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Thread: Roll rates in RCVD

  1. #11
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    For the Z bars... packaging has changed a lot since the 1900's just because there is so much "stuff" in a car now (especially sports cars with the engine behind the driver) that the only way to link the suspensions like you mentioned would be hydraulically.

    Anyway whether hydraulic or mechanical, a diagonal linking is an extra complication. Yes it will improve ride + body control but so will a lot of other complications. The manufacturers have decided its not worth the cost/complexity. I don't understand why this makes them a pack of morons...

    Regarding the spring rate calculations... you are really nit picking there. Both calculations (based on displacement or frequency) are very simple and quick. To me they are the same thing and I find it hard to justify wasting much breath to argue about which is better.

  2. #12
    Quote Originally Posted by Z View Post
    Jose,

    "120 views and no responses."

    Better now?

    Z


    Quote Originally Posted by Z View Post

    For your Spring-Damper design I suggest you look at U of Cincinnati's 2013 car (see links and discussion by Matt (mdavis) on the "FSAE Lincoln 2013" thread, in the Competitions section). They had direct-acting SDs (like Monash, winner of 2013-Oz), no ARBs, rubber spacers for spring-rate adjustments, and they did quite well! At most you might need three sets of springs (soft, medium, and stiff), with my earlier "static displacement" calc getting you in the ball-park for these rates.

    Z
    Regarding the push/pullrod vs direct acting dampers, I have a question. I've read that one of the advantages of the push/pullrod method, is the reduction in unsprung mass. Now, unsprung mass is defined as the mass of the wheel, (...) etc and half (or, an undefined portion) of the mass of the connecting elements (spring, damper, bellcrank, etc). My question is, what is that portion? How could I calculate it?
    I don't actually think that it is that important, for most purposes I guess 50 % is fine. I am only asking because, by that definition, a direct acting shock abosrber would actually reduce the unsprung mass, because you don't have the rocker or the pushrod. Right?

  3. #13
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    Interesting question...

    Remember that the unsprung mass is (another) simplification to represent the total "inertia" or "resistance to movement" of the total suspension. It is not a real mass but an "effective" mass which gives the same resistance to vertial motion as the complete system (of pushrods pullrods Z bars, dampers - only the moving part! etc etc...).

    The best way to measure it is to move the suspension and measure the forces required to do so. The effective mass (your unspring mass) is then the ratio of the acceleration from the movement to the force required to do so or as Newton said it: M=a/F.

    So anything that moves will be adding to the effective unsprung BUT some of those things will be contributing to it via its POLAR inertia not its mass. For example a rocker. Its mass is largely irrelevant because it doesnt move. Its inertia is what is resisting acceleration of the system.

    So you can see, its not such a trivial thing to calculate. But given that your calcs are already a simplification (i.e. they are representing the complete suspension as a single point mass at the wheel centre) then if you are 10-15% out in your unsprung mass determination, its not going to be a show stopper.

    Enjoy

  4. #14
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    Tim,

    For the Z bars... packaging has changed a lot since the 1900's ...
    ... the only way to link the suspensions like you mentioned would be hydraulically.
    NOoooo!!! They are just two long straight bars that fit under the car. Very similar to ARBs, but longitudinal rather than lateral (and Z rather than U ends).

    ... whether hydraulic or mechanical, a diagonal [actually "longitudinal"] linking is an extra complication.
    Again, noooo!!! Just three springs, namely 2 x longitudinal + 1 x lateral Z-bars, can provide the entire suspension for a car. Conventional suspensions have six springs, namely 4 x corner + 2 x ARBs. Which is more complicated? Which is simpler?

    Furthermore, as I have explained many times before, the difference in performance is like night and day. But until suspension engineers exerience this, THEY WILL NEVER KNOW.

    The manufacturers have decided its not worth the cost/complexity.
    And here is the key point. The manufacturers, for the most part, do not have a clue about these sorts of things. Where are the papers? The books? The seminars? The in-house design reports with extensive experimental results, cost-benefit analyses, etc., etc? Certainly, Claude is not teaching this stuff in his OEM seminars. Or are you, Claude?

    For the record, I am sure there were many French papers in the 1930s-50s. Also the Packard paper of the 1950s is a cracker, especially some of the comments in the post-presentation discussion. (Edit: Found it. "The New Packard Torsion Level Suspension", F. R. McFarland, SAE meeting June 1955, 560026, vol 64, 1956 SAE Transactions. Thanks Forbes!)

    So I am sure there are SOME people deep in the bowels of the big OEMs that have some idea of these systems. But I am equally sure that those people DO NOT make the big decisions.

