This document summarizes a technical paper about designing and tuning air suspensions for tractors. It discusses the challenges of achieving a balance between ride comfort and handling for tractor-trailer combinations, given their mass distribution and the influence of the trailer on dynamics. The paper presents the development of a new primary air suspension system for a Volkswagen tractor to improve comfort, protect fragile loads, and allow for different trailer configurations while maintaining other vehicle characteristics. A mixed suspension using springs in front and air bags in the rear was selected based on its benefits and ability to meet design goals. The tuning process and methods used to evaluate and optimize the suspension are also described.

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SAE TECHNICAL
PAPER SERIES 2002-01-3041
Tractor Air Suspension Design and Tuning
Leandro Pugliese de Siqueira, Felipe Nogueira, Cesar Coutinho Ramos,
João Guilherme Herrmann, Silas Sartori, Charles Villiger,
Gabriel Regis de Paula and Fernando Fuhrken
Volkswagen Truck & Bus Operation
Valter Martins de Vargas e João Felipe Araujo
Suspensys Automotive Systems
Reprinted From: Truck Vehicle Dynamics & Suspensions
(SP-1728)
International Truck and Bus
Meeting and Exhibition
Detroit, Michigan
November 18-20, 2002

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3. ABSTRACT
This paper aims to present the difficulties on
designing and tuning a tractor suspension, specially due to
its particular mass distribution and the trailer influence on
its dynamic behavior. It also presents the development of a
primary suspension for the new VW 18.310 Tractor.
The drivers for the development of this new primary
suspension were: improve comfort level; achieve a better
condition on fragile load transportation; and allow different
fifth wheel positions and heights (to take the most
advantages of legal limitations on GAWRs - Gross Axle
Weight Rating- composition height and length). This
should be accomplished without affecting other vehicle
dynamic characteristics such as handling.
Due to its benefits to the pavement, the additional
facility on tractor-trailer coupling operation and some
package limitation, a mixed suspension (parabolic springs
on the front axle and pneumatic foils on the rear) is chosen
on the cost vs. functional attributes trade off process.
The experimental approach to the ride & handling
tuning process is discussed, the subjective methods and
objective metrics utilized to guide the development process
are presented. Finally, the results obtained are compared to
multi-body computer simulation, considering cost, time and
precision.
INTRODUCTION
The constant search for new products which
exceed customer expectation, best fitting on their
necessities and capable of delighting then, leads
Volkswagen Truck and Bus Operation to develop a new
tractor truck based on its current VW 40-300.
As result of a detailed market research, including
customer insight and a study on future transportation
tendencies, a list of product requirements and attributes to
meet VW goals was generated. Besides the standard
equipment on the new VW 18-310 Tractor, including the
repowered Euro II Cummins C engine (303 hp), the ZF 16
speeds synchronized gear box, 24 volts electric system and
air conditioning, the following goals are pursued:
• Improvement of comfort level for both passenger and
driver;
• Minimization of losses on the transportation of fragile
loads by reducing the acceleration level transmitted to
the cargo;
• Increment of the front axle gross weight rating up to 6
ton without affecting Ride & Handling;
• High mileage comfort robustness;
• Use of legal limitations on trailer heights (4,40 meters)
and composition length (18,15meters).
THEORETICAL BACKGROUND
Truck ride is admittedly a complex subject, but
what is ride? Gillespie defines ride as “a subjective
perception normally associated with the level of comfort
experienced when traveling in a vehicle” [1].
A good ride to one driver may not be suitable to
another. While one driver might want a soft ride, with slow
motions, the other may prefer the ride of firmer suspensions
in order to have increased the handling capability.
2002-01-3041
Tractor Air Suspension Design and Tuning
Leandro Pugliese de Siqueira, Felipe Nogueira, Cesar Coutinho Ramos, João Guilherme
Herrmann, Silas Sartori, Charles Villiger, Gabriel Regis de Paula and Fernando Fuhrken
Volkswagen Truck & Bus Operation
Valter Martins de Vargas e João Felipe Araujo
Suspensys Automotive Systems
Copyright © 2002 SAE International

