I'd have thought we've got 7 different threads on RCH already, and most of them terminate with some sort of conclusion that it is just horses for courses.

I worked on a car that had either an IRS or a watts link live axle. We did not change the front suspension, and I'm damned if I could pick the difference between the two on public roads unless they were rough. The difference in RCH would be 200mm or so.

Cheers

Greg Locock


New here? Try reading these, they might help FAQ731-376: Eng-Tips.com Forum Policies http://eng-tips.com/market.cfm?

Zero (ground level) is a practical minimum. Going below in some conditions isn't the end of the world.

Independent suspension + high roll center = jacking. Jacking + soft spring rates = cars that are prone to rolling over by tripping over their own swing axles.

Beam axle will tolerate high roll center but it still has bad side effects in one-wheel bump.

You can make any bad suspension design work by not letting it move. (high spring and damping rates)

Pick something reasonable and run with it.

OK ,
i understand upper level when car start lifting,
but where is real bottom level? can be any lower value always compensate with extra roll bar stiffness?
i think then suffers opposite wheels independency
or coming transitient problems???

Yes, a lower roll center will increase the roll couple (based on the vertical distance between roll center and center of gravity) so if controlling roll is a priority, it would want stiffer springing and/or anti-roll bar rates.

Consider what happens to the contact patch in one-wheel bump if the instant center is below ground level and what that might do to the steering.

Roll centers / instant centers near ground level aren't the end of the world. Look at Formula 1 cars with their upper and lower wishbones out in plain view. (The arms are almost horizontal). There's not much camber change with roll, and the roll center at nominal ride height is near ground level. There's also not much functional suspension travel at all - the spring and antiroll rates barely let it move. And they're pretty fast on a smooth track. Might not work out so well if you tried to run one of those through a WRC gravel section, or go off-road desert racing with it.

And that begs the question ... what type of racing and on what type of track. I can guess, based on your posting history, but I don't want to guess. The design criteria for Formula 1 racing aren't the design criteria for WRC racing and those aren't the design criteria for the Baja 1000 racing.

Just saying "roadracing" doesn't fully address the question, because I'll guarantee that a F1 car will have difficulty with the rough, frost-heaved surfaces of accessible-to-amateur-level racetracks in my area (Canada). My thing is motorcycle roadracing, and I set mine up to have plenty of compliance and suspension travel. Too stiff beats up the rider (me) and sends the bike chattering off into the weeds. When you have an application that needs compliance and suspension travel, that's when it all starts getting tough.

Yes Brian,
let's say the interest is on the saloon cars on the asphalt track around 2-2,5 Hz ride frequency up to 100 mm total wheeltravel

"..........lower value always compensate with extra roll bar stiffness?"
In other words, designing in a lower "roll center" (that is, lower suspension roll resistance due to suspension geometry) to compensate for a too-stiff roll bar?

Buggar,
in other words:
when car is lowered, RCH go down.............is possible corrected with more stiffer anti roll bar? again and again?

(always is easyer increase ARB stiffness before restore RCH)

excuse my english

Given the ease of changing antiroll bars, I would think it's the other way around: having to use a very heavy antiroll bar in response to having a roll center too low (i.e. below ground level).

MacPherson front suspensions are rather well known for having lots of instant-center movement with suspension travel. The "tuners" who install suspension kits that slam the car to the ground end up placing the roll centers below ground level (and, often, with the instant-center on the "wrong" side of the car so that camber change on initial bump travel is the "wrong" way). 'Course, this is the same crowd that also installs lowering springs so low that the car is riding on bump stops thus having spring rates approach infinity, and antiroll bars that can only be described as "excessive", and ultra-low-profile tires with no sidewall, thus ending up with the aforementioned "you can make any bad suspension geometry work if you don't let it move" ... and a car that hops and skips and chatters over the slightest imperfections.

Don't do that.

Is a too-low geo RC the unavoidable consequence of lowering when there are competition rules prohibiting things like the relocation of suspension pivot locations and tall ball joints? Is there some other overriding reason?

One thing you clearly don't want is a RC so low that the camber gain starts going in the wrong direction on the outboard wheel (this being consistent with a badly placed front view instant center as noted by Brian). Camber gain that only slows to a small value in the right direction probably isn't going to be optimal either.


Norm

Lowering something with MacPherson geometry without relocating pivot points and/or dropped spindles will lower the roll center by more than the amount that the car is lowered, and can cause camber gain with bump travel to go the wrong way. What you can do about that, depends on careful reading of the rulebook. The easiest fix may be to not lower it by enough to cause trouble.

