A New Approach to In-Wheel Motor Solutions for Electric Vehicles

Abstract

The In-Wheel Motor represents a non-conventional propulsion architecture in which the electric motor is integrated into the wheel, offering advantages such as improved energy efficiency, individual torque control, and drivetrain simplification. In this study, two architectures, inboard and outboard, were developed using an original three-dimensional motor–brake–suspension–steering assembly model, in which disk brake position and In-Wheel Motor integration act as primary design drivers influencing vehicle dynamics. Both architectures were developed in CATIA V5 and exported to Altair Motion 2025 for multibody dynamics simulations. The study evaluates the impact of inboard versus outboard disk brake positioning on vehicle dynamics and provides a qualitative assessment of the associated architectures in terms of mechanical complexity, serviceability, sealing requirements, bearing load asymmetry, and packaging constraints. The results indicate that the inboard architecture exhibits more linear and stable kinematics and compliance (K&C) behavior compared to the outboard configuration, at the expense of increased mechanical complexity and reduced serviceability. By contrast, the outboard architecture preserves a simpler, more conventional MacPherson-like layout with a lower component count and improved service access but is dynamically outperformed under the imposed geometric constraints of the present study.

1. Introduction

The focus of decarbonization of road transport, sustained by the European Green Deal [1] and the Fit-for-55 program [2], has accelerated the development of high-efficiency electric powertrains and distributed propulsion concepts [3,4]. Within the context of advances in SiC (Silicone Carbide) inverters, AFPM (Axial Flux Permanent Magnet) maturity, IWMs (In-Wheel Motors) have re-emerged as a viable architecture because they simplify the drivetrain, reduce parasitic losses, act as an architectural enabler for precise wheel-torque control for stability, safety, and performance [5,6,7,8,9], and are increasingly relevant for FCEVs (Fuel Cell Electric Vehicles), as they can support modular powertrain layout that can accommodate complex hydrogen storage systems and energy management strategies [10]. When coupled with modern traction control and torque-vectoring strategies, IWMs improve vehicle dynamics and energy efficiency while enabling alternative vehicle-level architectures [11,12], consistent with experimental and numerical studies on distributed propulsion architectures [13,14,15].

The practical feasibility of IWMs has been largely enabled by the evolution of AFPM motors [16,17]. Their compact “pancake” geometry and high torque density are uniquely suited to the spatial constraints of a wheel rim [18,19], as corroborated by recent literature reviews of direct-drive motor architectures for EVs (Electric Vehicles) [20,21]. However, this high level of integration gives rise to two tightly coupled, fundamental challenges: a significant increase in unsprung mass, which can detrimentally impact ride comfort and tire-road contact [22,23], and the critical need for robust thermal management within a high-vibration environment [24,25,26].

These challenges directly affect the integration of ancillary systems, making the positioning of the disk brake a first-order architectural decision in IWM applications [27,28]. Two main architectures can be identified:

• The outboard brake architecture, where the disk is positioned between the wheel and motor, preserves conventional serviceability but risks exacerbating brake-to-motor heat conduction [23].
• The inboard brake architecture, where the disk is relocated toward the chassis, can improve thermal behavior by enabling less restricted airflow paths, but introduces complexity in suspension geometry and caliper mounting, potentially increasing unsprung mass [29,30].

While the literature provides a strong foundation in individual domains such as IWM vehicle dynamics [30,31], thermal management [32,33], and control [5,34], a clear gap remains. There is a lack of a systematic, space-constrained comparison of inboard versus outboard brake placements within a modern IWM system, using identical chassis hard point location and standardized kinematics and compliance procedures, while prior studies focus on either dynamics or thermal aspects separately [35]. Such a study is necessary to evaluate the brake positioning on vehicle dynamics, constrained by the other geometric variables.