4. Handling, on the other hand, refers to vehicle
qualities that feedback to the driver when changing direction
and sustaining lateral acceleration in the process. These
qualities affect the smoothness of the driving task or the
driver’s ability to maintain control.
The compromise between ride and handling on a
trailer-tractor combination is a particular hard task due to its
load conditions. Objectionable ride conditions on a tractor
vehicle are often described by drivers as a head snapping
action or as back slapping from the seat.
These driver reactions are primarily the result of
chassis pitching. Although pitching and bouncing motions
generally occur together, the pitch movement is predominant
on a tractor-trailer vehicle combination.
These effects can be clearly understood by
analyzing the Dynamic Index of tractors. This index
represents the relationship among the mass properties of
sprung weights, the dimensional values of wheelbase and
center of gravity location. It can be defined as:
(1)
Where:
i= √(I/m) = radius of gyration;
A, B= distance from the center of gravity to the front/rear
wheel respectively;
I= Momentum of inertia;
m= Total mass
Figure 1: Dynamic Index
The radius of gyration provides a useful way of
picturing the weight distribution of all components making
up the sprung weight of a vehicle. The sprung weight can be
considered to be equally divided and concentrated at two
points, one ahead and one behind the center of gravity at
distances equal to the radius of gyration.
As the concentrated weights are moved away from
the C.G., the inertia effect increases. This increment occurs
with the square of the distance and it is proportional to the
moment to rotate the weights around the center of gravity.
Thus, the dynamic index can be interpreted as the
ratio between the rotational inertia and the applied moment
developed by the suspension set. Consequently, high pitch
frequencies are associated with small dynamic indexes.
Because bounce involves basically inertia of translation, its
frequency changes little with dynamic index variation.
Passenger cars are designed with the center of
gravity located near the midpoint of the wheelbase, while
wheelbases are dimensioned in relation to the weight
distribution to approach a dynamic index of 1. This feature
combined with low rate and friction suspension, accounts for
the significant improvements made in passenger car ride
compared to earlier models.
The same principles contribute to the comparatively
good ride of buses. With the exception to the fact that on
buses, the difference between loaded and unloaded
conditions are considerable and must be taken into
consideration on the ride study.
AB
i
Di
2
=
By comparison, legal requirements stipulate truck
and trailer axle loads. Besides, the relative inflexibility in
locating heavy components generally results in low values of
dynamic index. That makes it a distant wish to achieve ride
comfort in tractor-trailer equipment comparable to that of
buses or passenger cars.
Even more, on this type of vehicle combination, the
trailer inputs are another source of disturbance to the tractor
ride. Because trailers usually have low values of dynamic
index (if considered as a vehicle unit), pitching is
predominant with higher frequencies and larger amplitudes
of vibration motion than in bounce. In trailers of low
dynamic index, the pitch axis is located behind the kingpin
and at an appreciable distance above the fifth wheel pivot.
When pitching occurs, the rotational oscillation
around the pitch axis produces a vibration at the fifth wheel
pivot that has both horizontal and vertical components of
motion with the same frequency. Since the tractor and trailer
are connected through the fifth wheel pivot, any motion the
trailer develops at the pivot must also occur in the tractor at
this point.
Motions created in the tractor also affect the trailer.
But, since the loaded trailer has generally five to six times
the weight of the tractor, the larger trailer weight has more
effect on the tractor than the smaller tractor weight on the
trailer.
A B
C.G.

5. Although neither changes on mass properties of
tractor nor wheelbase modifications on the VW 40-300 are
in the context of this project, the improvement on ride would
be achieved by:
• Increasing the amount of deflection either at front and
rear suspension, reducing consequently the frequencies
of both pitch and bounce movements;
• Reducing the amount of friction resisting the
suspension deflection. The primary suspension of the
VW 40-300 is composed of tapered leaves at front and
a rear suspension with tapered leaves on the first stage
and parabolic leaves on the second.
Also, the suspension project should be robust
enough to ensure a high mileage comfort, while maintaining
low friction on the suspension even with the severe
conditions of usage.
Improving the comfort level is also the key to allow
increments on the front axle load rating. Beginning with the
fifth wheel positioned at the rear tractor suspension center,
any fifth wheel setting forward of this position reduces the
radius of gyration of the sprung weight. The center of gravity
of the sprung weight moves forward with increased fifth
wheel settings, resulting in increased AB factor (Equation 1)
for a given wheel base.
Both factors combine to promote an appreciable
reduction of dynamic index that increases the pitch
frequency. By improving comfort level, it would be possible
to reposition the fifth wheel on a forward location in order to
achieve a 6-ton load on the front axle without affecting ride.
From the original list of product’s attributes
generated by Marketing, the one missing on the
considerations above is the necessity to take most
advantages of maximum legal limitation on trailer heights
(4,40 meters).
This goal would be achieved by minimizing the
suspension static deflection between loaded and unloaded
conditions, and reducing the current fifth wheel height. To
reduce the fifth wheel height it is necessary to lower the fifth
wheel sub-frame.
This modification brings to the discussion another
important variable on the vehicle dynamic study: the chassis
stiffness and its vibration modes. If frame bending and truck
pitch frequencies are close then ride will be affected. The
modal desalignment must be taken into consideration during
the design process.
Figure 2 presents the first chassis vertical mode of
the VW 40-300 and the new VW 18-310. Observe that the
bending characteristics, i.e. frequency and modal shape, are
very similar regardless the fifth wheel sub-frame differences.
VW 40-300 bending mode – 6,9 Hz
VW 18-310 bending mode – 6,4 Hz
Figure 2: Chassis Vertical Mode.
Once the project targets and the direction have been
established, a VW suspension system supplier (Suspensys
Automotive System) was invited to come aboard and co-
design the new suspension system, which is detailed in the
next section.
Meanwhile, a Volkswagen supplier of CAE services
(T-Systems) joined the team to perform a multi-body study
with two major objectives:
1. Comparisons with the experimental approach to design
and tun the vehicle suspension;
2. Provide indication of future modifications that could
further improve ride & handling characteristics of the
tractor.
DESIGN PHASE
As stated in the introduction, the ride of a 4x2
tractor is to be refined due to comfort and fragile load
handling implications. Moreover, since the experimental
approach is followed, no simplifications or linearizations are
necessary, and the physical phenomena can be analyzed
directly through the response of a prototype.

