how to design a bell crank for push rod suspension

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Originally Posted by Jay Lawrence

Where did you get "effective" damper travel from? There is wheel motion, and this is translated to the damper. There isn't really anything "effective" here. Don't confuse yourself with motion ratio. It doesn't necessarily make whatever suspension mode stiffer; it can make it stiffer or softer or leave it the same, or it can be rising/failing rate.

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Searched for details of damper to get more in depth. Came across this article by "Kaz Technology".
http://www.kaztechnologies.com/fileadmin/user_upload/Kaz_Tech_Tips/FSAE_Damper_Guide-_Jim_Kasprzak_Kaz_Tech_Tip.pdf

While reading I was finding what effect damper travel will have on its characteristics. When damper travels more, the fluid is pushed more and since the cross section area of valve is same, better damping is achieved. (This was taught to us in no further detail). But after going through the article, first of all its not just travel but its travel per sec which affects damping. In it they have explained how fast, medium and slow speed dampers differentiate in rebound and compression. Then I came across this:

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There are several reasons for using high motion ratios. The first is higher motion ratios require lower spring rates for the same wheel rates. Lower spring rates are also lighter, and result in less spring and shock friction as well as lower component loads. The other reason is greater damper travel and higher shock velocities. Since dampers perform better at higher velocities, and the wheel displacements are quite small on a FSAE car, higher motion ratios produce better shock performance.

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By higher motion ratio they meant as close to 1.0 i.e the spring displacement or damper displacement is same as wheel displacement.

So I concluded, if we can bring motion ratio very close to 1 by direct mounting the damper, the benefits of easy frame design, load paths etc can be achieved without losing on shock performance.

Ashir,

You have enlightened me to the sheer "genius" of modern engineering.

I used to think that the modern approach to engineering, as seen in a lot of FSAE, simply involves COPYING whatever the current fashion happens to be, and then justifying your "engineering decisions" by offering up some lame, half-baked, CALCULATIONS as support.

But, no, I was completely wrong. Absolutely NO calculations are required! Not even any sort of rational, or well-reasoned, qualitative analysis. Instead, you only have to regurgitate a few HALF-TRUTHS, and then back them up with a mountain of IRRELEVANCIES, and BLATANT BULLDUST. Sheer genius!!!
~o0o~

For example...

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Originally Posted by Ashir

I went in more depth of search results and visited this topic
http://www.fsae.com/forums/archive/i...hp/t-408.html?

... conclusion seems good for me.

... my objective is to design the suspension based on the damper we already have. This will save us lot of money (more than what it will be saved on using direct damper mounting by removing bell crank and pushrod costs).
So now I have to make my suspension stiff without changing damper, I will use bell crank to chance motion ratio.

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It seems that your decision (in that post) is to use "bell cranks" (= pushrod-and-rockers) because your previous car already has them, so it is easier for you to keep doing the same. Is this your thinking?

(Edit: I just read your above post more closely, and it now seems that you may move to direct-acting SDs, but only if they have MR close to 1?)

So far, in your decision making, you seem to be happy with the following arguments. (These are taken from the above-linked thread, and are my main reason for ranting here... :)).
~o0o~

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For the shock to work well, and also to allow lighter springs (both in rate as well as in mass), the MR needs to be somewhere near 1:1. Modern formula cars strive for 1:1 for the front (if not higher), and even higher for the rear. This allows the springs to be of lighter rate and mass, ...

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Pure BULLDUST!!! (And, frankly, pure STUPIDITY!) For a given wheel-rate and travel, more spring travel (ie. higher MR = damper/wheel motion) requires a lower spring-rate, but also a proportionally LONGER SPRING!

Thinking about it "big-picture-wise", the integral of wheel-force x wheel-travel = strain-energy to be stored in the spring. So, for springs of equal strain-energy/mass capability (ie. for the same quality of spring steel), a given wheel-force x wheel-travel requires the same mass of steel, regardless of MR!

(Of course, fibreglass, or rubber-band, or gas, (or other...), springs ARE lighter for the same total strain energy storage, again regardless of MR...)
~o0o~

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[the magical MR=1] ... also pumps more fluid for every increment of wheel movement.

The more fluid that the shock can pump - especially with the high wheel rates we have on modern cars - the easier it is for the shock to be fine tuned for good spring and tire contact patch control.

