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Design of Electromechanical Height Adjustable Suspension
Article    in  Proceedings of the Instit ution of Mechanic al Engineer s Part D Journal of A utomobile Engineering · Oct ober 2017
DOI: 10.1177/0954407017728633
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Design of Electromechanical Height
Adjustable SuspensionJournal Title
XX(X):1–15
c
The Author(s) 2016
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DOI: 10.1177/ToBeAssigned
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Nicola Amati, Andrea Tonoli, Luca Castellazzi, and Sanjarbek Ruzimov
Abstract
In the general context of vehicles’ fuel consumption and emissions reduction, the minimization of the aerodynamic drag
can offer not negligible benefits regarding the environmental issues. The adjustment of the vehicle height is one of the
possible ways to provide a reduction of the resistances to vehicle motion, in addition to consequent aspects regarding
the increased versatility of the vehicle. The aim of this paper is to present in a systematic way to the state of the art
of height adjustment systems for passenger vehicles, summarizing the main modes of operations, working principles
and architectures. Particular attention is then given to electromechanical systems, which represent the next trends for
future vehicles due to their high reliability and relatively low costs. A design methodology for electromechanical height
adjustment systems with the purpose of optimizing their performance is presented. Such procedure is able to reach
the most efficient working point even in presence of constraints of different nature. Prototypes have been designed,
produced and tested to demonstrate the potentialities of electromechanical height adjustment systems. Furthermore,
potential benefits and drawbacks of using such systems are highlighted.
Keywords
Vehicle suspension systems, self levelling, height adjustment, aerodynamic drag, fuel consumption reduction, emission
reduction.
Introduction
Road transport is one of the most significant consumer
of fossil fuels and source of air pollution. According to
the data of the International Energy Agency,1in 2013 the
road transport was the second largest sector of Carbon
Dioxide ( CO2) emissions. Its share in global world CO2
emissions was 20% (12% in Europe2). This justifies the trend
towards more stringent legislative restrictions and standards
on vehicle emissions such as, for example, those introduced
by European Commission to set targets for new cars.2Such
targets require that new cars have to emit less than 130
gCO 2=kmby 2015, and less than 95 gCO 2=kmby the
end of 2020. Considering that between the period 2010-
2014 the average emission level decreased2of 17 gCO 2=km,
an additional decrease of 35 gCO 2=kmfor the same time
frame seems to be challenging for vehicle manufacturers.
The penalties for exceeding the limits on one side and the
incentives towards lower emission vehicles2on the other,
encourages the application of innovative technologies to
reduce emissions.
Technologies that are widely used to reduce fuel
consumption and exhaust gas emissions are based onimproving the internal combustion engine (ICE) and the
drive train efficiencies, on lowering vehicle weight and on
reducing the resistances.3;4The latter can be achieved by
optimizing the aerodynamic shape and reducing the frontal
area.5;6Since the lowering of the vehicle height could
give benefits in both the aerodynamic drag coefficient and
reference area, it is then considered a promising feature to
be used on a modern vehicle to reduce fuel consumption
andCO2emission. Considering that one of the trends of
the last decade shows a wide use of 4×4 crossover SUV
cars (13 %share in European market in 20157), especially in
city conditions, the implication of height adjustment systems
seems to give interesting opportunity to reach a compromise
between versatility of the vehicle and fuel consumption.
Mechatronics Lab, DIMEAS, Politecnico di Torino, Italy
Corresponding author:
Ruzimov Sanjarbek, Mechatronics Lab, Department of Mechanical and
Aerospace Engineering, Politecnico di Torino, Duca degli Abruzzi, 20,
Torino, Italy
Email: sanjarbek.ruzimov@polito.it
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2 Journal Title XX(X)
In spite of the above mentioned benefits, the design
of the height adjustment suspension as a standalone
system was not studied previously in the literature.
Rather, it is more addressed as an intrinsic feature of
an active suspension.8–13To fill this gap, the present
paper investigates the environmental benefits of using such
systems and gives detailed critical analysis of existing
height adjustment suspensions. Furthermore, the paper
describes a methodology to design an electromechanical
height adjustment system and to optimize its performance.
The prototypes of height adjustment suspension system
comprising electric motor, speed reducer and screw-nut
mechanism have been built and tested on a vehicle. The
validation of the design methodology is lastly described.
Environmental Benefits of Vehicle Height
Adjustment Systems
The benefits of adjusting vehicle height are analyzed by
using a vehicle model including the longitudinal vehicle
dynamics.14An average A-segment vehicle was chosen
for the analysis. The vehicle data are given in Table 1.
The engine torque and the specific fuel consumption maps
implemented in the model are experimentally measured
by the manufacturer. The aerodynamic drag coefficient of
the vehicle at nominal road clearance is 0.35. It has been
measured numerically and experimentally, that decreasing
the road clearance by 20 mmand 40 mm, this coefficient will
reduce to the values indicated in Table 1. Running the vehicle
simulator at different road clearances, fuel consumption and
CO2emission values for New European Driving Cycle
(NEDC) and Worldwide Harmonized Light vehicles Test
Procedures (WLTP) are obtained. During the simulations,
the value of the aerodynamic drag coefficient is decreased
from nominal to one of the reduced values (0.33 or 0.31)
only on extra urban parts of the homologation cycles. Fuel
consumptions obtained by the simulations are then converted
inCO2emission by using the conversion factor suggested in
EC Regulation No 443/2009 for gasoline engine (i.e., 2330
gCO 2per one lof petrol).15
The results of simulations are depicted in Figure 1. The
fuel consumption reduction at different road clearances
compared to the nominal value is in the range of 0.037-0.22
l=100km (0.88 – 4.41 %), which corresponds to reduction of
CO2emissions by 0.87 – 5.17 g=km. These results justify
the use of height adjustment systems on vehicles to achieve
lower CO2emissions. Besides that, added benefits include
possible reductions of body roll16;17and improved vehicle
accessibility. By rising the vehicle height, higher versatility,
Figure 1. Vehicle fuel consumption at different values of
aerodynamic drag coefficient while driving on two homologation
cycles.
i.e. adaptability of the vehicle to different road conditions can
be achieved.17
Table 1. The vehicle data used in the simulations.