6. In other words, any nonlinear deformations, axle
kinematics or elasticity implications, tire/wheel
characteristics, chassis elasticity and other variables are
measured. So their influence in ride is taken into account in
the tuning process.
MS
Sprung
Mass
Z
SuspensionKS CSNonetheless, some hypotheses need to be presented
in order to narrow the focus of the problem. With that intent,
this paper presents the ride improvement in a tractor-trailer
arrangement, portraying a 4x2 tractor connected to a three-
axle trailer driving over a paved surface.
MU ZU Unsprung
Mass
The cabin suspension and seat characteristics are
not an object of study in this paper. They would improve
passenger comfort but not improve the fragile load handling.
Variations of road surface, another factor that could affect
ride, are also not considered. It is impossible to control the
quality of the pavement over which the tractor will roll.
TireKT
ZR
Road
The trailer is also not a variable since nowadays it is
difficult, although desirable, to have a fixed tractor-trailer
configuration. Now that the hypotheses that govern the study
have been laid out, the next sections attempt to define ride in
an engineering point of view.
PRIMARY SUSPENSION SYSTEMS
Ride motions have both frequency and amplitude
governed by four basic characteristics related in some
manner to the primary suspension:
1. Amount of larger total deflection either at front or rear
suspensions;
2. Ratio of larger to smaller suspension travel;
3. Amount of friction resisting suspension travel; and
4. Weight distribution of the vehicle related to wheelbase
and CG location.
There are many factors that affect ride, as shown in
the introduction, but no significant improvement or
development can be achieved without an adequate
suspension. From these four characteristics, one must
comment that the suspension must have enough deflection to
reduce frequencies of body movements (bounce and pitch),
and the hysteretic energy loss must be minimal so that the
suspension can be effective in small deflections.
Any study or improvement of a vehicle's dynamic
response must start with the knowledge of the basic
suspension system. Figure 3 shows a typical quarter-car
model. That figure shows a sprung mass supported on a
primary suspension that connects the unsprung and the
sprung masses. Then the entire system is put into contact
with the road surface by the tire (simplified by a spring).
Figure 3: Typical quarter-car model.
The major goals when designing such a suspension
system are basically to isolate sprung mass vibration and to
maintain the contact between tire and road surface. The
objectives behind these two main goals are improvement of
both passenger comfort and overall security, by minimizing
the normal load variation on tires.
These two objectives are a function of the elastic
and dissipative properties of the suspension in the vertical
direction. So the next subsections describe the vertical
elastic properties (ride rate) of the suspensions and the
dissipative (hysteretic behavior) of springs.
Ride Rate
The ratio of a loaded spring to its spring deflection
is termed “rate”. The so-called ride rate (RR) is the overall
rate of the suspension spring connected in series with the tire
to form the primary suspension.
A stiffer spring will transmit more vibration to the
chassis, but improve handling. On the other hand, a softer
spring will most likely improve comfort levels but it will
also require more suspension travel. So the designer is
always faced with compromises when designing a
suspension system.
It is worth also mentioning that, since the
suspension spring is connected in series with a relatively
higher rate tire spring, the suspension spring governs the ride
rate of the entire system. The ride rate is given roughly by:
KK
KK
TS
TS
RR
+
×
=
(2)

