Printed Development of Analytical Process To Reduce Side Load in Strut-Type Suspension
Printed Development of Analytical Process To Reduce Side Load in Strut-Type Suspension
Printed Development of Analytical Process To Reduce Side Load in Strut-Type Suspension
www.springerlink.com/content/1738-494x
DOI 10.1007/s12206-009-1103-z
(Manuscript Received May 7, 2009; Revised August 17, 2009; Accepted September 16, 2009)
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Abstract
Methods have been developed to reduce the side load and friction force acting on the shock absorber inherent in the MacPherson strut
system, popularly used in vehicle suspension systems. Reducing this friction force is one of the most important issues in improving the
ride comfort of a car. The side load of the shock absorber can be reduced by controlling the force line of the coil spring. To reduce the
side load, we designed an S-shaped coil spring. For the design of the side load spring, we also developed an analytical process, which
utilizes finite element analysis and mechanical system analysis. All analysis results for the stiffness, stress, fatigue life, and spring force
line were validated through experiments.
Keywords: MacPherson strut suspension; Ride comfort; Optimal design; Side load; Spring force line
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absorber, as shown on the left side of Fig. 1. The load W act- Point L and point U:
ing from the ground causes a bending moment in the shock ⎡ f ( z − c) f ( z − c) ⎤
[ xl , yl , zl ] = ⎢ x l + a, y l + b, zl ⎥
absorber. During the piston movement of the shock absorber, fz fz
⎣ ⎦
the friction is increased due to this bending moment, which
⎡ f ( z − c) f ( z − c) ⎤
results in increased riding discomfort. One common approach [ xu , yu , zu ] = ⎢ x u + a, y u + b, zu ⎥
to reducing this friction is by using an offset strut, shown on ⎣ f z f z ⎦
the right side of Fig. 1. With this type of strut, the friction and
side load can theoretically be eliminated; however, the actual 2.2 Determination of the ideal spring force line
side load, in practice, does not become zero because (1) an
eccentric load may be created by the deviation between the We used the MSC.ADAMS software, which can solve
load axis and geometric center axis of the coil spring, (2) it is multi-body dynamics problems, to calculate the side load ap-
difficult to keep the required amount of offset because of a plied to a strut-type suspension and find the spring force line
lack of space around the suspension, and (3) the load axis that would minimize the side load.
position is greatly influenced by the boundary condition of the Fig. 3 shows the front suspension kinematic model used to
coil spring find the ideal spring force line for the target vehicle, and de-
scribes the terms used to define a spring force line. The spring
2.1 Calculation of the spring force line force line is defined as the line that connects the spring upper
hard point to the lower hard point. The ideal spring force line
To calculate the spring force line from FE analyses or ex- can be found by moving the spring lower hard point to the
periments, the reaction forces (F) and moments (M) have to be location which minimizes the side load generated at the strut
measured at the location (point A) where the spring lower seat rod top mount during vertical movement of the wheel.
is fixed on the strut tube. This location can be the same with Fig. 4 shows the change in side load at the strut rod top
the intersecting point (point B) of the spring lower seat plane mount for various inclination angles. A vertical wheel travel,
and the strut axis. The spring force line means the line con- in the general driving conditions, between -20 and 20 mm is
necting two points (point L and point U) on the spring upper assumed. The side load can be minimized in this range most
and lower plane at which no reaction moments are exerted.
The two points can be calculated as follows: (Fig. 2)
Measured forces: F = [ f x , f y , f z ]
Measured moments: M = [mx , my , mz ]
Point O where the moments are zero:
⎡ m y mx ⎤
[ a , b, c ] = ⎢ − , , zo ⎥
⎣ z f f z ⎦
The line equation passing point O:
x−a y−b z −c
= =
fx fy fz
(a) Front suspension kinematic model
Z (strut axis)
U ( xu , y u , z u ) Strut rod top mount
C (location for side load measurement)
(spring upper - XY plane)
F (M=0) Spring upper hard point
Z-axis [mm]..
Vertical wheel travel [mm]
(Strut axis)
100
Fig. 4. Side load at strut rod top mount.
50
(Lower seat) 0
-10 0 10 20 30 40
Y-axis [mm]
-50
Rebound
Rebound Unladen
Unladen Laden
Laden Bump
Bump
(a) Conventional shape (b) C-shape (c) S-shape (d) L-shape
Stiffness [N/mm]
Calculation method Error [%]
Experiment FEA
Linear regression 22.86 22.78 0.3
2 Point at both ends 23.39 22.91 2.1
500
450
400
350
Side load [N]
300 optimal_lh
250 optimal_rh
conventional_lh
200 .
conventional_rh
150
100
50
Fig. 10. Experimental setup for measuring spring force line.
