Additively Manufactured Wheel Suspension System with Integrated Conductions

The additive manufacturing process selective laser melting (SLM) is chosen for the production of the chassis components. The reason for this is the good availability of the equipment, acceptable costs, and high component quality that can be achieved. In SLM, a thin layer of powder is applied to a build platform and melted locally by a laser, according to the layer information from the GCode. The build platform then moves down by one layer thickness in the z-direction, and another powder layer is applied. As the layer information is imprinted into the new powder layer, cohesion with the previous layer is simultaneously established. This process is repeated until a full component height is reached. After powder metallurgical production, the components are cleaned. If necessary, functional surfaces can be reworked by machining, and any support structures are removed. This results in components whose properties are comparable to those of the forged components [4]. In order to reduce residual stresses and to obtain almost completely dense components, thermal post-treatment can be used in conjunction with hot isostatic pressing [4]. Table 1 gives an overview of the constructive constraints of SLM.

Table 2 summarises the geometric boundary conditions. The connection points to the subframe (Figure 2), the position of the wheel hub motor, and the connection point of the shock absorber to the lower wishbone are already given at the time of construction of the double-wishbone front axle presented.

The arrangement of the chassis components takes into account that the intended bolted suspension joints have a significantly higher load capacity in the compression direction compared to the tension direction [6]. The available installation spaces for the individual components are initially determined based on the requirements for the kinematics to be mapped and the steering angles to be realised. However, the results of the topology optimisation are continuously analysed during the course of the development work, and the findings on the sub-areas critical for the component rigidity are thus incorporated into the installation space specification. Figure 3 shows the final design spaces available for the upper control arm, lower control arm, and wheel carrier components.

First, the necessary load capacity of the wheel suspension is defined. In the automotive industry, the strength design of suspension components is commonly based on standard load cases (Table 3). The direction of acceleration refers to the transmission of force from the vehicle to the road. These are derived from driving manoeuvres and can be used for quasi-static FEA. The presented load cases 1 to 10 are used for quasi-static strength design, and the load cases 11 to 16 are used to verify the fatigue strength of chassis components. Since the load cases are expressed in terms of wheel accelerations, they can be applied to a wide range of vehicles [3]. The load cases Vertical impact (2), Longitudinal impact (3), and Lateral impact (4) represent high quasi-static loads on the chassis components in all spatial directions. The U-Shift prototype is intended for a low driving speed, but the selected load cases can also potentially occur at low speeds (attenuated). As the wheel accelerations represent limit cases that only rarely occur, the yield strength is used below as the material parameter for the mechanical design of the components. A check of the fatigue strength of the components is recommended when the components are pursued further.

Multiplying the wheel accelerations by the wheel load provides the forces at the wheel contact point acting on the vehicle. While 40% of the vehicle mass is on the front axle when the driveboard is empty, the proportion of the total mass on the front axle is reduced to 30% when the vehicle is fully loaded with a passenger capsule (Table 4).

The maximum wheel load on the front wheel is 750 kg, according to Table 4. Since the unsprung masses are directly supported by the road surface, this can be neglected in the strength design of the double-wishbone front axle. The weight of the wheels, wheel hub motors, and other components proportionally belonging to the unsprung masses is 150 kg per wheel, which is a good approximation. Accordingly, the relevant proportion of the wheel load is set to 600 kg for the design of the chassis components. This assumption results in forces acting on the centre-wheel contact point, as shown in Table 5.

Based on the available installation space shown in Figure 3, topology optimisation of the lower wishbone was carried out in HyperWorks 2022. Since the lower wishbone is a component that is subjected to high bending stress, the light metal material Ti6Al4V is selected. This material combines high strength with low density and good corrosion resistance. Alternatively, stainless steel and nickel-based alloys are also suitable for the highly stressed component. Due to the high modulus of elasticity of the steels mentioned, a higher component stiffness can be achieved with the same available installation space. However, the simulation shows that the component mass is minimised by using Ti6Al4V, and the resulting displacements are acceptable. In order to allow the highest possible degree of plastic deformation, a thermal post-treatment of the wishbone is performed. Ti6Al4V has a yield strength of ≥970 MPa and a yield strength of ≥14% in all spatial directions after a two-hour heat treatment at 800 °C in a vacuum (Table 6). An additional post-treatment with HIP is helpful for increasing the fatigue strength of the components [4,7]. However, since the vehicle under consideration is a demonstrator designed for low mileage, this additional effort is not necessary.

