When suspension and chassis are stressed to a G load what is this value based on?
The unsprung mass multiplied by g?
The total mass on the corner multiplied by g?
The mass of the car multiplied by g?
Some other value?

Incidently I know of an actual test in which g was measured in a truck driving over potholes and speed bumps. The aim was to check what forces fragile cargo would be subjected to. The answer was 3g.

If you're analysing a whole vehicle then it would be the whole vehicle weight. However if your attempting to analyse a suspension 'corner' it would be whatever weight is being carried by that corner.

It's also direction dependant; can apply during cornering, straight line acceleration etc... 'g' is simply a unit of acceleration based upon the acceleration due to gravity and allows you to multiply up the loadings on a component in any situation.

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Originally posted by cymtriks
When suspension and chassis are stressed to a G load what is this value based on?
The unsprung mass multiplied by g?
The total mass on the corner multiplied by g?
The mass of the car multiplied by g?
Some other value?

Incidently I know of an actual test in which g was measured in a truck driving over potholes and speed bumps. The aim was to check what forces fragile cargo would be subjected to. The answer was 3g.

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Cymtriks,

I'm sure there will be a few differing opinions on this one.

I've always taken the worst case load on a corner ie max cornering and braking therefore most loaded wheel is the front outside. Calculate weight transfer front/rear and then left/right.

Then you need to do a simultaneous equation to resolve wishbone forces by balancing linear forces left to right and taking moments about the contact patch.

You will then need to apply a reserve factor of your own.

Damper mount loads. Well you have the reaction of the weight on this loaded corner vertically through the tyre, so you will need to resolve for the load in the pushrod/damper. Shock/g loads are in addition to these. Whatever weight is on that corner, you are trying to accelerate (not just the unsprung mass) as the damper should be considered solid. So you multiply the corner reaction force by the shock acceleration (whatever g you choose).

I would say the shock values are rather more than 3g. The parcels acceleration would be damped as it has no mechanical link to the truck.
Do a quick calc on the vertical acceleration when hitting a 2inch bump at 60mph, that'll give you an idea- this is damped by the tyre though

Sorry if that was all a bit brief, but if you agree with this approach and have any questions, drop me an email and I'll try to help. I know you are a very capable engineer and so I'm guessing I wont be telling you anything new though


Cheers

[Edited on 11/1/08 by nasty-bob]

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Originally posted by nasty-bob
Shock/g loads are in addition to these. Whatever weight is on that corner, you are trying to accelerate (not just the unsprung mass) as the damper should be considered solid. So you multiply the corner reaction force by the shock acceleration (whatever g you choose).

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Originally posted by nasty-bob
I would say the shock values are rather more than 3g. The parcels acceleration would be damped as it has no mechanical link to the truck.
Do a quick calc on the vertical acceleration when hitting a 2inch bump at 60mph, that'll give you an idea- this is damped by the tyre though

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It's easy to get confused when people start talking about loads in terms of 'g' levels. A fairly standard 3g vertical bump loadcase would be 3x the corner weight applied vertically - this does not mean that the wheel's acceleration is 3g for such a load case. The 'g' term for loadcases is a convenient way to express that the load is normalised to the static load on that wheel.

3g is also a typical maximum whole body acceleration that you would measure on any vehicle with suspension be it a passenger car or a truck. This is what was measured on the back of your truck an acceleration NOT a load.

Have a look at my post at the end of this thread there is a link to a pdf of a presentation with some fairly industry standard loadcases

Chris,
Is that corner weight the total weight on the corner or just the sprung weight?

I'd have thought that if a load is present then, by F=ma, there must be an acceleration and vice versa.

That's a very interesting table in the link but I can't see any reference to how the "g" factor is used, i.e. a factor applied to corner weight or to some other value.

Where do you work btw?

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Is that corner weight the total weight on the corner or just the sprung weight?

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I'd have thought that if a load is present then, by F=ma, there must be an acceleration and vice versa.

