Home

Current Status

L6B Specifications

Simulation Results

Kit Description

Features

Design Articles

   Electrical System

   Suspension System

Photo Gallery

You are visitor

since October 2005

LAWRENCE TECHNOLOGIES Brighton, Ontario, Canada

Design Article: Suspension Systems and the L6B
A Bit of History, A bit of Hype, and a Lot of How To
by Pete Lawrence

Of the many questions received by e-mail over the past few months, the most popular subject is the Suspension System. Questions have covered topics ranging from the basic configuration and geometry to mechanical design and implementation. The following description provides answers to most of those questions (and a few more we haven’t received yet). In this article, we will not only tell you how the L6B suspension is designed, we’ll tell you why it was designed that way.

Design Philosophy

Perhaps surprisingly, the design of the L6B suspension does not use the latest state-of-the-art techniques. There are no active elements or feedback loops; there are no adaptive control systems and no embedded computers (although a desktop computer was used extensively in the design and optimization of the suspension geometry). And you won’t find any rear wheel steering (intentional or otherwise). The design philosophy is based on three fundamental principles: (a) keep it simple, (b) use the best proven and understood techniques that don’t violate the first principle, and (c) continue refining the design until it is optimized.

Suspension Systems: History and Development

A study of the history of automotive suspension design shows a fairly gradual improvement in performance from the days of the horseless carriage until about 1960. Then, during the ‘60s and ‘70s, suspension system design improved rapidly and the results were seen in both Formula 1 and the beastly but marvelous CanAm cars that dominated the North American racing scene at that time. Since then, suspension systems have been refined and many new concepts have been developed, but the resulting performance improvements have been relatively small. Interestingly, in a head-to-head comparison between the 1970 McLaren M6 GT and the ultra-modern, million-dollar McLaren F1, the thirty year old M6 came out on top in virtually every performance category except top speed (Ref. Article by Jacques Duval & Paul Frère). In another interesting example of the superiority of good basic design over highly touted technological advances, the lap record at Mosport, set by Hans Stuck driving a Porsche 962C in 1985, stood unbroken for over fifteen years before it was finally surpassed by the remarkable Audi R8 just a few months ago.

While the monster engines that proliferated the CanAm era certainly contributed to the rapid rise in performance, the most significant factor was the unequal length wishbone suspension – not its invention (that occurred many years earlier), but its development and optimization. Improvements came at first by experimentation and trial-and-error and later (as designers began to understand the effects of subtle changes in geometry) by the application of mathematics and science. In the many years since, most successful race car suspensions have been firmly based on the unequal length wishbone designs of thirty years ago. During that same period, production car suspensions have migrated toward the ubiquitous McPherson strut, a clever compromise that permits reasonable performance at low cost, but one that can never match the performance or adjustability of dual wishbones.

Four developments in the years since those glorious CanAm days have made it much easier to design a high performance suspension system. The first is the continually growing proliferation of information. Thanks to writers like Carrol Smith, anyone with enough ambition to try, can find all kinds of data and recommendations that were just not available to the legends like Colin Chapman and Bruce McLaren. Having a good starting point for a design can easily eliminate ninety percent of the problems.

Secondly, the advent of the personal computer has made it possible to analyze suspension geometry without cutting and welding (and without resorting to the paper dolls of a decade ago). Ideas that at one time would have taken several days or even weeks to assess, can now be evaluated and accepted or rejected in a matter of minutes. The third advancement, the emergence of computer based design (rather than analysis) tools is, essentially, a combination of the first two. By merging the available knowledge with computer simulations, we no longer have to dream up a configuration and then analyze it to check its performance, then change something and check again to see if we improved it. We can now decide, up front, how we want the suspension system to behave, then ask the computer to reassign the variables to make it happen.

