Damping ratios

[Paterson]

Reading milliken it appears that desireable damping ratios for chassis are between .7 and 1 while unsprung ratios are greater than 1 for overdamping.

Upon analysis of the or current vechile, i can only get a damping ratio for unsprung mass that is much lower than that of the chassis.

From analysis of the car weight, is it fair to say this is likely to be due to the small mass ratio between unsprung and sprung mass and the low spring rates used when compared with a standard heavier race car.

Is this a standard problem within Formula SAE design constraints or can someone give me ideas for solving this problem.

[Z]

Can you post some of your numbers/calculations? Approximately what your masses are, spring rates, estimated tyre rates, etc. Only need your rough numbers to see if you're doing it right - you don't have to give any secrets away.

Remember, sprung mass bounces on springs and tyres in series, while "unsprung" mass bounces between spring and tyre in parallel (which, to me, makes the name "unsprung" pretty silly).

Damping ratio of 1 or more is much too stiff. For racing maybe 0.5-0.7. For comfort more like 0.25. With conventional dampers you will have different ratios for body bounce, pitch, and roll, and for wheel hop - you can't "optimise" them all at the same time.

But, in practice, for FSAE I don't see too many problems with damping.

Z

[Paterson]

unsprung weight: 15kg
sprung weight each front wheel: 43kg
spring rate: 1010lb/in approx.
tyre spring rate: 760 lb/in (Goodyear 13x6.5's)

[Denny Trimble]

What's your installation ratio (shock motion / wheel motion)? You'll need that to calculate your wheel rate (Kw = Ks * IR^2)

What are your damper forces at the speed you care about (perhaps 1 to 5 in/s)?

[Paterson]

motion ratio of 1:1 wheel to shock.

[Paterson]

we haven't decided on a final damping co-ef yet. we are doing prelim calcs for this purpose. The current shocks (Penske units) are hopefully going on a dyno next week. Last years team didn't have enough time to evaluate any damping stuff, they are just stock off the self configuration.

[Denny Trimble]

OK, do you have an extra zero in there somewhere?

1010 lb/in wheel rates are a little high for FSAE, unless you have some sort of downforce device I haven't heard of.

Does the fact that the tire is softer than your spring raise a red flag?

[Paterson]

just slightly.

110 lb/in springs sorry.

[jack]

my shot in the dark for damping ratio was .65. how does that sound? anybody want to explain how to find this value without just guessing?

here is some info:
motion ratio: 1:1
front spring rate: 130 lbs/in
rear spring rate: 160 lbs/in
chassis Hz: 3.2
critical damping constant (chassis): 39.657 lbs/in/sec
front wheel Hz: 24.43
rear wheel Hz: 23.5
front corner critical damping constant: 15.893
rear c.c.d.c.: 19.103
front damping coeff: 10.33
rear damping coeff: 12.417

does anyone want to check if i have a mistake in my spread sheet?? http://fsae.com/groupee_common/emoticons/icon_wink.gif

i also have a spot to plug in damper velocities, and it gives me the force on the damper, what is a good damper velocity in in/sec to use for this?

[fade]

It seems like the damper velocites would vary depending on the situation. you need some sort of disturbance profile to estimate the effect it will have on your wheel load ie a pothole velocity curve will be different than say the shaft speeds from the car rolling in a corner. I am aware that most dampers incorperate shims for blow off which serves to reduce the damping rate at higher speeds but where should this "nose" be? I would suspect that LVDTs on your current suspension should give you a good idea on the speeds.

[jack]

yes, of course the dampers are going to move with all sorts of velocities, i would like to know a "top speed" so i can calculate the maximum force that the dampers put into the chassis.

...yeah i guess i could figure a speed out, but that would be a little tedious.

[Z]

Paterson,

Here is how I figure it. I have converted your units to metric, and added a bit to body mass in case you didn't include the driver (or do you have a very light car?).