Parameters Value
Vehicle mass, kg 1090
Engine volume cm3900
Nominal power kW 60
Cxat nominal vehicle height, 0.35
Cxat 20 mmlowered height, 0.33
Cxat 40 mmlowered height, 0.31
Analysis of Existing Systems on Vehicle
Height Adjustment.
The main components of a generic suspension systems are
primary and secondary elastic and damping members. They
damp out the oscillations and isolate the vehicle body from
impacts coming from road irregularities, improving ride
comfort and to guaranteeing the contact between tire and
road surface (i.e. road holding capability).8–13;16;19;20;23;24;26
Based on the operation modes of damping element,
vehicle suspensions can be divided into: passive, semi-active
and active.8–13Passive suspensions’ components have non-
adjustable characteristics, whilst in semi-active suspensions,
the damping of the system can be varied according to an
input signal. Active suspensions differ from semi-active ones
for what the energy injected into the system concerns. In
general, in semi-active suspensions the required energy is
limited to be enough for actuating control valves.8Vehicle
height adjustable suspensions are classified into active
suspensions with small actuation bandwidth.8;26A specific
type of height adjustable suspension is the self leveling
suspension, an architecture in which the vehicle height is in
advance set to optimal level and the system maintains this
level under different loading circumstances (allowing proper
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Amati et. al. 3
weight distribution). Basically, height adjustment is fulfilled
by increasing or decreasing the relative distance between the
vehicle body and the wheel hub. This can be mainly realized
by acting on different parts of the suspension strut, such
as: upper spring seat, lower spring seat or shock absorber
tube (by moving telescopically). Depending on the type of
actuation, height adjustment systems can be of a pneumatic,
a hydraulic, a hydro-pneumatic and an electromechanical
kinds. Here below, summary description and analysis of the
state of the art of different typologies of height adjustment
systems are reported.
Hydraulic.
Hydraulic height adjustment system (Figure 2) comprises
of linear (hydraulic cylinder) or rotary (hydraulic motor)
actuators, hydraulic pump (standalone or combined with
already existing one), pipes and hydraulic fluid, control
valves and sensors. Relative displacement of the piston and
cylinder of the actuator unit allows to modify the height of
the vehicle.
To improve ride comfort and vehicle stability, Mercedes
Benz AG18–20uses an hydraulic active suspension system
that features height adjustment capability. A hydraulic linear
actuator mounted in series to the helical spring and in parallel
to the damper of a traditional suspension system is used
(schematically shown in Figure 2(a)). Pressures up to 200
bar are supplied by an engine driven pump (in schemes,
Mis used to denote a driving motor or an engine). The
fluid stored in accumulators ensures faster actuation times.
Hydraulic servo valves guarantee the independent levelling
of each vehicle corner. The upper spring plate is attached to
the cylinder of the hydraulic actuation unit while the piston is
connected to the vehicle body18;19. Thus, by pumping fluid
into the cylinder the piston moves and the relative distance
between the vehicle body and the upper spring plate changes.
Van der Knaap21patented an active suspension system
in which the linear hydraulic actuator is integrated with
the shock absorber. Active control of damping force is
accomplished by means of controlled directional valves and
an electric motor driven pump, which actuates the piston-
cylinder unit of the shock absorber (Figure 2(b)). Using this
system, the adjustment of the vehicle height is performed
acting on the shock absorber piston, that moves relative to the
tube fixed on the wheel hub. Integrating the piston-cylinder
unit with the one already existing in the shock absorber,
reduces the system cost. However, additional control valves
are required to decouple height adjustment feature from
damping purposes.ZF Sachs Nivomat shock absorbers are used in rear
suspensions to restore vehicle height to predefined levels
at different loading conditions9;22. The system does not
require an external pump (hence, motor and pump in Figure
2(b) should not be considered). Upper and lower volumes
are separated by the piston of the damper and they are
interconnected by means of pipes and valves. The pumping
of the hydraulic fluid from one volume to another is realized
owing to the relative motion between vehicle body and
wheel hub (sprung and unsprung masses, respectively). It
is obvious that changing of the level occurs only after
traveling some distance22depending on the load. While the
system of Figure 2(a) uses additional actuator unit (cylinder-
piston) to conventional spring and shock absorber, other
solutions provided by BMW and ZF Sachs Nivomat realize
height adjustment systems integrated into the damper (Figure
2(b))21;22.
Figure 2. Hydraulic height adjustment system schematics: (a)
with additional cylinder-piston unit; (b) with cylinder-piston unit
integrated into damper.
The hydraulic height adjustment systems are advantage-
ous when fast actuation of the system is the main requi-
rement. As they deliver control force at high rate allowing
fast levelling, thus the active feature can be provided. In
addition, the main components of the system can be fixed
on vehicle sprung part, leading to slight modification of
the unsprung mass21;22. The installation size of hydraulic
actuation unit is more compact. However, the presence of
hydraulic fluid and high pressures in the system represent
main drawbacks of the hydraulic suspension systems. As the
hydraulic fluid may cause a corrosion of metallic parts due
to fluid moisturization, the prevention of latter requires use
of special fluids. Furthermore, due to high pressures in the
system the tight tolerances in manufacturing of the cylinder-
piston group is mandatory. As a result, high cost of the
system components is expected. In fact, the adaptation of
hydraulic height adjustment system as a constituent part of
a multi-functional active system is more practical.
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4 Journal Title XX(X)
Pneumatic.
