Abstract
Vehicle axle vertical characteristics strongly affect ride comfort and handling, but the influence of individual components is not well quantified. It remains unclear to what extent individual components contribute to the stiffness, friction, and damping characteristics of a vehicle axle. This study aims to quantify the contribution of individual components to the axle characteristics, with particular emphasis on axle damping. To address this question, a MacPherson strut axle was tested on an axle test rig in 19 configurations. In addition to the intact axle, components were removed, deliberately degraded, or mechanically decoupled to prevent force transmission between selected components. The axle was excited quasi-statically and dynamically, using both single-sided excitation and harmonic in-phase excitation of both wheel carriers. For each axle configuration, stiffness, damping, and friction were evaluated. The results show that removing either the main springs or the anti-roll bar substantially reduced axle friction, indicating that these components introduce preload within the axle assembly that increases friction. In addition, a sand-contaminated lower ball joint significantly increased friction in the linear evaluation region. The damping analysis revealed that draining the shock absorbers of oil reduced axle damping by approximately 80% relative to the intact axle for all investigated excitations.
Introduction
The vertical characteristics of vehicle axles play a crucial role in both vehicle dynamics and ride comfort. 1 In the early phase of the development process, the key axle characteristics are optimized and defined in terms of characteristic parameters such as stiffness, damping, and friction, independent of the specific axle concept and geometry. 2 After the target axle characteristics have been defined, the axle concept and the component properties are specified. In this context, it is essential to understand how individual components influence the overall axle characteristics. Studies have shown, for example, that the shock absorber in a MacPherson strut axle accounts for approximately 70% of the axle friction. 3 Further investigations indicate that the overall axle spring rate is composed of contributions from the main spring, the auxiliary spring rate, and the gas stiffness of the shock absorber.4,5 Many studies, however, focus exclusively on individual axle components, such as ball joints and bump stops.6,7 As a result, it is often not entirely clear to what extent the individual component properties actually influence the overall axle characteristics. 3 Rather, studies show that the effect of individual components is strongly dependent on the system context. 8 This work investigates the influence of the individual components of a MacPherson strut axle on its characteristics.
A second motivation for this work relates to road safety. With Vision Zero, the European Union has set the goal of reducing the number of road traffic fatalities to zero. 9 One potential measure to reduce road traffic fatalities is the introduction of mandatory axle damping tests as part of periodic technical inspections.
Several studies have examined the effect of mandatory periodic technical inspections on road safety. A study conducted in Texas, United States, found that the share of fatal crashes involving vehicles with defects is significantly higher than the share involving vehicles without defects. 10 Another study from the United States showed that states with mandatory periodic technical inspections have, on average, 5.5% fewer traffic fatalities per 100,000 registered vehicles. 11 A further study reported that after periodic technical inspections were introduced in Turkey in 2007, the number of crashes decreased substantially in the following years. 12 To date, a mandatory axle damping test as part of periodic technical inspections, based on defined test methods, is required only in Belgium. 13 In other countries, only a visual inspection is performed in this area as part of periodic technical inspections.14,15
Multiple studies indicate that degraded axle damping increases the braking distance of passenger cars and adversely affects their lateral dynamic behavior.16–24 Many investigations focus on shock absorber degradation in these analyses.25–29 For shock absorbers, the most common degradation mechanism is oil and gas loss caused by leakage at the piston rod seal. 30 However, these studies raise the question of how large the shock absorber contribution to overall axle damping actually is. Is it sufficient to focus on shock absorber oil loss, or do other axle components contribute a similarly large share to axle damping?
The objective of this work is to quantify the contributions of the individual components of a strut axle to its stiffness, damping, and friction characteristics. The investigation was conducted on a strut axle because this axle concept is the most common in vehicles in Germany. 31
In the “Methods and Materials” section, the test axle and its investigated configurations are described, along with the test rig, measurement instrumentation, and defined excitation profiles. The subsequent sections present the evaluation and interpretation of the results. This is followed by a dedicated discussion of the findings. Finally, the key results are summarized, and the main conclusions are highlighted.
