I had a look at your drawing of the Rotorway system George and I have to say I don't understand at all what they were doing. Since there is still a rigid radial needle bearing at both ends nothing has been gained in terms of
flexibility. All the elastomer does is replace the thrust bearing!?? Maybe I'm missing something but I don't see what else it could do. The main point of the elastomeric bearing is to allow the blade to flex in the radial direction.

Rotorway's is 2.115" OD and 1.0" ID. by .575 thick.

Older hinge designs relied on conventional metal bearings. By basic geometry, this precludes a coincident flapping and lead-lag hinge and is cause for recurring maintenance. Newer rotor systems use elastomeric bearings, arrangements of rubber and steel that can permit motion in two axes. Besides solving some of the above-mentioned kinematic issues, these bearings are usually in compression, can be readily inspected, and eliminate the maintenance associated with metallic bearings.

I am trying to analyse a laminated elastomeric bearing pad for a bridge. This consists of alternate layers of Teflon and steel. Can anyone please recommend a simplified analysis/design procedure?

I assume that the teflon layers are thin and present only for the purpose of limiting the shear force that can be transmitted between plates. As such the plates' composite bending stiffness would be approximately the sum of individual bending stiffnesses (if contact between steel plates is maintained). The equivalent plate could be calculated based on an equivalent single thickness or an increased Youngs modulus (keeping the thickness of a single layer).

The added compliance do to total thickness of steel could be combined with compliance of the foundation. The results, of such an approximation would be progressively less realistic with more uneven loads or supports.

I am surprised you are not looking at Fabrica or some other fiber reinforced rubber to provide the compliance.

The Teflon layers are _not_ thin. They are there to permit a slight rotation and a slight translation between the bridge girder and the pier. They are not simply placed in between the steel plates, but bonded to them. Thus, some shear _can_ be transmitted.

The problem involves very large material nonlinearities, and my aim is to find out whether the steel plates will touch or the Teflon layers will tear away from their bonding, and under what loads.

You haven't mentioned FEA as a design method, however, a simple 2D model of the cross section would allow you to focus your efforts on material property verification and behavior. All three layers would be modeled, the teflon, the bonding and the steel. The contacting body(s) would then be configured to impart the necessary loads with appropriate surface contact stiffness and friction. I've performed similar analyses, and read the results of another 2-3 years ago where a threaded insert was bonded inside of a hole. The entire assembly was modeled for contact, focusing on the shear strength of the bonding.

It sounds like your materials are all isotropic, so this avoids some of the tedium associated with material model verification of anisotropic property application found in many composites. This is a perfect, 2D nonlinear problem... 1-2 days max, given accessible material data for the teflon and bonding! If your material data is in question, then you could spend another 1-2 days testing your material models. But, then when you're finished, you've got a model to use for all kinds of other studies.

I can see one problem with modeling the rotation in 2D, depending upon its plane... but after successful 2D modeling, a 3D model would be in order anyway!

Hope I've interpreted your situation accurately,

I have an old research paper from the American Helicopter Society. It was on design of elastomeric high load feathering bearings. Said best material was latex rubber. Does anyone know of a source of stock bearings we could use in homebuilt (to carry about 30,00 #.) Would like to experiment on my heli.

Elastomeric bearings would make things real easy but one guy who tests them sent me a e-mail and recommended that I not try to make them for aircraft use since there was a lot to them that you couldn't see like heat dissipation. Since nobody wants to sell you them I figure I better stay away. However, if you have information on the exact rubber compound they use than I could make them but if a bearing crumbled and locked up in the rotorhead, it would be fatal.

The patent I saw said you can use any rubber but the metal disks must be strong to prevent tearing since under pressure there must be a lot of stretching forces on the metal plates.

I can make a few elastomeric bearings using washers and polyurethane rubber but how would we know if it was safe? The big problem that worries me is getting my helicopter done and not knowing if it is a death trap. On paper it
could look good but until you wear out your parts you never know for sure how long anything is going to last.

After dissecting my Rotorway elastomeric bearing it looks like all we need to find is suitable glue.

I can buy .004 or thicker latex natural rubber sheets and brass shim stock. I take the shim stock and glue the latex on one surface and punch out zillions of round discs. They can then be stacked in a tube with steel spacers every half-inch to keep the discs from sliding out of position.

Should work unless there is a need to glue the discs together but not sure it would be necessary with a tube on the O.D.

Latex cannot come in contact with oil so the tube will act as a protective housing.

Getting the rubber soft enough is probably going to be the big problem. Most rubbers start at shore 60 and go harder. An O-ring is shore 80 durometer.

I recall reading, in one of the Internet pages, that 2 different resilient materials are used. The harder one is located at the outside diameter of the disks and the softer one is located in the middle. This may be done so that the harder resilient material restricts the "bulging out" of the resilient material when the bearing is subjected to compressive loading. It also would appear, to me, that there might be no reason for the inner resilient material to be bonded to any thing, just that it is not compressible.

I did some calculations and made some assumptions to come up with a basic bearing design. For starters, I assume a bearing that has a 2" OD and a 3/4" ID with a conical profile of 40 deg from flat. Assuming a cent force of 6000 lbs
this gives a compressive stress of 2200 psi in the elastomer in the thrust direction. If you start with ten .032 thick rubber layers separated by .016 brass or steel shims and assume that the rubber will bulge into a semi-spherical shape out the side under load you get .010" of overall compression of the bearing under load. The semi-spherical thing is a total assumption but is the best I have to go on at this point. If you have ten layers of the rubber/brass and you assume the rubber can offset in shear by its own thickness under reasonable force from the pitch horn then you will get 18 deg of rotation in either direction (total 36 deg) from the bearing which will easily cover the approx 24 deg max that you need out of the feather bearings. The overall thickness of this bearing would only be .48".We could be better assured of the bearing stability if we lowered the thickness of each elastomer layer to .016 and used 20 layers which would then lower the overall compression under load to .005". The bearing would be .64" thick overall.
This is for the bearing on the inside of the hub.

The bearing on the outside is a different story all together. If it were conical (pointing the opposite direction)it would have to be prestressed when loaded into the blade grips or it would come loose under centrifugal load and be useless. If it were bonded to the restraining blocks it would be under tension to the same order as the compressive bearing was under compression and would probably tear apart. So I don't think a conical is the way to go on the outer bearing.
I think the outer bearing could be a standard radial needle bearing with one thicker, say .062 layer of perhaps fiber reinforced rubber around it (such as found in reinforced rubber hose) pressed into the block. The rubber layer would allow for movement of the inner conical bearing and would act as the pivot point and so would not move very much.

Another option would be to press the radial needle bearing into a Teflon lined spherical bearing to allow for the rotation. This would be simple but the static friction of the Teflon each time it oscillates may create some feedback, but I think it would be quite small and maybe/hopefully negligible.

A third option would be a radial elastomeric bearing but this would require nesting metal tubes pressed together with nesting layers of rubber and would be more difficult to manufacture. It would be the best option if it could be done reasonably easily.

Elastomeric bearings are bearings and are susceptible to failure although the failure manifests itself in different ways as opposed to a metal bearing. In most cases they are limited to 5000 hours but very few ever reach that limit. Elastomeric bearings are susceptible to Ozone damage unless they are impregnated with wax and they can be destroyed if they come in contact with hydrocarbon material and it is not cleaned off. They are also susceptible to solar radiation in that they are limited to an exposure of 160-degrees F for one hour in their total lifetime. Any thing in excess of this can reduce the life by 50%.