TABLE OF CONTENTS
Acknowledgement………………………………………...........................................................................…i
Introduction……………………………........................................................................……………...…...…ii
chapter 1. OVERVIEW OF SUSPENSION SYSTEM................................................................................1
1.1 Functions of the Suspension System....................................................................................................1
1.2 Requirements for the Suspension System............................................................................................1
1.3 Classification of Suspension Systems...................................................................................................2
1.3.1 Classification by Guidance Components............................................................................................2
1.3.2 Classification by Elastic Elements......................................................................................................3
1.4 Structure of the Suspension System.....................................................................................................3
1.4.1 Elastic Element...................................................................................................................................4
1.4.2 Guidance Element..............................................................................................................................5
1.4.3 Damping Element...............................................................................................................................5
1.5 Analysis of Suspension Systems in Electric Vehicles versus Conventional Vehicles ...........................6
1.5.1 Different Weight Distribution...............................................................................................................6
1.5.2 Absence of Internal Combustion Engine.............................................................................................7
1.5.3 Optimization for Energy Efficiency.......................................................................................................7
1.5.4 Ride Comfort and Quietness...............................................................................................................8
chapter 2. DESIGN OPTION SELECTION.................................................................................................9
2.1 Suspension System on VinFast VF8......................................................................................................9
2.2 Selection and Analysis of Suspension Design Option..........................................................................10
2.2.1 Front Suspension...............................................................................................................................10
2.2.2 Rear Suspension................................................................................................................................11
chapter 3. FRONT SUSPENSION DESIGN..............................................................................................15
3.1 Basic Data for Calculation.....................................................................................................................15
3.2 Determination of Basic Parameters of the Front Suspension System..................................................15
3.2.1 Determining Spring Stiffness and Required Static Deflection............................................................16
3.2.2 Checking the Dynamic Travel of the Wheel.......................................................................................16
3.3 Kinematics of the MacPherson Suspension.........................................................................................17
3.3.1 Determining the Length of the A-Arm and Joint Positions.................................................................17
3.3.2 Graphical Method for Constructing the Kinematic Relationship of the Suspension System ............19
3.3.3 Geometric Relationship of the MacPherson Suspension System.....................................................20
3.4 Kinetics of MacPherson suspension system........................................................................................22
3.5 Determining Spring Stiffness, Deflection, and Damper Stroke............................................................23
3.5.1 Spring Stiffness and Working Travel.................................................................................................24
3.5.2 Damper Stiffness and Working Travel..............................................................................................24
3.6 Determining Reaction Forces and Loads on the Front Suspension....................................................25
3.6.1 Dynamic Load Only...........................................................................................................................25
3.6.2 Maximum Lateral Force Only............................................................................................................28
3.6.3 Maximum Traction or Braking Force.................................................................................................30
3.7 Selection and Strength Verification of Main Components....................................................................32
3.7.1 Lower A-Arm (Control Arm)...............................................................................................................32
3.7.2 Ball Joint Strength Calculation..........................................................................................................38
3.8 Shock Absorber Calculation.................................................................................................................40
3.8.1 Calculating Shock Absorber Damping Force.....................................................................................40
3.8.2 Determining the Damping Coefficient of the Shock Absorber............................................................41
3.8.3 Determining the Dimensions of the Front Shock Absorber................................................................42
3.8.4 Determining the Orifice Size of the Shock Absorber Valve................................................................43
3.8.5 Determination of damper spring dimensions.....................................................................................45
3.9 Spring Calculation for Front Suspension System.................................................................................48
3.9.1 Force Acting on Spring......................................................................................................................48
3.9.2 Spring Design Procedure..................................................................................................................49
chapter 4. REAR SUSPENSION DESIGN................................................................................................52
4.1 Determine the basic parameters of the rear suspension system.........................................................52
4.1.1 Determine Spring Stiffness................................................................................................................52
4.1.2 Determine the static wheel travel (static deflection)...........................................................................52
4.1.3 Determine the dynamic wheel travel (dynamic deflection)................................................................52
4.1.4 Check dynamic wheel travel..............................................................................................................53
4.1.5 Determine the average damping coefficient of the damper...............................................................53
4.1.6 Design approach for the rear suspension system.............................................................................53
4.2 Multilink suspension kinematics (2 lateral arm only).............................................................................54
4.3 Multilink suspension dynamic...............................................................................................................56
4.3.1 Maximum traction and braking force conditions................................................................................56
4.3.2 Maximum lateral force condition........................................................................................................56
4.3.3 Dynamic load case.............................................................................................................................57
4.4 Spring Calculation.................................................................................................................................57
4.5 Damper Design......................................................................................................................................57
4.5.1 Damper Selection...............................................................................................................................57
4.5.2 Characteristics of the damper.............................................................................................................58
4.5.3 Determine the external dimensions of the shock absorber.................................................................59
4.5.4 Determine the dimensions of the valves.............................................................................................60
4.5.5 Determine the dimensions of the rebound valve.................................................................................62
4.5.6 Determine the dimensions of the compression valve..........................................................................65
4.6 Check the strength condition..................................................................................................................66
4.6.1 Piston rod strength check....................................................................................................................67
4.6.2 Lower control arm (I-shaped arm).......................................................................................................67
chapter 5. STRUCTURE SIMULATION OF THE SUSPENSION SYSTEM ..............................................73
5.1 Overview of Simulation Methods...........................................................................................................73
5.1.1 Structure Simulation Methods............................................................................................................73
5.1.2 Types of Structural Durability Analysis Problems...............................................................................75
5.1.3 Detail Steps in general structural analysis.........................................................................................77
5.2 Structural simulation of Front Suspension System...............................................................................77
5.2.1 Lower Control Arm (LCA)...................................................................................................................78
5.2.2 Vertical C-Arm....................................................................................................................................84
5.3 Structural simulation of Rear Suspension System................................................................................89
5.3.1 Lower Horizontal Arm (LHA)...............................................................................................................90
chapter 6. EXPERIMENTS OF THE SUSPENSION SYSTEM.................................................................. 96
6.1 Overview of Vehicle Experiement..........................................................................................................96
6.1.1 Objectives of the Experiment..............................................................................................................96
6.1.2 Requirements for Measuring Equipment ...........................................................................................96
6.2 Overview of the Kistler RoaDyn Test System........................................................................................96
6.2.1 Structure of the Kistler RoaDyn Test System.....................................................................................96
6.2.2 Experimental Procedure and Installation of the Kistler RoaDyn........................................................98
6.3 Overview of the Strain Gauge Sensor.................................................................................................103
6.3.1 General Overview of the Strain Gauge Sensor................................................................................103
6.3.2 Strain Gauge Construction...............................................................................................................104
6.3.3 Signal Acquisition and Data Processing Unit ..................................................................................107
6.3.4 Measurement Diagram.....................................................................................................................107
6.3.5 Experimental Procedure...................................................................................................................107
6.3.6 Sensor Installation Procedure..........................................................................................................108
6.3.7 Bridge Circuit Configuration.............................................................................................................108
6.4 Experimental Scenarios......................................................................................................................109
6.4.1 Scenario 1 - ISO Wave Road Test...................................................................................................110
6.4.2 Scenario 2 – Maximum Braking Test................................................................................................111
6.5 Experimental Results and Analyzes....................................................................................................112
6.5.1 ISO Wave Road Test........................................................................................................................112
6.5.2 Maximum Braking Test.....................................................................................................................113
6.6 Extended Recommendation for Experimental Testing Using Kistler RoaDyn on VinFast VF8 ...........114
chapter 7. MAINTENANCE AND REPAIR OF THE SUSPENSION SYSTEM .......................................116
7.1 Front Suspension System...................................................................................................................116
7.1.1 Common Failures in the Suspension System..................................................................................116
7.1.2 Methods for Detecting Suspension Failures....................................................................................117
7.1.3 Maintaining Suspension Stability.....................................................................................................118
7.2 Rear suspension system....................................................................................................................119
7.2.1 Visual Inspection.............................................................................................................................119
7.2.2 Disassembly of the Suspension System.........................................................................................119
7.2.3 Cleaning and Component Inspection..............................................................................................120
CONCLUSION.........................................................................................................................................121
REFERENCES.........................................................................................................................................123
Acknowledgement
Under the guidance of Assoc. Prof: …………….. and the lecturers of the Automotive and Special Vehicle Research Group of the Faculty of Automotive Engineering, School of Mechanical Engineering, as well as the assistance from fellow students, we have been able to complete the assigned research contents of this project.
