How much margin over B16.5 allowable pressure do you follow in equivalent pressure leakage check ?

Some people follow 3 times including the effect of external forces & moments due to thermal expansion?

regards,

sam

If you are doing a leakage check using the Kellogg equation, then take a step further and use the equivalent pressure in an ASME VIII Div 2 flange calculation.

Thank You, Sir.

Yes, I am doing a leakage check using the Kellogg equation.

Is there any way to consider this computed equivalent pressure Pe = P + 4F/PI*G^2 +16M/PI*G^3 with any allowable in an ASME VIII Div 2 flange calculation?

I understand that in ASME Sec VIII Div1 Appendix-S we have flange rigidity J which has to be kept below 1 in gasket seating & operating condition both to avoid leakage. This is done in Caesar-II flange leakage check module.

But, my question is still not answered.

regards,
sam

'How much margin over B16.5 allowable pressure do you follow in equivalent pressure leakage check ?'

Sam,

ASME SEC VIII Div 1 APP 2 eqn can be used with pressure replaced by equivalent pressure, something which CAESAR II does.

Regarding your question, I don't think there can be any answer , simply because( to my knowledge) anyone has done any kind of correlation. The Equivalent pressure is an way to keep the total axial stress ( membrane+flexural effects) on a flange to a very conservative value. However, this may not have any relevance w.r.t leakage as leakage is a function of many parameters where if imposed loading is less is defintely good but may not be good enough ( the role of gasket is not addressed as well as the relative deformation of bolts, gaskets and the ring, the torque sequencing, creep relaxtation of the gaskets etc).Flange leakage analysis is a very fuzzy area and as MoverZ has suggested, when one method fails, it is best to go to the next available method in the market.Even if someone does an FEA analysis there is too much to go in the modelling with too many assumptions.

Regards

Anindya,

I second your opinion with one remark.
The Equivalent Pressure method presumes that "the bolt load changes very little when a moment is applied whereas the gasket loading changes appreciably" (Kellogg book, chapter 3.11).

Best regards.

One paper on this topic is-

BOLTED FLANGE JOINTS UNDER EXTERNAL MOMENTS:
AN ANALYSIS USING THE COMPOUND GASKET APPROACH FOR SPIRAL WOUND GASKETS
PVP2007-26841

Please don't anyone ask me to email them a copy as my refusal may offend. The conclusion (using EN1591-1 method) was that P+Peq=2xPr was probably acceptable, but they did use vey high bolt-up torques, higher than is usual practice so I'm a little sceptical.

Another oft-used approach ASME III NC3658.3 can result in P+Peq = 2 to 6 times the rated pressure (and dynamic load can be much higher again) ..I wonder if engineers working to this code apply the formula blindly or if they do more analysis.

Anindyo,

regarding 'ASME SEC VIII Div 1 APP 2 eqn can be used with pressure replaced by equivalent pressure, something which CAESAR II does.', the same shortcoming of considering stress in kellog's equivalent pressure method remains.

sam

Yes, you are spot on. The same shortcoming remains.That is why, when one method fails, you go to the next available one in the market. Both the Eq. pressure and DIV 1 stress analysis method attempt to control the applied loading on the joint by keeping its resistance i.e. stress within aspecified value. The DIV1 APP2 method is not as difficult to meet as the Eq. pressure method is.

Regards

Sam, / Anindya,

My reference was to ASME VIII Div 2, not Div 1 App. 2. The 2007 edition of Div. 2 has a new flange design section (4.16) which allows for direct consideration of external loads without recourse to the Kellogg approach, although it is similar. I understand that there is long term work going on at present to replace the Div. 1 method with another, new routine.

Personally I think the EN1591 method has the capability of being the best approach, since it considers the entire elastic system under pressure and external loads. The authors will first need to join the rest of us in the real world, and make the Code rules intelligible.

And Martin (MPB), I fully agree that taking the ASME III one-liner out of context is thoroughly dangerous.

