Vehicle having rolling compensation
The present invention relates to a vehicle, in particular a rail vehicle, having a car body, a
first running gear, and a second running gear arranged at a distance from the first running
gear in the direction of a vehicle longitudinal axis, wherein the car body is supported on the
first running gear in the direction of a vehicle height axis by means of a first spring device,
and the car body is supported on the second running gear in the direction of the vehicle
height axis by means of a second spring device. The car body is coupled to the first running
gear by means of a first rolling compensation device, and is coupled to the second running
gear by means of a second rolling compensation device. The first rolling compensation
device and the second rolling compensation device counteract rolling motions of the car body
toward the outside of the curve about a rolling axis parallel to the vehicle longitudinal axis
during travel in curves. The present invention also concerns a method for setting rolling
angles on a car body of a vehicle.
On rail vehicles - but also on other vehicles - the car body is generally supported on the
wheel units, for example wheel pairs and wheelsets, via one or more spring stages. The
centrifugal acceleration generated transversely to the direction of motion and thus to the
vehicle longitudinal axis means that as a result of the comparatively high position of the
centre of gravity of the car body the car body has a tendency to roll towards the outside of
the curve in relation to the wheel units thus causing a rolling motion about a rolling axis
parallel to the vehicle longitudinal axis.
Such rolling motions detract from the travel comfort when they exceed certain limiting values.
In addition they also constitute a danger of breaching the permissible gauge profile and, in
terms of the tilt stability and thus also the derailment safety, a danger of inadmissible
unilateral wheel unloading. In order to prevent this, as a rule, rolling support mechanisms in
the form of so-called rolling stabilisers are used. The job of these is to offer a resistance to
the rolling motion of the car body in order to reduce the latter, but at the same time not
hindering the rising and dipping motion of the car body in relation to the wheel units.
Such rolling stabilisers are known in various hydraulically or purely mechanically operating
designs. Often a torsion shaft extending transversely to the vehicle longitudinal axis is used,
as known from EP 1 075 407 B1, for example. On this torsion shaft, on either side of the

vehicle longitudinal axis, levers secured against rotation are located, extending in the vehicle
longitudinal direction. These levers are in turn connected to rods which are arranged
kinematically in parallel with the suspension devices of the vehicle. When the springs of the
suspension devices of the vehicle deflect, the levers located on the torsion shaft are set in a
rotational motion by means of the rods to which they are connected.
If during travel in curves a rolling motion occurs with varying spring deflections of the
suspension devices on either side of the vehicle, this results in differing angles of rotation of
the levers located on the torsion shaft. The torsion shaft is thus loaded by a torsional
moment, which - depending on its torsional stiffness - at a certain torsional angle, it
compensates by a counter-moment resulting from its elastic deformation, thus preventing a
further rolling motion. On rail vehicles fitted with bogies the rolling support mechanism can
also be provided for the secondary suspension stage, i.e. between a running gear frame and
the car body. The rolling support mechanism can also be applied in the primary stage, i.e.
operating between the wheel units and a running gear frame or - in the absence of
secondary suspension - a car body.
Such rolling stabilisers are also used in generic rail vehicles, such as those known from
EP 1 190 925 A1. On the rail vehicle known from this document the upper ends of the two
rods of the rolling stabilisers (in a plane running perpendicularly to the vehicle longitudinal
axis) are displaced towards the centre of the vehicle. As a result of this the car body, in the
event of a deflection in the vehicle transverse direction (as is caused, for example, by the
centrifugal acceleration during travel in curves) is guided in such a way that a rolling motion
of the car body toward the outside of the curve is counteracted and a rolling motion directed
toward the inside of the curve is impoed upon it.
This rolling motion in the opposite direction serves, inter alia, to increase the so-called tilting
comfort for the passengers in the vehicle. A high tilting comfort is normally understood here
to be the fact that, during travel in curves, the passengers experience the lowest possible
transverse acceleration in the transverse direction of their reference system, which as a rule
is defined by the fixtures of the car body (floor, walls, seats, etc.). As a result of the tilting of
the car body towards the inside of the curve caused by the rolling motion the passengers
(depending on the degree of tilting) experience at least part of the transverse acceleration
actually acting in the earth-fixed reference system merely as increased acceleration in the
direction of the vehicle floor, which as a rule is perceived as less annoying or uncomfortable.

The maximum admissible values for the transverse acceleration acting in the reference
system of the passengers (and, ultimately, the resultant setpoint values for the tilt angles of
the car body) are as a rule specified by the operator of a rail vehicle. National and
international standards (such as for example EN 12299) also provide a starting point for this.
Here, with the vehicle from EP 1 190 925 A1, it is possible to create a purely passive system,
in which the components of the suspension and of the rolling stabilisers are adapted to each
other in such a way that the desired tilting of the car body is achieved solely by the
transverse acceleration acting during travel in curves.
For such a passive solution, firstly the rolling axis or the instantaneous centre of rotation of
the rolling motion must be comparatively far above the centre of gravity of the car body.
Secondly, the suspension in the transverse direction must be designed to be comparatively
soft, in order to achieve the desired deflections solely with the acting centrifugal force. Such
a transversely soft suspension also has a positive effect on the so-called vibration comfort in
the transverse direction, since impacts in the transverse direction can be absorbed and
dampened by the soft suspension.
These passive solutions have the disadvantage, however, that because of the transversely
soft suspension and the elevated instantaneous centre of rotation in normal operation, but
also in unplanned situations (e.g. an unexpected stopping of the vehicle on a curve with a
high cant) comparatively high transverse deflections in the transverse direction also result
meaning either that the typically specified gauge profile is breached or (in order to avoid this)
only comparatively narrow car bodies with reduced transport capacity can be constructed.
The problem of large deflections in order to achieve a certain rolling angle can indeed be
mitigated by shifting the rolling axis or the instantaneous centre of rotation. But this allows
only even lower rolling angles to be achieved passively. Consequently the system stiffens in
the transverse direction so that not only reductions in tilting comfort but also reductions in
vibration comfort have to be accepted.
The rolling motion adjusted for the bend of the curve currently being travelled and the current
running speed (and consequently also the resultant transverse acceleration) on the vehicle
from EP 1 190 925 A1 can also be influenced or set actively by an actuator connected
between the car body and the running gear frame. Here, from the current bend of the curve

and the current vehicle speed, a setpoint value is calculated for the rolling angle of the car
body, which is then used for setting the rolling angle by means of the actuator.
While this variant offers the opportunity of creating more transversely stiff systems with lower
transverse deflection, it has the disadvantage that the vibration comfort is impaired by the
transverse stiffness introduced by the actuator so that, for example, transverse impacts on
the running gear (for example when travelling over switches or imperfections in the track) are
transmitted to the car body with less damping.
In order to compensate for at least the disadvantages regarding vibration comfort by
transversely stiff suspension, in WO 90/03906 A1 for a passive system it is proposed that,
kinematically in series with the rolling compensation device, a comparatively short transverse
supplementary suspension stage is introduced. The disadvantage of this solution, however,
is that firstly due to the additional components it increases the installation space required,
and secondly the problems described above of large transverse deflections or reduced
transport capacity are present here again.
A further problem in connection with the use of such rolling support mechanisms is the
sensitivity of the vehicle to side winds. In particular in the area of a forward vehicle, and there
in particular in the area of the forward running gear, under the effects of the flow against the
vehicle transversely to the direction of travel caused by side wind, there is a force action on
the vehicle, the effective point of application of which (in the direction of travel) is usually
located in front of the centre of gravity (which is usually located in the longitudinal centre of
the car body).
This force action caused by side wind brings about a so-called yaw motion of the car body
(thus a rotation of the car body about its height axis), wherein the forward part of the car
body is deflected by the side wind, while the trailing part is rotated against the side wind. The
deflection continues until the restoring forces of the support of the car body on the running
gear balance out the yaw moment from the side wind.
The problem here is that this yaw motion of the car body generally results in a reduction in
the wheel loadings (and thus in a so-called wheel unloading) on each of the running gear
sides respectively of the two running gears. With the forward running gear there is a wheel
unloading to the side of the running gear turned towards the side wind (thus the windward

side of the running gear), which is intensified even further by the lift that normally acts in this
area.
Particularly when the rolling support mechanisms described are used, due to the opposing
transverse deflections and the opposing rolling deflections generated in the area of the two
running gears, torsion of the car body is caused which further intensifies the wheel
unloading. In particular on double-decker vehicles, due to the large impact surface area for
the side wind and the comparatively high position of the centre of gravity, a considerable
wheel unloading can occur, which should, however, for reasons of derailment safety, not
exceed specified limits.
In order to reduce the risk of derailment, for an existing vehicle up until now there has only
been the possibility of detecting the wind strength using suitable means and adapting the
vehicle speed accordingly. Alternatively a correspondingly low maximum vehicle speed has
been set, until the risk of derailment under the side wind intensity to be expected on the route
remains within the specified limits. Such reductions in the vehicle speed are naturally highly
undesirable for the vehicle operator.
The object for the present invention was therefore to provide a vehicle or a method of the
type mentioned initially, which does not have, or only to a limited extent, the disadvantages
mentioned above and in particular which, in a simple and reliable manner allows a reduced
sensitivity to side wind and, eventually, a high travel comfort for passengers with a high
transport capacity of the vehicle.
The present invention solves this problem on the basis of a vehicle according to the
preamble of claim 1 by means of the features indicated in the characterising part of claim 1. It
also solves this problem on the basis of a method according to the preamble of claim 24 by
means of the features indicated in the characterising part of claim 24.
The present invention is based on the technical teaching that, in a simple and reliable
manner, a reduced sensitivity to side wind or an increase in the permissible speed of the
vehicle can be achieved despite the use of rolling compensation devices, if a component of
the wheel unloading resulting from the torsion of the car body, as it results, for example, from
side wind, is at least reduced by active intervention on one of the two rolling support devices
and/or a coupling of the two rolling support devices. It has turned out that, by means of such
an active intervention, at one of the rolling support devices or a suitable mechanical and/or a

control system coupling of the two rolling support devices, in a simple manner, a reduction in
the torsion of the car body (as far as to a value of zero) is possible.
In this way it is possible in an advantageous fashion to at least in part compensate for the
disadvantageous properties of such rolling compensation devices from the side wind
sensitivity point of view, or possible even to completely eliminate them. In other words the
advantageous effects of such rolling compensation devices in terms of greater travel comfort
for passengers and high transport capacity of the vehicles can be readily achieved, without
significant reductions in terms of side wind sensitivity or the permissible maximum speed
having to be made.
According to a first aspect the present invention therefore relates to a vehicle, in particular a
rail vehicle, having a car body, a first running gear, and a second running gear arranged at a
distance from the first running gear in the direction of a vehicle longitudinal axis, wherein the
car body is supported on the first running gear in the direction of a vehicle height axis by
means of a first spring device, and the car body is supported on the second running gear in
the direction of the vehicle height axis by means of a second spring device. The car body is
coupled to the first running gear by means of a first rolling compensation device, and is
coupled to the second running gear by means of a second rolling compensation device. The
first rolling compensation device and the second rolling compensation device counteract
rolling motions of the car body toward the outside of the curve about a rolling axis parallel to
the vehicle longitudinal axis during travel in curves. The first rolling compensation device is
designed in such a way that a torsional load on the car body about the vehicle longitudinal
axis, which is in particular caused by wind loads acting on the car bodies, is counteracted.
Additionally or alternatively, the first rolling compensation device and the second rolling
compensation device are coupled in such a way that such a torsional load, in particular
caused by wind loads, is counteracted.
The torsional load on the car body can basically be counteracted in any suitable manner by
vehicle internal measures in the area of at least one of the two rolling compensation devices.
The first rolling compensation device is preferably designed in such a way as to impose upon
the car body under a first transverse deflection of the car body in relation to the first running
gear in the direction of a vehicle transverse axis a first rolling angle about the rolling axis,
while the second rolling compensation device is designed in such a way as to impose upon
the car body under a second transverse deflection of the car body in relation to the second
running gear in the direction of a vehicle transverse axis a second rolling angle about the

rolling axis. The first rolling compensation device is then designed to reduce the torsional
load on the car body in such a way that a deviation between the first transverse deflection
and the second transverse deflection and/or a deviation between the first rolling angle and
the second rolling angle is counteracted. In addition or as an alternative, the first rolling
compensation device and the second rolling compensation device are coupled together in
such a way that such a deviation between the first transverse deflection and the second
transverse deflection and/or a deviation between the first rolling angle and the second rolling
angle is counteracted.
At this point it is mentioned that, depending on the design of the rolling compensation device,
as a rule there is a specified relationship between the transverse deflection concerned and
the associated rolling angle, so that consideration of the transverse deflections and
consideration of the rolling angle can as the case may be represent equivalent or equal
measures.
The reduction in the torsional load on the car body can, as explained in more detail in the
following, be carried out by purely passive measures (i.e. measures without external supply
of energy). In advantageous variants of the vehicle according to the invention, however, an
active solution is realized. To this end, it is preferably provided that the first rolling
compensation device has a first actuator device with at least one first actuator unit controlled
by a control device. The first actuator device is preferably configured to contribute, controlled
by the control device, to the setting of the first transverse deflection in order to at least
reduce the deviation between the first transverse deflection and the second transverse
deflection and/or the deviation between the first rolling angle and the second rolling angle.
Additionally or alternatively, the second rolling compensation device has a second actuator
device with at least one second actuator unit controlled by the control device, wherein the
second actuator device is then preferably configured to contribute , controlled by the control
device, to the setting of the second transverse deflection in order to at least reduce the
deviation between the first transverse deflection and the second transverse deflection and/or
the deviation between the first rolling angle and the second rolling angle.
The active reduction or elimination of the torsional load is preferably achieved in that the
control device has at least one detection device to detect at least one detection variable,
which is representative of the torsional load applied to the car body. The control device is in
this case designed to control the first actuator unit and/or the second actuator unit in such a