Pneumatic suspensions are mainly used in public transpor-
tation buses, heavy-duty vehicles and premium cars where
increased passenger comfort is required. They feature height
adjustment systems to allow easy entry/exit of passengers.
Figure 3(a) lists main components used in pneumatic suspen-
sions. They consists of air springs (usually, reinforced rubber
sleeve/bag inflated with pressurized air), high pressure reser-
voir (tank), air compressors, air dryer (to remove moisture)
and control system (including control valves, height and
pressure sensors). Adjustment of the ride height is straight-
forward and can be accomplished by pumping the air into or
out of air spring bag.
Hirose et al. in23describes Toyota’s Electronic Modulated
Air Suspension (TEMS) system, which controls suspension
spring rate, damping force and vehicle height by means
of controlling the amount of air in the air spring as a
function of vehicle traveling conditions. During braking and
cornering spring rate and damping force are increased to
reduce pitch and roll. At high speeds the vehicle height
is lowered to increase stability. Height is increased while
driving on rough roads to avoid vehicle hits bumps. The
system complexity (and cost as a consequence) is high, as it
includes sensors (height, steering and throttle position) and
actuators to control the spring rate in addition to already
expensive pneumatic components.
The application of a pneumatic suspension to the
McPherson strut is presented by Tener.24Rolling lobe
variable rate air spring, in parallel with hydraulic damper
is used to find a compromise between ride and handling
performance with three different settings (Track, Sport and
Touring). The height of the vehicle can be decreased to
improve the handling performance and by varying the air
volume in the rubber sleeve.
Fiat Chrysler Automobiles uses Quadra-Lift rair suspen-
sion system on its top line of Grand Cherokee model.25It
features height adjustment system with five different height
modes (Normal, Park, Aero, Off road 1 and Off road 2).
To allow easy entry/exit of the passengers and loading-
unloading of the cargo, the vehicle height is lowered by 40
mmwith respect to normal operating height when parked. At
high speeds the vehicle height is lowered of 13 mm to reduce
aerodynamic drag, which leads to improved fuel economy
on highways. Two off road modes increase ground clearance
by 25 mm and 53 mm to overcome obstacles and guaranty
improved off-road capability.
Self leveling suspension system by BMW is based on a
pneumatic actuation.26When the vehicle loading conditions
change, the vehicle height is restored to the preset positionby controlling the air pressure in the pneumatic cylinders
mounted on the rear axle. Two electrically operated external
compressors work independently from each other, allowing
independent actuation of left and right corners.
Figure 3. (a) Pneumatic and (b) Hydro-pneumatic height
adjustment system schematics.
Off the shelf retrofittable solutions available on the
market are mainly based on pneumatic suspensions. Height
adjustment is performed to improve vehicle stability, ride
comfort and handling under different loading conditions
(loaded, towing the trailer and etc.). On these solutions,
selection of rational vehicle height depends on driver’s
experience.
The main advantage of pneumatic suspensions is the
possibility of controlling the spring rate of the air
springs. It allows to find a good compromise between
improved handling performance and ride comfort.23–26Other
advantages are those of intrinsically owning the height
adjustment feature and a relatively fast actuation speed
in presence of stored pressurized air. However, high cost,
reduced reliability and robustness (mainly due to failure of
rubber air springs) and high maintenance requirements are
the main drawbacks of this kind of suspension systems.
Moreover, pneumatic suspensions are less efficient when
significant changes of the loading conditions take place or
fast and frequent height adjustment is required (especially,
when pressurized air store is depleted). Due to high cost
of the components used in pneumatic suspensions, they are
mainly used on top E- and F-segment models.
Hydro-pneumatic.
Hydro-pneumatic suspensions use elements of both hydrau-
lic and pneumatic suspensions. The hydraulic part of the
system is responsible for providing damping and can be
used for height adjustment, while pneumatic part with high
pressure accumulator provides elastic properties to the sus-
pension. Simple schematics of the system is depicted in
Figure 3(b).
Industrial pioneer in using hydro-pneumatic suspensions
has been Citroen, applying it on the vehicle back in 1950s
on it’s DS model.9The system is still used in luxury
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Amati et. al. 5
models of Citroen with some upgrades of it’s control logic
and components. In the early system proposed by Citroen,
a high pressure accumulator filled with nitrogen gas (is
used to avoid corrosion of components) works as a spring
with variable stiffness. The nonlinear behavior of the spring
ensures improved ride comfort both when fully loaded
(stiffer spring, gas in the accumulator is more compressed)
and unloaded (softer spring, gas in the accumulator is less
compressed). The accumulator is attached to the end of the
hydraulic actuator, where gas and liquid is separated by
means of a membrane (Figure 3(b)). Height adjustment is
performed hydraulically by pumping fluid into or out from
the hydraulic cylinder (thus, by varying volume of the fluid).
In addition to the suspension, braking and steering systems
are connected to the same pump.
Sarel F. and Els P.17studied an application of hydro-
pneumatic suspensions to prevent roll over on off-road
vehicles. However, the proposed system can be equally
adapted for the vehicle height adjustment. The proposed
configuration of the system is comparable to one of Citroen.
Nevertheless, the particular difference is two nitrogen
pneumatic accumulators, that allow a higher flexibility in
tuning elastic and damping properties of the suspension. A
smaller accumulator with volume of 0.1 lis used to have
stiffer spring, while by enabling larger volume accumulator
of 0.4 lsofter characteristics is obtained. The system uses an
increased number of valves (three per corner) which ensures
enhanced performance of the suspension.
Summarizing pros and cons of hydro-pneumatic suspen-
sions, it can be highlighted that they combine advantages
of both hydraulic and pneumatic suspensions9. They allow
faster height adjustment speed, higher reliability and robus-
tness of the components, more compact size and improved
performance. However, their use is limited to high end
segment cars due to high production cost and the need of
specialized service of the components.
Electromechanical.
