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Rigid coupling design - friction of bolted shaft flanges 3

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crazyjpeters

crazyjpeters

Mechanical
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I cannot seem to find an answer anywhere, so I'm hoping some bright individual will be able to enlighten me.

I'm analysing a large machine shaft coupling, which is a bolted flange design, clearance bolt holes (not fitted bolts), where torque is to be transmitted through friction of the faces. This is augmented through a friction modifier (more on this later).

Can anyone point me to a formula which calculates the torque capacity based on number of bolts, bolt force, type of coupling material (static friction factor), etc.

The only meaningful formula I can find is supplied by Loctite, but I'm not entirely sure it is correct. It is stated as follows:
Torque capacity = Mechanical contribution + Adhesive contribution, with details on their website at:

I know I've seen a similar formula without the adhesive contribution, but can't seem to locate it.

I'm up to my eyeballs in formulas and handbooks that seem to ignore friction of bolted joints as a torque transmission tool. Apparently fitted bolts are the only way to do it, so my machine is destined to end up failing.... :-(
 
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well a standard formula is torque = Preload*diameter/5.

btw, what's to stop your flanges from slipping (slightly) untill a pair of bolts (at least) bear up and transmit shear ?
 
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Can you c'bore the flanges (diameter a bit less than the bolt circle diameter) so that the torque is carried by a larger radius? Then just use:

torque = preload x (radius of c'bore) x (coef. of friction)
 
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maybe we're confusing "torque" ... in my equation it is the bolt installation torque (whereas brian is using torque as the torque reacted by bolt shear)
 
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Friction between flanges is the "only" shaft torque transmission. That's the only thing stopping coupling slippage before it encounters the shear of a bolt. The point is never encountering a bolt in shear.

I'm looking for something with a solid theoretical basis, or an empiracle formula that can account for the pressure from the bolting and what fraction of the theoretical flange friction could be achieved in reality. It's obviously not going to be 100% perfect contact with even pressure. There are 12 bolts on this flange, and it is spigoted, so there is an inner and outer radius to the flange contact surface.

The loctite formula could model this, I suppose, but I'm not sure I believe it correctly models it when I remove the adhesive portion of the calculation.
 
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yes, so you can determine the bolt preload (from the bolt installation torque) = Fn; and then the friction force (from metal-on-metal or maybe metal-on-fay sealant) = Ff = Fn*CF; and then the torque (= Ff*PCD).
 
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Here's a design that handles over 500 lb-ft of abusively applied torque
without a whimper.


ARP used to have a tech paper on their website claiming automotive flywheel bolts are subjected to shear at the flywheel crank interface, even though American flywheel bolts were not "fitted" for most of the modern era. I do not believe that is correct. And when I went to their website just now, I could not find that statement, so they may have changed their mind or perhaps their tech writer.

Six 7/16 inch bolts torqued to 60 lb-ft or so. Bolt circle 3.58" diameter.
(8000 lb est clamp load)(1.812" )( 6 bolts )(.2 guess u)( 1/12) = 14,500 LB-FT
hmmm.....

here's something for later, when the discussion changes to the utter unsuitability of torque as a means of establishing preload. About half way down the page here.
 
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crazyjpeters,

It is quite common to use friction in a fastener clamped interface to transmit shaft torque. The reason being that if more than two fasteners are needed, they cannot guarantee load sharing in shear unless they have interference fits.

The simple moment capacity of the joint is the fastener preload applied at the pitch radius times a coefficient of friction (assuming no misalignment or bending). It is important to design for sufficient joint friction such that there is never a chance of slippage. Otherwise the joint will suffer fretting.

The loctite compounds noted were correctly described as "friction modifiers" and not adhesives, since that is how they function. The loctite compound improves the torque capacity of the joint by enhancing the friction at the interface, and not by contributing adhesive bond shear strength.

Terry
 
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Hi Tmoose, I'm actually investigating this for a coupling theoretically subjected to what I think translates to 1.8million ft-lb of torque. I'm also well aware at how poor a torque wrech is at applying a set preload (and how little I can trust the stamped 8.8 grade chinese made bolts at this plant).

We actually are using bolt heaters to establish correct preload (bolt length measured pre-heat and post-cooling), so I have some degree of confidence in the preload values.

Tbuelna, I've also seen that fitted bolts are not always fitted, especially since the lengthwise deformation causes some reduction in cross-section. I believe there are often sleeve-bolt combinations that accomplish the tight fit, but don't rely on the bolt-shaft to transmit that shear.

What the OEM (in this case), had provided was a method to apply a sprayed on abrasive coating (boron-carbide in a lacquer spray-paint carrier) to act as the friction modifier. I have seen gaskets for wind-turbine hubs that are impregnated with boron-carbide, but I honestly don't know what value this would carry in terms of "static shear strength" according to the loctite formula.