We would like to express my sincere gratitude to our academic supervisors, to all the lecturers and staff members of the Automotive and Special Vehicle Research Group of the Faculty of Automotive Engineering, to our cooperation companies Viettel and VNPETRO, and to our peers in the project support team and members of the BK-AUTO Club.
Our special thanks go to: ……………., who enthusiastically guided, supported, and created favorable conditions throughout the project, helping us achieve meaningful results.
Due to time constraints and the wide scope of research across various fields, some shortcomings are inevitable. We sincerely welcome any comments and feedback from lecturers and fellow researchers to further improve the project in future studies.
Thank you very much!
Introduction
In this thesis, we carried out the design and calculation of a suspension system for a 5-seater electric vehicle, using VinFast VF8 as the reference model. Subsequently, the suspension system was simulated based on the input parameters obtained from the design results, and the model was validated through experimental tests measuring force, moment, and deformation at critical components within the suspension system.
The thesis consists of seven main chapters:
- Chapter 1: Overview of Suspension system
- Chapter 2: Design Option Selection
- Chapter 3: Front Suspension Design
- Chapter 4: Rear Suspension Design
- Chapter 5: Structure Simulation of The Suspension
- Chapter 6: Experiments of The Suspension system
- Chapter 7: Maintenance and Repair of The Suspension system
The thesis includes 8 technical drawings:
- General Layout
- Front Suspension System
- Front Shock Absorber
- Rear Suspension System
- Rear shock absorber drawing
- Drawing of typical components
- Simulation results drawing
- Experiment Results
Student
(Signature and Full name)
1./ ………………
2./ ………………
chapter 1. OVERVIEW OF SUSPENSION SYSTEM
1.1 Functions of the Suspension System
The suspension system has the following functions:
- Provides a flexible connection between the wheels and the vehicle body, reducing vertical dynamic loads acting on the body and ensuring smooth wheel rolling on the road surface,
- Transmits forces from the wheels to the body and vice versa, ensuring appropriate movement of the wheels relative to the body,
- Rapidly dampens vibrations from the road surface impacting the vehicle body.
1.2 Requirements for the Suspension System
The suspension system must meet the following requirements:
- Ensure appropriate natural vibration frequency for the suspended mass,
- Provide suitable dynamic deflection to prevent impact with rubber bump stops,
- Possess appropriate damping capability,
- Prevent the vehicle body from tilting during cornering or braking,
1.4 Structure of the Suspension System
The suspension system consists of three main components: elastic element, damping element, and guiding element:
- Elastic element: Converts vibrations from the road surface into oscillations with frequencies suitable for human comfort and cargo safety.
- Damping element: Dampens relative oscillations between the chassis and wheels generated during vehicle operation.
- Guiding element: Transmits longitudinal and lateral forces, as well as torques, between the wheels and the vehicle body.
1.4.1 Elastic Element
The metallic elastic element consists of one or more metal components, commonly using leaf springs, coil springs, and torsion bars. Non-metallic elastic elements typically include rubber, pneumatic, or hydraulic components.
- Leaf Spring: A set of steel plates of varying sizes curved with different radii and bound together into a "leaf spring pack." The plates are clamped tightly and restrained longitudinally by a center bolt and laterally by clamps. The top plate has a greater radius than the lower ones, creating residual stress when assembled. This residual stress counteracts the operational stress under load, enhancing the durability of the leaves. The leaf spring pack is mounted to the vehicle frame via brackets. One end is fixed, while the other is attached to a movable bracket to allow deflection during operation. Besides acting as an elastic component, the leaf spring also contributes to damping by friction between the plates, converting mechanical energy into heat. Additionally, it can transmit both longitudinal and lateral forces between the wheels and the chassis, thus functioning as a guiding component.
- Coil Spring: Mainly used in passenger cars. It transmits only vertical forces, so suspension systems with coil springs require additional guiding and damping components. Coil springs come in cylindrical, conical, or barrel shapes. Their ease of manufacture makes them widely used. Conical and barrel springs are applied where space is limited, allowing the spring to compress flatly.
- Pneumatic (Air Spring): This elastic element comprises a compressed air chamber made of reinforced rubber, typically in two configurations: bellows and convoluted types. A key advantage is the ability to vary stiffness by adjusting internal air pressure. This is achieved via an air supply system and automatic pressure adjustment equipment, which responds to load changes detected by air-based or sensor-based mechanisms.
1.4.2 Guidance Element
In trucks and buses, the guiding component is typically the leaf spring, which also serves as the elastic element. In contrast, passenger cars use control arms or linkage systems as guiding components to better satisfy handling, stability, and ride comfort requirements.
1.5 Analysis of Suspension Systems in Electric Vehicles versus Conventional Vehicles
1.5.1 Different Weight Distribution
In modern electric vehicles (EVs), the battery pack is typically heavy and mounted beneath the vehicle floor, resulting in an even weight distribution between the front and rear axles. This setup not only lowers the vehicle's center of gravity but also enhances cornering stability and reduces the risk of rollover, particularly during hard braking or sudden maneuvers.
Since the battery mass is located low and spreads along the floor pan, the suspension system must be optimized to accommodate this significant load while maintaining ride comfort. This requires springs with appropriate stiffness and specially designed dampers to ensure both stability and high load-carrying capacity.
1.5.3 Optimization for Energy Efficiency
The suspension systems in electric vehicles are often engineered to minimize friction and reduce energy losses caused by unnecessary movement of joints or mechanical components. This is especially critical as EVs aim to maximize range per charge.
Many premium EVs today adopt air suspension or electronic suspension systems, which can vary stiffness and ride height based on road conditions and load. This not only enhances adaptability but also improves aerodynamic efficiency, contributing to better energy savings.
1.5.4 Ride Comfort and Quietness
Without engine and transmission noise, the suspension system plays a larger role in isolating road, wind, and chassis noise. Advanced noise-insulation materials, suspension mounts, and high-quality dampers are used to enhance cabin quietness.
chapter 2. DESIGN OPTION SELECTION
2.1 Suspension System on VinFast VF8
The reference vehicle selected is the VinFast VF8, which features a suspension structure that uses the Smart Axle system (a modified version of the MacPherson suspension) for the front suspension (Figure 2.1) and the Control Blade system (a variation of the multi-link suspension) for the rear suspension (Figure 2.2).
The suspension system of the VinFast VF8 has the following characteristics:
* Elastic Component: Coil Spring Type
- Advantages:
+ High durability and excellent shock absorption capability.
+ Compact design with reasonable cost.
+ High stability, easy stiffness adjustment to suit different vehicle types.
- Disadvantages: Poor lateral load resistance; only transmits vertical loads.
* Damping Component: Twin-Tube Shock Absorber
- Advantages:
+ Reasonable cost and long service life.
+ Adjustable damping valves to suit various vehicle types and operating conditions.
+ Provides smooth and comfortable ride quality.
- Disadvantages:
+ Poor heat dissipation and low response sensitivity.
+ Reduced performance under high-load operations.
* Guidance Component
- Front Suspension: “Smart Axle” with integrated guiding components from both MacPherson and double wishbone systems:
+ Advantages: Better lateral force resistance than classic MacPherson, occupies less space than traditional double wishbone.
+ Disadvantages: Requires more space than standard MacPherson, less stable than double wishbone.
2.2 Selection and Analysis of Suspension Design Option
The front suspension design option is chosen similarly to the reference vehicle. For the rear suspension, due to time constraints and the project scope, a torsion beam suspension is selected, which differs from the reference model.
2.2.1 Front Suspension
- Elastic Element: A coil spring is selected, suitable for independent suspension systems in passenger cars, offering high durability and effective shock absorption. Additionally, it has a compact design, is easy to manufacture, cost-effective, and highly stable.
- Damping Element: A monotube shock absorber is chosen, appropriate for passenger cars requiring enhanced performance and heat dissipation, especially under normal and moderately demanding driving conditions. This type also offers consistent damping and improved responsiveness, particularly under the heavy loads from the battery and other vehicle systems.