Car details:
=============
Wheel Assembly = 15kg
1/4 Body = 50kg
Kspring = 20 kN/m (~20 kg/cm)
Ktyre = 130 kN/m

So;
Kbody-bounce = (Ks x Kt)/(Ks + Kt) = 17 kN/m (springs in series)
Kwheel-hop = Ks + Kt = 150 kN/m (springs in parallel)

Resonant Frequencies (assuming no damping):
===========================================
Given by;

N = 1/2pi x sqrt(K/mass)

So;
Nbody-bounce = 0.16 x sqrt(17,000/50) = 2.9 Hz

This is only for symmetric body "bouncing" (or "heaving"). In practice bounce and pitch are coupled and somewhat more complex (I'll leave for another time...). Real "bounce mode" will be similar frequency to above while pitch and roll will be higher.

Nwheel-hop = 0.16 x sqrt(150,000/15) = 16 Hz

This should be reasonably accurate, but only while tyre on ground. Once tyre starts to leave ground (ie. severe wheel-hop) frequency and required damping go down, but grip disappears!

Critical Damping Coefficients for above:
========================================
Note that these are only "critical" in the mathematical sense of lowest damping for no overshoot. Given by;

Cc = 2 x sqrt(K x mass)

So;
Ccb-b = 2 x sqrt(17,000 x 50) = 1.8 kN/(m/s)
Assumes damper between body and ground! (See notes below)

Ccw-h = 2 x sqrt(150,000 x 15) = 3 kN/(m/s)

So what dampers do you need?
============================
If you use dampers with C = 2 kN/(m/s) then Damping Ratio for wheel-hop is 0.67. This is more than enough, possibly too much. Same dampers will give DR > 1 for control of the body in bounce, so VERY stiff control of body bounce, and even stiffer in pitch and roll. Too stiff dampers on a bumpy track feel as if you are driving through thick mud - lots of drag and not a good thing. Much too stiff dampers take the springs out of action, so the car ends up bouncing on its tyres, which is as bad, maybe even worse than, too soft dampers. TOO STIFF IS NOT GOOD!

Bottom line is that C = 2 kN/(m/s) is the stiffest you will need with the above set-up. This is about the softest range for normal racecar dampers.

So really there is no justification for using push/pullrods-&-rockers to get the motion ratio up so that you can get "good control from the dampers". The extra motion ratio just increases the effective shaft stiction forces which is also bad (plus all the other disadvantages of rockers that I mentioned elsewhere).

Z

[Z]

Jack,

Maximum damper force depends on the maximum bump size! Denny suggested damper velocities of 1 - 5 in/sec., which is about maximum range for body motions. Any velocities above this come from bumps, hence your damper stiffness should be reduced (ie. blow-off) above this range. Maximum bump velocity in FSAE?? Say, 20 in/sec would be a VERY severe pothole. Typically, the "knee" can start at 2 - 4 in/sec.

Your figures? Hmmmm. Compare them and your calcs to my above post...

Z

PS. (Edit) I've had another look at your figures, and converting units and using masses ~same as Paterson's, your damping coefficients come out about the same as mine.

Your guess at DR = 0.65 is ok, based on what seems to work in practice. IE., if you experiment with different dampers until the car "feels good", then do the calcs later, you find that DR is from 0.2 (for comfort, but a bit "floaty") to 0.7 (for very stable platform, but a bit "draggy").

[D J Yates]

Interesting topic and something i wish i'd had more time to investigate this year.

Obviously, a damping ratio is given by c/cc where c is the coeficient of damping and cc the critical damping for a mode of vibration. Now, perhaps in my youth and inexperience i'm making a mistake, but i could swear that non of the equations in this thread correctly express terms for the natural frequencies of the various suspension modes (In the side view alone there area 4 modes of vibration, sprung mass pitch and translation, front unsprung mass and rear unsprung mass!). If these are not correctly expressed, then neither will the critical damping and damping ratios etc. I can apreciate the need for simplifying the situation, but i have neither the ability nor patience to solve for more than two modes. Instead, i found the best way to select damper rates is as follows:

Equations of motion -> simulink model -> input damping rates and iterate until your satisfied.

This allows you to model non-linear dampers (i.e. different compression and rebound rates) and different road disturbances at different speeds. For example, here's a pic of the response in the side view to a single road disturbance at around 30mph. The dashed lines are the unsprung mass displacements and the dotted the lines the suspension travel.

http://img.photobucket.com/albums/v208/white.noise/response4.bmp

I found that the best damping ratio in pure translation was not the best for combined pitch and translation (the front had an adverse affect on the rear). Decreasing the damping ratio resulted in a faster net recovery of the front and rear (i.e. faster return to normal load on tyre before disturbance, and therefore lateral force etc). Something to think about.