I'm a bit uncomfortable with the spray-paint idea, and one of the loctite adhesives may be a recommendation I make. The exact one may not exist, though, as I need a temporary resistance to applied heat (bolt heaters), but would like the adhesive to dissolve/relax with heat later if we need to remove it (think loctite over a 1 m2 area. No fun!
 
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Having something crush-able in a bolted joint sounds like trouble to me, as loss of a few thousandths bolt stretch is likely to sigNIFicantly reduce the preload and all-important clamping force.
Uneven distribution and unevenly sized grit sounds like an invitation to induce misalignment, too.

Loctite to fortify an interference/clamped metal-to-metal fit - In the semi modern era Harley factory recommendation for crankshaft assembly was to use high temp, high strength, slow curing loctite (620 ?)on the tapered joints used to lock the mainshafts and crankpin to the flywheels. Previously nut torque alone was used to lock the tapered shafts to the flywheels.
General description here on page 4 -
 
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to me a friction only joint is so full of variables/unknowns that the smallest one is preload which'll varying from bolt to bolt, but CoF is one, contact area another, the distribution of contact force, ...

and you've got one awfull lot of torque.

obviously the design works, hope you've got a big SF to absorb all the varibles.

but then you've got the bolts working as a reserve, assuming the joint slips a little ...
 
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Fairly recent reference to recommended Harley Loctite.

===========================

Actually, I think a torque wrench can be used pretty effectively to apply preload. The ARP link showed behaviour of bolted joints under repeated assemblies, and lubrication methods to achieve better repeatability.
For some fastener/joint designs torque is the most practical method by far. It was "good enough" for the Navy for decades. My preference is never to unnecessarily paint myself or the manufacturing folks into a corner, and that would definitely include requiring applying preload surgically. Much of the stuff I've ever done had loading known at best to an order of magnitude. Makes it tuff Justifying efforts to create clamping force to 1 or 2 significant figures. Throw more big bolts at it, torque them hard, and use the time I saved to concentrate on the important but subtle details of the joint and local features that ensure good clamping, and must endure cyclic loading since they are caught outside of the protective zone of preload.
 
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Hi crazyjpeters

I think hydtools link answers your question directly, the link I have posted allows for the bolts to carry the shear from the torque.


I wouldn't recommend relying on the friction between two bolted faces, have you considered using keyways in the couplings?
The torque your transmitting seems very high and I'm wondering how the half coupling is attached to its respected shaft to transmit such a high torque.

desertfox
 
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The design is the OEM's. I'm just trying my best to verify their design, and perhaps suggest improvements to the plant (27-30MW @720rpm, maybe my math is wrong on the torque).

As the coupling is 750mm wide, and the friction coating is just a single pass with a paint gun, maybe a couple of mils, I'm not too concerned with preload loss from the coating. The thought is that the grit embeds into the faces, providing shear strength increase. Pretty difficult to quantify without labratory data.

The studs are not overly long, which is part of my frustration. I would have liked a much longer bolt, which could have been tensioned up into the upper elastic range with a bit of breathing room.

Keyways are exactly what got us to use a larger flanged coupling. Stresses due to poor machining, led to cracking at a shoulder.
 
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750mm is flange diameter and thickness ? What is bolt circle diameter? stud diameter?

30 MW ~ 40,000 HP.
(300,000 lb-ft)(720 rpm)/5252 = 41,000 HP

still looking at ~100,000 lb clamping per bolt with u=0.2
 
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hello crazyj...

Pay attention to Tmoose's methodology, he has it right.

On another note, from a design standpoint, what is the nature of the torque? Is it unidirectional, and caused by some sort of gravity load, or is it reversing, going negative to positive? It makes a huge difference as to whether the assembly can rely on the shear capacity of the fasteners unidirectionly, or if you have to design for everything to be carried by bolt clamp times coefficient of friction in a fatigue application.

As an example, the attachment of counterweight bolts in vibratory screens is a static, unidirectional load, even though the counterweights are spinning. On the other hand, the bolts attaching the bearing housings see reversing loads, even though the bearing housing does not rotate. It makes a big difference on how the connection must be designed.

Another note: depending on the nature of the loads, you may not want paint, foreign objects such as dirt, grease, etc in the joints.

Another question regarding loads...is the joint subject to torsion only, or is there anything else, transverse loads, axial loads, bending loads, and what are there nature, reversing or static?

 
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Hi crazyipeters

I calculate the torque to be 292751 lbsft but we need the bolt cicle radius to get stresses on bolts etc..

desertfox
 
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"The studs are not overly long, which is part of my frustration. I would have liked a much longer bolt, which could have been tensioned up into the upper elastic range with a bit of breathing room."

maybe i'm not thinking this morning, but what's the connection to preload and length ? ... sure there is bolt extension (bigger for a longer bolt) but the same preload can be applied to a short or a long bolt; no?

"studs" ...

we need PCD, and number of bolts to figure out how much torque is reacted by each pair of bolts and so the required preload.

 
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