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Design and Modelling Methodologies of an Efficient and Lightweight Carbon-fiber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car

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The design of an electric urban car's chassis for the "Shell Eco Marathon" competition takes into account the usage and the type of the vehicle. The most critical factors of designing the new chassis are: the reduction of the weight, improvement of strength and stiffness and reduction of material and manufacturing cost. Towards this direction, a new design approach for a lightweight carbon-fiber reinforced epoxy (CFRE) monocoque chassis, is proposed, which conforms to structural, ergonomic and aesthetic requirements. For the development of this innovative approach, the parametric design method was chosen, in order for the design to be modified easily. The chassis efficiency, in terms of high strength in low mass, was obtained by following appropriate design steps and rules which conform to the vehicle structural and dynamical constraints and by choosing the composite material CFRE. Additionally, a method that calculates the mechanical properties of the composite material CFRE is presented. Furthermore, a model has been created, which calculates automatically the total loads applied on the vehicle's chassis. Worst case stress scenario was chosen and the model's output was evaluated for the new chassis design.
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Material Science and Engineering with Advanced Research
Design and Modelling Methodologies of an Ecient and Lightweight Carbon-ber
Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car
Evangelos Ch. Tsirogiannis1*, Georgios E. Stavroulakis2 and Sofoklis S. Makridis3
1School of Mechanical Engineering, Naonal Technical University of Athens, Athens, GR15780, Greece
2School of Producon and Management Engineering, Technical University of Crete, Chania, GR73100, Greece
3Department of Environmental and Natural Resources Management, University of Patras, Agrinio, GR30100, Greece
www.verizonaonlinepublishing.com
Mater. Sci. Eng. Adv. Res 2(1). Page | 5
*Corresponding author: Evangelos Ch. Tsirogiannis, Mechanical Engineering, Naonal Technical University of Athens, Athens, GR,
Greece; Tel: +306947089008; E mail: vaggelis.tsirogiannis@gmail.com
Arcle Type: Research, Submission Date: 12 December 2016, Accepted Date: 9 February 2017, Published Date: 20 February 2017.
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of
an Ecient and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res
2(1): 5-12. doi: hps://doi.org/10.24218/msear.2017.21.
Copyright: © 2017 Evangelos Ch Tsirogiannis. This is an open-access arcle distributed under the terms of the Creave Commons
Aribuon License, which permits unrestricted use, distribuon, and reproducon in any medium, provided the original author and
source are credited.
Vol: 2, Issue: 1
Abstract
e design of an electric urban car’s chassis for the “Shell Eco
Marathon” competition takes into account the usage and the
type of the vehicle. e most critical factors of designing the new
chassis are: the reduction of the weight, improvement of strength
and stiness and reduction of material and manufacturing cost.
Towards this direction, a new design approach for a lightweight
carbon-ber reinforced epoxy (CFRE) monocoque chassis,
is proposed, which conforms to structural, ergonomic and
aesthetic requirements. For the development of this innovative
approach, the parametric design method was chosen, in order
for the design to be modied easily. e chassis eciency, in
terms of high strength in low mass, was obtained by following
appropriate design steps and rules which conform to the vehicle
structural and dynamical constraints and by choosing the
composite material CFRE. Additionally, a method that calculates
the mechanical properties of the composite material CFRE
is presented. Furthermore, a model has been created, which
calculates automatically the total loads applied on the vehicles
chassis. Worst case stress scenario was chosen and the model’s
output was evaluated for the new chassis design.
Keywords: Parametric design, Lightweighting, Chassis, Com-
posites, Carbon-ber, CFRE, Monocoque, CAD, CAE, Vehicle
dynamics, FEM, Modelling, Electric car, Racing, COG, Shell eco
marathon, Eciency, Chassis design, Stress scenario.
Introducon
e main goal of this publication is to demonstrate a strategy
plan concerning the designing process and guidelines, the
materials, the worst case stress scenario and the loads, for the
maximization of the car structure eciency, in terms of high
strength and performance in low mass [1]. By applying the new
strategy plan, the chassis of an electric car can have less weight
and become more durable. Historically, the studied prototype of
the electric car has been employing an aluminium space frame
and has already won four trophies in six years, in the Shell Eco
Marathon, being placed among the best cars in this European
competition.