This paper closes that gap by providing a quantitative and qualitative comparison of the two brake architectures integrated into a MacPherson strut suspension with a virtual pivot, a layout chosen for its relevance to IWM packaging. We hypothesize that the inboard architecture will yield superior dynamic stability and linear K&C behavior due to a more symmetrical weight distribution, while the outboard layout will exhibit sharper responses, due to design constraints of shorter strut and shorter suspension arms, at the cost of increased compliance and nonlinearities [29,34,35,36].

The objective of this study is not to create a kinematically optimized suspension for each powertrain architecture, but to quantify the first-order effects induced by brake disk relocation under fixed chassis fixation points. The imposed constraints retained realistic production packaging limitations without compensatory re-tuning suspension, in order to preserve comparability and reveal intrinsic architectural trade-offs.

To evaluate this, detailed three-dimensional models of both architectures in CATIA V5 were developed, and a set of multibody dynamics K&C simulations was executed in Altair Motion 2025. The results provide a crucial framework for assessing IWM-based vehicle architectures, offering clear guidance on the trade-offs between dynamic performance, mechanical complexity, and cost.

The novelty of this work lies in the evaluation of inboard and outboard brake architectures coupled effects within an identical MacPherson suspension with identical chassis fixation, without compensatory re-tuning. This approach isolates the intrinsic architectural effects of brake placement on K&C behavior, bearing loads, and system complexity on a fixed suspension topology in a defined space constraint.

The paper is structured as follows: Section 2 details the methodological approach, including geometric modeling and simulation setup. Section 3 presents the comparative results of the K&C analysis. Section 4 discusses the broader implications, including mechanical and cost trade-offs. Finally, Section 5 concludes with design recommendations and outlines directions for future work.

2. Materials and Methods

This section details the systematic approach employed to model, simulate, and analyze the two IWM architectures. The methodology encompasses geometric modeling in CAD, with the input parameters defined in Table 1 and Table A1, the definition of multibody dynamics parameters, which can be found through Appendix A Table A2, Table A3, Table A4, Table A5 and Table A6, and the specific K&C simulation protocol.

Table 1. Vehicle data used as input for 3D and simulation.

Parameter	Value	Comment
Vehicle weight (static load equivalent)	22,250 N	Direct measurement on D-segment ICE vehicle
Motor torque	400 Nm (continuous) 790 Nm (peak @450ARMS)	Use data from available mass market AFPM available into data sheet specification [37,38]
Tire model	Rigid	Used for K&C isolated suspension geometry and compliance
Wheel size	235/45 R18	Direct measurement on D-segment ICE vehicle

Table 1 summarizes the input data used for the 3D modeling and multibody simulations. A rigid wheel representation was adopted to avoid confounding effects from tire compliance and to isolate suspension and packaging constraints geometric effects.

Multibody dynamics simulation is a time-dependent kinematic and dynamic analysis that determines the response of a system under time-dependent loading [39]. In this study, the elements of the suspension are considered rigid, except for joints/bushings, spring, and shock absorber of the MacPherson strut, which are defined in Appendix A. Also, the vertical rate of the wheel is defined in Appendix A.

The governing equations of motion for the multibody system are expressed as follows:

$$M·\ddot{u}=f\left(t\right)$$

(1)

$${\dot{u}}_{t=0}=v_0$$

(2)

where $M$ is the mass matrix, the vector $f$ is the vector of external forces, and the vector $u$ represents the generalized coordinates.

Stiffness, damping, constraint forces, external loads, and gravity are all included in the external force vector $f$. An initial and maximum integration time step, an end time, and integrator tolerance need to be defined [39].

The time step definition follows the standard K&C simulation protocol, a set of tests defined into Altair MotionSolve library, presented below.

Both models were subjected to a standardized set of K&C tests to evaluate suspension behavior under controlled conditions. The simulation sequence included:

• Ride (0–10 s): Assessed vertical jounce and rebound.
• Roll (10–20 s): Analyzed behavior during opposing wheel displacement.
• Contact Forces (20–80 s): Measured responses to lateral, aligning, braking, and acceleration forces.
• Steering (80–90 s): Evaluated input-induced toe variation.

