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Nozzle Flange Load Calculation 4
- Thread starter weeeds
- Start date
Usually the flange is not the weak link. The neck-to-shell joint is typically the limiting factor. To check the flange stresses I usually make a quick comparison of the bolting loadss. I first determine the max bolt load under the max pressure permitted on the flange. I also compute the max bolt load due to the design pressure plus nozzle loads. If the bolt loads at the flange rating pressure far exceeds the operationg design loads of the vessel, I don't go any further.
Joe Tank
Joe Tank
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1
- #5
gr2vessels
Mechanical
widla,
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
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
There's the Kellogg's method of converting the external moment or force to an equivalent pressure, then add it to the internal pressure, & compare that to the B16.5 allowable.
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
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
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- #8
arto - just to play devil's advocate, why would you compare the value (equivalent pressure + internal pressure) to the design pressure in B16.5?
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...)
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...)
In TEMA code, there is a para RCB 7.165 effective differential design pressure; it was stated that "under certain circumtances the code and other regulatory bodies permit......" Can anybody advice what exactly the circumtances TEMA mentioned? What are the advantages and disadvantages if we using this rule (differential design pressure) for fixed type heat exchanger design.
Thanks in advance
Thanks in advance
ksquare, you should really start another thread for this, but...
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
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
gr2vessels
Mechanical
TGS4,
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
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
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- #12
I'll have to go with the devil on this one. I have never seen a nozzle or flange fail when designed according to code. We lost two vessels to internal explosions that littlerly tore them apart and scattered the remains over a very large area. All the lines of failure were in the shell and head plates proper none crossed the limits of the nozzles or broke the flanges. Some of the pipe attached by flanges to the vessels was torn apart a short distance from the flange. The flanged joint with the short section of pipe traveled together.
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.
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.
The Kellogg's book pg.78 sez to apply the equivalent pressure to an ASME flange analysis.
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.
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.
gr2vessels - I have read that lien many times. However, it does not say "how" to do that. Why would you compare the value (equivalent pressure + internal pressure) to the design pressure in B16.5. Why not 1.1*Pdesign? Why not 2*Pdesign? Why not 20*Pdesign? Why not 0.95*Pdesign? Where does it say compare to 1*Pdesign?
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?
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?
gr2vessels
Mechanical
TGS4,
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
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
So, what you're saying is that if my piping system design pressure is equal to the rating pressure of the flange (something, BTW, which is almost the norm now for new/greenfield construction), then the system has exactly zero margin for external loads? I can't reconcile that with article 2.5.2 of B16.5.
I will also note a paper where the author performed several FEAs on flanges that showed that they had substantially higher capacity w.r.t. stress than what could be assumed per the rating.
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...)?
I will also note a paper where the author performed several FEAs on flanges that showed that they had substantially higher capacity w.r.t. stress than what could be assumed per the rating.
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...)?
gr2vessels
Mechanical
The issue of external nozzle loads is related to the pressure vessels rather than piping joints. I am currently working on a separation column, 42' tall, c/w an integral heater, 30" ID, fitted at the bottom with an ASME B16.47 - 30" NPS 300# WN flange, bolted on the top nozzle of a large horizontal accumulator. The column and accumulator design pressure / temperature is the equivalent of ASME B16.5 / B16.47 Class 150#. Except for the 30" WN Class 300# flanged connection, all other nozzles are rated 150#. The reason for the 300# rated connection is the additional external loading on the column bottom flange (wind, seismic, column weight and piping loads acting on the column) and the recommendation of Clause 2.5.2 of taking care of those additional loads, which pushed the flange rating above Class 150#, in the Class 300#. Also, I used fully detailed and tested gasket factor and seating stress values, given by the manufacturer for the design of the joint. I would prefer the detailed engineering to any interpretation of the code or assume the large margin of capacity on standard flange. I fully agree with you that the B16.5 is an ambiguous reference for bolting and gasket stress calculations, is also inconsistent on ratings with B16.47 and with ASME VIII requirements. It is a prescritive mess. Good perhaps for piping design, but definitely not good enough for pressure vessel designers.
gr2vessels
gr2vessels
That's an interesting scenario. A couple of factors come to mind:
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.
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.
TGS4 and other,
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.
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.
- Thread starter
- #20
There are some great and enlightening points here for which I thank you.
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.
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.
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