However, using aluminium as structural material, additional
aluminium was required to meet stiness and strength demands.
Furthermore, with the use of space frame as chassis type, there
was not enough space to install mechanical and electrical parts.
Accordingly, it is necessary to create extra housings on the chassis
in order to t in the mechanical and electrical parts. Moreover,
space frame is dicult to be manufactured because it is made out
of many parts that are assembled together.
Consequently, according to the strategy plan, a CFRE monocoque
chassis design, is proposed, that oers great design freedom and
is lighter, stier, stronger, easier to manufacture and more spa-
cious than the previous one. To achieve a high quality design, the
design specications were compromised with the team’s targets,
the ergonomic and safety issues were evaluated, the structural
possibilities and limitations regarding the available materials
were taken into account, the structural engineering constraints
regarding a lightweight, sti, strong and easy to manufacture de-
sign were investigated and the loads that act on the axles were
analyzed and calculated. Simulation and manufacturing proce-
dures were outside the scope of this publication. For the three-di-
mensional design, the ProEngineer Wildre 5 soware was used.
Ergonomics
e driver can be aided in his performance by ensuring that all
controls can be easily reached, he/she has a comfortable seating
position and that visibility over the front of the chassis is sucient
[2]. e variables for a good seating position are the vertical
and horizontal position of the steering wheel, the horizontal
position and angle of the seat with respect to the horizontal, the
horizontal and vertical position of the pedal assembly, the height
and horizontal position of the dashboard and front roll hoop.
Besides being comfortable, the driver must be safe at all times
[3]. is mainly involves that many rules are followed in order to
design a safe car. Some major regulations of Shell Eco Marathon
are the existence of a roll bar that withstands 700 N (applied in all
directions) and extends 5 cm around driver’s helmet, a bulkhead
that secures the driver, a wide and long enough chassis design
to protect the driver’s body and dimensional demands for the
chassis to allow for quick driver egress in case of accidents or re
[4]. ese regulations are oen with respect to a so called 95th
percentile male [5] as shown in Figure 1.
Mater. Sci. Eng. Adv. Res 2(1). Page | 6
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21. doi: hps://doi.org/10.24218/msear.2017.21.
Figure 1: 95th percenle male driver template on our new chassis
New chassis’ design
e boundary dimensions of the new chassis design are
1740x730x740 mm. Working with parameters and relations in
Pro Engineer, the new chassis design is generated parametrically
[6]. It can be updated by changing one or more values of its
parameters. To do so, variables and algorithms are used to
generate a hierarchy of mathematical and geometric relations [7].
Changing the parameters values, some optimal steps are created
for the nal design stage of the new chassis as shown in Figure 2.
Figure 2: Final design stage
Mechanical properes of the new chassis
e CFRE was chosen as construction material for the principal
characteristics strength/light weight, durability, stiness, fatigue,
thermal expansion, energy dampening, corrosion resistance and
production exibility [8,9].
Groups of unidirectional plies cut at various angles from the con-
tinuous sheet of “prepreg” are stacked with a sequence that is de-
scribed by formulae [0/30/0/-15/15/0/-30/0] – 8 plies arranged.
Unidirectional pliesmeans that a large percentage of the fibers
has the same orientation allowing higher specific moduli in the
main fiber direction.
e volume fraction of ber/resin, can be calculated:
Vf = (Wf ⁄ df ) ⁄ [(Wf ⁄ df ) + (Wm dm )] (1)
Vm= (1-Vf ) (2)
where df is the density of the ber, dm is the density of the resin,
Wf is the weight of the ber, and Wm is the weight of the resin.
Assuming that the structure is a simple beam with length L, con-
sisting of ber and resin that deform together and the deforma-
tion is time independent, a method of estimating the stiness of
a unidirectional composite is performed (rule of mixtures).
Ec = Ef Vf + Em (1-Vf ) (3)
where Ef is the elastic modulus of the ber, and Em is the elastic
modulus of the epoxy.
Assuming an anisotropic thin composite lamina with the bers
aligned in the x1 direction, transverse to the x2 direction and
vertically to the x3 direction, Young’s modulus E, shear modulus
G and Poisson ratios v, in all three axes, are required for its
characterization [10].
