Abstract:
The present invention provides a passive vehicle suspension system providing optimal camber gain and avoiding the “bump camber” problem. This vehicle suspension system utilizes a conventional sway bar and a kinematic system to rotate the tire(s) of a vehicle in a direction opposite the roll of the vehicle during turning or cornering, for example. The kinematic design allows for adjustability and optimization, providing optimal camber gain. Although an automotive vehicle suspension system is illustrated and described, and, specifically, a racing or everyday automotive vehicle suspension system, the concepts of the present invention are capable of much broader applicability.

Description:
CROSS-REFERENCE TO RELATED APPLICATION(S) 
     The present patent application/patent is a continuation-in-part of U.S. patent application Ser. No. 11/875,128, (James Franklin CUTTINO et al.), filed on Oct. 19, 2007 now U.S. Pat. No. 7,896,361, and entitled “PASSIVE VEHICLE SUSPENSION SYSTEM PROVIDING OPTIMAL CAMBER GAIN,” which claims the benefit of priority of U.S. Provisional Patent Application No. 60/853,791 (James Franklin CUTTINO et al.), filed on Oct. 23, 2006, and entitled “PASSIVE SUSPENSION DESIGN FOR OPTIMAL CAMBER,” the contents of both of which are incorporated in full by reference herein. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates generally to vehicle suspension systems and the like. More specifically, the present invention relates to a passive vehicle suspension system providing optimal camber gain. 
     BACKGROUND OF THE INVENTION 
     Vehicle suspension systems and the like are designed to enhance the performance of modern vehicles, while maintaining a comfortable ride for passengers. Obviously, this performance-enhancement aspect is most important in racing applications, for example, while this comfort-maintenance aspect is most important in everyday applications, for example. These vehicle suspension systems constrain the position and orientation of a tire relative to a road surface. The maximum force that the tire can generate in a desired direction of travel is greatly influenced by the camber of the tire. Camber is the lean angle of the tire relative to the road surface, with the tire being perpendicular to the road surface under zero camber conditions. In general, the tire can generate the most straight-line acceleration and braking force under zero camber conditions, but can generate the most turning or cornering force with some amount of camber gain—the tire essentially “leaning into” the turn or corner. By maintaining zero camber conditions while traveling straight and leaning the tire into turns or corners, tire grip can be maximized and tire wear can be minimized. 
     Nearly all of the vehicle suspension systems currently available and used provide misapplied, insufficient, or no camber gain while turning or cornering. As a result, the tire rolls relative to the road surface—in the same direction that the vehicle rolls. This provides non-optimal camber for optimal tire grip. In addition, the majority of vehicle suspension systems currently available and used experience “bump camber.” Bump camber is a condition wherein the camber gain provided to compensate for vehicle roll, when provided, also results in the tire cambering during purely vertical motion of the vehicle, such as when the vehicle goes over a bump. Previous attempts to address these camber gain “issues” have typically involved the use of active devices, adding energy to the vehicle suspension systems via actuators and the like. Such vehicle suspensions systems are inherently complex and expensive. 
     Thus, what is needed in the art is a vehicle suspension system that passively provides optimal camber gain when a vehicle turns or corners—this camber gain being opposite in direction to the vehicle roll. What is also needed in the art is a vehicle suspension system that avoids the related bump camber problem. Ideally, this passive vehicle suspension system providing optimal camber gain and avoiding the bump camber problem would be relatively simple and inexpensive, making it suitable for both racing and everyday applications. 
     BRIEF SUMMARY OF THE INVENTION 
     In various exemplary embodiments, the present invention provides a passive vehicle suspension system providing optimal camber gain and avoiding the bump camber problem. This vehicle suspension system utilizes a conventional sway bar and a kinematic system to rotate the tire(s) of a vehicle in a direction opposite the roll of the vehicle during turning or cornering, for example. The kinematic design allows for adjustability and optimization, providing optimal camber gain. It should be noted that, although an automotive vehicle suspension system is illustrated and described herein, and, specifically, a racing or everyday automotive vehicle suspension system, the concepts of the present invention are capable of much broader applicability. 
     In one exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a first kinematic system associated with a first tire of a vehicle, wherein the first kinematic system is configured to adjust the camber of the first tire relative to a road surface in response to forces experienced by the vehicle; and wherein the passive vehicle suspension system operates without adding energy to the first kinematic system through an actuator. The passive vehicle suspension system also includes a second kinematic system associated with a second tire of the vehicle, wherein the second kinematic system is configured to adjust the camber of the second tire relative to the road surface in response to forces experienced by the vehicle. Typically, the first kinematic system is coupled to the second kinematic system. The first and second kinematic systems are configured to camber the first and second tires relative to the road surface into a turn or corner encountered by the vehicle, substantially opposite the direction of chassis roll. Typically, the camber of the first and second tires relative to the road surface is substantially equal. The first and second kinematic systems are also configured to maintain a substantially-parallel relationship between the first tire and the second tire during chassis heave. 
