RACK AND PINION STEERING ASSEMBLY FOR RIGID FRONT
SUSPENSION VEHICLE
TECHNICAL FIELD OF THE INVENTION
[0001] The present invention generally relates to vehicle steering systems, and more particularly relates to a rack and pinion steering assembly for a rigid front suspension vehicle with an ideal rack position for optimum vehicle performance.
BACKGROUND OF THE INVENTION
[0002] The primary function of a steering system is to convert driver's input to angular rotation of front tires through a steering wheel. Additionally, the steering system must enable the driver ease in turning of the front tires in a certain path as expected. Commercial vehicles that are intended for off-road application incorporate a solid front axle for improved vehicle performance and life, hence a recirculating ball (RCB) steering system with rigid front suspension (RFS) is used. Similarly, passenger vehicles use a rack and pinion steering system with independent front suspension (IFS), in which the rack is coupled at respective ends to the front wheels. In IFS vehicle, the suspension moves to and fro through circular path when viewed from front side of the vehicle, whereas in RFS vehicle the wheel takes linear movement. Typically, it is observed that there is about 4 - 6mm more lateral movement of tie rod ends while the wheel moves through rebound to bump conditions in RFS vehicle compared with equivalent IFS vehicle.
[0003] Most prior art rack and pinion steering system is designed such that it follow a suspension path in a circular manner only and it cannot accommodate the tie rod lateral movements in RFS vehicle due to its inherent design. This leads to conflict in movements between the steering and suspension hard points. Additionally, there is variation in the tie rod true length which primes to issues like, predominant toe change, rack and pinion gear failure, oil leak due to seal damage, non-conformance Ackerman geometry, steering wheel kick back, vehicle pulling, reduced tire life and

higher force transfers to chassis which leads to reduced chassis life. In some prior systems, the rack and pinion steering gear is mounted on an un-sprung mass portion of the vehicle, hence the gear is subjected to heavy vibrations and direct impact loads from ground which leads to premature failures in the rack casing. Also, during vehicle dynamic condition through bump to rebound, the wheel experiences disturbances in toe variation due to difference in path followed by the tie rod and the front suspension which leads to major issues like the tire wear and a steering wheel kick back.
[0004] Therefore, it is necessary to provide for an improved rack and pinion steering assembly for a rigid front suspension vehicle with an ideal rack position along with ideal vehicle hard points for steering, suspension and front axle geometry for optimum vehicle performance, as described in greater detail herein.
SUMMARY OF THE INVENTION
[0005] The following summary is provided to facilitate an understanding of some of the innovative features unique to the disclosed embodiments and is not intended to be a full description. A full appreciation of the various aspects of the embodiments disclosed herein can be gained by taking the entire specification, claims, drawings, and abstract as a whole.
[0006] A primary objective of the present invention is to provide a rack and pinion steering assembly with rigid front suspension for small commercial vehicles, with an ideal rack position for optimum vehicle performance.
[0007] Another objective of the present invention is to provide a rack and pinion steering assembly for a rigid front suspension vehicle, which is capable of providing ideal hard points for front suspension and a front axle geometry based on a mathematical model.

[0008] Another objective of the present invention is to provide a rack and pinion steering assembly for a rigid front suspension vehicle, which is capable of providing an optimum suspension travel for improving the vehicle ride and handling.
[0009] Another objective of the present invention is to provide a rack and pinion steering assembly for a rigid front suspension vehicle, which is capable of providing an optimum tie rod and ball end angles for provided suspension movements in vehicle dynamic conditions based on a mathematical model.
[0010] According to the embodiment of the present invention to achieve the objective of the invention, a rack and pinion steering assembly for a rigid front suspension vehicle is disclosed which includes a tubular housing having an elongated rack with a plurality of rack teeth that extends along a longitudinal axis of the tubular housing. The rack is movable axially in the tubular housing in response to rotation of a vehicle steering wheel. A pinion shaft with a plurality of pinion teeth is attached to a steering shaft and carried in the rack inside the tubular housing and the pinion teeth is meshed with the rack teeth. A pair of mounting brackets connects the tubular housing on a chassis cross member in a sprung mass portion of the vehicle. A tie rod connects to each end of the rack that is protruding from the tubular housing via a ball joint. Each end of the tie rod is connected to a stub axle arm on an unsprung mass portion of the vehicle through the ball joint. The steering wheel is rotated so that a rotary motion is transferred to the tubular housing through a bevel gearbox and the steering shaft. The rotary motion at the pinion shaft is converted into a linear displacement of the rack which causes tires to rotate in a desired direction. The position of the rack along with vehicle hard points for steering, suspension, front axle geometry and design of the tie rod and ball ends are optimized based on a mathematical model.
[0011] The pinion teeth engaged with the rack teeth is varied by adjusting a screw that is engaged with the tubular housing. The rigid front suspension vehicle is composed of a stiffer leaf spring for holding a front axle beam in position to provide a

