DESC:FIELD
The present disclosure relates to a front suspension assembly of a vehicle. More specifically, the present disclosure relates to a front suspension architecture with lower control arms for a sports utility vehicle (SUV).
DEFINITION
Independent Lower Control Arm (iLCA) or intelligent link (iLINK): the term iLCA or iLINK stands for independent lower control arm or intelligent link. It refers to a specific type of suspension component used in vehicles, particularly in the context of front suspension systems. The iLCA configuration is part of a split-link architecture, where each split lower control arms are configured to function independently for each wheel, which provides greater flexibility and control over the wheel's movement during various driving conditions.
McPherson strut suspension architecture: A McPherson strut suspension is an independent suspension system that combines a coil spring, damper, and structural strut into a single unit. It features at least one lower control arm that connects the steering knuckle to the vehicle subframe.
Kingpin axis: The kingpin axis is an imaginary axis that passes through the upper and lower pivot points of a vehicle's steering system, about which the wheel assembly pivots during steering. In the front suspension assembly, particularly in a McPherson strut configuration, the virtual kingpin axis is dynamically defined by the mounting interfaces of the strut, steering knuckle, and lower control arms.
Instantaneous Center: The instantaneous center (IC), also known as the instant center of rotation, is a dynamically shifting point of rotation at which a link or component in a mechanical system experiences zero velocity at a given instant. In the context of vehicle suspension and steering systems, the instantaneous center is the point around which a wheel or linkage effectively rotates at a specific moment. The position of the instantaneous center varies dynamically based on the movement of suspension components, which influences parameters such as wheel camber, steering geometry, and overall vehicle handling characteristics.
BACKGROUND
The background information herein below relates to the present disclosure but is not necessarily prior art.
The front suspension assembly of a vehicle plays a pivotal role in determining the ride comfort, handling performance, and overall manoeuvrability. A key component of this system is the lower control arm (LCA), which connects the wheel assembly to the vehicle’s chassis or subframe. Over the years, various configurations have been developed for LCAs to improve the performance of the suspension system, such as stability, load distribution, structural durability, and packaging efficiency. However, the conventional configuration of the LCA has significant limitations in terms of construction, assembly, and performance.
Generally, a McPherson strut suspension system is configured with a single lower control arm (LCA), which is constructed by welding sheet metal panels together. One end of the LCA is mounted to the front subframe (FSF) or chassis using two rubber bushes for articulation, while the other end is fitted to the knuckle of the wheel assembly via a ball joint. The ball joint allows pivoting motion, to facilitate suspension articulation during dynamic conditions.
While the McPherson strut suspension system with the single LCA is simple and cost-effective, it presents significant limitations in terms of ride and handling due to the constrained flexibility in positioning the hardpoints of the LCA. The ball joint placement and the rearward positioning of the steering tie rod on the FSF contribute to suboptimal steering behaviour, particularly toe-in during cornering, leading to undesirable oversteer tendencies. Further, the turning circle diameter (TCD) of vehicles using the conventional single LCA configuration, with a wheelbase of approximately 3,000 mm, is around 12 meters, which compromises vehicle manoeuvrability, especially in urban or tight spaces.
Additionally, in some vehicles, the weight distribution is more biased toward the rear, which affects the understeer gradient. The understeer gradient is a measure of how much the vehicle resists turning when subjected to a steering input. When the understeer gradient is not achieved, the vehicle tends to oversteer, causing it to steer more than the input given by the driver. To compensate for this inherent, oversteer tendency, the steering rack is positioned at the front of the vehicle's front wheel center. However, this positioning reduces Ackermann geometry, which negatively affects steering precision and stability during cornering.
To address the ride and handling limitations of the conventional single LCA configuration, a double wishbone suspension system was developed, which is characterized by a split LCA construction. The split LCA configuration utilizes four links, two links at the top and two links at the bottom to form the front suspension assembly, wherein the comfort link is positioned rearward, and the handling link is placed frontward.
While split LCA configuration enhances ride and handling characteristics by distributing the suspension loads across multiple links, such kind of four-link construction is complex and requires additional packaging space, whilst the loads from the suspension articulation due to road undulations are taken up and shared by these four links. Hence, such kind of the conventional four-link double wishbone type of construction is costly, complex in construction, and assembly as well. Further, the packaging requirements for these four links demand larger space, and thus, it tends to increase assembly time and manufacturing costs. Therefore, the vehicles utilizing the conventional split LCA or the double wishbone suspension configuration with a wheelbase of about 2.9m, exhibit a TCD of approximately 12.2 meters, which further exacerbates manoeuvrability challenges as compared to the single LCA configuration.
Further, to overcome the drawbacks of the above two conventional configurations. i.e., the McPherson strut configuration and the double-wishbone configuration of the LCA, a hybrid approach was developed, which employs a split two-link LCA configuration in a McPherson strut suspension system. In this configuration, a comfort link is positioned frontward, and a handling link is placed rearward. The comfort link placed forward might negatively impact front end crash load path. This will deteriorate crash safety of the vehicle. While the above mentioned split two-link LCA configuration offers a compact suspension configuration, however, it has several drawbacks. The integration of ball joints in the links increases overall costs and renders them non-serviceable, which mandates entire link replacement in the event of ball joint failure. The vehicles utilizing the split two-link LCA configuration with a wheelbase of about 2.8m, exhibit a TCD of around 12 meters, which limits manoeuvrability and fails to achieve the desired improvement over the other configurations.
Overall, the conventional configuration of the LCA exhibits several drawbacks, including limited flexibility in hardpoint positioning, inefficient packaging, reduced structural durability, comparatively high TCD, and suboptimal NVH performance.
Therefore, there is a need for a front suspension assembly of a vehicle that alleviates the aforementioned drawbacks.
OBJECTS
Some of the objects of the present disclosure, which at least one embodiment herein satisfies, are as follows:
It is an object of the present disclosure to ameliorate one or more problems of the prior art or to at least provide a useful alternative.
An object of the present disclosure is to provide a front suspension assembly of a vehicle with lower control arms that optimizes hardpoint placement.
Another object of the present disclosure is to provide a front suspension assembly with lower control arms that minimizes the turning circle diameter (TCD).
Still another object of the present disclosure is to provide a front suspension assembly that enhances structural durability.
Yet another object of the present disclosure is to provide a front suspension assembly that improves packaging efficiency.
Still another object of the present disclosure is to provide a front suspension assembly that reduces NVH (Noise, Vibration, and Harshness) levels and improves performance.
Yet another object of the present disclosure is to provide a front suspension assembly that provides serviceable components.
Still another object of the present disclosure is to provide a front suspension assembly that simplifies manufacturing and assembly.
Yet another object of the present disclosure is to provide a front suspension assembly that improves handling dynamics.
Still another object of the present disclosure is to provide a front suspension assembly that enhances component load distribution.
Other objects and advantages of the present disclosure will be more apparent from the following description, which is not intended to limit the scope of the present disclosure.