     In another exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a first upright member coupled to a first tire of a vehicle, wherein the first upright member cambers with the first tire relative to a road surface; a first lower control arm pivotably coupled to a chassis of the vehicle and the first upright member; a first bell crank pivotably coupled to the first upright member and the first lower control arm; and wherein rotation of the first bell crank causes the first upright member and the first tire to camber relative to the road surface. The passive vehicle suspension system also includes a first connector link pivotably coupled between the first bell crank and the first upright member. The passive vehicle suspension system further includes a second upright member coupled to a second tire of the vehicle, wherein the second upright member cambers with the second tire relative to the road surface; a second lower control arm pivotably coupled to the chassis of the vehicle and the second upright member; a second bell crank pivotably coupled to the second upright member and the second lower control arm; and wherein rotation of the second bell crank causes the second upright member and the second tire to camber relative to the road surface. The passive vehicle suspension system still further includes a second connector link pivotably coupled between the second bell crank and the second upright member. The passive vehicle suspension system still further includes a substantially-rigid sway bar rotatably coupled to the first and second bell cranks, wherein the substantially-rigid sway bar is rotatably coupled to the chassis of the vehicle. The passive vehicle suspension system still further includes a sway bar connector coupled between the substantially-rigid sway bar and each of the first and second bell cranks Finally, the passive vehicle suspension system includes an equal-camber bar rotatably coupled to the first and second upright members. 
     In a further exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a first kinematic system associated with a first tire of a vehicle, wherein the first kinematic system is configured to adjust the camber of the first tire relative to a road surface in response to forces experienced by the vehicle; a second kinematic system associated with a second tire of the vehicle, wherein the second kinematic system is configured to adjust the camber of the second tire relative to the road surface in response to forces experienced by the vehicle; and wherein the passive vehicle suspension system operates without adding energy to the first kinematic system through an actuator. Typically, the first kinematic system is coupled to the second kinematic system. The first and second kinematic systems are configured to camber the first and second tires relative to the road surface into a turn or corner encountered by the vehicle, substantially opposite the direction of chassis roll. Typically, the camber of the first and second tires relative to the road surface is substantially equal. The first and second kinematic systems are configured to maintain a substantially-parallel relationship between the first tire and the second tire during chassis heave. The first kinematic system includes: a first upright member coupled to the first tire of the vehicle, wherein the first upright member cambers with the first tire relative to the road surface; a first lower control arm pivotably coupled to a chassis of the vehicle and the first upright member; a first bell crank pivotably coupled to the first upright member and the first lower control arm; and wherein rotation of the first bell crank causes the first upright member and the first tire to camber relative to the road surface. The first kinematic system also includes a first connector link pivotably coupled between the first bell crank and the first upright member. The second kinematic system includes: a second upright member coupled to the second tire of the vehicle, wherein the second upright member cambers with the second tire relative to the road surface; a second lower control arm pivotably coupled to the chassis of the vehicle and the second upright member; a second bell crank pivotably coupled to the second upright member and the second lower control arm; and wherein rotation of the second bell crank causes the second upright member and the second tire to camber relative to the road surface. The second kinematic system also includes a second connector link pivotably coupled between the second bell crank and the second upright member. The passive vehicle suspension system also includes a substantially-rigid sway bar rotatably coupled to the first and second bell cranks, wherein the substantially-rigid sway bar is rotatably coupled to the chassis of the vehicle. The passive vehicle suspension system further includes a sway bar connector coupled between the substantially-rigid sway bar and each of the first and second bell cranks Finally, the passive vehicle suspension system includes as equal-camber bar rotatably coupled to the first and second upright members. 
     In a still further exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a left housing and a right housing, wherein each of the left and right housings includes a shock paddle configured to pivotably engage a shock or the like that is directly or indirectly coupled to a chassis of a vehicle; a torsional splitter bar disposed substantially concentrically within the left and right housings, wherein the torsional splitter bar includes a torsional splitter bar paddle on each end, wherein each of the torsional splitter bar paddles is coupled to and selectively biases a first linkage via a first elongate member; an equal camber bar disposed substantially concentrically within the left and right housings, wherein the equal camber bar includes an equal camber bar paddle on each end, wherein each of the equal camber bar paddles is coupled to and selectively biases the first linkage via a second elongate member; and a pivotable axle assembly coupled to the first linkage via a second linkage. Optionally, the torsional splitter bar is disposed substantially concentrically within the equal camber bar. Preferably, the equal camber bar paddles are disposed on opposite top and bottom surfaces of the equal camber bar. 