Most of the available literature until late 90s is devoted to
hydraulic, pneumatic and hydro-pneumatic vehicle height
adjusting devices and they are available commercially on the
market. Starting from the mid 2000s, new solutions using
electromechanical actuators and electronic control units
started to appear in literature, specifically in form of patents.
This could be the result of car manufacturers’ attempts
to find low cost height adjustment suspension systems
applicable also on low cost car segments. So far, no solutions
using electromechanical system are present on the market.
In principle, all the electromechanical height adjustmentsystems make use of linear translational part, usually by
means of threaded screw-nut mechanism connected to the
suspension components and actuated by electric motors
directly or through speed reducer (Figure 4).
AUDI AG’s patented a solution for electromechanical
height adjustment system with ad-hoc electric motor and
ball-screw mechanism27. The stator of the electric motor
is fixed to the vehicle body through a rubber element
comprising low friction bearing while the rotor of the motor
is connected to the ball-screw. The nut is connected to
upper spring seat (Figure 4(a)) and translates along the
axis of the shock absorber providing height adjustment
for the vehicle body. Additional to the main spring, a so
called compensating spring (not shown in Figure 4(a)) is
included between the vehicle body and the nut. It prevents
the system components from shock loads. Internally, the
hollow ball-screw and the shock absorber tube can move
telescopically, relative to each other. The presence of the
ball-screw increases the overall system efficiency. Therefore,
overhauling may occur under the load of the vehicle body
weight. As the locking feature is not intrinsically included
in the system, additional locking device must be used (even
if is not mentioned in the patent). The presence of the ad-
hoc electric motor, the additional spring and the ball-screw
sophisticates the system construction, possibly resulting in
higher cost of the system.
Figure 4. Electromechanical height adjustment system
schematics applied to independent suspension: (a) Upper
spring seat actuation; (b) Lower spring seat actuation; (c)
Actuation by means of lever mechanism; (d) Shock absorber
tube actuation.
Hakui T. et al.34presented a solution of a height
adjustment system using electromechanical actuation. The
speed reduction stage takes advantage of differential gear
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6 Journal Title XX(X)
principle to obtain large reduction ratio using two spur gear
pairs. The actuation principle of this solution is similar to one
depicted in Figure 4(a)). A small locking torque is required
to provide self locking owing to the presence of the reduction
stage. In this case, the locking torque is provided by shunting
the electric motor windings.
Another solution patented by AUDI AG uses a ball-screw
connected to an electric motor through reduction stage of
gears and a nut, connected to the lower spring seat28. The
schematic view of the solution is shown in Figure 4(b).
The clearance between a rubber bump stop and the shock
absorber tube is varied by connecting the surface to the
translating nut to avoid over compression of the spring
during bumps. The system is less complex compared to the
previously mentioned one, however, the employment of a
ball-screw still increases the cost.
In patents29and30, the actuator group is fixed to the
vehicle body as shown in Figure 4(c). The solution described
in29uses a rotary actuator with reduction gear stage and
an arm to move lower spring seat. A screw-nut mechanism
connected to lower spring seat by means of a leverage
mechanism is used in30. Self-locking feature can be included
intrinsically in reduction gear stage on the electric motor
shaft29, in the screw-nut30or using the electric motor
braking.29;30The main advantage of this kind of system is the
compactness. In fact, there is no need to embrace the shock
absorber envelope or that of providing further amplification
of the lifting force due to presence of leverage mechanism.
At the same time, the presence of additional spring and lever
system may lead to increased cost.
Kim et al. of Hyundai Motor Company have patented a
vehicle suspension system with electronic control to adjust
vehicle height in real time during cornering, braking and
acceleration.31;32As shown in Figure 4(d), the hollow ball-
screw is fixed on the shock absorber tube and moves
relatively to wheel hub in a telescopic way changing the
vehicle height.31The nut is connected to the output of
the electric motor through a reduction gear stage and is
supported by radial bearings fixed on the strut brackets of
the wheel hub. The nut takes its rotation from the electric
motor. The ball-screw translates along the axis of the shock
absorber, changing the distance between wheel hub and the
vehicle body. Similar solution is described in,32where a
grooved shaft (instead of the ball-screw in previous solution)
is connected to the shock absorber tube and a roller element
is guided in the groove. The stator of the motor is mounted
on the strut brackets fixed on the wheel hub. Interpretation of
how overhauling is avoided in the systems is not discussed
in the patents. The main drawback of these systems is thatthey transmit all the force flowing from the vehicle body to
the wheels (and vice versa), leading to accelerated wear of
the components. In addition, involvement of ad-hoc electric
motor, of rollers and of additional bearings, increase the cost
of the system.
An electromechanical system to adjust vehicle height
applied to torsion beam suspension is described in patent33.
Usually, torsion beam suspensions are characterized by
non coaxial shock absorber and spring. Therefore, actuator
dimensions can be more compact compared to the case
of McPherson or double wishbone suspensions as the
actuator can be located inside the helical spring. As in29;30,
suspension coil spring is divided into two parts separated
by movable spring seat to prevent the actuator system from
impact loads. Different diameters of springs are used to
operate in a telescopic way and make the system more
compact. The stator of the electric motor slides axially on
a splined surface (integrating thereby anti-rotation system).
The rotor of the motor is connected to the ball-screw while
the nut is fixed on a movable plate and translates axially
(Figure 5). Both upper and lower spring seat actuation
(Figures 5(a) and (b), respectively) can be applied in this
architecture and indeed, both are mentioned in the patent.33
The system is simple in construction and may be retrofitted
to existing suspension relatively easy. However, splitting the
spring into two parts may alter suspension characteristics.
Figure 5. Electromechanical height adjustment system
schematics applied to semi-independent twisted beam
suspension: (a) Upper spring seat actuation and (b) Lower
spring seat actuation.