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The "high wheel rates" come when incompetent suspension engineers eventually discover that "any suspension will work, if you don't let it...". However, these incompetents believe that they got the suspension working by "fine tuning" it.

As for "pumping more fluid", just how much fluid needs to be pumped (ie. how much energy needs to be dissipated) to control these teeny-weeny racecars, as they race around their billiard-table smooth tracks? (Please do the calculations, and see that it = SFA.)

Also, if you double the MR, then you double the "static friction" forces felt at the wheelprint due to stiction in the damper seals, etc. Generally, this is a bad thing.
~o0o~

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... the higher the MR the less force that is put into the frame where the Coil-over dampener attaches.

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This is one of those half-truths that obscures several more important facts. Namely, for given Fz wheel forces the stresses felt by the main portion of the chassis are, quite obviously, the same regardless of suspension details (as explained by The-Man). A pushrod at the same angle as a DASD, has, quite obviously, the same forces acting on it. The addition of a rocker with high MR simply transfers this same force to two, hopefully smaller, forces at two separate chassis nodes (ie. the rocker node, and the SD node).

"So, hey, why not 10 x pushrods, rockers, and dampers per corner!? Yeah, then the force into each of the, err... 10, no that's 20 chassis nodes, is only, err... 1/20 of what it was before... Genius!!!" :)
~o0o~

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If you practice with 39 C ambient temperature, and 53 C track temperature (like we do here at our country), radiator flow is really critical, and a shock in the way is a big problem.

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Groooaaannnn... See below...
~o0o~

The above-linked thread had some discussion (by Big Bird ++) along the lines of,
"If you are in close competition with Stuttgart at the front of the field, then tiny details such as the exact MR of your rockers might make the difference between 1st and 2nd place.".

Sounds plausible... Except that Monash won the most recent FSAE event, Oz-2013, with direct-acting SDs. And they won by the proverbial country mile! And they have an aero car, so it seems that DASDs don't mess up aero flows too much. And Monash have a side-mounted radiator, and they test and race in Oz, "the sunburnt country", with the FSAE comp held in summer. So DASDs don't seem to kill radiator flows too much either...

But Monash are not a Northern Hemisphere team like Stuttgart. Does that make a difference? Should it? They were briefly at the top of the FSAE ladder (IIRC, sometime last year?). And they have been hovering around there for quite some time...

Anyway, on that linked thread from 2010 there was a post from Fil (who seems to have been on the Monash team?) who said that he did a study of DASDs and concluded that they "actually made sense". I recall Monash at Oz-2013 pointing to a bucket full of pushrods and rockers (= dead weight and $$$s) as justification for their move to direct-acting SDs.

So, does anyone know how long Monash have been running DASDs, and what their competition record with them is?

(Important point: Your decisions should NOT be swayed by what some other team does, except, perhaps, as confirmation of your calculations!)
~o0o~

Bottom line;
Know thyself!

Are you doing what you are doing because it really makes sense, or simply because it is easier to follow the flock?

Z

For those interested in Motion Ratios, I should add some more comments here, because the situation is even worse than suggested above.

(Note that in the following,
MR = spring-damper-change-of-length/wheelprint-vertical-displacement.
So, for given size bump, larger MR means more spring-damper movement. (Sometimes MR is defined as the inverse of this.))

"Expert" opinion nowadays has it that, in general, a higher MR is better (see all the lame reasons in above post). In fact, one expert FSAE Design Judge has advised students several times that by increasing MR the "unsprung mass may be DECREASED...". This advice applies when using the same damper. And as noted in above post, for given wheel-rate and wheel-travel, the spring's mass also stays the same, regardless of MR.

So, regardless of MR, the same mass of spring and damper must be accelerated whenever the wheel hits a bump and is itself accelerated. But, with an increased MR, the acceleration of the SD must also increase. Doesn't this effectively INCREASE the "unsprung mass", by adding to the wheel's mass, the SD's mass, err.... multiplied by the MR?

Well, dear students, it is actually even worse than that! Let's do the calcs.
~o0o~

Consider a Wheel-Assembly of mass Mwa (= tyre/wheel/upright/hub/axle/moving-bits-of-suspension...), which is the nominal "unsprung mass".

At the end of some push/pullrod&rocker linkage (which we consider massless, or roll into Mwa) is a Spring-Damper that has a part with mass Msd (= perhaps the piston/piston-rod/spring-seat/first-few-coils-of-spring...) that must move at a velocity relative to the wheelprint that is determined by the above definition of Motion Ratio.