If anyone would like a copy of the simulink model i used, just PM me. It's far from perfect so i don't mind sharing.

BTW, what mode of vibration is recomeded as 3Hz. Is it an actual mode, or a simplified mode (i.e sqrt[corner mass(kg)/wheel spring rate(n/m)/(2*pi)])? Might be worth a seperate thread altogether.

[Paterson]

thanks Z,

that is what my spreadsheet has been indicating. I an unsure as to the advantages of a overdamped chassis. I though a slightly underdamped chassis would be better to drive.

[Kevin Hayward]

Only just started reading this post. Lots of really good stuff.

Jack, the speeds we have seen in data acquisition seem to peak at about 100mm/s for "smooth" tracks. In this sort of setup the damper spends most of its time under 25mm/s. However on rougher conditions (ie Detroit) we see peaks of 150-175mm's with much more spread out velocities. We end up working on a damper shape based on the rougher conditions mainly because the "knee" location really does not matter as much for smooth surfaces. The upper end Z's suggestions of around 50-100 mm/s is around what we aim for.

One of the things that should be noted looking at these speeds is that they are quite low. These are for 1:1 motion ratios. This is one of the reasons mountain bike shocks suck. In order to get an unmodified MBD working well you need motion ratios less than 1 (wheel moving more than the shaft), around 0.6 works with the Fox shocks. This further reduces the shaft speed of the damper. Hysteresis ends up being a big problem for them and the lower the speeds the more noticeable the effect. Unfortunately it is the lower speeds that you want accurate control as they deal with the handling issues. Basically the MBD's do not shift enough fluid for control at these speeds. Just thought I'd mention that as the damping coefficients that are calculated etc do not end up being what the car sees under real conditions. Possibly one of the reasons to look into the whole digressive vs. progressive damper curves.

We also end up designing for very similar force-velocity curves for both rebound and compression. The valves we use allow us an adjustment range of damping ratios between around 0.4 and 1. We can change this range by changing some of the internals ... but it hasn't been necessary. Once again Z's suggestion of around 0.6 is pretty much how it pans out after track tuning.

David's pitch analysis is definitely on the right track. The dampers become incredible tuning tools for the transitional pitch effects such as throttle off oversteer. Also has to make you wonder whether there should be different settings for different events ... http://fsae.com/groupee_common/emoticons/icon_smile.gif

Cheers,

Kev

[Nick McNaughton]

Kev,

Seriously. Stop procrastinating and get back to working on your PhD.

A friendly reminder from a NY airport,
Nick & Jase

[Paterson]

Good information,

we have just been working out our roll gradients for the system and are geeting results in the order of .2 deg/g. Seems very low to me! Anyone got any advice as to what is the standard for the SAE competition, i would guess around 1-1.5 deg/g? Thoughts?

[fade]

that does seem very low (stiff). >1 deg/g would be more suitable for low speed. are you running very close to the ground? the parking lot is pretty bad in some spots.

[vivy]

ihave been working on our shocks spring rate , and have calculated , like 260 lb/inch in front and 430 lb/inch rear . with installation ratio of arnd 1.23:1 , pretty much same for rear and front and car weight arnd 320-330kg with driver , so wanna know , that is this too stiff for this much weighing car. our car was having a traction problem at the rear previous year,, if its too stiff , hw can we play with our adjustable MBD's , to lower our spring rate ..
the dyno is not available with us , to chk for differnet combinatiions,, thats a big prob for us too and roll gradient from the springs comin out to be 1.125 .....
and wats the general spring rate value with fsae teams.,, and at this spring rate in front, we can observe that our nose can touch durin brakin g(its lukin too losse to our likin)

vivek bithar
Team defianz racing team
D.C.E[/quote]

[JHarshbarger]

Comparing a 320-330kg car's spring rate to most FSAE car's spring rate is not going to be accurate because most FSAE cars weigh considerably less than that. You can, however, compare the natural frequency which incorporates the spring rate and the mass. Most FSAE cars will run natural frequencies in the range of 2.6 to 3.2 Hz. Some will run higher natural frequencies in the front some will run higher natural frequencies in the rear, but not usually varying more than .4 Hz. The variation of 170 lb/in between the front and rear appears to be quite excessive, unless your weight is strongly biased to the rear and you want a higher natural frequency in the rear.