Ex = Ef Vf + Em (1-Vf ) (4)
Ey = Ef Em ⁄ (Em Vf + Ef (1-Vf ) ) (5)
Ez=Ey (6)
νxy = νf Vf + νm (1-Vf ) (7)
νyz = (νxy Et ) ⁄ Ec (8)
νxz = νxy (9)
Gxy = Gm Gf ⁄ (Gm Vf + Gf (1-Vf ) ) (10)
Gyz = Ey ⁄ (2(1-νyz )) (11)
Gxz = Gxy (12)
Respectively, the longitudinal tensile strength, the transverse
tensile strength and the compression strength on the composite
are listed.
σx = σf Vf + σm (1-Vf ) (13)
σy= σm (1-√(4Vf ⁄ π) (14)
σcompx = Gm ⁄ (1-Vf ) (15)
where σf is the bers stress levels, and σm is the resin stress levels.
For the multi-ply laminates, the tensile modulus, the shear
modulus and the Poisson ratio of a random continuous-ber
composite can be calculated by:
E = (3 ⁄ 8) Ε1 + (5 ⁄ 8) E2 (16)
G = (1 ⁄ 8) E1 + (1 ⁄ 4) E2 (17)
ν = (E-2G) ⁄ 2G (18)
where E1 is the longitudinal modulus, and E2 is the transverse
modulus for a unidirectional lamina.
e Krenchel model is utilized for the approximation of the
strengths of multi-ply laminates. e eciency factor, nθ, is used
in a mixture-rule calculation [11]:
nθ=∑an cos4 θ (19)
σc = nθ σfu Vf + σm (1-Vf ) (20)
Table 1 and Table 2 provides the properties of unidirectional
CFRE and multi-ply laminates.
Table 1: Properes of unidireconal CFRE
VALUE UNIT
Elasc modulus Ex 380.100 Gpa
Elasc modulus Ey 28.269 Gpa
Elasc modulus Ez 28.269 Gpa
Poisson rao νxy 0.336
Poisson rao νyz 0.025
Poisson rao νxz 0.336
Shear modulus Gxy 4.213 Gpa
Shear modulus Gyz 13.790 Gpa
Shear modulus Gxz 4.213 Gpa
Tensile strength σx 2539.400 Mpa
Tensile strength σy 8.251 Mpa
Compressive strength σcompx 4.722 Gpa
Mater. Sci. Eng. Adv. Res 2(1). Page | 7
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21.
Table 2: Properes of mul-ply laminates
PROPERTIES VALUE UNIT
Elasc modulus E 160.206 Gpa
Shear modulus G 54.580 Gpa
Tensile strength σc971.400 Mpa
Poisson rao ν 0.468
Mass and center of gravity denion
Pro Engineer gives that the new chassis mass value is 5.38 kg. us,
the total vehicles weight (with the driver on seat) is 149.64 kg. It
can also provide the center of gravity of the chassis. ereaer,
the dierent centers of gravity (battery light, steering system,
driver+seat, chassis, fuel cell, electric motor) are added up, and
form the main center of gravity [12-16] as shown in Table 3.
Table 3: New vehicle’s center of gravity
COGx (mm) COGy (mm) COGz (mm)
867.900 391.178 419.625
Determine the worst case stress scenario
e loads of a chassis structure are divided into crash, ride,
towing, aerodynamic, cornering, braking and tractive loads.
Crash cases are oen the most difficult and critical to design.
ese are outside the scope of this publication, since the structure
moves out of the elastic regime into deep collapse. e ride loads
is one of the more important criteria by which people judge the
design and construction “quality” of a car. However, in this case,
the track road is not rough, but quite smooth [17]. Towing loads
cannot be neglected, but the vehicle will not need to tow another
vehicle. Aerodynamic loads also stress the vehicle structure.
Nevertheless, these loads are negligible because the forward
velocity is very small (20 km/h – 30 km/h). Consequently, static,
cornering and braking loads were taken into account.
