Material in creep range 1

[Ehiman1]

Hello everyone,
I am checking a calculation of a pressure vessel, where the operating temperature is 480 C and design temperature is 510 C for SA-387 gr.11 cl.1.
In this case In accordance with ASME code the design temperature for this material is in creep range, so in my opinion design temperature should updated with Larson Miller formula to consider this thing.
In this case, with 20 years of utilization we have a design temperature of 520 C that means an higher shell thickness.
Could you please confirm my reasoning?
Because in the calculation vendor didn’t consider this modification of design temperature.

Thanks

 

[TGS4]

The allowable stress in Section II, Part D (I am assuming that you are fabricating this to ASME Section VIII, Division 1 - even though you didn't say) include for the effects of creep. No additional considerations are necessary.

Plus, it should be noted that you never operate the equipment at he design temperature - any life calculation should be based on the operating temperature.

 

[GD2]

Ehiman1,
Don’t you think your provided info is incomplete? What’s the design pressure at those design temperatures?

GDD
Canada

 

[DriveMeNuts]

The allowable stress in Section II, Part D are for 100,000 hours (or 11 years). What is the ASME requirement beyond this time limit to 20 years as required in the OP?
Are physical tests and measurements mandatory at 11 years even if in the original design report you calculate a lower allowable stress from fundamentals corresponding to a longer life of 20 years?
Would this 20 year vessel (without inspections at 11 years) be able to have a U-Stamp?
What if you only wanted a life of 5 years and wanted to push the material to its limit? That involves calculating a higher stress than required by Part D. Can this 5 year vessel have a U-Stamp?

 

[GD2]

ASME Committee has set the following criteria's for equipment operating in creep regime. The max allowable stress is the minimum of the following:
(1) 100% of the average stress to produce a creep rate of 0.01%/1,000 hr
(2) 100 F[sub]avg[/sub]% of the average stress to cause rupture at the end of 100,000 hr
(3) 80% of the minimum stress to cause rupture at the end of 100,000 hr.

Note criteria 2 and 3 that talks about the end of 100,000 hrs. Up to 815 C, Favg= 0.67. above 815C, Favg = 1/n. The multiplier 0.67 increase the life above 100,000 hrs by a factor of 1.5n, where n ranges from 5 to 10 buy may not exceed 0.67. the F-factor is developed by MPC.


GDD
Canada

 

[DriveMeNuts]

GD2, I am aware of Mandatory Appendix 1 of Part D.
Multiplying the average stress to cause rupture at the end of 100,000 hr by the design margin F[sub]avg[/sub] of 0.67 is like dividing yield stress by 1.5 for plastic collapse. F[sub]avg[/sub] is a design margin against creep failure of 1.5. I don't see how it allows you to increase the creep life.

My questions are:
What if you want to design equipment for 200,000 hours? Do you just recalculate a lower allowable stress than specified by Part D and run the equipment for 200,000 hours without inspections? This time is more than Part D seems to intend for. This seems closely related to what the OP is asking. Surely you wouldn't design the equipment with the Part D stress for 100,000 hours and then run the equipment for 200,000 hours?

What if you want to design equipment for 50,000 hours? Do you just recalculate a higher stress and run the equipment for only 50,000 hours? This stress is higher than what Part D mandates. Are you stuck with the Part D stress and just have to accept wasting material? Very expensive material.

 

[MJCronin]

Lets review the meaning of an ASME U-Stamp:

ASME U Stamp requirements:

The pressure vessels under ASME U stamp requirements are specifically inspected by a third-party authorized inspector. The inspector must review and approve the calculations as well as witness the ASME hydro test. Such inspectors are commissioned by the National Board of Boiler and Pressure Vessel Inspectors. A complete data report is furnished in the form U-1 containing the signature of the authorized inspector. The manufacturers of such pressure vessels need to be registered to the National Board for the production of ASME U stamped pressure vessels. Also, they need to maintain a permanent data record of all pressure vessels.

Also:

The ASME U Stamp is an indication that the pressure vessels adhere to ASME’s guidelines including design, fabrication, inspection, and testing. It is used for the certification and acceptance of pressure vessels. Seeing the “U” symbol on a product ensures that it meets the latest edition of the Code. It also affirms that the vessel is designed and manufactured according to the standards of ASME.