higher spring constant and arrest change in length of the tie rod. The leaf spring suspension is composed of a rubber eye bush that is placed between frame bracket and frame spring eye. The rack is composed of a stopper portion at its end and a cavity is created in the stopper portion for long travel of the steering rack and the pinion shaft. An outer wheel cut angle for a given inner wheel cut angle is determined from Ackerman condition of a typical steering system with a rack and pinion steering gear configuration.
[0012] The steering and the front axle hard points with the toe change value in the vehicle with respect to position and suspension travel of the rack is determined based on a steering arm length, a tie rod length, a rack casing length, a rack ball joint center to center length, a travel of rack, a distance between the front axle and the rack center axis, an Ackerman angle, distance between a kingpin center to an arm hard point. The rack and the pinion shaft linkage geometry at the inner wheel position and the outer wheel position is determined by considering symmetry of the center of vehicle. The toe change arising due to angle variation in the tie rod is determined based on the tie rod length, an initial inner and outer ball joint point, an optimized inner and outer ball joint, a tie rod lateral movement when the rack and the pinion is moved vertically, an initial and modified tie rod angle, an initial tie rod height across vertical direction, a downward movement in vertical direction, a toe change/side, a tire radius, an outer lock angle and an knuckle arm length.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] The disclosed embodiments may be better understood by referring to the figures, in which like reference numerals refer to identical or functionally-similar elements throughout the separate views, further illustrate the present invention and, together with the detailed description of the invention, serve to explain the principles of the present invention.
[0014] FIG. 1 illustrates a perspective view of a rack and pinion steering assembly, in accordance with an exemplary embodiment of the present invention;

[0015] FIG. 2 illustrates a perspective view of the rack and pinion steering assembly attached to a chassis cross member in a sprung mass portion of a rigid front suspension vehicle, in accordance with an exemplary embodiment of the present invention;
[0016] FIG. 3 illustrates an isometric view of the rack and pinion steering assembly attached to the vehicle, in accordance with an exemplary embodiment of the present invention;
[0017] FIG. 4 illustrates a perspective view of a tie rod with ball joints, in accordance with an exemplary embodiment of the present invention;
[0018] FIG. 5 illustrates a graphical representation of a spring characteristics comparison of independent front suspension (IFS) and rigid front suspension (RFS) vehicles, in accordance with an exemplary embodiment of the present invention;
[0019] FIG. 6 illustrates a graphical representation of a bump stopper characteristics comparison of the independent front suspension and rigid front suspension vehicles, in accordance with an exemplary embodiment of the present invention.
[0020] Fig 7 shows the toe change in the Rigid Front Suspension system with respect to rack position, in accordance to the present invention.
DETAILED DESCRIPTION OF THE INVENTION
[0021] The particular values and configurations discussed in these non-limiting examples can be varied and are cited merely to illustrate at least one embodiment and are not intended to limit the scope thereof.

[0022] In the following, numerous specific details are set forth to provide a thorough description of various embodiments. Certain embodiments may be practiced without these specific details or with some variations in detail. In some instances, certain features are described in less detail so as not to obscure other aspects. The level of detail associated with each of the elements or features should not be construed to qualify the novelty or importance of one feature over the others.
[0023] The claimed subject matter has been provided here with reference to one or more features or embodiments. Those skilled in the art will recognize and appreciate that, despite of the detailed nature of the exemplary embodiments provided here; changes and modifications may be applied to said embodiments without limiting or departing from the generally intended scope. These and various other adaptations and combinations of the embodiments provided here are within the scope of the disclosed subject matter as defined by the claims and their full set of equivalents. Like numbers refer to like elements throughout.
[0024] The present invention relates to a rack and pinion steering assembly with rigid front suspension in small commercial vehicles, with an ideal rack position. The rack and pinion steering assembly for a rigid front suspension vehicle provides ideal hard points for a front suspension and a front axle geometry. The present invention mainly focuses on optimum tie rod and ball end angles for provided suspension movements in vehicle dynamic conditions. Also, the present invention is capable of providing an optimum suspension travel for improving the vehicle ride and handling. The vehicle parameters such as rack position along with vehicle hard points for steering, suspension, front axle geometry, spring travel from unladen to overload, bump stop characteristics, suspension bush design and design of the tie rod and ball ends are optimized based on a mathematical model for better performance of the vehicle.

