts throughout 
the several views. The exemplification set out herein illustrates one 
preferred embodiment of the invention, in one form, and such 
exemplification is not to be construed as limiting the scope of the 
invention in any manner. 
DETAILED DESCRIPTION OF THE INVENTION 
Referring now to the drawings and particularly to FIG. 1, there is shown an 
example transmission, including a first hydraulic transmission section 2 
in the form of a hydrodynamic clutch and a second mechanical transmission 
section 3, the functional mode of the method in accordance with the 
invention. The first hydraulic transmission section 2 includes at least 
two bucket wheels--the first bucket wheel and a second bucket wheel. The 
first bucket wheel is described as primary bucket wheel 4 and the second 
bucket wheel as secondary bucket wheel 5. The primary bucket wheel 4 and 
the secondary bucket wheel 5 together, form at least one toroidal work 
chamber 6, which can be filled with operating medium. For this purpose, an 
operating medium supply unit which is not shown in detail, is allocated to 
the toroidal work chamber 6. The gear box unit 1 is designed so that the 
hydrodynamic transmission section 2 and the mechanical transmission 
section 3 can be connected in series for the purpose of power transfer 
from input shaft E to the transmission output shaft A. The possible 
coupling of hydrodynamic transmission section 2 and mechanical 
transmission section 3 in series for the purpose of power transfer between 
the transmission input shaft and the transmission output shaft A describes 
the gearing branch for the power transfer between transmission input shaft 
E and transmission output shaft A. A second gearing branch is described, 
for transfer of power from the transmission input shaft E to transmission 
output shaft A, in this instance bypassing the hydrodynamic transmission 
section 2. In the following, the comments regarding the method relate to 
an operational mode which enables hydraulic-mechanical power transfer 
between the transmission input shaft and the transmission output shaft, 
that is, in the first gear/gearing branch via the hydrodynamic clutch 2 to 
the mechanical transmission section 3. The constructive design of the 
example transmission is described below: The secondary bucket wheel 5 is 
permanently connected with the mechanical transmission section 3 through a 
connecting shaft 7. Connecting shaft 7 can be connected with the 
transmission input shaft E through a bridge coupling UK which may also be 
described as a "through-coupling." The secondary bucket wheel 5 can also 
be connected with transmission input shaft E through this coupling. The 
primary bucket wheel 4 can be connected by use of a primary bucket wheel 
coupling PK with the transmission shaft E. The primary bucket wheel 4 is 
mounted on a connecting shaft 8, whereby connecting shaft 8 can be 
connected through the primary bucket wheel coupling PK with the 
transmission input shaft E. A brake component which, in this instance, is 
described as primary bucket wheel brake PB, is allocated to the connecting 
shaft 8. This primary bucket wheel brake PB is mounted rigidly on a 
stationary transmission component, preferably at the gear box 9, as 
indicated here. 
The secondary bucket wheel 5 can be coupled with the second mechanical 
transmission section 3 through a connecting shaft 7. In the illustrated 
example, the second mechanical gearing part 3 includes three planetary 
gear sets--a first planetary gear set PRI, a second planetary gear set 
PRII and a third planetary gear set PRIII. The individual planetary gear 
sets each include at least a first sun gear, identified with Ia for the 
first planetary gear set PRI, with IIa for the second planetary gear set 
PRII, and with IIIa for the third planetary gear set PRIII. The individual 
planetary gear sets also each include a ring gear, planetary gears and a 
fixed link. The ring gears for the individual planetary gear sets are 
identified here with Ib for the first planetary gear set PRI, with IIb for 
the second planetary gear set PRII and with IIIb for the third planetary 
gear set PRIII. In the driving mode, six speed levels could, for example, 
be achieved with this transmission unit, whereby at least two speeds are 
free from the operation of the connecting coupling. In the first gear of 
the first operational mode, the so-called start-up gear, the primary 
bucket wheel clutch PK is operated and thereby connects the transmission 
input shaft E with the primary bucket wheel 4 through connecting shaft 8. 