If the springs you already have are too stiff, you can modify your motion ratio by modifying parts or making new ones, or you can buy new springs.

[B Hise]

or throw some spring rubbers on the front.

[vivy]

our car is 40(front ) and 60(rear ) .. can u elaborate little more on a higher natural frequency rear required for strongly biased to rear cras or its effect on car response if it doesnt happen means front gettin higher frequency for strongly biased rear car ...

on an optimum g tutorial i read it sates that higher frequency at front will help at cornering response and lower frequency at rear for grip on exit of a turn .. that will avoid traction loose at rear...
thnx for the reply joel
hope to get ur reply soon

[JHarshbarger]

Running a slightly higher (~.2 Hz) natural frequency in the front is fairly common in FSAE. I didn't mean to imply that you should have a higher natural frequency in the rear, it just appeared like that was what you were going for with the high spring rate in the rear. I don't have experience with strongly rear baised cars, so I don't know if you should compensate for that. Natural frequencies do compensate for weight difference though. For instance, even if you run a lower natural frequency in the rear, you rear wheel rate will still be much higher than the front when you have that kind of weight distribution.

Make sure you are focusing on the wheel rates when doing your calculations. The spring rate should only be needed when you are ordering the springs.

Sorry if this isn't the answer were looking for. Maybe someone with more experience with strongly rear biased cars can elaborate.

[The Stigg]

Hi Guys,

I know that this was an old post, but in my optinion very important.
As damping ratios and damping curves are one of the key factors in tuning transient behaviour.

For setting up our system I startet with the Spring&Damper Tech Tips from Optimum G. Damping curves looke like the yellow ones in the graph:

Wheel Assembly = 13kg
1/4 Body = 50kg
Kspring = 20 kN/m (~20 kg/cm)
Ktyre = 130 kN/m

http://img406.imageshack.us/img406/5905/dampingcurves02hk2.jpg
http://img406.imageshack.us/img406/6356/damping03tw8.jpg

The red and blue curves are the ones that I would prefer when regarding the simulation and testing results.
But why are these curves so different?
Did I make a mistake?

Regards Andy

[B Hise]

Andy,

As you said, the optimum g numbers are a baseline considering only heave (and no tire damping). Once you get the car into the 3D world, add roll and pitch inertias, everything to consider in the yaw plane, suspension geo. etc. the "best" damper settings can change from what you come up with in those calculations.

Using these calculations is a good tool for a starting point , but testing is always the best way to figure out your preferred settings. If you get a shock with a good adjustment range say .3 -1.2 damping ratio, a few sessions in the parking lot should do you some good. The section in Tune to Win (I think) by Jan Zuijdijk is a very good design of experiment for getting the car right subjectively. And theres always the histogram...

Bryan

[The Stigg]

Thx Bryan,

I know that simplified models ar far away from the real world, but thats exactly the reason for reopen this thread.
It's not so important to know the exact numbers but i want to see a trend in these models.

So real life testing shows me that this is the way our damper curves need to be, but in theory i'm not able to find the reason for that.
And this could be a pain in the ass!

Regards Andy

[B Hise]

You can go further by moving from the quarter car model to the half car model (roll and/or pitch and ignore the unsprung masses) and that should help you find some numerical justification to what you see on track. I think someone posted generalized inertia numbers a while back. I don't remember and am too lazy to search. You can also try to approximate it from your CAD model or you can measure it.

Note that your drivers lap time might improve with adjustments that contradict theory. The driver plays a large role in the end shock setup and the curve doesn't always match up with what the book says it should. Of course the judges aren't really keen to hear that answer so it might take you a little bit to connect the dots.

Bryan

[JHarshbarger]

One thing I noticed is that your compression and rebound damping curves are identical. The rebound damping should be much greater (about twice as much) than the bump damping. Use the yellow lines from your graph as an approximation.

What are you basing your ideal damping curve off of? Do you have a particular damper in mind?

[B Hise]

The rebound damping should be much greater (about twice as much) than the bump damping.

That is not true in practice. Yes the spring also absorbs energy and this is why people stick to this application of theory.

