Gasket Seating Force

[engpes]

I am trying to determine the gasket seating force for a 3" 150# RF flange with a 316 SS spiral wound flex gasket (y=10,000 & m=3.0). The design pressure is 285 psig @ 100F (max for 3" 150# flange).

I know that ASME B16.5 says that this flange and gasket are good for this pressure, but for other reasons I need to determine the gasket seat force (Wm2) per ASME VIII, div 1 / ASME B16.5. I also know how appendix II calculates the gasket seating force, but this sometimes requires larger bolts than ASME B16.5 requires.

My question is:

1.) How can I determine the gasket seating force for an ASME B16.5 3" 150# RF Flange? Is there a standard chart for this?
2.) Is there an equation similar to ASME VIII, Div 2 appendix 2 that is used for standard flanges?

Thank you in advance.

 

[SnTMan]

engpes, actual applied seating force is P/A.

P = Bolt load = Bolt area * applied bolt stress (not necessarily allowable stress per Sec II, Part D, may be higher)

A = gasket gross area Pi/4 8 (OD^2 - ID^2). NOT using N, b, bo, etc, etc from Appendix 2.

Somewhat complicated if a centering ring is present, usually ignored as the winding is thickner that the ring.

Now if you want Wm2, just calculate Wm2 per Appenidx 2. Wm2 doesn't care what your flange is.

Regards,

Mike

 

[engpes]

Mike,
If I used P/A this will give me a pressure (lb/in^2), wont it? I am looking for the net force applied to the gasket.

Also, if I calculate Wm2 for a 3" 150# Flange from appendix 2 I get the following:

Wm2 = 3.14 x b x G x y
Wm2 = 3.14 x 0.1875 x 4.375 x 10,000
Wm2 = 25,758 lbs

If I use Wm2 / Sa to calculate the bolt area required(based on IID 25 ksi allowable) I get 1.0303 in^2. This requires (4) 0.75" studs when a 3" 150# flange only has 0.625" studs.

I am trying to figure out the bases for a 3" 150# RF flange only having (4) 0.625" studs, and am inevitably looking for the seating force.

Thanks and sorry if I am overcomplicating this.

 

[SnTMan]

engpes, sorry, of course P/A is STRESS. Beating my brains out today on a god-awful series of Part UHX calcs

The net force applied to the gasket is the bolt load unless something is going somewhere.

B16.5 flanges and Appendix 2 flanges are much different animals, you will rarely be happy running Appendix 2 calcs on B16.5 flanges. (Hint - up the bolt stress)

Much discussion has occured in the past, do a site search.

Regards,

Mike

 

[racookpe1978]

What does the manufacturer of your specific spiral wound gasket require for this gasket in this size flange?

You MUST follow that specific requirement for torque sequence, torque amounts, and bolt diameter and length. I don't care what you might calculate, only the gasket specifications matter.

 

[XL83NL]

Refer to ASME PCC-1 for more details on these topics. Reviewing that post construction code may take away a lot of questions.

 

[engpes]

Thank you SnTMan and racookpe1978! That is exactly what I needed to know and I verified through a few other sources.

One more question. Being that the gasket alone will govern the required seating force, will the required bolt tension/torque values be the same for a A182 Gr. F316 stainless flange?

This will use SA 193-B8 studs (ASME B16.5 derates these flanges to 175 psi). Does this mean that SA 193-B8 studs are adequate to apply the same torque value as carbon flanges? Will they be adequate to apply the same gasket load (25,758 lbs) per ASME B16.5?

Thank you!

 

[SnTMan]

engpes, w/ regard to gasket seating force, flange material will rarely matter. However you must keep an eye on bolting yield strength, as many of the stainless bolting materials have a low yield compared to carbon steel, etc.

No use planning on a 45K assembly stress if bolting yield is 30K.

Regards,

Mike

 

[engpes]

Yes I know that they have much lower yields, but I am trying to establish an allowable. Appendix II does not seem to be an option because the studs will never pass (especially with B8 lower yields).

I am basically trying to determine the stud and gasket materials to use for a SA182-F316 150# RF flange? ASME B16.5 says that SA193-B8 Cl.2 (strain hardened) can "be used with all listed gasket materials, provided it has been verified that a sealed joint can be maintained under rated working pressure and temperature."

I am trying to use a RF spiral wound flex gasket (316 SS, m=3.0, y=10,000 psi). What guideline can I use to determine that the bolt stresses are acceptable to make the required seal? I know that in this specific case IID lists allowable stresses the same as SA193-B7, but if it is not?

Thanks!

 

[SnTMan]

engpes, 10K is a good seating stress. Calculate gasket winding area, forget "b", just use pi/4 * (OD^2 - ID^2). Now you have a required bolt load. You can use an assembly bolt stress of maybe 70% or 80% (or perhaps more, but see below) of yield at ambient.

You should consider thermal expansion effects at design / operating temperature, to insure that bolting is not stressed beyond yield at design / operating temperatures.

The more pre-load you can establish in the joint, the less effect external forces have on it. With spiral wound gasket and centering ring you do not have to worry about crushing or over-compressing the gasket.

B7 bolting is often used with stainless flanges and can be a good choice if the external service environment permits.

Regards,

Mike

 

[XL83NL]

Dont go to high on preload though, just my 2 pennies. Ive seen a customer spec requiring at least somwhere around 70% the bolt yield at ambient. This resulted in a lot of cupping of SPWND gaskets, simply becuase the customer too much preload. Itd say 50-60% may be a better number ..

 

[SnTMan]

XL83NL brings up a point. Whatever preload is selected all the members of the joint must be adequate for it, hopefully with some margin.

Regards,

Mike

 

[engpes]

Thanks guys. Do you know of any RF spiral wound flex gasket chart that lists the max allowable gasket force before crushing will occur? Force at which gasket will fail and leak?

 

[gasketguru]

The 3"-150 (and 8-150) is one of the the worst case flanges in terms of bolt load vs. gasket area - thankfully the max system pressure is only low. However the traditional ASME VIII / Taylor-Forge calculation will under-cook the loads for you. A limit analysis calculation such as EN1591 tends to give a more realistic target value as it uses real load-leak data. Both load-compression and load-leak tend to give exponential decay curves. The load you want (need) might depend on whether you are sealing water or toxic gas, and what you can achieve will depend on the flange and bolt material grades. Assuming you have A105 or stronger with B7 bolts and standard pipe schedule then anything up to about 9 tonnes is possible, but torque tightening is not accurate enough to allow you to go to that level reliably.

Most gasket load charts will be a compromise - for this particular size joint I usually suggest at least 4.5 tonnes / 10,000 lbs per bolt. This might not get "full" compression to the guide ring in every case, but the sealing element stress should be enough for decent tightness against only 20 bar system pressure. You can get quite a spread of variance in torque vs. bolt tension as friction varies a lot - thus most people add a bit to aim a little to the high side. Thus increasing this to say 5.5 tonnes per bolt (12,000 lbs = 57,900 psi on the root of the bolt) would give better compression and some allowance against the inaccuracies of torque - again assuming materials are strong enough.

Some manufacturers try to make soft gaskets for class 150 so they compress easier, but you don't get something for nothing - you need densification of the sealing element to get gas tightness. Use gaskets which have an inner ring as well as outer ring - those without inner support can allow radial spread inwards and loss of tightness - the price difference is only pennies usually.