SUMMARY
The present disclosure envisages a front suspension assembly of a vehicle. The vehicle comprises a front subframe (FSF) mounted on a vehicle chassis, and a steering rack positioned at the front side (F) of the vehicle and mounted on the front subframe. The steering rack is configured to convert a steering input in lateral movement of tie rods, wherein the suspension assembly is configured to optimize steering geometry and reduce the turning circle diameter (TCD). The front suspension assembly comprises a pair of steering knuckles, a pair of telescopic struts, and a split lower control arm configuration. Each of the steering knuckles is configured to be pivotally coupled to a respective front wheel and operatively connected to the steering rack via the tie rods. Each of the steering knuckles defines a virtual kingpin axis about which the steering knuckles pivot. Each of the telescopic struts has an operative upper end configured to be coupled with an operative section of the vehicle structure and an operative lower end configured to be mounted on a predefined mounting interface of the corresponding steering knuckle. The split lower control arm configuration has a pair of first link arms and a pair of second link arms. The split lower control arm is configured to be pivotally mounted between the front subframe and the corresponding steering knuckles. Each of the first link arms and each of the second link arms defines a control arm axis. An instantaneous centre is defined by the intersection of each of the control arm axes with each of the kingpin axes, which dynamically shifts in response to the steering input to alter the steering geometry and reduce the turning circle diameter (TCD).
In an embodiment, each of the steering knuckles includes first knuckle hardpoints and second knuckle hardpoints. First end of each of the first link arms are configured to be pivotally mounted on corresponding first sub-frame hardpoints and first end of each of the second link arms are configured to be pivotally mounted on corresponding second sub-frame hardpoints of the subframe.
In an embodiment, second end of each of the first link arms are configured to be coupled to an operative section of the corresponding steering knuckles via first ball joints at the first knuckle hardpoints and second end of each of the second link arms are configured to be coupled with the corresponding steering knuckles via second ball joints at the second knuckle hardpoints. The control arm axis extends through the centers of their respective the first sub-frame hardpoints, the first knuckle hardpoints, the second sub-frame hardpoints, and the second knuckle hardpoints.
In an embodiment, the first ball joint is integral with the first link arm, and the second ball joint is assembled in the steering knuckle to pivotally mount the second link arm, thereby facilitating independent articulation of the first link arm and the second link arm.
In an embodiment, the shifting of the instantaneous center along the control arm axis varies the knuckle arm length between the center of tie-rods and the virtual kingpin axis. The knuckle arm length dynamically varies as the steering knuckle pivots about the virtual kingpin axis in response to steering input. The variation in the knuckle arm length dynamically adjusts the Ackerman geometry by altering the relative steering angles of the inner and outer front wheels relative to the turning radius. The dynamic adjustment in Ackerman geometry increases the differential steering angle between the inner and outer front wheels, thereby generating a controlled toe-out effect during cornering to achieve a turning circle diameter (TCD) of less than 10 m.
In an embodiment, the split lower control arm configuration and the positioning of the virtual kingpin axis dynamically vary the knuckle arm length. The variation in knuckle arm length changes the mechanical advantage of the steering linkage to reduce the steering torque required to pivot the steering knuckle about the virtual kingpin axis, and wherein the variation in the knuckle arm length increases the Ackermann percentage to reduce the turning circle diameter, thereby facilitating manoeuvrability for vehicles having a gross weight of 2.9 tonnes or higher while maintaining the required steering response and directional stability.
In an embodiment, the first link arm is configured as a handling link, the first link arm is positioned rearward of the steering rack and the second link arm is configured as a comfort link, the second link arm is positioned rearward of the first link arm.
In an embodiment, the knuckle arm length varies dynamically as the steering knuckle pivots about the virtual kingpin axis in response to steering input. The variation controls the relative angular displacement between the inner and outer front wheels for maintaining the Ackerman steering geometry to facilitate progressive steering response and directional stability.
In an embodiment, the shifting of the instantaneous center along the control arm axis causes a corresponding adjustment in the inclination of the virtual kingpin axis. The adjustment regulates the effective knuckle arm length, thereby maintaining a predetermined Ackerman steering percentage to maintain consistent cornering characteristics.
In an embodiment, the forward-mounted steering rack induces toe-out characteristics during cornering, wherein the relative positioning of the first sub-frame hardpoints, the first knuckle hardpoints and the second sub-frame hardpoints, the second knuckle hardpoints is configured to maintain Ackerman geometry.
In an embodiment, the split lower control arm configuration is configured such that the first link arm and the second link arm define a controlled pivoting angular displacement about their respective the first sub-frame hardpoints, the first knuckle hardpoints and the second sub-frame hardpoints, the second knuckle hardpoints in response to a longitudinal force acting on the wheel or tire contact patch. The angular displacement of the first link arm and the second link arm allows the wheel center to shift rearward, thereby reducing the transmission of impact forces to the vehicle structure. The relative movement between the first link arm and the second link arm enables gradual force absorption, thereby mitigating the effect of road irregularities. The controlled pivoting angular displacement of the lower control arms maintains the stability of the suspension geometry, thereby preventing abrupt load variations on suspension components.
In an embodiment, the controlled pivoting angular displacement of the first link arm and the second link arm in response to a longitudinal force acting on the wheel or tire contact patch causes a rearward displacement of the wheel center. The rearward displacement of the wheel center results in a dynamic adjustment of the wheel recess within the wheel arch under vertical loading conditions and the suspension geometry regulates the turning circle diameter (TCD) by maintaining the required steering angles during wheel displacement.
In an embodiment, the suspension assembly is a McPherson-type configuration. The telescopic struts are configured as a structural load-bearing member, to provide both damping and lateral support to the steering knuckle. The upper end of the telescopic strut is mounted to a hardpoint on the vehicle body to allow articulated movement of the steering knuckle about the kingpin axis. The first link arm and the second link arm collectively define a lower guiding mechanism for the steering knuckle, wherein their pivoting angular displacement facilitates controlled variation of the instantaneous center, thereby reducing the turning circle diameter (TCD).
In an embodiment, the first ball joint has a diameter in the range of 30 mm to 33 mm.
In an embodiment, the second ball joint has a diameter in the range of 34 mm to 36 mm.
In an embodiment, the distance between the centre of the first ball joint and the second ball joint is in the range of 40 mm to 100 mm.
BRIEF DESCRIPTION OF THE ACCOMPANYING DRAWINGS
The front suspension assembly of the vehicle, of the present disclosure will now be described with the help of the accompanying drawing in which:
Figure 1 (prior art) illustrates the positioning of the steering tie rod on the rearwards on the Front subframe (FSF) of a conventional suspension system.
Figure 2 illustrates an isometric view of the front corner module sub-assembly (knuckle and lower control arm), depicting the single LCA coupled to the knuckle of Figure 1.
Figure 3 illustrates the conventional single LCA construction of Figure 1.
Figure 4 (prior art) illustrates the isometric view of the front axle of another conventional suspension system.
Figure 5 illustrates the front corner module subsystem (knuckle and split links) of Figure 4, wherein the four-front suspension LCA links form the front suspension.
Figure 6 (prior art) illustrates the isometric view of the conventional front suspension system in which the comfort link is placed towards the front of the vehicle, and the steering tie rod and the stab bar assembly are also placed at the front of the vehicle.
Figure 7A illustrates an isometric view of the left-side sub-front corner module assembly of Figure 6 (knuckle and split links).