2.1. IWM Model Overview and Technology Background

The core of the assembly is an AFPM motor, selected for its high torque density and short axial length, which are critical for integration within the wheel rim envelope. The motor design is based on the principles of YASA (Yokeless and Segmented Armatures), which shorten thermal paths to the coolant and reduce axial length [16,17,18,25,37,38,40,41]. The model incorporates integrated power electronics, integrating the inverter with the motor within the same housing to reduce DC-link inductance and improve power density, though concentrating heat sources in the wheel rim [42].

To address thermal and fault-tolerance challenges, a redundant sub-motor and sub-inverter segmentation strategy is employed, inspired by commercial implementations. This design allows for selective torque derating and limp-home functionality but increases connector count, influencing the overall packaging constraints [43].

2.2. Geometric Modeling and Design Constraints

A key challenge addressed in this study is the geometric packaging of an axial-flux electric motor within the wheel envelope. The motor construction and design fill the confined space in the wheel rim, making it unavailable for classic knuckle and lower ball joints. This fundamental constraint led to the development of two distinct architectures for integrating the disk brake components:

• Outboard Architecture: The disk brake is placed in the same position as on a classic suspension design.
• Inboard Architecture: The electric motor is placed nearest to the wheel, and the disk brake is positioned inboard of the motor.

Both architectures were developed using a classic MacPherson suspension, selected for its wide availability, cost efficiency, and the design space it allows for other drivetrain components [44].

A critical packaging constraint was imposed by the position of the lower ball joint, which was limited by physical interference with the motor housing. To address this limitation while preserving a conventional MacPherson suspension geometry, a “virtual” ball joint concept was implemented by dividing the single lower arm into two separate arms, front and back, thereby defining a virtual pivot point, as illustrated in Figure 1 and Figure 2. This solution enabled the steering axis to closely replicate that of classic MacPherson suspension used in ICE (Internal Combustion Engine) architecture, maintaining an identical kingpin offset of 10 mm and avoiding detrimental effects on steering feel and straight-line stability [45,46].

Figure 1. Back view of classic MacPherson suspension physical ball joint, on the left-hand side, and construction “virtual” lower ball joint for inboard architecture, in the middle, and outboard architecture, on the right-hand side. Dashed lines shows the geometrical construction of suspension geometry and the line show the middle of the tyre.

Figure 2. Top view of classic MacPherson suspension physical ball joint, on the left-hand side, and a virtual ball joint construction of the MacPherson suspension for inboard architecture, in the middle, and outboard architecture, on the right-hand side. Dashed lines shows the geometrical construction of suspension geometry.

The kingpin offset, defined as the lateral distance in the ground plane between the projection of the steering axis and the wheel centerline, is therefore identical for all shown architectures. With a rigid tire model employed, the kingpin offset directly characterizes the effective steering geometry without confounding effects from tire compliance.

A key aspect of both architectures is the integration of the fixation points for the suspension strut and the arm ball joints into the motor interior cover or motor housing. Components used for both architectures are presented in Table 2 with different numbers in each architecture construction.

Table 2. Bill of Materials for both architectures.

Component	No. of Components
	Outboard Architecture	Inboard Architecture
Motor housing	1	1
Rotor with magnets	2	2
Coils	24	24
Ferrite of coils	24	24
Power electronics	1	1
Shaft	1	1
Bearings for shaft	1	1
Carrier plastic for holding the coils on the shaft	1	1
Carrier plastic for holding the coils on the housing	1	1
O-ring seal	1	2
Cover (wheel side)	1	1
Cover (suspension side)	1	1
Screws for fixing the carrier plastic to the housing	24	24
Screws for fixing the electronics	36	36
Screws for fixing the rotor to the shaft	12	12
Screws for fixing the covers	48	48
Screws for shock absorber	1	1
Flange for wheel/disk brake	1	2
Screw for fixing the flange	1	2
Studs for wheel/disk brake fixation	5	10

To ensure a straight and accurate comparison, both suspension architectures, shown in Figure 3 and Figure 4, were designed using the same fixation points on the chassis level with specific constraints given by the architecture or design. One fundamental difference was the strut length and the suspension arm length; due to packaging constraints, the outboard solution required a 30 mm shorter strut than the inboard solution.