Sam,

Kellogg method transforms the external loads as reaction on gasket and defines the effect as done by an equivalent pressure.
For tensile force this "transform" procedure is easy: a tensile F force has the same effect on gasket as a pressure:
P1=F/(PI/4*G^2)=4*F/(PI* G^2).
For a bending moment, the peak effect on gasket is a force of 4M/G magnitude and that force has the same effect on gasket reaction as a pressure:
P2=(4M/G)/(PI/4*G^2)=16*M/(PI* G^3).

So the EQP is:
operating pressure+ 4*F/(PI* G^2)+ 16*M/(PI* G^3), under the assumption the external F and M act on gasket, while bolts are not participating.
The method has noting with flange stress, in my understanding.


The method is conservative for two reasons. First, the external loads effect is not only on gasket. Second, the limit for total pressure is assumed to be as those in rating pressure- however we don’t know how much is the stress gasket, when flange is loaded by rating pressure. Simple questions without simple answers: could gasket carry on more than rating pressure? Increasing the pressure, the gasket load is reduced- which is the limit? and where is counted the initial gasket stress- after "bolt-up" procedure?

Sec III Div NC subpara 3658.3 has a complete different assumption: external loads are transferred as bolt loads, while the gasket is not counted as participating. The bolts are assumed to react to a bending moment M load. Bolts are not equally loaded, and the peak "local" bolt load is 4M/C.
Now this load can be counted as giving the bolt stress 4M/(C*Ab)< S and that leads to M<(S/4)*C*Ab.
It is supposed that 12.5 ksi- one half of the allowable bolt stress 25ksi- is available to compensate the external load…
So basically, the method has nothing common with flange stress and gasket reaction. The bolts being generously dimensioned, the method concludes that the external bending moment can be high.

ASME VIII is the Taylor- Forge method counting flange stress. As an improvement of the method, the external loads have been considered in an equivalent pressure- say Kellogg formula. However, a deviation has been applied and, as per NC- 3658.1(b), the equivalent pressure is used only to compute the Hydrostatic End Load H while the operating pressure is used in the remaining calculations.

ASME VIII Div 2 has introduced directly the effect of external loads. However, the "root" of the method is still Taylor Forge procedure which is a little bit strange…. It assumes "seating" requirements and "design pressure" requirements. But following blind method math, the results for design pressure near zero and "seating" (with pressure zero) are complete different! In my opinion this shows that we are incorrect applying the formulas when apply them for checking stress flange…
It is interesting that the Taylor Forge method can be re-written counting an unique bolt load for both "bolt-up" and "design pressure", which –BTW- it was the original intention of Taylor-Forge bulletin… this gives more realistic results (has got results as total pressure would be 1.8...2.4 rating pressure), but this interpretation is not a formal one, and, in addition the results are dependent on this "bolt-up" load that is somehow arbitrary- exactly how it is in field!

About EN 1591. For sure it is an accurate method, hard to be followed "by hand". The only doubt I have is the fact "Mechanical Model details" paragraph which is saying "Only circumferential stresses and strains in the ring are considered. Radial and axial stresses and strains are neglected".
Have no idea PVP2007-26841 deals with this aspect when apply method for ASME Flanges.

Regards,

Codeti allows you to use a margin of 1.5 times table rated pressure for extreme occasional conditions like extreme wind and earthquake.
I have attached the page for refrence in other post.

In my opinion, for a "stress in flange" method, it makes sense to consider a coefficient as 1.5 for occasional stress.
But for load on gasket, this is very questionable unless you know very well that the gasket load under pressure rating is not near "leakage"! Do you? And do you know at 1.5*p_rating the gasket is ok?
So I would imagine a gasket can survive in occasional load with EQP=1.5*P_rating if is working satisfactorily in regular service at 1.5*P_rating as well … Just my opinion, of course!

Sam,

Just one opinion

If your code is B31.3, you can do as follows:

1. Check equivalent pressure against B16.5 allowable pressure.

if fail go to step 2

2. Perform a ASME SEC VIII Div 1 APP 2 check with 1 x B31.3 allowable (as this is specified by B31.3).

if fail go to step 3

3. FEA



If your code is B31.1, this one instruct you how to use ASME SEC VIII Div 1 APP 2.