7. Where KS is the suspension vertical rate and KT is the tire
vertical rate.
It is always desirable to keep the natural frequency
as close to 1Hz as possible due to the fact that, as the
frequency increases, so do the acceleration peaks associated
with road irregularities. Gillespie points out that at 1Hz,
these acceleration peaks are at a minimal level [1, 3].
Since the spring rate governs the ride rate, it is
common to use a suspension rate as low as possible in order
to maintain a smaller ride rate and a smaller natural
frequency. The natural frequency of the quarter-car may be
expressed in terms of the ride rate as (in rad/sec):
M
RR
n
=ϖ
(3)
But along with lower natural frequencies, what
implies greater vibration isolation, comes a need for more
suspension travel as stated earlier. The next subsection
presents a definition that ties ride isolation characteristics
(ϖn) and suspension travel.
Suspension Travel
There are several ways to analyze the effect of
suspension travel on ride. For instance, the static deflection
of the vehicle is one of the earlier ways to evaluate vibration
isolation characteristics and design envelope. It represents
the ratio of sprung mass to suspension rate:
(4)
The initial study of a primary suspension must be a
compromise between suspension travel and natural
frequency. The smaller the natural frequency, close to 1Hz,
the better for vibration isolation in the range that causes
discomfort to human beings, as explained earlier. But that
would require more working space for the suspension, and
therefore drive up costs considerably besides compromising
handling and cornering.
Figure 4 is a graph that relates static deflection to
natural frequency. The importance of this graph is that it
shows that the trade-off between spring rates and design
envelope discussed above.
Figure 4: Static deflection versus natural frequency [1].
Moreover, there is no improvement in ride
characteristics if one attempts to compensate a smaller
deflection in one axle with a greater one on the other. That
is, it has been shown experimentally that increasing front
suspension deflection cannot substantially reduce the pitch
frequency associated with a smaller rear deflection in a
tractor [5].
In general, larger front deflections tend to create a
diving motion in deceleration and rearing in acceleration.
Also, a small deflection of a tractor or truck rear suspension
will likely increase both pitch frequency and force
transmission, being detrimental to ride and to the pavement.
K
M
S
S
g
SD
×
=
Another point worth mentioning is that the rear
deflection should be smaller than the front one. By doing so,
the designer places the pitch axis forward of the center of
gravity. The next section portrays the various types of
springs available to designers and their peculiarities.
Spring Types
There are two basic types of springs used in truck
suspensions: leaf springs and air springs. These two types of
springing perform well with the large load variations
encountered in typical truck operation. The gross vehicle
weight (GVW) may be three times higher than the curb
weight of a vehicle. That is why most truck suspensions,
specially rear ones, posses variable rates.
Figure 5 shows a typical suspension travel curve for
a rear leaf spring. Note the increment in stiffness showing
the variable rate.

8. Note also that the area between the loading and
unloading portions of the graph represent the energy lost due
to friction, or the so-called hysteretic behavior.
-20
0
20
40
60
80
100
120
0 10 20 30
Displacement (mm)
Load(x100kgf)
Loading Unloading
Figure 5: Load x Displacement for a typical trapezoidal leaf
spring.
There are two types of leaf springs, trapezoidal and
parabolic springs. Trapezoidal springs have the following
characteristics [3]:
• Equivalent trapezoidal geometry (shown on Figure 6);
• Each transverse cross-section is subjected to the same
bending moment;
• Most effective stored energy by volume ratio;
• High hysteresis;
• Low maintenance costs.
Parabolic springs present thickness variations along
the various transverse cross-sections (see Figure 7) and they
have the following characteristics:
• Equivalent simply supported beam geometry;
• Constant stress distribution;
• Lower hysteresis;
• Higher maintenance costs.
In almost all Hotchkiss type suspensions, either
trapezoidal or parabolic leaf springs are applied. Each has
certain characteristics that must be explored by the designer
while defining a new suspension.
Figure 6: Equivalent trapezoidal geometry [3].
Figure 7: Typical parabolic spring [3].
Usually leaf springs are heavier in mass than other
types of springs, but they are often used as an attaching
linkage or structural member. So, the designer must be aware
of this advantage in order to provide an economically
competitive suspension.
In this type of suspensions, the static friction is a
function of the vertical load. In fact, both vertical rate and
friction are proportional to the amplitude of the suspension
travel and not dependent on the forcing frequency.
The other type of often used suspension is the
pneumatic spring suspension. This type of spring is basically
“a column of confined gas in a container designed to utilize
the pressure of the gas as the force medium of the spring”
[4].
Although air springs have been applied since 1953
in other countries, only recently have they been increasingly
applied in the Brazilian market. This is due to the fact that
costs of this suspension are decreasing over the last few
years, and major manufacturers are practicing more
emphasis on comfort.