     In a still further exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a first system associated with a first tire of a vehicle, wherein the first system is configured to adjust the camber of the first tire relative to a chassis of the vehicle in response to forces experienced by the vehicle, and wherein the first system includes a link of fixed length and a link of variable length coupled to the first tire; and wherein the passive vehicle suspension system operates without adding energy to the first system through an actuator. The passive vehicle suspension system also includes a second system associated with a second tire of the vehicle, wherein the second system is configured to adjust the camber of the second tire relative to the chassis of the vehicle in response to forces experienced by the vehicle, and wherein the second system includes a link of fixed length and a link of variable length coupled to the second tire. Optionally, the first system is coupled to the second system. Optionally, the first and second systems are configured to camber the first and second tires relative to the road surface into a turn or corner encountered by the vehicle, substantially opposite the direction of chassis roll. Optionally, the camber of the first and second tires relative to the road surface is substantially equal. Optionally, the first and second systems are configured to maintain a substantially-parallel relationship between the first tire and the second tire during chassis heave. Optionally, the link of fixed length includes a plurality of connected members. Optionally, the link of variable length includes a plurality of connected members that one or more of translate, rotate, and pivot with respect to one another. In one specific embodiment, the first and second systems collectively include: a left housing and a right housing, wherein each of the left and right housings includes a shock paddle configured to pivotably engage a shock or the like that is directly or indirectly coupled to the chassis of the vehicle; a torsional splitter bar disposed substantially concentrically within the left and right housings, wherein the torsional splitter bar includes a torsional splitter bar paddle on each end, wherein each of the torsional splitter bar paddles is coupled to and selectively biases a first linkage via a first elongate member; an equal camber bar disposed substantially concentrically within the left and right housings, wherein the equal camber bar includes an equal camber bar paddle on each end, wherein each of the equal camber bar paddles is coupled to and selectively biases the first linkage via a second elongate member; and a pivotable axle assembly coupled to the first linkage via a second linkage. Optionally, the torsional splitter bar is disposed substantially concentrically within the equal camber bar. Optionally, the equal camber bar paddles are disposed on opposite top and bottom surfaces of the equal camber bar. 
     In a still further exemplary embodiment, the present invention provides a passive vehicle suspension system, including: a first system associated with a first tire of a vehicle, wherein the first system is configured to adjust the camber of the first tire relative to a chassis of the vehicle in response to forces experienced by the vehicle, and wherein the first system includes a link of fixed length and a link of variable length coupled to the first tire; a second system associated with a second tire of the vehicle, wherein the second system is configured to adjust the camber of the second tire relative to the chassis of the vehicle in response to forces experienced by the vehicle, and wherein the second system includes a link of fixed length and a link of variable length coupled to the second tire; and wherein the passive vehicle suspension system operates without adding energy to the first and second systems through an actuator. Optionally, the first system is coupled to the second system. Optionally, the first and second systems are configured to camber the first and second tires relative to the road surface into a turn or corner encountered by the vehicle, substantially opposite the direction of chassis roll. Optionally, the camber of the first and second tires relative to the road surface is substantially equal. Optionally, the first and second systems are configured to maintain a substantially-parallel relationship between the first tire and the second tire during chassis heave. Optionally, the link of fixed length includes a plurality of connected members. Optionally, the link of variable length includes a plurality of connected members. Optionally, the link of variable length includes a plurality of connected members that one or more of translate, rotate, and pivot with respect to one another. In one specific embodiment, the first and second systems collectively include: a left housing and a right housing, wherein each of the left and right housings includes a shock paddle configured to pivotably engage a shock or the like that is directly or indirectly coupled to the chassis of the vehicle; a torsional splitter bar disposed substantially concentrically within the left and right housings, wherein the torsional splitter bar includes a torsional splitter bar paddle on each end, wherein each of the torsional splitter bar paddles is coupled to and selectively biases a first linkage via a first elongate member; an equal camber bar disposed substantially concentrically within the left and right housings, wherein the equal camber bar includes an equal camber bar paddle on each end, wherein each of the equal camber bar paddles is coupled to and selectively biases the first linkage via a second elongate member; and a pivotable axle assembly coupled to the first linkage via a second linkage. Optionally, the torsional splitter bar is disposed substantially concentrically within the equal camber bar. Optionally, the equal camber bar paddles are disposed on opposite top and bottom surfaces of the equal camber bar. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The present invention is illustrated and described herein with reference to the various drawings, in which like reference numbers are used to denote like system components and/or method steps, as appropriate, and in which: 
         FIG. 1  is a graph illustrating lateral force as a function of slip angle for various tire cambers, including zero camber, a tire leaning into a turn or corner, and a tire leaning out of a turn or corner, as well as the occurrence of camber thrust; 
         FIG. 2  is a graph illustrating maximum lateral force, F y max, as a function of camber in the direction of a turn or corner, highlighting the maximum lateral force peak for a light-weight vehicle, a medium-weight vehicle, and a heavy-weight vehicle, respectively; 
