Suspension kinematics detail design challenges

[747heavy]

Marcus, with all due respect, are we not mixing some things up here?

we have come from, being worried about carbon dust ingresssion, to high pressure water/steam ingression and corrosion damage.
I´m not quite sure if it relates to the suspension rocker in an F1/race car.

Don´t get me wrong, I´m not totally opposing your view, but I feel we have a bit of
context drift here, as far as the application goes.

So, if needle bearings are at large unsuited for a rocker application, what do we make of there se in engine rockers (valve train) - all incompetent engineers @ work?

[riff_raff]

747heavy,

Engine valvetrain rockers mostly don't use rolling elements for the pivots, only for the roller followers which have a constant rotary motion, not an oscillatory motion.

To understand why a plain bearing pivot can perform as well as a needle bearing pivot, you need to consider that the plain bearing can have a much smaller journal diameter than the needle bearing for an equivalent load capacity. So the moment radius that the friction occurs at is much smaller, and thus the friction torque can be similar even with a greater Mu. With high oscillating frequencies in the rocker, and the relatively high polar inertia in a steel rolling element bearing complement and retainer, the rollers can experience lots of skidding/sliding at the points of rotation reversal.

As for teflon-lined spherical bearings on suspension arm pivots, their frictions can vary widely between new and used. If you've ever tried to rotate one of these teflon-lined spherical bearings when new, they are very tight and stiff. After a few laps, their clearances open up and their friction drops appreciably. One single spherical bearing would not be an issue, but with 8 a-arms, two toe links, two steering tie rods, 4 pushrods, 4 dampeners, and 2 sets of roll bars linkages, the total number of sphericals is probably close to 50. That's potentially lots of friction.

Regards,
riff_raff

[marcush.]

thanks riff, I was reluctant to answer on that one but tbh the little engine devlopment I did revealed that needle bearings are not the best of ideas in valvetrains.I undersrtand that even bad ideas can work with considerable invest and
expensive development( for example having the engine behind your rear axle) but usually it pays of to get the basics correct first and look for the exotic materials
and finer points in engineering to optimise and not to tweak a bad idea into something workable..I´m too cheeky her considering my knowledge and I´m aware of it.

[mep]

Friction doesn't hurt, so long as it is consistent & accounted for in the set-up. The problem with plain bearings, in my experience, is that friction varies with load, & the rockers in an open wheeled vehicle tend to be highly loaded. I have seen a vehicle with plain bearing rockers that really didn't require dampers....

As a foot to earth, I have just looked up the total friction (links, rocker, damper, etc.) in the suspension of an F3 vehicle (no flexures), selected more-or-less at random. It was around 120 N at the front axle, & 150 N at the rear axle (both per side). I would expect the damper contribution to be no more than 1/3 of the total (although that can vary with charge pressure & side load/bending moment).

I think Dave has a good point here. We should take a closer look at friction.
In the end its the defining factor for performance.

Some figures I found:

plain bearings mu: 0,1 - 0,05
ball bearing mu: 0,0015
needle bearing mu: 0,0025

friction moment
M=F*mu*0,5*d

So when the mu values are valid for our application the ball bearings give us in fact lower friction.
Some argue the diameter could be smaller on plain bearings. In principle that’s true but we are speaking about rather small loads here for a bearing so the difference is not that big. At least not up to 100 times like the difference between the friction values.

I think it’s a matter of wheel load or better bearing load something like a scale effect.
For the very small 1:4 scale car its best to use plain bearing. A FSAE car with its low 200kg and no downforce might be in the border area with all solutions possible.
For highest loads, much downforce but low car mass it should be better to have bearings with lower friction with taking the disadvantage of low duration of life.

[747heavy]

Hi Terry,

Thanks for your reply. You have a point, but I think, it´s a bit more complex then just that. And there are cases where the shaft diameter is not only defined by bearing load capacity.

I don´t want to take the thread off it´s aim, so I will keep it short.
I was not talking about cam follower rollers. I was talking about rocker bearings. Now, why it was wrong to say ,almost any engine uses them, there use is not that uncommon either.
A GM LS7 engine or a BMW motorbike engine come to mind, you will also find many aftermarket kit´s for US big/small block V8´s containing this type of bearing (for better or worse), and the fact that larger OEM suppliers using this in there products, makes me think, that it is at least not gross engineering oversight to use them.



There are surely advantages for plain bearings, especially when they can be used in an hydrodynamic application, such as 4-stroke ICE which provides a clean and pressuriezed oil supply.
If this is not the case, I think meedle bearings are quite commonly used in 2-stroke engines as piston pin bearing and in tripods and yoke joints for drive train applications.
Here is an excerped out of "Advanced Automotive Technology" in regards to the general matter.


So at least it seams to be still open to discussion.

