More here... I think.... at least easier to search through.. about halfway down this page..
http://www.auto-ware.com/techref/lib_index.htm
TUNING TRANSIENT BEHAVIOR OF FRONT-WHEEL-DRIVE AUTOCROSS CAR
http://www.auto-ware.com/ortiz/Chass...--July2010.htm
YAW MOMENTS FROM DIFFS WITH FRONT-WHEEL DRIVE
http://www.auto-ware.com/ortiz/Chass...vember2008.htm
If anyone is getting any funny ideas about trying THIS... (four spot welds and a bit of grinding to pop the brake line mount, the a short weld off the outside, maybe 7 minutes work TOPS)
You can really only go up just more than about 5/8 of an inch before you run out of clearance for the axles..(thanks for looking out Reaper1) I plan on using the bobble strut mount and the front mount to squeeze myself just a wee bit more room...
But I went from this...
To this...
No problem! I'm interested to see how this comes out!
Couldn't you shim the rear of the k-member down that same 5/8" as an experiment to see what happens? Or am I being an idiot again? (don't answer that!)
I imagine you probably should shim the rack back up a touch to combat bump steer if you try this.
I went a bit more than 5/8... and yes you could shim the rear of the K frame and probably effect the rest of the suspension geometry less... (maybe) I figured moving the front would help raise the roll center up a tiny bit as well... So I did that... and since in my application I may end up even changing that angle even more.. I can still drop the back a bit if I decide I want to...
Moving either end is going to effect bumpsteer.... But how much I dunno... It may not really change it all that much at all given that you are moving the LCA right along with the rack... It will change... But with some experimenting there might be a place in there where bumpsteer with stock stuff could be nearly eliminated and you could have your pro lift... But I could be wrong.. honestly I personally had not given that any thought because It doesn't matter in my application... I have to build all that stuff anyway...
I think I mentioned it in this thread already, but Mark Ortiz has a pretty cool newsletter he sends out once a month. most of the stuff is circle track type stuff.. But pretty cool anyway...
This months..
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: markortizauto@windstream.net. Readers are invited to subscribe to this newsletter by e-mail. Just e-mail me and request to be added to the list.
ANTI-DIVE, AND THE LOTUS REACTIVE RIDE HEIGHT SYSTEM
Current F1 news includes reactive anti-dive front suspension being declared illegal. Why don't they simply employ enough (100%?) anti-dive geometry to do the job. In other words, what are the negative side effects of running high levels of anti-dive geometry in road racing?
Also, can you please verify for me if this is an accurate statement:
With anti-dive the axes of the upper and lower control arms are not parallel. When the suspension compresses, the upper ball-joint moves further negative on the X axis of the car than the lower ball-joint. This leans the spindle rearward relative to the chassis (adds caster). The torque of the brake caliper tries to rotate the spindle forward relative to the chassis (reduce caster) which results in the suspension extending. These countering forces are what makes anti-dive work: compressing suspension adding caster versus caliper torque trying to reduce caster.
Taking the last part first – what makes anti-dive as we know it work – what the questioner is describing is sometimes referred to as torque anti-dive. It is indeed related to the instantaneous rate of caster change with respect to suspension displacement. There is also thrust anti-dive, which relates to longitudinal, or x-axis displacement of the hub or wheel center.
This conceptual framework applies if we think of the forces as acting at the wheel center. I find it simpler to think of the forces as acting at the contact patch. In that case, all anti-dive relates to the instantaneous rate of longitudinal displacement with respect to suspension displacement, for the contact patch center, with the brake locked. This rule covers all cases, even including an inboard brake with drop gears in the upright.
The rule can be expressed by the following equation:
dFz/dFx = dx/dz (1)
where:
Fz is jacking or anti-dive force induced in the suspension
Fx is longitudinal force at the contact patch
x is longitudinal displacement at the contact patch
z is vertical displacement of the suspension, at the contact patch
dFz/dFx can also be called the jacking coefficient for braking. On a kinematics and compliance (K&C) rig, we can measure dx/dz by cycling the suspension freely with the brakes locked and the sprung mass constrained longitudinally and laterally as it is moved vertically, and measuring longitudinal displacements of the wheel support pad. We can measure dFz/dFx by holding the sprung structure at fixed ride height (or a series of fixed ride heights) with the brakes locked, applying rearward force at the contact patch, and measuring change in vertical load at the wheel support pad. Measured results will not follow the equation exactly. The differences between predicted and measured values will give us some indication of the compliances, clearances, and frictional effects in the system.