         FIG. 3  is a graph illustrating the ratio of tire lateral force to tire normal load as a function of tire normal load; 
         FIG. 4  is a schematic diagram illustrating both positive and negative camber caused by the vertical displacement of the body of a vehicle (a double A-arm suspension being illustrated); 
         FIG. 5  is a schematic diagram illustrating the tire camber caused by body roll when a double A-arm suspension is used in a vehicle, the body roll undesirably exceeding the tire camber when both are in the same direction; 
         FIG. 6  is a semitransparent perspective view of one exemplary embodiment of the passive vehicle suspension system of the present invention; 
         FIG. 7  is an isometric front view of the passive vehicle suspension system of  FIG. 6 ; 
         FIG. 8  is a partial isometric front view of the passive vehicle suspension system of  FIGS. 6 and 7 ; 
         FIG. 9  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-8 , the passive vehicle suspension system operating while the vehicle is in a turn or corner; 
         FIG. 10  is a semitransparent perspective view of the passive vehicle suspension system of  FIGS. 6-9 , highlighting the installation of a sway bar (anti-roll bar) that forms an important component of the passive vehicle suspension system; 
         FIG. 11  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-10 , highlighting the installation of an equal-camber mechanism that forms an important component of the passive vehicle suspension system; 
         FIG. 12  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-11 , the passive vehicle suspension system operating while the vehicle is in a turn or corner, highlighting the travel of the passive vehicle suspension system during body roll; 
         FIG. 13  is a partial isometric front view of the passive vehicle suspension system of  FIGS. 6-12 , the passive vehicle suspension system operating while the vehicle is in a turn or corner, highlighting the travel of the left side of the passive vehicle suspension system during body roll; 
         FIG. 14  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-13 , the passive vehicle suspension system operating while the vehicle is in a turn or corner, shown with the body orientation fixed and the substantially-rigid sway bar in place; 
         FIG. 15  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-14 , the passive vehicle suspension system operating while the vehicle is in a turn or corner, showing the torsional displacement of a substantially-non-rigid sway bar and the corresponding relative displacement of the bell crank ends; 
         FIG. 16  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-15 , the passive vehicle suspension system operating while the vehicle is in a turn or corner, showing the (theoretical) torsional displacement of a substantially-non-rigid sway bar and the corresponding relative displacement of the bell crank ends; 
         FIG. 17  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-16 , highlighting the non-installation of the equal-camber mechanism that forms an important component of the passive vehicle suspension system; 
         FIG. 18  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-17 , also highlighting the non-installation of the equal-camber mechanism that forms an important component of the passive vehicle suspension system; 
         FIG. 19  is an isometric front view of the passive vehicle suspension system of  FIGS. 6-18 , further highlighting the non-installation of the equal-camber mechanism that forms an important component of the passive vehicle suspension system; 
         FIG. 20  is a series of isometric front views of the passive vehicle suspension system of  FIGS. 6-19 , the passive vehicle suspension system operating while the vehicle experiences heave conditions, highlighting the rotation of the bell cranks and sway bar such that zero camber is achieved; 
         FIG. 21  is a semitransparent perspective view of an alternative exemplary embodiment of the passive vehicle suspension system of the present invention; 
         FIG. 22  is a top isometric view of the alternative exemplary embodiment of the passive vehicle suspension system of  FIG. 21 ; and 
         FIG. 23  is a front isometric view of the alternative exemplary embodiment of the passive vehicle suspension system of  FIGS. 21 and 22 . 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     The vehicle suspension system of the present invention is designed and configured to optimize tire contact with the road surface. This vehicle suspension system constrains the tires (left and right, whether front or back) to be flat on the road surface when traveling straight or braking, and leans the tires into a turn or corner during lateral acceleration, for example, without the use of active devices. In addition, the left and right suspensions can displace independently. The vehicle suspension system is compared to conventional double A-arm suspensions, McPherson strut suspensions, and the like, as appropriate. A specific conceptual design is presented herein. This specific conceptual design is one possible conceptual design and illustrates how some of the various packaging issues can be resolved. It will be readily apparent to those of ordinary skill in the art, however, that other conceptual designs can be utilized. 
     In general, high-performance vehicle suspension systems are designed and configured to maximize tire forces in order to obtain the most acceleration in a desired direction of travel, for example. These vehicle suspension systems must, however, be designed and configured to accommodate a given tire&#39;s performance characteristics. Most pneumatic tires have similar performance characteristics, such as the effect that camber has on maximum lateral force. For example, if the tire is leaning into a turn or corner, it will generate more lateral force than if the tire is leaning out of the turn or corner. This is illustrated in  FIG. 1 , a top-down view, which shows zero camber  10 , the tire leaning into the turn or corner  12 , and the tire leaning out of the turn of corner  14  as a function of slip angle and lateral force. Tire camber will often generate a lateral force even if the tire is steered in a forward direction of travel with zero slip angle, an effect which is referred to as “camber thrust”  16 . It should be noted that the slip angle  18  is the angle between the direction of travel  19  of the tire  15 , about the axis of rotation  20 , and the forward tire vector  21 . The lateral force  22  lies along the axis of rotation  20  of the tire  15 . 