Figure 3. MacPherson suspension with disk brake inboard architecture.

Figure 4. MacPherson suspension with disk brake outboard architecture.

The applied K&C test is a standard test from Altair MotionSolve library, which perform ride test simulating a 100 mm vertical displacement for the inboard solution and 70 mm for the outboard solution, due to the design constraint mentioned above, over 10 s to assess jounce and rebound characteristics. The roll test applied ±50 mm, respectively, ±35 mm, again due to strut constraint, opposing displacements to the wheels to evaluate camber and toe changes during cornering. Contact force tests included lateral forces up to 3000 N and longitudinal forces up to 4000 N to replicate braking and acceleration scenarios. Steering tests measured toe-angle variation under a 360° steering wheel input to quantify responsiveness.

Afterwards, for inboard and outboard architecture, the mechanical challenges derived from the design and constraints established previously will be addressed.

Although the outboard architecture requires a strut that is 30 mm shorter than the inboard configuration, this difference is a direct consequence of brake placement under fixed chassis hard-point constraints rather than an independent design choice. No compensatory suspension re-tuning or geometric optimization was applied in order to preserve identical chassis fixation points and ensure architectural comparability. As a result, the observed differences reflect coupled architectural effects due to shorter suspension arms associated with brake and IWM placement, rather than optimized suspension performance.

3. Results

The K&C simulations revealed distinct dynamic profiles for each proposed architecture.

The ride test demonstrated fundamentally different vertical force characteristics between the two architectures. As shown in Figure 5 and Table A7, the inboard architecture exhibits an approximately linear vertical-force trend over the investigated travel range, whereas the outboard configuration shows a more progressive (nonlinear) rate in jounce, which correlates with the shorter effective control arm geometry imposed by packaging constraints.

Figure 5. Comparison of evolution of vertical wheel force during jounce and rebound between inboard architecture and outboard architecture.

Roll test outcomes, summarized in Figure 6 and Table A8, revealed significant differences in camber gain. The inboard solution produced a wider, more gradual camber curve, promoting consistent tire contact patch retention. Conversely, the outboard architecture exhibited a sharper, parabolic camber response, reaching a higher peak camber angle under maximum roll.

Figure 6. Camber angle as a function of roll angle during the standardized roll test.

The response to braking and acceleration forces highlighted a critical divergence in longitudinal compliance. As shown in Figure 7, the inboard architecture exhibited minimal toe change, with less than 0.75° deviation under a 4000 N braking force. This stability is attributed to its symmetrical load paths. The outboard design demonstrated progressive toe-out behavior, reaching 0.8° at peak load, indicating geometric compliance that could affect straight-line tracking during hard deceleration.

Figure 7. Toe angle variation on longitudinal acceleration test.

The toe compliance, $C_{toe}$, is defined as the rate of change in the toe angle, $\partial {\theta }_{toe}$, with respect to the applied force, $F_x$, [47]:

$$C_{\mathrm{t}\mathrm{o}\mathrm{e}}={\displaystyle \frac{\partial {\theta }_{toe}}{\partial F_x}}$$

(3)

The variation trends shown in Figure 6 correspond to the analytical definition of toe compliance provided in (3).