[0025] Referring to FIG. 1 a perspective view of a rack and pinion steering assembly (10), in accordance with an exemplary embodiment of the present invention. In general, a rack and pinion is a type of linear actuator that comprises a pair of gears which convert rotational motion into linear motion. A circular gear called "the pinion" engages teeth on a linear "gear" bar called "the rack"; rotational motion applied to the pinion causes the rack to move relative to the pinion, thereby translating the rotational motion of the pinion into linear motion. Note that in FIGS. 1-6, identical or similar parts or components are generally indicated by identical reference numerals.
[0026] In this present invention, the rack and pinion steering assembly (10) includes a tubular housing (12) having an elongated rack (14) with a plurality of rack teeth that extends along a longitudinal axis of the tubular housing (12). A pinion shaft (32) with a plurality of pinion teeth attached to a steering shaft (36) and carried in the rack (14) inside the tubular housing (12). The pinion teeth is meshed with the rack teeth and varied by adjusting a screw (35) that is engaged with the tubular housing (12). A tie rod (26) connects to each end of the rack (14) that is protruding from the tubular housing (12) via a ball joint (28).
[0027] Referring to FIG. 2 a perspective view of the rack and pinion steering assembly (10) attached to a rigid front suspension vehicle (15) is illustrated, in accordance with an exemplary embodiment of the present invention. The assembly (10) includes a pair of mounting brackets (20) that connects the tubular housing (12) in a sprung mass portion (24) of the vehicle (15). Generally, in a vehicle with a suspension, sprung mass (or sprung weight) is the portion of the vehicle's total mass that is supported above the suspension. The sprung weight typically includes the body, frame, the internal components, passengers, and cargo. In a preferred embodiment, the tubular housing (12) is mounted on a chassis cross member (22) via the mounting brackets (20), based on design consideration. The chassis cross member (22) is mounted to the vehicle (15) using a chassis mounting bracket (52).

[0028] Similarly, each end of the tie rod (26) is connected to an unsprung mass portion (31) of the vehicle (15) through the ball joint (28). Generally, in a ground vehicle with a suspension, the unsprung mass (or the unsprung weight) is the mass of the suspension, wheels or tracks (as applicable), and other components directly connected to them, rather than supported by the suspension (the mass of the body and other components supported by the suspension is the sprung mass). Unsprung mass includes the mass of components such as the wheel axles, wheel bearings, wheel hubs, tires, and a portion of the weight of springs, shock absorbers, and suspension links. In a preferred embodiment, ends of the tie rod (26) is connected to a stub axle arm (30), based on design consideration.
[0029] The rack (14) is movable axially in the tubular housing (12) in response to rotation of a vehicle steering wheel (18). The rack and pinion steering assembly (10) further includes a sealing bellows (45) fastened on one side to the housing (12) and on another side to the steering tie rods (26) at its both ends. The sealing bellows (45) axially expand and contract in response to the axial movement of the rack (14). The steering wheel (18) is rotated so that a rotary motion is transferred to the tubular housing (12) through a bevel gearbox (40) and the steering shaft (36). The rotary motion at the pinion shaft (32) is then converted into a linear displacement of the rack (14) which causes tires (44) to rotate in a desired direction. The rack (14) and the pinion shaft (32) converts the rotational motion of the steering wheel (18) into linear motion needed to turn the tires (44).
[0030] The rack (14) and the pinion shaft (32) provides a gear reduction, making it easier to turn the tires (44). In a preferred embodiment, the parameters of the vehicle (15) such as position of the rack (14) and design of the tie rod (26) and the ball joints (28) are optimized based on a mathematical model for better performance of the vehicle (15). The rack and pinion steering assembly (10) is mounted on the sprung mass portion (24) of the RFS vehicle (15) and ends of the tie rods (26) are connected to the un-sprung mass (31) through effective ball joints (28) having suitable degrees of freedom for better performance of the vehicle (15) and to arrest high toe change, as shown in FIGS. 3-4.