The connection between the hydraulic transmission section 2 and the 
mechanical transmission section 3 is achieved through the first coupling 
component K1. The second and third braking elements B2 and B3 are also 
actuated. The hydraulic component 2, particularly the toroidal work 
chamber 6, is filled with operating medium in this stage. The power flow, 
or power transfer, requires a transmission input shaft E which is coupled 
indirectly with a drive motor (not shown), the primarily bucket wheel 
coupling PK, the primary bucket wheel 4, the secondary bucket wheel 5 
through the first coupling element K1, the first planetary gear set PRI, 
particularly the ring gear Ib of the first planetary gear set onto fixed 
link IIId of the third planetary gear set and thereby the transmission 
output shaft A, which can be connected at least indirectly with a unit 
that is to be driven, for example the wheels of a vehicle. 
The hydraulic transmission section 2 works as a hydrodynamic clutch during 
the start up process. During the transition into the second gear, the 
first clutch element K1, as well as the third brake element B3 remain 
actuated, as does the primary bucket wheel coupling PK. The second brake 
element B2 is released and the first brake element B1 is actuated. Since 
at least in the second gear the connecting coupling UK is also not 
actuated, the hydraulic gearing element 2 takes over the function of the 
hydrodynamic clutch within the second gear. Preferably, the connecting 
coupling is also not yet actuated in the third gear. However, the primary 
bucket wheel coupling PK, the first coupling element K1 as well as the 
second coupling element K2, and the third brake element B3 are actuated. 
All other load shifting elements are not engaged. The power flow 
therefore, occurs through the transmission input shaft E, the hydraulic 
transmission section 2, the connecting shaft 7 through the first coupling 
element K1 to the first planetary gear set PRI, particularly to the ring 
gear Ib of the first planetary gear set PRI. A further power segment is 
transferred through the second coupling element K2 to the sun gear Ia of 
the first planetary gear set PRI. The power segments that are again 
brought together at the fixed link Id of the first planetary gear set PRI 
are transferred through fixed link IIId of the third planetary gear set 
PRIII to the transmission output shaft A. 
In fourth gear, the first coupling element K1, the third coupling element 
K3, the connecting coupling UK, the primary bucket wheel coupling PK, as 
well as the second brake element B2 are actuated. None of the other power 
shifting elements are engaged. The fixed link IId of the second planetary 
gear set PRII is established by use of the second brake element B2. The 
ring gear IIIb of the third planetary gear set PRIII is connected with 
fixed link IIId of third planetary gear set PRIII. The power transfer 
occurs from the transmission input shaft E through the connecting coupling 
UK, the connecting shaft 7, the first coupling element K1 to the ring gear 
1b of the first planetary gear set PRI, through the second planetary gear 
set PRII, the connecting shaft 10 to the planetary gears IIIc of the third 
planetary gear set PRIII which drive the fixed link IIId which is coupled 
with ring gear IIIb, to the transmission output shaft A. 
In fifth gear, only the second brake element B2 is released and the first 
brake element B1 is actuated, meaning that the connection between the 
first planetary gear set PRI and the second planetary gear set PRII, 
connecting shaft 11, is established. Sun gears Ia and IIa of the two 
planetary gear sets PRI and PRII are therefore stationary. Power transfer 
occurs again through the input shaft E, the bridging coupling UK, the 
connecting shaft 7, the first coupling element K1, the ring gear Ib of the 
first planetary gear set PRI and from there through the fixed link Id of 
the first planetary gear set PRI to the fixed link IIId of the third 
planetary gear set PRIII and thereby, the transmission output shaft A. 
The sixth gear differentiates from the fifth gear in that all three 
coupling elements K1, K2 and K3 are actuated, while all brake elements B1, 
B2 and B3 are released. 
The shifting quality when changing from one gear to the next, substantially 
depends on the actuation of the individual control elements K1, K2, K3, 
B1, B2 and B3 in the mechanical transmission section. In a particular 
instance, the chronological sequence of the actuation of the "coming 
control element" and the release of the so-called "leaving control 
element" is decisive, as is the extent of the actuating forces. The 
"coming control elements" are those control elements or components which, 
during gear changes are to be engaged between two different gears, i.e. 
control elements that are actuated in the following selected gear. The 
"leaving control elements" are those control elements which, at the time a 
gear change is selected are still engaged or actuated in the previous gear 
and which are to be released at gear change. 