Longitudinal displacement measurements during braking tests, presented in Figure 8, further underscored this difference. While both systems behaved similarly up to 1500 N, the outboard design exhibited 15% less wheel center movement at the 4000 N peak load. This nonlinear divergence correlates with the shorter effective length of the outboard control arms, which amplifies elastic deformations under high stress. The inboard architecture displacement curve remained linear, confirming its superior stiffness in resisting brake torque reactions.

Figure 8. Longitudinal displacement of the wheel center during the longitudinal braking test.

Steering response analysis uncovered fundamental differences in handling linearity. Figure 9 illustrates that the inboard architecture keeps a near-linear relationship between steering wheel input and toe angle across the full ±360° range, conceptually defined as a function of $y_{rack}$—steering rack travel, $∆z_{wheel}$—vertical wheel travel, and $C_{compliance}$—compliance of the elastic elements, with a maximum variation of ~15°:

$${\delta }_{toe}=f(y_{rack},∆z_{wheel}, C_{compliance})$$

(4)

Figure 9. Toe angle variation under steering test.

The outboard design, however, displayed pronounced nonlinearity, with toe changes accelerating beyond ±200 steering wheel input. This behavior stems from the compromised Outer Tie Rod Ball Joint (OTRB) position, a direct consequence of the outboard brake packaging constraints.

The qualitative summary of the K&C tests is presented in Table 3, highlighting the fundamental trade-off between stability and responsiveness. −, °, and + indicate undesired, moderate, and desired behavior, respectively.

Table 3. Overview of test results outcome.

Performance Metric	Inboard	Outboard
Straight-Line Stability	Lower bump steer and less wheelbase change +	Quicker yaw response in bumps −
Cornering and Turn-In	Lower camber gain for even tire wear °	Higher camber gain improves cornering grip, but increases wear +
Braking and Acceleration	Smaller toe changes lead to a more stable line under acceleration or braking +	Larger toe changes can cause more yaw on acceleration or braking −
Steering Feel	Smoother toe progression and consistent torque °	Stronger toe variation gives sharper, more sensitive feedback −

For a comprehensive overview, data are provided in Appendix B Table A7, Table A8, Table A9, Table A10 and Table A11 to eliminate the influence of unequal maximum suspension travel, and a normalized comparison of K&C gradients was performed within the common metrics, presented in Table 4.

Table 4. Normalized K&C gradients.

Metric	Inboard Architecture	Outboard Architecture
Vertical wheel rate	≈97 N/mm	≈132 N/mm
Camber gain (ride)	≈0.008 deg/mm	≈0.017 deg/mm
Toe gradient (ride)	≈0.64·10−4 deg/mm	≈0.62·10−4 deg/mm
Camber gain (roll)	≈0.01 deg/mm	≈0.015 deg/mm
Toe compliance (braking)	≈1.9·10−4 deg/N	≈2·10−4 deg/N
Longitudinal wheel center compliance	≈34·10−4 mm/N	≈33·10−4 mm/N

In summary, despite its geometric constraints, the outboard architecture exhibited toe compliance under braking similar to the inboard design, as shown in Table 4. However, its significantly higher vertical wheel rate and camber gain during ride are direct consequences of the shorter suspension arms and strut.

4. Discussion

4.1. Vehicle Dynamics

This study provides a systematic, simulation-driven comparison of inboard versus outboard disk brake architectures for an AFPM-based IWM system [48]. The results reveal a fundamental trade-off between dynamic performance, mechanical feasibility, and economic viability, with direct implications for the practical implementation of IWM vehicle architectures.

The analysis indicates that neither architecture represents a universally optimal solution. Instead, each architecture offers specific advantages while introducing specific drawbacks, resulting in a design compromise that must be resolved at each vehicle design level.

The inboard architecture exhibits superior dynamic stability, characterized by its linear K&C behavior, minimal compliance steer, and symmetrical bearing loads, establishing it as the superior dynamic linearity and stability within the investigated constraints. This aligns with classical vehicle dynamics theory, where predictable responses are paramount for stability and comfort. Furthermore, its layout may offer potential thermal advantages, subject to dedicated thermal validation, by distancing the motor from the primary brake heat source.