Figure 7B illustrates an isometric view of the right-side sub-front corner module assembly of Figure 6 (knuckle and split links).
Figure 8 illustrates a front suspension assembly using a McPherson strut configuration with a split two-link lower control arms arrangement in accordance with the present disclosure.
Figure 9A illustrates the isometric view of the left-side front sub-assembly of the iLCA type of front suspension assembly of Figure 8 in accordance with the present disclosure.
Figure 9B illustrates the isometric view of the right-side front sub-assembly of the iLCA type of front suspension assembly of Figure 8 in accordance with the present disclosure.
Figures 10A-10B illustrate a comparison of the forward placement of the steering rack of Figure 8 in accordance with the present disclosure and the rear-mounted steering rack of the conventional suspension system.
Figure 11 illustrates a top view of the proposed suspension assembly depicting formation of virtual intersection point by first link arm and second link arm in accordance with the present disclosure.
Figure 12 illustrates the knuckle arm length in rear view and side view of the front suspension assembly and the king pin axis inclination in accordance with the present disclosure.
Figure 13 illustrates dynamic variation in knuckle arm lengths to achieve a low turning circle diameter (TCD) in accordance with the present disclosure.
Figure 14 illustrates a graphical comparison of the longitudinal travel of the wheels during bump events with the proposed suspension assembly having iLCA configuration and the conventional suspension system with solid LCA configuration.
LIST OF REFERENCE NUMERALS
20 First conventional front suspension system
12 Stab bar assembly of the first conventional front suspension system
14 Lower control arm (LCA) of the first conventional front suspension system
14A Ball joint
14B Front inner bush
14C Rear inner bush
16 Knuckle of the first conventional front suspension system
18 Steering tie rod of the first conventional front suspension system
22 Chassis / front subframe
54 Second conventional front suspension system
42 Stab bar of the second conventional front suspension system
44 Steering tie rod of the second conventional front suspension system
46 Double-wishbone or upper links of the second conv. front suspension system
48A Knuckle LH of the second conventional front suspension system
48B Knuckle RH of the second conventional front suspension system
50A Handling link LH of the second conventional front suspension system
50B Handling link RH of the second conventional front suspension system
52A Comfort link LH of the second conventional front suspension system
52B Comfort link RH of the second conventional front suspension system
84 Third conventional front suspension system
70 Stab bar of the third conventional front suspension system
72A Comfort link LH of the third conventional front suspension system
72B Comfort link RH of the third conventional front suspension system
74 Steering tie rod of the third conventional front suspension system
80A Handling link LH of the third conventional front suspension system
80B Handling link RH of the third conventional front suspension system
81 Ball joint of the third conventional front suspension system
82A Knuckle LH of the third conventional front suspension system
82B Knuckle RH of the third conventional front suspension system
86 Chassis / front sub frame of the third conventional front suspension system
200 Front suspension assembly of the present disclosure
102A, 102B Tie-rod
105A, 105B Second sub-frame hardpoints
106A, 160B First sub-frame hardpoints
107A, 107B Second knuckle hardpoints
108A, 108B First knuckle hardpoints
109A, 109B Tie-rod hardpoints
110A, 110B Hardpoints on the vehicle body
111A, 111B First ball joint
112A, 112B First link arm
113A, 113B Second ball joint
114A, 114B Second link arm
115A, 115B Telescopic strut
116A, 116B Steering knuckle
140 Front sub-frame (FSF) of the present disclosure
150 Steering rack
160 Stabilizer bar
A1, A2 Knuckle arm length
K1, K2 Virtual kingpin axis
W1, W2 Instantaneous center
TCD Turning circle diameter
F Front of the vehicle
M Line depicting the assembly of the present disclosure
N Line depicting the conventional system (single or solid LCA)
DETAILED DESCRIPTION
The present disclosure relates to a front suspension assembly of a vehicle. More specifically, the present disclosure relates to a front suspension architecture with lower control arms for a sports utility vehicle (SUV).
Embodiments are provided so as to thoroughly and fully convey the scope of the present disclosure to the person skilled in the art. Numerous details are set forth, relating to specific components, and methods, to provide a complete understanding of embodiments of the present disclosure. It will be apparent to the person skilled in the art that the details provided in the embodiments should not be construed to limit the scope of the present disclosure. In some embodiments, well-known processes, well-known apparatus structures, and well-known techniques are not described in detail.
The terminology used, in the present disclosure, is only for the purpose of explaining a particular embodiment and such terminology shall not be considered to limit the scope of the present disclosure. As used in the present disclosure, the forms "a,” "an," and "the" may be intended to include the plural forms as well, unless the context clearly suggests otherwise. The terms "comprises," "comprising," “including,” and “having,” are open ended transitional phrases and therefore specify the presence of stated features, elements, modules, units and/or components, but do not forbid the presence or addition of one or more other features, elements, components, and/or groups thereof.
Typically, the front suspension system (20) significantly influences a vehicle's ride comfort, handling, and manoeuvrability. A crucial component of the suspension system (20) is the lower control arm (LCA (14)), which links the knuckle (16) of the wheel assembly to the chassis (24) of the vehicle. Conventional LCA (14) configurations, including single and split configurations, exhibits limitations in terms of performance, packaging, and durability.
The McPherson strut suspension system (20) with the single LCA (14) is simple and cost-effective, however, it presents significant limitations in terms of ride and handling due to the constrained flexibility in positioning the hardpoints of the LCA (14). The placement of the ball joint (14A) and the rearward positioning of the steering tie rod (18) on the FSF contribute to suboptimal steering behaviour, particularly toe-in during cornering, leading to undesirable oversteer tendencies. Further, the turning circle diameter (TCD) of vehicles using the single LCA (14) configuration, with a wheelbase of approximately 3,000 mm, is around 12 meters, which compromises vehicle manoeuvrability, especially in urban or tight spaces. Figure 1 illustrates the positioning of the steering tie rod (18) on the rearwards on the Front subframe (FSF) of the conventional suspension system (20).
Figure 2 illustrates an isometric view of the front corner module sub-assembly (knuckle (16) and the lower control arm (14)), depicting the single LCA (14) coupled to the knuckle (16) of Figure 1. The coupling is achieved through two bushes (14B, 14C), which connect the single LCA (14) to the front subframe (FSF) or chassis (24), as illustrated in Figure 2. The positioning of these two bushes (14B, 14C) on the FSF forms the frame-side hardpoints, which are critical to optimize the ride and handling characteristics of the vehicle. In the conventional single LCA (14) configuration, however, the flexibility in positioning these hardpoints at the frame end, along with the single ball joint (14A) hardpoint at the knuckle (16) end, is significantly constrained. This limitation arises due to the inherent packaging constraints and the single LCA (14) construction, which restricts the optimal placement of these key hardpoints. Figure 3 illustrates the conventional single LCA (14) construction of Figure 1.