     The amount of camber gain, in the direction of a turn or corner, required to maximize the tire&#39;s lateral force depends upon many factors, such as the given tire&#39;s performance characteristics and the normal load on the tire. The addition of camber, leaning the tire into the turn or corner, typically will not increase the heat generated by the tire, but will improve the tire&#39;s performance. Overheating a tire can cause increased tire wear and degraded tire performance. Thus, the addition of too much or too little camber can result in less than optimal lateral force. This is illustrated in  FIG. 2 , which shows maximum lateral force as a function of camber in the direction of a turn or corner, highlighting the maximum lateral force peak for a light-weight vehicle  24 , a medium-weight vehicle  26 , and a heavy-weight vehicle  28 , respectively. 
     Maximum tire longitudinal force, required during acceleration or braking, is typically produced with nominally zero tire camber. The maximum tire longitudinal force is a function of many parameters, including, but not limited to, tire normal load, F z , tire temperature, and road conditions. As the normal load increases, the maximum lateral and longitudinal forces also increase. However, the ratio between maximum lateral force and normal load (F y /F z ), often considered the coefficient of friction of the tire, is not constant, as the normal load varies and almost always has a negative slope. This is illustrated in  FIG. 3 . 
     Therefore, as the mass of the vehicle decreases, the maximum lateral or longitudinal acceleration of the vehicle typically increases while using the same set of tires. This same trend can be seen for tires used off-road, on dirt or gravel surfaces. In the case of an unsmooth level road surface, the normal force is changing but the mass of the vehicle remains the same. If the normal force can be kept more constant, maximum lateral or longitudinal forces can be achieved on average. A more constant normal force over unsmooth level road surfaces can be achieved by utilizing a relatively light-weight unsprung suspension mass (i.e. wheels, tires, etc.) relative to the relatively heavy-weight sprung mass of the vehicle (i.e. chassis, etc.) and utilizing relatively low spring stiffness, assuming the suspension does not reach its travel limits. 
     Referring to  FIG. 4 , camber is defined as the angle  31 , 33  between the tire  15  and the road surface  17 . Zero camber means that the tire  15  is perpendicular to the road surface  17 . A positive camber scenario  30  exists when the tops of the tires  15  are leaning outwards and away from the vehicle. A negative camber scenario  32  exists when the tops of the tires  15  are leaning inwards and towards the vehicle. The camber illustrated is, in this case, caused by vertical displacement of the of the body  34  of the vehicle (a double A-arm suspension  36  being illustrated). Some negative camber is usually utilized by performance vehicles for increased lateral force potential. 
     Double A-arm suspensions, McPherson strut suspensions, and live-axle with leaf spring suspensions are very commonly used suspensions for vehicles, the last being very commonly used in truck rear ends. DeDion suspensions, swing arm suspensions, and multi-link suspensions are also used, but are somewhat less common. Each suspension has it own advantages and disadvantages, making it more or less desirable in a given application. 
     Some suspensions maintain zero tire camber during body roll and heave, thereby optimizing tire capabilities during acceleration and braking Live-axle suspensions and DeDion suspensions both exhibit this behavior. An additional advantage of DeDion suspensions is that they have less unsprung mass, which is beneficial to tire performance over unsmooth level road surfaces. Live-axle suspensions, however, use fewer parts and are cheaper to produce. An equal-length double A-arm suspension also maintains zero tire camber during heave, but, unlike live-axle and DeDion suspensions, the tires camber in the wrong direction during body roll. Double A-arm suspensions are also relatively light weight as compared to live-axle suspensions. 
     Thus, most vehicle suspension systems camber the tire(s) during body roll. It is possible to camber the tire(s) in the direction of the turn or corner using a double A-arm suspension, but this camber gain is typically less than that which would be optimal (i.e. less than the body roll). This is illustrated in  FIG. 5 , which shows the body  34  of the vehicle rolling to a greater degree than either of the tires  15  are cambering, although the two are linked by the double A-arm suspension  36 . The disadvantage of the double A-arm suspension geometry is the excessive camber during heave, as previously shown in  FIG. 4 . The double A-arm either performs well during body roll or during heave, but typically does not perform both well during both. Thus, most double A-arm suspension geometries are compromised to perform less than optimally during both body roll and heave. 
     McPherson strut suspension systems typically camber the tire(s) in the wrong direction during body roll, and also camber the tire(s) during heave. However, McPherson struts allow for a variety of mounting configurations, allowing for more packaging options. McPherson strut suspension systems typically do not optimize tire performance, and are similar to the compromised double A-arm suspension systems in this respect. McPherson strut suspension systems rely on a strut, which is a translational joint, to resist the bending moment created by tire forces. The bending moment on this strut creates frictional forces which resist the motion of the suspension, resulting in decreased performance and poor ride quality. Design improvements have been proposed, but create additional packaging issues. 
