21 replies to this topic

[seryt]

I'm trying to understand how load transfer to each tyre can be estimated for different (steady state) conditions. A few books I've looked at, including Staniforth's Comp Car Suspension, imply that the roll centre is a key parameter. A search of the archives of this Forum revealed some interesting discussion suggesting that the force based roll centre can be different from the geometric roll centre (the usual parameter used). Does this mean that using the geometric roll centres - as in a lot of "accepted" literature - will result in incorrect estimates of tyre loads, chassis translation and rotation even for steady state conditions? If this is true, then what is the significance of the geometric roll centre that is often talked about? Is it often a good approximation of the actual roll centre?

Thanks.

[Greg Locock]

Milliken in RCVD seems to walk you through it in section 18.4, and offers some simplified equations as well.

I see that in the notes he cunningly specifies the RCH by a force based method.

For 'normal' suspensions with small deflections FBRC and GBRC height is fairly similar, for the few suspensions I have checked. I generally use GBRC for linear range work, and don't really think about it at all for limit handling.

[Goran Malmberg]

Roll centre is what have kept my busy for about a year now. That’s the reason for my first tread. What we want to know is how the transferred weight hit the tire contact patch to the ground. What I am saying now is a result of my own experiment with physical models, and must be read with some care.

Rollcentre is commonley defined as where the forcelines intercept vertically under cgh at rest. Under load the Fl intercept may occur elsewhere from the centre of the car, or in theory Fl may become parallel resulting in no interception and therfore no Rc, then where is the actual point of roll?
However, if we only look at Fl of the loaded wheel, the car still roll;s about where this line intercept with the under cgh vertical centre line. The Fl of the other side wheel is to unloaded to come in to play. In order to calculate the tire load from weight transfer we must know the total weight transfer for the situation in question, might be say 90 % to the loaded side, and we must know the forceline angle of each wheel under the same load. Then we may calculate Wt for each wheel from Fl angle tan* F.

There are a lot things that enters the picture like weight distribution, rollaxis angle, Tw front rear and last but not least the spring rate where we must know how to calculate the proper wheelrate.
Goran Malmberg

[Ben]

The fact of the matter is that the force line intercept is not the centre of rotation of the sprung mass. There are a variety of reasons for this, but basically as Greg has already stated the force intercept and the centre of rotation are close enough in the linear range for small suspension movements for the concept to be useful at an early design stage.

When it comes down to it, the roll centre height is an arbitrary decision within a sensible range. There are good reasons for having a higher rear roll centre (more rear anti-roll %) but beyond that. I don't think looking in detail at kinematic roll centre migrations is worthwhile. So without a multibody simulation, a sensible initial design with some adjustability followed by a lot of testing is the only effective development path.

Ben

[Goran Malmberg]

Ben,
I have performed extensive testing that is impossible to verify in a few lines, and the reason is that I am interested in the phenomenon. While performing such tests, one "by accident" is getting the eyes on other effects that is valuable.

For tuning we may use a simplified model like using the "resolution line" principle of Mike Ortiz as a tool for the matter.

Goran Malmberg

[seryt]

Many thanks for the replies. Its interesting that Milliken pretty clearly defines the roll centre on the basis of forces yet this seems to be ignored a lot when people talk about roll centres. I remember a conversation with someone who once ran Formula Altantic cars and had the view that one of the best handling cars he'd come across had an awful lot of (geometric) roll centre migration for small chassis movements. Maybe the real roll centre was less affected?

Based on what Ben said, a full solution of the kinematic and equilibrium equations is needed to get the tyre loads? So the tyre parameters (e.g sensitivity to vertical load of slip angle for peak lateral force) become potentially quite important in this?

[Ben]

Quote

Originally posted by seryt
So the tyre parameters (e.g sensitivity to vertical load of slip angle for peak lateral force) become potentially quite important in this?

Exactly.

Ben

[Goran Malmberg]

Ben,
Calculating car balance is a complicated matter. Contact patch is getting loaded from weight transfer and is diffrentley distributed over each wheel by the suspension system. But it is also distributed unevenley over the one single wheel contact patch, by inflation and camber etc. the trick is to keep all loads distributed the same over all square inches of road contact area.

The angle of the forceline define the difference between geometric and sprung weight distribution.
And shock absobrber and spring is part of the sprung side where the shock is working over movement and spring by location, (if location is a good word to use). The forceline location is not very hard to difine, but whats more tricky is motion-ratio for the spring. Wheelrat is Mr^2*springrate. And there is a lot factors involved in findin Mr. Mr also affect the shock setting by changing movement of speed. There has been fare to little attention payed to Mr in my eyes.