If you can install adjustable upper strut mounts, you can get some static camber and you can play with strut inclination angle (hopefully both fore/aft and in/out). If you can get some clearance in the bolt holes where the strut attaches to the spindle, you can get some static camber. Sometimes there may be a different car (which may be platform-related) that has parts that will interchange but have different geometry that can be swapped - Audi TT versus VW Golf, for example. Sometimes you can play with the way the subframe attaches to the unibody (which has the effect of moving the control arm attachment points and the steering rack together). You can have "collision damage".

World Rally cars have MacPherson geometry that can be ... interesting.

Norm,
Yes rules
and
Relocation of suspension pivot locations often is not easy,
Usually used the upright extensions brings new problems with upright extra load and compliance or is not possible due to upright configuration (steering rack relocation not possible also)

There exist only one right way which is only new uprights which is really expensive,
that is this other reason

From that point of view seems stiffer ARB only one way to maintain car out of bumpstop

I'm trying to find this acceptable RCH bottom level

The camber gain seems like a reasonable boundary,
But really we want to have CoG height above when chassis to ground gap or wheel arch area still are sufficient?

With MacPherson, when the front-view angle between the chassis-end pivot, the ball joint, and the strut mount is at 90 degrees, that is the spot where the camber gain with further bump travel starts going the wrong way. With normal geometry (a few degrees of front-view inclination of the strut), this will already have the roll center below ground level. If the lower arm is horizontal (ball joint center same height above ground as the chassis-end control arm pivot) that will place toe roll center a little above ground level and still have a little bit of camber gain left in it with further suspension compression. In the VW world, the latter scenario is commonly regarded as being as low as you can go without starting to cause trouble.

I betcha there is not a whole lot of difference in ride height between these scenarios ... I'll guess 20-30mm ...

Yes Brian,
I know McPherson problems and kinematic,
and yes I know TT vs Golf differences this can little help but is insufficient for lowering over 50mm

WRC is different (money) level

TT Front roll center HF = 0, Rear HR = 150 mm
1999 Golf HF = +2 mm HR = 12.4 mm
2000 Golf HF = -2 mm HR = 11.5 mm (Probably the same as 1999 but higher roll stiffness f & r same distribution.

Hyundai (N=16) all have HF ~ 100 mm, HR = ~100 mm

Perhaps it would help the rest of us if we knew exactly which vehicle we are talking about here, in case someone knows something more about it.

At 50mm lowering ... How much reserve travel in compression do you need (to allow for braking, cornering, and bumps)? With most normal cars, 50mm lowering won't leave much left. You would need high spring rates ... except that high spring rates with shorter springs might lead to zero preload in jounce, or even the spring coming loose entirely.

The fix may be "don't lower it that much" ...

Brian,
is general question

no problem go 50mm down with shorten dampers
around 30mm bupstop gap is sufficient with high rate springs of course helper springs is need

excuse me Cibachrome,
i do not understand
although Your data are always interesting
what Your opinion? is camber gain realy limit?

You mentioned TT and Golf. I stated their measured roll centers at 2 pass load. My approach is always different. Start with: "What roll per g can you tolerate ?", then, "What suspension travel do you have ?". Then: "What ride rates do you need to accomplish track bumps ?", "What caster do you need to align tire MZ peak with tire FY peak for a pair of tires?", "What extra roll stiffness do you need to fulfill your roll per g target, given the roll per g from springs ?", then its front and rear bar sizes that are not heavy, easy to change, efficient and fit. What's left is roll centers to provide missing roll stiffness AND +- 5% tunable TLLTD with the bars in your hauler. Very simple actually. Consider moving some weights around to help you: battery, gas tank, driver seat, etc.

Perfect!
then TT vs Golf have different ride height (have different uprights)

for clarification:
"What suspension travel do you have ?"
i think direct related to "What ride rates do you need to accomplish track bumps ?"
"What roll per g can you tolerate ?"
less roll=less camber gaing requirement?
"What caster do you need to align tire MZ peak with tire FY peak for a pair of tires?"
is resultant (max acceptable) static camber due to camber gain (or camber loss?)

"What extra roll stiffness do you need to fulfill your roll per g target, given the roll per g from springs ?"
i understand

"then its front and rear bar sizes that are not heavy, easy to change, efficient and fit."
this probably i not understand, this mean for ARB stiffness no exist limit? (or maximum % stiffness contribution?)

"What's left is roll centers to provide missing roll stiffness AND +- 5% tunable TLLTD with the bars in your"
OK

seems that real limit is maximum acceptable static camber related to roll
yes?