The main disadvantages of the electromechanical height
adjustment systems are lack of ability to vary elastic behavior
and lower leveling speed. Since the former disadvantage
is inherent to all the modern passive suspensions, it can
be omitted. The height adjustment is required only when
the driving conditions are changed, for example, from low
speed city conditions to high speed highways or from smooth
to rough terrains. In these cases, the leveling speed is not
critical, as actuation can be performed in a larger time. The
following particularities of these systems can be attributed to
their advantages:
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Amati et. al. 7
Higher reliability and robustness (when failure of
the system occurs, the system behaves much like
traditional suspension system)
Lower initial cost of the system
No special experience needed for the service
Overall compact size of the components
Possibility of having modular design (the system can
be retrofit to the traditional suspension)
Therefore, it is reasonable to use electromechanical height
adjustment systems when a trade-off between high reliability,
low cost, compact size and fast actuation is required.
Design Methodology of Electromechanical
Systems
Literature review presented in the previous section shows
that electromechanical systems represent the last trend in
vehicle height adjustment systems. Mechanical components
such as threaded screws and gear trains offer high reliability
and reasonable responsiveness, while the possibility of using
commercial electric motors for automotive applications
could lead to an acceptable balance between cost and
benefits. The literature is focused on constructive and
functional aspects of electromechanical systems, and does
not adequately address the optimization of the main design
parameters. To fill this gap, the aim of the present section is to
propose a design methodology to optimize the performance
while being compliant with imposed constraints. Such
constraints can be of an electrical, a mechanical or
a geometrical nature and may vary depending on the
suspension type. McPherson (front) and twisted beam (rear)
suspension schemes are considered in the following due
to their wide use in different vehicle segments. However,
the proposed methodology can be equally adapted to other
suspensions by considering the relevant constraints. In case
of McPherson scheme, the actuation of the upper or the lower
spring seats (Figures 4(a) and (b)) is more advantageous
than the movement of the shock absorber tube, as only the
loads supported by the springs act on the height adjustment
system. Additionally, this system does not affect the role
of the shock absorber in the kinematics of the suspension.
However, solution with upper seat actuation may require
significant modification of the vehicle frame around the
suspension tower with a relevant impact on a region that is
very critical for pedestrian protection. This could represent
a critical issue especially in case of retrofitting the system
to existing suspension. Therefore, due to the absence of
the above mentioned issues, lower spring seat actuation isconsidered in case of McPherson suspension scheme as
shown in Figure 6.
Working Principle of the System
Figure 6. Schematics of the height adjustment system with
lower spring seat actuation.
An electric motor 4bcoupled with mechanical speed
reducer 3a, 3b and3cdrives the nut 5b(Figure 6). Power
is supplied by onboard electric battery 4a. The speed
reducer can be of different kind like planetary or parallel
axis gear trains, belt and pulleys or chain and sprockets.
Even if in all the cases the functionality does not change,
reliability, weight, cost and transmission efficiency can be
driving factors in choosing one instead of the other. In
the arrangement shown in Figure 6, the speed reducer is
composed of a planetary gear head 3band a parallel axis
gear stage with pinion 3aand gear 3c. Rotational motion is
then transformed into linear by means of screw 5aand nut
5bthat is part of gear 3c. The screw 5ais internally hollow
and mounted on the shock absorber tube 2b. The compound
bearing 6(axial thrust and radial bearings) decouples the
axial motion from the rotations of the nut ( 3cand5b). In
case of extreme deflection of the spring 2a, the upper part of
the shock absorber tube 2bcomes into contact with rubber
bump stop 8to avoid failure of the former.
Definition of Optimal Parameters of the
Electromechanical Components
The procedure for defining the optimal parameters of the
electromechanical components includes the following steps:
Determination of the required power to lift the load;
Application of geometrical and irreversibility con-
straints;
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8 Journal Title XX(X)
Selection of electric motor;
Application of current and time constraints;
Definition of geometrical and mechanical parameters.
The power required for the system to behave as desired
is the first fundamental parameter to design a power
transmission system. It affects the selection of the electric
motor and further components design.
The efficiency of the transmission from the shaft of the
electric motor 4b(Figure 6) to the lower spring seat 7can be
computed as:
total=srs=ggbs (1)
where,gbis efficiency of the gearhead 3b,gis efficiency of
pinion 3aand gear reducer 3c, andsis efficiency of power
screw-nut mechanism (components 5aand5b). Usually,gb
andgare given in the manufacturers datasheets. The power
screw efficiency can be computed as:
s=tan
tan (+)(2)
whereis the friction angle, function of the friction
coefficientfand thread angle (= arctan(f=cos ). The
parameteris the screw lead angle and can be defined as
function of the screw pitch pand its mean diameter d. For
single thread power screws this angle can be computed as:
tan=p
d(3)
The average power required from the electric motor can be
computed for a given external load Fn, travel stroke uzand
actuation time tas:
Preq;em=1
totalFnuz
t(4)
Isolines of the screw efficiency scomputed using
Equation 2 are plotted in Figure 7 as a function of screw
pitch and diameter. It is showing that the smaller is the screw
diameter and the larger is the screw pitch, the higher is the
efficiency and, as a result, the required power is smaller.
The procedure of design optimization includes finding the
different pairs of dandpwhich satisfy all the imposed
constraints. These constraints are related to the geometry and
the performance of the system and can be summarized as:
1. Irreversibility s<50%
2. Geometry dmin>dtube+s
3. CurrentIss<Iss;max
Figure 7. Isolines of screw efficiency s,%. Friction coefficient
f=0.2.
4. Time (or Speed) t50 mm<tmax
wheredminis minor (root) diameter of the screw, Issis the
current absorption at steady state and t50 mm is the time to
cover a given distance (in this case 50 mm).
The first constraint is related to the need of having an
intrinsically irreversible actuator at the screw-nut mechanism
level. To satisfy this constraint the screw efficiency s
has to be less than 50 % that is in general the threshold
of irreversibility of mechanical transmission systems. The
small area indicated with number 1in Figure 8 must be
omitted from further consideration as it does not satisfy this
constraint.