     Multi-link suspensions are more complicated than the other typical suspensions, and still do not optimize tire camber during both body roll and heave. Multi-link suspensions offer benefits in terms of the various roll center and pitch center locations, and incorporate compliance or camber steer, if so desired. 
     In general, the vehicle suspension system of the present invention provides optimal camber gain in the direction of a turn or corner during body roll and provides substantially zero camber during heave. This vehicle suspension system weighs less than conventional live-axle suspensions, for example. The camber gain of the vehicle suspension system can also be changed by changing the associated mounting locations, without the need for additional parts. 
     Referring to  FIGS. 6-8 , in one exemplary embodiment, the passive vehicle suspension system  40  of the present invention is shown in a half-car model. It will be readily apparent to those of ordinary skill in the art that this passive vehicle suspension system  40  can be implemented in either the front or the rear of a vehicle with drive axles, steering, brakes, etc. (none of which are shown for purposes of clarity). Preferably, the passive vehicle suspension system  40  is symmetrical from left to right. It is possible, however, to design an asymmetrical configuration for specific applications, and this possibility is contemplated by the present invention. Each individual part of the passive vehicle suspension system  40  is described in detail herein below in order to demonstrate how the passive vehicle suspension system  40  is assembled and works in operation. The sprung mass of the vehicle, or chassis  42 , is represented conceptually by a box in the center of the vehicle. The left and right tires  15  (front or back) are also shown ( FIGS. 6 and 7 ). 
     A lower control arm  44  connects an upright member  46  and a bell crank  48  to the chassis  42 . Both the upright member  46  and the bell crank  48  can rotate relative to the lower control arm  44 . Although the lower control arm  44  is represented conceptually as a substantially-triangular structure, it can have any suitable shape and can consist of multiple parts. Similarly, although the bell crank  48  is represented conceptually as a substantially-triangular structure, it can have any suitable shape and can consist of multiple parts. A spring (not shown) is also connected to the lower control arm  44  which produces a moment on the lower control arm  44  that supports the weight of the vehicle. In operation, the bell crank  48  changes the camber of the tire(s)  15  by rotating. The bell crank  48  is attached to the lower control arm  44  by a pivot joint  50  that allows one degree of freedom between the bell crank  48  and the lower control arm  44 . The bell crank  48  also has an upright connecting link, referred to herein as the upright member-bell crank connector  52 , which has one degree of freedom relative to the bell crank  48 . A rotational joint  54  ( FIGS. 6 and 7 ) with three degrees of freedom is used to attach the sway bar-bell crank connector link  56  ( FIGS. 6 and 7 ) to the bell crank end  58  ( FIG. 8 ). 
     Referring to  FIG. 9 , the left bell crank end  58  is being pushed upwards by the left sway bar-bell crank connector link  56 , causing the left tire  15  to lean into the turn or corner. Conversely, the right bell crank end  58  is being pushed downwards by the right sway bar-bell crank connector link  56 , causing the right tire  15  to also lean into the turn or corner. 
     Ideally, the tire(s)  15  have one degree of freedom relative to the upright member  46 , allowing rotation only about the spindles&#39; axes. Each upright member  46  is connected to both the lower control arm  44  and the upright member-bell crank connector  52 . The upright member  46  is typically coupled to a steering mechanism, etc., which is/are not shown because it/they are not necessary for understanding the operation of the passive vehicle suspension system  40 . Pivot joints  60 , 62  connecting the upright member  46  to the lower control arm  44  and the upright member-bell crank connector  52 , respectively, have only one degree of freedom in the model presented to simplify things and to, in operation, keep the tire(s)  15  steering straight ahead. The upright members  46  are also connected to an equal-camber mechanism  64 , which is described in greater detail herein below. 
     Referring to  FIG. 10 , a sway bar (anti-roll bar)  66  is connected to the bell cranks  48  with perpendicular sway bar connectors  68 . The sway bar  66  is a torsional spring, with lever arms located at each end of the sway bar  66 . When a force is applied to one of the lever arms, the force is counteracted by an opposing force at the other lever arm. This opposing force, which creates a moment in the torsional spring, causes the torsional spring to twist. This rotation results in a deflection of the lever arms. The sway bar  66  can only rotate relative to the chassis  42  because it is attached to the chassis  42  with rotational joints (not illustrated). For understanding the operation of the passive vehicle suspension system  40 , it is assumed that the sway bar  66  has a very large torsional stiffness, and, therefore, the deflection of the lever arms can be considered to be negligible. The sway bar  66 , as designed and configured herein, does not “constrain” the suspension like a conventional sway bar does. Differences in left/right suspension travel are possible even with this rigid sway bar  66  because the bell cranks  48  can rotate. A conventional sway bar is typically connected to the lower control arms  44 . 