9. Air springs present the following differences in
regards to leaf springs [3 and 4]:
• Vertical rate increases automatically along with load
increments (nonlinear progressive rate);
• Low hysteretic behavior, therefore they are usually
coupled with shock absorbers;
• Relative constant natural frequency over a broad range
of loading
These three characteristics, which ultimately affect
positively the ride perception, are the advantages of
pneumatic springs. But the need of extra control valves and
piping to provide air for the pneumatic springs result in cost
increases over leaf springs, which is an important set back.
Leaf Spring Nonlinearities
Usual designations of spring rate and suspension
deflection imply zero friction. So the nominal spring rate of
leaf springs differ significantly from the instant rate due to
interleaf friction. Experience shows that pitch frequency may
be reduced by as much as 2 Hz with low rate, low friction
suspension [5].
This is due to the fact that most trapezoidal leaf
springs present a hysteretic behavior similar to that of the
leaf spring shown on Figure 8.
Large amounts of friction, as portrayed in the
Figure, are detrimental to ride comfort and can render
useless any effort to properly design a suspension system or
improve an existing one. The underlying reason for this
effect is shown in that Figure.
Actual ride displacements are only a fraction of the
total suspension travel, so in regular use the actual
displacements fall into the inner loop of the spring hysteresis
curves. When in one of these smaller hysteresis loops, the
suspension presents a much higher effective stiffness as the
suspension undergoes small displacements
The mechanism of hysteretic behavior is well
known, and so is its determination. But the designer should
keep in mind at all times that this phenomena is nonlinear in
nature, and that it is dependent on displacement amplitude
rather than on forcing frequency.
In other words, if the interleaf forces are not
sufficient to overcome the static friction between them, the
spring itself will not deflect causing an instant rate increase.
This in its turn results in poor vibration isolation and ride
discomfort.
Figure 8: Load deflection characteristics of a hysteretic leaf
spring [1].
As Gillespie points out, “this type of behavior is
very characteristic of heavy trucks, thus it is often observed
that some heavy trucks ride better on rough roads than on
smooth roads” [1]. This is true since in rough roads the
suspension travel increases forcing the springs to deflect
more, beyond the static friction, and determining spring rates
closer to the nominal values.
So spring friction should always be kept to a
minimum in order to provide a comfortable ride. Figure 9,
below, shows the comparison of the spring rates of a
trapezoidal rear leaf spring in various conditions of use.
0.0
50.0
100.0
150.0
200.0
250.0
300.0
0 10 20 30 4
Alternated course around GVW position [mm]
Rate[kgf/mm]
Used without lubrication
Used w/ lubrication
New
Nominal
0
Figure 9: Rate comparisons for the same spring in different
friction situations.

10. There, it is possible to see the differences between
the nominal rate and to compare it with the actual spring
rates. Note that the rate varies according to the amount of
friction present in the system. As a conclusion from the
information presented on that Figure, one can note the
importance of keeping friction values to a minimal for ride.
This was one of the problems encountered by the
design team in the present work. The hysteretic behavior, as
shown on Figures 8 and 9, resulted in spring rates three
times higher than nominal values. So the vehicle showed
comparatively better ride in rough roads than in paved roads.
Now that the reader possesses the necessary
background on what is the problem at hand, the next sections
provide insights on the design concept. It presents the ideas
followed and then the experiments carried out in order to tun
this new suspension for improved ride, handling, cornering
and braking.
DESIGN CONCEPT
Normally trucks present a 3:1 load variation from
empty to GVW in the rear springs, that is, the load supported
by the rear springs may be three times as high for a loaded
truck than for an empty one. In a tractor, this relationship
may increase to values close to a 6,5:1 ratio!
In front suspensions, this ratio is not quite as high,
but it may be close to a 2:1 increase. Furthermore, there is no
significant difference between a normal truck and a tractor in
regards to front axle load.
Any suspension design must take this load variation
into account, either using a variable rate spring (with an
auxiliary spring) or with a progressive rate spring, when
conceiving a new concept. In this study, the rear suspension
had a high nominal rate and presented a very noticeable
hysteretic behavior. The front suspension presented a
comparatively low rate but also had high hysteretic behavior.
So the problem at hand was to improve the ride of
such a system with an optimal cost x benefit guideline, while
taking into account the targets presented by the clients and
showed in the introduction. In order to achieve these goals,
the points singled out in the paragraph above had to be
addressed. It should be remembered that it is not possible to
solve a lack of spring travel in the rear suspension with a
greater front travel. Thus, solutions found locally had to be
analyzed for their inter-relationships.
Putting together all the ideas presented in the
previous sections, a new primary suspension based on
improvements over the existing one was designed for the
VW 18.310 Titan Tractor. This new design needed to
improve comfort levels and also comply with the design
envelope, cost limitations and customer’s desires.
Among the requirements of this new design, it is
worth pointing out:
• Improve ride comfort;
• Increase payload (by decreasing suspension weight);
• Achieve even lower service requirements and
maintenance costs;
• Transmit less forces to the pavement; and
• Provide easier trailer connection.
A variety of factors that interact with each other in a
very complex manner need to be tuned so as to achieve the
goals set. These factors, as explained in the previous
sections, had to be quantified in some engineering manner in
order to assess the sensibility of the design to each variation.
From the basic observations, it was obvious that the
tractor’s pitching frequency was high due to excessive
suspension hysteresis and lack of suspension travel. Both
these factors are linked with each other and also directly
connected to the spring rate.
Initially, the hysteretic behavior was studied by
introducing lubrication in the springs. As shown in Figure 9,
the lubrication is quite effective in lowering dynamic rates.
But, as the springs remain in service for long times and loose
their lubrication, the ride rate increases causing comfort
problems. Moreover, there is no way to guarantee that all
drivers will lubricate the suspensions, so a new solution had
to be found.
For the rear suspension, an analysis of the rate,
deflection and hysteretic behavior pointed towards a
pneumatic spring. This is due to additional factors, such as:
• Lower spring rate, thus improving comfort;
• Light weight, improving cargo capacity;
• Relative constant natural frequency over a broad range
of loading conditions;
• Low hysteretic behavior; and
• Ability to facilitate trailer hook up by altering spring
height easily;
Furthermore, since it was desirable to keep vehicle
weight down, the pneumatic spring was place behind the
axle with the aid of a “Zeta Spring”. By doing so, more roll
control is achieved, eliminating the need for stabilizer bars.
Figure 10, on the next page, shows a side view of the final
layout of the rear suspension.