To address the issues of the single LCA (14) configuration of the first suspension system (20), a double-wishbone suspension system (40) with a split LCA was developed, which uses two upper (46) and two lower links i.e., a handling link (50A, 50B) and a comfort link (52A, 52B) which are operatively mounted on a chassis (54). Figure 4 illustrates the isometric view of the front axle of another conventional suspension system (40). The handling link (50A, 50B) is placed forward of the vehicle, whereas the comfort link (52A, 52B) is placed rearward of the vehicle. The ball joint of the comfort link (52A, 52B) is integral part of the comfort link (52A, 52B), whereas the ball joint of the handling link (50A, 50B) is fastened in the knuckle (48A, 48B) with the help of a pinch bolt. Figure 5 illustrates the front corner module subsystem of Figure 4 wherein the four-front suspension LCA links (50A, 50B, 52A, 52B) and double wishbone links (46) form the conventional front suspension system (40). This configuration provides enhanced load distribution and improved ride and handling. As the conventional suspension system (40) is of four links construction (50A, 50B, 52A, 52B) with double wishbone (46) type of system (40), the packaging envelope is relatively larger, expensive, and complex assembly.
Further, a front suspension system (84) was developed with a hybrid split two-link LCA (72A, 72B, 80A, 80B), by combining the features of the McPherson strut and split configurations. Figure 6 (prior art) illustrates the isometric view of the conventional front suspension system (84) in which the comfort link (72A, 72B) is placed towards the front of the vehicle and the steering tie rod (74) and the stab bar (70) is also placed at the front of the vehicle. The ball joint (81) for handling and comfort links (80A, 80B, 72A, 72B) are integrated in the handling and comfort link (80A, 80B, 72A, 72B) respectively as shown in Figures 7A and 7B. While the hybrid split two-link LCA configuration provides a more compact suspension, however it introduces drawbacks, such as:
• the comfort and handling link ball joints (81) are integrated in their respective links (72A, 72B, 80A, 80B), their total cost is relatively expensive as compared to non-integrated ball joint suspension links;
• as both ball joints (81) are integrated in their respective links, they are non-serviceable. Any ball joint concern or failure during the vehicle’s service life will require replacement of the entire link;
• furthermore, placement of the comfort link hardpoint at the front end is challenging due to the presence of the stab bar (70) mounting and steering rack (74) mounting at the front end itself, which constraints the packaging space significantly, thereby also compromising the suspension performance.
To overcome the aforementioned drawbacks, the present disclosure envisages a front suspension assembly (200) of a vehicle with lower control arms or especially independent lower control arms (iLCAs) (hereinafter referred to as iLCA’s (112A, 112B, 114A, 114B)).
Embodiments, of the present disclosure, will now be described with reference to the accompanying drawing.
A front suspension assembly (200) of a vehicle, of the present disclosure, will now be described in detail with reference to Figure 8 through Figure 14.
The present disclosure relates to the front suspension assembly (200) of a vehicle as illustrated in Figure 8. The front suspension assembly (200) is configured to optimize steering geometry and minimize the turning circle diameter (TCD), thereby improving manoeuvrability and stability. The vehicle includes a front subframe (FSF) (140) and a steering assembly. The FSF (140) is configured to be securely mounted on the vehicle chassis at a front side (F) and is configured to act as a framework for mounting the front suspension assembly (200), the steering assembly, and drivetrain components. The FSF (140) defines multiple hardpoints i.e., first sub-frame hardpoints (106A, 106B), and second sub-frame hardpoints (105A, 105B) that facilitate the pivotal mounting of various components of the front suspension assembly (200) as illustrated in Figures 9A-9B and Figure 13. Further, the front subframe (FSF) (140) has a ladder frame structure having a left-side arm (145A) and a right-side arm (145B). Each of the hardpoints (105A, 106A) are configured to be positioned on the left-side arm (145A), while the hardpoints (105B, 106B) are configured to be positioned on the right-side arm (145B). The first and second sub-frame hardpoints (105A, 106A, 105B, 106B) on the left and right arms (145A, 145B) are arranged in a mirrored configuration.
The steering assembly includes a steering rack (150) and a pair of tie rods (102A, 102B). The steering rack (150) is positioned at the front side (F) of the vehicle and is configured to be mounted on the FSF (140). The steering rack (150) has a first end and a second end. Each end of the steering rack is configured with corresponding tie rods (102A, 102B). Each of the tie rods (102A, 102B) is configured to be movably connected to each of the front wheels by means of a respective knuckle (116A, 116B). The steering rack (150) is configured to translate a driver’s steering input into lateral movement of tie rods (102A, 102B), thereby controlling the manoeuvrability of the front wheels.
In an embodiment, the tie rod (102A) is configured to be movably connected to the left-hand front wheel via the respective knuckle (116A), while the tie rod (102B) is configured to be movably connected to the right-hand front wheel via the knuckle (116B).
Further, the FSF (140) includes a stabilizer bar (160) having a first end and a second end, which is configured to be mounted on the FSF (140). The first end of the stabilizer bar (160) is operatively connected to the left-side sub-assembly (200A), while the second end is operatively connected to the right-side sub-assembly (200B). The stabilizer bar (160) is mounted on the FSF (140) such that the first end is positioned between sub-frame hardpoints (105A, 106A) on the left-side arm (145A), and the second end is positioned between the sub-frame hardpoints (105B, 106B) on the right-side arm (145B).
Further, the front suspension assembly (200) comprises a left-side front sub-assembly (200A) as illustrated in Figure 9A and a right-side sub-assembly (200B) as illustrated in Figure 9B. The left-side front sub-assembly (200A) is configured to be mounted on a left-side arm (145A) of the FSF (140), while the right-side sub-assembly (200B) is configured to be mounted on a right-side arm (145B) of the FSF (140) in a mirrored configuration. Each of the left-side and right-side front sub-assemblies (200A, 200B) comprises a respective steering knuckle (116A, 116B), a telescopic strut (115A, 115B), a first link arm (112A, 112B), and a second link arm (114A, 114B) as illustrated in Figure 13.
An operative section of each of the steering knuckles (116A, 116B) is configured to be pivotally coupled to the respective front wheels, thereby facilitating rotational movement about a defined steering axis. The steering knuckles (116A, 116B) are configured with a plurality of hardpoints. i.e., first knuckle hardpoints (108A, 108B), second knuckle hardpoints (107A, 107B), tie rod hardpoints (109A, 109B), each hardpoints (107A, 107B, 108A, 108B, 109A, 109B) serving as a designated attachment location for securing various components of the steering assembly and the suspension assemblies (200A, 200B) to the FSF (140) as illustrated in Figure 8. The steering knuckles (116A, 116B) are configured to be operatively connected to the steering rack (150) via respective tie rods (102A, 102B) at the corresponding tie-rod hardpoints (109A, 109B) as illustrated in Figure 11. Each of the steering knuckles (116A, 116B) defines a virtual kingpin axis (K1, K2) about which the steering knuckles (116A, 116B) pivot.
In an embodiment, the steering knuckles (116A, 116B) are configured to be operatively connected to the respective tie rods (102A, 102B) at the corresponding tie-rod hardpoints (109A, 109B) via respective third ball joints (117A, 117B).
Each telescopic strut (115A, 115B) has an operative upper end and an operative lower end and functions as a load-bearing member providing damping and lateral support. Each of the upper ends is configured to be coupled with an operative section of the vehicle structure at a predefined vehicle hardpoint (110A, 110B) as illustrated in Figure 12 and each of the lower ends is configured to be mounted on a predefined mounting interface of the corresponding steering knuckle (116A, 116B).