Now, let's assume the Wheel-Assembly of mass Mwa hits a bump, and after a short time dt, it is moving upward with extra velocity dVwa.

By Newton II, the WAs "rate of change of quantity of motion", namely,
dP/dt = Mwa*dVwa/dt,
is proportional to, and caused by, an upward acting force Fwp (force from-ground-to-wheelprint).

So, in appropriate units,
Fwp = Mwa*dVwa/dt.

(Note: Newton's "quantity of motion" is nowadays called "momentum" (ie. P = m*V, or for "rate of change", dP/dt = d(m*V)/dt). And while NII is, strictly speaking, F ~ P-dot = dP/dt, in this case you can use F = m*A if you find it easier...)

In the same small time period dt, the moving part of the SD must also increase in velocity, but now by MR*dVwa (from the definition of MR). The force needed to cause this change of quantity of motion of the SD is,
Fsd = Msd*(MR*dVwa)/dt.
This is the force acting from the rocker, to the moving-part-of-SD, via the ball-joint connecting these two parts.

BUT!!! As the reaction to this force, namely -Fsd, is passed backward from the SD, through the rocker and pushrod linkage, to the WA, it is necessarily multiplied by MR (Law of Levers).

So, whenever the WA is "changing its quantity of motion" (ie. accelerating), it is also feeling the inertial reaction of the SDs "change in motion". The total force at the wheelprint required to cause BOTH these "changes in motion" (ie. accelerations) is thus,
Fwp(total) = (Mwa + MR*MR*Msd)*dVwa/dt.

Bottom line here, the effective "unsprung mass" is INCREASED by the mass of the moving-part-of-Spring-Damper, multiplied by the MOTION RATIO SQUARED!
~o0o~

Are the above theoretical calculations backed up by any empirical examples? I can think of quite a few in the automotive world, but here are two from further afield.

1. To survive, an animal must be able to accelerate its feet very rapidly (and not just when running, but also when kicking out at predators, etc.). The typical Motion Ratio of an animal's muscles to whatever they are driving/controlling is very low. For example, a "gluteus maximus" (= bum muscle) might only move 0.1 metres while it drives the animal's foot 1 metre backwards. So here MR = 0.1 (with the muscle = the SD, and the foot = the wheelprint).

If the moving part of the "glute" has mass of, say, 1 kg, then the above calcs with MR = 0.1 suggest that this only adds 0.01 kg (= 0.1*0.1*1 kg) to the effective "unsprung mass" of the foot. If the Motion Ratio was 10, as suggested by extrapolating some of the lame reasoning in the above post, then the foot would be effectively burdened with an extra 100 kg (= 10*10*1 kg) of "unsprung/glute mass"!

2. A "trebuchet" is a Medieval siege engine that throws stones at castles. It has a large wooden lever-arm pivotted several metres above ground. The short end of the lever-arm attaches to a large basket full of several tons of stones, and the long end of the arm (plus a sling) does the throwing. Gravitational force acting on the heavy basket powers the machine (a bit like a muscle, or SD). But, quite obviously, the basket itself can't accelerate any faster than "G" (~9.8 m/s.s).

Do the calcs and you should see that the only way to throw stones a long way (ie. with high speed) is to have a very low MR (= basket-motion/throwing-end-motion). Incidentally, the same principles apply to the similar, but much older, torsion-spring powered "ballista".

In suspension terms, the above examples suggest that if you want the wheel to rapidly move up and down so that it can follow a bumpy road, and if the stuff controlling this movement (eg. the SD, muscle, etc.) has significant mass, then a LOWER MOTION RATIO IS BETTER.
~o0o~

So, bottom line is that hundreds of millions of years ago Nature used trail-and-error to figure out that low MRs are good. A thousand years ago Medieval turnip farmers had the common sense (ie. inspiration from Nature) to do the same. But today some Engineers in the uppermost echelons of motorsport do not seem to have figured it out yet. Well, not the "experts"...

Maybe some of you students can help them? :)

Z

(PS. Some F1 teams have figured it out, and have their springs and dampers at quite low MRs. (Hint: torsion springs and rotary dampers attached directly to the rockers.)

For teams with direct acting dampers, how do you balance your cross weight? I've found that by adjusting a pushrod about a turn, I can change the cross weight by about 0.5-1%, whereas I've not found much change in the cross weight when adjusting a spring perch. Granted, I've been dead tired every time I've tried it with the perches and never bothered trying afterwards.