Methods and materials
To investigate the influence of individual components on the vertical characteristics of a strut axle, the front axle of a Volkswagen Golf VI was measured on an axle test rig using multiple axle configurations and several excitation types. Figure 1 shows the axle and the test rig. The axle is connected to the vehicle via eight mounting points and through the driveshafts. The eight mounting points are distributed as follows: two at the brackets, two at the subframe, two at the upper strut mounts, and two at the control arms. To geometrically define these mounting points, an identical axle was removed from a test vehicle and the locations were measured using a coordinate measuring device (Romer Absolute Arm, Hexagon). The axle was subsequently mounted to the test rig at these exact locations.
Axle test rig.
The two wheel carriers can be excited independently of each other in the three translational degrees of freedom and in the rotational degree of freedom about the vertical axis, both quasi-statically and dynamically, at frequencies of up to 30 Hz. For the investigations presented in this publication, the axle was excited in the vertical direction under displacement control. The displacement measurement system used has an accuracy of 0.01 mm. The force sensors used have an accuracy class of 0.1 and therefore a maximum error of 2 N. The measurement data were recorded at a sampling frequency of 500 Hz. During all measurements, the shock absorbers were cooled using a blower. The shock absorber temperature was monitored using temperature sensors mounted on the outer tubes of the shock absorbers. Table 1 summarizes the maximum excitation levels of the axle test rig. The maximum excitation velocity of the test rig in the vertical direction is 4 m/s.
Overview of the maximum excitation levels of the axle test rig.
Figure 2 shows all axle components individually. Table 2 provides an overview of the tested axle configurations. The different configurations were assigned to the components shown in Figure 2. After baseline testing of the intact axle, individual components were degraded while all remaining axle components were kept in an uncompromised condition. The degradation of the components was carried out on new components, which made it possible to measure the intact axle at the end of all measurements. After the final manipulated configuration had been measured, the axle was measured once more in the intact condition to obtain a final reference measurement.
Components of the MacPherson strut axle under investigation.
Overview of the tested axle configurations.
For the component degradation, only extreme conditions were considered, which are not encountered in real-world applications. Consequently, the results presented here should be interpreted as the maximum expected effect of component degradation on the axle characteristics. For configurations without a specific component like axle shaft (2), main spring (3), bump stop (4), and anti-roll bar (6), the component is simply uninstalled. To achieve the configuration with degraded shock absorbers (5), two holes were drilled into the shock absorber housings (Figure 3b) and the shock absorbers were fully drained. This eliminates both the stiffness contribution of the pressurized shock absorber and the hydraulic damping caused by the interaction of the hydraulic fluid with the shock absorber piston.
(a) Intact shock absorber, (b) manipulated shock absorber, and (c) tie rod with removed polymer bearing insert.
Ball joint degradation may manifest itself in two distinct ways. Damage to the bearing shells can introduce play, thereby interrupting force transmission within a certain displacement range. This condition, referred to as “no force at bearings” at the anti-roll bar link (7), tie rod end and inner bearing (9, 11), and lower ball joint (15), is reproduced by removing the polymer bearing insert, as illustrated in Figure 3(c).
Slipping out of the metal ball is prevented by crimping the edge of the bearing housing. The method proved to be efficient since no transmission of forces between the remaining metal parts of bearings was observed during testing.
The second degradation mode of the ball joints is characterized by increased roughness of the contact surfaces and, consequently, higher force transmission within the joint. This “high friction” condition at the anti-roll bar link (8), tie rod end and inner bearing (10, 12), and lower ball joint (16) was induced by removing the bearing seal, flushing out the lubricant with solvent, and introducing fine sand with a grain size of 0.063 to 0.2 mm until friction force saturation was reached. An example of a manipulated ball joint is illustrated in Figure 4(c).
Manipulation stages of the ball joint: (a) untreated condition, (b) degreased ball head, and (c) sand-treated bearing, showing the resulting change in ball stud surface quality.
The strut mount is degraded (13) by removing the lower strut mount component shown in Figure 5(b). The main springs are degraded (14) by cutting of its lowest winding as illustrated in Figure 5(c). Rubber bushings at the front and rear of the transverse link (g) (17, 18) and for the anti-roll bar (19) are manipulated to transmit no force by removing material at the perimeter of the bushings in order to introduce a gap to allow for mechanical play. By removing the elastomer, the transmission of forces and moments about the rotational and longitudinal degrees of freedom of the bushings is prevented. The method also proved to be effective, as no transmission of forces in these directions was observed at the manipulated bushings during testing.
