Fatigue Screening
Fatigue Screening
Fatigue Screening
Here an ASME PTB-3 Validation for “Example E5.5.3 – Elastic Stress Analysis, and Equivalent
Stresses” to check for protection against failure from cyclic loadings per ASME Sec VIII Div
2 is carried out using ANSYS.
Problem Statement:
Evaluate the vessel top head and shell base metal regions given in Example E5.2.1 in accordance
with the fatigue methodology provided in paragraph 5.5.3. Note that the nozzle to head weld is
machined and subjected to full volumetric examination and both ID and OD surfaces receive
MT/PT and VT. The shell to head weld is in the as-welded condition and the OD surface receive
the same inspection as above. The ID surface receives only full volumetric examination. The
cyclic loading design requirements given in the Users’ Design Specification are provided below:
Operating pressure = 380 psig at 125oF
Corrosion Allowance = 0.125 inches
Cyclic Life Requirement = 20000 full pressure cycles
Number of Shutdowns/Startups = 20
Solution:
Typically before performing a detailed fatigue analysis the equipment is checked to see if it
meets fatigue screening criteria per sec 5.5.2. Three methods for fatigue screening criteria are
provided in the code. We’ll see each one of them.
The allowable number of cycles can be obtained as N(C1S) = 10X (3-F.21) where X is obtained
from Annex 3-F. A summary is shown in the above table.
Since NΔFP > N(C1S) for both components, fatigue evaluation for both are required.
Fatigue Evaluation – Elastic Stress Analysis, and Equivalent Stresses
Finite Element Model from Example E5.2.1 is used in this analysis. The only cyclic load here is
the pressure load which varies from 0 to 380 psi. Typically to find alternating stress amplitude
one needs to solve two load cases:
Case 1: The maximum cyclic load (in our case pressure of 380 psi)
Case 2: The minimum cyclic load (in our case pressure of 0 psi)
However for a linear elastic analysis like the one which we would be doing, the response
(deformations, stresses) will be directly proportional to the applied loads. Hence Case 3 can be
directly obtained by applying ΔP load = Pmax – Pmin = 380 – 0 = 380 psi in a single load case.
Note, this approach would work even if we had a non zero Pmin.
ANALYSIS SUMMARY
Internal Pressure of ΔP = 380 psi was applied on the inner surfaces of the model.
Pressure Thrust of 888.75 psi was applied on the flange face as negative pressure.
Axial displacement was arrested at the shell base.
ANALYSIS RESULTS
Since Membrane + Bending Stresses are less than SPS, fatigue penalty factor Ke = 1
Computation of alternating stresses @ SCLs
Salt = Kf*Ke*Total Von-Mises Stress Range/2
Here Kf, the fatigue strength reduction factor is taken from Tables 5.11 and 5.12.
The below table summarizes the results (values in red are from this analysis and those in black
are from ASME PTB-3.
Finally allowable number of cycles, need to be calculated for Salt using Annex 3-F. Calculation
for allowable number of cycles are not being done for this analysis, however reference table from
ASME PTB-3 is produced below.
Knowing the allowable number of cycles N and the design cycle life of the component n fatigue
damage factor can be calculated as D = n/N. In case multiple cyclic loads are present, the miner’s
rule D = n1/N1 + n2/N2 … can be used.
This results in calculated fatigue damage for the limiting region (nozzle outside radius) of 0.073
(from ASME PTB-3). Similarly, a fatigue damage of 0.039 (from ASME PTB-3) is calculated
for the head knuckle.