We've used pushrods and bellcranks for the longest time and I've never really given direct acting dampers much serious thought. I thought that it would be difficult to get the motion ratio I wanted and difficult to package an ARB other than a passenger car style torsion-and-bending tube. The ideas I came up with mostly involved smaller bellcranks to connect a conventional u-bar, which sort of defeated the purpose of eliminating the bellcranks.

One solution I've seen is what looks like a shuttle bar ARB on the rear of the new Monash car. Our team has had a shuttle bar in the past, but we ditched it because its motion ratio varied quite a bit through bump travel, although maybe they've been able to fix this since theirs is mounted over a much larger width than ours was.

While I'm in favour of simplifying the system by eliminating the pushrods and bellcranks if the same or similar performance can be achieved, I wouldn't want to do so at the cost of losing roll stiffness distribution adjustment with the ARB. Changing springs is quite a PITA compared to adjusting an ARB blade or swapping a torsion bar if you want to reset your cross weight as well. I think the weekend autocrosser would certainly much prefer it, even though that mindset went out the window a long time ago.



Regarding motion ratios: I've honestly never really thought too much about picking a motion ratio. If you have a target wheel travel, a coilover travel fixed from the manufacturer, and want a roughly constant motion ratio, then it's already defined for you. Anything other than that will require you to put a really thick bumpstop on the damper or a hardstop somewhere else in the system, unless you're cool with the possibility of your suspension bottoming out on something other than a bump stop. It sort of makes sense to use all the travel you have in the damper though.

The arguments I've always heard for "high" motion ratios have been the same; lighter springs, lessened effects of friction, hysteresis and play, and better "use" of the shock. I bought the lighter springs argument without actually thinking about it, but it seems that's not actually a valid argument. As Z mentioned, a higher MR amplifies any frictional force felt at the damper, although it does attenuate any play in the damper(not that there should be any).

On the con side, increasing the speed of the damper will increase the mass flow/dynamic pressure though the piston, which will decrease the minimum pressure in the piston for a fixed minimum Cp. This increases the potential for cavitation and requires a higher shock pressure, increasing friction in the damper. Then there's the increased amplification of the damper/bellcrank mass effect that Z mentioned.

A crazy idea I just had is that maybe, the usual recommendation for "high" motion ratios stem from digressive damping knee speeds. I've noticed that the knee speeds for most digressive damping curves I've seen have fallen in the range of 25-50 mm/sec, with higher damper speeds being labelled as controlling road input and lower damper speeds being labelled as controlling chassis motions. For a fixed knee speed, you don't want your motion ratio getting too low, as while you can still increase the damping rate to keep the wheel damping rate the same, the knee speed at the wheel increases. This results in your damping blowing off later under larger bumps, or maybe even not at all, and would probably make tuning the dampers give unexpected results.

This is pure speculation, as I don't exactly have experience with a wide range of damping curves. It may explain why some F1 teams use the low motion ratio rotary dampers, since they probably design their own damper anyways and could change the knee speed to be in line with their lower motion ratio.

We adjusted corner weight with spring perch position. It did take a lot of adjustment to make a big difference, but that was the price we paid.

We didn't run an ARB last year, but we had one designed. It used some aluminum elements in bending, and I wasn't all that fond of it. It wasn't completely manufactured when we did our first test, and the results were good enough without it that we simply developed the package that was done, rather than trying to change things up. To run the front ARB would have meant so much dive under brakes that it would not have been funny.

As for quick adjustments to roll stiffness distribution, we used spring rubbers (similar to NASCAR) that we cut and modified to fit our spring diameters/wire diameters. They were for much larger springs, but some time on a bandsaw took care of that. We had 3 different durometer spring rubbers, and could interchange them as we pleased. Some people had a lot of doubts about them staying in place, but we ran 2 competitions with the same rubbers in place and had no troubles.

-Matt

Dear Z,

I know it has been over 3 years since you put up this post. But, I really seem to find something fundamentally wrong with what you said. I am new to the world of FSAE and please do pardon me if I sound stupid.

Let me first state the things I agree upon.
dP/dt = Mwa*dVwa/dt
Fsd = Msd*(MR*dVwa)/dt.
BUT!!! As the reaction to this force, namely -Fsd, is passed backward from the SD, through the rocker and pushrod linkage, to the WA, it is necessarily multiplied by MR (Law of Levers).