Although, in many cases, a high level of compression with respect to rebound yeilds the best performance on the 7-post, track, and in the sim. The car is much more complex than the heave model suggests and it would be dangerous to adhere to the numerous assumptions it makes when trying to develop a competitive car.

[murpia]

Originally posted by B Hise:
<BLOCKQUOTE class="ip-ubbcode-quote"><div class="ip-ubbcode-quote-title">quote:</div><div class="ip-ubbcode-quote-content">The rebound damping should be much greater (about twice as much) than the bump damping.
...in many cases, a high level of compression with respect to rebound yields the best performance on the 7-post, track, and in the sim. </div></BLOCKQUOTE>
Low-speed or high-speed compression? And improvements where, body control-related handling balance or load fluctuations over bumps (better grip)?

Regards, Ian

[B Hise]

Low speed, high speed, and the shocks fundamental characteristic plays a big role as well. All shocks are not created equal and the steady state curves people calculate and post are not exactly whats happening.

Body control, load control and response can all be improved with the right compression level. It also making tuning with rebound straightforward.

Sometimes a good way to look at things is with pair analysis. Roll the car with no compression and tons of rebound and the inside tire unloads quickly for example.

[Kurt Bilinski]

Originally posted by fade:
that does seem very low (stiff). >1 deg/g would be more suitable for low speed. are you running very close to the ground? the parking lot is pretty bad in some spots.
Fade, what the heck is going on in your avatar?

[HenningO]

You are talking high speed, low speed compression, rebound and all the whats and ifs and don't...

It might be earlier in this thread, I didn't go through all of it, but, first, figure out what shock speeds you are going to see on the track (taking your specific MR into account). Then you'll see if you have to worry about the damping at 200 mm/s.

http://fsae.com/groupee_common/emoticons/icon_smile.gif

[Chris_S]

Originally posted by Kurt Bilinski:
Fade, what the heck is going on in your avatar?

Its the Milliken camber car.

[mtg]

The baselines in the Spring & Damper Tech Tips are exactly that- baselines. I just used them myself to come up with baselines for a race bike I'm building of which no baseline exists.

If you don't know where to start, start with those baselines. If you find that you have better performance with a different setup, by all means use it.

Soundjewel: I'm curious about your simulation. Has it been accurately calibrated to real track data?

[The Stigg]

Hi,

well to be honest these simulations came primarily from a quater vehicle model.
But I added some acceleration data (Upright) and laptimes to be able to find out the differences in simulation an real-life.
What I saw was that the simulation data didn't fit the recorded ones.
Now I'm working on a full vehicle model, with tire damping and recorded road characteristics.
What I can see until now is that the new modell fits quite well what we see on the testtrack (thx to B Hise)
I think its a matter of combinations in roll, heave and pitch that messes the quater modell results up.

[Hub]

Hi,

This post is quit old but it's interesting.
Like The Stigg I've made damping curve with the help of the tech tips section from optimumg and I obtain similar curve. If I've understand it's more a baseline and it's better to have a symmetrical curve like the red and blue curves.

For damping force I'm confuse, because if I compare to DB and penske 7800 curve the lowest setting is higher than needed? For example 100 N at 250 mm/s for me and 500 N at 250 mm/s for the penske 7800.
http://www.motorsportsspares.c...adjustment_sweep.jpg (http://www.motorsportsspares.com/images/db1_adjustment_sweep.jpg)
http://kaztechnologies.com/Pen...ure%202009-09-22.pdf (http://kaztechnologies.com/Penske%207800%20Series%20Technical%20Info%20Brochu re%202009-09-22.pdf) (page 9)

Do you have an advice?

[DJHache]

Keep in mind that your calculations are based on solving linear equations with constant coefficients while you are most definitely not using strictly linear dampers.

The kaz technology pdf mentions some suspension parameters (weight, natural frequency, motion ratio), are your parameters the same?

[Hub]

Higher weight in our car but the natural frequency and motion ratio are in the same range.
Like B Hise said before optimumg formula are baseline but the number differ a lot!

[Crispy]

Does anyone have a good explanation of why damper curves should be asymmetric? I have heard this many times, but in the the math I don't see anything that is different between moving up or down from equilibrium, and I haven't heard a practical/real world explanation that convinces me either.