My definition of Rc is the point where the roll is separated from jacking forces, for a simple explanation. A car with high Rc, like the guy with the Formula Altantic car, may experience this as very good handling as great road contact. And the car might have appear balanced, but such car become more sensitive to road obstructions and jacking.
I should put the advantage of such car on the personal drivers positive list. But I dont know how high the Rc was located then, but the static height compared to dynimic is not that big. The difference in dynamic calculation is that we calculate each wheel more independentley, where the unladen wheels get - jacking while the loaded get + J. The unloaded wheel pic up less force, but if the geometry tilt the Fl up - jacking become influencial anyway.

Goran Malmberg

[Goran Malmberg]

Sorry, I meant Seryt, not Ben!
Goran Malmberg

[gary76]

One of the better explanations I have come across is in an SAE paper titled "Roll centres and jacking forces in independent suspensions" 1999-01-0046

[Ben]

Quote

Originally posted by gary76
One of the better explanations I have come across is in an SAE paper titled "Roll centres and jacking forces in independent suspensions" 1999-01-0046

Is that the one by the guy from Randle Engineering? If so I've also read it and would agree that it's worth a read.

My annoyance at accepting the kinematic roll centre = motion centre argument for many years is te reason I discouraged people from reading Competition Car Suspension on the recommended reading thread.

I'll try and pinch the work copy of Blundell and Harty's book and post their comments on "roll centres" because I feel it's a much better point to start at.

Ben

[McGuire]

There is nothing wrong with the concept of geometric roll center...except perhaps that it is entirely misnamed.

For those who are not familiar with him, below are the views of Mark Ortiz, whose name some may recognize from his columns in RCE. Personally, I can't fall in behind all of it but it is well worth examination...


The Mark Ortiz Automotive
CHASSIS NEWSLETTER
PRESENTED FREE OF CHARGE
AS A SERVICE TO THE
MOTORSPORTS COMMUNITY

August 2004


WELCOME

Mark Ortiz Automotive is a chassis consulting service primarily serving oval track and road racers. This newsletter is a free service intended to benefit racers and enthusiasts by offering useful insights into chassis engineering and answers to questions. Readers may mail questions to: 155 Wankel Dr., Kannapolis, NC 28083-8200; submit questions by phone at 704-933-8876; or submit questions by e-mail to: markortiz@vnet.net. Readers are invited to subscribe to this newsletter by e-mail. Just e-mail me and request to be added to the list.


ROLL CENTER MIGRATION, SOME MORE

The discussion of roll center definition in the June newsletter prompts this question from a reader:

Would you be able to discuss the effects of front lateral roll center migration for an oval track car with a solid axle rear end (NASCAR style) – perhaps an example on a short track where there are low speeds and aerodynamic loads, and higher amounts of vehicle roll where the left side suspension could be travelling into rebound?


The questioner mentions that he is an engineer for a major car manufacturer, and expresses a desire to remain anonymous, which is my usual practice in any case. Knowing that the questioner is an engineer, I am going to assume here that the reader is conversant with the basics of roll center theory as usually understood, and will not start at square one.

As those who read the June newsletter will know, I do not believe that the intersection of the front view projected force lines can properly be considered the roll center, or moment center, or anything of the kind. There are situations where you don’t get big modeling errors if you use the force line intersection as a roll or moment center, and other cases where you get huge errors. This merely illustrates that an incorrect analysis method can coincidentally produce correct or nearly correct answers in certain cases, despite the incorrectness of the method itself.

I have also said that the roll center, properly assigned, should be considered a point in side view (of the car), and its lateral position should be considered undefined. It lies in the transverse plane containing the wheel center in all cases, or, in side view, it lies straight down or straight up from the wheel center. So we really need only one number to define its position, namely its height. This height is not the same as the height of the force line intersection. Rather, it is the mean height of the two force line intercepts on a line I call the resolution line.

The resolution line is a vertical line in the front view, positioned according to distribution of lateral force generated by the two tires. For example, if the right front tire is generating 75% of the front lateral force, the front suspension resolution line is 75% of the track width away from that tire.

