Nozzle Flange Load Calculation
Nozzle Flange Load Calculation
(OP)
Is it common to also examine the stresses on the WN flange attached to the nozzle when calculating the ability of a nozzle to withstand specified nozzle loads?
Thank you
Thank you





RE: Nozzle Flange Load Calculation
Joe Tank
RE: Nozzle Flange Load Calculation
RE: Nozzle Flange Load Calculation
The equation I am thinking of is found in ASME III, I believe.
Regards,
Mike
RE: Nozzle Flange Load Calculation
The external loads on the nozzle flane are quite frequently overlooked, because of the lack of confidence in the calculation method in Div. III, because those calculations are based on vague assumptions and sporadic data provided by the salesmen of the gasket company, because of the vague definition of the flange failure (first the flange will leak, then break off from the nozzle or pipe!) and the difficult task of assessing the real limit of stresses in the flange. However, most of the people, including myself, feel safer calculating the equivalent pressure and including it for the check of flange rating.
Cheers,
gr2vessels
RE: Nozzle Flange Load Calculation
Pf=4F/(pi*G²)
Pm=16M/(pi*G^3)
ref: M.W. Kellogg Co., "Design of Piping Systems" 2nd ed., (New York:John Wiley & Sons), 1956.
the thing is, it doesn't just apply to your nozzle, it should apply to the other flanges in the line too
RE: Nozzle Flange Load Calculation
RE: Nozzle Flange Load Calculation
First off - where does even the Kellogg method say to do that?
Second - B16.5, in article 2.5.2 only says that care must be taken to avoid "severe" external loads, and then, it is only a warning for Class 150 above 400°F and for all the other classes above 750°F. So, why even bother worrying about external loads on flanges?
(Again, just playing devil's advocate here...)
RE: Nozzle Flange Load Calculation
Thanks in advance
RE: Nozzle Flange Load Calculation
ASME Sec VIII, Div 1 really places no restrictions on WHEN you do differential design, just on HOW you do it, mostly in terms of marking, testing, etc.
The reason to do a diff design is to SAVE RESOURCES (somebodys' money).
In your example it might allow a thinner tubesheet or tubes, the only components that see both pressures. You would have to restrict the operation of the exchanger such that the differential is NEVER exceeded, and hydrotest also needs looked at.
I once had a U-tube exchanger that, during the sales stage, the effect of external pressure at high temperature on SS tubes was overlooked such that the min tube wall was not sufficient for the shellside pressure with vacuum applied to the tubeside at the design temp.
You can: 1) Change tube materials - affects cost, nobody likes that.
2) Restrict design temperature of the tubes - no cost but can be messy due to operating conditions.
3) Restrict the presure differential across the tube wall - also no cost but can be messy due to operating conditions.
With the PERMISSION of the client, option 3 was chosen for this job.
On other jobs that had high pressures the customers specified diff design of alloy tubesheets to save those resources.
BTW, ALL PV design is differential pressure design. We just don't call it that when the reference is atmospheric pressure;)
Regards,
Mike
RE: Nozzle Flange Load Calculation
Refer to ASME VIII, Appx 2, clause 2-1(a) "...Proper allowance shall be made if connections are subject to external loads other than external pressure". I think it's self explanatory.
Cheers,
gr2vessels
RE: Nozzle Flange Load Calculation
One one of the vessels it was stated that we could weld the plate back to gather and we would have a good vessel.
Our excursions(?deflagrations?)are extremely rapid, supposedly as fast as the speed of Nitro.
RE: Nozzle Flange Load Calculation
Using the B16.5 chart would be instead on analyzing each ANSI flange.
There were some good articles by Walther Stikvoort in Chemical Engineering Magazine in July1986 & June1994.
RE: Nozzle Flange Load Calculation
If you choose 1*Pdesign, are you not saying that the B16.5 flange is essentially unable to handle ANY external load at the design pressure? From what's in B16.5, I would say that is doubtful. Again - I'm not concerning myself with Sect. VIII, only B16.5 flanges.
Oh - and arto - what's an ANSI flange? I wasn't aware that ANSI had a current flange standard?
RE: Nozzle Flange Load Calculation
To me this issue is very simple;- the total flange capacity to withstand stresses, is the limit stated in the ASME B16.5 (B16.47)(that is not design pressure, is rating; - for the design pressure definition, please review the relevant clause of ASME VIII Div 1, for example). The total compounded stresses affecting a flange are the internal stresses (due to pressure and temperature) and external loads (compressiom, bending, etc..) due to piping loads (thermal expansion of attached pipes), bolting loads, wind and earthquake loads, weight of equipment bolted to the flange, including asymmetric loads, etc..If you can estimate all the external loads to an "equivalent internal pressure at a given temperature", then it will be obvious that the sum of the internal pressure at temperature and the "equivalent pressure" at temperature cannot exceed the pressure / temperature rating of the flange. How to estimate the "equivalent pressure", refer above posts.
Cheers,
gr2vessels
RE: Nozzle Flange Load Calculation
I will also note a paper http://st
I also want to point out one other "flaw" in your argument - what value do you use for the assembly bolt stress/load? You certainly won't find that value in B16.5. In fact, B16.5 is purposely silent on that issue because that assembly bolt load is a function of the gasket selected, which can range from a self-energizing gasket to a solid metal gasket. Both extrema require very different bolt loads, and yet both "satisfy" B16.5. How is that possible, unless there is additional margin in the B16.5 ratings to allow for external loads, even when the design pressure equals the flange rating pressure (at a given temperature...)?
