The advent of in-wheel motor (IWM) technology represents a paradigm shift in the architecture of the modern electric vehicle car. By integrating the traction motor directly into the wheel hub, this configuration offers significant advantages such as enhanced powertrain efficiency, simplified vehicle chassis design, and unparalleled potential for independent wheel torque control. However, this innovative propulsion method introduces substantial challenges for the vehicle’s suspension system. The incorporation of the motor within the wheel assembly significantly increases the unsprung mass and occupies critical space traditionally reserved for suspension components. Consequently, the suspension system for an electric vehicle car equipped with IWMs cannot be a mere adaptation of conventional designs; it requires a fundamental re-evaluation and redesign to maintain, or even improve, vehicle dynamics, ride comfort, and handling stability.
This article details a comprehensive methodology for the design and optimization of a front MacPherson strut suspension system tailored specifically for an electric vehicle car driven by in-wheel motors. The primary objective is to develop a suspension layout that accommodates the spatial constraints imposed by the IWM while preserving the hardpoint interfaces with the vehicle’s lower body structure. This constraint ensures compatibility with existing vehicle platforms, minimizing broader architectural changes. The core of the development process leverages multi-body dynamics simulation using ADAMS/Car software to construct a precise virtual prototype, analyze its kinematic and compliance characteristics, and iteratively optimize hardpoint locations to achieve desired performance targets.
1. Introduction and Design Challenges
The transition to electrification has unlocked numerous vehicle layouts, with the in-wheel motor drive being one of the most radical. In a typical electric vehicle car with a central motor, the suspension design can often inherit principles from internal combustion engine vehicles. In contrast, an IWM-driven electric vehicle car presents unique hurdles:
- Increased Unsprung Mass: The motor, brake, and associated components become part of the unsprung mass. This increase can severely degrade ride comfort, as the suspension’s ability to isolate the body from road irregularities is compromised, and worsen wheel hop dynamics, potentially leading to a loss of tire contact.
- Packaging Constraints: The physical volume of the IWM intrudes into the space typically used for the suspension’s lower control arm, steering linkage, and damper assembly. This necessitates a complete re-packaging of these components, often requiring novel geometric layouts.
- Altered Load Paths and Load Cases: The suspension must react to high driving and regenerative braking torques directly at the wheel hub, introducing new load cases not present in conventional vehicles. Components like the lower control arm and strut must be designed to handle these additional moments and forces.
- Kinematic Compromise: The need to route the steering mechanism around the motor can force compromises in suspension kinematic design, potentially leading to undesirable changes in parameters like camber and toe angle during wheel travel and steering.
The design challenge, therefore, is to synthesize a suspension geometry that mitigates the drawbacks of high unsprung mass (through intelligent kinematics) while fitting within the tight packaging envelope and maintaining acceptable levels of steering effort and feedback for the electric vehicle car.
2. Theoretical Foundation of Suspension Kinematics
The performance of a suspension system is fundamentally governed by its kinematic behavior—the precise motion of the wheel assembly relative to the vehicle body. For a MacPherson strut suspension, the wheel carrier is connected through a lower control arm (LCA) and a telescopic strut that acts as both a guiding element and a spring-damper unit. The kinematics are described by the variation of key wheel alignment parameters as functions of wheel vertical travel (jounce/rebound) and steering input.
The primary parameters are:
- Camber Angle (γ): The inclination of the wheel plane from the vertical. Negative camber during cornering improves lateral force generation. The camber change with wheel travel, ∂γ/∂z, is critical.
- Toe Angle (δ): The angular deviation of the wheel’s longitudinal plane from the vehicle’s direction of travel. Minimal variation with travel is desired for straight-line stability.
- Kingpin Inclination (KPI) & Caster Angle (τ): Define the steering axis orientation. They influence steering feel, self-centering torque, and camber change during steering.
- Scrub Radius (rs): The distance at the road surface between the tire contact patch center and the point where the steering axis intersects the ground. A small or negative scrub radius is often targeted for reduced steering disturbance during braking.
- Roll Center Height (RCH): The instantaneous center about which the vehicle body rolls. Its location affects the roll moment distribution and the vehicle’s roll behavior.