Don't fall into the trap of attempting to eliminate all body roll in the interest of attempting to keep the tire upright at all costs. Unless you are using tires with excessively low aspect ratio, they will tolerate being leaned over more than a lot of people think. It is possible to establish the relationship between camber and lateral force ... if you are serious about this, find out what that relationship is.

Attempting to eliminate body roll will result in excessive spring and antiroll rates, which will make the vehicle less able to keep the tires in contact with the ground over bumps and dips. It is ALL a tradeoff. Once again ... Formula 1 cars can use extremely high spring rates that will not work in WRC. Different conditions.

Quote (sierra)

"What extra roll stiffness do you need to fulfill your roll per g target, given the roll per g from springs ?"
i understand

"then its front and rear bar sizes that are not heavy, easy to change, efficient and fit."
this probably i not understand, this mean for ARB stiffness no exist limit? (or maximum % stiffness contribution?)

What you need to do with those two thoughts is tie them together. Once you've chosen your spring rates (on whatever basis), how much more roll stiffness do you need? After that, then it becomes a question of how you should distribute this additional roll stiffness.

Unless you're looking to define some maximum/minimum value envelope of RC heights to work within, I'm not sure how productive chasing a single aspect of suspension design - and having to make substantial compensation for the associated downside(s) - is really going to be. Even chasing a lower sprung mass CG height for its value might not be as productive as many people think. Smaller changes with smaller compensations might work better overall.


Consider that the tires you're going to be running - which may be dictated or otherwise limited by the class rules - will define an upper limit on lateral g. So there may not be much point in holding roll down to something really low like 1°/g or so if the tires aren't good for much more than 1g.

Maybe use damping to improve things like subjectively feeling that the car has taken a set.

Don't forget that after you've gone through all of these seemingly endless calculations, you'll still have to drive the real car rather than your simulation of it to see if the results lived up to the predictions.


Norm

OK,
Thank you for all your valuable comments!!
now some supplementary questions

What difference between vertical force from RCH or ARB? (both works against roll, but exist difference for contact patch vertical load over road bump when different % contribution is used?)
What transitional behavior difference if car controled his roll with RCH or ARB different % contribution?(slow versus fast load transfer),
We can expect problems when RCH axis have high angle?

another ideas?

Radek

There is a factor which tends to make a car with excessively high RC "skatey".

A car is cornering and a transient lateral load is applied - lets say a sudden increase. This could be a sudden steering input or sliding tyres suddenly finding grip. Lets compare two cars - one with high RC and low roll stiffness and the other with low RC and high roll stiffness, both with the same roll in deg/g. Although both will settle at the same roll angle in steady state cornering, the car with high RC will roll more gradually - call it "roll inertia" if you like. On the high RC car the transient is laterally transferring vertical contact patch load instantly, whereas the low RC car transfers load in proportion to roll angle. This sudden load transfer compromises the tyre's ability to maintain grip in these circumstances and make the car slower and considerably more difficult to drive.

je suis charlie

It is informative to plot the build up in forces in the various load paths as the car enters a corner. Milliken has a useful but limited series of plots of this. these show that RCH is a much 'faster' way of transferring forces than ARB. I have done this and written a report about it. Sadly it is behind a firewall.

Cheers

Greg Locock


New here? Try reading these, they might help FAQ731-376: Eng-Tips.com Forum Policies http://eng-tips.com/market.cfm?

I can imagine as the car enters a corner, then high RCH build vertical forces faster "similar" like a hard damping" (for example extra lower front RCH will be give corner entry oversteer)

But when car already in steady state and hit road bump, this have similar effect?

Damping is a very slow path for forces, in cornering.

Cheers

Greg Locock


New here? Try reading these, they might help FAQ731-376: Eng-Tips.com Forum Policies http://eng-tips.com/market.cfm?

Thought experiment.

Bear in mind that with a suspension design having a high instant-center, the contact patch moves "out" as it moves "up".

If you are cornering then the tire will be at a certain slip angle in top view with a certain force being applied due to the lateral acceleration. What the relationship is between that slip angle and lateral force will depend on the tire characteristics, the inflation pressure, etc.

Now, combine the two statements above. If it goes over a bump, it forces that tire to a greater slip angle and it applies a greater normal force and it applies a greater force in the direction of the lateral acceleration. Going down the other side, the opposite happens.

Maybe that results in a "jolt" laterally, felt by the driver. (I hate the ride motions of Jeeps with solid axles because of this.) Maybe it results in the tire momentarily losing grip (on either side of the bump). Who knows.

Quote:

But when car already in steady state and hit road bump, this have similar effect?

Why wouldn't it? Forces due to kinematics always build more quickly than forces that develop via displacements involving elastic components. Relative magnitude would be a separate matter.