(a) Strut mount, (b) lower strut mount component, and (c) manipulated main spring.
The axle’s stiffness, damping, and friction characteristics were investigated. Table 3 summarizes all measurements performed. Before each measurement, a conditioning procedure was performed. For the quasi-static measurements, the conditioning consisted of three repetitions of the respective quasi-static test. For the dynamic measurements, three repetitions of the corresponding dynamic test were conducted at the maximum amplitude of ±50 mm. The quasi-static tests were performed in two excitation modes. In the first mode, both wheel carriers were excited simultaneously in-phase. For the evaluation of stiffness and friction under simultaneous in-phase excitation of both wheel carriers, the results from the two sides were averaged. In the second mode, only the left wheel carrier was excited while the right wheel carrier was held fixed, and the force was measured at the left wheel carrier only. All quasi-static tests used a ramp input with a velocity of 1 mm/s. Figure 6 shows the force displacement characteristic measured at the left wheel carrier during simultaneous in-phase excitation of both wheel carriers for the intact axle. Axle stiffness and friction were evaluated in two regions of the characteristic. The first region was the linear range around zero displacement. For this purpose, axle stiffness and friction were evaluated in the rebound region over a displacement range from −45 mm to −1 mm and in the compression region over a displacement range from 1 mm to 27 mm. To characterize axle stiffness and friction in the linear region, the values from the rebound and compression branches were averaged. The second evaluation region is the progressive region of axle stiffness, which was evaluated over a displacement range from 35 mm to 45 mm. The progressive region is defined by the compression of the bump stop. Both evaluation regions are illustrated in Figure 6.
Summary of the measurements performed.
Force–displacement characteristic of the axle under quasi-static excitation, highlighting the evaluation regions used to determine stiffness and friction.
To determine the axle friction characteristics, half of the hysteresis width was calculated in each of the two evaluation regions. Axle friction depends on the excitation displacement amplitude. 32 Large displacements that extend into the progressive region of the axle response tend to result in higher axle friction values. Deubel proposed a method for determining axle friction that uses substantially smaller amplitudes of 3–5 mm. 3 However, the approach used here is suitable for a relative comparison of the individual configurations. In general, it should be noted that knowledge of axle friction is particularly important when investigating the effect of degraded axle damping on overall vehicle behavior. 33
To investigate the effect of individual components on axle damping, the axle was excited in-phase using harmonic inputs with amplitudes of 10, 30, and 50 mm and peak velocities of 0.1, 0.25, and 0.5 m/s. These values were selected on the basis of measurement runs with a similar test vehicle. The displacement of the front axle shock absorbers of a Volkswagen Passat B8 was measured, and the corresponding displacement velocities were determined during a drive at a constant speed of 30 km/h on a class D rough road section. The measurement runs are described in a separate publication by the author. 25 Figure 7 shows the measured shock absorber displacement and shock absorber displacement velocity from this measurement run. The figure indicates a maximum damper displacement of ±50 mm. Displacement velocities reach values of up to 1.4 m/s, although the majority remain below 0.6 m/s. On this basis, the displacement amplitudes and displacement velocities used for the axle tests were selected as a compromise, since the motion ratio between wheel displacement and shock absorber displacement in the MacPherson axles investigated is also close to 1.
Displacement and displacement velocities of the front axle shock absorbers of a VW Passat B8 test vehicle during passage over a class D rough road section at 30 km/h. 25
To determine damping, the hysteresis area of the force displacement curve for one oscillation, ΔW, was evaluated. The damping constant d was then calculated using Equation 1 based on the excitation frequency f and the amplitude
A comparison of the measurements obtained for the intact configurations at the beginning and at the end of the investigation revealed a maximum deviation of 1% for stiffness, 0.1% for friction, and 2.5% for damping.
Left-side force displacement curve under simultaneous in-phase harmonic excitation of both wheel carriers with a 50 mm amplitude and a peak velocity of 0.5 m/s, comparing the intact axle with configuration five featuring degraded shock absorbers.