The motion can be analyzed using vector loop methods. Defining position vectors for hardpoints, one can derive the functional relationships. For instance, the wheel center position $\vec{r}_{WC}$ relative to the body can be expressed as a function of the LCA inner joint $\vec{r}_A$, LCA outer joint $\vec{r}_B$, strut top mount $\vec{r}_T$, and strut-to-knuckle connection $\vec{r}_S$. A simplified constraint equation for the strut length $L_s$ is:
$$ L_s = | \vec{r}_S – \vec{r}_T | $$
As the wheel moves, $\vec{r}_B$ moves on a sphere (if the LCA is a simple link) or a more complex curve (if it is an A-arm), and $\vec{r}_S$ moves with the knuckle. The solution of these constraint equations for various wheel travel positions yields the alignment curves. The derivatives of these curves are the key performance metrics. For example, the camber gain per unit of body roll is crucial for handling and is targeted to be in a specific range, such as:
$$ \text{Target Camber Gain: } \frac{\partial \gamma}{\partial \phi} \approx -0.6 \text{ to } -0.8 \,\, \text{deg/deg}$$
where $\phi$ is the body roll angle.
3. Virtual Prototyping and Optimization Methodology
The design process follows a structured virtual development cycle. The baseline is an existing front suspension from a prototype vehicle. The IWM dimensions and mounting requirements are defined as new boundary conditions.
Step 1: ADAMS/Car Model Construction. A precise multi-body dynamics model of the modified MacPherson suspension is built. Key components—the lower control arm, tie rod, knuckle (wheel carrier), strut, and the IWM unit (modeled as a rigid mass attached to the knuckle)—are created with accurate mass and inertia properties. The joints (revolution, spherical, fixed) are defined at the initial hardpoint locations. The tire is modeled using a PAC2002 or similar semi-empirical model for force calculation. The model for one corner of the electric vehicle car is then assembled into a full vehicle template for subsequent analyses.
Step 2: Kinematic and Compliance (K&C) Analysis. A series of standardized simulations are run on the quarter-car or full-vehicle model:
- Parallel Wheel Travel: Both wheels move vertically in phase to analyze alignment parameter variations with ride motion.
- Opposite Wheel Travel: Wheels move out of phase to simulate roll and analyze roll center migration.
- Steering Sweep: At different ride heights, the steering wheel is turned to analyze steering-related parameters like toe-out-on-turn, Ackermann error, and caster trail variation.
- Compliance Analysis: Forces and moments are applied at the tire contact patch to simulate the effects of bushings and component flexibility on wheel alignment under load (braking, cornering, acceleration).
Step 3: Design Optimization Formulation. The initial hardpoint locations, chosen to package around the IWM, rarely yield ideal kinematics. Therefore, a mathematical optimization problem is formulated.
Design Variables (x): The coordinates (x, y, z) of movable hardpoints, such as the LCA inner and outer ball joints, the strut top mount, and the tie rod inner and outer points. Typically, 10-15 variables are considered. Their movement is bounded by packaging constraints.
Objective Function (f(x)): A weighted sum of squares of deviations from target performance curves. For example:
$$ f(\mathbf{x}) = w_1 \sum ( \gamma(\mathbf{x}, z) – \gamma_{target}(z) )^2 + w_2 \sum ( \delta(\mathbf{x}, z) – \delta_{target}(z) )^2 + w_3 ( r_s(\mathbf{x}) – r_{s,target} )^2 + … $$
where $w_i$ are weighting factors reflecting the importance of each target.
Constraints (g(x), h(x)): These include packaging limits (clearance to IWM, chassis), joint angle limits (to prevent binding), and specific performance bounds (e.g., scrub radius must be between -10mm and +5mm).