Unfortunately, we do not know this distribution of lateral force exactly, in most cases. We have to estimate it. That means our modeling of the suspension’s behavior is only as good as this estimate. This is unfortunate, but ignoring the fact doesn’t make it go away. The behavior of the suspension really does depend on the distribution of lateral force. To predict the jacking forces each of the individual wheels generates, and thereby calculate an anti-roll or pro-roll moment, we must not only know the suspensions’ geometry, but also the forces at the contact patches. Any analysis method that takes this into account, even using an estimate for the lateral force distribution, is better than a method that ignores this factor altogether.

What we’re doing here is directly analogous to modeling longitudinal anti effects – anti-dive, anti-squat, anti-lift – in side view. It is widely recognized that when modeling longitudinal anti effects, we have to know the front/rear distribution of longitudinal force, or try to estimate it with reasonable accuracy. For example, for braking short of lockup, we use the calculated brake bias. If the front brakes make 70% of the rearward force, we construct our resolution line 70% of the wheelbase back from the front wheel center. We then look at where the front wheel side view force line intercepts this resolution line. We take the height of this intercept as a percentage of sprung mass center of gravity height, and that is our percent anti-dive. We can do the same for the rear wheel, and that’s our percent anti-lift. When these are both 100%, the car will not pitch at all in braking, regardless of wheel rates, nor will the whole car jack up or down.

We can likewise define a percent anti-roll for the right and left wheels in an independently suspended front or rear pair, and we may also average these to define a percent anti-roll for the wheel pair. The average height of the intercepts makes a good value to use for roll center height – much better than using the force line intersection height, though in some cases the two values may be similar. The average height of the intercepts, or roll center height, may also be described as the sprung mass c.g. height times the percent anti-roll for the wheel pair. Also, the height of each of the intercepts, as a percentage of sprung mass c.g. height, is that wheel’s percent anti-roll.

This definition of the roll center provides a number that can be accurately used for load transfer and roll angle calculations, for it is a valid measure of the suspension’s geometric anti-roll properties.

Assigning a roll center location is useful not only for modeling or analysis, but also for discussion. To use the method I’m advocating for discussion, it is useful to have default assumption for lateral force distribution. I think assuming that the outside wheel generates 75% of the force is appropriate, absent better information.

Note that suspensions only generate geometric anti-roll or pro-roll moments in response to car-horizontal (lateral or longitudinal) forces. Force line slopes, force line intersections, and force line intercepts of the resolution line do not affect any tendency to roll, or resist roll, in response to car-vertical forces. (Spring splits, or wheel rate splits, do affect this. So does an offset c.g., or static left percentage other than 50%.)

Of course, the questioner here is not asking about effects of changes to the location of the roll center as I prefer to define it. He is asking about effects of lateral migration of the roll center as most people conceive it, namely the intersection of the front view force lines.

And in fact we can say some things about the position of the force line intersection, and the conclusions we can draw from it, for particular classes of situations. We can’t necessarily say the car has more tendency to roll, other things held constant, if the intersection moves to the inside of the turn, nor that the car has less tendency to roll if the intersection moves in the direction of roll, even
in a banked turn. However, we can make some more complex and qualified statements, for particular sets of conditions or assumptions.

To begin with, there are certain situations where we don’t know much at all from the force line intersection. If the force line intersection is at the contact patch center for either of the wheels, we know that the opposite wheel’s force line is horizontal, and therefore the opposite wheel has no anti-roll or pro-roll. However, the force line for the wheel on top of the intersection could be at any angle,
and therefore this suspension could have any amount of anti-roll or pro-roll. In this situation we can’t say anything about the overall amount of anti-roll or pro-roll in the geometry from the location of the force line intersection, nor can we infer the location of the roll center as I define it, without additional information.

Parallel lines do not intersect. When the force lines are parallel, there is no force line intersection. In this situation, users of force line intersection as the roll center will either say the roll center is undefined, or that it has disappeared, or – arbitrarily – that it is on the vehicle centerline, at the average height of the two force line intercepts of the centerline, which will be at ground level. However, the parallel force lines could be at any angle, relative to car-horizontal. We don’t know what that angle is; all we know is that it’s the same for both of them. We know that one wheel has anti-roll and the other has pro-roll, but we don’t know how much. We know that the anti-roll and pro-roll forces are equal if the tires are making equal car-lateral force, which would equate to a roll center at ground level. But if the tire forces are unequal – and they usually will be – we cannot say how much overall anti-roll or pro-roll the system has, and we cannot define the roll center my way, without additional information.