RE: Nozzle Flange Load Calculation
gr2vessels
RE: Nozzle Flange Load Calculation
1) The external compressive load is tending to keep the flange "closed" - this would tend toward staying with the Class 150 flange. However, I would be a little concerned about crushing the gasket.
2) The wind, seismic, and piping loads (very very good that you are considering the piping loads - I am appalled by how many vessel engineers don't, but that's the subject of a future thread...) will tend to "open" the flange on one side.
3) Depending on the relative magnitudes of the above, the result could go either way. However, the gasket crushing on the "closed" side of the flange is a real issue - I would calculate the compressive stress in the gasket at the "closed" side, and compare it to the maximum allowable compressive stress for that gasket.
Now for my rant - I am thoroughly convinced that the calculations contained in ASME Section VIII, Division 1, Appendix 2 (relating in particular to stress in the flange) have nothing whatsoever to do with the successful operation of a flange. The biggest issue for a flange is leakage, and these calculations don't address leakage at all. Therefore, I put 0.00% faith in relying on these calcs. I view them as a mere hurdle for a situation where the calculation is mandated by law, but they have no bearing on how the flange will actually perform. Rant over.
RE: Nozzle Flange Load Calculation
You're right Sect1, Appendix 2 design is not for design to prevent leakage during operation. It says in paragraph 2-1(a) "These rules are for hydrostatic end load and gasket seating." The purpose of the code is only to ensure "safety". However, since the 05 addendum, paragraph 2-14 was introduced to address flange rigidity to prevent in-service leakage. It is a simple calculation to check whether or not your flange will leak. How good is it? I wish I know. Anyone care to share some experience with the rigidity criterion?
To really check flange leakage I would want to use the flange leakage FEA software from Paulin.
RE: Nozzle Flange Load Calculation
But, the fundamental issue for me is whether any additional testing needs to be performed on a B16.5 flange that is attached to a nozzle, when the nozzle is evaluated for external loads. This is outside the scope of Appendix 2 but the intent is to meet the requirements of Sect. VIII Div. I.
RE: Nozzle Flange Load Calculation
Your original question does not ask about "testing". When you say testing, I interprete it mean NDE or some mechanical testing. If so, my humble opinion for you is no, you don't do any testing when the B16.5 flange is evaluated for external loads.
But you are still wondering if it is normal to "also" check the stresses in the flange when checking nozzle stresses due to external piping loads, the answer is No. Having said the general answer, you need your own engineering judgement to deceide whether you need to do it or not depending on the design conditions you are looking at.
Good luck.
RE: Nozzle Flange Load Calculation
RE: Nozzle Flange Load Calculation
My recent wording was not correct. I did not mean "testing" as in NDE, I meant evaluating using calculations, FEA, or something like that.
RE: Nozzle Flange Load Calculation
to keep their damn loads off my vessel. We have always found this to be easily accomplished by using the inherent flexibility of a piping system. I personally have never seen where the initial loads cannot be mitigated by a careful study of the piping. If it takes more that a couple of 500 lb. chain falls during installation something is amiss.
As I've stated before we have systems that a highly cyclic service at 2000 psig @ 600F where we used Class 2500 flanges with spiral wound gaskets. The Class 2500 flanges are for gasket seating. Parts of this system is on a make and break cycle of two weeks. These flanges do not leak. This system has also proven that if the proper flange is used and made up properly they can be very forgiving. Sometime ago when the problem of flange leaks got on the front burner we had teams to attack the problem, which we didn't have. The jest of the story was that after reams and reams of nebulous numbers were rolled out for all the site piping systems the only meaningful results were that gasket seating ruled with the primary stress coming from thermal expansion. Take care of these two conditions using any method and you are pretty certain that you can maintain the integrity of the piping system.
Flanges on our site were like the flavor of the month on 3-4 year cycle.
The above mentioned piping was once converted to a system of lap joint flanges. We used Schedule 80 stub ends where we welded the edges of flange and used 4 bolt backup flanges. This was to eliminate the gasket. We also made the raised face 1" high on some Class 2500 flanges by cutting the flange face back and doing the same thing but again only using 4 bolts to handle the hydraulic end force. This concept evolved from an older process that operated operated at 6500 psig where all flanges were thread on 2" pipe or tube with the gasket being a modified Bridgman ring, self energizing. The flange was four bolt two inch thick piece.
I shouldn't tell this but if you have enough flange for the job you can profane all the rules with no trouble at all. On each of polymer process units we have a 2 piece bypass line around a process filter that consists of 8 ft of pipe with a Grayloc on one end and Class 2500 flange on the other. The Class 2500 flanges are normally made up with no problem, but in time the pipe spools shorten or get mismatch where the gap exceeds the spacing for one gasket to seat. The answer initiated by the mechanics was to put two or three spiral wound gaskets to act like a dutchman. I can't tell the maximum number that have been used. The system has never leaked.
Again if you got a properly designed flange strange and wondrous things can be accommodated. I've mentioned this before where we have a system, 24", that operates at 250 psig @ 1200F where the spiral wound gasket filler has to be mica. On two occasions graphite filled gaskets have been used with no problem even though the graphite was oxidized almost instantly by the process. The spirals alone held.