When we are designing the car, or analyzing an existing car from point measurements, for outboard brakes the car has positive anti-dive if the side view instant center (SVIC) is either behind the wheel and above ground or ahead of the wheel and below ground, or if the SVIC is undefined (side view projected control arms parallel) and the side view projected control arms slope upward toward the rear.
For inboard brakes, assuming no gears in the upright, the rule is the same, with one important change: the car has positive anti-dive if the side view instant center (SVIC) is either behind the wheel and above wheel center or ahead of the wheel and below wheel center, or if the SVIC is undefined (side view projected control arms parallel) and the side view projected control arms slope upward toward the rear. Stated another way, the linkage or control arm system can only have torque anti-dive when torque reacts through it rather than directly to the sprung structure via a jointed shaft.
So anti-dive geometry in the suspension linkage must either create caster change with heave, or make the wheel move forward with heave. In the case of an inboard brake, only the latter of these will produce anti-dive.
It is worth noting that absence of caster change with heave (equal displacement at all four wheels; vertical translation of sprung mass) does not mean the caster never changes. Controlling caster change in both heave and pitch is very much like controlling camber change in both heave and roll: we can’t have zero change in both modes. The best we can do is compromise so we don’t have a lot of change in either mode. One recommendation I often make is to have the side view virtual swing arm length (SVSA) about equal to the wheelbase. That gives similar amounts of caster change per inch of wheel travel in heave (caster increase) and in pitch (caster decrease).
What limits how much anti-dive we can run? Two effects, basically. First, we tend to get wheel hop with large amounts of anti effect, of any kind, mainly at the point of wheel slip. Second, anything that makes the wheel move forward with respect to the sprung mass when the suspension compresses makes the suspension less able to absorb bumps. When the wheel hits a bump, it is best if it can move rearward as well as upward. If it has to move forward to move upward, that makes the suspension less compliant, and increases ride harshness and wheel load variation.
Briefly, that’s how anti-dive geometry as we have known it works. What is this reactive ride height control thing that Lotus came up with, that has been recently in the news for getting banned by the FIA? How does it work? Would it offer advantages over conventional anti-dive? Would it be applicable outside of Formula 1?
The device uses a small hydraulic system, actuated by brake torque, to raise the front end a little bit, compensating for compression under braking. Rather than being attached to the upright in a completely rigid manner, the caliper is mounted to the upright on a bracket that can rotate with respect to the upright – somewhat like a brake floater on a beam axle. The caliper bears against a piston, or pushrod acting on a piston, in a small master cylinder attached to, or built in unit with, the upright. Under braking, the master cylinder sends fluid under pressure through a short hose to a slave cylinder built into the lower end of the suspension pushrod. With sufficient hydraulic pressure, the pushrod extends, raising the front ride height.
An anti-dive effect is thus achieved, without any wheelbase or caster change in heave. If such a system is tested on a K&C rig, Equation (1) will not apply. There can be a positive jacking coefficient without having a locked wheel contact patch moving forward when the suspension is compressed. In fact, if the suspension geometry provides zero anti-dive as conventionally analyzed, the system will show slight rearward motion at the contact patch as the hydraulics work, when rearward force is applied to the contact patch. The system will create a compliance. However, this particular compliance will be accompanied by an increase in wheel load if the sprung mass is not allowed to rise, or a ride height increase compared to behavior without the system if the sprung mass is not constrained vertically.
The FIA banned the system under the rule prohibiting movable aerodynamic devices. Some have suggested that they should have used another rule that prohibits suspension devices primarily intended to influence aerodynamics. Personally, I think either of those is a reach, and the latter rule is highly ambiguous. Claiming that a suspension system is a movable aerodynamic device in the same sense as a movable wing or a sucker fan is baldly absurd. It may be that on a very smooth track, with ground-effect-critical wings and bodywork, the suspension does affect aerodynamics as much as it affects anything. But in that case, anything at all in the suspension is a device that influences aerodynamics. Certainly a third spring in the front suspension is that. Those were introduced after the advent of wings, to limit pitch, and control how far front wings are from the ground. But they’re legal.