I have a few ideas, but none of them convince me that we need asymmetric curves for FSAE cars.

-The tire leaving the ground could cause asymmetry, but if this is happening very often I think the design may have larger problems than damping. If the car is getting air off bumps (lots of cars at FSG this year) could having asymmetric damping help in this case?

-The road profile could be asymmetric, meaning bump movements are larger or harsher than rebound movements. I think this would be important for cars dealing with curbs, but we don't have curbs.

-Controlling load transfer transiently? I believe increasing compression or rebound damping should have the same effect on load transfer rate, so I think this is a symmetric effect as well.

Any thoughts?

@Hub: As far as 100 N at 250 mm/s. We mostly look at 125 mm/s for our high speed and 100N seems really low, even for 125 mm/s. I would think typical damping forces should be closer to 200N at 125 mm/s even for a lightweight car and higher for a heavy car (with 1:1 ratios).

[exFSAE]

I don't understand the reasoning given in the OptG tech tips.

How about controlling load transfer rates under forward pitch (braking) versus rearward pitch (accel)?

Or combination of braking, rolling, then accelerating..

[Mike Cook]

Good question. Its kinda funny, because if you look in the newest version of the shock absorber handbook, Dixon dedicates a few pages to this topic. He gives a number of reasons why damper curves tend to be asymmetric, however he does not present a very good argument.

My basic understanding is that per cycle of oscillation a certain amount of damping force is required to give your damping ratio (whatever it may be). You can balance the compression/rebound ratio any way you want. However in practice, for ride quality, I believe that having much more damping on the rebound side is better. This is mainly because because when the wheel hits a bump, the spring force is going to accelerate the chassis up and any damping forces on top of this is just going to increase the impact which will be felt as a harder hit to the driver. If you put the damping force on the rebound side you reduce this impact.

Anyways, for street cars this is generally the idea. I have seen some dampers that were almost 10x rebound as compression.

A while back nascar ran a ton of rebound, but this was mostly to get the car to 'pack' down the straights to reduce drag.

In fsae I don't have much experience with playing with the c/r ratio. In one case, we were trying to simulate a monoshock suspension by putting a ton of rear rebound into the car and taking all the front damping out. The car actually got worse on transient maneuvers which was the last thing I would expect to happen and I think we went past the point of having too much rebound.

Mike

[rkraft]

Crispy,

You're on the right track. But it seems you really haven't looked into the math. Sweep a nonlinear 1/4 car model through a range of biased curves and examine your TLV and other important metrics. You will find something. At Kaz Technologies we have seen this in data from many kinds of cars, including FSAE cars.

Also, your third idea has something to do with it. Changes in compression vs. rebound will only create identical responses for infinitely stiff, lumped parameter systems.

Mike, damping is a non-conservative force. This means it is dissipative. So you need to look at more than the force, you need to look at the system energy. If you attended our damper seminar after MIS 2009, you would know that there are some very telling metrics for energy dissipation that professional damper engineers use to examine just that.

[Hub]

Originally posted by Crispy:
@Hub: As far as 100 N at 250 mm/s. We mostly look at 125 mm/s for our high speed and 100N seems really low, even for 125 mm/s. I would think typical damping forces should be closer to 200N at 125 mm/s even for a lightweight car and higher for a heavy car (with 1:1 ratios).
Tks for your help.

I have low damping force but why?
I use the tech tip formula. I've made an error?

Initial Slope = 4*?*?_ride*?_ride*m_sm
Low Speed Compression Slope = 2/3*Initial Slope
High Speed Compression Slope = 1/3*Initial Slope

Low Speed Compression Slope
?_ride = 0.7
?_ride = 2.6 Hz
m_sm = 69 kg

Low Speed Compression Slope
?_ride = 0.3
?_ride = 2.6 Hz
m_sm = 69 kg

http://h.ds.free.fr/fs/Damping_curve.png

[Crispy]

I have never seen the abbreviation TLV. I'm guessing it means Tire Load Variation, the tires peak to peak load variation.