The optimization is executed within ADAMS/Car using its built-in Design of Experiments (DOE) and optimization tools (like Insight). An iterative process of DOE sampling, surrogate model creation, and gradient-based optimization is used to find the hardpoint set that minimizes $f(\mathbf{x})$ while satisfying all constraints.
| Hardpoint | Coordinate | Initial Value (mm) | Lower Bound (mm) | Upper Bound (mm) | Primary Influence |
|---|---|---|---|---|---|
| LCA Inner Front | X | 450.0 | 440.0 | 460.0 | Wheelbase, Anti-dive, Roll Center |
| Y | -280.0 | -290.0 | -270.0 | ||
| Z | 120.0 | 110.0 | 130.0 | ||
| LCA Outer | X | 780.0 | 775.0 | 790.0 | Camber Gain, Track Width |
| Y | -650.0 | -660.0 | -640.0 | ||
| Z | 150.0 | 140.0 | 165.0 | ||
| Strut Top Mount | X | 400.0 | 390.0 | 410.0 | Caster, KPI, Lateral Force Steer |
| Y | -550.0 | -560.0 | -540.0 | ||
| Z | 700.0 | 680.0 | 720.0 | ||
| Tie Rod Outer | X | 800.0 | 795.0 | 810.0 | Toe Curve, Ackermann, Bump Steer |
| Y | -620.0 | -630.0 | -610.0 | ||
| Z | 180.0 | 170.0 | 190.0 |
4. Analysis of Results and Performance Discussion
Following the optimization routine, the kinematic characteristics of the final IWM-adapted suspension are compared against the original design targets and, where relevant, the baseline suspension. The results demonstrate the efficacy of the process in mitigating the inherent challenges of the IWM electric vehicle car.
| Performance Metric | Unit | Baseline (Conventional EV) | Initial IWM Design | Optimized IWM Design | Design Target |
|---|---|---|---|---|---|
| Camber Gain (Roll) | deg/deg | -0.65 | -0.35 | -0.72 | -0.6 to -0.8 |
| Toe Change (60mm Jounce) | deg | 0.15 (Toe-in) | -0.40 (Toe-out) | 0.08 (Toe-in) | 0.0 to 0.2 Toe-in |
| Static Scrub Radius | mm | +12.5 | +32.0 | -3.2 | -10 to +5 |
| Roll Center Height (Static) | mm | 65 | 42 | 58 | 50 – 70 |
| Caster Angle (Static) | deg | 4.0 | 2.5 | 3.8 | 3.5 – 4.5 |
| Kingpin Inclination | deg | 10.5 | 14.2 | 11.0 | 9 – 12 |
| Bump Steer Range (±50mm travel) | deg | ±0.25 | ±0.80 | ±0.18 | ≤ ±0.30 |
The table reveals critical insights. The initial IWM design, constrained by packaging, performed poorly: low camber gain, excessive toe-out in bump, and a large positive scrub radius. This combination would lead to unstable cornering, nervous straight-line behavior, and significant torque steer/brake pull in the electric vehicle car. The optimized design successfully brought all parameters within the target ranges.
Camber Gain: The optimized value of -0.72 deg/deg is excellent. During cornering, as the body rolls, the outside wheel gains negative camber, helping to maintain a larger tire contact patch and generating higher lateral force to counter the increased roll moment from the higher unsprung mass. This is a vital compensation mechanism for the IWM-driven electric vehicle car.
Toe Characteristics: The optimization minimized bump steer and achieved a slight toe-in during jounce. This promotes straight-line stability and understeer characteristics, which are generally desirable for predictable handling, especially in a high-torque application like an IWM electric vehicle car.
Scrub Radius: Reducing the scrub radius to a slightly negative value (-3.2mm) is a major achievement. A negative scrub radius helps mitigate steering wheel pull during asymmetric braking (e.g., on a µ-split surface), a crucial safety and comfort feature. It also reduces the steering disturbance caused by longitudinal forces from the in-wheel motor’s torque output.
Roll Center and Caster: The Roll Center Height was raised to 58mm, providing a better lever arm for the suspension to resist roll, partially offsetting the higher CG often associated with battery packs in an electric vehicle car. The caster angle was restored to near 4°, ensuring good straight-line stability and steering self-centering feel.
The graphical outputs from ADAMS/Car further illustrate these improvements. For example, the plot of camber vs. wheel travel shows a near-linear and steeper slope for the optimized design compared to the shallow curve of the initial design. Similarly, the toe curve is flattened and kept in the toe-in region across the travel range. The migration path of the roll center is more controlled and remains within a desirable vertical window.
