Cycle Life Analysis
File: PVE-3380
Last Updated: October 20, 2009
DV
Commentary:
When is a Cycle Life Analysis required?
VIII-1 does not provide a hard and fast guideline for when fatigue analysis is required. The older VIII-2 code specified that it would not be required if the number of cycles is less than 1000 (VIII-2 2001 AD-160.2 Condition A). The new VIII-2 states that fatigue life analysis is required above 100,000 cycles and often required at lower cycles (VIII-2 2008 5.5.2.1(c)).
From our experience, we know that a vessel designed to VIII-1 has maximum stresses between 2x and 4x the allowable design stress Sa. These maximum stresses are found in "local" areas such as flange necks, knuckles or heads, nozzle to shell junctions or other changes in shape. Some of this stress is caused by flange seating loads or support loads that do not fluctuate so do not affect the cycle life. Other stresses change with each applied pressure cycle and need to be considered.
The difference in stress between unpressurized and pressurized conditions depends on the design features used in the vessel - what type of head, how the nozzles are attached, weld efficiency and more. For a SA-516 70 vessel at ambient temperature the maximum allowable stress Sa = 20,000 psi. The expected maximum stress is between 40,000 and 80,000 psi. This corresponds to an expected cycle life of 1,100 to 8,500 cycles.
Many pressure vessels are not expected to experience 1,100 cycles. Many other factors will affect the actual fatigue life: corner joint details; operating at less than the full design pressure; temperature difference effects; design of nozzle details; and type and quality of welding or other factors that change the surface finish and more.
Cycle Life Sample
For this sample filter vessel the operating pressure range cycles between a full vacuum and a pressure of 230 psi. The design has flanged connections. The flange stresses are part seating and part operating. To get an accurate fatigue analysis two runs are required - one at the pressure case, and the other at the vacuum case. When the two cases are subtracted (according to the rules of VIII-2 5.5.3.2) the operating stress range is obtained without flange seating effects. This stress difference is not the same as simply taking the peak stress from one the higher pressure case.
Step by Step
See sample #6 for an introduction to FEA on a simple shape. This cycle life study uses the same modeling and analysis methods on a more complex shape.
 The solid model used for the fatigue analysis |
|
 Mesh with refined areas where higher stresses are located. This model has over 1,000,000 nodes. |
 230 psi loading deformation results (deformation x100) |
|
 Full vacuum loading deformation results (deformation x200) Note that the vessel deformation shape is different than the 230 psi case especially note the deformed shapes of the heads) |
 Stress plot for the 230 psi case. The peak stress indicated is not the final maximum alternating stress used. |
|
|
Once the vacuum case is subtracted from the 230 psi case, the maximum stress range is found to be 42,000 psi. This is the difference in stress between the full vacuum and the 230 psi operating case without any flange seating effects. (It is also less than the 62,000 psi peak shown in the picture above which is caused by a combination of flange seating and operating loads). The alternating stress is computed to be 21,000 psi for a predicted cycle life of 49,000 cycles.
Fatigue life curve for the vessel - expected life is 49,000 cycles.
Conclusion
Cycle life analysis is often required for products subject to cyclic processes and as a requirement for provincial CRN registration for many classes of pressurized components. We have successfully used this FEA method to provide reports calculating pressure component fatigue life's accepted by many Authorized Inspectors and review engineers.
Downloads (pdf format):