    Regarding the spring rate calculations... you are really nit picking there. Both calculations (based on displacement or frequency) are very simple and quick. To me they are the same thing and I find it hard to justify wasting much breath to argue about which is better.
    Like I said, THE DIFFERENCE IS NIGHT AND DAY. The conventional calcs will forever keep the auto-industry in the dark. You can't even start to calculate a "front/rear ride frequency" using the conventional approach on an interconnected suspension. Too hard! (Ask Claude.) So NO progress. Daylight is much better, IMO. Which is why I am wasting my breath...
    ~~~~~o0o~~~~~

    Jose,

    Regarding the push/pullrod vs direct acting dampers, I have a question. I've read that one of the advantages of the push/pullrod method, is the reduction in unsprung mass.
    ...
    I am only asking because, ... a direct acting shock abosrber would actually reduce the unsprung mass, because you don't have the rocker or the pushrod. Right?
    Well, IMO, you are absolutely RIGHT.

    But I am not an expert, and the experts have so far been very reluctant to discuss this matter. Or to "defend their positions" on the matter, as they so often tell the students they must do. (In fact, one expert, Pat Clarke, has just recently notified me that I have received "1 infraction point" for my comments on this thread, with veiled threats that I will be banned from the Forum!. Yes, the Thought Police have learnt a lot from their trip to India! )

    Anyway, I recently discussed the effect of Motion Ratios on unsprung mass on this "How to design a bell crank for push rod suspension" thread. In brief, not only do the pushrod and rocker ADD their mass to the total unsprung mass for that corner, but their mass is multiplied by the MR SQUARED. So PR&Rs, especially when designed for high MRs, certainly do increase unsprung mass.

    Comments and criticisms from the experts that support push/pullrods&rockers with high MRs, yet claim they REDUCE unsprung mass, most welcome!

    Z

    (PS. Just read Tim's post above. Could add a lot more, but no time just now. Well, for rotational motions the inertia is ~ M*R*R, and note which term is squared! And kinetic energy is ~ M*V*V, and note which term is squared! Ie., parts which must move fast are much harder to get moving...)
    Last edited by Z; 03-06-2014 at 07:13 AM.

  5. #15
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    Quote Originally Posted by Z View Post
    NOoooo!!! They are just two long straight bars that fit under the car. Very similar to ARBs, but longitudinal rather than lateral (and Z rather than U ends).
    They aren't just 2 longitudinal bars...

    They also have the lever arm at each side, which on one of the axles needs to swap sides of the car. So you need longitudinal space to package the torsion bars (already getting tricky for a mid-rear engine car with no transmission tunnel) and then lateral space for the bars to swap sides of the car and swing an arc through the wheel travel movement without hitting anything. Its quite a significant increase in packaging space compared to an anti roll bar. And given that it needs to span the wheelbase of the car, I can imagine the diameter isn't going to be as compact as a typical anti roll bar either.

    As luck would have it, I have open now a CAD model of the back of mid-rear engined car. You've got buckleys chance of fitting anything in longitudinally. I dont think you would get far either asking the manufacturer to reduce the luggage space in the front in order to swing the lever arms of the Z bars either. For a front engined car maybe.

    For FSAE, I think it is a very viable solution which should be considered. I'm already considering a soft twist mode suspension for my own project car.

    I respect your theoretical talent Erik, but in my opinion, one of your weaknesses has always been the jump from theory to reality. A lot of what you propose simply isn't practical these days. Particularly for modern cars which are becoming increasingly tightly packaged.

  6. #16
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    One other thing I would add. The use of interters in F1 and some other racing categories is aimed at INCREASING the effective unsprung mass in order to get better control of the contact patch load variations.

    So what Z says is right (about MV^2 and MxR^2) BUT its not nessarily a bad thing. Its not such a straightforward thing to analyse though.

    I realised I didn't really address the question on the push rods either...

    If you compare 2 setups with the same spring/damper and same wheel rate, I suspect (though Im not sure) that the effective unsprung mass contribution from the damper will be the same (since the work = force x disp is the same). The push rod and rocker then will be in addition to it. So I agree with Z on this point...

    So unless your push rod setup allows you to use smaller dampers, I suspect you will gain unsprung mass.

    Anyway, in my opinion, the choice of motion ratio (and therefore whether to use direct or push rod) should depend on what dampers you use and what damping rate you want at the wheel. Springs are easily made to size, but damping cannot be easily adapted to all motion ratios.

  7. #17
    Interesting discussions going on here.