Apparently, interpretation of this rule depends on a judgement of what the primary intent of the device at issue is. I would still say it’s a big reach to say that control of aero properties is the
primary function of any kind of passive anti-dive strategy, when cars having no aerodynamic downforce devices at all use various strategies to limit dive, and so do motorcycles. F1 cars had anti-dive before they had wings. This is just a slightly different way of passively harnessing the forces in braking to reduce front suspension compression.
F1 legality aside, does this idea offer functional advantages? Can it do anything that ordinary anti-dive cannot?
Lotus is saying this is all much ado about nothing, because they tested the system and didn’t like the way it behaved, so they weren’t going to race it anyway. That may be so, but I don’t think the matter is as simple as that.
I would expect that, as with so many things, brake-reactive ride height modification can hurt or help the car, depending on the details of how it’s applied and its interaction with other design and setup elements.
Although I would defer to actual experimental results on this, I am inclined to suppose that anything that produces a large jacking coefficient will cause wheel hop or chatter at the limit of adhesion, and also exact some penalty in ride quality and wheel load variation when braking, even if the jacking coefficient is obtained without caster or wheelbase change in heave. I would also expect that this propensity would depend on the overall jacking coefficient of the front suspension system – meaning the total from the brake-reactive elements plus any conventional anti-dive or pro-dive.
Most F1 cars nowadays have little or no anti-dive or pro-dive. If we simply add a reactive anti-dive system to an existing F1 car, we would either end up with a fairly small effect (if the reactive system does not create a large jacking coefficient), or end up with wheel hop or chatter in limit braking.
But since everything in a chassis interacts with everything else, we don’t necessarily get a meaningful read on an idea’s true potential by simply bolting it onto an existing car and seeing if the driver likes it or if the lap times come down. What if we designed the car to have reactive anti-dive, and took advantage of the system’s properties by changing other things?
For example, suppose we designed the front suspension with a side view instant center near the vertical rear axle plane and well below ground. The wheel would then move rearward considerably as the suspension compresses, and the system would have considerable geometric pro-dive. Suppose we then combined this with enough reactive anti-dive to give us just a modest amount of net anti-dive. We would then have improved ride and reduced wheel load variation most of the time, with braking behavior about like other cars.
We would get this without the disadvantages of compliance struts as used in passenger cars to allow the wheel to move rearward on bumps. Compliance struts allow compliance caster change in
braking and also in cornering due to the rearward force the tire generates when it is steered and running at a slip angle. This in turn results in compliance camber change. Those effects are the reason compliance struts are not used on race cars. Even for a passenger car, where the benefits of compliance struts are deemed worth the penalties, the combination of pro-dive geometry with reactive anti-dive could allow really low impact harshness, and/or allow lower control arm compliance to be reduced, resulting in a handling improvement without ride penalty.
Reactive anti-dive is just as suitable for inboard brakes as for outboard ones. Ordinarily, the only way to get anti-dive with inboard brakes is to make the wheel move forward when the suspension compresses. With reactive anti-dive, we can have a wheel that moves rearward in compression, and any amount of anti-dive we want. In many cases, we might even be able to dispense with hydraulics. Hydraulics are pretty much inescapable if the brake has to steer with respect to the suspension, but if the brake does not steer and is part of the sprung mass, in many cases we will be able to get the effects we want with purely mechanical actuation.
Compared to conventional anti-dive, reactive anti-dive is a roundabout way of getting the job done, and it does inevitably involve using more parts. However, as with other roundabout ways of doing things (e.g. pushrod and rocker suspension), the extra bits afford a convenient way of introducing intentional nonlinearities into the system’s behavior. With either hydraulic or non-hydraulic actuation, we can use preload springs, limiting springs or stops, and linkages with rising or falling motion ratios to get all sorts of non-linear anti-dive. We can, for example, have anti-dive that is extremely aggressive for small amounts of brake torque, then drops away to near zero or even goes negative for the high ranges of brake force where we are likely to encounter chatter or wheel hop.
Therefore, it is my opinion that brake-reactive anti-dive has a future, perhaps most of all for street use. In outlawing it for F1, the FIA is passing up an opportunity to have racing justify its existence by improving the breed.
There is some very interesting things there. The biggest thing I took away from it in respect to our cars and suspensions is that by introducing pro-lift in the front as is done with Brian Slowe's tubular k-frame, that it may be helping reduce wheel hop in both braking AND acceleration, which is a VERY good thing!
another interesting one
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: markortizauto@windstream.net. Readers are invited to subscribe to this newsletter by e-mail. Just e-mail me and request to be added to the list.