Experimental Results
Figures 9 and 10 illustrate the effects of the different axle configurations on the vertical stiffness and friction of the axle for both in-phase excitation and single-sided excitation of the left side. In addition, Table 4 summarizes the absolute values of the measurements.
Effect of axle configuration on axle stiffness and friction in the linear and progressive evaluation regions under simultaneous quasi-static excitation of both wheel carriers; the values from the left and right sides were averaged.
Effect of axle configuration on axle stiffness and friction in the linear and progressive evaluation regions under quasi-static single-sided excitation of the left wheel carrier, with the right wheel carrier held fixed.
Absolute axle stiffness and axle friction in the linear and progressive evaluation regions under quasi-static two-wheel and single-wheel excitation.
Figure 9 shows the relative effect of the tested axle configurations on axle stiffness and axle friction in the linear and progressive evaluation regions for the simultaneous in-phase quasi-static excitation of both wheel carriers. The figure shows that removing the main spring strongly reduces both stiffness and friction in both evaluation regions. This result demonstrates that the suspension spring is the primary determinant of the vertical stiffness of the axle.
The next largest effect on stiffness in the linear region is observed for the configuration with the main spring shortened by one coil. This modification increases the linear axle stiffness by approximately 15%. The other investigated configurations show comparatively minor effects. It can be observed that configurations without the anti-roll bar, without shock absorber oil, without force transmission at the front transverse link bushing, and without force transmission at the anti-roll bar bushings each exhibit approximately 5% lower stiffness in the linear range. The reduction in stiffness for the configuration without the anti-roll bar can be attributed to the absence of the torsional stiffness provided by the anti-roll bar bushings. When the anti-roll bar bushings are manipulated, nearly the same axle stiffness is measured as in the configuration without the anti-roll bar. The anti-roll bar bushings can therefore be considered a secondary stiffness acting in parallel with the main spring.
When the shock absorbers contain no oil, they also lack gas preload. Both mono-tube and twin-tube shock absorbers in automotive applications typically include a pressurized gas chamber, which, among other functions, prevents oil cavitation at high excitation velocities. 34 The gas pressure typically ranges between 3 and 10 bar. 35 The results shown in Figure 9 indicate that the gas preload in the shock absorber also acts as a parallel spring on the vertical axle stiffness, contributing approximately 3% to the total axle stiffness. Manipulating the front transverse link bushing similarly reduces the vertical axle stiffness by approximately 3%. This bushing is arranged in series with the main spring; its modification can alter the preload conditions within the assembly and thereby influence the overall axle stiffness.
In the progressive stiffness region, the configuration without the main spring again exhibits a substantially lower stiffness. The configuration without the bump stop also exhibits a markedly reduced stiffness compared with the baseline configuration. This shows that the progressive region is significantly defined by the bump stop.
A substantially higher progressive stiffness is observed for the variant in which the rear transverse link bushing does not transmit force. For this configuration, progressive axle stiffness increases by about 40%. More detailed analysis of the stiffness curves shows that the manipulated configuration has lower stiffness in the linear region and, in the progressive region at maximum displacement, reaches almost the same maximum force as the standard variant. The stiffness in the progressive region is therefore significantly higher. The likely cause of this behavior is that the play in the manipulated rear transverse link bushing reduces the linear stiffness of the axle, since the bushing acts in series with the spring strut. At larger displacements, contact in the bushing is probably reestablished, after which the entire axle shows greater stiffness for these larger displacements.
With respect to friction, removing the main spring markedly reduces axle friction in both the linear and progressive regions by up to 25%. This is likely explained by the lack of preload between the various axle components when the main spring is absent. 3 The MacPherson axle is designed such that the shock absorber always transmits part of the horizontal axle forces. Increased horizontal forces lead to increased shock absorber friction. The main spring is already preloaded in the axle assembly, so that a horizontal force continuously acts on the shock absorber, which increases friction. If the spring preload is absent, the horizontal preloading of the shock absorber is also reduced, and axle friction decreases. Likewise, removing the main spring prevents the spring path of the top mount from being compressed, so that no hysteretic behavior is observed, which would normally add to the overall axle hysteresis.