* Advantages
- Maintains the simplicity and cost-effectiveness of the MacPherson suspension.
- Improves wheel alignment stability compared to a conventional MacPherson system.
- Reduces lateral wheel movement, enhances road grip and safety.
- Operational performance can be optimized depending on specific design considerations.
* Disadvantages
- Potentially more complex than the standard MacPherson system, leading to higher maintenance costs.
- Heavier than the MacPherson suspension, affecting fuel efficiency.
- May not match the double wishbone system in terms of stability and off-road performance.
2.2.2 Rear Suspension
- Elastic Element: The coil spring type is consistently selected, similar to the elastic component used in the front suspension system.
- Damping Element: A twin-tube shock absorber is chosen, appropriate for passenger cars requiring cost efficiency and durability, especially under normal driving conditions. This type also offers smoothness and comfort, particularly under the heavy loads from the battery and other vehicle systems.
- Guiding Component: two lateral links and one trailing arm-to connect each rear wheel to the vehicle chassis. This configuration allows for independent control of wheel movement in multiple axes, resulting in enhanced ride quality, precise handling, and improved road contact.
Specification of reference vehicle: VF8 ECO using SDI battery Table 1
* Tire Dimension Specifications:
- Front/Rear Tire: 225/55R19
- Design radius r0:
=> r0 = 365,1 (mm).
- Since the tire is of the low-pressure type, select λ = 0,93, thus the average effective rolling radius rb:
=> rb = 339,54 (mm).
chapter 3. FRONT SUSPENSION DESIGN
3.1 Basic Data for Calculation
The front suspension system is of the MacPherson type.
* Fundamental parameters for calculation:
- Front axle track width: B = 1700 (mm)
- Wheel radius: Tire specification 225/55R19. rbx = 365,1 (mm)
- Kingpin inclination angle: β = 15o
- Variation in kingpin inclination: Δβ = 3o
- Maximum mounting height of suspension strut: Htmax = 800 (mm)
- Camber angle: α = 0
- Caster angle (positive): γ0 = 5o
- Radius of wheel rotation around kingpin axis: r0 = -15 (mm)
- Minimum ground clearance when fully loaded: Hmin =100 mm.
- Static deflection: ft = 213 (mm)
- Dynamic deflection: fd = 150 (mm)
- Suspension deflection in unloaded condition: f01 = 180 (mm)
- Length of upright strut: Kr = 320 (mm)
3.2 Determination of Basic Parameters of the Front Suspension System
To evaluate the ride comfort of a vehicle during motion, parameters such as vibration frequency, vibration acceleration, and vibration velocity are commonly used. In suspension design calculations, vibration frequency is used as the basis for selecting the fundamental parameters of the elastic component. For passenger vehicles, the appropriate vibration frequency is typically in the range of 60–90 cycles per minute.
3.2.1 Determining Spring Stiffness and Required Static Deflection
The stiffness of the spring is calculated based on the condition that the vibration frequency must fall within the range of 60–90 cycles per minute.
Where: ft is static deflection (m), g is gravitational acceleration(m/s2).
Select vibration frequency n = 65 cycles/minute. Then, the static deflection of the elastic component is: ft = 21,3 (cm).
3.2.2. Checking the Dynamic Travel of the Wheel
The dynamic travel of the wheel is checked based on the condition that ensures the minimum ground clearance Hmin
=> H0 > 463 (mm)
Where :
M01: Load on each wheel in unloaded condition;
Mt1: Load on each wheel in fully loaded condition.
3.3. Kinematics of the MacPherson Suspension
3.3.1. Determining the Length of the A-Arm and Joint Positions
The graphical method is used to determine the length of the A-arm and the positions of the joints.
The specific steps are as follows:
- Draw a horizontal line representing the road surface plane: dd
- Draw the transverse symmetry axis of the vehicle A0m: A0m is perpendicular to dd.
- On A0-dd take A0B0 = 830 mm
- B0 is the contact point between the tire and the road surface
- At B0, draw a vertical line B0z perpendicular to dd.
- On line A0B0, extend outward and mark a segment A0B0 take B0C0= |r0| = 15 mm.
- At C0 draw C0n forming an angle with the vertical direction β = 15o.
- On B0z take B0B = rbx = 365.1 (mm).
- At B, draw a line perpendicular to B0z intersecting C0n at C2. C2 is the rigid connection point between the wheel upright and the steering knuckle.
Similarly, the outer joint position of the control arm in the fully loaded state is determined as follows: when the suspension reaches its maximum deformation, assuming the vehicle body remains stationary, the wheel moves upward to point B1.
If we assume that the track width at this state remains unchanged compared to the unloaded condition, then:
Then B0B1 = ft0 + ft – fd = 150 + 213 – 180 = 183mm.
- From B1 draw B1q // dd. On B1q take B1D1 = B0C0= |ro| = 15 mm.
- From D1 draw D1n’ forming an angle with the vertical plane β’ = β0 + ∆β = 18o, intersecting Cn at O2.
- Connect D1O2, D1O2 represents the centerline of the kingpin axis when the suspension is fully compressed. During wheel movement, the segment C0C1 remains unchanged. Therefore, on D1O2 t take D1D2 = C0C1. D2 is the location of the outer ball joint of the control arm under maximum suspension compression.
- Hence, C1 and D2 lie on an arc centered at the inner pivot joint of the lower control arm.
- Draw the perpendicular bisector kk’ of segment C1D2
- If line O1C1 is extended and a perpendicular is drawn from line O2C0, they intersect at point P (the instantaneous center of rotation of the wheel).
- Connect PB0 and extend it to intersect A0m at S (S is the instantaneous center of rotation for the axle as well as the vehicle body in the horizontal axle plane).
- Measuring the distance O1C1 gives the length of the ‘A’ of the suspension:
Ld = O1C1 = 300.69 (mm)
3.3.2. Graphical Method for Constructing the Kinematic Relationship of the Suspension System
When the suspension deforms, both the kingpin inclination angle and the track width (distance between tire contact patches) change. The contact points of the tire with the road surface are labeled: 0, 1, 2, 3, 4, 5
The corresponding kingpin inclination angles are: δ0, δ1, δ2, δ3, δ4, δ5
In this graph, we obtain the displacements ∆B (horizontal direction), ∆δ change in kingpin inclination angle) as functions of the wheel’s vertical movement ∆B = f(S), ∆δ = f(S). From these, the suspension kinematic diagram is constructed.
Compared to similar vehicles, the values of ∆B and ∆δ remain within allowable limits (based on Figure 3.3 of the textbook Automotive Chassis Structures).
3.4 Kinetics of MacPherson suspension system
3.4.1. Dynamic Load Only
In the diagram, only the vertical force Z is present (X, Y are absent).
Where:
G11: Load applied on the front axle (sprung mass).
kd: Dynamic load coefficient, kd = 1,8 – 2,5 for passenger cars on good roads, select kd = 1,8
We have: Z =13185,9 N
3.4.3. Maximum Lateral Force
In the diagram, forces Z and Y are present (X is absent).
Where:
B: Track width, B = 1,700 (m)
Gbx: Unsprung mass (includes tire, rim, and brake unit), Gbx = 20.10 = 200 (N).
hg: Height of the vehicle’s center of gravity, hg = 0,180 (m)
φy: Lateral adhesion coefficient, taken as 1
Forces are calculated as follows: Z = 8456,27 N
3.5 Determining Spring Stiffness, Deflection, and Damper Stroke
Elastic components may include coil springs, conical springs, or torsion bars. This section focuses only on calculating force and selecting the layout of coil springs.
Common angular placements include longitudinal and lateral tilt angles, depending on the spatial constraints of the vehicle.
3.5.1 Spring Stiffness and Working Travel
Where:
f = ft + fd: total working travel of the wheel,
Clx: Stiffness of the elastic element,
flx: Working stroke of the spring
3.5.2. Damper Stiffness and Working Travel
The damper arrangement typically found in vehicles is shown below: the damper axis does not coincide with the kingpin axis—common in vehicles where: ro (steering axis radius) is negative and the kingpin inclination angle δ is large.
Where:
Kgc: Stiffness of the damper element,
fgc: Working stroke of the damper.