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That's a very interesting table in the link but I can't see any reference to how the "g" factor is used, i.e. a factor applied to corner weight or to some other value.

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Where do you work btw?

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I'm not saying this is how the professionals do it, but this is how I applied G loads to an FEA model for an analyses at university (back in 1992 - ideas may have moved on). I recall at the time that there was no definitive "thats the way to do it" moment - but the approach did merit the award of a first for the final year dissertation. The car was a mid engined BEC of similar configuration to an Elise and in a pleasant coincidence the chassis was of similar configuration being twin torque tubes either side of the passenger space joined by a number of transverse bulkheads. My torque tubes were composite aluminium panels rather than the extrusions that Lotus employed.

2 load types were considered, those that resulted from accelerating the car (brake, cornering, acceleration and bump) and a torsion test.

The 1st set of load cases were modeled by:

* The loads were assumed to apply through the suspension pick up points
* A "global acceleration" was applied to the complete model. This introduce the loads created by accelerating the mass of the chassis panels themselves.
* The loads imposed on the chassis by accelerating the major car components were modeled by applying the forces that would result at the appropriat points on the chassis model. The major masses were:
o engine and gearbox unit
o driver and passenger

The panels were modeled using ANSYS' STIFF99 shell elements.

So considering brake loads the following constraints were applied:

Taking the road as being the XZ plane the front suspension pickup were fixed in the X, Y and Z planes
The rear pickup were fixed in the Z plane.

Reading back through the analyses a couple of simplifications jump out:

* Although the model did allow for weight transfer affects it did not model the loads on the chassis that result from the brake calipers trying to twist themselves around their hubs.
* The loads were assumed to be quite a bit lower than those that are accepted now, being:
o 2.5G braking
o 2G lateral acceleration
o 2G acceleration
o 5G bump (this load case created the highest stress levels)
* the target torsional stiffness was 6.8 kn (I think that this value came from Vehicle Dynamics by Barstow but it may have been Costin & Phipps))

As I say it was along time ago, and although I was able to get guidance as to how to use the FEA tool, I didn't find any standard methods of modeling the loads on a car chassis.

I'm currently designing a Lotus 7 using similar chassis format and hoping that the FEA tools within SolidWorks will prove adequate. So any construction criticism of this approach would be very welcome.

Cheers,
Puk

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Originally posted by chriscook
What type of vehicle are you looking at this for?

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Originally posted by Puk
Lots of interesting stuff

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Originally posted by cymtriks

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Originally posted by Puk
Lots of interesting stuff

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My FEA models are in Nastran. I do FEA for a living!

My models assume that the rear suspension mounts are fixed vertically and that loads, up and down, are applied at the front mounts to give a twisting couple.

I'm happy that I can design a STIFF chassis for a given weight but most of the stuff I can find on "g" loads, so that I can assess the STRENGTH is a bit obscure, hence the original post.

I wonder how people who stress chassis assume they are supported?

One way would be to fix three suspension mounts vertically (or other directions as required) and apply the "bumps" at the fourth.

However this has to be wrong, real chassis are not supported like that!

I suppose I could put spring elements in but then I'm adding complications and changes that might not be assumed in the g loads commonly quoted.

It's a while since I read Costin and Phipps (from the library) and I can't recall how they assumed loads were reacted.

I doubt that the g loads assumed in traditional methods are actually based on a fully dynamic FEA model with all the spring stiffnees and masses in, which is what is actually required.

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LOL - thanks for saving us from the first draft of your mail! But seriously Uphilll Racer if you can point me at a good reference then I'd be grateful.

Cymtriks I think that it was Costin & Fipps (library copy too) that suggests the torsion stiffens target. From memory I don't think they did cove the design of the structure to achieve this – but rather described testing the completed chassis. Still the torsional stiffness target has the advantage of being easy to comprehend an easy to model (be that on paper or 'puter). But there is always the nagging fear that hitting a kerb is going to crumple your pride and joy :-(

Kikiturbo - where do the "5 longditudionally, 4 verticaly, 3 side." accelerations come from? I've seen them quoted in the internet previously but without any explanation. Dennis "dpcars.net" Palatov suggests 4G in each direction – but again I'm not sure where the figures come from.