The front suspension assembly (200) of a vehicle has a split lower control arm (independent lower control arms (iLCAs)) configuration including the pair of first link arms (112A, 112B) and the pair of second link arms (114A, 114B). Each of the first link arms (112A, 112B) has a first end configured to be pivotally mounted on the corresponding first sub-frame hardpoints (106A, 106B) provided on the left-side arm (145A) and the right-side arm (145B) of the front subframe (140), respectively. The second end of each of the first link arms (112A, 112B) is configured to be coupled to an operative section of the corresponding steering knuckle (116A, 116B) via a first ball joints (111A, 111B). The first ball joints (111A, 111B) are mounted at corresponding first knuckle hardpoints (108A, 108B) on the left-side arm (145A) and the right-side arm (145B) of the front subframe (140), respectively.
In an embodiment, the first ball joints (111A, 111B) are mounted at corresponding first knuckle hardpoints (108A, 108B) on the left-side steering knuckle(116A) and the right-side steering knuckle (116B), respectively.
In an embodiment, the first link arm (112A, 112B) is configured as a handling link and is configured to be positioned rearward of the steering rack (150).
Similarly, each of the second link arms (114A, 114B) has a first end and a second end. The first end of each of the second link arms (114A, 114B) is configured to be pivotally mounted on the corresponding second sub-frame hardpoints (105A, 105B) provided on the left-side arm (145A) and the right-side arm (145B) of the front subframe (140), respectively. The second end of each of the second link arms (114A, 114B) is configured to be coupled to an operative section of the corresponding steering knuckle (116A, 116B) via second ball joints (113A, 113B). The second ball joints (113A, 113B) are mounted at corresponding second knuckle hardpoints (107A, 107B) on the left-side arm (145A) and the right-side arm (145B) of the front subframe (140), respectively. Each of the first link arms (112A, 112B) and each of the second link arms (114A, 114B) define a control arm axis (C1, C2) respectively.
In an embodiment, the second ball joints (113A, 113B) are mounted at corresponding second knuckle hardpoints (107A, 107B) on the left-side steering knuckle(116A) and the right-side steering knuckle (116B), respectively.
The control arm axis (C1, C2) is defined by lines extending through the centers of the first sub-frame and the first knuckle hardpoints (106A, 108A, 106B, 108B) of the first link arms (112A, 112B) and the second sub-frame and the second knuckle hardpoints (105A, 107A, 105B, 107B) of the second link arms (114A, 114B) as illustrated in Figures 11 and 13.
In an embodiment, the second link arm (114A, 114B) is configured as a comfort link, and is configured to be positioned rearward of the first link arm (112A, 112B).
In an embodiment, the first ball joint (111A, 111B) is integral with the first link arm (112A, 112B), and the second ball joint (113A, 113B) is integral with the steering knuckle (116A, 116B) to pivotally mount the second link arm (114A, 114B) to enable independent articulation of the first link arm (112A, 112B) and the second link arm (114A, 114B) to optimize steering geometry.
In an embodiment, the front suspension assembly (200) ensures optimal load transfer characteristics by maintaining an ideal relationship between the ball joints (111A, 111B, 113A, 113B). The distance between the centers of the first ball joint (111A, 111B) and the second ball joint (113A, 113B) is maintained within a predefined operational range of 40 mm to 100 mm. This configuration optimizes the positioning of the instantaneous center (W1, W2), thereby minimizing tire scrub and improving cornering efficiency.
In an embodiment, the steering knuckle (116A, 116B) is configured to accommodate the dynamic movement of the second ball joint (113A, 113B), wherein the configuration prevents interference between the first link arm (112A, 112B) and the second link arm (114A, 114B) while maintaining the required steering response.
In an embodiment, the mounting interfaces of the first ball joint (111A, 111B) and the second ball joint (113A, 113B) are positioned in close adjacency on the steering knuckle (116A, 116B), wherein the positioning enables a compact packaging arrangement while ensuring a lower turning circle diameter (TCD) of less than 10m.
The intersection of each control arm axis (C1, C2) with the corresponding virtual kingpin axis (K1, K2) determines an instantaneous center (W1, W2) as illustrated in Figures 11 and 13. Each of the kingpin axes (K1, K2) is defined by the line extending from the corresponding instantaneous center (W1, W2) to the corresponding hardpoint (110A, 110B). The instantaneous center (W1, W2) is configured to be dynamically shifted based on the steering input, such that the instantaneous centers (W1, W2) laterally shift along the control arm axis (C1, C2) in response to the steering input. Further, the shifting of the instantaneous centers (W1, W2) dynamically alters the instantaneous steering geometry, thereby optimizing the steering geometry and reducing the turning circle diameter (TCD).
The shifting of the instantaneous center (W1, W2) along the control arm axis (C1, C2) leads to a variation in the knuckle arm length (A1, A2). The knuckle arm length (A1, A2) is defined as the distance between the center of the tie-rod hardpoints (109A, 109B) on the steering knuckle (116A, 116B) and the virtual kingpin axis (K1, K2). The knuckle arm length (A1, A2) dynamically varies as the steering knuckle (A1, A2) pivots about the virtual kingpin axis (K1, K2) in response to steering input as illustrated in Figure 13. The variation in the knuckle arm length (A1, A2) dynamically adjusts the Ackerman geometry by altering the relative steering angles of the inner and outer front wheels to meet required turning radius. In addition, the variation in the knuckle arm length (A1, A2) increases the Ackermann percentage to reduce the turning circle diameter, thereby facilitating manoeuvrability for vehicles having a gross weight of 2.9 tonnes or higher while maintaining the required steering response and directional stability.
Furthermore, the dynamic adjustment of Ackerman geometry results in an increased differential steering angle between the inner and outer front wheels. This greater angle difference effectively induces the controlled toe-out effect during cornering, which enhances the vehicle’s ability to navigate tight turns with improved stability.
The functional effects of the variable knuckle arm length (A1, A2) and its influence on Ackerman geometry and turning performance are illustrated in greater detail in Figure 13. Figure 13 illustrates that the variation in knuckle arm length (A1, A2) varies continuously when the wheel turns. Furthermore, the knuckle arm length (A1, A2) differs significantly for the left and right wheel during turning in the event of steering the vehicle for cornering. Figure 13 illustrates a left turn taken by the vehicle wherein both the LH and RH wheels turn leftwards, wherein the tilt angle of the left wheel (LH) and the right wheel (RH) is not the same due to the difference in the knuckle arm length (A1, A2). This differential in the knuckle arm length (A2 > A1) between left-hand side (LH) and right-hand side (RH) results in a significant difference between outer and inner wheel angles during cornering of the vehicle when the steering wheel is turned, thereby resulting in higher Ackermann percentage. A higher Ackermann percentage indicates a higher difference between outer and inner wheel angles during steering. The greater the angle difference, the lower the turning circle diameter of the vehicle. However, the front placed steering rack (150) generally yields a lower Ackermann percentage thereby resulting in higher TCD. Considering understeer targets, the steering rack was placed at the front end. However, by strategically placing the first sub-frame hardpoints (106A, 106B) and the second sub-frame hardpoints (105A, 105B) of the iLCA configuration, a lower TCD is achieved despite a forward placed steering rack (150) position by virtue of hardpoint (105A, 107A, 105B, 107B, 106A, 108A, 106B, 108B) positioning of the iLCA. Further, this key feature of variable knuckle arm length (A1, A2) is controlled by the positioning of the first and second knuckle hardpoints (108A, 108B, 107A, 107B) on the knuckle (116A, 116B) and the front suspension assembly (200) ensures lower TCD with sufficient understeer and best in class ride and handling characteristics for the vehicle.