3.6 Determining Reaction Forces and Loads on the Front Suspension
3.6.1 Dynamic Load Only
- Only Force Z is present; forces X and Y are absent.
- With Z = 13185,9 N
- Then, calculate the forces:
Z1Z= 769,17 N
Z2Z= 12416,73N
- And moment MZ:
MZ = 191,04 N.m
With r0 = 15 mm is steering axis radius; = 15o is kingpin inclination angle
Load acting on Vertical C-Arm:
Fy1 = 3936,94 N
Fz1 = 769,17 N
Fy2 = 7349,7 N
Fz2 = 12476,13N
Flx = 15572,30N
3.6.2 Maximum Lateral Force Only
- There is no lateral force Y, only vertical force Z and longitudinal force X are present
- The effects of force Z and corresponding reactions are similar to the previous analysis.
- The lateral force Y acting on the fictitious upright AB generates reaction forces AY, BY:
YYB = 5442,27 N
YYA = 2814 N
3.7 Selection and Strength Verification of Main Components
3.7.1 Lower A-Arm (Control Arm)
The lower A-arm has an A-shaped structure and is mounted to the vehicle body via two pivoted joints. The outer end is connected to the C-arm through a pin joint. Using two inner hinged joints increases the stiffness of the suspension system.
The main stress states include tension, compression, and bending. The cross-section of the lower A-arm is referenced, and in the strength verification, it is assumed that one side of the A-arm bears the entire load. The calculation proceeds as follows:
3.7.7.1 Case of Dynamic Load Only
Fy = 7671,03 N
Fz =6,9.10-4 N
The lower control arm is subjected to axial tension and longitudinal bending.
- Fz acts as a shear force and causes longitudinal bending in the ZOY plane.
Substitute the values: tmax = 1,15.10-6 (MPa)
- With material being aluminum AlZnMgCu1,5F50: sb = 600 MPa
[t] = 200 (MPa)
Then: tmax< [t]
With n = 1,5: safety factor.
=> Shear strength condition is satisfied.
The component Fz also causes maximum bending moment at the joint connecting the control arm to the vehicle frame. Since the joint is a pivot, the moment at the center is zero. Verification is done at a nearby section (cross-section 1–1).
3.7.1.3 Maximum Traction or Braking Force
Fx = Cx = 3516,24 N
Fy = 1495,64 N
Fz = -2,16.10-3 N
Fz acts as a shear force and induces a longitudinal bending moment in the (zoy) plane:
+ Shear force: Qy = Fz = -2,16.10-3 N
+ Cross-sectional area: S = 900 mm2
+ Substitute into the equation: tmax = 4,6.10-6 (MPa)
=> tmax< [t]
With n = 1,5: safety factor.
=> The lower control arm satisfies the shear strength condition.
VonMises Stress will be calculated for further comparison with simulation results: sv = 20,89 MPa
3.7.2. 1Ball Joint Strength Calculation
The ball joint is a spherical joint that connects the control arm to the steering knuckle. Its working condition is mainly subjected to shear, bending, and bearing (crushing) loads.
3.7.2.1. Shear Strength Calculation
Where, Q: Shear Force
+ Case 1:
Qc1 = 4011,37 N
Qc2= 14428,09 N
+ Case 2:
Qc1 = 5361,39 N
Qc2= 12905,93 N
+ Case 3:
Qc1 = 7559,48 N
Qc2= 132136,24 N
Here, Case 1 yields the maximum shear force
3.7.2.2. Bending Stress Calculation
Mu: Bending moment;
h: Maximum distance from the neutral axis, h = 13 (mm)
=> M0 = 785,40 N.m
=> su = 238,83 MPa
3.7.2.3. Crushing Stress Calculation
The maximum bearing force occurs in Case 2: Q = Fz = 7964,74 N
Where:
Scd: Contact surface area, taken as 2/3 of the spherical surface area
R: Radius of the ball, selected as R = 15 (mm)
Then: δcd = 4,23 Mpa
- Also: [dcd] = 150 (MPa).
- Then dcd £ [dcd]. Satisfies the crushing strength condition.
3.8. Shock Absorber Calculation
3.8.1. Calculating Shock Absorber Damping Force
According to automotive theory, the damping effect is represented by a damping force Qc applied at the wheel. This damping force depends on the vibration velocity V and follows the rule:
Q = k.Vm
Where: k is the coefficient characterizing the damping resistance of the suspension system and m ≈1
=> k = 1973 Ns/m
3.8.3. Determining the Dimensions of the Front Shock Absorber
Select the following basic dimensions:
- Shock absorber length L: Chosen to match the suspension’s dynamic travel. Select L = 400 (mm).
- Piston diameter Dp: Chosen such that working pressure falls within 2.5–5 MPa. Select Dp = 40 (mm).
- Piston rod diameter Dc: Chosen based on piston diameter: Dc = (0,4÷0,5).Dp. Select Dc = 16 (mm).
Where kgn and kgt are damping coefficients calculated above, Vg is damping coefficients calculated above, select Vg = 0,3 (m/s)
Ntt = 190,13 W
Energy absorbed by the shock absorber is converted into heat and dissipated through the outer tube.
Where:
kτ is heat transfer coefficient, kτ = 45÷60 w/(m2.oC). Select kτ = 60 w/(m2.oC) Tg is shock absorber surface temperature Tg ≤ [Tgmax] = 120oC.
Tm is ambient temperature. Select Tm = 27 (oC)
=> Sg = 0,11 m2
=> Tg = 55,81 0C
=> Thermal condition satisfied
3.8.5. Determination of damper spring dimensions
With D3 = 13,28 (mm), D4 = 10,08 (mm)
D is the average diameter of spring, D = 24 (mm),
d is the wire diameter of the spring,
P2 is force acting on the spring when the valve is fully open
The allowable stress of the spring material is: [t] = 50 ¸ 70(MPa). Select [t] = 70
=> Select d = 3 (mm)
Spring length when valve is fully open:
Hg = .n.d + d.n0 = 8,8 mm
Where,
d is the gap between adjacent coils, d = 0,8(mm)
n0 is the total number of coils, n0 = n+1 = 5+1 = 6 (turns)
Spring pitch: t = 2,33 mm
Designed Parameter of the Damper Spring Table 2.
3.9. Spring Calculation for Front Suspension System
In a suspension system, spring is an elastic element responsible for absorbing and smoothing out motion. During operation, the spring is subjected only to vertical loads (Z-direction) and does not transmit longitudinal or lateral forces.
Based on the loading conditions analyzed in the dynamic section, the dynamic load Z reaches its maximum value; therefore, the spring should be designed for this critical load case.
3.9.1. Force Acting on Spring
The spring is designed to withstand the maximum dynamic load: Z = 13185,9 (N)
Where:
Z: is the maximum dynamic load
llx: is the moment arm of the spring placement, llx = l = 290 (mm)
ld: is the length of the suspension arm, ld = 300,69 (mm).
Minimum force acting on the spring:
=> Fms = 6452,39 N
Where, M10 is the front axle load when unloaded, M10 = 1244,6 (Kg).
3.9.2. Spring Design Procedure
Design Data: Fmax = 13671,96 (N); Fmin = 6452,39 (N)
Design Procedure:
* Step 1: Select Material and Basic Spring Parameters
Select material 50CrV4 with:
- Tensile strength of helical spring: [t] = 1600 MPa
- Wire diameter: d = 10 ¸ 20 mm
- Spring index: c = 10
* Step 3: Determine Spring Dimensions
With the compressive spring, total number of coils (n₀):
n0 = n + 1,5 = 7 + 1,5 = 8,5 (turns)
Spring length (fully compressed):
Hs = (n0 – 0,5). d = (8,5 - 0.5). 16 = 128 (mm)
Free length of spring:
H0 = Hs + n. (t - d) = 128 + 7. (101,71 - 16) = 727,97 (mm).
Spring length under minimum load: H0 = 0,489 (m).
* Step 4: Strength Check
Maximum torsional stress: tmax = 1550,37 MPa
Minimum torsional stress: tmin = 606,26 MPa
=> Both values satisfy the shear strength condition.