way that the torsional load is reduced or, as the case may be, even substantially completely
eliminated.
Thus, for example, the control device can be configured to control the first actuator unit
and/or the second actuator unit in such a way that, in the direction of a vehicle transverse
axis, a deviation between a first transverse deflection of the car body in relation to the first
running gear and a second transverse deflection of the car body in relation to the second
running gear is reduced. Here, it is self-evident, of course, also that the corresponding rolling
angle of the car body in relation to the respective running gear may be focussed on.
The necessary degree of reduction in the deviation between the transverse deflections or the
rolling angles depends, in particular, on the design of the vehicle. Relevant influencing
variables here include the torsional stiffness of the car body about the vehicle longitudinal
axis and the distance between the two running gears in the direction of the vehicle
longitudinal axis. The stiffer the car body or the smaller the distance between the two running
gears, the smaller the deviation must be between the transverse deflections or the rolling
angles in order to achieve a specified reduction in the torsional load.
In preferred variants of the vehicle according to the invention, it is preferred that the control
device controls the first actuator unit and/or the second actuator unit according to the
detection variable in such a way that the deviation between the first transverse deflection and
the second transverse deflection is less than 40 mm, preferably less than 25 mm, further
preferably less than 10 mm. Additionally or alternatively, the control device can control the
first actuator unit and/or the second actuator unit as a function of the detection variable in
such a way that a deviation between a first rolling angle of the car body in relation to the first
running gear and a second rolling angle of the car body in relation to the second running
gear is less than 2°, preferably less than 1°, further preferably less than 0.5°. Here it is self-
evident that, as a rule, of course the most extensive possible reduction in the deviation
concerned is advantageous or desirable.
For the detection variable basically any variable can be determined which allows conclusions
to be made on the current torsional load on the car body and, thus, ultimately the wheel
unloading resulting from this torsional load to be made. For example, it is possible to
determine directly at the car body (for example by means of one or more strain gauge strips
or similar) a representative variable for the actual torsional load on the car body and to use
this for the further control of the active components. In further preferred variants of the

vehicle according to the invention it is provided that the detection device as the at least one
detection variable detects a variable representative of the first transverse deflection of the car
body and/or a variable representative of the second transverse deflection of the car body,
which is then used for the further control of the active components.
Additionally or alternatively, the detection device can detect as the at least one detection
variable a variable representative of a deflection of a component of the first rolling
compensation device and/or a variable representative of a deflection of a component of the
second rolling compensation device, which is then used for further control of the active
components.
It is once again mentioned at this point that the use of an active component in the area of just
one of the two rolling compensation devices may be sufficient. Thus, for a reduction in the
torsional load it may be sufficient, for example, that through active intervention on the forward
running gear the yaw moment on the vehicle body resulting from the wind load can be
counteracted in that the deflection of the car body is counteracted by a corresponding force
action in the area of the rolling compensation device of the forward running gear, while the
deflection in the trailing running gear is allowed.
Of course it is likewise possible in the area of the trailing running gear by means of active
intervention to counteract the yaw moment on the car body resulting from the wind load in an
isolated manner, by counteracting the deflection of the car body by a corresponding force of
action in the area of the rolling compensation device of the trailing running gear, while the
deflection on the forward running gear is allowed.
Finally, a combination of both variants can of course be provided, in which in the area of both
rolling compensation devices a coordinated active intervention takes place. This is an
advantage in particular with regard to the design of the active components, since these must
then be designed for a correspondingly lower power.
It is further mentioned at this point that the control device can be designed through suitable
measures such that the influences caused by side wind described above can be
distinguished from other vehicle dynamics influences (e.g. entry to and exit from track
superelevations, changes in the radius of curvature of the track, etc.). For this corresponding
filters as well as previously generated models of the vehicle can be used. Here, in particular,
account can be taken of the fact that influences caused by side winds have a quasi-static

nature and consequently occur in a comparatively low frequency range, which is, as a rule,
less than 2 Hz, so that, in particular, a differentiation of higher frequency dynamic influences
is as a rule possible without problems.
As has already been mentioned above, additionally or alternatively to the active solution
described above, a passive reduction of the torsional load on the car body can also be
provided. This can be achieved by a corresponding mechanical coupling of the two rolling
compensation devices. In preferred variants of the vehicle according to the invention it is
provided that the first rolling compensation device and the second rolling compensation
device are coupled together mechanically by means of a passive coupling device, wherein
the coupling device, in order to reduce the torsional load on the car body, in the direction of a
vehicle transverse axis generates concurrent adjusting movements in the area of the first
rolling compensation device and the second rolling compensation device.
The mechanical coupling between the two rolling compensation devices can be created in
any suitable fashion. Thus, for example, any mechanical gearing can be used to create this
coupling. In particularly advantageous variants of the invention the coupling is at created at
least in sections by means of a fluidic operating principle, since in this way a particularly
simple, space-saving design for bridging the distance between the running gears is possible.
Preferably, the coupling device therefore comprises a fluidic coupling between the first rolling
compensation device and the second rolling compensation device.
In further advantageous variants of the invention the desired high travel comfort for the
passengers with high transport capacity of the vehicle is made possible by selecting an
active solution with an active first rolling compensation device, which in particular can be
arranged kinematically parallel to the first spring device. The first rolling compensation
device, in order to increase the tilting comfort, is designed to impose upon the car body, in a
first frequency range and under a first transverse deflection of the car body in the direction of
the vehicle longitudinal axis, a first rolling angle component of the first rolling angle, which
corresponds to a current curvature of a current section of track being travelled. Additionally
or alternatively, the first rolling compensation device can be designed to impose upon the car
body in a second frequency range, which at least partially lies above the first frequency
range, a second transverse deflection component (as the case may be, therefore, also a
second rolling angle component about the rolling axis). In this way, the transverse deflection
component resulting from the first rolling angle component, the setting of which ultimately
represents a quasi-static adaptation of the rolling angle and thus the transverse deflection to

the current track curvature and the current speed, can be overlaid with a second transverse
deflection component (as the case may be, therefore, also a second rolling angle
component), the setting of which ultimately represents a dynamic adaptation to current
disturbances introduced into the car body.
While by means of the first rolling angle component and thus the first transverse deflection
component in the first frequency range, an increase in the tilting comfort is achieved, by
means of the second transverse deflection component (and as the case may be the second
rolling angle component) in the second frequency range (which at least partially lies above
the first frequency range) in an advantageous manner an increase in the vibration comfort is
achieved. By the design of the rolling compensation device as an active system in at least
the second frequency range, in an advantageous manner it is possible to design the support
of the car body on the running gear in the transverse direction of the vehicle to be
comparatively stiff, in particular to position the rolling axis or the instantaneous centre of
rotation of the car body comparatively close to the centre of gravity of the car body, so that
firstly the desired rolling angle is associated with relatively low transverse deflections and
secondly in the event of a failure of the active components the most passive possible
restoration of the car body to a neutral position is possible. These low transverse deflections
in normal operation and the passive restoration in the event of a fault allow in an
advantageous manner particularly broad car bodies with a high transport capacity to be built.
The active solution here has the particular advantage that all functions, therefore the
reduction in the sensitivity to side wind, the increase in tilting comfort, and the increase in
vibration comfort, can be achieved by correspondingly designed, overlaid control algorithms
in the control unit, which as the case may be have to control just a single active device in the
area of at least one of the rolling compensation devices. In other words, this allows a high
level of functional integration and/or a very compact design to be achieved, which in
particular with regard to the limited space which is available in any case in modern running
gears is a particular advantage.
Mention is made at this point of the fact that the second rolling compensation device, as the
case may be, can also have a different design from the first rolling compensation device. In
particular, however, the first rolling compensation device and the second rolling
compensation device are substantially of the same design, so that the following statements
concerning the features, functions and advantages of the first rolling compensation device
can equally be made in relation to the second rolling compensation device.

In this connection it is further noted that the second transverse deflection component,
depending on the design and the connection of the rolling compensation device, as the case
may be, does not necessarily have to be associated with a second rolling angle component
corresponding to the (static) kinematics of the first rolling compensation device, which is
overlaid on the first rolling angle component in the second frequency range. This is because,
for example with a comparatively soft, elastic connection of the first rolling compensation
device to the first running gear and/or the car body, as a result of the forces of inertia in the
second frequency range, within certain limits a kinematic decoupling of the transverse
movements of the car body from the rolling motion specified by the kinematics of the rolling
compensation device (for slow, quasi-static motions) occurs. Therefore, the more rigidly the
connection of the rolling compensation device to the running gear is created and the more
inherently rigid the design of the rolling compensation device is, the less this decoupling
takes place. Therefore, the first rolling angle component, in a design with a rigid coupling to
an inherently rigid rolling compensation device, in the second frequency range is ultimately
overlaid by a second rolling angle component.
In further preferred variants of the invention the first rolling compensation device, in order to
increase the tilting comfort, is designed such that it imposes on the car body, in a first
frequency range under a first transverse deflection component of the first transverse
deflection of the car body, a first rolling angle component of the first rolling angle, which
corresponds to a current curvature of a current section of track being travelled. Furthermore,
the first rolling compensation device, in order to increase the vibration comfort, is designed
such that it imposes on the car body, in a second frequency range, a second transverse
deflection component overlaid on the first transverse deflection component, wherein the
second frequency range at least partially, in particular completely, lies above the first
frequency range.
The first rolling compensation device can thus be designed such that it is active only in the
second frequency range, and thus only actively sets the second transverse deflection
component or, as the case may be, the second rolling angle component, while the setting of
the first rolling angle component is brought about purely passively as a result of the
transverse acceleration or the resulting centrifugal force acting on the car body during travel
in curves. It is similarly possible, however, in both frequency ranges, to bring about an at
least partially active setting of the rolling angle and the transverse deflection, respectively, by
means of the rolling compensation device, which is, as the case may be, supported by the

centrifugal force. Finally, it can also be provided that the setting of the rolling angle or the
transverse deflection is performed exclusively actively by means of the first rolling
compensation device. This is the case if the rolling axis or the instantaneous centre of
rotation of the car body is positioned at or near the centre of gravity of the car body, so that
the centrifugal force cannot make any (or at least no significant) contribution to the
generation of the rolling motion and the transverse deflection, respectively.
The first rolling compensation device can basically be designed in any manner. The first
rolling compensation device preferably comprises an actuator device with at least one
actuator unit controlled by a control device, the actuator force of which provides at least part
of the force for setting the rolling angle or the transverse deflection on the car body. With an
at least partially active setting of the rolling angle or the transverse deflection in the first
frequency range, the actuator device is designed to make at least a majority contribution to
the generation of the first rolling angle component in the first frequency range, in particular, to
substantially generate the first rolling angle component and the first transverse deflection
component, respectively.
The first frequency range, preferably, is the frequency range in which quasi static rolling
motions corresponding to the current curvature of the section of track being travelled and the
current running speed. This frequency range can vary according to the requirements of the
rail network and/or the vehicle operator (for example due to the use of the vehicle for local
travel or long-distance travel, in particular high-speed travel). The first frequency range
preferably ranges from 0 Hz to 2 Hz, preferably from 0.5 Hz to 1.0 Hz. The same applies to
the bandwidth of the second frequency range, wherein this is of course matched to the
dynamic disturbances to be expected during operation of the vehicle (as the case may be
periodic, but typically singular or statistically scattered), which are noticed by the passengers
and perceived as annoying. The second frequency range therefore preferably ranges from
0.5 Hz to 15 Hz, preferably from 1.0 Hz to 6.0 Hz.
Basically it can be provided that the active setting that takes place (at least in the second
frequency range) of the rolling angle and the transverse deflection, respectively, takes place
via the rolling compensation device exclusively during travel in curves on the curved track,
and therefore the first rolling compensation device is active only in such a travel situation.
Preferably, it is however provided that the rolling compensation device is also active during
straight travel, so that the vibration comfort in an advantageous manner is also guaranteed in
these travel situations.

In preferred variants of the vehicle according to the invention, by means of the first rolling
compensation device, a limitation of the transverse deflections of the car body (thus the
deflections in the vehicle transverse direction) in relation to a neutral position is carried out.
The neutral position is defined by the position of the car body which it adopts when the
vehicle is at a standstill on a straight and level track. In this way it is possible in an
advantageous way, to build particularly wide car bodies with high transport capacity, which
are matched to the gauge profile specified by the operator of the rail vehicle. The limitation of
the transverse deflections can be performed by any suitable components of the rolling
compensation device. Preferably, an actuator device of the first rolling compensation device
provides the limitation of the transverse deflections, since in this way a particularly compact,
space-saving design can be achieved.
As mentioned, the limitation of the transverse deflections can be matched to the gauge
profile specified by the operator of the vehicle. Particularly advantageous designs result if the
first rolling compensation device, in particular an actuator device of the first rolling
compensation device, is designed in such a way that a first maximum transverse deflection
of the car body from the neutral position occurring toward the outside of the curve during
travel in curves in the vehicle transverse direction is limited to 80 mm to 150 mm, preferably
100 mm to 120 mm. While, with regard to complying with the specified gauge profile,
limitation of the transverse deflections in vehicles with (in the longitudinal direction of the
vehicle) running gears arranged centrally below the car bodies is of particular importance, in
vehicles with running gears arranged in the end area of the car bodies it is of particular
interest to correspondingly limit the transverse deflections toward the inside of the curve.
Preferably, therefore, additionally or alternatively, a second maximum transverse deflection
of the car body from the neutral position occurring toward the inside of the curve during travel
in curves in the vehicle transverse direction is limited to 0 mm to 40 mm, preferably 20 mm. It
is self-evident that, with certain variants of the invention, it can also be provided that a
second maximum transverse deflection of the car body from the neutral position toward the
inside of the curve during travel in curves can also have a negative value, for example -
20 mm. In this case the car body will therefore also be deflected on the inside of the curve to
the outside of the curve, in order, for example, to adhere to a specified gauge profile with
particularly wide car bodies.
As already mentioned, the limitation of the transverse deflections can preferably be
performed by an actuator device of the first rolling compensation device. Here it is preferably

