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Thread: Bushing Stiffness Values for MotionView Simulation

  1. #1

    Bushing Stiffness Values for MotionView Simulation

    Hi everyone,

    I am a current member of the suspension team on Michigan Tech's formula SAE team, and I have been working to complete an Altair MotionView Model of our vehicle's front and rear suspension. The purpose of the model is to predict load cases in different linkages in order to perform more accurate FEA, as well as how the suspension reacts under jounce and rebound scenarios.

    While our team has a good idea what most of the inputs for the program should be in order to get static load results, we are wondering if anyone has any advice on what to enter for translational/rotational stiffness values [N/mm & N*mm/rad] for the joints in the model (rod ends: https://www.mcmaster.com/#60645k12/=1a3jwbz). Based on some feedback I've gotten, my idea was to model the translational stiffness as a value in the ballpark of 1e5 N/mm; does anyone have any advice on what to model as the rotational stiffness? I would think we would want to model it as something much lower since the rod ends are free to rotate to a point. I've tried finding sources online, and I've called to the supplier and McMasterCarr but neither of them are able to supply this value.

    My concern is that if the rotational/translational stiffnesses are wrong, then the force predictions by model could be inaccurate as well. Any input and/or advice is appreciated.

    Thanks guys!

  2. #2
    You could always try to find the values with an experiment...

  3. #3
    Hi!

    I am Claude.

    And you are?

    Not worth to introduce your self?

    Just a question of courtesy; this forum is not a super market!
    Claude Rouelle
    OptimumG president
    Vehicle Dynamics & Race Car Engineering
    Training / Consulting / Simulation Software
    FS & FSAE design judge USA / Canada / UK / Germany / Spain / Italy / China / Brazil / Australia
    [url]www.optimumg.com[/u

  4. #4
    Mi906 or who ever you are.

    English is not my first language so I could be wrong but usually when I read bushing I think of the rubber stuff. In passengers car, not on race cars. That quite different than the image you have in your McMaster link; I am used to call that a rod end. Just to be sure we are speaking about the same thing.

    I would be a bit worry about the rod end you show in that link. I am not sure about its quality and reliability. Usually what is cheap is not good and what is good is not cheap..... If I was you I would try to get some info from Aurora on their website.

    I think it is worth for you to read: Compliance and Friction in Elastic and Mechanical Joints of Race Car Suspensions. 06AC-219. Matthew S. Zipfel and Albert R. George Sibley School of Mechanical and Aerospace Engineering Cornell University 2005.

    I bet modeling a rod end friction could be a big project on itself. They are so many parameters: wear, load, temperature, ...

    I also think you should think about frequency: There is a static and a dynamic friction. The friction could be low but it changes quite often. If you have 4 Hz of natural suspended mass frequency, the load changes direction 8 times per second and that will have a considerable influence on your tire load consistency.

    There is a reason why some race team use blades instead of rod end
    Claude Rouelle
    OptimumG president
    Vehicle Dynamics & Race Car Engineering
    Training / Consulting / Simulation Software
    FS & FSAE design judge USA / Canada / UK / Germany / Spain / Italy / China / Brazil / Australia
    [url]www.optimumg.com[/u

  5. #5

    Ferrari F1 top wishbone picture

    Look at the top wishbones "rod ends"

    Audi won Le Mans a few years ago with such blades on the lower front wishbones so reliability can be addressed.
    Attached Images
    Claude Rouelle
    OptimumG president
    Vehicle Dynamics & Race Car Engineering
    Training / Consulting / Simulation Software
    FS & FSAE design judge USA / Canada / UK / Germany / Spain / Italy / China / Brazil / Australia
    [url]www.optimumg.com[/u

  6. #6
    Titanium blades inserted in Carbon fiber wishbones: that is a nice university project!
    Claude Rouelle
    OptimumG president
    Vehicle Dynamics & Race Car Engineering
    Training / Consulting / Simulation Software
    FS & FSAE design judge USA / Canada / UK / Germany / Spain / Italy / China / Brazil / Australia
    [url]www.optimumg.com[/u

  7. #7
    Tim, I know people who model bushings (the rubber stiff) in frequency and in 6 D. Not that simple I can tell you! Even bushing manufacturer do not have a reliable, relevant model; car manufacturers do their own measurement they trust more than the ones from the manufacturers. But you will tell me that is also true for tires...
    Claude Rouelle
    OptimumG president
    Vehicle Dynamics & Race Car Engineering
    Training / Consulting / Simulation Software
    FS & FSAE design judge USA / Canada / UK / Germany / Spain / Italy / China / Brazil / Australia
    [url]www.optimumg.com[/u

  8. #8
    I heartily second Claudes reference to the Zipfel and George paper. It was a pleasure to interact with Mathew on the project, and it is well respected within the racing industry.
    In fact, i was not aware at the time that it had been published until a prominent American designer called me referring to it.
    John McCrory
    Racecar Product Manager
    Aurora Bearing Company
    jmccrory@aurorabearing.com
    630-859-2030 x 3088

  9. #9
    Hi MI906 (whoever you are...)

    I woudln't worry too much about modeling these bushing stiffnesses in an FSAE car. The only reason they're there is to "dampen" the whole system since the bodies used in this kind of analysis are inifitely rigid, which is not the case in the real world. So by combining a rigid body and a "bushing" you can replicate this behaviour. Concerning the rotational stiffness, there is another reason for this: If you look at a member of your suspension system that has ball joints at both ends (like the steering rod or the pushrod), in a simulation with ideal balljoints they would have a rotational DOF along their axis and so they'd start rotating, which makes the solver fail to converge, as there is nothing stopping that rotation. So, there needs to be a tiny bit of rotational stiffness and also damping to prevent that.

    The influence of this modeling technique should be negligible, but why don't you just try that by changing the values a bit?

    Regards, Luniz
    Last edited by Luniz; 12-11-2017 at 04:32 AM.
    Lutz Dobrowohl
    2008-2011
    Raceyard Kiel

    Now: Scruitineer, Design Judge, application engineer @Altair engineering

    Whatever you do, do it hard!

  10. #10
    Senior Member
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    Modena, Italy
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    "Bushing" is MBS-speak for a 6 d.o.f. joint. It can be as rigid or a flexible as you want in any direction. If you are just calculating FEA loads then setting the rotational stiffness' to zero and the translational stiffness' very high (>50kN/mm) it should get what you need if you are using ball joints everywhere. Watch out for the toe-link which might need a small amount of rotational stiffness otherwise the solver won't converge.

    Anyway, you should have some sort of intuition as to what level of compliance will change the geometry enough that it will affect your results.

    I don't know why Claude recommends looking at joint compliance and friction. Seems like an unnesscessary complication if all you want are FEA loads...

    Keep in mind that a model made in this was is not going to give you any decent idea of you toe and camber compliances because they are primarily due to part compliance not joint compliance (with the exception of the wheel bearing and the $3 joint from Mcmaster).

    Have to say I don't like the idea of a $3 component in the primary structure of the suspension. The issue with a $3 rod-end is that in the best case scenario it will very quickly develop free-play which will destroy your driver feedback and worst case it will just fail.

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