Cornering loads are maximized when the vehicles speed is
at maximum speed and its turning radius is minimized. At
the Shell Eco Marathon’s track, in Rotterdam, there are four
counterclockwise andone clockwise turns with approximately
the same angle (90°). Regarding to the driving strategy the vehicle
runs at high speed (30km/h) on rst turn [18]. erefore, the rst
turn has been investigated while the chosen racing line depends
on the characteristics of the car, the cornering strategies and the
conditions around. In the apex point of the corner, the maximum
speed and stress is reached. So, the apex of the rst turn is the
point where there is the cornering worst case stress scenario.
Braking loads cause larger loads than tractive loads [19]. us,
a real situation needs to be considered when the chassis is
overloaded during braking. Supposing that while the vehicle
moving on the track,with its maximum speed, 30 km per hour,
the preceding vehicle suddenly brakes. erefore, the driver is
forced to brake immediately, to avoid the collision. At this point,
it is needed to nd a realistic “deceleration scenario” for urban
cars, to determine the deceleration value. e “Autonomous
Emergency Braking” (AEB) test of Euro NCAP is chosen.
Randomly, the Fiat’s braking control system is selected to see
how it behaves in braking tests [20]. At speeds between 20 km/h
and 30 km/h, the brakes apply a maximum deceleration of 6 m/
s2. In our case it is also supposed, that the driver’sreexes during
braking, are as good as Fiat’s braking control system. us, the
vehicle will be subjected tobrake with a deceleration equal to 6
m/s2 from 30 km/h to 0 km/h [21-23].
In order to demonstrate the strength of the chassis, it only has
to be shown that it withstands the total load worst case stress
scenario that is the combination of cornering and braking worst
case stress scenario. us, it is needed to study the scenario
where the vehicle is turning in the 1st corner and while is
positioned in the apex with 30km/h, it encounters a stationary
preceding vehicle and decelerates immediately (6 m/s2), to avoid
the accident.
Dynamic axle loads
Presuming that the vehicle sits statically on level ground, the
vertical loads can be calculated [17].
Wf = Mg (c ⁄ L) (27)
Wr = Mg (b ⁄ L) (28)
where M is the vehicle mass, g is the gravity acceleration, b is the
distance from the front axle to the CG, and c is the distance from
the rear axle to the CG.
According to the lateral dynamics, the two front wheels can be
represented by one wheel at a steer angle δ, with a cornering
force equivalent to both wheels. e same assumption is made
for the rear wheels[17].
Fy = Fyf + Fyr = (MV2) ⁄ R (29)
where V is the forward velocity.
During cornering, a dynamic load transfer from the inside to the
outside wheels occurs (the second mechanism for this study is
zero, because the chassis has not springs) [24].
Fzo - Fzi = (2Fy hr ) ⁄ t + (2Κφ φ) ⁄ t (30)
where, hr is the roll center height, Kφ is the roll stiness of the
suspension, and φ is the roll angle of the body.
e torque generated by the rotor, for each wheel brake, as well
as the total braking force is dened [25-27].
Fbp = Fd {L2 / L1 }
Pmc = Fbp ⁄ Amc
Pcal = Pmc
Fcal = Pcal Acal
Fclamp = 2Fcal
F friction = Fclamp μbp
Tr = Ffriction Reff
Tt = Tw = Tr
Ftire = Tt ⁄ Rt
Ftotal = ∑ F(tireLF,RF,LR,RR) (31)
where Fbp is the force output of the brake pedal, Fd is the force
applied to the pedal pad by the driver, L1 is the distance from the
brake pedal arm pivot to the output rod clevis attachment, L2 is
the distance from the brake pedal arm pivot to the brake pedal
pad, Pmc is the hydraulic pressure by the master cylinder, Amc is
the eective area of the master cylinder hydraulic piston, Pcal is
the hydraulic pressure to the calliper, Fcal is the linear mechanical
force by the calliper, Acal is the eective area of the calliper hy-
draulic piston, Fclamp is the clamp force by the calliper, Ffriction is the
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21.
Mater. Sci. Eng. Adv. Res 2(1). Page | 9
frictional force by the brake pads, μbp is the coecient of friction
between the brake pad and the rotor, Tr is the torque generated
by the rotor, Re is the eective radius of the rotor, Tt is the torque
in the tire, Tw is the torque in the wheel, Ftire is the force in the tire,
and Rt is the eective rolling radius of the loaded tire.
During braking, a dynamic load transfer from the rear to the
front axle occurs [25,28].