I ran 5 test, ranging from compression/rebound=2 to rebound/compression=2. Regardless of the bias I use, the frequency response of the TLV remains more or less constant. The sprung mass jacks up (high compression bias) or jacks down (high rebound bias), but the peak to peak and average tire force values are more or less always the same. The sprung mass frequency response for opposite cases (highest compression bias case and highest rebound case) look exactly the same. All I gather from this is that the effects are mostly symmetric, other than the jacking of the vehicle. I can't say that I really want the car to jack either way, so symmetric seems like a good way to go based on this information.

The single bump case is interesting though, higher rebound bias most definitely reduces both sprung mass movement and tire force peaks. But this is only for an "up bump", the opposite is true for a "down bump", provided the tire stays in contact with the ground.

Perhaps my model is missing something, or I need to look deeper. Maybe I have a mistake in my model, but it reproduces the analytical solution for linear damping, so it can't be too far off.

Any thoughts or further suggestions?

@Hub: As near as I can tell your plot matches the first group of low speed compression numbers, but I don't understand where the numbers come from. I'm not saying they are wrong, but how did you come up with them? I'm not familiar with the process you are using (if it's OprimumG I haven't been). Obviously from my comments above I don't know why someone would use asymmetric damping. If you make your damping symmetric it matches pretty well with the values I mentioned in the earlier post.

[rkraft]

Yes, TLV is tire load variation, normally quantified as a RMS value.

If your TLV is not changing a lot with c/r bias, that's not unheard of. TLV is primarily determined by stiffness and damping ratio, not bias: like I mentioned, the energy dissipation is also critical, and the total energy dissipated can be thought of as the area under your curves (don't think about it too much, this isn't exactly correct, but it is an indication). If your curves don't change as you change bias, the area doesn't change.

Like you said, there are other things to look at. Single wheel to sprung mass force and motion transmissibilities are certainly important. But it also has to do with what's going on in pitch, roll, etc (what are commonly known as low-speed events). If you have asymmetry at low speed, and less or none at high speed, then your whole curve will still be asymmetric...

[Hub]

Originally posted by Crispy:
@Hub: As near as I can tell your plot matches the first group of low speed compression numbers, but I don't understand where the numbers come from. I'm not saying they are wrong, but how did you come up with them? I'm not familiar with the process you are using (if it's OprimumG I haven't been). Obviously from my comments above I don't know why someone would use asymmetric damping. If you make your damping symmetric it matches pretty well with the values I mentioned in the earlier post.

Thanks for the advices.
It's from optimumg tech tips.
http://www.optimumg.com/Optimu...mpers_Tech_Tip_4.pdf (http://www.optimumg.com/OptimumGWebSite/Documents/TechTips/Springs&Dampers_Tech_Tip_4.pdf)
The graph correspond to the number, but I found the force to be low.


What method do you guys use to found your damper curve?

[Crispy]

The transmissibility to the sprung mass is the same in terms of the oscillation amplitude regardless of bias. The center of oscillation is simply offset due to jacking up or down.

I would say, in the absence of bump asymmetries, there isn't evidence, from a quarter car model, that using bias has an advantage (unless you want to jack your car up or down).

Or is there more here I am missing?

The next step for me would be a two wheel model or a full car model to look at low-speed events, but that will take a bit longer...

[DJHache]

What would happen if you had a car with significant anti-squat but little or no rear anti-lift? Symetric damping would end up jacking your car up wouldn't it? I think we should be wary of creating general rules to apply to all cars.

[skidpadguy]

DJHache, I don´t understand why anti-dive or anti-squat would affect damper jacking in heave, especially without any longitudinal forces?

[rkraft]

anti geometry almost always results in wheel recession. this will certainly create jacking. There was a great paper on this written in England a while back, but I forget the name of the authors.

[Crispy]

I agree we shouldn't be trying to making rules that apply everywhere. My intent was to look at a simple case and try to understand the effects of biased damping. I would really like to understand the basics, and for me It doesn't make to much sense to move on to more complicated systems until the most basic one is reasonably understood. I apologize if this was not clear.

More clearly: Using a simple 2-dof quarter car model (linear springs, no anti effects), and assuming a symmetric road profile. Can we make a case for asymmetric damping?

[BilletB]

More clearly: Using a simple 2-dof quarter car model (linear springs, no anti effects), and assuming a symmetric road profile. Can we make a case for asymmetric damping?

Yes, yes you can. However, the strength of your case will depend on what your goals are. The case has already been mentioned in this thread.