This analysis highlights a critical downside of the outboard architecture, due to the placement of the IWM and its size, the required shorter suspension arm compared with the inboard architecture. This constraint is shown in Figure 10, which is the main contributor of the nonlinearities from K&C, surpassing the shorter strut implied by the space and interference constraint.

Figure 10. Left-hand side—outboard architecture with asymmetrical loading of the bearing. Right-hand side—inboard architecture with symmetrical loading of the bearing.

Figure 11 illustrates the stress induced into the system through the position of the disk brake. For the outboard architecture, on the left-hand side, bearings are placed at $X_{max}$, and in the inboard architecture, bearings are placed at $X_{min}$.

Figure 11. Cross-section view of inboard architecture with two rotational seals. One towards the wheel and the second one towards the disk brake, closer to the power electronics.

Brake moment, $M_{brake}$, will be the same for both cases:

$$M_{brake}={\mathrm{F}}_{\mathrm{b}\mathrm{r}\mathrm{a}\mathrm{k}\mathrm{e}}·\mu ·R_{eff}$$

(5)

but, the difference will be in the force on the bearing, $F_{bearing}$:

$$F_{bearing}\approx {\displaystyle \frac{{M_{brake}+(F}_z· X_{max,min})}{s}}$$

(6)

where $F_{brake}$ is the tangential braking force at the disk’s effective radius, $R_{eff}$ is the effective radius of the disk brake, $\mu$ is the friction coefficient of braking, $s$ is the span between the bearings, $F_z$ is the vertical force, and $X_{max,min}$ is the distance between the SAI (steering axis)/KPI (kingpin inclination) of the wheel and the middle plane of the bearing package.

Due to $X_{max}>X_{min}$, the $F_{bearing}$ in the outboard architecture will be proportionally larger than in the inboard architecture.

To evaluate bearing durability, the lever arm $X_{max,min}$ in the outboard architecture, which is larger due to positing of the disk brake, amplifies the reaction forces $F_{bearing}$ under combined vertical and braking loads. Conversely, the inboard architecture allows the bearings to be positioned closer to the steering axis, $X_{min}$, minimizing the bending moments and ensuring a more symmetrical load distribution.

4.2. Complexity and Serviceability Considerations

These performance benefits are accompanied by notable mechanical complexity and serviceability penalties. The BOM (Bill of Materials), presented in Table 2, is significantly larger, requiring an additional wheel flange, extra fastening elements, and, most critically, a second rotational shaft sealing, shown in Figure 11. This dual-seal requirement is not merely an added component; it introduces significant manufacturing complexity and potential points of failure, and is a direct driver of increased assembly complexity and cost [49]. Consequently, while dynamically preferable for a premium passenger vehicle, its high complexity and severe maintenance challenges (likely requiring full IWM disassembly for brake service) make it a difficult solution for mass production [49,50].

The primary advantage of the outboard architecture lies in its simple BOM and preserved serviceability, making it initially attractive from a complexity and assembly perspective. It avoids the complex dual-sealing system of the inboard design.

However, this apparent complexity is offset when lifecycle effects and mechanical limitations are considered. The asymmetrical bearing loads, presented in Figure 11, suggest an increased risk of accelerated wear, requiring durability validation, potentially necessitating the use of higher-capacity and more costly bearing solutions [26,51]. Simultaneously, the high bushing stress from the short control arm geometry, directly evidenced by the large compliance steer in the results, suggests an increased risk and inconsistent long-term handling, requiring durability validation. These factors imply increased durability risk and lifecycle cost, partially negating the initial BOM advantage. This configuration also introduces ground clearance constraints between the suspension arm and the ground.