provided that the actuator device is designed to act as an end stop device for definition of at
least one end stop for the rolling motion of the car body. To this end, a stop defined by the
design of the actuator device (for example a simple mechanical stop) can be provided.
Preferably, the actuator device is designed to define the position of the at least one end stop
for the rolling motion of the car body in a variable fashion. In other words, it can be provided
that this stop by actively restraining the actuator device (for example by corresponding
energy provision to the actuator device) and/or passively restraining the actuator device (for
example by deactivating a self-restraining design actuator device) is freely definable at any
position in the adjusting path of the actuator device.
The actuator device of the first rolling compensation device can basically be designed in any
suitable manner. Preferably, it is provided that the actuator device in the event of its inactivity
offers at most only slight resistance, in particular substantially no resistance, to a rolling
motion of the car body. Consequently the actuator device is preferably not designed to be
self-restraining, so that in the event of a failure of the actuator device inter alia a restoration
of the car body to its neutral position is ensured.
In preferred variants of the vehicle according to the invention the first rolling compensation
device is designed in such a way that, even in the event of failure of the active components
of the first rolling compensation device, emergency operation of the vehicle with, as the case
may be, degraded comfort characteristics (in particular with regard to tilting comfort and/or
vibration comfort) is still possible while complying with the specified gauge profile.
Preferably, therefore, it is provided that the spring device, when an actuator device of the first
rolling compensation device is inactive, exerts a restoring moment on the car body about the
rolling axis, wherein the restoring moment is dimensioned such that, in the event of an
inactive actuator device, a transverse deflection of the car body from the neutral position for
a stationary vehicle under a nominal loading of the car body and with a maximum permitted
track superelvation is less than 10 mm to 40 mm, preferably less than 20 mm. In other
words, the spring device (in particular its stiffness in the vehicle transverse direction) is
preferably designed so that a vehicle which for any reason (for example due to damage to
the vehicle or to the track) comes to a standstill at an unfavourable spot, as before complies
with the specified gauge profile.
Additionally or alternatively it can be provided that the restoring moment in the event of an
inactive actuator device is dimensioned such that a transverse deflection of the car body

from the neutral position, under nominal loading of the car body and with a maximum
permitted transverse acceleration of the vehicle acting in the direction of a vehicle transverse
axis, is less than 40 mm to 80 mm, preferably less than 60 mm. In other words the spring
device (in particular its stiffness in the vehicle transverse direction) is preferably designed so
that a vehicle, in emergency operation in the event of failure of the actuator device, when
travelling at normal running speed, as before complies with the specified gauge profile.
The stiffness, in particular the transverse stiffness in the vehicle transverse direction, of the
support of the car body on the respective running gear can have any suitable characteristic
as a function of the transverse deflection. Thus, for example, a linear or even progressive
behaviour of the stiffness as a function of the transverse deflection can be provided.
Preferably, however, a degressive behaviour is provided so that an initial transverse
deflection of the car body from the neutral position experiences a comparatively high
resistance, this resistance decreasing however as the deflection increases. With regard to
the dynamic setting of the second rolling angle in the second frequency range during travel in
curves, this is an advantage, however, since the first rolling compensation device has to
make available lower forces for these dynamic deflections in the second frequency range.
It is preferably provided, therefore, that the spring device defines a restoring characteristic
line, wherein the restoring characteristic line represents the dependence of the restoring
moment on the rolling angle deflection and the restoring characteristic line has a degressive
behaviour. The behaviour of the restoring characteristic line here can basically be adapted in
any suitable manner to the current application. Preferably, the restoring characteristic line, in
a first rolling angle range and a first transverse deflection range, respectively, has a first
inclination and, in a rolling angle range above the first rolling angle range and a transverse
deflection range above the first transverse deflection range, respectively, has a second
inclination that is less than the first inclination, wherein the ratio of the second inclination to
the first inclination is in particular in the range from 0 to 1, preferably in the range from 0 to
0.5. The two rolling angle ranges and transverse deflection ranges, respectively, can be
selected in any suitable manner. Preferably, the first transverse deflection range ranges from
0 mm to 60 mm, preferably from 0 mm to 40 mm, and the second transverse deflection
range, in particular, ranges from 20 mm to 120 mm, preferably from 40 mm to 100 mm. The
rolling angle ranges, as a function of the given kinematics, then correspond to the transverse
deflection ranges.

Here it is self-evident that the determination of the characteristic of the spring device is
predominantly directed towards the transverse deflections, which, in the event of a failure of
active components, should still be achieved. The first inclination here, as a rule, defines the
residual transverse deflection in the event of failure of an active component, while the second
inclination determines the actuator forces for larger deflections and is, as far as possible,
selected such that these actuator forces in the event of large deflections can be kept low.
The second inclination is therefore preferably kept as close as possible to the value of zero.
As the case may be negative values of the second inclination are even possible or may be
provided.
In order to achieve the described restoring of the car body to its neutral position, the support
for the car body on the running gear can have any suitable stiffness. Here a stiffness that is
substantially independent of the transverse deflection can be provided for. Preferably,
however, it is again provided that the respective spring device has a transverse stiffness in
the direction of a vehicle transverse axis, which is dependent upon a transverse deflection of
the car body from the neutral position in the direction of the vehicle transverse axis, so that
for deflections in the vicinity of the neutral position another stiffness (for example a higher
stiffness) prevails than in the area of larger deflections. In this way the advantages described
above in terms of dynamic setting of the second rolling angle during travel in curves can
again be achieved.
The respective spring device, preferably, in a first transverse deflection range, has a first
transverse stiffness, while, in a second transverse deflection range above the first transverse
deflection range, it has a second transverse stiffness, which is lower than the first transverse
stiffness. Here it is self-evident that the transverse stiffness can vary within the respective
transverse deflection range. In addition, the behaviour of the transverse stiffness according
to the transverse deflection can basically be adapted in any suitable manner for the current
application.
Preferably, the first transverse stiffness is in the range 100 N//mm to 800 N/mm, further
preferably in the range 300 N/mm to 500 N/mm, while the second transverse stiffness is
preferably in the range 0 N/mm to 300 N/mm, further preferably in the range 0 N/mm to
100 N/mm. The two transverse deflection ranges can likewise be selected in any suitable
manner adapted to the respective application. The first transverse deflection range preferably
ranges from 0 mm to 60 mm, preferably from 0 mm to 40 mm, while the second transverse
deflection range preferably ranges from 20 mm to 120 mm, further preferably from 40 mm to

100 mm. In this way, with regard to a limitation of the maximum transverse deflection of the
car body with the lowest possible use of energy, particularly good designs can be achieved.
The advantageous behaviour of the vehicle already described above in the absence of one
or more active components of the rolling compensation device can preferably be achieved by
means of a corresponding design of the respective spring device, in particular of its
transverse stiffness.
Preferably, therefore, for a favourable behaviour in such emergency operation of the vehicle,
it is provided that the respective spring device in the direction of a vehicle transverse axis
has a transverse stiffness, wherein the transverse stiffness of the spring device is
dimensioned such that, in the event of inactivity of an actuator device of the rolling
compensation device during travel in curves with a maximum permissible transverse
acceleration of the vehicle operating in the direction of a vehicle transverse axis, a first
maximum transverse deflection of the car body from the neutral position toward the outside
of the curve in a vehicle transverse direction is limited to 40 mm to 120 mm, preferably to
60 mm to 80 mm. Additionally or alternatively it is provided that a second maximum
transverse deflection of the car body from the neutral position toward the inside of the curve
in a vehicle transverse direction is limited to 0 mm to 60 mm, preferably to 20 mm to 40 mm.
The rolling angle ranges then again, as a function of the given kinematics, correspond to the
above transverse deflection ranges.
Furthermore, additionally or alternatively, (with regard to a favourable behaviour for a
stationary vehicle) it can be provided that the transverse stiffness of the spring device is
dimensioned such that, in the event of inactivity of an actuator device of the respective rolling
compensation device, a transverse deflection (and, thus, a corresponding rolling angle
deflection) of the car body from the neutral position under nominal loading and with a
maximum permitted track superelevation is less than 10 mm to 40 mm, preferably less than
20 mm.
The active components of the respective rolling compensation device can basically be
designed in any way. Preferably, (as already mentioned) at least one actuator device is
provided, which is connected between the car body and the running gear and performs the
setting of the rolling angle in the second frequency range. Due to their particularly simple and
robust design, preference is for the use of linear actuators, for which, preferably, the travel
and the actuator forces are limited in a suitable manner in order to meet the dynamics

requirements of the setting of the transverse deflection and the rolling angle in the second
frequency range, respectively, with satisfactory results.
In variants of the vehicle according to the invention with particularly favourable dynamic
properties, the rolling compensation device is designed in such a way that an actuator device
of the respective rolling compensation device, in the first frequency range, has a maximum
deflection from the neutral position of 60 mm to 110 mm, preferably 70 mm to 85 mm, while,
additionally or alternatively, in the second frequency range, from a starting position, it has a
maximum deflection of 10 mm to 30 mm, preferably 10 mm to 20 mm. Furthermore, with
regard to the maximum actuator force, it can be provided that the actuator device, in the first
frequency range, exerts a maximum actuator force of 10 kN to 40 kN, preferably 15 kN to
30 kN, while, in the second frequency range, it exerts a maximum actuator force of 5 kN to
35 kN, preferably 5 kN to 20 kN.
In preferred variants of the vehicle according to the invention, the distance (in the neutral
position of the car body) between the rolling axis of the car body and the centre of gravity of
the car body in the direction of the vehicle height axis is adapted to the respective
application. Thus, the centre of gravity of the car body, as a rule, has a first height (H1)
above the track (typically above the upper surface of the rail SOK), while the rolling axis, in
the neutral position, in the direction of the vehicle height axis has a second height (H2) above
the track. Preferably, the ratio of the difference between the second height and the first
height (H2 to H1) to the first height (H1) is a maximum of 2.2, preferably a maximum of 1.3,
further preferably 0.8 to 1.3. The difference between the second height and the first height
(H2 - H1), in particular, can be between 1.5 m and 4.5 m, preferably 1.8 m. This allows
designs to be realized which, with regard to the limitation of the transverse deflections
already mentioned above and thus the feasibility of wide car bodies with high transport
capacity, are particularly favourable.
The respective rolling compensation device can basically be designed in any suitable
manner, in order to carry out the setting of the rolling angle of the car body in the two
frequency ranges. In particularly simple design variants of the vehicle according to the
invention it is provided to this end that the respective rolling compensation device comprises
a rolling support device, which is arranged kinematically in parallel to the spring device and is
designed to counteract rolling motions of the car body about the rolling axis when travelling in
a straight track. Such rolling support devices are sufficiently known, and so no further details
of them will be provided here. They can in particular be based on differing operating

principles. Thus, they may be based on a mechanical operating principle. But fluidic (for
example hydraulic) solutions, electromechanical solutions or any combination of all these
operating principles are also possible.
In a particularly simple design variant, the rolling support device comprises two rods, each of
which at one end is connected in an articulated manner to the car body and each of which at
the other end is connected in an articulated manner to opposing ends of a torsion element,
which is supported by the running gear, as has already been described at the outset.
Additionally or alternatively the respective rolling compensation device can also comprise a
guiding device, which is arranged kinematically in series with the spring device. The guiding
device comprises a guiding element, which is arranged between the running gear and the car
body and is designed such that, during rolling motions of the car body, it defines a motion of
the guiding element in relation to the car body or the running gear. Again, the guiding device
can have any suitable design in order to perform the guidance described. Thus it can for
example be created with the sliding and/or rolling of the guiding element on a guideway.
In particularly simply designed and robust variants of the vehicle according to the invention
the guiding device, in particular, comprises at least one multilayered spring. The multilayered
spring can be created as a simple rubber multilayered spring, the layers of which are
arranged to be inclined with respect to the vehicle height axis and to the vehicle transverse
axis, so that they define the rolling axis of the car body.
Here, it is pointed out that the design of the respective rolling compensation device with such
a multilayered spring device for definition of the rolling axis of the car body constitutes an
individually patentable inventive idea, which is, in particular, independent of the setting
described above of the rolling angle in the first frequency range and the second frequency
range.
The present invention can be used in association with any designs of the support of the car
body on the running gear. Thus, for example, it can be used in connection with a single stage
suspension, which supports the car body directly on the wheel unit. Particularly
advantageously it can be used in connection with two-stage suspension designs. Preferably,
the running gear accordingly comprises at least one running gear frame and least one wheel
unit, while the spring device has a primary suspension and a secondary suspension. The
running gear frame is supported via the primary suspension on the wheel unit, while the car

body is supported via the secondary suspension, which is, in particular, designed as
pneumatic suspension, on the running gear frame. The rolling compensation device is then
preferably arranged kinematically in parallel to the secondary suspension between the
running gear frame and the car body. This allows integration into the majority of vehicles
typically used.
The stiffness of the respective spring device, in particular, its transverse stiffness can, as the
case may be, be determined solely by the primary suspension and the secondary
suspension. In particular, the spring device comprises a transverse spring device, which, in
an advantageous manner, serves to adapt or optimise the transverse stiffness of the spring
device for the respective application. This simplifies the design of the spring device
considerably despite the simple optimisation of the transverse stiffness. The transverse
spring device can be connected at one end to the running gear frame and at the other end to
the car body. Additionally or alternatively the transverse spring device can also be connected
at one end to the running gear frame or to the car body and at the other to the rolling
compensation device.
The transverse spring device is preferably designed to increase the stiffness of the .
respective spring device in the direction of the vehicle transverse axis. Here it can have any
characteristic adapted for the respective application. The transverse spring device,
preferably, has a degressive stiffness characteristic, in order to achieve an overall degressive
stiffness characteristic of the spring device.
In preferred examples of the vehicle according to the invention it is further provided that the
respective spring device has an emergency spring device, which is arranged centrally on the
running gear, in order that, even if the supporting components of the spring device fail,
emergency operation of the vehicle is possible. The emergency spring device can basically
be designed in any manner. Preferably the emergency spring device is designed such that it
supports the compensation effect of the rolling compensation device. To this end, the
emergency spring device can comprise a sliding or rolling guide which follows the
compensation motion.
The present invention also relates to a method for setting rolling angles on a car body of a
vehicle, in particular a rail vehicle, about a rolling axis parallel to the vehicle longitudinal axis
of the vehicle, in which a first rolling angle and/or a first transverse deflection of the car body
is set in relation to a first running gear and a second rolling angle and/or a second transverse

deflection of the car body is set in relation to a second running gear which, in the direction of
a vehicle longitudinal axis, is arranged at a distance from the first running gear. The car body
is coupled to the first running gear via a first rolling compensation device, while the car body
is coupled via a second rolling compensation device with the second running gear. The first
rolling compensation device and the second rolling compensation device, during travel in
curves, counteract rolling motions of the car body toward the outside of the curve about a
rolling axis parallel to the vehicle longitudinal axis. The first rolling angle and/or the second
rolling angle are set in a manner coupled together such that a torsional load on the car body
about the vehicle longitudinal axis is counteracted. Additionally or alternatively, the first
transverse deflection and/or the second transverse deflection are set in a manner coupled
together such that a torsional load on the car body about the vehicle longitudinal axis is
counteracted. In this way the variants and advantages described above in connection with
the vehicle according to the invention can be achieved to the same extent, so that in this
context reference is made to the above statements.
Further preferred examples of the invention become apparent from the dependent claims or
the following description of preferred embodiments which refers to the attached drawings. It
is shown in:
Figure 1 a schematic sectional view of a preferred embodiment of the vehicle according
to the invention in the neutral position (along the line l-l from Figure 3);
Figure 2 a schematic sectional view of the vehicle from Figure 1 during travel in curves;
Figure 3 a schematic side view of the vehicle from Figure 1;
Figure 4 a schematic perspective view of part of the vehicle from Figure 1;
Figure 5 a transverse force-deflection-characteristic of the spring device of the vehicle
from Figure 1;
Figure 6 a schematic sectional view of a further preferred embodiment of the vehicle
according to the invention in the neutral position;
Figure 7 a schematic sectional view of a further preferred embodiment of the vehicle
according to the invention in the neutral position;