Now, this is where I do not agree upon.
Bottom line here, the effective "unsprung mass" is INCREASED by the mass of the moving-part-of-Spring-Damper, multiplied by the MOTION RATIO SQUARED!

From what I understand, you are basically increasing the effective spring rate of the car as a whole ( i.e. ride rate, by changing your wheel rate) when you increase your MR.
This does not is any way increase the 'unsprung' mass of the car. Your argument about how the force needed to produce the same wheel centre deflection along the Z axis of the car increases with increasing MR cannot
be attributed to the 'effective' increase in mass . The equation to be considered here is not F=mx(double dot) but F= mx(double dot) + kx + bx(single dot), where b is the damping coefficient. So, increasing your k increases the F you have to apply for the same wheel displacement.

Moreover, I do understand how your suspension's response time would increase if you were to increase your unsprung mass ( f = squareroot(k/m), where f= undamped natural frequency of the suspension )
But, what is happening here is that you are effectively increasing your 'k' and thus your suspension's response time should decrease, and not increase.

Also a progressively increasing MR in my opinion would help as you would be able to take care of your pitch and dive issues without having to increase your spring stiffness considerably which would in turn enable you to get
your desired ARB Roll contribution.( As in the percent by which your ARB contributes to the roll rate).

The only fault that I can see with higher motion ratios is that they considerably increase xSD(single dot, velocity of SD)= xW(single dot, velocity of W)*MR and this causes heating up of the damper. This heating up of the damper fluid causes it to change
it's viscosity and other properties which leads to undesired hysteresis.

I know that I am no expert. So, please feel free to point out mistakes.

Adi

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Originally Posted by Adi_97

...please feel free to point out mistakes.