If you and Marcus want to draw the conclusion that a plain bush (DU bush) is the only or engineering wise superior solution for an suspension rocker, then that´s fine with me, but I won´t agree with that.
And that´s where we started out.
Before we discount an needle or roller bearing off hand as ignorant engineering, we may need to consider it a little bit more - IMHO.
I agree that contermination inside an roller/needle/ball bearing is not beneficial, and will harm the performance and shorten the life span.
But, I´m reasonable confident, that these problems can be overcome successfully, and don´post a too great challenge.
If the load is applied in radial direction, with only a small side load component, I don´t find much wrong with the use of a needle bearing. As you will find in many motorbike rear suspension rockers for damper actuation, even on off- road bikes.

If the sideload component is high, or we have a combined load vector axial and radial, other bearings, such as tapered roller or ball bearings, are a better choice - IMHO.
I would not have thought, that a plain bearing is a good choice under this conditions.
But I´m open to learn and will accept a good explanation, but contermination with carbon dust and possible water ingression and corrosion, did not convince me of it.
Sorry

Finally, to contribute something meaningful to the thread, these are types of spherical bearings which are used in push/pull rods and dampers to minimize friction, especially under loaded conditions.
They are only used, where the main load is transmitted radial to the bearing, as in a push rod or a damper.
I don´t know, if these bearings are used in F1 pushrods, they are used in sports/GT and Touring cars, if the teams can afford them.
I´m sure Terry has come across them in his field of work as well.




As Dave pointed out, friction per se, is not the problem, as long as it is constant and does not change with conditions. Here lies one of the draw backs of plain bearings and rod-ends/ball joints. There friction varies greatly with load and load direction, as well as with temperature etc.

Finally, I would like to thank Riff-Raff,Marcus,Dave,Belatti,JET,mep,pup,Flynfrog and many others for a nice time here on this forum, and for interesting and stimulating discussions.
Due, to some recent developments (not related to this thread)I wont be able to participate longer on here.
I hope you have a nice time & take care
Ciao
747heavy

[WilO]

747,

thank you for sharing your knowledge; I enjoyed learning from you...

Ciao.

[DaveW]

Bon voyage, 747, & thanks.

Regards.

[thisisatest]

full suspension mountain bikes use a variety of bearing types in rear suspension systems: plain bearings, radial bearings, double row bearings, angular contact, etc.
my experience with those has shown that, when the spring/damper is removed from the system, the force required to move a system of plain bearings is much, much higher than a system with rolling elements at the pivots. even if the plain bearings are completely worn. A bike with cartridge bearings can move very freely and smoothly, only to find out that, when removed, one or more of the bearings is so damaged that it cannot be turned by hand! This has led many companies to switch to cartridge bearings at the pivots, even if they are not "ideal" from a purely engineering perspective.

[Carlos]

747heavy - I've very much enjoyed your participation these last few months. You have helped elevate the forum to a technical Renaissance.

[riff_raff]

747heavy,

If you're still around, look closely at that rocker arm picture you posted. If you analyze the friction forces at the three load points of that rocker, you should conclude that there is not much benefit with the needle bearing fulcrum. While it has about 1.6x the force applied that the pushrod plain spherical end has, its friction loss is likely much higher. This because the friction on the pushrod spherical end is applied at a moment radius of about .16 inch, while the friction on the needle bearing fulcrum is applied at a moment radius of about .40 inch. Since both interfaces operate with oscillatory motion, both interfaces are subject to the same boundary contact frictions at the point of reversal.

As for spherical bearings in a-arms, DaveW points out the problem. A brand new teflon-lined spherical bearing can have a friction 3x or 4x higher than one that has been run-in. If that friction change amounts to say 50 in-lbs per bearing, multiplied by 30 bearings, that's a difference of 1500 in-lbs in suspension friction that can occur over a few laps.

So long and good luck.
riff_raff

[DaveW]

Agreed riff_raff. However, I was trying to point out that rocker post side loads can be very high. Mep's introductory diagram (post 1) shows that the radius arm of the push rod is very much smaller (by a factor of 3, perhaps?) than that of the spring/damper. If the overall motion ratio is unity, then the implication is that the rocker will react a side load that is >3 times the wheel load. That factor can rise quickly with spring deflection if the geometry has a rising rate. I think the bending stiffness of the rocker post & its supporting structure is often a major contributor to the overall "installation" stiffness of a suspension. This also has an implication for friction if the rocker bearing slides, rather than rolls.

[mep]

Strange I don’t really get the point you are arguing about.

Agreed riff_raff. However, I was trying to point out that rocker post side loads can be very high. Mep's introductory diagram (post 1) shows that the radius arm of the push rod is very much smaller (by a factor of 3, perhaps?) than that of the spring/damper. If the overall motion ratio is unity, then the implication is that the rocker will react a side load that is >3 times the wheel load. That factor can rise quickly with spring deflection if the geometry has a rising rate.

Yes its true the radial force on the bearing will be 3 times or more higher than the wheel force but that is a geometrical constraint which has nothing to do with the type of bearing you use. The radial bearing force is equal for all types. This should even speak against plain bearings as you said yourself in an earlier post:

Friction doesn't hurt, so long as it is consistent & accounted for in the set-up. The problem with plain bearings, in my experience, is that friction varies with load, & the rockers in an open wheeled vehicle tend to be highly loaded. I have seen a vehicle with plain bearing rockers that really didn't require dampers....