WT = (aν ⁄ g)×(hc g ⁄ L)×(M×g) (32)
where aν is the deceleration, and hcg is the vertical distance from
the CG to ground.
Chassis load calculator (CLC) model
e values of the forces acting on a vehicle structure change
depending on the characteristics of the structure.
erefore, for academic and research purposes, a model created
which automatically calculates the magnitude and the direction of
the loads acting on each vehicle, by importing the characteristics
of the vehicle. In order to validate the derived model, data such
as track w i dth, wheelbase, center of gravity, mass, et cetera
were used as inputs. e breakthrough in this research work is
the overcoming of the time consuming process to calculate the
correspondi ng forces of dierent design structures of the car
with the use of CLC model.e equations (27) to (32) that were
utilized to implement this model are derived from the theory of
Vehicle Dynamics. According to the chosen stress scenario and
the imported characteristics of the new chassis, the applied loads
are calculated.
First of all, the vertical loads are identied as shown in Table 4
and Table 5.
Table 4: The data used for the calculaon of vercal dynamics
Distance from the front axle to the CG 602.900 mm
Distance from the rear axle to the CG 692.100 mm
Distance from the le side of the chassis to the CG 391.178 mm
Distance from the right side of the chassis to the CG 338.822 mm
Table 5: Vercal dynamics
Stac load on the le front wheel 372.264 N
Stac load on the right front wheel 429.788 N
Stac load on the le rear wheel 324.286 N
Stac load on the right rear wheel 374.396 N
en, the lateral loads are calculated according to the cornering
worst case stress scenario as presented in Table 6 and Table 7.
Table 6: The data used for the calculaon of lateral dynamics
Inial forward velocity 8333 mm/s
Turn radius 10000 mm
Gravity acceleraon 9810 mm/s2
Roll center height 280 mm
Track width 910 mm
Distance of chassis on y axis 730 mm
Track width (-) Distance of chassis on y axis 180 mm
Track width (-) Distance of chassis on y axis (from one
side)
90 mm
Table 7: Lateral dynamics
Cornering force 1062.276 N
Load transfer on the right 653.708 N
Load transfer on the le -653.708 N
During cornering the mass distribution changes, as well as the
cent e r of gravity [15,16,29]. Assuming that there is no mass
transfer in the z axis, since the car has not shock absorbers as
well as the fact that if there is a mass transfer in the x axis, it will
be negligible, then the new COGshown in Table 8.
e m ass distribution changes during the ¼ turn as given in
Table 9.
Table 8: New center of gravity during cornering
COGx (mm) COGy (mm) COGz (mm)
867.900 704.022 419.625
Table 9: Mass distribuon on le and right wheels
Distribuon of mass on the right side 96.441 %
Distribuon of mass on the le side 3.559 %
e braking loads are calculated according to the braking worst
case stress scenario as presented in Table 10, Table 11, Table 12,
Table 13, Table 14, Table 15 and Table 16.
Table 10: Data used for the calculaon of braking loads
Final forward velocity 0 mm/s
Absolute value of velocity change 8333 mm/s
Braking me 10 s
Braking distance 83330 mm
Maximum deceleraon 833.300 mm/s2
Wheelbase 1295 mm
Front area of front axle 265 mm
Tyre coecient of fricon 0.0025
Table 11: Data for the brake system dimensions
Distance from the brake pedal arm pivot to
the output rod clevis aachment L1
Distance from the brake pedal arm pivot to
the brake pedal pad L2
Front Rear
Wheel radius 280 mm 280 mm
Master cylinder diameter 12.7mm 12.7 mm
Distance-pushrod to balance bar pivot 30 mm 40 mm
Th e eecve area of the calliper hydraulic
piston found on one half of the calliper body 800 mm2800 mm2
Pad coecient of fricon 0.35 0.35
Disc diameter 160 mm 160 mm
Pad depth 3 mm 3 mm
Gap between top of pad and disc 1 mm 1 mm
Mater. Sci. Eng. Adv. Res 2(1). Page | 10
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21.
Resulting all the above, a 6.19 m/s2 deceleration was achieved,
wh ose value is greater than the maximum value of the Fiat’s
deceleration (6 m/s2), which was rst set as a goal shown in Table 15.