The optimization also considered the compliance steer effects. Under high cornering or braking loads delivered by the IWM, the deflections of the bushings—particularly the lower control arm front and rear bushings—were analyzed. The final hardpoint configuration was chosen to minimize unwanted toe changes due to these compliance effects, ensuring consistent handling at the limits for the performance-oriented electric vehicle car.
5. Implications for Vehicle Dynamics and Future Work
The successful kinematic optimization of the MacPherson suspension for the IWM application forms a solid foundation. However, it is only the first layer of the development process for the complete electric vehicle car. The next critical steps involve:
Elasto-Kinematic (Compliance) Modeling: A more detailed model incorporating the full non-linear stiffness characteristics of all rubber bushings is required. This allows for accurate prediction of handling response under combined longitudinal (IWM torque) and lateral forces.
Damping and Spring Rate Tuning: The significant increase in unsprung mass necessitates a re-tuning of the damper and spring rates. Stiffer damping may be required to control the higher inertial forces of the unsprung mass, but this must be carefully balanced against ride comfort. The optimal force-velocity profile for the damper in an IWM-driven electric vehicle car likely differs from that of a conventional vehicle. The spring rate must be chosen to maintain a target ride frequency, considering the increased unsprung mass:
$$ f_{ride} = \frac{1}{2\pi} \sqrt{\frac{K_{sprung}}{M_{sprung}}} $$
While the sprung mass $M_{sprung}$ may be similar, the wheel hop frequency, dependent on unsprung mass $M_{unsprung}$ and tire stiffness $K_{tire}$, increases, potentially requiring dedicated tuning:
$$ f_{hop} \approx \frac{1}{2\pi} \sqrt{\frac{K_{tire} + K_{sprung}}{M_{unsprung}}} $$
Integration with Torque Vectoring: This is the ultimate synergy. The independently controllable torque at each wheel opens the door for advanced yaw moment control (torque vectoring). The suspension’s kinematic and compliance characteristics directly influence how effectively these control moments are translated into vehicle motion. A cooperative design approach, where suspension kinematics are slightly biased to complement the torque vectoring controller’s authority, could yield superior agility and stability. For instance, a specific toe compliance under longitudinal force could be designed to work in concert with differential torque application.
Durability and NVH Analysis: The loads on suspension components, especially the lower control arm and strut mount, are higher due to the reaction torques from the IWM. Finite Element Analysis (FEA) is necessary to ensure component durability. Furthermore, the IWM itself is a source of vibration and electromagnetic noise; the suspension bushings and top mount must also be designed to provide adequate isolation from these high-frequency disturbances to the cabin of the electric vehicle car.
6. Conclusion
The design and development of a suspension system for an in-wheel motor driven electric vehicle car is a complex, multi-disciplinary challenge that cannot be addressed by conventional design rules alone. This work has demonstrated a systematic, simulation-driven approach to this problem. By starting from a prototype platform and imposing strict packaging constraints around the IWM, a modified MacPherson strut configuration was developed. Through rigorous kinematic and compliance analysis and subsequent mathematical optimization of hardpoint locations in ADAMS/Car, a final suspension geometry was achieved that meets stringent performance targets.
The results show that it is possible to overcome the inherent kinematic drawbacks of initial packaging solutions. Key parameters such as camber gain, bump steer, and scrub radius were optimized to not only match but in some aspects exceed the performance of a conventional suspension, thereby actively compensating for the increased unsprung mass. This optimized kinematics provides the necessary foundation for stable handling, predictable steering feel, and good ride compromise—all essential attributes for a commercially viable and enjoyable electric vehicle car.
The methodology outlined, combining virtual prototyping with multi-body dynamics simulation and numerical optimization, is highly efficient and accurate. It significantly reduces the reliance on physical prototypes and trial-and-error, accelerating the development cycle. This process holds high reference value for the automotive industry as it advances towards next-generation electric vehicle architectures centered on distributed drive systems like in-wheel motors. The future of suspension design for the high-performance electric vehicle car lies in this deeper integration of mechanics, electronics, and control, with advanced simulation serving as the critical enabler.