This is backed up by the friction figures and the formula to calculate the moment I posted:

plain bearings mu: 0,1 - 0,05
ball bearing mu: 0,0015
needle bearing mu: 0,0025

friction moment
M=F*mu*0,5*d

The only things that are variable here are the diameter (d) and the friction coefficient (mu).
The friction coeffs are seen above (so far nobody questioned them) so let’s see what effect the diameter has (as this got mentioned by riff raff at least once).

Ok we need to know the force now. Let’s say we talk about a high downforce car with a max downforce of 30 000N. That is 7 500N per wheel. Let’s say the pushrod has an angle of 25° by this we can calculate the pushrod force.

Fpush=7 500N/sin 25°= 17 746N

Let’s further say the rocker has a motion ratio of 1:2 by this we can calculate the bearing force.

F=sqr(Fpush²+(Fpush/2)²)=19 840N

This gives us in fact a force ratio of 1:2,6 what is very close to the approximation of Dave.

Now let’s find bearings that can take that force either static or dynamic. We have something in between on the other side we used the maximum downforce in the above equation which we will only reach on straights. I use data of SKF for this.



Single row ball bearing:
The smallest I found that comes close to our forces is 6403 which can take 22,9kN dynamic and 10,8 static but its dimensions are d=17mm and D=62mm . Seems to be to big so I think we can forget about ball bearings and it can’t even take the static load.



Single row cylindrical bearings:
Smallest I could find is 2203 which can take 23,8kN dynamic and 21,6kN static. The dimensions are d=17mm and D=40mm. So its smaller than the ball bearing and can take our load. The difference between dynamic and static load capacity is much smaller than on a ball bearing.
Bigger cylinder contact diameter D1=32,4mm.

Friction moment:
M=F*mu*0,5*d
M=19 840 * 0,0025 * 0,5*0,0324m = 0,8Nm



Needle bearings:
I don’t have data yet, maybe later. Anyway it should be similar than the cylindrical bearings above.



Plain bushing:
They define a maximum contact pressure of around 210N/mm².

p=F/d*l

Where F is the load
d the diameter
l the length
p the surface pressure

When I put in the bearing force I get a required area of 94,5mm² or let’s simply say the bearing needs to have a diameter and a length of 10mm.

friction moment:
M=F*mu*0,5*d
M=19 840N*0,1*0,5*0,01m = 9,92Nm !!!

As you can see even when you take the higher diameter into account the plain bushing has almost 10 times higher friction than a roller bearing.

How does this look for a 1:4 scale car?

We want to look at the other extreme now. They weight 10kg and have no downforce. That’s 25N per wheel and (I use the above calculated ratio of 1:2,64 now) 66N on the bearing.

Other than on the high forces its possible to use ball bearings here. You can’t even find a ball bearing that is small enough for this low force. On the same time you must care that the pin doesn’t get to small. I think it should not go under 4mm diameter. Looking at the forces bearings of that size can carry you realize they are magnitudes above those 66N. When you want to use plain bearings you should still stay above 4 or 5mm pin diameter.

Somewhere between those two examples are borders (FSAE?) where it doesn’t make sense to use ball bearings anymore and you have to switch on cylindrical bearings (point contact vs linear contact). That’s what I called scale factor in an earlier post. Regarding the load it’s always possible to use plain bushings but you have to pay for it with much higher friction. On a 1:4 scale car the friction might be low enough compared to the dampers so that plain bushings make sense but maybe those cars are also over damped?

[DaveW]

Nice analysis mep, & I am happy to concede that bearing choices at full size don't necessarily scale well...

[PlatinumZealot]

I know that you use needle bearings if you have a very small space with a very high load and low misalignment. I think the choice to use needle bearings centres on these requirements more than anything else (small space, high load, low misalignment).

If it fits rights around your shaft(lets say you cannot change the shaft dia) and it can take the required load, the corresponding ball will be a bigger radius. And it might be case where you cannot go any bigger..



I really think it is up to the designer and the scenario. I wouldn't trust it on a component that is likely to miss-align even a small amount though.

[riff_raff]

mep,

The coeff. for breakaway in a side by side needle bearing w/ grease lube is around .015 (per "Concepts of Bearing Technology", by Tedric Harris). And the static Mu for a teflon lined plain bearing is around .03 (see pg.33 here http://www.delta-elkon.co.il/_Uploads/1 ... nGuide.pdf )

With teflon liner materials, the best Mu is achieved at unit loads of about 10ksi, and can can be safely used up to 20ksi. To get best Mu (10ksi unit pressure) a 19,840N load would require less than .50 sq in projected area (LxD).

The rocker pivot bearing journal minimum size will be established by post bending loads, and not by the bearing itself. So both plain and rolling element bearing would have the same minimum journal size. However, the narrower journal width possible with the plain bearing will reduce the bending moment applied to the cantilevered post.

You are correct that with the example of a cantilevered suspension rocker, a plain bearing will likely have higher peak frictions than a needle roller by at least 2x (but not 10x). But the plain bearing will likely give a stiffer and lighter installation. So which do you prefer?

Regards,
riff_raff