Figure 8 a schematic view of part of a further preferred embodiment of the vehicle
according to the invention.
First embodiment
In the following, by reference to Figures 1 to 5, a first preferred embodiment of the vehicle
according to the invention in the form of a rail vehicle 101, having a vehicle longitudinal axis
101.1, is described.
Figure 1 shows a schematic sectional view of the vehicle 101 in a sectional plane
perpendicular to the vehicle longitudinal axis 101.1. The vehicle 101 comprises a car body
102, which in the area of its first end is supported by means of a first spring device 103 on a
running gear in the form of a first bogie 104 and in the area of its second end is supported by
means of a second spring device 113 on a second running gear in the form of a second
bogie 114. The first bogie 104 and the second bogie 114 have an identical design, so that
the following will primarily deal with the features of the first bogie 104. The same applies to
the first spring device 103 and the second spring device 113. It is self-evident, however, that
the present invention can also be used with other configurations in which other running gear
designs are employed.
For ease of understanding of the explanations that follow, in the figures a vehicle coordinate
system xf, yf, zt (determined by the wheel contact plane of the bogie 104 or 114) is indicated,
in which the xf coordinate denotes the longitudinal direction of the rail vehicle 101, the yf
coordinate the transverse direction of the rail vehicle 101 and the zf coordinate the
perpendicular direction of the rail vehicle 101. Additionally an absolute coordinate system x,
y, z (determined by the direction of the gravitational force G) and a passenger coordinate
system xp, yp, zp (determined by the car body 102) are defined.
The bogie 104 comprises two wheel units in the form of wheelsets 104.1, each of which via
the primary suspension 103.1 of the first spring device 103 supports a bogie frame 104.2.
The car body 102 is again supported via a secondary suspension 103.2 on the bogie frame
104.2. The primary suspension 103.1 and the secondary suspension 103.2 are shown in
simplified form in Figure 1 as helical springs. It is self-evident, however, that the primary
suspension 103.1 or the secondary suspension 103.2, can be any suitable spring device. In

particular, the secondary suspension 103.2 preferably is a pneumatic suspension or similar
that is sufficiently well known.
The vehicle 101 also comprises in the area of the first bogie 104 a first rolling compensation
device 105 and in the area of the second bogie 114 a second rolling compensation device
115. Again the first rolling compensation device 105 and the second rolling compensation
device 115 have an identical design so that, in the following, it is primarily the features of the
first rolling compensation device 105 that will be considered. The first rolling compensation
device 105 works kinematically in parallel with the secondary suspension 103.2 between the
bogie frame 104.2 and the car body 102 in the manner described in more detail below.
As can be inferred, in particular, from Figure 1, the first rolling compensation device 105
comprises a sufficiently known rolling support 106, which on the one hand is connected with
the bogie frame 104.2 and on the other with the car body 102. Figure 4 shows a perspective
view of this rolling support 106. As can be inferred from Figure 1 and Figure 4, the rolling
support 106 comprises a torsion arm in the form of a first lever 106.1 and a second torsion
arm in the form of a second lever 106.2. The two levers 106.1 and 106.2 are located on
either side of the longitudinal central plane (xf,zf plane) of the vehicle 101 in each case
secured against rotation on the ends of a torsion shaft 106.3 of the rolling support 106. The
torsion shaft 106.3 extends in the transverse direction (yf direction) of the vehicle and is
rotatably supported in bearing blocks 106.4, which for their part are firmly attached to the
bogie frame 104.2. At the free end of the first lever 106.1 a first rod 106.5 is attached in an
articulated manner, while on the free end of the second lever 106.2 a second rod 106.6 is
attached in an articulated manner. By means of these two rods 106.5,106.6 the rolling
support 106 is connected in an articulated manner with the car body 102.
In Figures 1 and 4 the state in the neutral position of the vehicle 101 is shown, which results
from travelling on a straight track 108 with no twists. In this neutral position the two rods
106.5, 106.6 run in the drawing plane of Figure 1 (yfzf plane), in the present example inclined
to the height axis (zf axis) of the vehicle 101 in such a way that their top ends (connected in
an articulated manner to the car body 102) are displaced towards the centre of the vehicle
and their longitudinal axes intersect at a point MP, which lies in the longitudinal central plane
(XfZf plane) of the vehicle. By means of the rods 106.5, 106.6 in a sufficiently known manner
a rolling axis running parallel to the vehicle longitudinal axis 101.1 (in the neutral position) is
defined which runs through the point MP. The point of intersection MP of the longitudinal

axes of the rods 106.5,106.6 in other words constitutes the instantaneous centre of rotation
of a rolling motion of the car body 102 about this rolling axis.
The rolling support 106 allows in a sufficiently known manner synchronous dip by the
secondary suspension 103.2 on either side of the vehicle, while preventing a pure rolling
motion about the rolling axis or the instantaneous centre of rotation MP. Furthermore, as can
be inferred in particular from Figure 2, because of the inclination of the rods 106.5, 106.6 the
rolling support 106 kinematics with a combined motion of a rolling motion about the rolling
axis or the instantaneous centre of rotation MP and a transverse motion in the direction of
the vehicle transverse axis (yf axis) is predefined. Here, it is self-evident that the point of
intersection MP and thus the rolling axis because of the kinematics predefined by the rods
106.5, 106.6, when there is a deflection of the car body 102 from the neutral position, as a
rule will likewise experience a lateral shift.
Figure 2 shows the vehicle 101 during travel in curves on a track superelevation. As can be
inferred from Figure 2, the centrifugal force Fy acting upon the centre of gravity SP of the car
body 102 (because of the prevailing acceleration in the vehicle transverse direction) causes
on the bogie frame 104.2 a rolling motion toward the outside of the curve, which results from
a larger dip of the primary suspension 103.1 on the outside of the curve.
As can further be inferred from Figure 2, the described design of the rolling support 106
during the travel in curves of the vehicle 101 in the area of the secondary suspension 103.2
brings about a compensation motion, which counteracts the rolling motion of the car body
102 (in relation to the neutral position indicated by the broken contour 102.1 on a straight,
level track) toward the outside of the curve, which in the absence of the rolling support 106
because of the centrifugal force impinging on the centre of gravity SP of the car body 102
(similar to uneven suspension by the primary suspension 103.1) would arise from larger dip
of the secondary suspension 103.2 on the outside of the curve.
Thanks to this compensation motion that is predefined by the kinematics of the rolling
support 106, inter alia the tilting comfort for the passengers in the vehicle 101 is increased,
since the passengers (in their reference system xp, yp, zp defined by the car body 102) notice
a part of the transverse acceleration ay or centrifugal force Fy currently acting in the earth-
fixed reference system merely as an increased acceleration component azp and force action
Fzp, respectively, in the direction of the floor of the car body 102, which as a rule is perceived
as less annoying or uncomfortable. The transverse acceleration component ayp and

centrifugal component Fyp, respectively, acting in the transverse direction perceived by
passengers in their reference system as annoying is thus recued in an advantageous
manner.
The maximum permitted values for the transverse acceleration ayp,max acting in the reference
system (xp, yp, zp) for passengers are as a rule specified by the operator of the vehicle 101.
The starting points for this are also provided by national and international standards (such as
for example EN 12299).
The transverse acceleration ayp acting in the reference system (xp, yp, zp) for passengers (in
the direction of the yp axis) is comprised two components, namely a first acceleration
component aypsand a second acceleration component ayPd according to the equation:

The current value of the first acceleration component ayps is a result of travelling the current
curve at the current running speed, while the current value of the second acceleration
component aypd is the result of current (periodic or usually singular) events (such as for
example passing a disruptive part of the track, such as switches or similar).
Since the curvature of the curve and the current running speed of the vehicle 101 in normal
operation change only comparatively slowly, with this first acceleration component aypsis a
quasi static component. Conversely, the second acceleration component aypd (which usually
occurs as a result of impacts) is a dynamic component.
From the current transverse acceleration a^, according to the present invention it is
ultimately possible to determine a minimum setpoint value for a transverse deflection
dyN.soii mm of the car body 102 from the vehicle height axis (zr axis). This is the transverse
deflection (and thus as the case may be the corresponding rolling angle), which is the
minimum necessary in order keep below the maximum permissible transverse acceleration
ayp, max- Depending on how high the level of comfort for the passengers of the vehicle 101
must be (and thus depending on by how far this maximum permissible transverse
acceleration ayp, max it should be kept below), a setpoint value for the transverse deflection
dyw.soii of the car body 102 in the direction of the vehicle transverse axis (yr axis) can be
specified, which corresponds to the current vehicle state. Here, this setpoint value for the

transverse deflection of the car body 102 again comprises a quasi static component
and a dynamic component dywd.soii, wherein the following applies:

The quasi static component dyws.soii is the quasi static setpoint value for the transverse
deflection (and thus the rolling angle) that is relevant for tilting comfort and which is
determined by the current quasi static transverse acceleration aypS (which in turn is
dependent upon the curvature of the curve and the current running speed v). Therefore, here
it is the setpoint value for the transverse deflection, as is the case with vehicles known from
the state of the art with active setting of the rolling angle for regulation of the rolling angle.
The dynamic component dyWcuoiion the other hand is the dynamic setpoint value for the
transverse deflection (and thus as the case may be also for the rolling angle) relevant for the
vibration comfort, which is the result of the current dynamic transverse acceleration aypd
(which in turn is caused by periodic or singular disturbances on the track).
In order to actively set the transverse deflection dyw of the car body 102 with respect to the
neutral position (as shown in Figure 1 by the broken contour 102.2), the first rolling
compensation device 105 in the present example also has an actuator device 107, which for
its part comprises an actuator 107.1 and an associated control device 107.2. The actuator
107.1 is connected at one end in an articulated fashion with the bogie frame 104.2 and at the
other in an articulated fashion with the car body 102.
In the present example the actuator 107.1 is designed as an electro-hydraulic actuator. It is
self-evident, however, that with other variants of the invention an actuator can also be used
that works according to any other suitable principle. Thus for example hydraulic, pneumatic,
electrical and electromechanical operating principles can be used singly or in any
combination.
The actuator 107.1 in the present example is arranged in such a way that the actuator force
exerted by it between the bogie frame 104.2 and the car body 102 (in the neutral position)
acts parallel to the vehicle transverse direction (yr direction). It is self-evident, however, that
with other variants of the invention another arrangement of the actuator can be provided,

provided that the actuator force exerted by it between the running gear and the car body has
a component in the vehicle transverse direction.
The control device 107.2 controls or regulates the actuator force and/or the deflection of the
actuator 107.1 according to the present invention in such a way that a quasi static first
transverse deflection component dyWs of the car body 102 and a dynamic second transverse
deflection component dyWdOf the car body 102 are superimposed on one another so that
overall a transverse deflection dyw of the car body 102 results, for which the following
applies:
The setting of the transverse deflection dyw takes place according to the invention using the
setpoint value for the transverse deflection dywson of the car body 102, which is composed of
the quasi static component dy^^n and the dynamic component dyWd,SOii, as defined for
example in equation (2).
In order to increase the tilting comfort for the passengers the setting (supported by the
centrifugal force Fy) of the first transverse deflection component dyWs in the present example
takes place in a first frequency range F1 that ranges from 0 Hz to 1.0 Hz. The first frequency
range thus is the frequency range in which the quasi static rolling motions of the car body
corresponding to the current curvature of the curve travelled and the current running speed
take place.
In order to increase, in addition to the tilting comfort, the vibration comfort for the passengers,
the setting of the second transverse deflection component dyWd in the present example takes
place according to the invention in a second frequency range F2, ranging from 1.0 Hz to
6.0 Hz. The second frequency range is a frequency range which is adapted to the dynamic
disturbances (as the case may be periodic, typically however rather singular or statistically
scattered) expected during operation of the vehicle, which are noticed by passengers and
perceived as annoying.
It is self-evident, however, that the first frequency range and/or the second frequency range,
depending on the requirements of the rail network and/or the vehicle operator (for example