     Referring to  FIG. 11 , the equal-camber mechanism  64  is a series of four bar members which keep the left and right upright members  46  parallel. The four bar members include a left bar member  70 , a center bar member  72 , a right bar member  74 , and two bar members which are parallel to the upright members  46 , called equal-camber bar connectors  76 . All of the bar members have rotational joints on their ends, allowing each of the bar members to rotate relative to one another. It is important to note that the center bar member  72  is not fixed to the chassis  42 . The center bar member  72  is free to move left and right relative to the chassis  42 , which allows the upright members  46  to rotate relative to the chassis  42 . 
     Referring to  FIG. 12 , when a vehicle begins to turn or corner, the lateral acceleration normally causes the chassis  42  to roll. If the vehicle is turning left, the chassis  42  rolls out of the turn or corner. In order for this to happen, the left suspension must travel a different amount than the right suspension. Typically, the right suspension compresses the spring (not illustrated) causing the right tire  15  to travel upwards into the wheel well (not illustrated). The left suspension travels downwards in a similar manner.  FIG. 12  shows chassis roll viewed from behind with the vehicle turning left. 
     Suspension travel only occurs when the associated lower control arm  44  rotates relative to the chassis  42 . The amount of rotation of the lower control arm  44  is a convenient way to describe suspension travel for this design and configuration. Notice the displacement of the right tire  15  into the road surface in  FIG. 12 , which represents the deflection of the right tire  15  due to the increased weight transfer during lateral acceleration. 
     Referring to  FIG. 13 , viewing half of the passive vehicle suspension system  40  provides a clear understanding of how it works. The chassis  42  is held at a constant location and orientation so that movements relative to the chassis  42  are more easily identified. The left lower control arm  44  rotates counter-clockwise relative to the chassis  42 . This results in the downward displacement of the left upright member  46 , the left tire  15 , and the left end  78  of the left bell crank  48 . However, the right end  58  of the left bell crank  48  is unable to move downwards due to the left sway bar-bell crank connector link  56  causing the left bell crank  48  to rotate. This rotation of the left bell crank  48  results in the rotation of the left upright member  46 , and therefore camber gain is achieved. The right passive vehicle suspension system (not illustrated) works in exactly the same manner, except that the displacements are in the opposite direction. 
     Referring to  FIG. 14 , viewing the passive vehicle suspension system  40  as a whole, the ends  58  of the bell cranks  48  are attached to the sway bar-bell crank connector links  56 . For this example, a constraint is placed on both of the bell crank ends  58  by the substantially-rigid sway bar  66  to be at the same height relative to the chassis  42 . Although the height of the sway bar connectors  68  must be the same for an infinitely-rigid sway bar  66 , the height can vary, especially during vehicle heave. The moment in the sway bar  66  is caused by the displacement differences in height between the left and right bell crank ends  58 . 
     Referring to  FIG. 15 , if the sway bar  66  is allowed to twist, the left and right bell crank ends  58  will be at different heights relative to the chassis as  42 . In general, less camber gain is possible with a substantially-non-rigid sway bar  66  due to greater allowed deflections of the left and right sway bar connectors  68 , which results in height differences in the left and right bell crank ends  58  relative to the chassis  42 . Because the right bell crank  48  is pulling upwards on the right sway bar connector  68  and the left bell crank  48  is pushing downwards on the left sway bar connector  68 , the sway bar  66  deflects accordingly. 
     Referring to  FIG. 16 , if the sway bar  66  is allowed to twist in the opposite direction (which is not a likely scenario because the tire lateral forces are always pulling upwards on the right sway bar connector  68  while turning left), the bell cranks  48  would rotate the opposite direction. This would provide even more camber with respect to the road surface, but likely is not possible to achieve without the active controllers. It should be noted that camber gain is adjustable with bellcrank geometry. 
     Referring to  FIG. 17 , the removal of the equal-camber bar  64  ( FIG. 11 ) removes a constraint on the upright members  46 . Because the upright members  46  are no longer required to be parallel, both tires  15  can camber in and out together if the sway bar  66  rotates relative to the chassis  42 . Assuming that the vehicle is traveling straight, with no lateral accelerations, it is possible to have positive or negative camber. If the sway bar connectors  68  rotate upwards, then both tires  15  have positive camber. This positive camber is due to the fact that both bell crank ends  58  move upwards, causing both upright members  46  to rotate equal amounts, but in opposite directions. 
     Referring to  FIG. 18 , if the sway bar connectors  68  twist downwards, then both tires  15  have negative camber. Toe-in and toe-out can cause these situations to occur, as well as road surface variations, if the equal-camber bar  64  ( FIG. 11 ) is removed. FIGS.  17  and  18  illustrate the effects that the equal-camber bar  64  has on the height of the bell crank ends  58  with respect to the chassis  42 . 