0
0 5 10 15 20 25 (Upper seat) 200
Target
Unladen Wheel travel [mm] Laden Experiment
150 FEA
Fig. 9. Comparison of side loads at strut rod top mount.
Z-axis [mm]..
(Strut axis)
100
ventional spring ranged from 375 N to 460 N, while the opti-
mized spring generated maximum side loads of 80 N. The 50
results show that side load was reduced by 80% through de-
sign optimization. (Lower seat) 0
-10 0 10 20 30 40
2.6 Experimental validation for spring force line and stiff- Y-axis [mm]
-50
ness
Fig. 11. Comparison of spring force line between experiment and FEA
To validate the design and analysis process of the spring for in the design weight condition.
reducing the side load, we manufactured the optimized spring
and performed an experiment to validate the calculated spring 6000
Experimental result
force line. We used a load cell, which can measure 3-axes 5000
FEA result
forces and moments, for the experiment. Fig. 10 shows the
4000
experimental setup.
Force [N]
As shown in Fig. 11, the spring force lines, calculated using 3000
ment and analysis for spring force line. Fig. 12 and Table 2 Strut displacement [mm]
represent the comparison results for stiffness. In Fig. 12, the Fig. 12. Stiffness curves for experiment and FEA.
strut displacement means the compressed length of a spring.
Table 2 shows that the stiffness values are very similar to the
design requirement of 23 N/mm, and the difference between 3. Analysis and experiment for stress and fatigue life
the experiment and FEA is 2.1%. The spring force line and stiffness were validated experimen-
Y. I. Ryu et al. / Journal of Mechanical Science and Technology 24 (2010) 351~356 355
tally in the previous section. In this section, we validate the reli- Table 3. Strain and stress results for test and analysis.
ability of an analytical approach to the stress and fatigue life Experiment FEA Error [%]
through a series of experiments. Max. Principal Strain [ / ] 4.76E-03 5.12E-03 7.0
Max. Shear Strain [ / ] 8.42E-03 9.15E-03 8.0
3.1 Finite element model and constraints Max. Principal Stress [MPa] 809 864 6.3
Max. Shear Stress [MPa] 653 709 8.0
Fig. 13 shows the finite element model and its constraints
for a coil spring. The upper seat can only rotate and the lower
seat can only translate along the strut axis. The degrees of
freedom of the other directions are fixed. Therefore, the spring
is deformed only in the direction of the strut axis when a load
is applied at the lower seat. The upper and lower seats are
modeled as rigid bodies, with the contact conditions between
the spring and seats. Fig. 14 shows the input load time history
applied at the lower seat.
Upper seat :
• X,Y,Z trans. DOF - Fixed
• X,Y rot. DOF – Fixed
• Z rot. DOF - Free
Contact Definition :
• Upper seat to spring
Strut Axis
600
Fig. 13. Finite element model and constraints.
400
(Load)
Bump
200
GVW
Rebound 0
0 50 100 150 200 250
Free length Strut displacement [mm]
(Time)
Fig. 14. Input load time history at spring lower seat. Fig. 17. Comparison of stress at maximum principal stress location.
356 Y. I. Ryu et al. / Journal of Mechanical Science and Technology 24 (2010) 351~356
The following shows the setup for the fatigue analysis of a Young-Il Ryu received his B.S. in Me-
coil spring chanical and Automotive Engineering
(1) Analysis Method: Strain-Life approach from Kookmin University in 2002. He
(2) Mean Stress Correction: Smith-Watson-Topper then received his M.S. degree from the
(3) Surface Treatment: Shot Peening Graduate School of Automotive Engi-
(4) Load Condition: GVWÆBumpÆReboundÆGVW neering in Kookmin University in 2004.
As shown in Table 4, the difference in the fatigue life be- He is currently a Ph.D. candidate at the
tween the analysis and experimental result is 4.1 %, which is a graduate school. His research interests include vehicle dynam-
reliable analysis result with regard to the deviation of product ics, multi-body dynamics, and durability.
quality.
Seung-Jin Heo received his M.S. in
4. Conclusion Mechanical Design from Seoul National
We developed a series of analytical processes for the design University in 1981. He then received his
of a coil spring to reduce the side load, and performed the Ph.D. in Automotive Engineering from
following. Technical University of Aachen in 1987.
(1) We constructed a customized front suspension kine- He is currently a professor at Kookmin
matic model, which can calculate an ideal spring force University. His research interests in-
line to minimize the side load. clude vehicle passive & active safety, vehicle control, vehicle
(2) The spring force line design was optimized. dynamics, and durability.
(3) Experiments were carried out to validate the analysis
results for spring force line, stiffness, stress, and fa-
tigue life.