The reason you can not use any increasing factor is that the codes are silent in this mater.

In any case try not to mix the codes.

Best regards

Dan,
One more "lost" method is Blick theory; I would place it next to EQP.
It is based on the same assumption as the EQP theory (i.e. assuming that the gasket is carrying-on the external loads), just the boundary condition has been taken as following.

The gasket reaction
HG = W -p*PI/4*G^2 -4*M/G
must be greater than
2*b*PI*G*m*p
(which is the condition to have at least "mp" stress on gasket, the same as considered by Taylor Forge- ASME VIII)

This condition leads to:
M < G/4*W -PI/16* p*G^3 -PI/2*b* G^2*m*p
or
M < G/4*Sop*Ab-PI/2* p*G^2*(G/8+m*b)
with Sop "operating" bolt stress.

Many quotations of this formula replace Sop with Sb.
In fact Sop (I’ve maintained the original notation from article!) is the central point of the problem, not only here, but in every method (including FEA, where the user must provide a bolt initial stress, isn’t it?)

Just an additional remark- in fact an answer to the original question.
If you think Blick theory is an acceptable one (and I think it is, why not!) the same condition can be forced to give a "margin" for total pressure under EQP theory.
So under this approach the margin isn’t p_rating but it’s a limit that depends also on "p"- the operating pressure.

You can find it following this way:
W -p*PI/4*G^2 -4*M/G -F > 2*b*PI*G*m*p
or
W- p_total*PI/4*G^2 > 2*b*PI*G*m*p
[with p_total= p+ 4*F/(PI* G^2)+ 16*M/(PI* G^3)]
or
Sop*Ab - p_total*PI/4*G^2> 2*b*PI*G*m*p
etc….

Dear Mr. Richard Ay,

In my opinion, it would be interesting to have also an "on-line" checker following Blick method.

I think Blick method is anyway more conservative/ realistic than NC 3658.3- which is just an evaluation of bolts capacity to be loaded "reasonably" with the external loads. Blick method is more realistic than EQP in that the limit of gasket load is expressed as the requirement to have minimum "mp" stress on gasket under p and M loads, which is a criterion valid also under our days ASME approach.

Blick theory has been 1950 published in three numbers (parts) of Petroleum Refiner/ Gulf Publishing Company Publication.
If you are interested, I can send you the articles.

My best regards.

Mariog,

If you have these in electronic form you can e-mail them to techsupport@coade.com, and we'll take a look at this.

For a bending moment, the peak effect on gasket is a force of 4M/G magnitude
-----------------------
mariog,

can you please explaing in terms of basic mechanics, how this 4M/G derived ?

thanking you in anticiapation.

Let's consider the case of a moment M applied on a circular plate. The reaction on G diameter is 4M/G/(PI*G)*cos(theta) where theta is a polar-angular variable.
On theta=0 the peak reaction is 4M/G/(PI*G), on opposite side we say it is -4M/G/(PI*G).
We can apply these local distributed forces (depending on theta) on a "infinitesimal pie slice" belong to the circular plate and we must write the equilibrium conditions with distributed forces applied on infinitesimal arch lengths.

We can say the forces (that must be in equilibrium with M) appear as maximum -4M/G on tensile loaded part and 4M/G on compressed part (signs depends on reference system), but we have a continuous variation of the gasket reaction over the circumference.
So such forces can be written conventionally as "global" while they are local and dependent on theta defined direction- this is the ASME convention.

According to this convention, gasket load on "tensile" loaded part of flange –as circular plate- is:
HG = W -p*PI/4*G^2 -4*M/G-F, where W is Bolts Load and F,M are the external loads, under the assumption that gasket is carrying-on the external loads.
ASME Code doesn't provide an equation for gasket reaction, because the above formula is just a simplified tentative to describe the reality. I would consider the best description of this problem has been done by Blick himself: "It should be noted that despite the fact that the code goes to great lengths to avoid defining the gasket load, the implication is there". True, and I would add that one basic aspect is the gasket load is decreasing with pressure (p) and external loads (M and F) for most of practical applications and this is the main reason for which the leakage appears!