CURRENT F1 CARS DO HAVE ANTI-DIVE
I was reading your April article [based on February 2012 newsletter] in Racecar Engineering on, "Questioning the rule makers." You stated, "Most current F1 cars have little or no anti-dive or pro-dive."
I had intended to write to you last year regarding something I observed in a late season F1 race. I noticed that from a roll hoop mounted camera it was evident that the car nose was rising under braking. This would imply greater than 100% anti-dive. I don't recall which constructor it was; however, it seemed it was more than one of them. How is this possible, and what would be the benefit?
I was wrong.
Until recently, my statement would have been accurate, but it appears F1 has suddenly rediscovered anti-dive, and there is a massive “trick of the week” effect going on, with designers going to the opposite extreme and using outrageous amounts of it. In fact, this may have a lot to do with the “prominent nose bridge” (or maybe brow ridge?) noses that have appeared. The front inboard pickup points of the upper control arm are being raised to create anti-dive, and simply having a higher or more convex top surface on the nose may be undesirable for reasons of sight lines, side view area, and lift.
And it may well be that some cars now have more than 100% anti-dive, so that the front actually rises under braking. Why might that be?
I do not think it is desirable to have the car pitch rearward under braking. However, if the rear lifts under braking, it may be aerodynamically preferable to have the front lift similarly, so that pitch remains minimal, even though the ride height increases. It is entirely possible to have 100% anti-lift at the rear as well. That would mean we could have zero pitch without the front lifting.
My consulting clients are mainly hobby racers. Very occasionally, I get calls from top-level NASCAR engineers. I get absolutely no F1 work. Consequently, I have no idea how well F1 engineers actually understand vehicle dynamics. With the money and prestige the series has, one would think they’d be the best, but I don’t have the visibility to confirm or dispute that.
I do know that there is lots of incorrect information published about anti-dive and related effects. I’ve seen one graphic recently that apparently has circulated quite a bit that purportedly shows geometry for 100% anti-dive. It shows the control arm pivot axes in side view – the axes defined by the control arm attachment points to the tub – meeting at the sprung mass c.g. in side view. This replicates illustrations in a number of old chassis books, but it is incorrect in at least two ways.
First, the side-view geometric properties depend on the actual side-view projected control arms. These are the lines where the control arm planes intercept the wheel plane, not the control arm pivot axes as seen in side view.
Second, we do not have 100% anti-dive when the side-view projected control arms intersect at the c.g. We have 100% anti-dive when the side view force line intercepts the side-view resolution line at sprung mass c.g. height. The side-view force line is the line from the contact patch center through the side-view instant center. The side-view resolution line is a vertical line located rearward from the front axle line by a percentage of the wheelbase equal to the percentage of ground-plane retardation force exerted by the front wheel pair when braking. This will generally be a greater percentage than the static front weight percentage, so the resolution line will generally be aft of the c.g.
If the front suspension meets this criterion, the front suspension will neither compress nor extend in braking, as the questioner correctly understands. If the front end is lifting in braking, that implies that the anti-dive is more than 100%: force line slope, and jacking coefficient, are greater than described above.
But it is important to note that sprung mass pitch, and front wing height, also depend on what the rear suspension does. 100% anti-dive only results in zero pitch if the rear suspension has 100% anti-lift. It is quite possible to provide that. There are many production cars that have more than 100% anti-lift. Almost any car with trailing arm or semi-trailing arm rear suspension jacks the rear suspension down in braking. This potentially results in wheel hop or chatter, but in practice, as long as either ABS or the brake bias keeps the rear wheels from approaching lockup, such cars brake just fine.
With SLA or double wishbone rear suspension, 100% anti-lift requires a lot of inclination of the side-view projected control arms – more than is needed for 100% anti-dive at the front. The need for extreme-looking geometry results from the fact that the ground plane force we have to work with is smaller at the rear than at the front.
So it is possible that a designer could deliberately make the front end lift in braking because the rear lifts, or it is possible that even an F1 designer might have read the wrong literature, and miscalculated.