Whichever is chosen the approach suggested by Nasty Bob to derive the force applied to the outside front wheel looks straightforward.

That leaves the question of constraints unanswered.

Assuming that the road is the XZ plane, with X parallel to the longitudinal axis of the car and Z is up, does the following make sense?

Fix the rear wishbone mounts in X & Z, the damper mount fixed in Z.

The inside front wheel, it should be free to move upwards and longitudinally but would be pulled towards the inside of the curve. The force pulling it in would be a function of the load on it, slip angle, friction coefficient. There must be a way of simplifying that! So no constraints just the appropriate forces acting in Y and Z.

It would be interesting to see whether a chassis that hits the torsional stiffness target is strong enough to cope with the worst load case.

Finite element analysis is no substitute for good engineering. It is a tool for a good engineer, nothing more. It is up to the design engineer to determine which problems would benefit most from its application and apply it to the understanding and resolution of those problems. A finite element representation can certainly give you figures with great accuracy with likewise complexity but in the first instance how do you know if this figure/figures is what you want or indeed need. One has to first make an assumption that the integration of all the constituent parts are going to provide the anaylsis results representing true real world comparisons. IanS

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Originally posted by Ians
Finite element analysis is no substitute for good engineering. It is a tool for a good engineer, nothing more. It is up to the design engineer to determine which problems would benefit most from its application and apply it to the understanding and resolution of those problems. A finite element representation can certainly give you figures with great accuracy with likewise complexity but in the first instance how do you know if this figure/figures is what you want or indeed need. One has to first make an assumption that the integration of all the constituent parts are going to provide the anaylsis results representing true real world comparisons. IanS

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Puk... the 5,4,3 rule comes from Greg Locock who designes suspension and other automotive systems for a livving.. he can be found on various forums, namely Eng-tips (reccomended) and atlasf1... here is one of numerous desgn loads thread from atlasf1/autosport :
http://forums.autosport.com/showthread.php?s=&threadid=98827

or
http://www.eng-tips.com/viewthread.cfm?qid=203004&page=2

eng tips is worth a read.... great forum..

Kikiturbo thanks for the site links. I just logged into eng-tips, it will take an extraordinarily act of will power to avoid frittering the day away browsing the site!

Cheers,
Puk

Uphill racer
My intention was certainly not to make it sound easy, just the opposite. Lots of definitive "assumptions" have to be employed unless you can be 100% sure that, say in the case of a spaceframe structure, every weld has 100% integrity and build in to your FEA equation the fact that on composite monocoques even with high grade professional purchases delamination can be a factor. FEA when carried out on clearly defined and graded aluminium billet machined parts can offer a greater area of accuracy within FEA analysis.

As an initial design aid it works well but at the end of the day empirical testing is just as important. (I am sure all F1 tubs etc have been fully FEA processed but are still required to undergo crash testing.) When I was commissed to build Allan Staniforth's present hillclimb car, Allan sent me his handmade scale balsawood chassis. A simple torsional twist fractured one of glued joints. Was this a fault in design or simply not enough glue in that area???? So I have to reiterate, it is not as simple as plot it out on the computer and move a few nodes around. Ians

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Originally posted by Uphill RacerMy models assume that the rear suspension mounts are fixed vertically and that loads, up and down, are applied at the front mounts to give a twisting couple.

I'm happy that I can design a STIFF chassis for a given weight but most of the stuff I can find on "g" loads, so that I can assess the STRENGTH is a bit obscure, hence the original post.

I wonder how people who stress chassis assume they are supported?

One way would be to fix three suspension mounts vertically (or other directions as required) and apply the "bumps" at the fourth.

However this has to be wrong, real chassis are not supported like that!

I suppose I could put spring elements in but then I'm adding complications and changes that might not be assumed in the g loads commonly quoted.