Designed Parameter of Spring in Front Suspension System Table 3
chapter 4. REAR SUSPENSION DESIGN
4.1 Determine the basic parameters of the rear suspension system
To assess a car’s ride comfort during motion, oscillation frequency, oscillation acceleration and oscillation velocity are used. In design calculations, oscillation frequency serves as the basis for selecting the basic parameters of the elastic component. For passenger vehicles, the oscillation frequency should be chosen within the range of 60-90 cycles per minute.
4.1.1 Determine Spring Stiffness
Spring stiffness is calculated according to the condition that it must be compatible with an oscillation frequency in the range of 60-90 cycles/minute.
Choose an oscillation frequency n = 70 cycles/minute.
Natural frequency: w = 7,33 rad/s
The average stiffness of the suspension system is calculated by taking the average of two values: when the vehicle is fully loaded and when it is unloaded.
For the rear suspension, stiffness of the elastic element: C = 403,5 N/cm
4.1.4. Check dynamic wheel travel
While:
H0: Static ground clearance under load
Hmin: Minimum required ground clearance
Hmin = 0.1 to 0.16 (m). Choose Hmin = 0.155 (m)
=> H0 fd + Hmin = 0.147 + 0.155 = 0.302 (m)
=> H0 302 (mm)
4.1.5. Determine the average damping coefficient of the damper
Damping coefficient of the suspension system: D = 2,932 rad/s
Average damping coefficient of the shock absorber, referred to the wheel:
For the rear suspension: Kϕ1 = 1172,8
4.1.6. Design approach for the rear suspension system
Multilink suspension system is selected for the rear suspension system design due to the restriction of time and knowledge of the thesis
Database for calculation:
- Vehicle rear axle track width B = 1667 (mm)
- Wheel radius: Tire designation 245/45R19. rbx = 351.55 (mm)
- Kingpin inclination angle: β = 15o
- Maximum ear height of the vehicle: Htmax = 539 (mm)
- Wheel Camber angle: α = 0
- Caster angle: γ0 = 5o
- Wheel scrub radius r0 = -15 (mm)
- Ground clearance at full load: Hmin =155 mm.
- Static deflection ft = 184 (mm)
- Dynamic deflection fd = 384 (mm)
- Suspension deflection in the unloaded state f01 = 438 (mm)
- Kingpin length Kr = 150 (mm)
4.2. Multilink suspension kinematics (2 lateral arm only)
The specific steps are as follows:
- Draw a horizontal line representing the road surface plane: dd
- Draw the horizontal axis of symmetry of the vehicle A0m: A0m perpendicular to dd.
- On the A0m, put: A0A1 = Hmin = 155 (mm); A1A2 = fd = 384 (mm);
- A2A3 = ft = 184 (mm); A3A4 = f01 = 438 (mm).
-On A0-dd put A0B0 = B/2 = 833.5 (mm).
- B0 is the contact path of the tire with the road surface
- At point B0 build B0z perpendicular to the ground line dd.
- On line A0B0 extends outward beyond segment A0B0 and B0C0= |r0| =15 mm.
- At C0 built C0n make an angle with the vertical axis β = 15o.
- On B0z put B0B = rbx = 351.55 (mm).
- At point B construct a line perpendicular to B0z that intersects C0n at the C3. C3 is the rigid connection between the wheel strut and the steering axis.
Similarly, we can find the position of the outer joint of the control arm (or "transverse arm") in the fully loaded position as follows: When the suspension reaches its maximum deformation, if we consider the vehicle body to be stationary, the wheel will translate upwards to point B1.
If we consider the distance between the two-wheel tracks in this state to be constant compared to the unloaded state.
Then, B0B1 = fd + ft - f01 =184+384–438= 130mm.
- From B1 draw the line B1q // dd.
- On B1q put B1D1 = B0C0= |ro| =15 mm.
- From D1 built D1n’ with the vertical plane β’ = β0 + ∆β = 130, inersecting Cn at O2.
- On D1n’ put D1D2 = C0C1.
- If O1C1 is extended and a line perpendicular to O2C0 is drawn, they will intersect at P (the instantaneous center of rotation of the wheel).
- Connect PB0 and extend it to intersect A0m at S (S is the instantaneous center of rotation for both the axle and the vehicle body in the horizontal plane of the axle).
- Connect PO2 and extend it until it intersects line k1k1 at O2 so O2 then the inner pivot joint of the upper control arm. O2C2 = Lt is the length of the upper transverse suspension arm.
4.4. Spring Calculation
In the suspension system, spring is the elastic component responsible for cushioning movement. During operation, the spring is only subjected to vertical loads (Z) and does not transmit longitudinal or lateral forces.
Based on the load conditions analyzed in the dynamics section, we see that the dynamic load case has the largest Z value, so we need to design according to this load condition.
The spring is calculated for the maximum dynamic load case: Z = 7424,5 (N)
There is a maximum force acting on the spring: Fmax = 8995,1 N
While:
Z: Dynamic load
llx: Spring arm length llx = l = 335,8 (mm)
ld: Control arm length ld = 331,2 (mm).
Minimum force acting on the spring: FmIN = 3095,5 N
While M10 is the load on the front axle when unloaded (or unladen front axle load), M10 = 6277 (Kg).
4.5. Damper Design
4.5.1 Damper Selection
Through the analysis of damper structures, we chose to design and calculate a single-tube damper with a compressed gas chamber (Nitrogen N2 gas), where the gas pressure in this chamber equals the oil pressure. Furthermore, a single-tube damper has a simple structure, is easy to manufacture, repair, and maintain. Moreover, this type of damper is very sensitive to light compression and rebound. If two dampers have the same cylinder diameter, a single-tube damper can have a larger piston rod compared to a twin-tube damper.
4.5.3. Determine the external dimensions of the shock absorber
The most demanding working condition is defined as: v = 0,3 (m/s)
Preliminary selection of the shock absorber L = 400 (mm)
The total length of the shock absorber includes:
Ld the length of the shock absorber head
Lm the length of the sealing unit
Lp is the length of the shock absorber piston.
Lg is the maximum working stroke of the shock absorber.
If the piston diameter d is taken as the basic parameter, the other parameters are determined as follows:
D = 45 (mm); d = 35 (mm); dc = 10 (mm); dn = 38 (mm); D0 = 50 (mm)
Lp = 35 (mm); Ld = 50 (mm); Lm = 50 (mm); Lv = 30 (mm); Lg = 200(mm)
4.5.4. Determine the dimensions of the valves
When the shock absorber operates, the following cases may occur:
- Case of light rebound valve operation
- Case of strong rebound valve operation
- Case of light compression valve operation
- Case of strong compression valve operation
Section 1-1 refers to the cross-section of the fluid inside the piston. Therefore, the flow velocity at section 1–1 is the relative velocity between the piston and the cylinder.
Section 2–2 is the cross-section of the fluid at the valve orifice outlet.
The difference in geometric height Δz between the two sections is very small (equal to the height of the orifice), so it is neglected in calculations.
The fluid flows through the orifice in turbulent mode, hence the kinetic energy correction factor α = 1.
The average energy loss hw₁₋₂ along the flow represents the amount of kinetic energy converted into thermal energy due to friction between the fluid and the orifice wall, fluid-to-fluid interaction, and friction between the fluid and the cylinder wall.
4.6. Check the strength condition
4.6.1Piston rod strength check
Check the strength condition of the piston rod diameter under the maximum load acting on the wheel. When in operation, the wheel is subjected to dynamic loads, The maximum value of the dynamic load is approximately twice the static load. Therefore, the dynamic load is:
Zđmax = 2.Zbx = 17500 (N)
The maximum tensile (compressive) stress generated in the piston rod is: smax = 99 MPa
The piston rod material is selected as 40 steel with an allowable stress [σ] = 400 (MPa) The maximum stress generated in the piston rod is less than the allowable stress of the material [σ] = 400 (MPa). Therefore, the shock absorber's piston rod satisfies the strength condition.
4.6.2. Lower control arm (I-shaped arm)
The lower control arm has a I-shaped structure and is mounted to the vehicle body using bolts. The outer end is connected to the knuckle bracket.
The main load conditions include tension, compression, and bending. The cross-section of the lower control arm is referenced, and for durability analysis, it is assumed that one part of the I-shaped arm bears the entire load.