In the case of both braking and turning loads, the mass distri-
bution changes, as well as the center of gravity of the vehicle
[15,16,29]. Assuming that there is no mass transfer in the z axis,
since the car has not shock absorbers, as well as the fact that if
there is a mass transfer in the y axis, it will be negligible, then the
new COG shown in Table 17.
Table 17: New center of gravity during braking and cornering
COGx COGy COGz
599.260 704.022 419.630
is is the center of mass that the vehicle has during braking and
cornering coexistence. It is observed that aer such a sudden stop
in a ¼ turn the mass distribution changes presented in Table 18.
Table 18: Mass distribuon on front and rear axles
Distribuon of mass on front axle 74.188 %
Distribuon of mass on rear axle 25.812 %
With the new center of mass, the cornering force on each wheel
can be found in Table 19.
Table 19: The cornering force on each wheel with the new center of
gravity
Cornering force (front) 788.086 N
Cornering force (rear) 274.190 N
Cornering force on the le front wheel 28.045 N
Cornering force on the right front wheel 760.041 N
Cornering force on the le rear wheel 9.757 N
Cornering force on the right rear wheel 264.433 N
Summarizing, the loads that act on each semi-axle of the chassis,
in the z axis, are presented in Table 20.
Table 12: Data for the dynamic characteriscs of the vehicle
Cg Height 408.793 mm
Wheelbase 1295 mm
Front wheel rolling radius 280 mm
Rear wheel rolling radius 280 mm
Weight on the front axle 82.309 kg
Weight on the rear axle 72.791 kg
Total weight 155.1 kg
Percentage weight on the front axle 0.531 %
Percentage weight on the rear axle 0.469 %
Table 13: Force applied on the balance bar by the driver
Kgf applied to pedal 10 kgf
Force applied to pedal 98.1 N
Pedal rao 4:1
Force on balance bar 392.4 N
Table 14: Braking force calculaon
Front Rear
Balance bar proporon 0.571 0.429
Force on M Cyl piston 224.229 N 168.171 N
Master/cylinder area 126.613 mm2126.613
mm2
Line pressure generated by the mas-
ter cylinder
1.771 N/
mm2
1.328 N/
mm2
Line hydraulic pressure transmied
to the calliper
1.771 N/
mm2
1.328 N/
mm2
The one sided linear mechanical
force generated by the calliper
1416.785 N 1062.588 N
Clapping force on disc generated by
the calliper
2833.569 N 2125.177 N
The friconal force generated by the
brake pads opposing the rotaon of
the rotor
991.749 N 743.812 N
Fx=Fsin45 701.273 N 525.955 N
Fy=Fcos45 701.273 N 525.955 N
Disc eecve radius 77.5 mm 77.5 mm
Disc torque, the torque generated by
the rotor (both pads 1 wheel)
76860.566
Nmm
57645.425
Nmm
The torque found on the re =
torque wheel = torque by the rotor
76860.566
Nmm
57645.425
Nmm
The force reacted between the re
and the ground (assuming fricon
exists to support the force
274.502 N 205.877 N
Table 15: Deceleraon and stopping distance
Total force (4 wheels) 960.757 N
Deceleraon a 6194.436 mm/s2
Stopping distance 5604.940 mm
Table 16: Load transfer from braking
Front Rear
Weight transfer 31.735 kg -31.735 kg
Axle load under braking 113.493 kg 39.487 kg
Dynamic axle load 1113.371 N 387.363 N
Load transfer from braking on the le
side
11.079 N -11.079 N
Load transfer from braking on the right
side
300.240 N -300.240 N
Mater. Sci. Eng. Adv. Res 2(1). Page | 11
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21.
Table 20: Combinaon load case
Stac load on the le front wheel 372.264 N
Stac load on the right front wheel 429.788 N
Stac load on the le rear wheel 324.286 N
Stac load on the right rear wheel 374.396 N
Load transfer from cornering on the right front
wheel 484.976 N
Load transfer from cornering on the right rear
wheel 168.732 N
Load transfer from cornering on the le front wheel -484.976 N
Load transfer from cornering on the le rear wheel -168.732 N
Load transfer from braking on the le front wheel 11.079 N
Load transfer from braking on the right front wheel 300.240 N
Load transfer from braking on the le rear wheel -11.079 N
Load transfer from braking on the right rear wheel -300.240 N
E verytotal force applied to each semi-axle, in the upward
direction on the z axis and is calculated by the sum of the above
loads shown in Table 21.