Figure 8. Isolines of required power Preq;em,W. Numbered
areas are omitted due to imposed constraints ( 1- Irreversibility,
2- Geometry). Point Pindicates minimum power after applying
the constraints.
The second constraint is related to construction issues.
In order to install the screw on the shock absorber tube,
the minor diameter of the screw is substantially given. This
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Amati et. al. 9
consideration is valid for independent suspension schemes.
In case of semi-independent suspensions (e. g. twisted beam)
the screw can be placed separately from the damper (i.e not
coaxial). Therefore, the minimum screw diameter should be
defined based on admissible contact pressure developed on
the screw-nut contact surface. The screw body diameter can
be found as a sum of shock absorber tube diameter dtube
and wall diametral thickness sof the screw body. Imposing
this constraint (e. g. dtube = 46.5 mm ands= 2mm), the
region which does not satisfy it can be defined (the area
is indicated with number 2in Figure 8). As a result, the
minimum power of the electric motor can be defined, which
is indicated as point Pin Figure 8. Values of screw pitch and
nominal diameter are therefore obtained.
Based on the value of the minimum required power,
industrially available electric motor can be selected or a
custom one can be designed. Upon the selection of the
electric motor, its characteristics (torque and speed constants,
winding resistance and inductance, motor nominal speed and
rotor inertia) will be defined for further system performance
analysis.
The evaluation of the electric motor current absorption and
the time needed to cover 50 mm distance depends on the
dynamics of the electromechanical system. With reference to
the scheme shown in Figure 6, a dynamic model is developed
and described below.
The voltage equation for equivalent DC circuit shown in
Figure 6 can be written as:
VDCRILdI
dt=ke!em (5)
whereVDCis the supply voltage of the vehicle battery, Rand
Lare resistance and inductance of winding respectively, Iis
the current flowing in the windings, keis the motor speed
constant and !emis the electric motor angular velocity.
The torque balance on the electric motor shaft is:
ktI=Jeq_!em+Tr (6)
wherektis the motor torque constant, Jeqis the equivalent
moment of inertia of rotating and translating elements, _!em
is the angular acceelration of electric motor shaft and Tris
the resistant torque due to load Fn, reported on the electric
motor shaft.
The equivalent rotational inertia Jeqis:
Jeq=Jem+Jgb+Jp
i2
gb+Jn
i2
gbi2g+m
i2
gbi2gi2s(7)
whereJemis the electric motor shaft inertia, Jgbis the
gearbox inertia, Jpis the pinion inertia, Jnis the nut inertiaandmis the corner mass. igb,igandisare the speed
reduction ratios of the gearbox, of the parallel axis gear stage
and of the screw-nut mechanism, respectively.
Similarly, the resistant torque reported at the electric motor
shaft is:
Tr=Fn
igbigistotal(8)
To write Equations 5 and 6 in a matrix form, the system
states and inputs has to be defined. The states of the system
are motor shaft angular speed !emand angular displacement
em, and current I. The inputs are supply voltage VDCand
resisting torque Tr. Then, the matrix form is:
f_xg= [A]fxg+ [B]fug (9)
and in particular:
8
><
>:_!em
!em
_I9
>=
>;=2
640 0kt
Jeq
1 0 0
ke
L0R
L3
758
><
>:!em
em
I9
>=
>;+2
6401
Jeq
0 0
1
L03
75(
VDC
Tr)
(10)
The angular displacement of electric motor emcan be
translated to a linear displacement uzof the nut considering
transmission in between these two elements as:
uz=em
igbigis(11)
where speed reduction ratio isof the power screw-nut
mechanism can be found using Equation 12:
is=2
p(12)
While the speed reduction ratio of the power screw
mechanism is defined using Equation 12, the total reduction
ratio of the speed reducer isris unknown in Equations 7,
8 and 11. To define it, kinematic link between motor and
load can be used. In this case, nominal speed of the motor
!em;nom is given in motor specification sheets and linear
speed of the system can be derived from requirements (for
example speed required to cover 50 mm in 12 s). The
equation to define the total reduction is therefore defined
from Equations 11 and 12:
isr=igbig=!em;nomt
uzis(13)
Onceisris defined, values for igbandigare defined
depending on geometrical constraints keeping their product
constant and equal to isr.
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10 Journal Title XX(X)
Based on above considerations, the optimization proce-
dure can proceed further. By selecting a range of values for
screw pitch ( p=pi) and mean diameter ( d=dj), perfor-
ming loop cycle all the dependent variables of interest can
be computed in the following order:
1.is(i) = 2=p(i)
2.itotal(i) =igigbis(i)
3.total(i;j) =gbgs(i;j)
4.Jeq(i) =Jem+Jgb+Jp=i2
gb+Jn=[igbig]2+
+m=[igbigis(i)]2
5.Tr(i;j) =Fn=[igbigis(i)total(i;j)]
The dynamic model for the electromechanical system
described above, runs simultaneously inside the same loop
in order to compute the current Iss(i;j)and the actuation
timet50mm (i;j).
The third constraint refers to a limitation on the electrical
power. The steady state current Issflowing through the
electric motor windings has to be smaller than a limit
valueIss;max. The latter is set based on maximum current
capabilities of the electrical machine and its control board.
The fourth constraint affects values of both mechanical
and electrical parameters. The actuation time is an important
design parameter for an height adjustment system, especially
in case of active suspensions. The required power is
substantially affected by this value. If the travel uzis
considered to be equal to 50 mm, then time to cover this
distancet50 mm has to be less than imposed limit tmax.
Figure 9. Current isolines, A. Numbered areas are omitted due
to imposed constraints: ( 1- Irreversibility, 2- Geometry, 3-
Maximum current absorption, 4- Maximum actuation time).