In an embodiment, the forward-mounted steering rack (150) induces the toe-out effect during cornering, wherein the spatial arrangement of the hardpoints (105A, 105B, 106A, 106B, 107A, 107B, 108A, 108B) associated with the first link arm (112A, 112B) and the second link arm (114A, 114B) is configured to compensate for the reduction in Ackermann percentage. This configuration helps to maintain an optimal steering response and desired understeer gradient. Additionally, the split lower control arm configuration, formed by the combination of the first link arm (112A, 114A) and the second link arm (112B, 114B), facilitates dynamic load transfer during braking and cornering. The tendency of conventional suspension systems to induce oversteer due to increased rear axle load is mitigated by the forward placement of the steering rack (150) and the precise positioning of the first link arms (112A, 112B) and the second link arms (114A, 114B) (iLCA) hardpoints (105A, 105B, 106A, 106B, 107A, 107B, 108A, 108B). This arrangement enhances vehicle stability, ensuring controlled handling and precise manoeuvrability, particularly under high-load conditions.
In an embodiment, the knuckle arm length (A1, A2) varies dynamically as the steering knuckle (116A, 116B) pivots about the virtual kingpin axis (K1, K2) in response to steering input. The variation controls the relative angular displacement between the inner and outer front wheels for adjusting the steering geometry.
In an embodiment, the shifting of the instantaneous center (W1, W2) along the control arm axis (C1, C2) causes a corresponding adjustment in the inclination of the virtual kingpin axis (K1, K2). The adjustment regulates the effective knuckle arm length (A1, A2), thereby maintaining a predetermined Ackerman steering percentage to ensure consistent cornering characteristics.
In an embodiment, the inclination of the virtual kingpin axis (A1, A2) varies dynamically in response to steering input, wherein the variation regulates the steering angles of the inner and outer front wheels, thereby adjusting the Ackerman geometry to maintain progressive steering response and directional stability.
In an embodiment, the split lower control arm (iLCA) configuration including the first link arm (112A, 112B) and the second link arm (114A, 114B) defines a controlled pivoting angular displacement about their respective hardpoints (105A, 105B, 106A, 106B, 107A, 107B, 108A, 108B) in response to a longitudinal force acting on the wheel or tire contact patch. This angular displacement of the first link arm (112A, 112B) and the second link arm (114A, 114B) allows the wheel center to shift rearward, thereby reducing the transmission of impact forces to the vehicle structure. The relative movement between the first link arm (112A, 112B) and the second link arm (114A, 114B) enables gradual force absorption, thereby mitigating the effect of road irregularities such as uneven surfaces, speed breakers or bumps. The controlled pivoting angular displacement of the first and second link arms (112A, 112B, 114A, 114B) maintains the stability of the suspension geometry, thereby preventing abrupt load variations on suspension components.
In an embodiment, the controlled pivoting angular displacement of the first link arm (112A, 112B) and the second link arm (114A, 114B) in response to a longitudinal force acting on the wheel or tire contact patch causes a rearward displacement of the wheel center. The rearward displacement of the wheel center results in a dynamic adjustment of the wheel recess within the wheel arch under vertical loading conditions. The adjusted wheel recess limits the required wheel travel envelope, thereby controlling the upward displacement of the wheel while accommodating suspension compression. Furthermore, the suspension geometry regulates the turning circle diameter (TCD) by maintaining the required steering angles during wheel displacement.
In an embodiment, the front suspension assembly (200) is a McPherson-type configuration, wherein the telescopic strut (115A, 115B) is configured as a structural load-bearing member to provide both damping and lateral support to the steering knuckle (116A, 116B). The upper end of the telescopic strut (115A, 115B) is mounted to a hardpoint (110A) on the vehicle body to allow articulated movement of the steering knuckle (116A, 116B) about the kingpin axis (K1, K2). Additionally, the first link arm (112A, 112B) and the second link arm (114A, 114B) collectively define a lower guiding mechanism for the steering knuckle (116A, 116B), wherein their pivoting angular displacement facilitates controlled variation of the instantaneous center (W1, W2), thereby reducing the turning circle diameter (TCD). The positioning of the first knuckle and second knuckle hardpoints (108A, 108B, 107A, 107B) on the steering knuckle (116A, 116B) and the first sub-frame and second sub-frame hardpoints (108A, 108B, 107A, 107B) on the front subframe (FSF) (140) is configured to refine the suspension kinematics. The proposed mounting configuration enhances vehicle stability during high-speed manoeuvres while improving safety and handling characteristics. For example, the forward placement of the steering rack (150) generates the couple acting on the front wheels during cornering, thereby generating a controlled understeer behaviour. The controlled understeer maintains the stability of the vehicle during high-speed driving or abrupt steering manoeuvres. The precise placement of suspension hardpoints also optimize the virtual kingpin axis (K1, K2), which reduces lateral and longitudinal displacement of the wheel center and minimizes wheel travel within the wheel arch during steering. These features allow for efficient utilization of available packaging space without affecting dynamic performance.
In an embodiment, the first and second link arms (112A, 114A, 112B, 114B) may feature an aperture (profile) and a cross-sectional geometry configured to withstand dynamic packaging constraints with structural durability. The lower arms (112A, 114A, 112B, 114B) may be configured to resist critical buckling loads induced by lateral compressive forces, which may arise during high-speed cornering or abrupt steering inputs. The configuration of the lower arms (112A, 114A, 112B, 114B) allows the vehicle to deliver a comfortable ride while maintaining the ability of the suspension assembly (200) to handle extreme forces without deformation or failure. In an embodiment, the split lower control arm configuration and the positioning of the virtual kingpin axis (K1, K2) dynamically vary the knuckle arm length (A1, A2). The variation in knuckle arm length (A1, A2) results in increase in Ackermann percentage thereby reducing turning circle diameter. The adjusted steering input force facilitates controlled articulation of the front wheels, enabling manoeuvrability for vehicles with a gross weight of 2.9 tonnes or higher while maintaining the required steering response and directional stability.
In an embodiment, the first ball joint (111A, 111B) has a diameter in the range of 30 mm to 33 mm.
In an embodiment, the second ball joint (113A, 113B) has a diameter in the range of 34 mm to 36 mm.
In an embodiment, the distance between the first ball joint (111A, 111B) and the second ball joint (113A, 113B) is in the range of 40 mm to 100 mm.
In an embodiment, the steering rack forward placement combined with the split lower control arm configuration results in at least a 30% increase in Ackermann percentage compared to a conventional single-link McPherson front suspension system.