The fourth development to facilitate suspension design for specialty cars is a bit less obvious. It is, surprisingly, the universal popularity of front wheel drive, a phenomenon initiated by the much maligned, but always loved, Mini. This is certainly not intended as a commendation for front wheel drive (a concept that has few merits aside from low manufacturing cost) but its popularity has resulted in greatly improved designs, reliability and availability of constant velocity joints. With a plethora of CV joints to choose from, we can now design an unequal length wishbone suspension for the rear wheels almost as easily as for the front.

The L6B Suspension System

General

As you have no doubt guessed by now, the L6B uses an unequal length wishbone configuration at both the front and the rear. It is, in reality, quite simple and differs from a thousand other designs in only two significant ways: (1) it has been computer designed to optimize the performance of the L6B chassis and (2) it has been implemented in a way that allows the builder to adjust and tune the performance (within a limited range) to suit his driving preferences or a particular racing circuit.

Unequal Length Wishbone Suspension, Front View

Fundamentals

The L6B design started from the premise that long wishbones are better than short ones. Aside from their current popularity (have you looked at a Formula 1 car lately?), long (especially very long) wishbones allow a relatively large up or down wheel displacement with a correspondingly small change in the angular movement of the wishbone. This makes it much easier to design the geometry to obtain the desired smooth and predictable changes in camber as the wheel moves through its range of travel. A suspension designed with short wishbones may work reasonably well over a limited range of movement, then suddenly change its characteristics as it nears the end of its travel. The L6B suspension is designed for a vertical wheel displacement of plus and minus three inches, but it still functions smoothly at ± four. Anything beyond ± four is irrelevant, as the nominal ground clearance is four inches.

The small angular displacement afforded by long wishbones makes it possible to use aircraft grade spherical rodends for the upper and lower ball joints and (most significant in the front suspension) install them with their axes vertical. This permits a much smaller turning radius (important on the street if not on the track) and provides an easy and reliable method for adjusting the camber.

Perhaps the most important advantage of the long wishbone concept in a car like the L6B is its tolerance toward changes in ride height. Often, an attempt to lower a production car to go racing will alter the suspension geometry so much that the benefits of the lowered chassis are completely obliterated by the unintentional but inevitable changes to wheel camber gain and roll centers. Long wishbones allow the builder of an L6B to set the nominal ground clearance anywhere between three and five inches (by simply adjusting the spring perches on the shock absorbers) without any significant effect on the suspension geometry.

The only significant disadvantage of long wishbones is their relative lack of stiffness (and stiffness is crucial in achieving predictable performance). This potential problem can be overcome by using a slightly larger tube when fabricating the wishbones, adding jury struts to the wishbones, and by designing very tall uprights. All three have been utilized on the L6B. Obviously, larger tubes and jury struts will increase the weight, but only by a small amount (and only half of the wishbone contributes to the unsprung weight). The tall upright is used to maximize the distance between the ball joints. This can significantly reduce the forces in the wishbones and thereby minimize any deflection.

One performance related parameter that is often badly managed on production cars (and on more than a few racing cars) is bump steer (the tendency of a wheel to change the direction in which it is pointed as it moves up or down). For any given suspension geometry, there will be one particular combination of steering rack width (measured between the inner tie rod ends) and steering rack height that will produce the proper (near zero) bump steer. With short wishbones (and correspondingly short tie rods) even the slightest deviation from the ideal steering rack width and height can result in a very pronounced increase in bump steer. Long wishbones are much more forgiving.

Specifics

So, now that we’ve established some design criteria and preferences, let’s see how the L6B measures up. First, (surprise!!) it does have long wishbones. The effective length of the lower front wishbones on the L6B is 21 inches (more than double that of many similar cars). This can only be made possible by a drastic reduction in the width of the frame between the wishbones. Fortunately, with a mid engine configuration, this is not particularly difficult and on the L6B, this section of the frame is only 12 ½ inches wide. This does, however, require the use of a very narrow steering rack. Most racks used in production cars would be much too wide and the resulting bump steer would be completely unacceptable. To solve this problem, we selected a very narrow and very rugged race car steering rack from Sweet Manufacturing. The steering rack width then became the starting point for the whole suspension design. This makes it possible to get the bump steer virtually perfect just by assembling the parts the way they were designed. Once assembled, the bump steer can be measured and small adjustments to suit a driver’s preferences can be made by adding or removing shim washers on the steering rack mounting bolts.