Norm

As long as this thread still hangs on, I suggest you make yourself up a mechanical model (I was raised on Erector Sets) to OBSERVE the effects of low and high roll centers. We even used on in a legal case to help explain to a judge and jury some odd things about roll induced 'whatever'. Allow the model to be manipulated laterally (glass top table is nice) and in induced roll from a rolling road. You will need a couple of sets of springs (soft and hard) to try to maintain the same steady state roll gradient. Have a stop watch on hand, too.

The reason you should have a course in Mechanics is because the transient behavior for the 2 'roll center' options is usually (and sometimes even deliberately overlooked). Yet it's a player in this discussion. Especially so with a mid-80's Corvette.

You see, the yaw velocity peak frequency of a vehicle (let's just says it's 1 Hz) is the result of it's dimensions, mass distribution and cornering compliances (axle side-slip gradients). It's also speed dependent. One of the reasons we test vehicles at multiple speeds ranges is to observe these key critical metrics. Now roll peak frequency is a product of it's roll stiffness and inertia tensor, blah blah, blah. (OK, its a matrix with cross terms). Speed is not a player in roll dynamics for just about all vehicles. So what happens to this matrix when you move the roll axis up or down ? You have NOT changed the sprung center of gravity. What HAVE you changed ?

If you are so unfortunate as to have your roll frequency line up with the inherent yaw rate frequency at some speed, then typical roll/yaw coupling terms will light up, as in front and rear roll-steer and roll-camber (i.e. steer by roll for the vehicle I mentioned. This means that for example, rear roll understeer intended to reduce the rear cornering compliance and improve transient response, gets it's sign flipped as speed increases and a whole lot of bad shitT starts to happen.

One school (The Mechanics) throws a LOT of damping at the vehicle to calm it down while the other school (The Nerds) sets the roll-steer to roll oversteer to prove beyond all reasonable doubt that Roll Oversteer is good for you and the only way to do business. I wish I could get Louis Black to do a comedy sketch on this subject. It would be amazing, funny as Hell and technically correct. Watch him on YouTube. You'll get the idea...

"And so, Your Honor, the Idiot who read a book on vehicle dynamics and drew up some really pretty colorful lines on some pretty white Vellum had only book learning passed on by previous horse and buggy designers ..."

More or less i understand
Thank you all
only Cibachrome........ why "gets it's sign flipped" ?
something i overlooked?

probably I understand at last
only just don't understand why Roll Oversteer can be good

restore steady state balance?

Look up the meaning of "convolution of cascaded transfer functions". Roll does not get it's input from the driver/operator in a flat road turn. Yes, you can be a weeee bit from large caster settings on steered wheels, but for now, get your head around this notion: The driver creates a yaw velocity and a sideslip velocity by steering some tires. Lateral acceleration results from the yaw rate and the sideslip acceleration. Roll happens when the sprung mass wakes up and gets the message. So in the mathematical sense, an Ay by Steer mechanism (transfer function) is input to a Roll by Ay mechanism (transfer function). Transfer functions for dynamic systems involve complex variables to simplify the math. This multiplication in the complex frequency domain is call 'convolution'. There "can be/is" coupling feedback by means of the suspension designer's intent to have 'Roll Steer'. (Don't call it 'bump steer', that's a term used by wrench heads. This is New School). The convolution process feeds back steer from the roll 'black box' to the chosen axle with a phase angle resulting from its net frequency and damping traits. Since yaw rate frequency increases with speed, the phase angle for roll steer can change sign if the peak frequencies are pretty close and yaw rate peak is below that of roll at low speed. As speed increases, the yaw rate peak will climb and overtake that of roll and keep going. This will eventually result in reduced damping, high overshoot, long settling times and even an instability at VERY high speed when the driver's own transfer function starts leaking soft brownish material. Even the most simple controls models of Vehicle Dynamics show you this if done correctly.

This overall mechanism is also a boundary constraint for your roll axis and sprung mass center of gravity locations. And the R.A. inclination angle is a means of altering the convolution process to favor the dynamic *transient response, especially if the vehicle is intended to carry large differential payloads (passengers, building materials, and old school vehicle dynamics text books.

I think i can imagine well
(I have very good imagination...........mainly for conditions close to leak soft brownish material )

If i understand, roll axis and sprung mass center of gravity locations are in priority relation,

Seems that roll axis inclination value is MOST important for transient events before only roll axis height which extra low height can be more or less corrected with ARB

We can set general scheme? (60/40........50/50.......40/60 % front weight distribution)
You can give some specific examples or even some recommendations?