There is a third unique class of situation – or, if you like, a special case of the parallel force line situation – the one where both force lines are horizontal. In other words, the force lines are not only parallel, but coincidental. In this case, we cannot say what the lateral location of the force line intersection is. We may say there is an infinitely large number of them. We do know, however, that all these points are at ground level, and we can say with certainty that the suspension has no anti-roll or pro-roll, regardless of the magnitude of car-lateral forces at the contact patches. We can also say that the resolution line intercepts of both force lines are at ground level, no matter where the resolution line lies. Therefore we can define a roll center height my way, at ground level, despite the fact that we cannot define a single force line intersection. In this case only, we can do this without knowing, estimating, or assuming lateral force distribution.

For all cases except the above three classes, we can calculate the slopes of the force lines from their intersection. Knowing this, and a known, estimated, or assumed lateral force distribution, we do know enough to assign a roll center my way, and to say something about the overall anti-roll or pro-roll characteristics.

We can also say which direction the force line intersection will move, for a known roll displacement, if we know one more characteristic of the suspension: how the force line slope changes with suspension movement.

William C. Mitchell, in his SAE paper no. 983085 entitled Asymmetric Roll Centers, introduces a definable parameter that is useful in discussing this. He calls it the incline ratio. I find this nomenclature to be suggestive of a different meaning, and I think Bill deserves recognition for coming up with the idea, so I call it the Mitchell index.

By either name, we calculate this number as follows: We look at the centerline intercept of the force line, and we note its rate of height change as the suspension moves in ride. We express the rise and
fall of the intercept as a proportion of the ride motion, and that’s the incline ratio, or Mitchell index. If the intercept moves up and down at the same rate as the sprung mass, we have a Mitchell index of 1. If it doesn’t move at all, we have a Mitchell index of zero. If it moves up when the sprung mass moves down, we have a negative Mitchell index. If it moves down when the sprung mass moves down, by a lesser amount, we have a Mitchell index between 1 and 0. If it moves down when the sprung mass moves down, by a greater amount, we have a Mitchell index greater than 1.

The case the questioner raised in the June newsletter, where a short-and-long-arm suspension has the lower arms shorter than the uppers, illustrates a Mitchell index substantially greater than 1. Likewise, a strut suspension has a Mitchell index greater than 1. A pure trailing arm suspension has a Mitchell index of 0. Most short-and-long-arm layouts have Mitchell indices fairly close to 1 or a bit greater. With unusually short upper arms, stock car front ends can have a Mitchell index a bit less than 1. To get a Mitchell index of zero with a short-and-long-arm suspension, we need either very long lower arms, or very short uppers. The lengths have to be in accordance with Olley’s Rule: the lengths of the control arms have to be inversely proportional to their height above ground level, usually as measured at the ball joints. For typical stock car lower arm and spindle (upright) dimensions, that means upper arms somewhere around six to seven inches long, rather than the lengths of 9 inches or more commonly seen. Not surprisingly, Mitchell indices less than zero are uncommon.

The Mitchell index can be different for the right and left wheels, and in oval-track stock cars it usually is, though not by much. It also varies some as the suspension moves, but it does not undergo large, sudden changes. We also end up defining it differently if we take the centerline as being where the frame builder marked it, or as being at the midpoint of the front track, or as being the edge view of the longitudinal c.g. plane. These nuances aside, if we consider roll to be angular motion about the ground intercept of whatever centerline we’ve defined, then we can say certain things about how the
force lines and their intersection will move in particular combinations of ride and roll, based on the Mitchell indices of the two individual wheel suspensions.

If the Mitchell index is 1, the force line slope doesn’t change in roll. If the Mitchell index is 1 for both right and left wheels, the force line intersection doesn’t move in roll.

To take the most common category of cases first, suppose that, at static condition, the force line intersection is above ground level and between the wheels. In this condition, if the Mitchell index is greater than 1, the force line intersection always moves laterally opposite to the direction of roll. The force line for the outside wheel (right wheel in a left turn) loses inclination, while the force line for the inside wheel gains inclination. In the case the questioner cites, the intersection would move to the right. If the Mitchell index is less than 1, the force line intersection moves toward the outside wheel instead. The outside wheel force line gains inclination, while the inside wheel force line loses inclination.

In general, the former case implies a decrease in overall geometric anti-roll, and the latter implies an increase in overall geometric anti-roll, even with no change in force line intersection height, because the outside wheel generates more lateral force. Correspondingly, the roll center, defined my way, drops in the former case and rises in the latter case, even with no change in force line intersection height.

Now let’s change things a little. Let’s suppose the force line intersection is between the wheels but below ground level. This is actually not an uncommon condition in stock cars, especially drop-snout cars on banked turns.