11. Figure 10: Layout of the rear suspension.
For the front suspension, although the hysteretic
behavior was less pronounced, it still constituted a problem.
But cost x benefit studies showed that a pneumatic
suspension would not be as attractive as it was in the rear
suspension.
So, parabolic springs were introduced as means to
lower friction, thus lowering dynamic rate. Moreover, the
interactions between front and rear suspensions in the tuning
showed that there was a need to further reduce the damping
while increasing the nominal rate on the front axle.
The following section portrays the development of
this design concept and the tuning process in detail.
EXPERIMENTAL APPROACH
Due to the complexity of vehicle dynamic
phenomenon, which involves multi-body systems and non-
linear effects, the design and tuning process of a suspension
system at Volkswagen is based on the trial and error method
of the experimental approach.
With the improvement and cost reduction of
hardware and software capable of perform those simulation,
this scenario will change in a few years. Although, VW
understands that the experimental approach will be kept in
use as a final certification process since ride is subjective in
nature.
Due to the reasons exposed above, VW Truck and
Bus Operation developed a very robust process to perform
vehicle dynamic experimental tuning, based on subjective
and objective evaluation. This section will explain the tuning
steps, the subjective tests and how they correlated with
objective metrics.
Vehicle Dynamics Tuning Process
The steps of a tuning process are as follows:
1. Tune the springs to achieve good front to rear ride
balance and steady state body control in braking and
roll, but also note their effects on limit stability, steering
response and secondary ride (vibrations);
2. Tune the stabilizer bars for steady state roll control and
limit stability, but also note their effect on head toss,
vibrations and steering response;
3. Tune dampers for primary ride control (body motion),
secondary ride (vibrations), discrete event ride, transient
maneuver body control and transient handling but note
major interactions with steering response;
4. Tune cab suspension (or mounting) rates, anti-roll bar
and dampers for body motion, vibration and discrete
ride;
5. Tune bushings for front and rear axle response, steering
feel, transient stability, secondary and discrete ride but
note the effects with body motion;
6. Tune tires for secondary ride, transient stability, wet
handling, steering response and feel, lateral acceleration
and maximum braking but note interactions with
straight ahead stability and steering disturbances;
7. Tune steering system for steering response, efforts,
precision and torque feedback but note effects on
steering disturbances, straight ahead, cornering and
transitional stability;
8. Tune suspension and steering kinematics for steering
response at front and rear axles and note the effects on
transient roll control, straight ahead stability and
steering torque feedback;
9. Tune steering friction and suspension alignment for
steering response, feel and disturbances.
Observe that this is the most logical sequence on
this interactive process performed by VW technicians, but
flexibility and common sense should always be applied.
Also, those are just the usual interactions and many others
may come up during the tuning process.
A matrix correlating vehicle dynamic attributes and
vehicle systems and sub-systems characteristics helps on
identifying those interactions and the most effective
improvement opportunities. Evaluation of base line
configuration helps to ensure high confidence on the results.
Once an acceptable vehicle dynamic compromise
can not be found, the redesign process comes in action
restarting the tuning using the step by step logical sequence.