By contrast, the configuration with a reduced main spring length shows that axle friction in the linear range increases by up to 15%. As described above, shortening the main spring leads to higher vertical axle stiffness, which simultaneously increases the forces acting within the same evaluation range. These higher forces cause increased preloading of other components and thus higher axle friction.
A similar effect is observed for the configuration without the anti-roll bar, for the configuration in which no force is transmitted through the anti-roll bar link, and for the configuration with modified anti-roll bar bushings. All three configurations reduce axle friction by approximately 10%. It is likely that these configurations prevent torsional deflection of the anti-roll bar bushings, so that no hysteresis develops in these bushings, which would otherwise add to the measured overall hysteresis. This indicates that the anti-roll bar is responsible for approximately 10% of the axle friction in the linear range.
Furthermore, it can be observed that the configuration without shock absorber oil reduces axle friction by approximately 5%. When the shock absorber contains no oil, its measured friction force decreases as well. 25 This reduction in shock absorber friction is also reflected at the axle level.
A pronounced increase in friction in the linear evaluation region is caused by the manipulated lower ball joint. For this configuration, axle friction is found to increase by approximately 15%.
The configurations with sand-contaminated tie rod end and anti-roll bar link bushings each increase axle friction by approximately 10% in the linear range. These results suggest that all of these joints influence axle friction, although the lower ball joint appears to have the greatest effect among them.
Friction in the progressive region of the characteristics is also strongly reduced for the configuration without the bump stop, in addition to the configuration without the main spring.
The reduction in friction observed for the configuration without the main spring has the same cause in the progressive region as in the linear region of the axle characteristic. Table 4 shows that the measured axle friction in the progressive range is significantly higher than in the linear range, whereas the configuration without the bump stop yields values similar to those of the baseline measurement in the linear range. This suggests that friction in the progressive region is determined by the properties of the bump stop. The other configurations show effects similar to those observed in the linear evaluation range, with only a few exceptions. In particular, the friction in the progressive region is reduced by approximately 5% to 10% for the configurations with a degraded strut mount and without force transmission at the ball joint.
Figure 10 shows the relative effect of the tested axle configurations on axle stiffness and axle friction in the linear and progressive evaluation regions for quasi-static single-sided excitation of the left wheel carrier, with the right wheel carrier held fixed. The figure indicates that, under single-sided excitation, removing the main spring and removing the anti-roll bar have the largest effects on the evaluated quantities. When either of these components is removed, the measured stiffness and friction decrease markedly. The largest increase in linear stiffness is observed when one coil is removed from the main spring. The linear stiffness of the axle is reduced most noticeably when the elastomer in the anti-roll bar bushing is removed. For this configuration, the linear axle stiffness under single-sided excitation decreases by approximately 15%. When the anti-roll bar link joints are modified, the linear axle stiffness under single-sided excitation decreases by approximately 6%. Both configurations confirm that the anti-roll bar effect has a very strong influence on axle stiffness during single-sided jounce.
In the progressive region, the configuration with the modified rear transverse link bushing increases axle stiffness by approximately 30%. The cause of this effect is the same as for the two-sided excitation of the axle, which has already been discussed.
For the configurations with modified tie rod joints without force transmission, as well as with increased friction in the inner tie rod joint, the stiffness in the progressive region increases by approximately 10% in each case.
In the linear region, axle friction increases by about 15% for the configurations with increased friction in the anti-roll bar link and in the lower ball joint. The configurations with increased friction in the tie rod end joints also show a significant effect, increasing friction by approximately 10%.
In the progressive region, the variant in which no force is transmitted through the rear transverse link bushing increases axle friction by approximately 18 %. The configurations with increased friction in the anti-roll bar link and in the tie rod end joints increase friction in the progressive region by about 10%.
Figure 11 summarizes the damping constants d for all tested axle configurations under simultaneous in-phase harmonic excitation of both wheel carriers with an amplitude of 50 mm and a peak velocity of 0.5 m/s. The figure clearly shows that the configuration with shock absorbers drained of oil exhibits substantially lower damping than all other tested configurations. This effect is also evident in Figure 8, where the corresponding hysteresis area is markedly reduced.