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Adi,

You have made several mistakes. Don't worry, the experts also make those same mistakes, over and over again.
~~~o0o~~~

MAIN MISTAKE.
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Now, this is where I do not agree upon.
[Z quote->]"Bottom line here, the effective "unsprung mass" is INCREASED by the mass of the moving-part-of-Spring-Damper, multiplied by the MOTION RATIO SQUARED!"
From what I understand, you are basically increasing the effective spring rate of the car ... when you increase your MR.
This does not in any way increase the 'unsprung' mass of the car.

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Effectively, it does.

If you keep exactly the same spring (ie. with the same spring-stiffness K), and you then increase that spring's MR (wrt the wheel), then, yes, you do increase the "effective spring rate of the car" (= "ride-rate", or "wheel-rate"). In fact, and as I am sure you know by now, the "effective wheel-rate" goes up by ... MOTION RATIO SQUARED!

And if you keep exactly the same damper (ie. with exactly the same settings), then the damper forces also go up, but by an amount that depends on the shape of the damper curves, multiplied by ... MOTION RATIO SQUARED.

And similarly, the inertial resistance of the Wheel-Assembly+Spring-Damper to acceleration goes up by ... MOTION RATIO SQUARED (multiplied by mass of moving-parts-of-SD, as explained in my previous post).

It is true that the "net unsprung mass" of the car remains the same, being simply the sum of the masses of Wheel-Assemblies and Spring-Dampers. But by fitting the SDs at the end of a linkage with increased MR, you do, MOST CERTAINLY, increase the force required to accelerate the WA+SD.
~o0o~

In the previous post I explained this behaviour in terms of classical Newtonian "changes of momentum". I guess many students struggle to understand this approach these days (because failed eduation system), so below is another way to look at it.

This time consider two simple examples, with some simple numbers thrown in, and forget all about "Newton's Laws".

Instead of tricky "changes in momentum", think about the WORK and KINETIC ENERGIES involved. (Note, of course, that these two concepts are derived from N's Laws.)
Define:
Work = Force * Distance (with * being the Scalar-Dot-Product of the F and D vectors),
Kinetic-Energy = 1/2 x Mass x Velocity-SQUARED,
with both having units of Joules.
~o0o~

EXAMPLE 1 - Consider a "one-piece" WA+SD, with combined mass M = 10 kg. That is, consider a Motion Ratio = 1.

This one-piece mass M is forced to move upwards at constant acceleration A = 1 m/s/s, perhaps because the wheel hits a bump with parabolic profile. (<- Think about why "parabolic" implies "constant acceleration".)

After 1 second of this constant upward acceleration A, the mass M has an upward velocity V = 1 m/s/s x 1 s = 1 m/s, and has travelled an upward distance D = 0.5 metres. (<- These V and D being calculated with the very simple 1-D Kinematics that used to be taught in high-schools.)

So the Kinetic-Energy of the mass M is now,
KE = 1/2 x 10 kg x 1 m/s x 1 m/s = 5 Joules.

If (to keep this example simple) we assume no dissipative forces such as friction, and if we believe in olden-day concepts such as "CONSERVATION OF ENERGY", then we must conclude that this newly acquired KE of mass M must be the consequence of some Work done on the mass. That is, an upward force F must have acted on the mass M over an upward distance D.

So, for the books to balance, Work-done-on-mass must equal the mass's final Kinetic-Energy.

So, Work-done-on-mass = F * D = 5 Joules (= KE).

We conclude that the upward force F that acted on the mass M over the whole of the distance D = 0.5 m, is given by,
F = W/D = 5/0.5 = 10 Newtons.

(Note that this force is in addition to any "static wheel-load", say that due to gravity.)
~o0o~

EXAMPLE 2 - Now let's split the one-piece WA+SD mass into two separate masses.

The Wheel-Assembly part has mass Mwa = 9 kg, and is forced to move upwards exactly as before, namely with upwards acceleration of Awa = 1 m/s/s.

The moving-part-of-Spring-Damper has mass Msd = 1 kg, and is at the end of a pushrod & rocker linkage that causes it to move (in any direction!) 10 times as far as the WA. That is, the SD has Motion Ratio = 10.

So the SD's acceleration is necessarily Asd = MR x Awa = 10 m/s/s. (<-Think about it.)

After 1 second of the WA's constant upward acceleration of Awa = 1 m/s/s, it has upward velocity Vwa = 1 m/s, has travelled distance Dwa = 0.5 m, and has acquired Kinetic-Energy,
KEwa = 1/2 x 9 x 1 x 1 = 4.5 Joules.
This is almost the same as the first example, just slightly less because slightly less mass.

After the same 1 second the SD has a velocity of Vsd = 10 m/s (BECAUSE ITS MR = 10), so it has acquired Kinetic-Energy,
KEsd = 1/2 x 1 x 10 x 10 = 50 Joules!

So the TOTAL Kinetic-Energy of WA+SD is now KEwd+sd = 54.5 Joules!

Note that although the SD has the much smaller mass of only 1 kg against the WA's mass of 9 kg, the SD nevertheless carries by far the majority of the total Kinetic-Energy, because it has the greater VELOCITY SQUARED (<- Edit - added emphasis, because important!).

Again from Conservation of Energy, we conclude that the Work done, by an upward force Fwa+sd, acting on the Wheel-Assembly+SpringDamper, over the distance Dwa = 0.5 m, must be,
Fwa+sd = Work/Dwa = 54.5 J / 0.5 m = 109 N.

So the inertial resistance to acceleration of the COMBINED WA+SD has increased by more than ten times over Example 1, even though the total masses of both systems are identical.

By similar reasoning, we find the force Fwa to accelerate ONLY the WA is,
Fwa = 4.5 J / 0.5 m = 9 N.
Similar to Example 1, just a bit less because bit less mass (9 kg vs 10 kg).

And the force Fsd@wp, ACTING AT THE WHEELPRINT (and thus only moving distance D = 0.5 m) required to accelerate ONLY the moving-parts-of-SD, is,
Fsd@wp = 50 J / 0.5 m = 100 N!

Yikes!!! That is a lot of force to accelerate a small mass. And it will certainly increase wheel loads on the uphill side of bumps, and keep the wheel hanging in the air over the downhill sides!

(Note that the force Fsd, acting DIRECTLY on the moving-parts-of-SD, and thus moving the much greater distance of MR x D = 5 m, is given by,
Fsd = 50 J / 5 m = 10 N, as expected.
The MR, or "leverage", of the pushrod & rocker linkage multiplies this force to give the much higher force at the wheelprint.)
~~~o0o~~~

OTHER MISTAKES.
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...progressively increasing MR in my opinion would help as you would be able to take care of your pitch and dive issues without having to increase your spring stiffness considerably which would in turn enable you to get your desired ARB Roll contribution.

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1. "Progressive MR" (aka "rising-rates") on corner/wheel-springs is, in general, VERY BAD for circuit racing cars.