due to the use of the vehicle for local travel or long-distance travel, in particular high-speed
travel) can also vary.
By means of the solution according to the invention the first transverse deflection component
dyWsof the car body 102, the setting of which ultimately represents a quasi static adaptation
of the transverse deflection (and thus of the rolling angle) to the current curve bend and the
current running speed, is thus overlaid by a second transverse deflection component, dyWd of
the car body 102, the setting of which ultimately represents a dynamic adaptation to the
current disturbances introduced into the car body so that, overall, a higher comfort for the
passengers can be achieved.
The control device 107.2 controls the actuator 107.1 as a function of a series of input
variables, which are supplied to it by a higher level vehicle controller and separate sensors
(such as for example the sensor 107.3) or similar. The input variables considered for control
include, for example, variables which are representative of the current running speed v of the
vehicle 101, the curvature x of the current curved section being travelled, the track
superelevation angle y of the track section currently being travelled and the strength and the
frequency of disturbances (such as track geometry disturbances) of the track section
currently being travelled.
These variables that are processed by the control device 107.2 can be determined in any
suitable manner. In particular, in order to determine the setpoint value of the dynamic second
transverse deflection component dyWd,soii it is necessary to determine the disturbances or the
resultant transverse accelerations ay, the effects of which are to be at least attenuated via the
dynamic component dyWd> with sufficient accuracy and sufficient bandwidth (thus for example
to directly measure them and/or calculate them using suitable models of the vehicle 101
and/or the track generated in advance).
Here, the control device 107.2 can be realized in any suitable manner, provided that it meets
the safety requirements specified by the operator of the rail vehicle. Thus, for example, it can
be made as a single, processor-based system. In the present example, for the regulation in
the first frequency range F1 and the regulation in the second frequency range F2 different
control circuits or control loops are provided.
In the present example the actuator 107.1, in the first frequency range F1, has a maximum
deflection of 80 mm to 95 mm from the neutral position, while, in the second frequency

range, it has a maximum deflection of 15 mm to 25 mm from a starting position. In the first
frequency range F1 the actuator 107.1 also exerts a maximum actuator force of 15 kN to
30 kN, while, in the second frequency range, it exerts a maximum actuator force of 10 kN to
30 kN. In this way a particularly good configuration from the static and dynamic points of view
is achieved.
Through the design of the rolling compensation device 105 as an active system it is
furthermore possible in an advantageous manner to design the support of the car body 102
on the running gear 104 in the transverse direction of the vehicle 101 to be relatively stiff. In
particular it is possible to position the rolling axis and the instantaneous centre of rotation
MP, respectively, of the car body 102 comparatively close to the centre of gravity SP of the
car body 102.
In the present example, the secondary suspension 103.2 is designed so that it has a
restoring force-transverse deflection characteristic Iine108 as shown in Figure 5. Here, the
force characteristic line 108 is an indication of the dependency of the restoring force Fyf
exerted by the secondary suspension 103.2 on the car body 102, which acts during a
transverse deflection yf of the car body 102 in relation to the bogie frame 104.2. Similarly, for
the secondary suspension 103.2, a restoring characteristic line in the form of an moment
characteristic line can be indicated, which is an indication of the dependency between the
restoring moment Mxf exerted by the secondary suspension 103.2 on the car body 102 and
the rolling angle deflection aw from the neutral position.
As can be seen from Figure 5, the secondary suspension 103.2, in a first transverse
deflection range Q1, has a first transverse stiffness R1, while, in a second transverse
deflection range Q2 lying above the first deflection range Q1, it has a second transverse
stiffness R2 which is less than the first transverse stiffness R1.
Here, it is self-evident that the transverse stiffness (as can be seen from Figure 5 also from
the broken force characteristic lines 109.1, 109.2 of other embodiments) can vary (as the
case may be, considerably) within the respective transverse deflection range Q1 or Q2. The
respective transverse stiffness R1 or R2 is preferably selected so that the level of the first
transverse stiffness R1 at least partially, preferably substantially completely, lies above the
level of the second stiffness R2. Of course, a transitional area between the first transverse
deflection range Q1 and the second transverse deflection range Q2 can be provided in which
there will be an intersection or overlapping, respectively, of the stiffness levels. Basically the

behaviour of the stiffness according to the transverse deflection can be adapted to the
present application in any suitable manner.
In particular, in advantageous variants of the invention, in the second transverse deflection
range Q2 a second gradient at least in the vicinity of the value of zero, preferably equal to
zero, can be provided, as indicated in Figure 5 by the contour 109.3. Similarly, in other
variants of the invention, in the second transverse deflection range Q2, a negative second
gradient can be provided, as indicated in Figure 5 by the contour 109.4. In this way, the
actuator forces in the event of larger transverse deflections can be kept particularly low in an
advantageous manner.
In the present example the stiffness level in the first transverse deflection range Q1 is
selected so that the first transverse stiffness R1 is in the range 100 N/mm to 800 N/mm,
while the stiffness level in the second transverse deflection range Q2 is selected so that the
second transverse stiffness R2 is in the range 0 N/mm to 300 N/mm.
In the present example the force characteristic 108 in the first transverse deflection area Q1
accordingly has a first inclination S1 = dFyf/dyf(Q1) and in the transverse deflection area Q2 a
second inclination S2 = dFyf/dyt(Q2), which is less than the first inclination. The ratio
V = S2/S1 of the second inclination S2 to the first inclination S1 is in the range 0 to 3. It is
self-evident, however, that with other variants of the invention other values can also be
selected for the ratio V.
The two transverse deflection ranges Q1 and Q2 can likewise be selected in any way that is
adapted to the respective application. In the present example, the transverse deflection
range Q1 extends from 0 mm to 40 mm, while the second transverse deflection range Q2
extends from 40 mm to 100 mm. In this way, with regard to a limitation of the maximum
transverse deflection of the car body 102 with the lowest possible energy consumption for
the rolling compensation device 105, particularly favourable designs can be achieved.
As already mentioned, for the vehicle 101, similarly to the force characteristic 108, an
instantaneous characteristic can be defined. With this approach the restoring characteristic
line, in a first rolling angle range W1, has a first inclination S1 and, in a second rolling angle
range W2 lying above the first rolling angle range W1, a second inclination which is less than
the first inclination. With this approach also the ratio V = S2/S1 of the second inclination S2
to the first inclination S1 is in the range 0 to 3. The first rolling angle range W1 then,

depending on the specified kinematics, ranges, for example, from 0° to 1.3°, while the
second rolling angle range W2 ranges from 1.0° to 4.0°.
In other words, in the present example therefore a degressive behaviour of the transverse
stiffness of the secondary suspension 103.2 is provided, so that an initial transverse
deflection of the car body 102 from the neutral position is counteracted by a comparatively
high resistance.
The initial high resistance to a transverse deflection has the advantage that in the event of a
failure of the active components (for example the actuator 107.1 or the controller 107.2),
even when travelling a curve, (according to the currently existing transverse acceleration ay
or the centrifugal force Fy) an extensive passive restoration of the car body at least to the
vicinity of the neutral position is possible. This passive restoration, in the case of a fault,
allows in an advantageous manner particularly wide car bodies 102 and, consequently, a
high transport capacity of the vehicle 101 to be achieved. In order to prevent the actuator
107.1 impeding this passive restoration, the actuator 107.1 in the present example is
designed so that, in the event of its inactivity, it substantially presents no resistance to a
rolling motion of the car body 102. Consequently, the actuator 107.1 is not designed to be
self-restraining.
Thanks to the degressive characteristic line 108 the rise of the resistance to the transverse
deflection decreases as the deflection increases (with a negative inclination the resistance
itself can even fall). With regard to the dynamic setting of the second transverse deflection
dyWd in the second frequency range F2 during travel in curves of the vehicle101 this is an
advantage, since the rolling compensation device 105 must provide comparatively low forces
for these dynamic deflections in the second frequency range F2.
The degressive characteristic of the secondary suspension can be achieved in any suitable
manner. Thus, for example, as in the present example, the springs, via which the car body
102 is supported on the bogie frame 104.2, can be correspondingly designed so that this
characteristic is inherently achieved. In the case of air suspension this can for example take
place by a suitable design of the support of the bellows of the respective pneumatic springs.
It is self-evident, however, that the spring device 103 in other variants of the invention can
have one or more additional transverse springs, as indicated in Figure 1 by the broken
contour 110. The transverse spring 110 serves to adapt or optimise the transverse stiffness

of the secondary suspension 103.2 for the respective application. This simplifies the design
of the secondary suspension 103.2 considerably despite the simple optimisation of the
transverse stiffness.
The transverse spring 110 can, as shown in the present example, be connected at one end
with the running gear frame and at the other with the car body. Additionally or alternatively
such a transverse spring can also be connected at one end with the running gear frame or
with the car body, while at the other it is connected with the rolling compensation device 105
(for example with a rod 106.5, 106.6). Similarly, the transverse spring can also operate
exclusively within the rolling compensation device 105, for example between one of the rods
106.5,106.6 and the associated lever 106.1 and 106.2, respectively, or the torsion shaft
106.3.
The transverse spring 110 can be designed to increase the stiffness of the spring device in
the direction of the vehicle transverse axis. It can have any characteristic adapted for the
respective application. Preferably, the transverse spring 110 itself has a degressive stiffness
characteristic in order to achieve an overall degressive stiffness characteristic of the
secondary suspension 103.2.
The transverse spring 110 can be designed in any suitable manner and work according to
any suitable operating principles. Thus, tension springs, compression springs, torsion springs
or any combination of these can be used. Furthermore, a purely mechanical spring, an
electromechanical spring, a pneumatic spring, a hydraulic spring or any combination of these
may be involved.
The transverse stiffness of the secondary suspension 103.2, in the present example, is
dimensioned so that, in the event of inactivity of the actuator 107.1 (for example because of a
failure of the actuator 107.1 or the controller 107.2), on the car body 102, a restoring moment
Mxf is exerted about the rolling axis, which is dimensioned so that a rolling angle deflection
of the car body 102 from the neutral position for a nominal loading (e.g.
m = mmax) of the car body 102 and for a vehicle at a standstill (e.g. v = v0 = 0) on a maximum
permitted track superelevation is less than 2°. For the first maximum
transverse deflection of the car body 102 from the neutral position
toward the outside of the curve, in the present example, it is the case that it is limited to
60 mm. For the second maximum transverse deflection of the car body

102 from the neutral position toward the inside of the curve it is the case here that this is
limited to 20 mm.
In other words, the secondary suspension 103.2 is designed such that the vehicle 101, if for
any reason (for example due to damage to the vehicle or to the track) it comes to a standstill
at such an unfavourable spot, as before complies with the specified gauge profile.
Furthermore, the restoring moment M*, when the actuator 107.1 is inactive, must be
dimensioned so that a rolling angle deflection of the car body 102 from
the neutral position for a nominal loading of the car body 102 and for a
maximum permitted transverse acceleration acting in the direction of the transverse
axis of the vehicle of the vehicle is less than 2°. For the first maximum transverse deflection
of the car body 102 from the neutral position toward the outside of the
curve, in the present example, it is the case that this is limited to 60 mm. For the second
maximum transverse deflection of the car body 102 from the neutral
position toward the inside of the curve it is the case here that this is limited to 20 mm.
In other words, the spring device (in particular its stiffness in the vehicle transverse direction)
is preferably designed so that a vehicle, in emergency operation in the event of failure of the
actuator device, when travelling at normal running speed as before complies with the
specified gauge profile.
In any case it is thus ensured, with the present example, that even in the event of failure of
the active components of the rolling compensation device 105 emergency operation of the
vehicle 101 with as the case may be degraded comfort characteristics (in particular with
regard to tilting comfort and/or vibration comfort) is nevertheless possible while complying
with the specified gauge profile.
With regard to the high width of the car body 102 that can be achieved and, thus, in
connection with the high transport capacity a further advantageous aspect of the design
according to the invention exists in the present example in that, through the design and
arrangement of the rods 106.5, 106.6, the distance AH (that exists in the neutral position of
the car body 102) between the rolling axis of the car body 102 and the instantaneous centre
of rotation MP, respectively, and the centre of gravity SP of the car body 102 in the direction
of the vehicle height axis (zr direction) is selected to be comparatively small.

Thus the centre of gravity SP of the car body 102, in the present example, has a first height
H1 = 1970 mm above the rail, more accurately stated above the upper surface of the rail
SOK, while the rolling axis, in the neutral position (shown in Figure 1), in the direction of the
vehicle height axis has a second height H2 above the upper surface of the rail SOK, which in
the present example is in the range 3700 mm to 4500 mm. Accordingly, in the present
example the following relationship results

which gives the ratio of the difference between the second height H2 and the first height H1
to the first height H1, and which is in the range of approximately 0.8 to approximately 1.3.
This allows designs to be achieved which with regard to the abovementioned limitation of the
transverse deflections and, thus, the feasibility of wide car bodies with high transport capacity
are particularly favourable.
Thus, the comparatively low distance AH between the instantaneous centre of rotation MP
and the centre of gravity SP has the advantage that firstly, simply as a result of the
comparatively small transverse deflections of the car body 102, a comparatively high rolling
angle aw is achieved. In this way, during travel in curves, on the one hand, even at high
running speeds v or high curve bends, only comparatively low transverse deflections of the
car body 102 are necessary in order to achieve the quasi static component aWsof the rolling
angle awand the quasi static component dyws of the transverse deflection dyw, respectively.
Furthermore, as the case may be, even heavy transverse impacts can be compensated by
comparatively low transverse deflections of the car body 102, with which the dynamic
component aWdOf the rolling angle aw is created.
In other words, therefore, in normal operation of the vehicle 101 comparatively low
transverse deflections are required in order to achieve the desired travel comfort for the
passengers. Thanks to the low transverse deflections, in normal operation, a gauge profile
that is specified for the rail network on which the vehicle 101 is operated can be adhered to
in normal operation even with wide car bodies 102.
A further advantage of the low distance AH of the instantaneous centre of rotation MP from
the centre of gravity SP lies in the comparatively small lever arm resulting therefrom which
the centrifugal force Fy acting on the centre of gravity SP has to the instantaneous centre of

rotation MP. In the event of a malfunction of the active components of the rolling
compensation device 105 (for example in the event of a failure of the actuator 107.1 or the
controller 107.2), the centrifugal force Fy during travel in curves (according to the current
transverse acceleration ay) thus exerts a lower rolling moment on the car body 102, so that,
at least in the vicinity of the neutral position, an extensive passive restoration of the car body
102 by the secondary suspension 103.2 is possible.
In other words, therefore, even in the event of such a malfunction or an emergency operation
of the vehicle 101, comparatively low transverse deflections of the car body 102 occur.
Thanks to the low transverse deflections in emergency operation a gauge profile specified for
the rail network on which the vehicle 101 is operated can be adhered to even during such
emergency operation with wide car bodies 102.
It is self-evident that, with certain variants of the vehicle according to the invention with
particularly low transverse deflections, it can be provided (for example by a corresponding
design and arrangement of the rods 106.5,106.6) that the rolling axis or the instantaneous
centre of rotation of the car body is at or near the centre of gravity SP of the car body, so that
the centrifugal force Fy cannot make any (or at least no significant) contribution to the
generation of the rolling motion. The setting of the rolling angle awthen takes place
exclusively actively via the actuator 107.1.
Generally, therefore, it is to be noted that the contribution of the centrifugal force Fy to the
setting of the rolling angle aw is determined by the distance AH of the instantaneous centre
of rotation MP from the centre of gravity SP. The smaller this distance AH is the greater will
be the proportion of the actuator force of the actuator 107.1 that will be needed to set the
rolling angle aw (which corresponds to the current running situation and is necessary for the
desired travel comfort of the passengers).
In order to ensure adherence to a specified gauge profile in normal operation in any case, in
the present example, a limitation of the transverse deflections adapted to the gauge profile
specified by the operator of the vehicle is provided which comes into play in limit situations of
the operation of the vehicle 101. It is self-evident, however, that, with other variants of the
vehicle according to the invention, such a limitation can be used already in normal operation.
But, similarly, it can be provided that such a limitation is also absent so that in all possible
travel situations and load situations, respectively, of the vehicle no such limitation is active.