[0031] Referring to FIG. 4 a perspective view of the tie rod (26) with the ball joints (28), in accordance with an exemplary embodiment of the present invention. The steering rack (14) is connected to the tie rods (26) by the ball joints (28). The tie rods (26) are connected in a known manner by steering knuckles to respective tires (44) of the vehicle (15). The steering rack (14) is composed of a soft bump stopper portion (not shown) at its end and a cavity is created in the bump stopper portion for long travel of the steering rack (14) and the pinion shaft (32). The rigid front suspension vehicle (15) is composed of a stiffer leaf spring (46) that provides higher spring constant and arrest change in length of the tie rod (26). A rubber suspension eye bush (not shown) is placed between each leaf of the leaf spring (46) and compression of the bush take place when the leaf spring (46) operates in a lateral direction. The spring characteristics comparison of the independent front suspension (IFS) and the rigid front suspension (RFS) vehicles and the bump stopper characteristics comparison of the independent front suspension and rigid front suspension vehicles is shown in FIGS 6-7.
[0032] Fig 7 shows the toe change in the Rigid Front Suspension system with respect to rack position. The toe change with respect ot landen, unladen and overload onditions are analyzed for arriving the ideal rack position from the front axle.
[0033] The parameters of the vehicle (15) such as position of the rack (14), spring travel from unladen to overload, characteristics of the bump stopper, design of the suspension bushes and design of the tie rod (26) and ball joints (28) are optimized based on a mathematical model in order to arrest the high toe change. For example, typical layout of Ackerman Steering geometry which is most widely used in commercial vehicle is considered. Ackerman condition is satisfied when centers of both front tires (44) meet at a point on the rear axle which is turning point of the vehicle (15). Where, So = outer wheel angle, Si = inner wheel angle, W = front track width of the vehicle, B = distance between left and right kingpin centerline, L = wheel base of the vehicle.
[0034] Ideal Ackerman condition of typical steering system with a rack and pinion steering gear assembly can be expressed as shown below in equation (1):

[0035] From the above equation, the outer wheel cut angle (Sa) for a given certain inner wheel cut angle (Si) can be calculated. Mentioned below is a mathematical derivation of optimum hard points of the steering and the front axle (54) showing toe change values in the vehicle (15) with respect to position of the rack (14) and suspension travel. By considering symmetry of the center of vehicle (15), following expression can be derived for length of the tie rod (26):
[0036] Where, X= steering arm length, Y= tie rod length, p= rack casing length, p+2r = rack ball joint center to center length, q= travel of rack, d= distance between front axle and rack center axis, P= Ackerman angle, f=distance between kingpin center to arm hard point. The linkage geometry of the rack (14) and the pinion shaft (32) at inner wheel position is determined by considering symmetry of the center of the vehicle (15) and following expression can be derived for length of the tie rod (26) as shown in equation (3):
[0037] The linkage geometry of the rack (14) and pinion shaft (32) at outer wheel position is determined by considering symmetry of the center of vehicle (15) and following expression can be derived for length of tie rod (26) as illustrated in equation (4):
[0038] The mathematical model for toe change arising due to angle variation of the tie rod (26) is derived as follows. Where, ¥f= tie rod length when viewed from

front, h= initial inner ball joint point, °i= initial Outer ball joint point, h= optimized inner ball joint point, °:= optimized outer ball joint point, x=tie rod length across °iB, x\.= tie rod lateral movement when R&P is moved vertically, a= initial tie rod angle, KL = improved/modified tie rod angle, Z= initial tie rod height across vertical direction, S= downward movement in vertical direction, y= toe change/side, R= radius of tire, So =outer lock angle, X= knuckle arm length. Considering h50i and ^°2 following equations can be derived:
[0039] And deriving an expression for Hi :
[0040] By substitution of all the variables in above equation, toe change per side can be arrived as shown below in equation (7 and 8),
[0041] Note that the vehicle level refinements from the suspension bushes to upgradation of the tie rod (26) and the ball joints (28) is carried to adopt the rack and pinion steering assembly (10) in the RFS small commercial vehicle (15). Physical trails are conducted in the RFS vehicle (15) and optimized hard points are derived from the above mathematical equations along with the stated refinements and compared with that of equivalent IFS vehicle. It is proved that all the parameters and life of RFS vehicle are similar to that of IFS vehicle. The actual inner and outer cut angles of the RFS vehicle (15) are calculated, measured and compared with that of equalant IFS vehicle. The comparison of IFS and RFS wheel cut angles defines that the ideal angle for IFS and RFS are inline.