Damping constants for all axle configurations under simultaneous in-phase harmonic excitation of both wheel carriers with an amplitude of 50 mm and a peak velocity of 0.5 m/s.
To assess the generality of this finding, Figure 12 summarizes the effect of each configuration on axle damping across all dynamic excitation inputs investigated. For each configuration and each excitation condition listed in Table 3, the calculated damping was normalized by the corresponding value of the intact axle under the same excitation. The results are presented as box plots, where the central line indicates the median of the values and the upper and lower edges of the box represent the 75th and 25th percentiles, respectively. The whiskers indicate the extreme values, and the crosses mark the outliers.
Relative effect of each axle configuration, normalized to the intact axle, for all measured harmonic in-phase tests with simultaneous excitation of both wheel carriers (Table 3).
Across all excitation conditions, the configuration with shock absorbers drained of oil exhibits the largest reduction in axle damping. Axle damping decreases by at least 78%, with a median reduction of 83%. The influence of the shock absorber on axle damping becomes more pronounced as the peak excitation velocity increases, since this also increases the shock absorber operating velocity.
The configurations with no force at the outer tie rod end bearing, no force at the lower ball joint, high friction at the lower ball joint, no force at the front transverse link bushing, and no force at the anti-roll bar bushing also show an influence on the damping coefficient of approximately 30% for individual excitation conditions. However, these effects occur especially for excitation signals with small amplitudes of around 10 mm and lower maximum excitation velocities of 0.1 m/s. For larger excitation amplitudes and higher maximum excitation velocities, the deviations in damping measures between the modified configurations and the baseline configuration are very small and remain in the low single-digit percentage range. The axle configuration without shock absorber oil is therefore the only configuration that exhibits a clearly reduced axle damping for all excitation signals, and particularly for excitations with large amplitudes and high maximum excitation velocities.
Discussion
The results first reveal several expected effects of individual components on the investigated axle characteristics. The main spring provides the dominant contribution to axle stiffness in both the linear and progressive evaluation regions, for both simultaneous excitation of both wheel carriers and single-sided excitation. Under single-sided excitation, the anti-roll bar has the largest influence on stiffness in both evaluation regions. It becomes apparent that other components also contribute, to a lesser extent, to defining the vertical stiffness of the axle. In particular, the gas pre-charge of the shock absorbers and the anti-roll bar bushing act like springs connected in parallel with the main spring.
Furthermore, the results show that removing the main spring or removing the anti-roll bar substantially reduces axle friction in both evaluation regions and for both excitation types. This implies that the main spring and the anti-roll bar introduce preload within the axle assembly and thereby increase axle friction.
Regarding the anti-roll bar, it can also be observed that the torsional deflection of the anti-roll bar bushing accounts for up to 10% of axle friction in both the linear and progressive regions of the characteristic curve.
Furthermore, the lower ball joint appears to have the largest effect on axle friction in the linear range after the shock absorber. In the progressive region, the configuration in which no force is transmitted through the rear transverse link bushing produces the largest increase in axle friction, while removing the bump stop leads to the largest reduction in the measured friction in the progressive region.
The damping analysis indicates that the hydraulic oil in the shock absorbers is the primary determinant of axle damping. When the shock absorber oil is removed, axle damping is reduced by approximately 80%. Within the investigated excitation range, the other configurations did not produce damping changes comparable to the oil-loss case. Excitations of larger amplitudes can be considered safety-relevant. Reduced axle damping leads to increased vibration amplitudes of the sprung and unsprung masses around their natural frequencies. Increased vibrations of the unsprung masses result in larger wheel load variations, which reduce the average transmissible horizontal tire forces. 23 This, in turn, can lead to unstable vehicle behavior and longer braking distances, particularly on uneven roads.16,24
It can be shown that the residual damping of the axle without the fluid force of the shock absorber oil accounts for approximately 20% of the axle’s original dissipated energy for the investigated dynamic excitations. The residual damping most likely results from the superposition of several small dissipation mechanisms, such as friction in the shock absorber assembly, hysteresis in elastomer bushings, and dry-contact interactions in the axle system.