The limitation of the transverse deflections can be achieved by any suitable measures, such
as for example corresponding stops between the car body 102 and the bogie 104, in
particular the bogie frame 104.2. Similarly, a corresponding design of the rolling
compensation device 105 can be provided. Thus, for example, corresponding stops for the
rods 106.5, 106.6 can be provided.
In the present example, the actuator 107.1 is designed so that a first maximum transverse
deflection dyamax of the car body 102 from the neutral position occurring during travel in
curves toward the outside of the curve in the vehicles transverse direction (yf axis) is limited
to 120 mm. Since the bogie 104 is arranged on the vehicle 101 in the end area of the car
body 102, it is of particular interest to accordingly limit the transverse deflections toward the
inside of the curve. The actuator 107.1 therefore also limits a second maximum transverse
deflection dyiimax of the car body 102 from the neutral position toward the inside of the curve
occurring in the vehicle transverse direction during travel in curves to 20 mm.
This different limitation of the maximum transverse deflection toward the inside of the curve
(dyi,max) and toward the outside of the curve (dya,max) is achieved in the present example via
the control device 107.2. The control device 107.2 controls the actuator 107.1 for this
purpose (according to the direction of the curve currently being travelled) such that, when the
respective maximum transverse deflection (dyiimax and dya,max, respectively) is reached, a
further transverse deflection beyond the maximum value is prevented.
Furthermore, it can be provided that the control device 107.2 varies the maximum transverse
deflection toward the inside of the curve dyiimax(P) and/or toward the outside of the curve
dya,max(P) according to the current position P of the vehicle 101 on the rail network travelled.
Thus, for example, in certain track sections toward the inside of the curve and/or toward the
outside of the curve a lower maximum transverse deflection of the car body 102 can be
permitted than in other track sections. It is self-evident here that the control device 107.2
then must have available corresponding information on the current position P.
According to the invention it is further provided that the first rolling compensation device 105
and the second rolling compensation device 115, in order to reduce the side wind sensitivity
and to increase the permitted speed of the vehicle 101, respectively, are control-wise
coupled together, in that the control device 107.2 controls both the actuator 107.1 of the first
rolling compensation device 105 and the corresponding actuator 117.1 of the second rolling
compensation device 115 in such a way that, for example, under the effect of a side wind

load SW, a reduction of the torsional moment MTx acting on the car body 102 (as the case
may be as far as zero) is carried out.
In a design, in which the vehicle 101 for example forms the head of the train, in the event of
occurrence of side wind, a resultant side wind load SW in relation to the centre of gravity SP
of the vehicle, arranged (as a rule) approximately centrally in the vehicle longitudinal
direction, acts on the car body 102 in a manner displaced towards the head end and above
the centre of gravity SP of the vehicle (as shown in Figure 1).
In the event of inactivity of the actuators 107.1, in the event of an off-centre, displaced
towards the centre of gravity SP attack of the side wind load SW (with the forces or moments
shown in Figure 1) on the car body 102, as a result of the yawing moment, in the area of the
forward first bogie 104 (arranged at the head end), as a result of the design of the first rolling
compensation device 105, a first transverse deflection of the car body 102 would occur in
relation to the first bogie 104, as indicated in Figure 1 by the dash-double dotted contour
102.3. Conversely, on the trailing second bogie 114, as a result of the design of the second
rolling compensation device 115, a second transverse deflection of the car body 102 in
relation to the second bogie 114 running contrary to the first transverse deflection would
occur, as shown in Figure 1 by the broken contour 102.2.
From the force equilibriums and moment equilibriums the following values for the vertical
wheel contact forces Fzr, Fzl on either side of the running gear result here:

From equations (5) and (6) it is clear that, through the deviation dy between the first
transverse deflection (of the car body 102 in relation to the first bogie 104) and the opposing
second transverse deflection (of the car body 102 in relation to the second bogie 114), a
torsion of the car body and thus the torsion moment MTx results, which leads to a
considerable reduction on the amount of wheel contact force Fzr on the right side.

The controller 107.2 controls the actuator 107.1 of the first rolling compensation device 105
and the corresponding actuator 117.1 of the second rolling compensation device 115 such
that they reduce the deviation dy, in order to achieve in this way a reduction in the torsional
moment MTx acting on the car body 102 (as the case may be as far as to a value of zero).
This makes it possible to at least reduce a component of the wheel unloading resulting from
the torsion of the car body 102, and, as the case may be, to even eliminate it completely.
It is once again mentioned at this point that, depending on the design of the rolling
compensation device, as a rule a specified relation between the transverse deflection
concerned and the associated rolling angle exists, so that a consideration of the transverse
deflections and a consideration of the rolling angle may represent measures that are
equivalent or equal to one another.
The active reduction or elimination of the torsional load MTx is in the present example
achieved in that the control device 107.2 has at least one detection device to detect at least
one detection variable, which is representative of the torsional load MTx applied to the car
body 102. The control device 107.2 is in this case designed to control the actuator 107.1 of
the first rolling compensation device 105 and the corresponding second actuator 117.1 of the
second rolling compensation device 115 in such a way that the torsional load MTx is reduced
or, as the case may be, even substantially completely eliminated.
In the present example the control device 107.2 is designed in order to control the first
actuator 107.1 and the second actuator 117.1 in such a way that both the first transverse
deflection and also the second transverse deflection are reduced, so that, overall, a reduction
in the deviation dy results.
In the present example it is provided that the control device controls the first actuator 107.1
and the second actuator 117.1 as a function of the detection variable such that the deviation
dy between the first transverse deflection and the second transverse deflection is less than
10 mm.
For the detection variable basically any variable can be determined which allows to draw
conclusions on the current torsional load MTx on the car body 102 and, thus, ultimately the
wheel unloading resulting from this torsional load MTx. For example, it is possible to
determine directly at the car body 102 (for example by means of one or more strain gauge

strips or similar) a representative variable for the current torsional load on the car body and
to use this for the further control of the active components. In further preferred variants of the
vehicle according to the invention it is provided that the detection device of the control device
107.2, as the at least one detection variable, detects a variable representative of the first
transverse deflection of the car body 102 and a variable representative of the second
transverse deflection of the car body 102, which are then used for the further control of the
first actuator 107.1 and of the second actuator 117.1. This is an advantage to the extent that,
as the case may be, the variable to be set anyway is directly determined via the active
components. In the simplest example the detection device here can be realized by a
deflection sensor or similar integrated in the respective actuator 107.1, 117.1.
It is once again mentioned at this point that the use of an active component in the area of just
one of the two rolling compensation devices may be sufficient. Thus, for a reduction in the
torsional load it may be sufficient, for example, that through active intervention on the forward
running gear 104 the yaw moment on the vehicle body 102 resulting from the wind load SW
can be counteracted in that the deflection of the car body 102 is counteracted by a
corresponding force action in the area of the rolling compensation device 105 of the forward
running gear 104, while the deflection in the trailing running gear 114 is allowed.
Of course, it is likewise possible, in the area of the trailing running gear, to counteract by
means of active intervention the yaw moment on the car body resulting from the wind load in
an isolated manner, in that the deflection of the car body is counteracted by a corresponding
force of action in the area of the rolling compensation device of the trailing running gear,
while the deflection on the forward running gear is allowed.
It is further mentioned at this point that the control device 107.2 can be designed through
suitable measures such that the influences caused by side wind described above can be
distinguished from other vehicle dynamics influences (e.g. entry to and exit from track
superelevations, changes in the radius of curvature of the track, etc.). For this corresponding
filters as well as previously generated models of the vehicle can be used. Here, in particular,
account can be taken of the fact that influences caused by side winds have a quasi-static
nature and, consequently occur, in a comparatively low frequency range, which is, as a rule,
less than 2 Hz, so that, in particular, a differentiation from higher frequency dynamic
influences is as a rule possible without problems.

Furthermore it can consequently be provided that the control device 107.2 limits the
difference

between the rolling angle aw1 on the forward bogie 104 and the rolling angle cw on the
trailing bogie 114 or limits the difference

between the transverse deflection dywi on the forward bogie 104 and the transverse
deflection dyw2 on the trailing bogie 104. Here also, a similar active setting of the limitation
can be carried out, as the case may be, dependent upon the current section of track and/or
other variables (such as for example the rolling speed in the area of the respective bogie
104).
As has already been mentioned above, additionally or alternatively to the active solution
described above, a passive reduction of the torsional load on the car body 102 can be
provided, as shown schematically in Figure 8. This can be achieved by a corresponding
mechanical coupling of the two rolling compensation devices 105 and 115. According to
Figure 8, to this end it is provided that the first rolling compensation device 105 and the
second rolling compensation device 115 are coupled together mechanically by means of a
passive coupling device 120, wherein the coupling device 120, in order to reduce the
torsional load MTx on a car body 102, in the direction of a vehicle transverse axis generates
concurrent adjusting movements in the area of the first rolling compensation device 105 and
the second rolling compensation device 115.
The mechanical coupling between the two rolling compensation devices can be created in
any suitable fashion. Thus, for example, any mechanical gearing can be used to create this
coupling. In the present example, the coupling is carried out at least section-wise by means
of a fluidic operating principle, since in this way a particularly simple, space-saving design for
bridging the distance between the two running gears is possible.

To this end, the coupling device 120 comprises hydraulic cylinders 120.1,120.2 which are
coupled to the car body 102 and the respective rods 106.6 or 116.6 of the first rolling
compensation device 105 and the second rolling compensation device 115. The working
rooms of the hydraulic cylinders 120.1, 120.2 are contrarily coupled via a hydraulic line, in
order to achieve the desired concurrent setting motions.
As can be seen from Figure 1, the spring device 103 also has an emergency spring device
130.3, which is arranged centrally on the running gear 104.2 in the vehicle transverse
direction, in order that, even if the secondary suspension 103.2 fails, emergency operation of
the vehicle 101 is possible. The emergency spring device 103.3 can basically be designed in
any manner. In the present example the emergency spring device 103.3 is designed so that
it supports the compensation effect of the rolling compensation device 105. To this end, the
emergency spring device 103.3 can comprise a sliding and/or rolling guide which (in the
event of it being used, thus in emergency mode) can follow the compensation motion of the
rolling compensation device 105.
Basically it can be provided that the active setting of the rolling angle and of the transverse
deflection, respectively, via the rolling compensation device 105 takes place exclusively
during travel in curves on the curved track, and therefore the first rolling compensation
device 105 is active only in such a travel situation. In the present example, the rolling
compensation device 105 is also active during straight travel of the vehicle 101, so that in
any travel situation at least a setting of the transverse deflection dyw and, as the case may
be, the rolling angle ctw, respectively, takes place in the second frequency range F2 and,
thus, the vibration comfort in an advantageous manner is also guaranteed in these travel
situations.
Second embodiment
A further advantageous embodiment of the vehicle 201 according to the invention is shown in
Figure 6. The vehicle 201, in its basic design and functionality, corresponds to vehicle 101
from Figures 1 to 5, so that here merely the differences will be dealt with. In particular,
identical components are provided with identical reference numerals, while similar
components are provided with reference numerals incremented by a value of 100. Unless
otherwise stated in the following, regarding the features, functions and advantages of these
components reference is made to the above statements made in connection with the first
embodiment.

The difference from the example in Figures 1 to 5 lies in the design of the rolling
compensation device 205. Unlike in vehicle 101 the latter is arranged kinematically in series
with the spring device 103 via which the car body 102 is supported on the wheel units 104.1
of the respective bogie 104.
The rolling compensation device 205 comprises a guiding device 211, which is arranged
kinematically in series with the spring device 103. The guiding device 211 comprises two
guiding elements 211.1, which are supported at one end on a support 211.2 and at the other
on the car body 102, respectively. The support 211.2 extends in the vehicle transverse
direction and for its part is supported via the secondary suspension 103.2 on the bogie frame
104.2.
During rolling motions of the car body 102, the guiding elements 211.1 define the motion of
the support 211.2 in relation to the car body 102. The respective guiding element 211.1 is
designed as a simple multilayered spring device comprising a multilayered rubber layer
spring 211.3.
The rubber layer spring 211.3 is constructed from a plurality of layers, wherein for example
metal and rubber layers are interleaved. The rubber layer spring 211.3 is compressively rigid
in a direction perpendicular to its layers (so that the layer thickness under loading does not
change significantly in this direction) while, in a direction parallel to its layers, it is flexible (so
that under axial loading a significant deformation in this direction takes place). The layers of
the rubber layer spring 211.3, in the present example, are arranged at an inclination to the
vehicle height axis and to the vehicle transverse axis, so that they define the rolling axis and
the instantaneous centre of rotation MP, respectively, of the car body 102.
In the present example the layers of the rubber multilayered spring 211.3 are designed as
simple flat layers and such that the point of intersection of their mid-normals 211.4 defines
the rolling axis and the instantaneous centre of rotation MP, respectively, of the car body
102. It is self-evident, however, that, with other variants of the invention, another singly or
multiply curved design of these layers can be provided. In particular, it can be a case of
concentric cylinder sleeve segments whose centres of curvature lie in the instantaneous
centre of rotation MP.