2. I see nothing "desirable" in ARBs.

Z

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Originally Posted by Z

2. I see nothing "desirable" in ARBs.

Z

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Hi Z,

Loooooooong time lurker, first time interlocutor, to protect myself from Claude's etiquette wrath I'll tell you that I'm an old guy, not a current student, forgotten most of the higher math I learned in getting a degree in it, club racer, kart racer, life long student and power user of suspensions.

How undesirable do you find ARB's? Are they a packaging complication? Are they practically unnecessary and an abomination unto The Lord? Or something in between to the extent that they're not the same thing (been there, thought that).

I thought it had been proven that for a given roll stiffness, achieving it with Spring And ARB gives a lower single wheel rate than achieving it with Spring only. Correspondingly then the warp stiffness too is lower with the spring/ARB combination.

Here's an entertaining thread in which the subject of ARB's is taken up: https://honda-tech.com/forums/road-r...asure-2761537/ A consolidated treatment appears in this thread: https://honda-tech.com/forums/road-r...-know-2903722/

Can you expand on your position on the subject?

Thanks,

Scott

Z, Yes you Z i am speaking to you! :)

1. IMO the main reason why variable motion ratio are used is getting the most of aeromaps. You need a "soft" car at low speed and a "stiff" car at high speed to make a compromise between mechanical grip and exploitation of the aeromaps which often are very front ride ride height sensitive.
Other solutions such as bump rubber or variable spring rate can be used. Test made on 7 post rigs in wind tunnel and on track has shown (at least to me and many of our customers) that sometimes yes you lose a bit in mechanical grip but you gain much more in aerodynamic downforce and aerobalance consistency.
Do you have any experience with such car.

2. Simplified problem: you driver complains about oversteer. 2 solutions: soften the rear or stiffen the front. To decide which way to go one more question to the driver: is the car too lazy or too nervous? If the car is too lazy you will stiffen the front. if the car is too nervous you will soften the rear.
Lets' say your diver says he as oversteer and he finds the car too nervous. So yo decide to soften the rear but oh surprise there isn't any rear ARB. OK so you will do it with the soften rear springs but if you do so you change the car pitch and heave frequencies (which affect both mechanical grip and aerodynamic platform)
For me front and rear ARB gives you the possibility to modify the TLTTD without affecting ride and aero ride height control Your comments?

3 Unless I miss something.... in your last post Example 2 you give a Mass wheel assembly of 9 Kg and a Spring damper assembly of 1 Kg. These are good generic realistic number.... but a motion ratio of 10?

Claude

Claude,

I think just about anyone with experience in professional motorsports would agree with you that you can possibly exploit an advantage from properly designed rising rate motion ratios, front and rear antiroll bars, etc if you have you have a good understanding of your vehicle. The important things you mentioned like having a good aeromap, ability to quantify mechanical grip changes, access to a 7 post or virtual 7 post software, and you can definitely improve performance of the car.

However consider the context of a hypothetical FSAE team, one that-

-Has no idea where their wings are dynamically (no simulation tools or post-processing methods to calculate wing positions from sensors)
-Has no idea where their wings should be (no aeromap)
-They are vaguely aware that a concept called "mechanical grip" exists, but they don't know how to define it or calculate any metric associated with it.
-They have no access to a 7 post rig or virtual 7 post rig software, and wouldn't have the first clue what to do with it even if they did.
-They know that generally "softer ARB in rear = better rear grip, stiffer ARB in rear = worse rear grip" but can't explain why.

I'm curious to know your opinion on this-
Do you believe such a team gains any advantage from having pushrod/pullrod actuated bellcranks, front and rear ARBs, etc? Are those parts worth their weight to a team like that?

Do you think it is more productive for a team like that to dedicate 2-3 members spending the whole design stage debating pushrod vs pullrod vs monoshock vs 3rd heave spring style suspensions (with none of the tools or knowledge required to make the best choice and fully exploit it's advantages), repackaging the ARB, and "optimizing" the bellcranks. Or would they be better off leaving all that crap alone and work on filling in the gaps in their knowledge so they can actually make better choices in the future?

And lastly -
How would you score a team in design judging with the lack of understanding that I listed above, if they had non adjustable dampers, direct mounted from A-arm to chassis, and no ARB's?
How would you score a team with the same lack of understanding, but they have the most expensive 4 way adjustable dampers money can buy, front and rear titanium blade electronically adjustable ARBs, pushrod or pullrod bellcrank actuation, and their design report says that their ride and roll frequencies are somehow "optimized"?