4.6.2.1. Maximum traction/braking load
We have: Z = 4454,7 ; Xmax = 3341 N
And Mx = 1174,53
We also have: Z = 4454,7
In the maximum braking scenario, only vertical (Z) forces are present, with no lateral (Y) force. Because the longitudinal arm has absorbed all longitudinal force, the force is considered negligible , FX = 0, FY = 0, FZ = Flx
Fz: Acts as a shear force and generates a longitudinal bending moment in the plane (zoy).
We have maximum shear stress: tmax = 22,58 (N/mm2)
With the material being aluminum alloy AlZnMgCu1,5F50, it has the following properties:
sb = 510 (Mpa)
[t] = sb / 2n = 170 (Mpa) ³ tmax
=> Satisfies the strength condition
Mu = Z.( - (L – 250)) = 6955,77.130 = 904250 (N.mm)
The cross-section is U-shaped with the following dimensions: h = 40 (mm), b = 60 (mm), t = 3 (mm)
Jx =172600 (mm4)
y: taken at the point with a vertical coordinate of max: y = 36,2 (mm)
sb = 189,65 (N/mm2)
=> Satisfies the strength condition.
4.62.3 Dynamic Load Only
Only force Z, Fx = 0, Fy = 0, Fz = Flx
Fz: Acts as a shear force and induces a longitudinal bending moment.
We have: Flx = 11591,2 (N)
The cross-section is U-shaped with the following dimensions: h = 40 (mm), b = 60 (mm), t = 3 (mm)
Jx =172600 (mm4)
y: taken at the point with a vertical coordinate of max: y = 36,2 (mm)
sb = 316,04 (N/mm2)
=> Satisfies the strength condition.
VonMises Stress will be calculated for further comparison with simulation results: sv = 304,27 (N/mm2)
chapter 5. STRUCTURE SIMULATION OF THE SUSPENSION SYSTEM
5.1 Overview of Simulation Methods
5.1.1 Structure Simulation Methods
5.1.1.1. Rayleigh-Ritz Approximation Method (Rayleigh-Ritz)
The Rayleigh–Ritz method is an approximation technique used to solve variational problems, especially those involving the extremization of energy functionals in mechanics. It is commonly applied in problems of beam bending, mechanical vibrations, and linear elasticity.
This method is based on variation principles—typically the principle of minimum potential energy or Hamilton’s principle—where the exact solution of the problem corresponds to a function that minimizes the total potential energy functional.
5.1.1.2. Finite Difference Method (FDM)
The Finite Difference Method is a numerical technique that approximates solutions to ordinary differential equations (ODEs) and partial differential equations (PDEs) by replacing derivatives with discrete difference expressions. This method transforms a continuous problem into a discrete one by dividing the domain into grid points and applying different formulas.
Derivatives are approximated by finite differences between function values at adjacent grid points, resulting in a system of linear or nonlinear algebraic equations that can be solved computationally.
5.1.2 Types of Structural Durability Analysis Problems
5.1.2.1. Static Strength Analysis
The main purpose of this analysis is to ensure that the structure can withstand all static loads encountered during the manufacturing, transportation, and launch without excessive deformation.
A static analysis calculates the effects of steady loading conditions on a structure, while ignoring inertia and damping effects, caused by time-varying loads. A static analysis can include steady inertia loads (such as gravity and rotational velocity), and time-varying loads that can be approximated as equivalent static loads (such as static equivalent wind). A static analysis can be either linear or nonlinear. All types of nonlinearities are allowed: large deformations, plasticity, creep, stress stiffening, contact elements, and so on. This section focuses on linear static analyses only.
- Externally applied forces and pressures
- Steady-state inertial forces (such as gravity or rotational velocity)
- Imposed (nonzero) displacements
- Temperatures (for thermal strain)
5.1.2.3. Harmonic Response Analysis
The main purpose of this analysis is to investigate whether the structure can withstand all dynamic loads encountered during transportation and operations. The results of harmonic analysis are used in the next section to perform stress fatigue damage analysis for the structure due to mechanical dynamic vibration.
Any sustained cyclic load will produce a sustained cyclic response (a harmonic response) in a structural system. Harmonic response analysis gives the ability to predict the sustained dynamic behavior of any structure, thus enabling to verify whether structural designs will successfully overcome resonance, fatigue, and other harmful effects of forced vibrations.
5.1.2.5. Spectrum Analysis
The main purpose of this analysis is to provide investigations into whether the structure can withstand all dynamic random vibration loads encountered during transportation and operation. The entire finite-element model of the structure is used during the spectrum analysis process to calculate displacements and stresses in the structure modules.
- Response Spectrum
- Dynamic Design Analysis Method (DDAM)
- Power Spectral Density (PSD)
5.1.3. Detail Steps in general structural analysis
The 6-steps instruction below is used in the thesis to conduct the simulation for the suspension system.
Step 1: Build the 3D model of the structure, import the geometry into the simulation application
Step 2: Conduct geometry treatment and meshing
Step 3: Define materials, properties
Step 4: Define boundary conditions
Step 5: Run analysis
Step 6: Analyze the simulation results
5.2. Structural simulation of Front Suspension System
5.2.1. Lower Control Arm (LCA)
5.2.1.1. 3D Modelling
3D Model of LCA as shown 5.2.
5.2.1.3.Material and Properties Defining
Identification of Material and Property of LCA as shown 5.5.
5.2.1.6. Results Analyzing
The stress in all cases is represented as von Mises equivalent stress.
LCA Simulation Results Table 4
The finite element analysis (FEA) of the Lower Control Arm (LCA) subjected to three distinct loading scenarios confirmed that the von Mises stresses remained within acceptable thresholds for the chosen material, AlMnZnCu1,5F50. Simulated maximum stress values ranged between 26.5 mm and 27.7 mm, while calculated stress values were slightly lower, resulting in deviations of 5.33% to 5.5% in Cases 1 and 2, and a minimal 1.4% in Case 3. These discrepancies fall within the acceptable limits for structural validation, thereby ensuring the credibility of the simulation.
The maximum displacement values observed across all loading cases were notably low, with the peak value recorded at 2,6.10-1 mm. When contrasted with the overall dimensions of the LCA (minimum length approximately 300 mm), these displacements are negligible, which reflects the component’s high structural rigidity and favorable deformation behavior under stress.
5.3. Structural simulation of Rear Suspension System
5.3.1. Lower Horizontal Arm (LHA)
5.3.1.1. 3D Modelling
3D Model of LHA as shown 5.26.
5.3.1.3. Material and Properties Defining
Identification of Material and Property of LHA as shown as shown 5.29.
5.3.1.4. Boundary Condition Defining
Identification of Boundary Conditinos of LHA - RBE2 Elements as shown 5.30.
5.3.1.6. Results Analyzing
The stress in all cases is represented as von Mises equivalent stress.
Lower Horizontal Simulation Results Table 6
The finite element analysis of the Lower Horizontal Arm (LHA) under three distinct loading conditions demonstrated that the von Mises equivalent stresses remained within the allowable range for the selected material, AlMnZnCu1,5F50. The calculated maximum stress values were slightly higher than the simulated results, with deviations ranging from 1.4% to 7%. Such discrepancies are minor and fall within acceptable thresholds for structural verification, validating the accuracy of the simulation approach.
The observed maximum displacements across all cases were minimal, with the highest value recorded at 2.3 mm. When compared to the overall structural dimension of the LHA (with a minimum length of approximately 300 mm), these displacements are negligible. This reflects a high degree of structural stiffness and confirms the component’s resistance to deformation under operational loads.
In summary, the simulation results affirm the mechanical soundness and durability of the current LHA design under various loading scenarios. While the overall performance is satisfactory, localized design refinements in high-stress zones may further enhance the arm's reliability and fatigue resistance.
chapter 6. EXPERIMENTS OF THE SUSPENSION SYSTEM
6.1 Overview of Vehicle Experiement
6.1.1 Objectives of the Experiment
The purpose of the experiment is to evaluate and identify the advantages and disadvantages of component assemblies, systems, and the entire vehicle during the stages of design, manufacturing, and operation, focusing on the following aspects:
- Technical specifications and fundamental operating characteristics
- Operational reliability (such as trajectory stability during braking, steering, etc.).