Now, if the Mitchell index is greater than 1, the force line intersection moves toward the outside wheel in roll! The outside wheel is still losing anti-roll, or should we say gaining pro-roll. The inside wheel is still gaining anti-roll, or losing pro-roll. So the change in roll resistance is still the same as when the force line intersection was above ground, but the lateral migration of the force line intersection is in the opposite direction – toward the outside wheel.

If the Mitchell index is less than 1, again the change in roll resistance is the same as with an above-ground intersection – it increases. And again, the lateral migration of the intersection is in the opposite direction – toward the inside wheel.

This illustrates that we cannot infer the change in roll resistance knowing only the direction of lateral migration of the force line intersection, even supposing that the intersection height isn’t changing.

The wildest migrations of the force line intersection occur when the force lines are close to horizontal, and close to parallel. Small changes in force line angle will make the intersection move

all over the place. Small ride motions can make it move from above ground and way out to the right to below ground way off to the left. Does this mean the geometric anti-roll or pro-roll moment is varying all over the place, or that the car’s properties in a banked turn are varying all over the place? Not at all, because the force line slopes and individual wheel anti-roll and pro-roll are not changing much. And, correspondingly, the height of the roll center as I define it doesn’t change much.

All of this holds true regardless of whether the turn is banked, and regardless of what kind of suspension is at the other end of the car.

[Greg Locock]

Excellent stuff. There's more in that than most SAE papers, I certainly haven't gone through it thoroughly yet.

[Goran Malmberg]

Mark Ortiz is good at these stuff, thats why I talked about his "resolution line" method. THis particular newsletter is ayear old and I recieved this from Mark at the time. I also have his Video tape from a class speach "Minding your anti", together with a few private explanation letters from him. I had this in mind while doing my own experiments and testings. Those SEA papers is a little strange sometimes as they dont come up with very good explanations, espesially when making full scale car tests.
Goran Malmberg

[Supercar]

Quote

Originally posted by Ben When it comes down to it, the roll centre height is an arbitrary decision within a sensible range. There are good reasons for having a higher rear roll centre (more rear anti-roll %) but beyond that. I don't think looking in detail at kinematic roll centre migrations is worthwhile. So without a multibody simulation, a sensible initial design with some adjustability followed by a lot of testing is the only effective development path.[/B]

So, without an ADAMS model and lap time simulation, how would you objectively through testing pick the correct location for roll centers? I am saying *objectively* because I would like to rely on data, which is objective, rather than driver feedback, which is subjective.

Let's say I am now in the process of relocating my roll centers. Currently the front one is below the ground (I think) and the rear one is only a few mm above. I want to raise them both to speed up the steering response. I probably can set the front one 2" above the ground and retain a good front suspension geometry. Should I set the rear one at 3, 4, or 5"? The rear is totally adjustable and I can hang any data acquisition on the car that is needed.

What are the tests that I should run and what should I look for? Somehow I suspect that outside of the circle track world not that many engineers tune cars with roll centers, as opposed to tuning with springs, bars and dampers. I may be wrong though.

Philip

[Greg Locock]

Down here we tune roll centres a lot. The taxi cab racers adjust RCH during the race, I'm a bit dubious of that, and can't believe it has enough effect to matter, within the likely range of adjustment. Still, that's racing.

I look at initial roll control in transients and the frequency response in pulse type events. Generally this drives us towards higher rather than lower RCH, within the other constraints of our suspension.

I've seen measured numbers for several cars with respectable handling whose FBRCs are higher at the front than the rear.

Hmmm. The only guidelines I know of are (1) make sure that if it is below ground level, keep it below ground level, and vice versa and (2) the rear should be at least 50mm higher than the front. I don't really know how much faith I have in those - it seems odd that the recommendations haven't changed when we switched from cross plies to radials, for example.

[McGuire]

On the ovals the NASCAR taxi chassis is fairly sensitive to RC adjustment. The rear end is a live axle with trailing arms and coil springs of course, located in lateral plane by a Panhard bar with vertical adjustment via jack screw at the chassis mount. (With this setup the RC is said to be where the bar on the level bisects the axle's lateral CL.)

The adjustment is typically used to mess with the car in mid-corner out. Raising the bar will tend to loosen the rear (induce oversteer/reduce understeer) while lowering the bar will tighten up the car (reduce oversteer/induce understeer). A good driver can feel one turn on the jack screw. If you ask these guys what they are doing they may say they are moving the rear RC closer/further to the CG.