The rear suspension presents a bit more of a problem when we try to use long wishbones because there’s an engine in the way (and a V8 with large radius exhaust headers can take up a lot of room!). This problem is exacerbated by the incredibly fat rear tires which eat up yet another significant portion of the width we would like to devote to wishbones. In spite of this, however, by using a few cunning tricks, we’ve been able to fit wishbones that are almost 17 inches in effective length – not as long as on the front, but still much better than most.

To point the rear wheels in the right direction and to facilitate bump steer control, the rear upright incorporates a steering arm which is connected to the frame via tie rods. This arrangement is virtually identical the steering linkage at the front, except that there is no steering rack. The inner tie rod ends are bolted to slotted brackets on the frame so they may be adjusted up or down to fine tune the bump steer.

Camber Control

So far, we’ve looked at some general concepts and implementation, but we haven’t discussed what, specifically, we want the suspension geometry to do. For many years, the primary goal of a chassis designer was to keep the wheels perpendicular to the road. With an independent suspension system, this requires that we design the suspension system so it will automatically adjust the camber to compensate for chassis roll. If this were all we had to do, it would be a relatively easy task and we could achieve near perfect results. Unfortunately, however, the chassis doesn’t just roll, it can also pitch (dive or squat). A system that has been designed to compensate perfectly for roll changes will almost certainly respond badly to changes in pitch. The standard solution to this dilemma is a compromise between the two requirements. Choosing the right balance between roll compensation and pitch compensation must be based on several factors, including the physical characteristics of the car and the way in which it will be driven. For racing and other high performance driving, the performance of the front suspension system is most important during a turn, when the chassis tends to roll, so the design should favour good roll compensation. A bit of imperfect camber is not a serious problem in a straight line deceleration, so we might be tempted to ignore pitch compensation altogether. However, there will be times when, like it or not, we might have to brake in the middle of a corner, so we still need to accommodate some pitch change along with the roll change. In fact, in high performance driving, "trail braking" is quite normal - continue braking into a corner to keep the car balanced, smoothly trailing off, and then smoothly accelerating later in the corner up to full throttle before the exit, all resulting in pitch change through a large percentage of the corner.

To reach a reasonable compromise for the L6B front suspension we have assumed that, when the wheel moves up or down, 75 percent of that motion is due to chassis roll (so that, as the left wheel moves up, the right wheel moves down) and 25 percent of the movement is due to chassis pitch (when the left wheel moves up, the right wheel also moves up) and this was the basis for the design. While this may appear somewhat arbitrary and unscientific (probably because it is) it will allow us to achieve excellent performance over a very wide range of driving conditions. Fortunately, several other factors lessen the need for perfect camber control. The L6B, with its wide track and low center of gravity, even without anti-roll bars (or sway bars), will develop only about 3 degrees of chassis roll in a 1 g turn. The installation of anti-roll bars can cut this in half. This results in a vertical wheel displacement (due to a 1 g turn) of less than one inch, so any small imperfection in the camber compensation is not likely to be noticed. Similarly, any braking-induced dive will be much less severe than it would be in a large front engined car, so again, less than perfect camber compensation can be tolerated. Tire designers have also made a large contribution. Even though tires have become wider and wider over the years (and this should make camber control more important) those same tires have become much more tolerant to imperfect camber.

Roll Center Control

We started this discussion with a comment that, for many years, camber control was the primary goal when designing a suspension system. However, as the understanding of the relationship between suspension geometry and car performance grew, it became increasingly more evident that, for cars that will be driven at or near their performance limits, proper placement and control of the roll centers is paramount.