In summary, the aim of vehicle experimentation is to comprehensively assess the quality of components, systems, and the entire vehicle, thereby proposing improvements to enhance product design, manufacturing, and diagnostics for better vehicle performance.
6.1.2. Requirements for Measuring Equipment
- Ensuring the required accuracy for the experiment, being unaffected by vibrations
- Having linear or near-linear characteristics for easier extrapolation and interpolation (y = f(x))
- Being compact and lightweight, being resistant to weather and environmental influences
6.2. Overview of the Kistler RoaDyn Test System
6.2.1Structure of the Kistler RoaDyn Test System
The Kistler RoaDyn test system is a force and torque measurement system applied directly to the wheel during vehicle operation. It utilizes force sensors mounted at the wheel hub to accurately capture data on longitudinal forces, lateral forces, vertical forces, and torque. This equipment is essential for vehicle dynamics research, suspension system design, tire testing, and performance evaluation under real-world driving conditions.
The Kistler RoaDyn system is a high-precision wheel force transducer used in vehicle dynamics research to measure the complete set of forces and moments acting at the wheel center. Structurally, the system integrates multiple functional components arranged concentrically around a load-bearing rim.
The core sensing elements are the load cells, which are symmetrically arranged between the inner part (hub-side structure) and the rim. These load cells can capture three orthogonal force components (longitudinal, lateral, and vertical) as well as the associated moments (pitch, roll, and yaw), thus enabling six-component force/moment analysis in real-time.
6.2.2. Experimental Procedure and Installation of the Kistler RoaDyn
6.2.2.1. Required devices to operate KiRoad Performance:
- The wheels equipped with Load Cell sensors and accompanying devices (In/Outboard Transmission, HUB Electronics, magnetic mounts, and wire guiding components) have been installed on the test vehicle (two wheels RR and RL mounted on a Toyota Hiace). The KiRoad Performance system and necessary connection cables (power cable, LAN cable, USB cable, wheel sensor cables, WLAN, CAN to DTI cable) are also set up.
- The laptop has been installed with the KiCenter software.
- Power supply unit from Kistler and a power battery (12V, over 50 Ah).
- Spirit level or water leveling hose.
- Accompanying technical manuals and user guides.
6.2.2.2. Steps connecting ECU to the KiRoad Performance
Working Principle of the KiRoad Wheel Dynamics Measurement System:
The Load Cell sensors operate based on the strain gauge principle, in order to ensure accurate display values, the sensor angles must first be calibrated correctly, followed by force calibration.
* Step 1: Connecting the ECU to the Computer
Connect the KiRoad Performance (abbreviated as ECU) as shown with the devices (power supply, wheel sensor outputs, computer, WLAN, and DTI Logger if necessary) according to the diagram via the connectors shown. Turn on the power switch, monitor the display screen, and verify the ID, IP address, RR + RL (confirm connection with the two rear wheels: right rear - RR and left rear - RL).
Once the connection is completed, click on the KiRoad Performance tab (5413698), and all information should be fully displayed.
* Step 2: Sensor calibration from the beginning is performed
Place the vehicle on a lift so that the wheels can rotate freely. Spin the wheels smoothly for 2–3 rotations and ensure the error between the actual rotation angle and the displayed rotation angle in the Measurement Display tab is minimal. Rotate the wheels to the zero-degree position as specified in the user manual. Sensor F should be at the topmost position, and sensors A and E should lie on the same horizontal plane. This can be done using a spirit level or water leveling hose. Then gently pull the handbrake to fix the position of the wheels (note: do not rotate the wheels while pulling the handbrake).
Select Sensor adjustment → Adjustment → Angle Offset → Set angle offset (wheel 1 and wheel 2). Enter the angle 330°.
Or select Sensor adjustment → Adjustment → Angle Offset → Set angle offset by Internal Inclinometer.
After angle calibration is done, place the vehicle back on the lift to allow free wheel rotation. Spin the wheels 2–3 times to verify the displayed angle in the Measurement Display.
6.3.3. Signal Acquisition and Data Processing Unit
The output signals from the strain gauge sensors are typically very small, requiring amplification and bridge balancing when necessary. To meet these signal processing requirements, the project team utilizes the SIRIUS signal processing system from DEWEsoft
The signal processing unit is connected to a personal computer installed with DEWEsoft's specialized software for data acquisition and experimental storage.
6.3.5. Experimental Procedure
The procedure for stress measurement on the vehicle frame includes the following steps:
- Preparing the test vehicle
- Test surface: flat road surface (static testing only)
- Installing strain gauges
- Configuring and calibrating the measurement system
- Conducting the experiment according to the planned test scenario
6.3.6. Sensor Installation Procedure
The installation of the strain gauge follows these steps:
- Use a jack to lift the cabin, engine, and transmission.
- Determine sensor installation positions.
- Clean and polish the bonding surfaces using a grinder and sandpaper.
- Apply epoxy adhesive to the back of the sensor and press it into the pre-marked position for 1 minute.
- Cover the sensor with silicone adhesive and wrap it with tape to secure both the sensor and signal cables.
6.4. Experimental Scenarios
6.4.1. Scenario 1 - ISO Wave Road Test
Due to practical constraints, our team conducted the Kistler RoaDyn experiment on an alternative reference vehicle - the Mitsubishi Outlander PHEV-whose fully loaded weight is approximately equal to the curb weight of our primary reference vehicle, the VinFast VF8. The purpose was to obtain parameters of torque and forces acting on the wheels as the vehicle passes over a bump, to be used as input data for the simulation model.
During this experiment, key parameters were measured using the Kistler RoaDyn wheel force transducer system. These included the vertical load distribution (Fz) acting on each of the four wheels over time, as well as the dynamic wheel torque (My) and rolling moment (Mx) generated as the vehicle traversed the undulating surface. These measurements provided crucial insights into how effectively the suspension system absorbs vertical oscillations and maintains load balance across the axle during transient excitations.
6.4.2. Scenario 2 - Maximum Braking Test
The Low-Speed Maximum Braking Test aims to evaluate the suspension system’s behavior and wheel load response under sudden braking at low speeds, typically ranging from 5 to 10 km/h. Although not representative of high-speed emergency braking, this test is particularly important for assessing low-speed dynamics, validating simulation models, and calibrating sensors or control systems such as autonomous driving algorithms, electric vehicle regeneration braking, or advanced driver-assistance systems (ADAS).
The test procedure involved propelling the vehicle to a steady low speed within the specified range, followed by the application of full braking force while maintaining a straight trajectory with the steering wheel in a neutral position. The test surface was flat and of high friction to ensure consistent conditions.
6.5. Experimental Results and Analyzes
The experiments are conducted with Strain Gauge Sensor only due to the lack of facilites and time, the Kistler RoaDyn will be conducted in the near future and will be presented later.
Do treo sau không thiết kế theo đúng hệ thống treo trên xe tham khảo VinFast VF8, đồ án sẽ chỉ so sánh kết quả thí nghiệm và kết quả mô phỏng của hệ thống treo trước (with the scenario maximum braking test), còn kết quả thí nghiệm của hệ thống treo sau (with the scenario Maximum Load Test) và phần thí nghiệm còn lại của hệ thống treo trước (with the scenario ISO Wave Road Test) sẽ được sử dụng để tham khảo cho các nghiên cứu trong tương lai.
6.6. Extended Recommendation for Experimental Testing Using Kistler RoaDyn on VinFast VF8
Due to the lack of knowledge, time and facilites, the following recommendation will be mentioned.
To improve the accuracy and reliability of the dynamic simulation model of the VinFast VF8, it is strongly recommended to conduct an experimental test utilizing the Kistler RoaDyn wheel force transducer system. This advanced measurement system is capable of capturing six-component wheel forces and moments (Fx, Fy, Fz, Mx, My, Mz) during real driving conditions with high precision and resolution.
The objective of this test is to obtain real-world force data under specific vehicle operating scenarios-such as driving over standardized bump roads (e.g., ISO wave road) or during maximum braking at low-to-medium speeds (5–60 km/h). These scenarios should be carefully chosen to represent critical loading conditions for the vehicle’s suspension system. The measured data will include vertical loads, longitudinal and lateral forces, as well as moments at the wheel hubs, providing comprehensive insight into how the suspension behaves under actual dynamic conditions.