Figures 9 and 10 show the steady state current and the
actuation time, respectively, when all four constraints are
imposed (numbers indicate constraints). Current absorbed by
electric motor increases mainly with the increase of mean
screw diameter and is less influenced by change of screw
pitch (Figure 10). On the contrary, the actuation time is less
influenced by the mean screw diameter, but the effect of
Figure 10. Actuation time isolines, s. Numbered areas are
omitted due to imposed constraints: ( 1- Irreversibility, 2-
Geometry, 3- Maximum current absorption, 4- Maximum
actuation time).
screw pitch is more evident (Figure 10). A larger pitch causes
a reduction on the actuation time due to an increment of the
screw transmission ratio is.
Experimental validation
To evaluate the potentiality of using height adjustment sys-
tems in the vehicle and to validate the proposed methodology
has been adopted to design an electromechanical height
adjustment system integrated in a compact vehicle with
front McPherson strut and rear twisted beam suspensions.
The main design choices were: (1) Retrofit capability of
the system; (2) Use of power screw – nut mechanism to
intrinsically obtain the system irreversibility and (3) Low
cost of the system.
Prototypes of actuators for front and rear suspension
The design methodology was followed to define the
optimal values of the system parameters. The main input data
for the optimization procedure are reported in Appendix A1.
For the front suspension, the minimum power to lift the load
Fn= 3700 Nto the distance uz= 50 mm int= 13 swith
the shock absorber tube diameter dtube = 46.5 mm can be
calculated using the Equation 4. The point Pin Figure 8
shows that this power is about 83 W. It can be obtained at
point where screw pitch p= 10 mm and diameter d= 53.5
mm. One possible off the shelf motor is Maxon brushed DC
RE35 with 90 Wnominal power.35This motor was selected
with consequent modification of the screw pitch to p= 8
mm (hence,dis decreased to 52.5 mm). Similarly, the main
specifications of the actuator unit for the rear suspension
applying the load Fn= 2500 Nwas defined (see Appendix
A1). The minimum diameter of the screw in this case was 25
mm, defined by admissible contact pressure developed on
the screw thread surface. Maxon brushed DC RE30 with 60
Wnominal power was chosen to actuate the rear system.
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Amati et. al. 11
Based on the initial data and imposed constraints, the
design parameters of the systems have been defined. The
Appendix A2 shows the values for these parameters. Four
prototypes (two for each axle) then have been built.
Figure 11. Section view of the front actuator assembly. 1-
Power screw, 2- Casing cover, 3- Pinion, 4- Planetary
gearhead, 5- DC motor, 6- Anti rotation element, 7- Integrated
gear and nut, 8- Actuator casing, 9- Compound bearing and 10
– Lower spring seat.
Figure 12. Experimental prototype of the front actuator: (a)
components and (b) its integration in the front suspension.
Figures 11 and 12 show, respectively, a section view of the
electromechanical actuator and its integration in the vehicle
front suspension. The power screw 1consists of a hollow
cylindrical body with trapezoidal threads, fitted and welded
on the shock absorber tube (see Figure 12(b)). The DCelectric motor 5transmits the torque to the pinion 3by means
of a three stage Maxon GP42C planetary gearhead 4. The
pinion 3is mounted on the motor casing cover 2and is
supported radially by a needle bearing to reduce the friction
losses (for industrialized solution low friction bushings can
be used). The pinion (has 32 teeth) is engaged with part 7
which integartes the nut (inner part of the component) and
gear (outer part, has 71 teeth). The rotation of the nut 7is
transformed in axial displacement of the lower spring seat 10.
To decouple the rotation of the gear 7and the lower spring
seat 10, a combined radial and thrust bearing 9is installed
between them. A tooth sliding in a groove of the power screw
works as an anti-rotation. It prevents unwanted rotation of the
system due to reactive torque during the actuation.
Figures 13 and 14 show a section view of the
electromechanical actuator and its integration in the vehicle
rear suspension. The power screw 1is fixed on the interface 9
welded to the longitudinal rail of the chassis close to luggage
compartment. The DC electric motor 11transmits torque to
the pinion gear 8by means of three stage Maxon GP32C
planetary gearhead 10. The pinion 8is installed in the motor
casing cover 4(integrated with upper spring seat) and by
a needle bearing to reduce the friction losses. The pinion
(32 teeth) is engaged with part 3including the nut (inner
part of the component) and the gear (outer part, 71 teeth)
of final stage. The rotation of the nut 7is, in this case,
transformed into translation of the upper spring seat 4. A
combined bearing 5decouples the rotation of the gear 3and
the upper spring seat 4. Two anti-rotation elements 7attached
to the casing 5slide along the groove of the power screw
to prevent unwanted rotation of the system due to reactive
torque.
The prototypes were assembled on front (Figure 12(b))
and rear (Figure 14(b)) axles of an Asegment, 4×4 vehicle.
Five electronic control boards based on Freescale micro-
controllers have been used to control the four corners
individually. Four of them act as slave boards equipped
with an H-bridge and current sensors, to measure current on
each corner. One master board has been connected via CAN
Bus to manage the slave boards. On the front suspensions
ON/OFF switches were installed to limit the travel of
actuators. Instead, on the rear suspensions this function is
realized by means of mechanical end stops. The height of
the vehicle was measured by means of one rotational sensor
per each corner. The required position was set by the driver
manually on the control panel. In this prototype, the vehicle
height could be continuously varied (depending on the need
of the driver) and no specific mission profile was defined. A
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12 Journal Title XX(X)
total stroke of 70 mm was obtained (lifting by 20 mm and
lowering by 50 mm).
Figure 13. Section view of the rear actuator assembly. 1-
Power screw, 2- Casing cover, 3- Integrated gear and nut, 4-
Casing of the actuator (combined with upper spring seat), 5-
Compound bearing, 6- Bump stop, 7- Anti rotation element, 8-
Pinion, 9- Chassis interface , 10- Planetary gearhead and 11-
DC motor.