With regard to realistic degradation states, it should be noted that a completely oil-empty shock absorber is highly unrealistic. Studies show that the work performed by a shock absorber is strongly nonlinear with respect to its oil fill level. It has been observed that the dissipated work of a shock absorber decreases stepwise as the oil level drops, and that this work still depends strongly on the shock absorber excitation. Shock absorbers with reduced oil fill are therefore particularly sensitive to long-duration excitations with higher excitation velocities above 1 m/s. At these higher excitation velocities, foaming and cavitation are highly likely. Under such conditions, the dissipated work of a shock absorber with 70% oil fill can already decrease to 30% of its initial value. If the shock absorber still contains 40% of its oil, these excitations can reduce the dissipated work to nearly zero. The results of this study show that the fluid force of the shock absorber accounts for approximately 80% of axle damping. Combined with the component-level investigations, it can therefore be concluded that an oil loss of approximately 30% in both shock absorbers of an axle can reduce axle damping to about 45% of the intact axle’s original damping under large-amplitude excitations and high frequencies. By contrast, the other investigated configurations do not have a comparably large effect on axle damping.
The investigations presented in this work were carried out on a MacPherson axle. Among other features, the MacPherson axle is characterized by a high motion ratio between wheel travel and shock absorber travel, as well as by the fact that a large proportion of the horizontal wheel forces is supported by the shock absorbers. Other axle concepts do not exhibit these properties. Multi-link axles, for example, are designed such that the shock absorbers transmit as little horizontal force as possible. 36 This generally leads to lower shock absorber friction and axle friction, and thus to lower residual damping. 37 Overall, it can therefore be assumed that, for a large number of different axle concepts, the shock absorbers account for the major share of axle damping. To support this hypothesis, however, similar bench tests should be repeated for other axle concepts.
Conclusion
The objective of this study was to investigate how individual components affect the relevant vertical characteristics of a MacPherson strut axle. A particular focus was placed on identifying which axle components make a meaningful contribution to axle damping.
To this end, the front MacPherson axle of a Volkswagen Golf VI was tested in 19 different states. In addition to the intact axle, individual components such as the main spring and the anti-roll bar were removed. Other components were deliberately degraded, for example by draining the shock absorber oil or by increasing friction in ball joints by introducing abrasive particles. In addition, force transmission at selected connections was eliminated.
The axle was excited on an axle test rig under quasi-static and dynamic conditions using both single-sided excitation and simultaneous in-phase excitation of both wheel carriers. Stiffness and friction were evaluated in the linear and progressive regions of the axle force–displacement characteristic. Axle damping was quantified from the hysteresis of the force–displacement loop.
The results show that the main springs and the anti-roll bar dominate axle stiffness. In addition, removing either the main springs or the anti-roll bar leads to a significant reduction in axle friction. This observation can be explained by the loss of preload within the axle assembly that is otherwise introduced by the springs and the anti-roll bar. Furthermore, the lower ball joint was found to significantly influence axle friction in the linear evaluation region. In the progressive region, the configuration in which no force is transmitted through the rear transverse link bushing produced the largest increase in axle friction.
The damping analysis revealed that the configuration with shock absorbers drained of oil exhibited an approximately 80% reduction in axle damping across all investigated excitations. The other investigated configurations showed an effect on axle damping only for excitations with small amplitudes and low maximum excitation velocities. It can therefore be concluded that the shock absorbers are the dominant contributors to axle damping within the investigated MacPherson axle and excitation range. None of the other axle manipulations produced a significant effect on axle damping across all investigated excitations. The investigations in this work were carried out on a MacPherson axle, which is the most widely used axle concept. It can be assumed that dampers are also the dominant component of axle damping in other axle concepts. However, to support this hypothesis, similar investigations should also be carried out on other axle concepts.
Footnotes
Acknowledgements
The authors acknowledge the use of ChatGPT (OpenAI) for assistance with the translation and language refinement of the manuscript; all scientific content and interpretations remain the responsibility of the authors.
Handling editor: Chenhui Liang
Funding
The authors received no financial support for the research, authorship, and/or publication of this article.
Declaration of conflicting interests
The authors declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.