12. This process, as state before, is based on subjective
and objective evaluations of the vehicle dynamic metrics.
Vehicle Dynamic Metrics
Vehicle dynamics at VW is divided in the four
usual groups, as showed in the tree structure presented in
figure 11.
RIDE
Primary ride / body motion
Secondary ride / vibrations
Impacts
STEERING
Parking / Maneuvering
Straight Ahead Controllability
Cornering Controllability
Steering Disturbance
Steering Noise
HANDLING
Straight Ahead Stability
Cornering Stability
Transitional / Lane Change Stability
BRAKING
Braking Behavior
Brake Operation
Brake Response
Braking Disturbances
Braking Noise
Figure 11: Vehicle Dynamics Tree Structure.
Each one of the attributes is subdivided up to four
levels of sub attributes. This paper will present the ones
which most help VW to develop the new suspension and the
engineering metrics correlated to them (Table 1).
Attribute/Sub attribute Engineering Metrics
RIDE
BODY MOTION
Bounce displacement Bounce frequency and
amplitude (0.5 – 5 Hz)
Pitch balance Pitch frequency and
amplitude (0.5 – 5 Hz)
Head toss “’B” pillar lateral
acceleration (Peak / 0.5 – 5
Hz)
VIBRATION
Vertical shake Seat track vertical accel. PSD
envelop (5 – 25 Hz)
Longitudinal & lateral
shake
“’B” pillar lateral &
longitudinal accel. PSD
envelop (5 – 25 Hz)
Harshness Seat track vertical accel. PSD
envelop (25 – 200 Hz)
Fatigue limit(*) Exposure Limits Curve
according to ISO 2631
DISCRETE EVENTS
Abruptness Seat track vertical accel. peak
(0.5 – 5 Hz)
Bump trough/topping Axle vertical acceleration
After shake decay Number of oscillation after
initial input
Impact noise Sound level peak [(B(A)]
HANDLING
STRAIGHT AHEAD
STABILITY
Straight running Steering wheel angle
distribution around straight
position @ 80 km/h
Acceleration lift-squat Lift angle @ maximum accel

13. Acceleration wheel
hop/tramp
Wheel hop/tramp frequency
and amplitude decay
Bump steer Straight path deviation @ 80
km/h
CORNERING
STABILITY
Under/oversteering Yaw rate delta [deg/s]
Road holding Maximum chassis lateral
acceleration
Roll angle Cab roll angle @ maximum
chassis lateral acceleration
TRANSITIONAL
STABILITY
Yaw overshoot Lane change yaw rate @ 80
km/h [g/90° Steering wheel
angle (SWA)]
Roll control angle Roll Gradient [deg/g]
Roll control damping Peak to steady state roll gain
ratio
Control activity Lane change steering wheel
angle distribution @ 80 km/h
(90° SWA initial input)
Linearity Lateral acceleration phase lag
time @ 60 km/h
STEERING
STRAIGHT AHEAD
CONTROLABILITY
Center feel Effort versus SWA chart @
80 km/h
STEERING
DISTURBANCES
Wheel fight Steering wheel tangential
acceleration (peak)
Wheel nibble Steering wheel tangential
acceleration PSD envelop
Shimmy Self excited steering wheel
tangential acceleration
frequency and RMS level
BRAKING
BRAKING BEHAVIOR
Dive Dive angle @ 0.6g
deceleration rate
Brake pull Straight path deviation on a
braking from 60 to 0 @ 0.6g
deceleration rate
* The ISO 2631 is used as a comparative guide for body
motion and vibrations study and constitute one more metric
on the vehicle dynamic study
Table 1: Vehicle Dynamics Attributes and Metrics.
Each measurement process is conducted following a
specific test procedure, but the common sense is key to
define additional conditions where the attributes can be
better evaluated. The subjective performance of vehicle
attributes is based on a rate point scale (0 to 10), that
represents the level of customer perception.
To guide the redesign cycles and tuning process, the
engineering metrics were compared against the targets,
established with the benchmarking of both VW current
products and competitor vehicles.
Figure 12 presents the acceleration level at chassis
front and rear end on a highway course (VW 40-300 tractor
with metallic suspension).
Figure 12: Chassis Acceleration on a Highway Course.

14. From this measurement we can observe the
following metrics:
Mode Description Frequency [Hz]
Tractor Bounce 2,3
Tractor Pitch 3,0
Tractor pitch induced by the
trailer(*)
2.6 – 3.2
* Depends on trailer configuration. Values measured with
Stake Load Trailer and Tilting Bucket Trailer respectively
(Figure 13).
Table 2: Bounce and Pitch Metrics.
Figure 13: Tilting Bucket Trailer configuration.
Figure 14 presents the vertical exposition limit for
the current 40-300 on a highway route, measured at the
passenger seat position, according to ISO 2631.
Figure 14: vertical exposition limit for the current 40-300.
The exposition limit for the best in class vehicle
(including full air cab suspension) used as benchmarking is
printed on figure 15.
After several redesign cycles and the tuning
process, the exposition limits for the final suspension setup
are presented on figure 16. Table 3 presents several
comparative results of comfort level obtained during the
development process.
Figure 15: vertical exposition limit for the best in class
vehicle.
Figure 16: vertical exposition limit for the 18-310 pneumatic
suspension.
Test Conditions Exposition Limits
(ISO 2631)
Subjective
Evaluation
18-310 with metallic
suspension
4:59 h 5.5
18-310 with metallic
suspension w/o
lubrication
3:37 h 4.5
18-310 metallic w/o
lubrication suspension
with current 40-300 fifth
wheel set
3:26 h 4.5
Top sold competitor 4:29 h 5.5
18-310 with new primary
suspension set
8:44 h 8.0
Best in class vehicle 10:32 h 9.0
Table 3: comparative results of comfort level