In the present example, the mid-normals 211.4 lie in a common plane, which runs
perpendicular to the vehicle longitudinal axis (xr axis). Accordingly the arrangement of the
two rubber layer springs 211.3, in the vehicle transverse direction, can also transmit
comparatively high forces without additional aids, while in the direction of the vehicle
longitudinal axis only limited forces can be transmitted without considerable shear
deformation. Accordingly, as a rule between the car body 102 and the bogie frame 104.2 a
longitudinal articulation is provided, which allows a corresponding transmission of forces in
the direction of the vehicle longitudinal axis.
It is self-evident, however, that, with other variants of the invention, another design of the
rubber multilayered springs 211.3 can be provided, which allows the transmission of such
longitudinal forces. Thus, for example, doubly curved layers can be provided. Similarly,
however, more than two rubber layer springs can be provided which are not collinear and are
thus spatially arranged so that their mid-perpendiculars and their radii of curvature,
respectively, intersect in the instantaneous centre of rotation MP of the car body.
As can further be inferred from Figure 6, the rolling compensation device 205 again
comprises an actuator device 207 with an actuator 207.1 and a control device 207.2
connected thereto. In a similar manner to the actuator 107.1, the actuator 207.1 acts in the
vehicle transverse direction between the support 211.2 and the car body 102.
Under the control of the control device 207.2, via the actuator 207.1, the rolling angle aw and
the transverse deflection dyw, respectively, is set (as shown in Figure 6 by the broken contour
102.2). The control device 207.2, in the present example, operates similarly to the control
device 107.2. In particular, the control device 207.2 controls or regulates the actuator force
and/or the deflection of the actuator 207.1 according to the present invention in such a way
that a quasi static first transverse deflection dyWs of the car body 102 and a dynamic second
transverse deflection dyWd of the car body 102 are overlaid on one another so that, overall, a
transverse deflection dyw of the car body 102 results, for which the above equation (2)
applies. Here also, the quasi static first transverse deflection dyWs is again set in the first
frequency range F1, while the dynamic second transverse deflection dyWd is set in the
second frequency range F2.
In the event of inactivity of the active components (thus, for example, of the actuator 207.1 or
the controller 207.2) of the rolling compensation device 205, the passive restoration of the
car body takes place via the elastic resetting force of the rubber layer springs 211.3. The

rubber layer springs 211.3 can be designed in such a way that they have a similar
characteristic to the secondary suspension 103.2 from the first embodiment, so that in this
regard reference is made to the statements above.
As can further be inferred from Figure 6, between the bogie frame 104.2 and the support
211.2 (kinematically in parallel with the secondary suspension 103.2) a conventional rolling
support 206 with rods 206.5, 206.6 running parallel to one another is provided, which
counteracts an uneven dipping of the secondary suspension 103.2. Additionally, between the
bogie frame 104.2 and the support 211.2, in the vehicle transverse direction, a further
actuator 212 of the rolling compensation device 205 operates, via which the transverse
deflection of the support 211.2 and thus also of the car body 102 in relation to the bogie
frame 104.2 can be influenced. It is self-evident, however, that, in other variants of the
invention, on the one hand such an additional actuator can, as the case may be, be
dispensed with and, on the other hand, that also again an inclined arrangement of the rods
can be provided.
The actuator 212 is likewise controlled by the control device 207.2 so that the control device
207.2, by controlling the actuators 207.1 and 212, can bring about an operational behaviour
of the rolling compensation device 205 like that which has already been described above in
connection with the first embodiment for the rolling compensation device 105.
Here again it is pointed out that the design of the rolling compensation device with such a
layer spring device for definition of the rolling axis of the car body constitutes an individually
patentable inventive idea, which is, in particular, independent of the setting, as described
above, of the transverse deflection (and as the case may be the rolling angle, respectively) in
the first frequency range F1 and the second frequency range F2.
Third embodiment
A further advantageous embodiment of the vehicle according to the invention 301 is shown in
Figure 7. The vehicle 301, in its basic design and functionality, corresponds to vehicle 201
from Figure 6, so that here merely the differences will be dealt with. In particular, identical
components are provided with identical reference numerals, while similar components are
provided with reference numerals incremented by a value of 200. Unless otherwise stated in
the following, regarding the features, functions and advantages of these components
reference is made to the above statements in connection with the first embodiment.

The difference from the example of Figure 6 lies merely in the arrangement of the rolling
compensation device 305. Unlike vehicle 201 the latter is arranged kinematically in series
between the primary suspension 103.1 and the secondary suspension 103.2, via which the
car body 102 is supported on the wheel units 104.1 of the respective bogie 104.
The rolling compensation device 305 again comprises a guiding device 311 with two guiding
elements 311.1, which are supported, on the one hand, on a support 311.2 and, on the other
hand, on the bogie frame 104.2. The car body 102 is supported via the secondary
suspension 103.2 on the support 311.2, which extends in the vehicle transverse direction.
The guiding elements 311.1 are designed like the guiding elements 211.1 and, during rolling
motions of the car body 102, define the motion of the support 311.2 in relation to the bogie
frame 104.2. The respective guiding element 311.1 is again designed as a simple
multilayered spring device, which comprises a rubber layer spring 311.3, with a design
similar to the rubber layer spring 211.3.
As can further be inferred from Figure 7, the rolling compensation device 305 again
comprises an actuator device 307 with an actuator 307.1 and a control device 307.2
connected thereto, which operate in a manner analogous to the actuator 207.1 and the
control device 207.2.
As can be further inferred from Figure 7, between the car body 102 and the support 311.2
(kinematically in parallel with the secondary suspension 103.2) a conventional rolling support
306 with rods 306.5, 306.6 running parallel to one another is provided, which counteracts an
uneven dipping of the secondary suspension 103.2. Additionally, between the car body 102
and the support 311.2, in the vehicle transverse direction, a further actuator 312 of the rolling
compensation device 305 acts, via which the transverse deflection of the car body 102 in
relation to the support 311.2 and, thus, also in relation to the bogie frame 104.2 can be
influenced.
The actuator 312 is likewise controlled by the control device 307.2 so that the control device
307.2, by controlling the actuators 307.1 and 312, can bring about an operational behaviour
of the rolling compensation device 305 like that which has already been described above in
the context of the first and second embodiment.

The present invention has been described above exclusively using examples for rail vehicles.
It is further self-evident that the invention can also be used in connection with any other
vehicles.
*****

Claims
1. Vehicle, in particular a rail vehicle, having
- a carbody (102),
a first running gear (104), and
a second running gear (114) arranged at a distance from the first running gear
(104) in the direction of a vehicle longitudinal axis, wherein
the car body (102) is supported on the first running gear (104) in the direction of a
vehicle height axis by means of a first spring device (103),
the car body (102) is supported on the second running gear (114) in the direction
of the vehicle height axis by means of a second spring device (113),
the car body (102) is coupled to the first running gear (104) by means of a first
rolling compensation device (105; 205; 305),
the car body (102) is coupled to the second running gear (114) by means of a
second rolling compensation device (115; 215; 315),
the first rolling compensation device (105; 205; 305) and the second rolling
compensation device (115; 215; 315) counteract rolling motions of the car body
(102) toward the outside of the curve about a rolling axis parallel to the vehicle
longitudinal axis during travel in curves,
characterised in that
the first rolling compensation device (105; 205; 305) is designed in such a way
and/or the first rolling compensation device (105; 205; 305) and the second
rolling compensation device (115; 215; 315) are coupled to each other in such a
way that a torsional load on the car body (102) about the vehicle longitudinal
axis, which is caused, in particular, by a wind load acting on the car body (102),
is counteracted.
2. Vehicle according to claim 1, characterised in that
the first rolling compensation device (105; 205; 305) configured to impose upon
the car body (102), under a first transverse deflection of the car body (102) in
relation to the first running gear (104) in the direction of a vehicle transverse axis,
a first rolling angle about the rolling axis;
the second rolling compensation device (115; 215; 315) is configured to impose
upon the car body (102), under a second transverse deflection of the car body
(102) in relation to the second running gear (114) in the direction of a vehicle
transverse axis, a second rolling angle about the rolling axis;

the first rolling compensation device (105; 205; 305) is designed in such a way
and/or the first rolling compensation device (105; 205; 305) and the second
rolling compensation device (115; 215; 315) are coupled together in such a way
that a deviation between the first transverse deflection and the second transverse
deflection and/or a deviation between the first rolling angle and the second rolling
angle is counteracted.
3. Vehicle according to claim 1 or 2, characterised in that
the first rolling compensation device (105; 205; 305) has a first actuator device
(107; 207; 307) with at least one first actuator unit (107.1; 207.1; 307.1)
controlled by a control device (107.2; 207.2; 307.2), wherein the first actuator
device (107; 207; 307), in particular, is designed to contribute to the setting of the
first transverse deflection,
and/or
the second rolling compensation device (115; 215; 315) has a second actuator
device (117; 217; 317) with at least one second actuator unit (117.1; 217.1;
317.1) controlled by the control device (107.2; 207.2; 307.2), wherein the second
actuator device (117; 217; 317), in particular, is designed to contribute to the
setting of the second transverse deflection.
4. Vehicle according to claim 3, characterised in that
the control device (107.2; 207.2; 307.2) has at least one detection device to
detect at least one detection variable, which is representative of the torsional load
applied to the car body (102), and
the control device (107.2; 207.2; 307.2) is configured to control the first actuator
unit (107.1; 207.1; 307.1) and/or the second actuator unit (117.1; 217.1; 317.1) in
such a way that the torsional load is reduced, wherein
the control device (107.2; 207.2; 307.2), in particular, is configured to control the
first actuator unit (107.1; 207.1; 307.1) and/or the second actuator unit (117.1;
217.1; 317.1) in such a way that, in the direction of a vehicle transverse axis, a
deviation between a first transverse deflection of the car body (102) in relation to
the first running gear (104) and a second transverse deflection of the car body
(102) in relation to the second running gear (114) is reduced.
5. Vehicle according to claim 4, characterised in that

- the control device (107.2; 207.2; 307.2) controls the first actuator unit (107.1;
207.1; 307.1) and/or the second actuator unit (117.1; 217.1; 317.1) asafunction
of the detection variable in such a way that the deviation between the first
transverse deflection and the second transverse deflection is less than 40 mm,
preferably less than 25 mm, further preferably less than 10 mm,
and/or
the control device (107.2; 207.2; 307.2), as a function of the detection variable,
controls the first actuator unit (107.1; 207.1; 307.1) and/or the second actuator
unit (117.1; 217.1; 317.1) in such a way that a deviation between a first rolling
angle of the car body (102) in relation to the first running gear (104) and a second
rolling angle of the car body (102) in relation to the second running gear (114) is
less than 2°, preferably less than 1°, further preferably less than 0.5°.
6. Vehicle according to claim 4 or 5, characterised in that
the detection device, as the at least one detection variable, detects a variable
representative of the first transverse deflection of the car body (102) and/or a
variable representative of the second transverse deflection of the car body (102)
and/or
the detection device, as the at least one detection variable, detects a variable
representative of a deflection of a component of the first rolling compensation
device (105; 205; 305) and/or a variable representative of a deflection of a
component of the second rolling compensation device (115; 215; 315).
7. Vehicle according to one of the preceding claims, characterised in that
the first rolling compensation device (105; 205; 305) and the second rolling
compensation device (115; 215; 315) are coupled together mechanically by
means of a passive coupling device, wherein
the coupling device, in order to reduce the torsional load on the car body (102), in
the direction of a vehicle transverse axis generates concurrent adjusting
movements in the area of the first rolling compensation device (105; 205; 305)
and the second rolling compensation device (115; 215; 315), wherein
the coupling device, in particular, comprises a fluidic coupling between the first
rolling compensation device (105; 205; 305) and the second rolling compensation
device (115; 215; 315).
8. Vehicle according to one of the preceding claims, characterised in that

the first rolling compensation device (105; 205; 305), in order to increase the
tilting comfort, is designed to impose, in a first frequency range and under a first
transverse deflection component of the first transverse deflection of the car body
(102), upon the car body (102), in the direction of the vehicle transverse axis, a
first rolling angle component of the first rolling angle about the rolling axis, which
corresponds to a current curvature of a current section of track being travelled,
and/or
the first rolling compensation device (105; 205; 305), in order to increase the
vibration comfort, is designed to impose, in a second frequency range, upon the
car body (102) a second transverse deflection component overlaid to the first
transverse deflection component, wherein
the second frequency range at least partially, in particular, completely, lies above
the first frequency range.
9. Vehicle according to claim 8, characterised in that
the first rolling compensation device (105; 205; 305) has a first actuator device
(107; 207; 307) with at least one first actuator unit (107.1; 207.1; 307.1)
controlled by a control device (107.2; 207.2; 307.2), wherein
the first actuator device (107; 207; 307), in particular, is designed to make at
least a majority contribution to the generation of the first rolling angle in the first
frequency range, in particular, to substantially generate the first rolling angle.
10. Vehicle according to claim 8 or 9, characterised in that
the first frequency range ranges from 0 Hz to 2 Hz, preferably from 0.5 Hz to
1.0 Hz,
and/or
the second frequency range ranges from 0.5 Hz to 15 Hz, preferably from 1.0 Hz
to 6.0 Hz
and/or
the first rolling compensation device (105; 205; 305) is also active during straight
travel.
11. Vehicle according to one of claims 8 to 10, characterised in that
the car body (102) has a neutral position, which it adopts when the vehicle is
stationary on a straight, level track, and