The roll center can be loosely defined as the center of rotation of the chassis during roll and its position is important because it directly influences the amount of chassis roll that will develop during cornering. The position of the roll center is determined entirely by the geometric design and setup of the suspension system. Depending on the design, the roll center could be a foot above the road, or a few inches above the road or even a few inches below the road. Of utmost importance during high speed cornering is the position of the roll center relative to the center of gravity. The roll moment (or torque) that causes the chassis to roll is directly proportional to the vertical distance between the roll center and the center of gravity.

Therefore, it seems obvious that, if we design our suspension system to put the roll center in the same location as the center of gravity, the roll moment will be reduced to zero and we will have no chassis roll. If only it were that simple! Unfortunately, the location of the roll center will tend to move as the suspension components move. Generally, if the nominal roll center position (i.e. with the wishbones in their normal position) is somewhere close to the road, the movement of the roll center can be made relatively small but, if we design for a nominal roll center close to the center of gravity (14 inches or so above the road), a small wheel displacement can cause a large movement in the roll center position. This can result in rapid and unexpected changes in the roll moment, making the car very difficult to control. From a driver’s perspective, a small variable roll moment can be much worse than a larger, but constant, roll moment.

Combining the Design Goals

The ultimate design goal then, is to produce a suspension system that minimizes the movement of the roll centers. But while doing so, we should keep the roll moment to some reasonable value so we don’t get excessive chassis roll, and we still have to maintain camber control. Also, experience has shown that its best to set the rear roll center a bit higher than the front.

With all these parameters to control, and keeping in mind that every time we change the geometry to improve one function everything else changes, its not difficult to see why we need the help of a computer to optimize the design.

Many years ago, when commercial suspension design software was way too expensive for anyone but General Motors (and perhaps Roger Penske), an analysis program, written in Basic, was put together to help in the design of the L5. The software was slow and cumbersome but it was much faster than cutting and welding and considerably more precise than cardboard models. Eventually, the program was converted into a spreadsheet format and playing what-if became so easy it was fun. Gradually, more inputs and outputs were added and the little program that grew became a real tool. It was, however an analysis tool, not a design tool. It was still necessary for the user to input a complete description of the suspension components and all their attachment locations, then the program would calculate the performance. Later, when Tom Pawlowski provided a convenient method of solving for the intersection of two circles, it became possible to turn the program into a design tool – one that can tell us where to locate the attachment points to achieve the required performance.

In its current state of development, this program allows us to move the wheel through its full range of vertical travel and see the effect on camber, the position of the roll center and the bump steer. A sample printout of some of the output data for the L6B front suspension is provided in the table below.

At first glance, this may seem like a meaningless list of numbers but, upon closer examination, it shows how well we’ve accomplished some of the design goals described in this article. Read on and we’ll walk you through the table………

The left hand column titled WHEEL DISPLACEMENT shows the vertical position of the front wheel as it is moved from its normal (at rest) position (0.00) to the normal limits of 3 inches up (3.00) and 3 inches down (-3.00) and to the extreme limits of ±4 inches. The numbers in the other columns, in the same row as the 0.00 wheel displacement, indicate the value of the other parameters when the suspension system is in its normal rest position.

The next column CAMBER CHANGE shows how the camber angle (in degrees) changes as the wheel is moved away from its rest position. This shows that as the wheel is moved up, the camber becomes more negative (the top of the wheel moves inward). As the wheel moves down, the camber becomes more positive (the top of the wheel moves outward) but by a smaller amount than the change due to upward movement. This difference in camber gain is achieved by making the upper wishbone shorter than the lower wishbone and is one of the main characteristics of the unequal length wishbone suspension.

Most car and tire combinations generally perform best with a bit of negative camber, even at rest (check any photograph showing the front view of a Formula 1 or Champ car). The column marked CAMBER shows that we have adjusted the suspension for a static (nominal) camber of –1 degree. The numbers in this column are just the arithmetic sum of this –1 degree bias and the CAMBER CHANGE column to the left. The values shown indicate the angle between the wheel axis and the (normally horizontal) plane of the chassis.