It's a while since I read Costin and Phipps (from the library) and I can't recall how they assumed loads were reacted.

I doubt that the g loads assumed in traditional methods are actually based on a fully dynamic FEA model with all the spring stiffnees and masses in, which is what is actually required.

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Originally posted by Puk
That leaves the question of constraints unanswered.

Assuming that the road is the XZ plane, with X parallel to the longitudinal axis of the car and Z is up, does the following make sense?

Fix the rear wishbone mounts in X & Z, the damper mount fixed in Z.

The inside front wheel, it should be free to move upwards and longitudinally but would be pulled towards the inside of the curve. The force pulling it in would be a function of the load on it, slip angle, friction coefficient. There must be a way of simplifying that! So no constraints just the appropriate forces acting in Y and Z.

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Hmm indeed!

I'll have a ponder on this and see if sketching some free body diagrams doesn't make thing clearer. At least to me!

Cheers,
Puk

The most accurate way is to restrain and load by the hubs with included suspension links. This is how a car is tested in real world situations. It is illogical to do tests anywhere other than from the hub mountings using a fixed shock. the degree of freedom of FEA is the same as a physical test. Ians
Rescued attachment torsion_screenshot_DT-5-10-04.jpg

Hi IanS - I can see how that would allow to assess the combined torsional rigidity of the chassis and its wishbones. But does that help when trying to establish whether introducing say a YG upward acceleration one one wheel combined with a XG longitudinal deceleration and a ZG lateral acceleration (outer wheel hits pothole whilst bleading of brakes before apex of a bend) will cause plastic deformation?

Some theoretical data whilst we are always expanding our personal knowledge of a subject can often be superfluous unless backed up with physical testing. Given that we find an answer to our problem, does this answer provide us with the solution. Determining accurate loading, is a complicated process that yet again to use my favourite terminology requires some assumption making. In this case the projected (as yet undefined) loads should indicate a worse case scenario. In the scenario you indicate, pot hole, braking, turning moment, would require some tyre data as compression and aligning torque have to be included. The forces must also include vehicle weight and achievable weight transfers. Don't forget as measurable loads it should also contain moments when the tyre is cantelevered off the wheel/upright. A tyre cannot experience 100% of longitudinal and lateral loads at the same time. As these loads have a shock absorbing effect on the upright/wishbone/chassis mount etc where is this data obtainable from (another assumption). Given that we cannot include one of the most important aspects of vehicle design/dyanmic loadings the next best solution is to apply the loads at the spindle bearing centre. I feel that the best starting point is as you suggest, starting with a free body drawing. Whatever ball parkyou end up in is going to be a large ball park. Sometimes a guesstimate on strengths required for certain components has to be just that. Ians

I'd forgotten how long it was since I had done any FBDs. I'm sitting surrounded by disintegrating text books!

Ians... with all due respect, I have to disagree with your torsional rigidity test... for two reasons..

pure chasis torsion test, with rear shock mounts and possibly rear suspension mounts being fixed points, and front shock or rocker bearing (depending on the type of suspension) being input nodes is more relevant to the chasis rigidity case, as pure chasis rigidity is an indicator, for me, to as how succesfully the chasis will transmit suspension loads to opposite wheels.

secondly, complete chasis test, as you suggest, is relevant of course, because we are also interested in how our suspension geometry will change with respect to wheel loadings, but in such a case, your testing model is not fully correct as road loads can not be applied at spindles directly, because in real life they are applied at tyre contact patch, and then you have a completely new set of forces on your spindle.. (basically you do not have a single upwards or rearward force, but a moment on a spindle so your suspension components are loaded differently...

If you are stressing the suspension then hold the suspension mounts rigidly - or bits of chassis nearby rigidly and you can check out the mounting brackets.

The problem with constraints for modelling the whole chassis is that you are trying to make a dynamic situation into one which is in static equilibrium. I think you may need to resort to inertia relief.

Whatever you do make sure you don't over constrain your model.


[Edited on 24/1/08 by chriscook]