    Z, your input is always very interesting, above all from theoretical perspective. I always got some very good food for though reading your comments, but here i have to say i am pretty much with Tim. At least when calculating Wheel Rates or Suspension Frequencies, i don't see practically a big difference between the results we could get with your (formally more correct) approach and the one from RCVD or Tim.

    Moreover, above all when we talk about suspension movements in the order of 20 mm, my experience has always showed me that the results i could calculate with the "classical" vehicle dynamics methods were very close to the ones measured on track (at least with properly calibrated sensors). Of course, as shown in the discussion regarding Anti jacking effects, this method is formally not correct (or, anyway, a simplification) but to be honest i think it could be considered correct within reasonable approximation boundaries.

    Regarding the discussion about the unsprung mass definition, if the car was already built i would suggest to simply disconnect the pushrod/pullrod (or damper) from the rocker (chassis) and measure with the scale under the wheel your unsprung mass, keeping the car away from ground with jacking supports, for example (this would of course ignore the mass of the rocker and of the damper, but according to any theory we should use only a portion of these). Unfortunately, when the car is still on paper (or in excel, or in CAD) it is a bit more difficult, but i guess that the 50% approach gets close.

  8. #18
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    Tim,

    I respect your theoretical talent Erik, but in my opinion, one of your weaknesses has always been the jump from theory to reality. A lot of what you propose simply isn't practical these days. Particularly for modern cars which are becoming increasingly tightly packaged.
    As I explained on another thread (somewhere?) I modified a car to "3-Z-bar" suspension about 30 years ago. It took about a weekend, or maybe two. I used an angle grinder, stick welder, and a bunch of miscellaneous junk, working under a tree. Sure, it was only a "paddock basher", but, I repeat, the difference in performance was like night and day.

    I will happily bet $1M with any OEM, or smaller "Supercar" manufacturer, or anyone else, that I can successfully modify their "modern" densely-packaged car to 3-Z-bar suspension. The end result will be simpler, lighter, cheaper to make, and have far superior performance, to what they originally had. (I would happily bet $100M, but I only have $1M available right now.)
    ~o0o~

    And here is where my practical experience trumps your theory (to continue the gambling analogy ).

    They aren't just 2 longitudinal bars...

    They also have the lever arm at each side, which on one of the axles needs to swap sides of the car...
    Once again, NOooooo!!!

    Longitudinal-Z-bars connect side-pairs of wheel (which is why these, and longitudinal-U-bars, are often referred to as "side-pair" springs).

    What you are referring to are "diagonal-Z-bars". These resist (ie. control by their spring stiffness) Heave and Twist motions, but provide NO resistance to Roll or Pitch motions. Briefly, such springs would be terrible for a car's performance (though not quite as bad as ARBs/lateral-U-bars!).
    ~o0o~

    Given that this subject is so poorly understood in the automotive community, I will post more on it in the next few days, but on another thread.

    For now I note that throughout the Packard paper referenced above they talk a lot about the "bounce and pitch ride frequencies" of the car, and also about the "centres" for these motions. This was, no doubt, following on from the work of people like Rowell and Guest in the 1920s. And these people only had pencil and paper for their calculations, or maybe slide rules. Back then, no computers to do all the work! (Note, however, that the calculations are extremely simple.)

    Sadly, due to the decline of educational standards over the last 50+ years, the majority of you seem to think that the much more puerile concepts of "front and rear ride frequencies" are just as good.

    (Oh, and BTW Pat, in case you don't have a dictionary, "puerile" means "childlike silliness". Hardly "inappropriate language"!)
    ~~~~~o0o~~~~~

    Silente,

    My main point is that FSAE is supposed to an EDUCATIONAL exercise for student engineers. As such, I reckon the students should be made fully aware of the limitations of whatever "theories" they are being taught, and they should also be pointed in the direction of any more accurate analyses, should they want to go further. In fact, RCVD has a section that covers this idea of a "ladder of abstraction" quite well.

    I repeat that the more accurate "bounce and pitch frequency" analysis is extremely simple (for those of you who might want to calculate actual "frequencies", though hardly important in FSAE). But, sadly, very few people are even aware of it these days.

    Also, considering this from the opposite direction, Jose started this thread because he thought (because he was taught?) that he had to FIRST choose a "ride frequency", before he could THEN calculate his "roll-rate". This is nonsense (which in polite English is often referred to as "codswallop", Pat), because it overcomplicates a much simpler matter. As I explained, the roll-rate can be calculated in one very simple step (in your head), with no need for any fictional "frequencies".