Notice that the U-Stamp references a specific edition of the ASME Code ...... Not a Code edition with a certain assorted group of new allowable stress changes or other rule changes desired by the purchaser !!!

A third-party has a very specific function here; the "third-party authorized inspector" must agree and certify that the vessel under review meets a specific edition of the ASME Code. Its a third party, not a third world inspector

He does not review unique calculations and either agree or disagree with the approach and assumptions that go beyond the rules of ASME

As a purchaser, you can have your pressure vessel competently designed, inspected, certified and tested to whatever rules you want ... Finite Element Analysis is very powerful. Your Vessel can have a five year life or five hundred year life. Its all your business..

But, under the current set of rules, it will not have an ASME U stamp ...

My opinion only

MJCronin
Sr. Process Engineer

 

[DriveMeNuts]

Yeah, so with respect to the OP, simply using Larson Miller to adjust the allowable stress in the original calcs for a life of 20 years is not enough. Once the vessel exceeds 100,000 hours the Engineer needs to consult the "ASME Fitness for Service" Code to extend its life. From memory this code will require that the Larson Miller life adjusted calc be carried out. It also requires that inspections be carried out. If the measured circumference of the vessel has creeped by 1% then the vessel has come to the end of its life according to "(1) 100% of the average stress to produce a creep rate of 0.01%/1,000 hr" from ASME II Part D MA1. ASME FFS-1 probably has other requirements as well.

In terms of a 50,000 hour operating life, I think material just needs to be wasted if I want it to be U-Stamped. The allowable stress in ASME II Part D can't be increased.

Was worth a try.

 

[GD2]

DriveMeNuts,
The 0.67 factor has been applied to the stress value (S[sub]Ravg[sub])[/sub][/sub] for the rupture life of 100,000 hrs. The allowable stress is therefore 0.67S[sub]Ravg[/sub], not Yield strength. ASME believes that this stress reduction would increase the life by a factor of 1.5n, where n value is between 5 to 10. If you apply this factor, the rupture life be expected from 700,000 hours to over million hours. However, this is not proven.

MPC have developed a procedure where the F-Factor can be lowered to a value that offers the average material a life margin of 10 beyond 100,000 hours.

In my opinion, if the designer has designed the vessel based on allowable stress Table 1-100 Sec II Part D Appendix 1, the 0.67 factor shouldn't be a problem for 20 years life.

On another observation, where equipment operates below 25 degrees below the design temperature, an additional increase in life of at least a factor of 2 can be expected (in our case).

Obviously, these life expectancies are in a situation where there is no in-service damage from corrosion, fatigue, hot spots, environmental cracking and creep damage from temperature excursions.
What I don't undersatnd is why Ehiman1 is stating two design temperatures of 510 and 520 C.

GDD
Canada

 

[DriveMeNuts]

GD2,
Yes, so F=0.67 is a design margin on the average Creep failure Stress, similar to the way a margin is applied to the yield stress for elastic design. F=0.67 is also an unproven approximate 1/n design margin on average life (not minimum life).

I find that expressing an opinion/belief from unproven science that it "shouldn't be a problem" for this real-world OP problem is not particularly helpful.

All that the MPC science does is give confidence that periodic monitoring beyond 100,000 hours is effective for safely managing the remaining life of the equipment. In my opinion, this monitoring should be mandatory (especially for U-Stamped vessels).

 

[GD2]

DriveMeNuts,

I agree.

Ehiman1 can always calculate the design lives based on 510 and 520C at the design pressure with the help of Larsen-Miller Parameter (LMP) equations and see what rupture life he gets. It's not very difficult.

GDD
Canada

 

[NRP99]

I find that expressing an opinion/belief from unproven science that it "shouldn't be a problem" for this real-world OP problem is not particularly helpful.

Please read this-

The values in Tables 1A and 1B are established by the Committee only. In the determination of allowable stress values for materials, the Committee is guided by successful experience in service, insofar as evidence of satisfactory performance is available. Such evidence is considered equivalent to test data where operating conditions are known with reasonable certainty.