[0042] The measured actual wheel cut angle in RFS system is within 5% variation to the one as in IFS compared with vehicle. This clearly demonstrates Ackerman principle is matched while adopting the rack and pinion steering assembly (10) in the RFS vehicle (15). Toe change in the RFS vehicle (15) at different position of the rack (14) and the pinion shaft (32) shows that the lowest toe variation is observed at 5mm rack offset position from the front axle (54). Also toe variation is compared in physical vehicle and found closely matching with theoritical values. Tire life is monitored for initial and optimized position of the rack (14) and found that the tire life with 5mm rack position is inline with the IFS vehicle. Steering effort measuremet is carried out at vehicle level with power assist ON and OFF conditions for IFS and RFS configurations. The steering effort in RFS vehicle is 20% higher than IFS due to its suspension configuration. Rack force measurement in RFS vehicle is almost equal to IFS vehicle during dry park steering conditions.
[0043] The rack (14) and the pinion shaft (32) is mounted on the sprung mass portion (24) of the vehicle (15) to overcome the issues arising due to road shock loads and vibrations. Premature system failures, geometry issues and pilot fatigue is arrested by proper suspension and steering linkage design and compliance. The mathematical model is developed to attain optimum steering and suspension geometry and by solving these equations, optimum hardpoints is arrived to meet best Ackerman error, steering effort, toe change and rack force.
[0044] It will be appreciated that variations of the above-disclosed and other features and functions, or alternatives thereof, may be desirably combined into many other different systems or applications. Also that various presently unforeseen or unanticipated alternatives, modifications, variations or improvements therein may be subsequently made by those skilled in the art which are also intended to be encompassed by the following claims.

We Claim:
1. A rack and pinion steering assembly (10) for a rigid front suspension vehicle
(15), comprising:
a tubular housing (12) having an elongated rack (14) with a plurality of rack teeth that extends along a longitudinal axis of the tubular housing (12), the rack (14) is movable axially in the tubular housing (12) in response to rotation of a vehicle steering wheel (18);
a pinion shaft (32) with a plurality of pinion teeth attached to a steering shaft (36) and carried in the rack (14) inside the tubular housing (12) in such a way that the pinion teeth is meshed with the rack teeth; charactarized in that
a pair of mounting brackets (20) connects the tubular housing (12) on a chassis cross member (22) in a sprung mass portion (24) of the vehicle (15);
a tie rod (26) connects to each end of the rack (14) that is protruding from the tubular housing (12) via a ball joint (28) where each end of the tie rod (26) is connected to a stub axle arm (30) on an unsprung mass portion (31) of the vehicle (15) through the ball joint (28),
wherein upon rotation of the steering wheel (18) a rotary motion is transferred to the tubular housing (12) through a bevel gearbox (40) and the steering shaft (36) and the rotary motion at the pinion shaft (32) is converted into a linear displacement of the rack (14) which causes tires (44) to rotate in a desired direction.
2. The assembly of claim 1, wherein the rigid front suspension vehicle (15) is composed of a stiffer leaf spring (46) for holding a front axle beam (54) in position to provide a higher spring constant and decrease change in length of the tie rod (26).
3. The assembly of claims 1 and 3, wherein the leaf spring (46) is composed of a rubber suspension eye bush that is placed between frame bracket and eye compression of the bush take place when the leaf spring (46) operates in a lateral direction to arrest a high toe change.

4. The assembly of claim 1, wherein a part of the bump stopper is made a cavity for long travel of the rack and pinion assembly and improving the ride of the vehicle.
5. The assembly of claim 1, wherein an outer wheel cut angle for a given inner wheel cut angle is determined from Ackerman condition of a steering system with a rack and pinion steering gear configuration.
6. The assembly of claim 1, wherein the front axle (54) and steering hard points with a toe change value in the vehicle (15) with respect to position and suspension travel of the rack (14) is determined based on a steering arm length, a tie rod length, a rack casing length, a rack ball joint center to center length, a travel of rack, a distance between the front axle and the rack center axis, an Ackerman angle and distance between a kingpin center to an arm hard point.
7. The assembly of claim 1, wherein the rack (14) and the pinion shaft (32) linkage geometry at the inner wheel position and the outer wheel position is determined by considering symmetry of the center of vehicle (15).
8. The assembly of claim 1, wherein the toe change arising due to angle variation in the tie rod (26) is determined based on the tie rod length, an initial and optimized inner and outer ball joint point, a tie rod lateral movement when the rack and the pinion is moved vertically, an initial and modified tie rod angle, an initial tie rod height across vertical direction, a downward movement in vertical direction, a toe change, a tire radius, an outer lock angle and an knuckle arm length.