15. As showed on Table 3, with the new primary
suspension, the comfort level of VW 18-310 is among the
class 8 leaders of the Brazilian market.
In order to achieve this comfort level without
affecting the other attributes, specially handling, it was
necessary to go through the interactive tuning and redesign
process for several times, including the major design
changes listed below:
• Change the rear “Z” spring eye position in order to
improve transitional stability and acceleration lift-squat
behavior;
• Change on original design front suspension rate in order
to achieve the best compromise between vibration / body
movements and control activity / linearity;
• Select new fifth wheel positions which could take most
advantage of legal GAWRS without compromising the
ride and handling levels established on the initial
program design assumptions;
• Evaluate the adoption of a Cellasto stopping as
auxiliary spring on the front suspension in order to
improve stability without affecting comfort level;
• Reduce the original design loads of shock absorber
(improve harshness) due to the impossibility of changes
in the cab mounting system.
COMPUTER SIMULATION
As already stated, a multi-body computer
simulation utilizing the ADAMS software was performed by
a third partner. In that study, each subsystem; namely cab,
powertrain and tractor primary suspension; were modeled
and their dynamic characteristics calibrated experimentally
through modal analysis. Then, the complete tractor trailer
computer model was assembled and validated against a
single event experimental test, as describe in reference [6].
The study was concentrated on ride characteristics,
without taking into consideration elastic deformations of
rigid components, such as the frame. Then, secondary ride
results could not be taken in consideration, neither handling
behavior.
The output of the simulation indicates elastic and
dissipatives characteristics compatible with the conclusions
of the present work. That is, a parabolic suspension in the
front axle and a pneumatic suspension on the rear.
The major difference between both studies is the
rate of the rear spring. The computer simulation results
proposed a 25% higher rate in the rear suspension that would
improve even more the comfort level of the 18-310.
But this recommendation must be proved
experimentally due to the limitations of this study, as
commented before.
CONCLUSIONS
The objectives set forth in the introduction were
completely fulfilled. These objectives were:
1. An improvement in comfort that would place the VW
18.310 Tractor among the leaders in its class;
2. Improvements in the handling of fragile loads;
3. Obtain a 6 ton GVW loading condition on the front axle,
without affecting ride;
4. Obtain high mileage comfort robustness; and
5. Provide conditions for most advantage of legal
limitations on trailer height.
Other benefits that were brought about by the
implementation of a new suspension concept include less
accelerations transmitted to the pavement; improved load
capacity through suspension weight reductions; and
improved trailer coupling by the ability of reducing rear
suspension height at the operation.
Regarding the computational analysis and
experimental approach comparisons, there are some
conclusions to be pointed out. Although computational
analyses of vehicle dynamic response are becoming
increasingly common, it is still very much dependent in
experimental validations and on skilled labor.
Moreover, the majority of the work in the literature
uses this tool as means for an initial assessment. This initial
assessment provides the elastic and dissipative
characteristics to be used in a suspension tuning.
By comparing the results of each approach, the
computational analysis and the experimental trial and error
technique, it was possible to see that both arrived at
comparable results. But it is the belief of the authors that a
proper design cycle should combine the advantages of both
methods, blending the computational analysis and the
experimental approach on each step of the tuning process
presented in this paper. The results will be more precise and
less time and cost will be demanded.
The experimental tuning based on a given set of
initial conditions, and following the metrics and steps
described in this paper proved successful. The results of the
tuning arrived at a new primary suspension with better
elastic and dissipative characteristics, which affect positively
the vehicle dynamics characteristics of the new VW 18.310
Tractor.

16. REFERENCES
1. “Fundamentals of Vehicle Dynamics” – Gillespie, T. D.;
Society of Automotive Engineers, Inc.; Warrendale, PA,
USA; 1992.
2. SAE International; “SAE HS-J788. Manual on Design
and Application of Leaf Springs”; SAE, Inc.;
Warrendale –PA; 1982.
3. Costa Neto, A.; “Dinâmica Veicular – Suspensão e
Dirigibilidade”; SAE Brasil – S. Carlos; 1996.
4. SAE International; “SAE HS-1576. SAE Manual for
incorporating Pneumatic Springs in Vehicle Suspension
Designs”; SAE, Inc.; Warrendale – PA; 1994.
5. LeFevre, W. F.; “Truck Ride Guide”; Rockwell-
Standard Corporation; Detroit – MI; 1967.
6. Nogueira, F; Russo, F; et. al.: “Estudo de comforto e
vibração de um veículo cavalo-carreta”; 2o
Colloquium
Internacional de Suspensões; SAE Caxias do Sul - RS;
2002.
7. ISO 2631; “Evaluation of human exposure to whole-
body vibration” – Parts 1, 2 and 3.
8. SAE J1490; “Measurement and Presentation of Truck
Ride Vibrations”/ SAE Recomende Practice