Furthermore, this approach will support the validation and refinement of simulation models by comparing measured and simulated results. Any significant discrepancies can be used to adjust suspension parameters, damping coefficients, or tire models, thereby improving the model’s predictive capabilities. The outcomes of this process will be beneficial for ongoing vehicle development, particularly in optimizing ride comfort, handling, and structural durability.
chapter 7. MAINTENANCE AND REPAIR OF THE SUSPENSION SYSTEM
7.1 Front Suspension System
7.1.1 Common Failures in the Suspension System
- Torn or cracked dust boots
- Torn or cracked strut mount bushings
- Broken springs
- Oil leakage from shock absorbers, failure to rebound after compression
- Damaged or scratched piston rods
So, Suspension system faults can often be detected through unusual noises or other noticeable symptoms. However, to ensure safety, the suspension system should be routinely inspected by professional technicians—ideally during every oil change-to promptly detect issues and apply appropriate solutions such as maintenance or replacement.
7.1.1.1. Related to Shock Absorbers
The shock absorber is one of the most failure-prone components in the suspension system. It serves to limit the motion of leaf springs or coil springs when the vehicle encounters road obstacles by utilizing the resistance of oil flowing through an orifice within the piston. At the same time, it provides ride comfort by absorbing and dissipating body vibrations.
However, shock absorbers frequently suffer from oil leakage or clogged orifices. In such cases, the outer tube becomes wet and attracts dust and dirt. Immediate replacement is necessary to avoid prolonged oscillation, resulting in a bouncy ride and discomfort for the driver.
7.1.1.3. Related to the Elastic Components
Springs in the suspension system tend to weaken or "sag" over time, particularly in vehicles frequently subjected to overloading. In such cases, the elastic components of the suspension fatigue, losing stiffness and elasticity.
This degradation reduces ride comfort, as the suspension becomes less effective in absorbing road irregularities, resulting in decreased ground clearance.
To prevent spring breakage or consequential damage to shock absorbers and other components, users should have the springs inspected and replaced in a timely manner.
7.1.2 Methods for Detecting Suspension Failures
7.1.2.1. While Driving and During Test Drive
Conduct a focused test drive, paying close attention to any signs of malfunction.
Lower the vehicle windows and listen carefully for any unusual sounds. If any noises are heard, attempt to identify their origin. Common suspension-related noises include:
- Knocking or clunking sounds: Often heard when hitting bumps, indicating possible issues with the piston rod, mounting nuts, or ball joints.
- Continuous humming or growling sounds: These intensify as speed increases and are usually caused by faulty wheel bearings or tire-related problems.
- Metallic clanking or rattling sounds: These may result from damaged bolts or broken connecting elements, sounding like metal parts striking each other.
7.1.2.3. Checking for Suspension Free Play
Use a jack to slightly lift the vehicle until the tire is off the ground while ensuring it remains stable. Firmly grip the tire and shake it with hands positioned at 9 o'clock – 3 o'clock and 12 o'clock – 6 o'clock. If any unusual movement is detected, it may indicate wear in some suspension components.
Be aware that abnormal movement could stem from various sources, so careful inspection and accurate diagnosis are necessary.
7.1.3. Maintaining Suspension Stability
Maintaining the stability of the suspension system is essential for ensuring safe vehicle operation. Regular inspection of suspension components such as shock absorbers, upright ball joints, and stabilizer links should be part of routine maintenance, especially after scheduled oil changes.
Key maintenance points include:
- Inspect rubber bushings such as A-arm bushings, stabilizer bar bushings, shock absorber mounts, and strut mounts.
- Check for cracked dust boots, delaminated rubber, oil leaks from shocks, or any missing parts in the suspension system.
- Look for debris or oil leakage on or around the bushings. Oil leaks should be repaired immediately.
- Inspect ball joints and connecting points of the stabilizer bar and steering outer tie rods.
- Ensure proper lubrication: if grease is visible in unintended areas, make sure it is reapplied to the correct locations during periodic maintenance.
- Check for oil leaks from shocks and ball joints. Any component with visible oil contamination should be replaced promptly.
7.2 Rear suspension system
7.2.1 Visual Inspection
The visual inspection is the first and crucial step to identify any abnormalities in the suspension system before proceeding with disassembly or replacement. The following checks should be performed:
* Visual observation:
- Inspect the connection points between the lateral links, trailing arm, and wheel hub for any signs of cracks, corrosion, deformation, or bending.
- Pay special attention to rubber bushings, checking for cracks, aging, tearing, or off-center displacement. These are common causes of noise and misaligned wheel geometry.
* Play check:
- Manually shake the rear wheels in both lateral and longitudinal directions (after lifting the vehicle) to detect any excessive play at ball joints or bushings.
- If play exceeds the allowable limit, evaluate the wear of each component to determine if replacement is necessary.
7.2.2. Disassembly of the Suspension System
If damaged components are identified, the suspension system should be disassembled for further inspection or replacement. Proper disassembly technique is required to avoid secondary damage:
- Lifting the vehicle: Use a two-post lift or hydraulic jack to elevate the rear end of the vehicle, ensuring the rear wheels are off the ground.
- Removing rear wheels: Use an impact wrench or torque bar to remove the wheels, creating a workspace for disassembly.
- Removing the shock absorber (if mounted to the trailing arm): Unbolt the upper and lower mounts of the damper to detach it from the suspension assembly.
7.2.3. Cleaning and Component Inspection
After disassembly, clean and inspect each component carefully to determine whether it should be reused or replaced.
- Cleaning:
+ Use a wire brush or clean cloth with brake cleaner to remove dirt, grease, and debris.
+ Pay close attention to link ends and rubber bushings for clear visibility during inspection.
- Component inspection:
+ Ball joints: Check for excessive play, dry movement, seized joints, or damaged dust boots.
+ Bushings: Examine for off-center displacement, cracks, wear, or uneven deformation.
+ Control arms (links): Place on a flat surface to check for bends or cracks. Use rulers or dial gauges to measure deflection if necessary.
CONCLUSION
This graduation thesis has systematically addressed the analysis, design, simulation, and testing of a vehicle suspension system through seven interconnected chapters. Each chapter contributed to the overall objective of evaluating and improving the structural performance and practical applicability of the system. The outcomes, along with identified limitations, are summarized as follows:
- Chapter 1 provided an overview of suspension systems, their classifications, functions, and importance in vehicle dynamics. While comprehensive in scope, this chapter was limited to a literature review without in-depth analysis of current advancements in active or semi-active suspension technologies.
- Chapter 2 presented the process of selecting appropriate design options for both the front and rear suspension. Although the choices were justified based on standard engineering criteria, the evaluation could be further enhanced with more comparative simulation data or experimental benchmarks.
- Chapter 3 and Chapter 4 detailed the design of the front and rear suspension systems, respectively. These chapters succeeded in laying out clear design parameters and modeling logic. However, constraints on time and computational resources limited the exploration of multiple design iterations or optimization techniques.
- Chapter 5 focused on the simulation of the suspension components under different loading conditions. The Finite Element Analysis (FEA) results showed that the simulated stress distributions closely approximated theoretical calculations, indicating that the digital models were built with a high degree of accuracy and reliability.
- Chapter 6 involved experimental testing of the suspension system. Although the test results exhibited significant deviation from theoretical and simulated outcomes-primarily due to limited time, project scope, and the unavailability of advanced measuring tools like the Kistler RoaDyn system-they still offered valuable insights into the actual mechanical behavior of the suspension components. These experiments highlighted real-world responses that are not fully captured by idealized models, reinforcing the importance of physical validation.
- Chapter 7 addressed the maintenance and repair aspects of the suspension system. This practical focus added useful guidelines for post-design considerations, although it remained general due to the project's emphasis on design and analysis.
In conclusion, this thesis successfully achieved its primary goals in design, modeling, and preliminary validation of a vehicle suspension system. Despite the noted limitations, especially in experimental precision, the work has laid a solid foundation for future research. Further improvements should involve more extensive testing with advanced instrumentation and expanded simulation cases to fully optimize both design and durability aspects of the system.
REFERENCES
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