    I have often been puzzled by students asking me "how do we calculate the spring rates". I say "puzzled", because FSAE conditions are so simple that I couldn't see where the problem was. I only became aware of the source of the student's difficulties when they started mentioning "ride frequencies". That is BAD education!
    ~~~~~o0o~~~~~

    To sum up, in sociological circles there is the expression "fear of the new" to describe H. Sapien's common and irrational aversion to anything unfamiliar. (Well, to a certain degree it makes sense in the "wild", but it is still irrational.) In the modern automotive world, "bounce and pitch frequencies" and "interconnected suspensions" are such things.

    But this behaviour should more accurately be described as "fear of the old, but forgotten"!

    More on the Packard paper soon...

    Z

  9. #19
    I do fully agree with Tim. Just as with modelling the very basic theory of vehicle dynamics can be grasped "rather" quickly but the more one learns or the more one refines a model the more questions come up. There is absolutely nothing wrong with learning traditionally all topics the hard way. It just takes time and shortcut's leave by definition gap's that sooner or later are going to catch you. There are many good books about vehicle dynamics and suspension design but there is only one bible: http://books.google.it/books/about/R...AJ&redir_esc=y .The book is very expensive and extremely tough literature but once understood of a fascinating simplicity that takes away all the glory of "guru's". Anybody that claims to be an expert in Dynamics must have read this book and if he did will confirm this statement.

    At the end I would like to add my 2 cent's to a longitudinal Z-bar .... the most famous (european) car for having a (hydraulic) Z-bar concept was the Citroen 2CV and drove like a giant pile of c...p. Brake and Acceleration Pitch were amplified and any roll center height jacking on either end of the car raised the other of the car.... so much for the inherent disadvantages of that system.


    Quote Originally Posted by Tim.Wright View Post
    In the interests of a balanced discussion, I will counter some of Z's points.

    The methods outlined in RCVD are simplifications and yes simplifcations require some cutting of corners. In this case, following the "traditional" method, you are largely constrained to end up with a traditional suspension with 4 springs and 2 anti roll bars. For passenger cars, this arrangement is a very good compromise between performance and complexity. Packaging anti roll bars is very simple. Packaging a longitudinal Z bar is a nightmare, and for what exactly?

    At the end of the day, ANY model or method (including Z's) is necessarily a simplification and the traditional methods are quite a reasonable way to arrive at an adequate (but not optimum) solution without a lot of time or large amounts of input data. As you have mentioned, you don't have the time to do something out of the box so following a well established (and sucessful) route is a good choice from a risk management point of view. Even Z would agree, its more important to get the car built 1 month earlier than to agonise for 1 month more in the design phase about suspension rates.

    The main problem though, from an educational point of view, is that often the traditional methods are taught as though they are the first principles of a suspension system whereas they absolutely are not. They are a (quite heavily reduced) simplification of how a suspension operates in roll and pitch. Additionally, they only have a limited range where they are valid (and this unfortunately does not include limit behaviour) and some suspension geometries cause the calculations to crash.

    Your intuition should be telling you this anyway since these methods simply a complex system of 2 wheels + elastic tyres, 12-20 links located in 3D space, connected by a compliant chassis by compliant joints down into an equation with only 9 or so parameters (hrcF, hrcR, KwhlF, KwhlR, KrollF, KrollR, mass, mass_dist, cgh).

    Regarding the book, there is too much good solid theory in RCVD (particularly the chapter on steady state stability and control) to simply brush it off as codswallop. My advice would be to follow whats in RCVD or in the Optimum G tips by all means but recognise you are cutting some corners by doing so.

  10. #20
    Quote Originally Posted by dynatune View Post
    At the end I would like to add my 2 cent's to a longitudinal Z-bar .... the most famous (european) car for having a (hydraulic) Z-bar concept was the Citroen 2CV and drove like a giant pile of c...p. Brake and Acceleration Pitch were amplified and any roll center height jacking on either end of the car raised the other of the car.... so much for the inherent disadvantages of that system.
    It’s worth mentioning that the 2CV was designed to be a cheap car capable of carrying eggs across a rough field without breaking them, not a sports car.

    The 2CV suspension springing arrangement is different to what Z is advocating. It does effectively have longitudinal Z-bars to resist heave and roll, but it then has what is effectively a pair of longitudinal U-bars which resist pitch and warp. Note that it is the longitudinal U-bars that control the pitch issues you mentioned.

    This arrangement is quite different to the 2 longitudinal + 1 lateral Z-bar arrangement which would result in zero warp stiffness. Then again, the 2CV Z-bar implementation is not hydraulic so perhaps you are thinking of a later model Citroen?
    Nathan

    UNSW FSAE 07-09

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