     Referring to  FIG. 19 , during lateral acceleration, the removal of the equal-camber bar  66  results in an overall less stable system. Tire forces have a major influence on the reaction of the vehicle suspension system  40 , as well as initial camber conditions before entering a turn or corner. Without the equal-camber bar  66 , it is possible to have positive camber for both tires  15  during lateral acceleration. Positive camber is not desirable for the outside tire  15 , for example, but is desirable for the inside tire  15  for maximum performance. By understanding what happens if the equal-camber bar  66  is absent, it is easier to predict when force arises in the equal-camber bar  66 . 
     Referring to  FIG. 20 , vertical chassis displacement relative to the road surface is called “heave.” Heave typically occurs when traveling over a speed bump, cresting a hill, etc. Unlike most vehicle suspension systems, the design and configuration of the present invention allows for the tires  15  to maintain zero camber change during heave. Track-width changes do occur with heave; however, they can be minimized by utilizing a long, horizontal lower control arm  44 . Notice the rotation of the bell crank  48  during heave. The upright member  46 , lower control arm  44 , bell crank  48 , and upright member-bell crank connector  52  create a four-bar mechanism. As the lower control arm  44  rotates closer to the upright member  46 , this moves the bell crank  48  closer to the upright member  46 , resulting in bell crank rotation. A smaller bell crank vertical length would exaggerate this rotation, which can be desirable in certain applications. 
     An important consideration for this design and configuration is the direction that the chassis  42  rolls during lateral acceleration. Normally, vehicle suspension systems roll out of the turn or corner, but it is possible, under extreme conditions, for the chassis  42  to roll into the turn or corner. As long as the roll-center height of the vehicle is located below the center of gravity of the chassis  42 , the vehicle will roll out of the turn or corner. However, if the roll-center height of the vehicle is located above the center of gravity of the chassis  42 , the vehicle will roll into the turn or corner. If the chassis  42  rolls into the turn or corner, the tire camber will be in the wrong direction when using the vehicle suspension system  40  illustrated and described. 
     As the center of gravity height increases while the static roll-center remains at the same height, the roll moment will increase for a given lateral acceleration. The increased roll moment will result in more roll. As roll increases, the camber also increases in the preferred direction. Controlling the amount of chassis roll also controls the amount of camber during lateral acceleration. The particular geometry chosen greatly influences the camber per degree of roll. 
     It is a common misconception to believe there is more weight transfer to the outside tire  15  in a turn or corner if a vehicle rolls more. Steady-state weight transfer to the tires  15  is purely a function of center of gravity location and track width for a half-vehicle model. Roll can cause the center of gravity to move slightly, but generally the effect is negligible. Therefore stiffening springs or the sway bar  66  in a vehicle can decrease roll, but will not significantly affect the steady-state weight transfer of a symmetric half-vehicle model under steady-state conditions. Springs, sway bars, and dampers only affect the rate at which weight transfer occurs, not final steady-state weight transfer. 
     Referring to  FIGS. 21 and 22 , in an alternative exemplary embodiment of the passive vehicle suspension system  80  of the present invention, an alternative mechanical configuration, or mechanical “package,” that performs the functions and achieves the results described above includes a housing  82  including an attached or integrally-formed shock paddle  84  ( FIG. 21 ) that is configured to pivotably engage a shock  86  or the like. This shock  86  is directly or indirectly coupled to the chassis of the vehicle. It should be noted that, at present, only one “side” of the passive vehicle suspension system  80  is being described. The other “side” of the passive vehicle suspension system  80  is substantially identical. A torsional splitter bar  88  and an equal camber bar  90  are disposed within the pair of housings  82  (as illustrated in  FIG. 22 ). In the alternative exemplary embodiment illustrated, the torsional splitter bar  88  is disposed concentrically within the equal camber bar  90 , both of which are disposed concentrically within the housing  82 . Each of the torsional splitter bar  88  and the equal camber bar include an attached or integrally formed paddle  92  and  94  ( FIG. 21 ), with both of the torsional splitter bar paddles  92  being disposed substantially on the “top” side of the torsional splitter bar  88 , while one of the equal camber bar paddles  94  is disposed substantially on the “top” side of the equal camber bar  90  and the other of the equal camber bar paddles  94  is disposed substantially on the “bottom” side of the equal camber bar  90  (as illustrated in  FIG. 22 ). The torsional splitter bar paddles  92  and the equal camber bar paddles  94  are each connected, via an elongate member  95  and  95 , to and selectively bias a first linkage  96 , which is, in turn, connected to and selectively biases a second linkage  98 . This second linkage  98  is operable for selectively cambering an axle assembly  100  ( FIG. 21 ) that rotates about a pair of hinge pins  102  ( FIG. 21 ). 
       FIG. 3  illustrates the passive vehicle suspension system  80  in a positive camber scenario. 
     Although the present invention has been illustrated and described herein with reference to preferred embodiments and specific examples thereof, it will be readily apparent to those of ordinary skill in the art that other embodiments and examples can perform similar functions and/or achieve like results. All such equivalent embodiments and examples are within the spirit and scope of the present invention, are contemplated thereby, and are intended to be covered by the following claims.