In an embodiment, the split lower control arm (iLCA) (112A, 114A, or 112B, 114B) configuration offers several additional benefits, which include increased longitudinal compliance, which enhances the vehicle's ride comfort. The first link arm (114A, 114B) and the second link arm (112A, 112B) within the iLCA may work synergistically to optimize both ride comfort and handling dynamics.
The longitudinal compliance of the suspension assembly (200) thus allows the wheels to exhibit greater longitudinal travel during bump events, as illustrated in Figure 14. As illustrated in Figure 14, when the vehicle travels straight ahead and when the LH and RH wheels are maintained parallel to each other, the longitudinal wheel travel from the virtual wheel centre for the proposed configuration of iLCA (112A, 114A, or 112B, 114B) is higher when compared to the conventional solid configuration of LCA for the convention suspension system (200). As depicted in Figure 14, in the event of vehicle traversing a bump, the higher longitudinal travel of the wheel in the iLCA type (M) configuration ensures increased comfort to the passengers. This increased comfort to the passengers due to the higher wheel longitudinal travel would not have been possible with the solid LCA type (N) of configuration, the longitudinal wheel recess during the bump event is much less as depicted in Figure 14. Additionally, the rearward placement of the second link arm (112A, 112B) relative to the first link arm (114A, 114B) may refine suspension kinematics further by reducing the X-Y displacement of the wheel center during steering. Such improvements minimize the tire packaging envelope, offering significant advantages in packaging efficiency, especially for vehicles with longer wheelbases and higher axle loads. The rearward placement of the second link arm (112A, 112B) relative to the first link arm (114A, 114B) and the close placement of their ball joints(111A, 111B, 113A, 113B) may refine suspension kinematics further by reducing the X-Y displacement of the wheel center during steering. Such improvements minimize the tire packaging envelope, offering significant advantages in packaging efficiency, especially for vehicles with longer wheelbases and higher axle loads.
The foregoing description of the embodiments has been provided for purposes of illustration and not intended to limit the scope of the present disclosure. Individual components of a particular embodiment are generally not limited to that particular embodiment but are interchangeable. Such variations are not to be regarded as a departure from the present disclosure, and all such modifications are considered to be within the scope of the present disclosure.
TECHNICAL ADVANCES AND ECONOMCAL SIGNIFICANCE
The present disclosure described hereinabove has several technical advantages including, but not limited to, the front suspension assembly of a vehicle that;
• allows greater flexibility in positioning of hardpoints to achieve improved ride comfort, handling performance, and alignment with vehicle dynamics;
• reduces the TCD, and therefore enhances the manoeuvrability of vehicles, especially in tight spaces and urban driving conditions;
• offers compact suspension assembly which effectively utilizes available space, and ensures seamless integration with the chassis of the vehicle without increasing complexity;
• minimizes noise, vibration, and harshness by reducing the component wear and optimizing the geometry of the assembly;
• offers serviceable components to reduce maintenance costs and extend the operational lifespan of the suspension assembly;
• minimizes manufacturing complexity to reduce the number of components, and simplifies the assembly process while maintaining or enhancing functional performance;
• enhances vehicle stability during cornering, braking, and acceleration by addressing the limitations of existing single-piece and split-link configurations;
• adapts to vehicles with varying wheelbase lengths, providing consistent performance regardless of vehicle size or configuration; and
• evenly distribute loads across the suspension components to reduce localized stress and extend the life of the suspension assembly.
The foregoing disclosure has been described with reference to the accompanying embodiments which do not limit the scope and ambit of the disclosure. The description provided is purely by way of example and illustration.
The embodiments herein and the various features and advantageous details thereof are explained with reference to the non-limiting embodiments in the following description. Descriptions of well-known components and processing techniques are omitted so as to not unnecessarily obscure the embodiments herein. The examples used herein are intended merely to facilitate an understanding of ways in which the embodiments herein may be practiced and to further enable those of skill in the art to practice the embodiments herein. Accordingly, the examples should not be construed as limiting the scope of the embodiments herein.
The foregoing description of the specific embodiments so fully reveal the general nature of the embodiments herein that others can, by applying current knowledge, readily modify and/or adapt for various applications such specific embodiments without departing from the generic concept, and, therefore, such adaptations and modifications should and are intended to be comprehended within the meaning and range of equivalents of the disclosed embodiments. It is to be understood that the phraseology or terminology employed herein is for the purpose of description and not of limitation. Therefore, while the embodiments herein have been described in terms of preferred embodiments, those skilled in the art will recognize that the embodiments herein can be practiced with modification within the spirit and scope of the embodiments as described herein.
Any discussion of devices, articles or the like that has been included in this specification is solely for the purpose of providing a context for the disclosure. It is not to be taken as an admission that any or all of these matters form a part of the prior art base or were common general knowledge in the field relevant to the disclosure as it existed anywhere before the priority date of this application.
While considerable emphasis has been placed herein on the components and component parts of the preferred embodiments, it will be appreciated that many embodiments can be made and that many changes can be made in the preferred embodiments without departing from the principles of the disclosure. These and other changes in the preferred embodiment as well as other embodiments of the disclosure will be apparent to those skilled in the art from the disclosure herein, whereby it is to be distinctly understood that the foregoing descriptive matter is to be interpreted merely as illustrative of the disclosure and not as a limitation. ,CLAIMS:WE CLAIM:
1. A front suspension assembly (200) of a vehicle, the vehicle comprising a front subframe (FSF) (140) mounted on a vehicle chassis, and a steering rack (150) positioned at a front side (F) of the vehicle and mounted on the front subframe (140), the steering rack (150) configured to convert a steering input in lateral movement of tie rods (102A, 102B), wherein said suspension assembly (200) configured to optimize steering geometry and reduce the turning circle diameter (TCD), said front suspension assembly (200) comprising:
• a pair of steering knuckles (116A, 116B), each of said steering knuckles (116A, 116B) configured to be pivotally coupled to a respective front wheel and operatively connected to the steering rack (150) via the tie rods (102A, 102B), each of said steering knuckles (116A, 116B) defining a virtual kingpin axis (K1, K2) about which steering knuckles (116A, 116B) pivot;
• a pair of telescopic struts (115A, 115B), each of said telescopic struts (115A, 115B) having an operative upper end configured to be coupled with an operative section of the vehicle structure and an operative lower end configured to be mounted on a predefined mounting interface of said corresponding steering knuckle (116A, 116B); and
• a split lower control arm configuration having a pair of first link arms (112A, 112B) and a pair of second link arms (114A, 114B), said split lower control arm configured to be pivotally mounted between the front subframe (140) and said corresponding steering knuckles (116A, 116B), each of said first link arm (112A, 112B) and each of said second link arm (114A, 114B) defining a control arm axis (C1, C2),
wherein an instantaneous center (W1, W2) is defined by the intersection of each of said control arm axes (C1, C2) with each of said kingpin axes (K1, K2), which dynamically shifts in response to the steering input to alter the steering geometry and reduce the turning circle diameter (TCD).
2. The front suspension assembly (200) as claimed in claim 1, wherein each of said steering knuckles (116A, 116B) includes first knuckle hardpoints (108A, 108B) and second knuckle hardpoints (107A, 107B), and wherein first end of each of said first link arms (112A, 112B) are configured to be pivotally mounted on corresponding first sub-frame hardpoints (106A, 106B) and first end of each of said second link arms (114A, 114B) are configured to be pivotally mounted on corresponding second sub-frame hardpoints (105A, 105B) of the subframe (140).