the first rolling compensation device (105; 205; 305), in particular a first actuator
device (107; 207; 307) of the first rolling compensation device (105; 205; 305), is
configured in such a way that
a first maximum transverse deflection of the car body (102) from the neutral
position occurring toward the outside of the curve during travel in curves, in a
vehicle transverse direction, is limited to 80 mm to 150 mm, preferably
100 mm to 120 mm,
and/or
a second maximum transverse deflection of the car body (102) from the
neutral position occurring toward the inside of the curve during travel in
curves, in a vehicle transverse direction, is limited to 0 mm to 40 mm,
preferably to 20 mm.
12. Vehicle according to one of claims 8 to 11, characterised in that
a first actuator device (107; 207; 307) of the first rolling compensation device
(105; 205; 305) is configured to act as an end stop device for definition of at least
one end stop for the rolling motion of the car body (102), wherein
the first actuator device is designed to define the position of the at least one end
stop for the rolling motion of the car body (102) in a variable fashion.
13. Vehicle according to one of claims 8 to 12, characterised in that a first actuator device
(107; 207; 307) of the first rolling compensation device (105; 205; 305), in the event of
its inactivity, offers at most only slight resistance, in particular substantially no
resistance, to a rolling motion of the car body (102).
14. Vehicle according to one of claims 8 to 13, characterised in that
the car body (102) has a neutral position, which it adopts when the vehicle is
stationary on a straight, level track,
the first spring device (103), in the event of inactivity of an actuator device (107;
207; 307) of the rolling compensation device (105; 205; 305), exerts on the car
body (102) a restoring moment about the rolling axis, wherein
the restoring moment, in the event of an inactive actuator device (107; 207; 307),
is dimensioned such that
a transverse deflection of the car body (102) from the neutral position for a
stationary vehicle under a nominal loading of the car body (102) and with a

maximum permitted track superelevation is less than 10 mm to 40 mm,
preferably less than 20 mm,
and/or
a transverse deflection of the car body (102) from the neutral position, under
a nominal loading of the car body (102) and with a maximum permitted
transverse acceleration of the vehicle acting in the direction of a vehicle
transverse axis, is less than 40 mm to 80 mm, preferably less than 60 mm.
15. Vehicle according to claim 14, characterised in that
the first spring device (103) defines a restoring characteristic line, wherein
the restoring characteristic line represents the dependence of the restoring
moment on the rolling angle deflection and
the restoring characteristic line has a degressive behaviour, wherein
the restoring characteristic line, in particular, in a first rolling angle range, has a
first inclination and, in a second rolling angle range above the first rolling angle
range, has a second inclination that is less than the first inclination, wherein
the ratio of the second inclination to the first inclination, in particular, lies in
the range from 0 to 1, preferably in the range from 0 to 0.5, further preferably
in the range from 0 to 0.1,
and/or
the first transverse deflection range, in particular, ranges from 0 mm to
60 mm, preferably 0 mm to 40 mm, and the second transverse deflection
range, in particular, ranges from 20 mm to 120 mm, preferably from 40 mm
to 100 mm.
16. Vehicle according to claim 15, characterised in that
the car body (102) has a neutral position, which it adopts when the vehicle is
stationary on a straight, level track, and
the first spring device (103), in the direction of a vehicle transverse axis, has a
transverse stiffness, which is a function of a transverse deflection of the car body
(102) in the direction of the vehicle transverse axis from the neutral position,
wherein
the first spring device (103), in particular, in a first transverse deflection range,
has a first transverse stiffness and, in a second transverse deflection range lying
above the first transverse deflection range, has a second transverse stiffness,
which is lower than the first transverse stiffness, wherein

the first transverse stiffness, in particular, lies in the range from 100 N//mm to
800 N/mm, preferably in the range from 300 N/mm to 500 N/mm, and the
second transverse stiffness, in particular, lies in the range from 0 N/mm to
300 N/mm, preferably in the range from 0 N/mm to 100 N/mm,
and/or
the first transverse deflection range, in particular, ranges from 0 mm to
60 mm, preferably from 0 mm to 40 mm, and the second transverse
deflection range, in particular, ranges from 20 mm to 120 mm, preferably
from 40 mm to 100 mm.
17. Vehicle according to one of claims 8 to 16, characterised in that
the car body (102) has a nominal loading and a neutral position, which it adopts
when the vehicle is stationary on a straight, level track, and
the first spring device (103), in the direction of a vehicle transverse axis, has a
transverse stiffness, wherein
the transverse stiffness of the spring device (103) is dimensioned such that, in
the event of inactivity of a first actuator device (107; 207; 307) of the first rolling
compensation device (107; 207; 307), during travel in curves with a maximum
permissible transverse acceleration of the vehicle acting in the direction of a
vehicle transverse axis,
a first maximum transverse deflection of the car body (102) from the neutral
position toward the outside of the curve in a vehicle transverse direction is
limited to 40 mm to 120 mm, preferably to 60 mm to 80 mm,
and/or
a second maximum transverse deflection of the car body (102) from the
neutral position toward the inside of the curve in a vehicle transverse
direction is limited to 0 mm to 60 mm, preferably to 20 mm to 40 mm.
18. Vehicle according to one of claims 8 to 17, characterised in that
the car body (102) has a neutral position, which it adopts when the vehicle is
stationary on a straight, level track, and
the first rolling compensation device (105; 203; 305) is designed in such a way
that an actuator device (107; 207; 307) of the first rolling compensation device
(105; 205; 305),
- in the first frequency range, has a maximum deflection from the neutral
position of 60 mm to 110 mm, preferably 70 mm to 85 mm,

and/or,
- in the second frequency range, from a starting position, has a maximum
deflection of 10 mm to 30 mm, preferably 15 mm to 25 mm,
and/or,
- in the first frequency range, exerts a maximum actuator force of 10 kN to
40 kN, preferably 15 kN to 30 kN,
and/or,
- in the second frequency range, exerts a maximum actuator force of 5 kN to
35 kN, preferably 5 kN to 20 kN.
19. Vehicle according to one of claims 8 to 18, characterised in that
the car body (102) has a neutral position, which it adopts when the vehicle is
stationary on a straight, level track,
the car body (102) has a centre of gravity which, in the neutral position, in the
direction of the vehicle height axis has a first height above the track,
the first rolling compensation device (105; 205; 305) is configured in such a way
that the rolling axis, in the neutral position, in the direction of the vehicle height
axis has a second height above the track, wherein
the ratio of the difference between the second height and the first height to the
first height is a maximum of 2.2, preferably a maximum of 1.3, further preferably
0.8-1.3.
20. Vehicle according to one of claims 8 to 19, characterised in that
the first rolling compensation device (105) comprises a first rolling support device
(106), which is arranged kinematically in parallel to the first spring device (103)
and is designed to counteract rolling motions of the car body (102) about the
rolling axis during straight travel, wherein
the first rolling support device (106), in particular, comprises two rods (106.5,
106.6), each of which, at one end, is connected in an articulated manner to the
car body (102) and each of which, at the other end, is connected in an articulated
manner to opposing ends of a torsion element (106.3), which is supported by the
first running gear (104),
and/or
the first rolling compensation device (205; 305) comprises a guiding device (211;
311),

the guiding device (211; 311) is arranged kinematically in series with the first
spring device (103),
the guiding device (211; 311) comprises a guiding element (211.1; 311.1), which
is arranged between the first running gear (104) and the car body (102), and
the guiding device (211; 311) is configured so that, during rolling motions of the
car body (102), it defines a motion of the guiding element (211.1; 311.1) in
relation to the car body (102) or the first running gear (104), wherein
the guiding device (211; 311), in particular, comprises at least one layer spring
device (211.3; 311.3).
21. Vehicle according to one of claims 8 to 20, characterised in that
the first running gear (104) has a running gear frame (104.2) and at least one
wheel unit (104.1) and
the first spring device (103) has a primary suspension (103.1) and a secondary
suspension (103.2), wherein
the running gear frame (104.2) is supported via the primary suspension (103.1)
on the wheel unit (104.1), and the car body (102) is supported on the running
gear frame (104.2) via the secondary suspension (103.2), which is in particular
designed as pneumatic suspension, and
the first rolling compensation device (105) is arranged kinematically in parallel to
the secondary suspension (103.2) between the running gear frame (104.2) and
the car body (102).
22. Vehicle according to claim 21, characterised in that
the first spring device (103) comprises a transverse spring device (110), wherein
the transverse spring device (110)
is connected at one end to the running gear frame (104.2) and at the other to
the car body (102),
and/or
is connected at one end to the running gear frame (104.2) or to the car body
(102) and at the other to the first rolling compensation device (105)
and
the transverse spring device (110), in particular, is configured to increase the
stiffness of the first spring device (103) in the direction of a vehicle
transverse axis, wherein the transverse spring device (110), in particular, has
a degressive stiffness characteristic.

23. Vehicle according to one of claims 8 to 22, characterised in that
the first spring device (103) has an emergency spring device (103.3), which, in
the vehicle longitudinal direction, is arranged centrally on the first running gear
(104), wherein
the emergency spring device (103.3), in particular, is configured so that it
supports the compensation effect of the first rolling compensation device (105).
24. Method for setting rolling angles on a car body (102) of a vehicle, in particular a rail
vehicle, about a rolling axis parallel to a vehicle longitudinal axis of the vehicle, in
which
a first rolling angle and/or a first transverse deflection of the car body (102) is set
in relation to a first running gear (104), and
a second rolling angle and/or a second transverse deflection of the car body
(102) is set in relation to a second running gear (114), which, in the direction of a
vehicle longitudinal axis, is arranged at a distance from the first running gear
(104), wherein
the car body (102) is coupled to the first running gear (104) via a first rolling
compensation device (105; 205; 305),
the car body (102) is coupled to the second running gear (114) via a second
rolling compensation device (115; 215; 315),
the first rolling compensation device (105; 205; 305) and the second rolling
compensation device (115; 215; 315), during travel in curves, counteract rolling
motions of the car body (102) toward the outside of the curve about a rolling axis
parallel to the vehicle longitudinal axis,
characterised in that
the first rolling angle and/or the second rolling angle are set in a manner coupled
together in such a way that a torsional load on the car body (102) about the
vehicle longitudinal axis is counteracted
and/or
the first transverse deflection and/or the second transverse deflection are set in a
manner coupled together in such a way that a torsional load on the car body
(102) about the vehicle longitudinal axis is counteracted, wherein
the torsional load, in particular, is caused by wind loads acting on the car body
(102).

25. Method according to claim 24, characterised in that
a deviation between the first transverse deflection and the second transverse
deflection and/or a deviation between the first rolling angle and the second rolling
angle, is counteracted, wherein
the first transverse deflection and/or the second transverse deflection, in
particular, at least in part is set actively by means of an actuator unit controlled by
a control unit (107.2; 207.2; 307.2).
26. Method according to claim 25, characterised in that
at least one detection variable is detected which is representative of the torsional
load applied to the car body (102), and
the active setting of the first transverse deflection and/or the second transverse
deflection by the control device (107.2; 207.2; 307.2) takes place as a function of
the detection variable, wherein
as the at least one detection variable, in particular, a variable representative of
the first transverse deflection and/or a variable representative of the second
transverse deflection is detected
and/or
as the at least one detection variable a variable representative of a deflection of a
component of the first rolling compensation device (105; 205; 305) and/or a
variable representative of a deflection of a component of the second rolling
compensation device (115; 215; 315) is detected.
27. Method according to claim 25 or 26, characterised in that
the deviation between the first transverse deflection and the second transverse
deflection, in particular, is set in such a way that it is less than 40 mm, preferably
less than 25 mm, further preferably less than 10 mm,
and/or
the deviation between the first rolling angle and the second rolling angle is set so
that it is less than 2°, preferably less than 1°, further preferably less than 0.5°.
28. Method according to one of claims 24 to 27, characterised in that
the first rolling compensation device (105; 205; 305) and the second rolling
compensation device (115; 215; 315) are coupled together mechanically by
means of a passive coupling device, wherein

via the coupling device, in order to reduce the torsional load on the car body
(102), in the direction of a vehicle transverse axis, concurrent adjusting
movements in the area of the first rolling compensation device (105; 205; 305)
and the second rolling compensation device (115; 215; 315) are generated,
wherein
the coupling device, in particular, comprises a fluidic coupling between the first
rolling compensation device (105; 205; 305) and the second rolling compensation
device (115; 215; 315).
29. Method according to one of claims 24 to 27, characterised in that
the first rolling angle is actively set, wherein,
during travel in curves, rolling motions of the car body (102) toward the outside of
the curve about the rolling axis are counteracted and,
in order to increase the tilting comfort, the car body (102), in a first frequency
range and under a first transverse deflection component of the first transverse
deflection, has a first rolling angle component of the first rolling angle imposed
upon it, which corresponds to a current curvature of a current section of track
being travelled,
and/or
the car body (102), in order to increase the vibration comfort, in a second
frequency range, has a second transverse deflection component of the first
transverse deflection overlaid to the first transverse deflection imposed upon it,
wherein
the second frequency range at least partially, in particular completely, lies above
the first frequency range.
30. Method according to claim 29, characterised in that the first rolling angle, in the first
frequency range, at least predominantly, in particular substantially completely, is
generated actively.
31. Method according to claim 17 or 18, characterised in that
the first frequency range ranges from 0 Hz to 2 Hz, preferably from 0.5 Hz to
1.0 Hz,
and/or
the second frequency range ranges from 0.5 Hz to 15 Hz, preferably from 1.0 Hz
to 6.0 Hz.

32. Method according to claim 17 or 18, characterised in that the setting of the second
transverse deflection component, in the second frequency range, for increasing the
vibration comfort also takes place during straight travel.
* * * * *

The present invention relates to a vehicle, in particular a rail vehicle,
comprising a car body (102), a first chassis (104), and a second chassis
(114) arranged at a distance to the first chassis (104) in the direction of a
vehicle longitudinal axis, wherein the car body (102) is supported on the
first chassis (104) in the direction of a vehicle vertical axis by means of a
first spring device (103), the car body (102) is supported on the second
chassis (114) in the direction of the vehicle vertical axis by means of a
second spring device (113), the car body (102) is coupled to the first
chassis (104) by means of a first roll compensation device (105), the car
body (102) is coupled to the second chassis (114) by means of a second
roll compensation device (115), the first roll compensation device (105)
and the second roll compensation device (115) counteract roll motions of
the car body (102) toward the outside of the curve about a roll axis
parallel to the vehicle longitudinal axis during curved travel, wherein the
first roll compensation device (105) is designed in such a way and/or the
first roll compensation device (105) and the second roll compensation
device (115) are coupled to each other in such a way that a torsional load
on the car body (102) about the vehicle longitudinal axis, which is
caused in particular by a wind load acting on the car body (102), is
counteracted