Because the road surface is used as the reference for all the measurements and calculations, when we calculate the position of the roll center, it is expressed as shown in the column titled ROLL CENTER ABOVE ROAD. However, we’re really interested in the roll center position relative to the center of gravity, so this distance is shown in the next column, ROLL CENTER BELOW C of G. This shows that, as discussed in the text, when we get everything else where we want it, the roll center is not very close to the center of gravity (in this case, its almost a foot below the C of G).

The next column RELATIVE ROLL MOMENT is the one that shows, quite dramatically, how well we’ve achieved our goal of keeping the roll moment constant. The numbers in this column show that as the wheel moves through its normal range of six inches, the roll moment stays within a range of 97 to 100 percent. Few drivers would complain about (or even notice) a change this small.

The SWING ARM LENGTH column is included for academic rather than practical purposes. It shows that if the suspension were (heaven forbid!) a pure swing arm configuration, this would be the length of the arm. As the numbers show that the length changes as the wheel moves up and down, this is obviously not a swing arm suspension.

One of the most useful features of this program is its ability to calculate bump steer characteristics as shown in the TOE-IN column at the extreme right side of the table. Typically, most designers will aim for zero bump steer (i.e. the wheels will remain pointing straight forward as they are moved up or down) and this would probably be good enough. However, with the precision available using computer aided design, we can do even better. The non-zero bump steer, as indicated by the -0.43 degrees toe-in (or +0.43 degrees toe-out) with a 3 inch upward wheel deflection, is intentional. This provides a slight amount of negative feedback to improve steering stability.

The benefits from negative feedback are easier to understand if we look at the detrimental effects of positive feedback (the condition we would create if we designed for toe-in as the wheel moves upward). Picture this….. You’re entering a hard left turn. You turn the steering wheel to the left and the turn begins to develop normally but, as the lateral acceleration increases, chassis roll begins to compress the right front suspension. As it does so, bump steer causes the wheel to turn further to the left. This, of course, increases the lateral acceleration…. and the chassis roll…. and the suspension compression… and the bump steer…. and the steering angle…. And the lateral acceleration…. and…. (got the picture?). By designing in a little bit of negative feed back, you may have to turn the steering wheel a fraction of a degree more, but the improvement in the predictability and smoothness of the response will make the car much easier to control.

Anti-Things

Anti-dive (a technique used to reduce the tendency for the nose of the car to head south during heavy breaking) can be designed into a suspension by inclining the mounting axis of the front wishbones (as viewed from the side). However, experience has shown that in a mid-engine car, there is very little need for anti-dive and the detrimental side-effects (such as reduced suspension compliance) far outweigh the benefits. Therefore, the L6B has zero anti-dive designed into the front suspension.

On the other hand, anti-squat (to reduce the tendency of the rear end of the car to sink during acceleration) has fewer drawbacks and, in a high powered mid-engine car, is considered almost essential. Although chassis squat during straight line acceleration may be not much more than a minor annoyance, excessive squat while accelerating out of a turn can become a real problem if the subsequent camber change results in a significant reduction in adhesion. We can reduce this problem by designing our camber control to accommodate some chassis squat, but this will reduce our ability to tolerate chassis roll, so we can’t design for chassis squat only. You may recall that the description of the front suspension mentioned that the design was based on a typical wheel deflection resulting 75% from chassis roll and 25 % from dive. For the rear suspension, the design weighting is 60% roll and 40% squat. This can compensate for some of the chassis squat on corner exiting, but not all of it.

Therefore, the L6B does incorporate anti-squat. This is implemented by use of a slotted attach bracket at the forward end of the upper wishbone. To avoid the problems associated with too much anti-squat, the range of adjustability is limited to 10% to 45%.

Springs and Shocks

The L6B uses a very conventional (conventional for race cars, anyway) coil-over spring/shock arrangement with the shock absorber mounted directly between the outboard end of the lower wishbone and the chassis. In accordance with our keep-it-simple principle, there are no pushrods and no bellcranks.