The chassis components are optimised individually by determining the forces acting on the components using force and moment equilibria and applying them to the individual components via rigid body elements [8]. This procedure offers significant savings in computing time, as the individual components can be optimised with the help of a linear static model. The required safety against plastic deformation is already covered by the high requirements of the standard load cases. In order to sufficiently consider simplifications of the linear static optimisation setup of the individual components, a permissible maximum stress of 700 MPa is defined. This corresponds to a safety factor of 1.39, which exceeds the safety factor in the standard load cases. A final validation is carried out by a subsequent simulation of the overall model. The lower wishbone is used to guide the electrical conductors, which are necessary for the operation of the wheel hub motor. By optimising the component several times with different MINDIM settings (minimum member size), a component geometry is achieved that allows the conductors to pass through sensibly. This takes into account the recommendation from Altair Inspire, Inc. to select 3 to 12 times the average element size for the MINDIM setting. It is also possible to make MAXDIM (maximum member size) settings to further manipulate the resulting geometry. However, this is not done because it usually leads to poorer material utilisation [9]. The parameters used for the optimisation are summarised in Table 7.

Figure 4 shows the result of the topology optimisation. Due to the great design freedom of the AM, the force-flow-compatible component shape does not have to be reduced to simpler geometries as in conventional manufacturing, but only the design specifications discussed in the introduction have to be taken into account. In the following, the blue-coloured component areas represent the non-design volumes.

The iso-surfaces of the design space are fed back into a common CAD (computer-aided design) format using the PolyNURBS tool integrated with Altair Inspire. The non-design areas are added manually. The final component is shown in Figure 5.

The steering knuckle is exposed to high bending and torsional stresses. The component must be designed in such a way that no collisions occur between neighbouring components at the high steering angles required. Since the track width is limited and the position of the spreading axis is predetermined, a small component depth in the y-direction is available. Consequently, high-strength materials with the highest possible stiffness are also required. In analogy to the lower wishbone, it is checked whether the stiffness of Ti6Al4V is sufficient or whether the use of steel is required. Since the deformations of the connection points are acceptable when Ti6Al4V is chosen, and a low component mass is achieved by using this light metal, the titanium alloy is selected. The ductility of this material has already been discussed in the optimisation of the lower wishbone. The feed-throughs for information, material, and electrical cables are already planned in the design space. A recess is also required to allow the thread of the lower suspension joint to be recut. The optimisation of the topology focuses on improving the utilisation of the material. The parameters used for the optimisation are listed in Table 8.

As before, the optimised geometry (Figure 6) is fed back into a CAD model using a PolyNURBS tool (Figure 7).

The connections for the cooling water supply as well as the hydraulic connection of the brake system integrated in the wheel hub motor, are located at the upper part of the wheel hub motor. In order to avoid a collision between the connection hoses and the spring-damper system, the above-mentioned lines are integrated into sections into the upper wishbone. As the upper wishbone only has to support forces in the x- and y-directions, there is a significantly lower strength requirement for the material used than for the lower wishbone. Due to the high lightweight potential of aluminium alloys, the material AlSi10Mg, which is commonly used in powder metallurgy, was selected. This material has a high corrosion resistance, which is essential for contact with cooling water and brake fluid. In addition, the good thermal conductivity of aluminium provides a cooling effect on the fluids flowing through it. The disadvantage of the aluminium alloy is its low elongation at break in the as-built state. However, T6 heat treatment can increase the elongation at break to 10%, with a yield strength of ≥230 MPa (Table 9) [10].

For a force-flow-compatible design of the component, topology optimisation of the design space is carried out as with the previously discussed components. At this point, the positions of the material-conducting lines remain undetermined. To minimise the component mass, the position of the lines is adapted to suitable force-flow paths. The maximum stress is initially set to 100 MPa, so that the topology optimisation already results in wide struts for subsequent line integration (Table 10).

As expected, an arrangement of tension and compression bars is suitable for transferring the imposed loads to the subframe. At the point of force application from the suspension joint to the wishbone, a stiffener is needed to support the bending moments (Figure 8).

It is analysed whether this optimised shape of the wishbone allows for a reasonable arrangement of material-carrying lines. Since a suitable line configuration is possible, the isosurfaces from the optimisation are manually fed back into the CAD format in the PolyNURBS tool. A CAD environment is then used to model the lines in the component. In order to take into account the influence of the fluids flowing through, suitable fluid pressures must be specified for a design. The cooling water circuit is operated at a low pressure of well below 10 bar. To take into account an average safety factor of 1.5, a pressure of 15 bar is simulated on the cooling water channels. A maximum brake system working pressure of approximately 160 bar is often used to estimate deformations when designing brake systems. Pressure relief valves usually limit the system pressure to 204 bar [11]. To prove the strength of the wishbone, a pressure of 240 bar is applied to the brake line. This also corresponds to a safety factor of 1.5, which is also applied to this safety-relevant line, as further safety factors are included in the standard load cases and are also required in the validation of the overall model. This load spectrum is used to identify the weak points of the wishbone and to optimise it again manually. The final upper wishbone is shown in Figure 9.

To save computing time, linear static models are used to optimise the individual components. The interfaces with the other components are simplified in this simulation. In addition, a change in the lever arms of the applied wheel forces due to large deformations of the components is not taken into account. In order to verify the structural strength of the entire system, it is necessary to use a non-linear quasi-static simulation model for large displacements. This also allows the modelling of component contacts. For the validation, a neutral wheel position is first used (Figure 10).