Is there a penalty in the car’s ability to absorb bumps while braking when there is that much anti-dive? Yes. However, when the track is very smooth, that may not matter so much. And when the alternative is to use stiffer springing instead, some anti-dive may be deemed preferable to that.
My own default recommendation regarding anti-dive lies somewhere between the recent former practice of using little or none, and the current fashion of using a huge amount. I generally suggest that the side-view force line should have a slope at static of around four degrees, and no more than eight degrees in any condition. Depending on brake bias, wheelbase, and sprung mass c.g. height, this will generally result in somewhere between 25% and 60% anti-dive.
EFFECTIVE UPPER CONTROL ARM PLANE IN STRUT SUSPENSION
From all of the books I have read regarding finding the swing axle length on strut suspension, they say that you project a line from the top of the strut, square to the centerline of the strut until it intersects with the line of the lower control arm.
I understand this 'in the old days' when struts were in line with the lower ball joint; however I am confused with modern suspension where the strut is bolted to the side of the steering knuckle. Should the upper line project square to the strut? Or should it be square to the 'virtual' strut (line between strut top pivot and ball joint)?
To really be accurate in finding front view geometry, it is necessary to find the front-view projected control arm. This will not contain the top pivot of the strut, although it will come fairly close to doing so. The effective upper control arm plane is the plane perpendicular to the strut axis, containing the top pivot center of rotation. The front-view projected control arm is the line where that plane intersects the front axle vertical plane. Ordinarily, in front view this line will pass slightly above the top pivot center of rotation. However, with two-dimensional drafting, a line through the pivot center and perpendicular to the strut axis will do as an approximation.
Correspondingly, the line where the effective control arm plane intersects the wheel plane is the side-view projected control arm. In side view, this line will usually be further above the pivot center than the front-view projected control arm is in front view.
It is the strut tube axis that we use for this, and not the steering axis. The line containing the ball joint and the top pivot center is the steering axis. We use that for determining front-view steering axis inclination, front view steering offset or scrub radius, caster, trail, and pin lead or trail. We use the tube axis for determining camber change, caster change, anti-dive, and anti-roll properties.
The lower control arm plane is determined the same way as in SLA suspension. It is the plane containing the lower arm pivot axis and the ball joint center of rotation.
Using the upper and lower front-view and side-view projected control arms, thus defined, we are able to find the front-view and side-view instant centers.
THREE-LINK GEOMETRY TO COMPENSATE FOR DRIVESHAFT TORQUE
For a 3 link solid axle, how is the offset of the upper link determined in order to minimize the effects of drive shaft torque on tire loads?
The offset of the upper link relates to a number of other factors. Most commonly, we know the height of the lower links, the tire radius, the height of the upper link attachment to the axle, and how far we can offset the upper link laterally from vehicle center or spring center. We then have a means of adjusting the angle of the upper link, and possibly its height, and we need to determine what that adjustment needs to be.
There is an equation for this. We may define the variables as follows:
LyU is the lateral (y axis) offset of the upper link from vehicle center, or from spring center.
HL is the height of the lower links from ground level, at the axle vertical plane.
HU is the height of the upper link from ground level, at the vertical axle plane.
NRP is the ring and pinion ratio, or the overall rear end ratio in the case of a quick-change.
RT is the loaded radius of the tire.
ϴU is the angle of the upper link from horizontal.
Then:
Tan ϴU = (RT (HU – HL)) / (NRP * HL * LYU)
It will be apparent from the equation that we can get the required compensation many different ways, but the upper link angle needed increases as we spread the upper and lower links apart. It increases as we reduce the lateral offset of the top link. It increases as we go to taller rear end ratios.
We may want to use less upper link inclination than the equation calls for, so the suspension doesn’t jack so much diagonal percentage into the car under braking. The equation should be considered to define the upper limit for top link inclination, when the top link reacts braking torque from both rear wheels.
Rules permitting, it is possible to have full compensation for driveshaft torque under power, and also have symmetrical behavior in braking. This involves the use of a birdcage or brake floater on the left. Barring that, there is a necessity to compromise between the conflicting objectives of minimizing roll and diagonal percentage change under power and minimizing these in braking.
This is kind of old, but I think it helps explain the effects of changing the anti-lift geometry as well as a very rough estimation as to how much geometry change is needed to give a certain effect in a vehicle that resembles our own in the exact situations we are experiencing.
http://www.whiteline.com.au/articles...WL%20ALK_b.pdf
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