Figure 14. Experimental prototype of the rear actuator: (a)
components and (b) its integration in the rear suspension.
Tests for performance validation
Broad tests were carried out on the both front and
rear suspensions to assess the system functionality and
performance. Figures 15 and 16 show the required values
of current to lower and lift the vehicle. Mean values
of measured currents show good correspondence withnumerical calculations, when the value of friction coefficient
f= 0.2 is used. However, lower value of friction coefficient
was expected due to lubrication and thread surface coating.
The performance of the front and rear systems are
summarized in Table 2. Figure 15 shows that the steady
currentIssto lower the vehicle for the front actuator is about
9.3Acompared with 6.5 Aof the rear actuator. By converse,
to lift the vehicle the front actuator requires around 11 Aand
the rear one 9.3 A(Figure 16). The average power required
to lower the front suspension is around 117 Wand for the
rear 80 W. To lift the vehicle the front actuator requires 140
Wand the rear 117 W. In average to perform a cycle the total
height adjustment system requires around 500 Wof power.
EC Regulation No 443/2009 suggests procedure to
convert required power of accessories to CO2emissions.15
Assuming an average power of 500 Wrequired to lift and
lower the vehicle, the efficiency of the standard alternator
equal to 67% and considering the time to perform one cycle
of height adjustment equal to 30 s(15sto lift and the
same for to lower) the emission of CO2per cycle can be
calculated. The result is 3.83 gofCO2emitted in one
actuation cycle. For an NEDC cycle with around 11 km
of total distance, assuming one cycle of height adjustment
actuation, then the emission will be 0.35 gCO 2=km. This
number is useful to define the mission profile of height
adjustment system to maximize potential benefits of using
such systems.
To check the irreversibility of the system the road tests
were performed. The vehicle was driven at constant speed
(30km=h) on roads with different pavement conditions
(including speed bumps). During these tests the system
represented total irreversibility, as it was expected.
Figure 15. Current absorption to lower the vehicle measured at
inputs of front and rear actuators.
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Amati et. al. 13
Figure 16. Current absorption to lift the the vehicle measured
at inputs of front and rear actuators.
Table 2. The main specifications of the front and rear systems.
Parameters FRONT REAR
Mass of the system, kg 2.2 2.1
Dimensions LxWxH, mm 190x140x225 152x100x372
Current to lower, A 9.3 6.5
Current to lift, A 11 9.3
Power to lower, W 117 80
Power to lift, W 140 117
Lowering speed, mm=s 2.7 3.7
Lifting speed, mm=s 1.5 3.2
Total power to lower W: 395
Total power to lift W: 520
Conclusions and future work
Height adjustment systems are able to offer not negligible
reductions of fuel consumptions and emissions when applied
to passenger vehicles. The modification of the vehicle
road clearance can introduce benefits up to about 4% of
reduced CO2emitted with respect to the normal production
configuration. In this paper, a literature review regarding
different height adjustment technologies has been largely
described. The focused attention on height adjustment
systems of the electromechanical kind is motivated by the
need of meeting functional and economic aspects (high
reliability, robustness and compact size vs. low costs of
manufacturing). Nevertheless, the literature does not deal
with design and optimization issues. To this end, with
the aim of optimize the main system parameters while
meeting different nature constraints, a design methodology
for electromechanical height adjustment systems has been
presented. Particular attention has been devoted to its
adaptability to different suspension schemes. Based on the
described methodology, prototypes for front and rear axle
suspensions have been manufactured and installed on a
real vehicle. The performed experimental tests highlightedthe low power absorption of such systems, which leads to
the possibility of actuating the four corners together. This
is the one of the main features of the electromechanical
height adjustment systems compared with the hydraulic or
the pneumatic ones, where only one axle at a time can
be supplied due to power limitations of vehicle battery. In
addition, critical issues have also been identified. The anti-
rotation system represents a crucial component. For further
investigations it will be more advantageous if this feature
could be intrinsically present in the system, for example
by eccentrically mounting the screw-nut mechanism with
respect to the spring seat.
Acknowledgements
The authors would like to thank Ing. Michele Ieluzzi and Ing.
Patrizio Turco for providing necessary data during the design
process and Dr. Fabrizio Impinna of FLAG-MS for development of
the control board and giving support in performing the experimental
tests.
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Amati et. al. 15
Appendix
Table A1. The main specifications of the system.
Parameters Symbol Unit FRONT REAR
Initial data
Mass to lift m kg 370 240
Load to lift Fn N 3700 2500
Friction coefficient f 0.2 0.2
Imposed constraints
Irreversibility total % 50 50
Shock absorber tube diameter dtube mm 46.5 NA
Absorbed current Iss A 12 12
Time to cover 50 mm t s 13 13
Table A2. System parameters obtained by design methodology.
Parameters Symbol Unit FRONT REAR
Power screw
Thread angle (Acme) 14.5 14.5
Pitch p mm=rev 8 4
Nominal diameter d mm 53 27.5
Electric motor
Nominal power Pem W 90 60
Nominal torque Tem Nm 73.1e-3 51.6e-3
Nominal speed !n r=min 6500 7630
Winding inductance L H 85e-6 34.5e-6
Winding resistance R Ohm 0.314 0.196
Torque constant kt Nm=A 0.0195 0.0139
Speed constant ke V=rad 0.0195 0.0139
Rotor inertia Jem kgm268.1e-7 33.7e-7
Speed reducer
Gearbox ratio igb 81:1 51:1
Gearbox efficiency gb % 72 72
Gearbox inertia Jgb kgm29.4e-7 0.7e-7
Final gear ratio ig 2.22:1 2.21
Final gear efficiency g % 95 95
Pinion inertia Jp kgm213e-7 20e-7
Gear and nut group inertia Jn kgm2483e-6 480e-6
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