With the help of various diagrams, FIGS. 2a through 2d provide simplified 
plots of the rotational speed/torque progressions which could possibly 
apply for the transmission described in FIG. 1, without the method of the 
invention. In particular, the relationships between the pressures imposed 
on the "coming" control elements over a period of time, the occurring 
torque changes, as well as the rotational speed changes on the drive side, 
that is, on the side of the transmission input shaft, as well as on the 
output side are given in the example. From this it becomes apparent that 
considerable moment distortions may occur in area I. According to the 
hitherto known state of the art, attempts have been made to reduce these 
by matching pressure progressions on the coming and/or leaving control 
elements. 
In the method of the invention, as a solution of the problem and for the 
removal of the aforementioned disadvantages, a control or regulating unit 
15 is allocated to gear box unit 1. Control or regulating unit 15 has at 
least two inputs, including a first input 16 and a second input 17, and a 
first output 18. The control or regulating unit 15 serves to compensate 
for slippage during gear changes in the mechanical transmission section 3. 
In particular, this relates to the slippage between the individual clutch 
and brake elements which are coupled with the drive or output side and 
which connect these for the purpose of torque transfer. The drive side in 
this instance is to be understood to be the area of the gear box unit 1 in 
which, during traction operation, power is transferred from the 
transmission input shaft in the direction of the transmission output shaft 
to a first element of the "coming" control element. The second element of 
the "coming" control element, which is coupled at least indirectly with 
the transmission output shaft A, is on the output side. Particularly in 
pull-shiftup situations, a suppression of the rotational speed occurs both 
at the transmission input shaft E and in the drive unit coupled with it 
when changing from one gear to the next higher gear, for example from the 
first gear to the second gear. The drive unit can be a combustion engine. 
In the method of the invention, this rotational speed suppression is 
compensated by the hydrodynamic clutch 2. This, however, applies only to 
shift-up situations, that is from a first lower gear to a second higher 
gear, during which the transfer of power occurs at least partially through 
the hydrodynamic clutch 2. The slippage compensation takes place by 
changing the fill level in the work chamber 6 of the hydrodynamic clutch 
2. The fill level change is made in order to realize a substantially 
constant rotational pump wheel 4 speed, and, at least indirectly, a 
substantially constant rotational speed in the drive unit that is coupled 
with gear transmission 1. The hydrodynamic clutch 2 is, therefore, used 
for speed control. The fill level change itself may occur during an 
adjustment which is secondary to this control and/or speed control. To 
convert the speed control, an actual value that at least indirectly 
describes the rotational speed of the primary bucket wheel 4 of the 
hydrodynamic clutch 2 is measured and supplied to the first input of the 
control or regulating unit 15. The first input 17 can, for this purpose, 
be connected to a device for measuring the rotational speed of the pump 
wheel 4. Since, in the method of the invention, the rotational speed 
control is accomplished during gear changes for the purpose of slippage 
reduction, the second input 16 of the control or regulating unit 15 is 
connected to a device for establishing and/or measuring a signal for the 
presence of a gear change SW. In the present example, the rotational speed 
of transmission input shaft E is measured in order to determine the 
rotational speed of the pump wheel 4. The actual value n.sub.EIST is 
compared with a desired value n.sub.ESOLL by the control or regulating 
unit 15. The desired value n.sub.ESOLL is a determined actual value used 
to describe the rotational speed of the pump wheel 4 at least indirectly, 
and is a firmly definable value for varying operating conditions and/or a 
usable desired value. In the last instance, the desired value can, for 
example, be a rotational speed or a characterizing value which is present 
at the beginning of a shifting process or in the presence of a signal for 
a desired gear shift at the transmission input shaft E and/or the pump 
wheel shaft 8 of the pump wheel 4 of the hydrodynamic clutch 2. The 
comparison is performed in a comparator 19 whereby it is at least decided 
whether the difference between the desired and the actual value is zero or 
not equal to zero. Only in the last mentioned instance is a correcting 
variable given at output 18 of the control or regulating unit 15. The 
correcting variable affects a device 20 for at least indirectly 
influencing the fill level in work chamber 6 of the hydrodynamic clutch 2. 
In this example, device 20 acts, so to speak, as an actuator. 
The following options are available for regulating the fill level: 
a) Regulation of the fill level within the scope of rotational speed 
control; or 
b) Regulation of the fill level within the scope of a fill factor control 
which is subordinate to the rotational speed control. 