Equations (1)–(4) are introduced to support the physical interpretation of the multibody dynamics results rather than to replace numerical simulation. In particular, Equation (4) provides a conceptual explanation for the higher toe compliance observed in the outboard configuration under longitudinal loading, while Equations (5) and (6) illustrate the amplification of bearing reaction forces due to reduced effective lever arms.

Equation (6) represents a quasi-static approximation intended to explain the observed trends in bearing loading. Dynamic bearing reaction forces were not explicitly extracted from the multibody solver and will be addressed in future work.

The presented findings provide a quantitative foundation for previously qualitative discussions in the literature [12,16,20,26]. By directly correlating brake placement with measurable compliance behavior and identifiable cost drivers, this study establishes a structured decision-making framework for early-stage vehicle architecture development.

For premium vehicle segments, the inboard solution may justify its higher complexity due to its superior refinement and potential thermal stability.

For well-known solutions and high-volume production, the outboard solution is a feasible starting point; however, the results clearly indicate that substantial geometric and component-level optimization would be required to achieve acceptable durability and dynamic performance, thereby reducing its initial BOM and assembly advantage.

4.3. Limitations and Future Work

Several limitations of this study must be acknowledged. First, the exclusive reliance on multibody dynamics simulation incorporates simplifications that may influence the dynamic behavior of the vehicle. Validation using a physical prototype is therefore required.

Second, this study focused on dynamics, qualitative complexity, and serviceability considerations. Potential thermal advantages of the inboard layout and the NVH (Noise, Vibration and Harshness) implications of both architectures, while logically deduced, require quantitative verification through coupled multiphysics simulations [25,26,28,52].

No direct thermal, NVH, or fatigue simulations were conducted; all related implications should therefore be interpreted as hypotheses derived from mechanical load paths.

Third, the geometric constraints, such as shorter strut (30 mm) and shorter suspension arms, imposed on the outboard architecture are a direct consequence of the chosen fixed hardpoint methodology and significantly impact the results. Future work should explore compensation within these constraints.

Furthermore, extracting dynamic bearing reaction forces from the multibody model will enable precise durability analysis and bearing selection.

5. Conclusions

This study systematically evaluated the influence of disk brake placement on the performance and viability of an In-Wheel Motor system for Electric Vehicles, FCEVs, or other suitable vehicles. Through a detailed multibody dynamics analysis of inboard and outboard architecture, a clear and consequential trade-off has been quantified, revealing that the optimal choice is not a technical absolute, but a strategic compromise defined by the vehicle target application.

The inboard brake architecture was confirmed as the dynamically superior solution, offering linear and predictable handling, enhanced stability, and favorable mechanical load distribution. However, these advantages are critically offset by a more complex BOM, a challenging dual-shaft sealing requirement, and reduced serviceability, rendering it a high-complexity option difficult to justify for mass-market applications.

Conversely, the outboard brake architecture, while simpler to service and, initially, a more established architecture, was found to be fundamentally constrained by packaging geometry. Its geometry leads to significant dynamic drawbacks, including nonlinear steering response and excessive compliance steer, and poses a risk of premature failure due to asymmetrical bearing loads. Without a significant redesign, it cannot be considered a direct replacement for conventional suspension layouts.

The main contribution of this work is the provision of a holistic decision-making framework that requires the simultaneous consideration of dynamic performance, mechanical durability and complexity, and serviceability viability.

There is no universal solution; the inboard architecture caters to premium segments where refinement justifies complexity, while the outboard layout requires significant geometric and component-level adjustment to become a robust solution for volume production.

Additional future work must focus on bridging this gap. The main priority is the geometric optimization of the outboard solution, through the exploration of geometric compensation (suspension arm length), spring-damper retuning, OTRB adjustment to fix the nonlinearity after 200°, and changes on the chassis fixation points level with emphasis on complexity and serviceability, to mitigate its shortcomings.

Furthermore, comprehensive thermal, NVH, and total cost-of-ownership analyses are the essential next steps to fully inform the development of next-generation EVs equipped with a high-performance IWM system.