These cars run counterclockwise most all the time so the chassis mount is on the right. On clockwise (road) courses the mounts will be reversed. (You can also see this photo is from a few years ago when rear ARB was also in vogue. ARB is seen just behind the Panhard bar. )

[phantom II]

Listen here you stubborn old fool, if you lock the steering in the straight ahead position and place a mechanic dood on the trunk lid armed with a jack screw wrench on this here taxi and you drive it forward.... when the dood turns the jack screw to the right the car will go left and visa versa. The CG movement is insignificant. It is the thrust line change that does it. On the road car, the Panhard Rod or track bar is in the horizontal position where rhe radius has less effect. A Watt's link will cure the side to side movement of the axle and if you use a 3 trailing link set up, you can induce roll understeer. That's how I make 'em anyhoo.

Quote

Originally posted by McGuire
On the ovals the NASCAR taxi chassis is fairly sensitive to RC adjustment. The rear end is a live axle with trailing arms and coil springs of course, located in lateral plane by a Panhard bar with vertical adjustment via jack screw at the chassis mount. (With this setup the RC is said to be where the bar on the level bisects the axle's lateral CL.)

The adjustment is typically used to mess with the car in mid-corner out. Raising the bar will tend to loosen the rear (induce oversteer/reduce understeer) while lowering the bar will tighten up the car (reduce oversteer/induce understeer). A good driver can feel one turn on the jack screw. If you ask these guys what they are doing they may say they are moving the rear RC closer/further to the CG.

These cars run counterclockwise most all the time so the chassis mount is on the right. On clockwise (road) courses the mounts will be reversed. (You can also see this photo is from a few years ago when rear ARB was also in vogue. ARB is seen just behind the Panhard bar. )

[Fat Boy]

Ummmmmmm, Phantom, are you for real? I honestly can't tell whether you're being serious or humorous.

[McGuire]

Quote

Originally posted by phantom II
Listen here you stubborn old fool, if you lock the steering in the straight ahead position and place a mechanic dood on the trunk lid armed with a jack screw wrench on this here taxi and you drive it forward.... when the dood turns the jack screw to the right the car will go left and visa versa. The CG movement is insignificant. It is the thrust line change that does it. On the road car, the Panhard Rod or track bar is in the horizontal position where rhe radius has less effect. A Watt's link will cure the side to side movement of the axle and if you use a 3 trailing link set up, you can induce roll understeer. That's how I make 'em anyhoo.

No you listen, you stupid jerk. Drain that cheap gin out of your ears and pay attention.

The adjuster does not move the bar right/left, it moves the link's chassis mount up/down. The lateral movement will be in bump/roll. And anyway, as long as four tires are parallel the vehicle will track on a straight path. It is when you turn the vehicle that interesting things begin to happen. I would never say that geometric RC and CG are not significant to each other. One could say that is what it is all about. Next I suppose I will have to illuminate you about the proper martini as the current recipe is causing brain damage.

[Supercar]

Quote

Originally posted by McGuire The adjustment is typically used to mess with the car in mid-corner out. Raising the bar will tend to loosen the rear (induce oversteer/reduce understeer) while lowering the bar will tighten up the car (reduce oversteer/induce understeer). A good driver can feel one turn on the jack screw. If you ask these guys what they are doing they may say they are moving the rear RC closer/further to the CG.[/B]

Could you explain why the rear bar is used to adjust the balance mid-corner out? Why not anywhere else in the corner? Like the turn-in, for example.

Philip

[McGuire]

Quote

Originally posted by Supercar
Could you explain why the rear bar is used to adjust the balance mid-corner out? Why not anywhere else in the corner? Like the turn-in, for example.
Philip

On corner out out they are also working on forward bite.

In NASCAR they have essentially three tools to work with in tuning the chassis during a race: rear wedge, Panhard link height and tire pressure. (They may also have some additional wheel rate to play with if they are running spring rubbers.) The general tendency is for the track to tighten up over the race as it rubbers up so they are chasing that. But there are other factors too, such as if the track temp rises/falls with direct sunlight... or commonly these days, starts in the afternoon and runs into nightfall. There you can easily see a 20+ degree swing in track temp. The current tires also have considerable falloff deliberately built into them and the car will change big over a long green flag run. Sorry, didn't mean to hijack this into a NASCAB thread.

On ovals forward bite is a major issue. If you want to see some wild, imaginative stuff, check out some of the Panhard, Watts link, J-bar, Jacob's Ladder and other setups used on other short-track oval cars. (NASCAR allows only the Panhard bar.)