For details on the execution of rotational speed control, we refer to the 
comments in "Voith--Hydrodynamic in drive technology", Krauskopf 
Engineering Digest 1987, Chapter "Control and adjustments", which is 
incorporated herein by reference. The adjustment of the fill level or 
realization of rotational speed adjustments is not limited by a concrete 
implementational example. It is, however, important that a value which 
represents at least an indirectly characterized rotational speed of the 
pump wheel is set during the initial speed levels, and that this value is 
held constant. The embodiment of the transmission unit illustrated in FIG. 
1 is only an example variation. 
The control and regulating unit 15 which serves to effect the fill level 
change may also be designed such that additional elements, particularly 
control elements, are triggered in the mechanical transmission section 3. 
Preferably, the control or regulating unit 15 would be part of an 
overriding control or regulating unit. The overriding control or 
regulating system in vehicles would preferably be the central drive 
control, which serves to trigger various components in the drive train. 
Corresponding to individual input values at the control or regulating unit 
15, regulation of the rotational speed of the pump wheel 4 begins at the 
beginning of the shifting process, preferably due to appropriate fill 
factor control of the hydrodynamic clutch 2, whereby the rotational speed 
level is maintained prior to the shifting process. Shifting of gears would 
normally lead to a suppression or reduction of the rotational speed of the 
pump wheel 4 and the motor that is coupled with the input shaft E. 
However, the hydrodynamic clutch 2 becomes softer due to the control 
process and permits the necessary slippage in the hydrodynamic clutch 2. 
On pull upshifting for example, a fill level change in the form of a 
reduction occurs to compensate for the slippage. The fill level change is 
determined by the extent of slippage. 
FIG. 3 clearly illustrates the dependency between filling level and 
slippage, which may be described by the extent of the relationship of 
rotational speeds of pump wheel and turbine wheel or primarily bucket 
wheel 4 and secondary bucket wheel 5 of the hydrodynamic clutch 2. The 
shifting operation does not cause changes in the motor speed. Shifting of 
gears can be swift, without great pressure build-up. Reduction of slippage 
is achieved by forced hardening of the hydrodynamic clutch 2, considering 
the minimum acceptable motor speed. The rotational speed control is 
disabled only when the slippage has been reduced to a minimum. The drive 
motor that is coupled with transmission input shaft E can accelerate 
further, that is, the hydrodynamic clutch 2 works again with maximum 
efficiency. 
The method according to the invention in mechanical-hydrodynamic compound 
transmissions with clutches having bevelled blading, is primarily relevant 
for realization of pull-upshifting situations. In push-operation or in 
pull-reverse shifting, such characteristics in this type of hydrodynamic 
clutch are usually not important. Apart from this, that is, in clutches 
that have non-beveled blading, the method of the invention is relevant for 
all shifting operations. 
FIG. 4 illustrates in simplified form the motor output versus the 
acceleration. The dot-dash line indicates the behavior in a conventional 
transmission unit with a hydrodynamic clutch and a mechanical transmission 
section. The hatched area 1 is the operational area which emphasizes the 
operation of the hydrodynamic clutch, that is, the filling and the 
transfer of power through the hydrodynamic transmission element. The 
dotted line applies to the transmission unit and how it adjusts itself in 
the method of the invention. The dotted line shows that the unit operates 
with constant rotational speed during operation of the hydrodynamic clutch 
and that the differential between the two lines represents a loss in 
efficiency, which can be compensated by the method of the present 
invention. 
The illustration clearly shows a transmission design in which the 
hydrodynamic clutch is active in the transmission of power in the first 
and second gear. However, other designs are also feasible. 
The method of the present invention is suitable for any transmission unit 
that includes a hydrodynamic transmission section and a mechanical 
transmission section, wherein the hydrodynamic transmission section takes 
the form of a hydrodynamic clutch and wherein transfer of power occurs 
through the hydrodynamic clutch when changing gears from any optional 
first gear to any optional second gear. It is irrelevant if further tasks 
are assigned to the individual transmission components, for example, if 
the functions of the hydrodynamic clutch and the hydrodynamic retarder are 
assigned to the hydrodynamic transmission section, as is described in 
concrete terms in the German patent application no. 297 00 605. 
While this invention has been described as having a preferred design, the 
present invention can be further modified within the spirit and scope of 
this disclosure. This application is therefore intended to cover any 
variations, uses, or adaptations of the invention using its general 
principles. Further, this application is intended to cover such departures 
from the present disclosure as come within known or customary practice in 
the art to which this invention pertains and which fall within the limits 
of the appended claims.