What GD2 said is correct. The margin on creep rupture stress for 100,000hr life will ensure that the stresses in the component undergoing creep will not go beyond creep rupture stress for 100,000hr life and in the process, conservatively, we have extended creep life by using that margin (less stress than creep rupture stress means higher life). See ASME B31.3-Appendix V for simple example considering creep damage of 1 with case of normal operation period and temperature excursion period.

Ehiman1 can always calculate the design lives based on 510 and 520C at the design pressure with the help of Larsen-Miller Parameter (LMP) equations and see what rupture life he gets. It's not very difficult.

No. Creep phenomenon occurs at the current state of component with "continuous" stress in the component and continuously/consistently exposed to high temperature. And not at an imaginary design temperature which is taken for design purpose.

Ehiman1 So IMO, vendor has done the correct calculation if he has taken the operating temperature. In a scenario, where design temperature is in creep phase but operating temperature is not, its logical to consider that vessel is not operating at creep regime temperature and failure mode will never be creep. It would be better to consult with the process licensor/owner/approver as to what will the actual temperature vessel will experience during operation/excursion and make decision based on that.

 

[jtseng123]

How many of you have really designed a reactor using 1 1/4 Cr-1/2Mo at design tempo 520C or even higher and in service for years ? I did.
20 years of life is the most commonly specified in the industry. The trend is up to 25 years now. Don't understand what is the argument point here for 100,000 or 200,000 hours.
I have been designed many reactors using code allowable stress without stress modification and in service for decades. For sure all have U stamps. Do you think it is no good ?

 

[DriveMeNuts]

NRP99, I don't understand your point. When I read your points from a theoretical average life point of view it makes some sense. However we are not discussing theory and average life. I am referring to the real world safe operation and its associated U-Stamp certification requirements.

What GD2 said is correct. The margin on creep rupture stress for 100,000hr life will ensure that the stresses in the component undergoing creep will not go beyond creep rupture stress for 100,000hr life and in the process, conservatively, we have extended creep life by using that margin (less stress than creep rupture stress means higher life). See ASME B31.3-Appendix V for simple example considering creep damage of 1 with case of normal operation period and temperature excursion period.

You correctly initially refer to the safe operation of a "single item" of equipment with the F factor applied. However in the second part of your statement you incorrectly mix up a single component with a statistical sample of components. You can not extend the creep life by using the F factor because if you increase F even slightly then a small percentage of identical items of equipment coming off the production line will fail before they reach 100,000 hours. If you increase F to 1.0, then 50% of items of equipment coming off a production line will fail before 100,000 hours. How do you know if the "single item" of equipment you refer to at the beginning of your comment will survive 100,000 hours with an increased F value? The answer is you don't know, therefore applying the F value doesn't increase the creep life. This is a real world Engineering Forum. We are not discussing theoretical average life science.

The ASME B31.3-Appendix V calculation has the F factor at its core and will therefore reveal a maximum allowable life of 100,000 hours for an item of U-Stamped equipment designed to utilise all of its material thickness. The only way to extend the life of this equipment beyond 100,000 hours into the statistically dangerous failure region while complying with the law, is to conduct periodic physical tests and measurements in accordance with ASME FFS-1.

Where the design of an item of U-Stmamped equipment has excess shell thickness, at 100,000 hours the user is required to conduct a life extension assessment. To claim at this point that the F factor extends the life and therefore do nothing is misleading and dangerous. The actual operating stress can be can be multiplied by F=0.67 and then a life can be calculated using ASME FFS-1 or perhaps ASME B31.3-Appendix V. Because the F factor is applied to the stress, this extended life to perhaps 200,000 hours does not extend into the statistically dangerous failure region. After 200,000 hours, the only way to extend the life of this equipment while complying with the law, is to conduct periodic physical tests and measurements in accordance with ASME FFS-1.

jtseng123,
I assume your equipment has excess thickness which you utilise at the 100,000 hour mark to calculate an extension of the life of your equipment to 25 years? Perhaps you have added extra thickness and calculated a life of 25 years using ASME FFS-1 at the design stage before fabrication. Or perhaps you are conducting periodic inspections and measurements to predict the failure life? If you are just winging it after 100,000 hours without any life calc or inspections, then that doesn't sound right.