Table 21: Total dynamic loads on each wheel (z axis)
Total dynamic load on the le front wheel -101.633 N
Total dynamic load on the le front wheel 1215.004 N
Total dynamic load on the le front wheel 144.475 N
Total dynamic load on the le front wheel 242.888 N
e cornering forces Fy are transferred from the contact patch
to the center of the axle. e equivalentsystem will consist ohe
cornering forces (Fy) plus the moments (Mx) that are created
from the cornering forces. ese moments are the result of the
cornering forces multiplied by the vertical distance, which is
z=280mm as presented in Table 22.
Table 22: Cornering forces Fy and moments Mx from contact patch to
the center of axle
Fy on the le front wheel 28.045 N
Fy on the right front wheel 760.041 N
Fy on the le rear wheel 9.757 N
Fy on the right rear wheel 264.433 N
Mx on the le front axle 7852.662 Nmm
Mx on the right front axle 212811.430 Nmm
Mx on the le rear axle 2732.090 Nmm
Mx on the right rear axle 74041.136 Nmm
ebrakingforceneeds to be analyzed in x and z axis (Fx, Fz) as
shown in Table 23.
Table 23: Braking forces
Fx Fz
Braking force on the le front axle 701.273 N 701.273 N
Braking force on the right front axle 701.273 N 701.273 N
Braking force on the le rear axle 525.954 N 525.954 N
Braking force on the right rear axle 525.954 N 525.954 N
e vertical distance of Fz from the end of the axle is calculated
as well as the vertical distance of Fx from the center of the axle
presented in Table 24.
Table 24: Vercal distances
x=(cos45)*0.0075 + 0.0075 12.803 mm
y=(sin45)*0.0075 5.303 mm
Fx, Fz are transferred to the axle. e equivalentsystem willconsist
ohe braking forces (Fx, Fz) plus the moments (My1, My2) that are
created from the braking forces as shown in Table 25.
Table 25: Braking forces Fx, Fz and moments My1, My2 from disc
eecve radius to the axle
Braking forces Fx Fz
Le front axle 701.273 N 701.273 N
Right front axle 701.273 N 701.273 N
Le rear axle 525.954 N 525.954 N
Right rear axle 525.954 N 525.954 N
Moments from braking forces My1 My2
Le front axle 3719.060
Nmm 8978.604 Nmm
Right front axle 3719.060
Nmm 8978.604 Nmm
Le rear axle 2789.295
Nmm 6733.953 Nmm
Right rear axle 2789.295
Nmm 6733.953 Nmm
Conclusion
e new chassis is extremely light, only 5.38 kg and consequently
less energy is consumed to move it, comparing to the previous
one that weights 10.85 kg. Τhis energy decrease is signicantly
high taking into account that the previous one was the lightest
chassis of the competition. Furthermore, the ergonomics and
the aesthetic acceptance of the new chassis is better than the
previous one.
Consequently, the breakthrough in this research work was not
only the achievement of the lightest chassis in the Shell Eco
Marathon competition that combines ergonomics, aesthetic and
strength demands but also the overcoming of the time consuming
process to calculate the corresponding forces of dierent design
structures of the car with the creation and use of the CLC model.
In the future, a FEM model will be developed and used in order
to demonstrate the resistance of the new chassis design under the
aforementioned extreme stress scenario.
Acknowledgements
Open access fees were covered by the Municipality of Agrinio,
Western Greece. Authors are grateful to the Mayor of Agrinio,
Mr. George Papanastasiou.
Citaon: Evangelos Ch. Tsirogiannis, Georgios E. Stavroulakis and Sofoklis S. Makridis (2017) Design and Modelling Methodologies of an Ecient
and Lightweight Carbon-ber Reinforced Epoxy Monocoque Chassis, Suitable for an Electric Car. Mater. Sci. Eng. Adv. Res 2(1): 5-12. doi: hps://
doi.org/10.24218/msear.2017.21.
Mater. Sci. Eng. Adv. Res 2(1). Page | 12
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