3. The front suspension assembly (200) as claimed in claim 2, wherein second end of each of said first link arms (112A, 112B) are configured to be coupled to an operative section of said corresponding steering knuckles (116A, 116B) via first ball joints (111A, 111B) at said first knuckle hardpoints (108A, 108B) and second end of each of said second link arms (114A, 114B) are configured to be coupled with said corresponding steering knuckles (116A, 116B) via second ball joints (113A, 113B) at said second knuckle hardpoints (107A, 107B), wherein said control arm axis (C1, C2) extends through the centers of their respective said first sub-frame hardpoints (106A, 106B), said first knuckle hardpoints (108A, 108B), said second sub-frame hardpoints (105A, 105B), and said second knuckle hardpoints (107A, 107B).
4. The front suspension assembly (200) as claimed in claim 3, wherein said first ball joint (111A, 111B) is integral with said first link arm (112A, 112B), and said second ball joint (113A, 113B) is assembled in said steering knuckle (116A, 116B) to pivotally mount said second link arm (114A, 114B), thereby facilitating independent articulation of said first link arm (112A, 112B) and said second link arm (114A, 114B).
5. The front suspension assembly (200) as claimed in claim 1, wherein the shifting of said instantaneous center (W1, W2) along said control arm axis (C1, C2) varies the knuckle arm length (A1, A2) between the center of tie-rods (102A, 102B) and the virtual kingpin axis (K1, K2), wherein:
o the knuckle arm length (A1, A2) dynamically varies as said steering knuckle (A1, A2) pivots about said virtual kingpin axis (K1, K2) in response to steering input;
o the variation in the knuckle arm length (A1, A2) dynamically adjusts the Ackerman geometry by altering the relative steering angles of the inner and outer front wheels relative to the turning radius; and
o the dynamic adjustment in Ackerman geometry increases the differential steering angle between the inner and outer front wheels, thereby generating a controlled toe-out effect during cornering to achieve a turning circle diameter (TCD) of less than 10m.
6. The front suspension assembly (200) as claimed in claim 5, wherein said split lower control arm configuration and the positioning of the virtual kingpin axis (K1, K2) dynamically varies the knuckle arm length (A1, A2), wherein the variation in knuckle arm length (A1, A2) changes the mechanical advantage of the steering linkage to reduce the steering torque required to pivot said steering knuckle (116A, 116B) about the virtual kingpin axis (K1, K2), and wherein said variation in said knuckle arm length (A1, A2) increases the Ackermann percentage to reduce the turning circle diameter, thereby facilitating manoeuvrability for vehicles having a gross weight of 2.9 tonnes or higher while maintaining the required steering response and directional stability.
7. The front suspension assembly (200) as claimed in claim 1, wherein:
o said first link arm (112A, 112B) is configured as a handling link, said first link arm (112A, 112B) is positioned rearward of said steering rack; and
o said second link arm (114A, 114B) is configured as a comfort link, said second link arm (114A, 114B) is positioned rearward of said first link arm (112A, 112B).
8. The front suspension assembly (200) as claimed in claim 5, wherein the knuckle arm length (A1, A2) varies dynamically as said steering knuckle (116A, 116B) pivots about said virtual kingpin axis (K1, K2) in response to steering input, wherein said variation controls the relative angular displacement between the inner and outer front wheels for maintaining the Ackerman steering geometry to facilitate progressive steering response and directional stability.
9. The front suspension assembly (200) as claimed in claim 8, wherein the shifting of said instantaneous center (W1, W2) along said control arm axis (C1, C2) causes a corresponding adjustment in the inclination of said virtual kingpin axis (K1, K2), wherein said adjustment regulates the effective knuckle arm length (A1, A2), thereby maintaining a predetermined Ackerman steering percentage to maintain consistent cornering characteristics.
10. The front suspension assembly (200) as claimed in claim 3, wherein the forward-mounted steering rack (150) induces toe-out characteristics during cornering, wherein the relative positioning of said first sub-frame hardpoints (106A, 106B), said first knuckle hardpoints (108A, 108B) and said second sub-frame hardpoints (105A, 105B), said second knuckle hardpoints (107A, 107B) is configured to maintain Ackerman geometry.
11. The front suspension assembly (200) as claimed in claim 3, wherein said split lower control arm configuration is configured such that said first link arm (112A, 112B) and said second link arm (114A, 114B) define a controlled pivoting angular displacement about their respective said first sub-frame hardpoints (106A, 106B), said first knuckle hardpoints (108A, 108B) and said second sub-frame hardpoints (105A, 105B), said second knuckle hardpoints (107A, 107B) in response to a longitudinal force acting on the wheel or tire contact patch, wherein:
• the angular displacement of said first link arm (112A, 112B) and said second link arm (114A, 114B) allows the wheel center to shift rearward, thereby reducing the transmission of impact forces to the vehicle structure;
• the relative movement between said first link arm (112A, 112B) and said second link arm (114A, 114B) enables gradual force absorption, thereby mitigating the effect of road irregularities; and
• the controlled pivoting angular displacement of said lower control arms (112A, 112B, 114A, 114B) maintains the stability of the suspension geometry, thereby preventing abrupt load variations on suspension components.
12. The front suspension assembly (200) as claimed in claim 11, wherein the controlled pivoting angular displacement of said first link arm (112A, 112B) and said second link arm (114A, 114B) in response to a longitudinal force acting on the wheel or tire contact patch causes a rearward displacement of the wheel center, wherein:
• the rearward displacement of the wheel center results in a dynamic adjustment of the wheel recess within the wheel arch under vertical loading conditions; and
• the suspension geometry regulates the turning circle diameter (TCD) by maintaining the required steering angles during wheel displacement.
13. The front suspension assembly (200) as claimed in claim 1, wherein said suspension assembly (200) is a McPherson-type configuration, wherein:
• said telescopic struts (115A, 115B) is configured as a structural load-bearing member, to provide both damping and lateral support to said steering knuckle (116A, 116B);
• said upper end of said telescopic strut (115A, 115B) is mounted to a hardpoint (110A) on the vehicle body to allow articulated movement of said steering knuckle (116A, 116B) about the kingpin axis (K1, K2); and
• said first link arm (112A, 112B) and said second link arm (114A, 114B) collectively define a lower guiding mechanism for said steering knuckle (116A, 116B), wherein their pivoting angular displacement facilitates controlled variation of said instantaneous center (W1, W2), thereby reducing the turning circle diameter (TCD).
14. The front suspension assembly (200) as claimed in claim 3, wherein:
• said first ball joint (111A, 111B) has a diameter in the range of 30 mm to 33 mm;
• said second ball joint (113A, 113B) has a diameter in the range of 34 mm to 36 mm; and
• the distance between the centre of said first ball joint (111A, 111B) and said second ball joint (113A, 113B) is in the range of 40 mm to 100 mm.

Dated this 25th Day of August 2025

_______________________________
MOHAN RAJKUMAR DEWAN, IN/PA – 25
OF R. K. DEWAN & CO.
AUTHORIZED AGENT OF APPLICANT