We had a failed Hydrogen reformer tube about five years ago. The Engineer responsible was hauled in front of the Chartered Engineering Tribunal to demonstrate that he did everything by the book. It was the worst day of his life, and thankfully the Tribunal cleared him. Our team in the US, doesn't keep a paper trail to protect themselves.....

 

[GD2]

ASME has set the stress rupture life at 100,000 hrs based on stress rupture tests at 10,000 hrs and extrapolating the stress and the rupture life on a log-log scale. The 100,000 hrs has been set as the design basis for the designers for materials with time-dependent behavior with a high confidence level.
ASME Committee also argues that the rupture life will be much longer than the 100,000 hrs in real life but how long is unrealistic to say.
Ehiman 1 case is an example. The vessel was designed for 510[sup]0[/sup]C but actually run for 20 years at design temperature of 520[sup]0[/sup]C.
Obviously, I don't bite his/her statement. Who runs a vessel for 20 years at design temperature? I would guess he/she means to say operating temperature.
So guys, don't stick to the rule of 100,000 hrs rupture life.
It is always best to do a FFS if you have issues with normal/environment corrosion, hot spots, fatigue and large temperature excursions.

GDD
Canada

 

[GD2]

NRP99,
The ASME B31.3 Appendix V example gives a good idea about how the rupture life is computed. But, it's in the most simplified form in a piping system where its evaluated for S[sub]L[/sub]. the LMP equation used is also a simplified one.
In a pressure vessel, the calculation is a bit different. LMP is a parameter that relates the stress, temperature and rupture life. The LMP Constant C varies remarkably on the minimum and average material properties. In turn, the rupture life also varies remarkably.
LMP is a very important parameter that was developed for materials used in creep service. API RP 530 - design of fired heater tubes gives a great understanding of how it has been used to determine the thickness of the tubes.


GDD
Canada

 

[BJI]

For code design, you take the allowable stress at DT, no adjustments are required. This is a conservative limit which has proven to be acceptable for design, with the justification of the allowable discussed many times already.

If you need to make some life estimation, then you can use the LMP and the operating temperature to determine the hours at OT. This should be greater than the 20 years required so I don't see the concern.

 

[DriveMeNuts]

BJI,
With respect to calculating the 'operating' life, FSS-1 says that a margin shall be applied.
Simply using the stress derived from LMP[sub]ave[/sub] or LMP[sub]min[/sub] isn't good enough, as if you have hundreds of items in service, some will fail.
After a failure, I'm not confident that you would get to keep your Engineering Chartership if you presented these raw values as your estimate at an Engineering Tribunal.
You will be covered if you use 0.8 x LMP[sub]min[/sub] or 0.67 x LMP[sub]ave[/sub] as per the way the ASME code calculates the design stress.

LMP[sub]min[/sub] is set at the 95% confidence interval. Perhaps there is enough information to calculate a new LMP line for a higher acceptable level of confidence.

 

[BJI]

We are still talking about design not FFS per se, so the margins are as discussed: secondary creep rate, 67% of avg stress or 80% of min stress to rupture. The LMP is used to determine equivalent time at temperature, therefore, converting from the design temperature at 100,000h to the equivalent time at the operating temperature. This all would still assume sustained operation at the design pressure, which therefore inherently includes additional margin at the operating pressure. You could get a similar result by performing the Level 1 FFS assessment, which is evaluated at the operating conditions.

If the creep assessment passes the elastic analysis, then it is relatively conservative, as discussed. When performing Level 3 FFS, using an inelastic method, such as MPC Omega, which includes strain accumulation and relaxation/hardening, the creep life determined is significantly increased over that determined from elastic analysis.

 

[DriveMeNuts]

BJI,
I understand and agree with everything you write. My concern is that when it comes to certifying your work you "assume" a lot. This seems very bodgy. Do you actually produce operating calculations with appropriate margins to prove your assumptions, and therefore cover yourself? Or do you just wing it and hope for the best?
If Design and operating conditions are the same or very similar, your assumptions become tight.