Since only limited computing capacities are available, the components are divided into tetrahedral elements with a linear approach function (TETRA4). In order to achieve usable results despite the small number of nodes per element, an average element size of 1.5 mm is selected, which is small compared to the component dimensions. The overall model is discretised with a total of 4.15 million TETRA4 elements. The contact between the lower ball stud and the lower wishbone is modelled as a friction slide contact. This allows the lower wishbone to be pushed onto the ball stud, resulting in a radial preload in addition to the axial preload. The procedure for the ball stud of the upper suspension joint is analogous. The contact between the ball studs and the bearing sockets is also modelled as movable or frictional. The contacts of the screw heads to the adjacent components are modelled as rigid (no relative movement) to save calculation time since no relevant relative movement occurs if the screws are sufficiently pretensioned. The model contains a total of 21,953 surface-to-surface elements that represent the force transmission between the components. The shock absorber and track rod are substituted by rigid body elements, since the deformations of the AM components are to be investigated separately. For a realistic representation, however, the screwed connection elements are modelled. The point of application of the tyre forces from the standard load cases is represented by rigid body elements. These engage with a flange, which is used for a realistic connection of the wheel hub motor to the steering knuckle. In order to be able to represent the isotropic fluid pressures on the upper wishbone, dummy plugs are constructed for the cooling water and brake lines. In the first load case, a preload force is applied to the bolts and tapered connections. For this purpose, one-dimensional beam elements (1D-beams) are modelled (Figure 11). The choice of the preload force is based on empirical values and is selected iteratively according to the results from the simulation at 25 kN to 70 kN, depending on the connecting element [12]. In another load case, hydraulic pressure is applied to the upper wishbone. Based on the resulting displacements and stresses, the selected standard load cases are then simulated.

Based on the results of the simulation, final weak points and singularities can be identified, and then the stress peaks can be reduced constructively. The global displacements are shown in Figure 12, Figure 13 and Figure 14 in the form of false-colour images. The fluid pressures are taken into account in the visualised deformations. The highest displacements occur in the standard load case (3) and longitudinal impact. This can be explained by the large lever arm in the y-direction due to the depth of the wheel hub motor. The large lever arm increases the effect of the tyre forces in the x-direction that occurs in the mentioned load case.

While the deformations of the wishbones are sufficiently small, large deformations of the steering knuckle occur, especially in the standard load case of longitudinal impact.

Figure 12 shows that these are areas of the components that are distant from the connection points. Consequently, these higher deformations are not passed on to the wheel and are therefore uncritical. The resulting deformation at the flange towards the wheel hub motor fulfils the quality criterion for the component stiffness defined in the introduction.

The uniaxial equivalent stress, according to Mises, is used to verify the structural strengths of the components. The quotient of the yield strength of the materials used and the occurring equivalent stress results in a degree of safety against plastic deformation, which goes beyond the safety contained in the load collectives. This safety covers, for example, simplifications in the FEM simulation and is, therefore, continuously marked S_FEM.

The maximum equivalent stress occurring in the steering knuckle, according to Mises, is 730.3 MPa (S_FEM = 1.33), and the maximum stress in the lower wishbone is 856.8 MPa (S_FEM = 1.13). These stresses occur in the standard longitudinal impact load case. Although these appear to be low safety factors, the design is still rated as load-bearing. The reason for this is that the longitudinal impact load case, which is particularly demanding for the chassis components, corresponds to a high acceleration of 2.5 g in the x-direction. As this general load case is designed for vehicles with a usually higher maximum driving speed, lower maximum accelerations can also tend to be expected at lower possible driving speeds. Thus, by conservatively retaining the standard load case, a higher level of safety is actually taken into account than the safety factor suggests. In order to support the assumptions made, it is recommended that the possible wheel accelerations on the real vehicle be recorded by sensors. The upper wishbone, on the other hand, is subjected to the greatest stress in the lateral impact. Here, it experiences a comparative stress of 123.8 MPa. (S_FEM = 1.85, Table 11). Simulated stresses for all components are below the yield strength of the materials used, and a small safety factor against simplifications is taken into account in the simulation so the components meet all strength requirements.

The selection of suspension joints is based on the loads derived from the force and moment balances, which are also used to optimise the single components. The selection of the joints is additionally validated by the FEM simulation of the entire model. The ball stud is identified as a critical component of the suspension joints. It is made of precipitation-hardenable stainless steel with a yield strength of 1550 MPa. The component experiences the highest stress in the standard load case longitudinal impact. Here, a comparative stress according to Mises of 1081 MPa occurs, resulting in a safety factor S_FEM of 1.43 (Figure 15). The exact stress values must be critically scrutinised, as there is always a deviation from the true stresses due to simplifications in the approach functions, the discretisation, and the numerical solution of differential equations in the FEA. In order to validate the simulation results, the aim is to investigate the convergence of the stress values in more detail as the project progresses.