In a Canadian case study, new piping under construction suffered some settlement in the underground portion, as evi- denced by soil movement and slight deflection of the above- ground section of the piping. One of the flange joints on the line was broken to assess the amount of strain in the joint. The joint sprung open with a wide gap that was far in excess of ac- ceptable flange alignment tolerances. This situation warranted an engineering assessment to evaluate the risk. The piping de- sign details are as follows:
• Design code: Canadian Standards Association (CSA) Z662
• Design pressure: 500 psig • Design temperature: 140°F
• Original minimum design metal temperature (MDMT): –49°F
• Material specification for the pipe: API X52 pipe (yield stress: 52,000 psi)
• Pipe size: 16-inch (in.) outside diameter (OD) × 0.375-in. thickness
• No post-weld heat treatment.
Piping and flange integrity concerns. The aboveground piping with the opened flange joint is shown in FIG. 1. Mea-
sured vs. allowable alignment tolerance at this flange is pre- sented in TABLE 1.
Bolting misaligned flanges together introduces residual loads and moments into the piping system. In this case study, flange leakage was recognized as a great risk, and so the engi- neering assessment was limited to aboveground piping with flange joints. Hydrotesting was considered an adequate check for the underground piping.
To evaluate the stresses in the piping and to assess the forces at the flange, the piping was modeled using proprietary stress- analysis software. If the aboveground piping is treated as two sections, A and B, then the total flange gap of 2.1 in. should be the sum of the deflections at the end flanges of A and B.
The deflections will be inversely proportional to the stiff- ness of the pipe sections A and B:
(DA + DB) = 2.1 in., and (DA ÷ DB) = (KB ÷ KA) (1)
where DA and DB are the displacements, and KA and KB repre-
sent the stiffness measurements.
DA and DB were computed using Eq. 1, and the displace-
ment stress at both A and B were evaluated. Section B, shown in FIG. 2, was found to have maximum stresses. For this reason,
detailed analysis was restricted to this section.
Several load cases were analyzed using stress-analysis software: 1. WNC + D1: Installation case (weight with no contents
and with displacement)
2. W + D1 + T1 + P1: Maximum operating temperature case with displacement
3. W + D1 + T3 + P1: Minimum operating temperature case with displacement (for brittle fracture calculation) 4. W + D1 + T4 + P1: Minimum operating temperature
case with displacement (for brittle fracture calculation) 5. Expansion case: (W + D1 + T1 + P1) − (W+ P1). In these load cases, T1 = 140°F, T3 = −49°F and T4 = −40°F. An installation temperature of 14°F was used for the assessment.
The additional stresses introduced in the piping by exter- nally applied forces, used in aligning the flanges, are secondary stresses. CSA Z662 uses Eq. 2 for estimating operating stress:
σ = [PD ÷ (4t) + MA ÷ Z + i × MC ÷ Z] (2) where:
P = Pressure
i = Stress intensification factor (SIF)
MA = Moments due to primary loads (sustained loads)
FIG. 1. Underground/aboveground transition section.
TABLE 1. Measured vs. allowable alignment tolerance at opened fl ange joint in aboveground piping
Item ASME PCC-1 tolerance, in. Measured tolerance, in. Gap, max./min. 0.03125 0.85
Max. lateral off set at fl ange OD 0.0625 0.96
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Maintenance and Reliability
MC = Moments due to secondary loads (displacement-
controlled loading)
D = Pipe outside diameter t = Pipe thickness
Z = Section modulus.
When SIF is equated to unity, the peak stress component will be removed, and the resulting stresses are primary and sec- ondary stresses.
Failure analysis. There are several possible reasons for piping failure, and a failure analysis must be performed for each case.
Plastic collapse. The aboveground piping did not reveal any visible deformation after the flange assembly. However, there could be residual stresses in the line resulting from dis- placement-controlled loading. These stresses are secondary stresses. Due to their self-limiting nature, secondary stresses cannot cause a plastic collapse.
Pipe buckling. Bending causes tensile stresses at the outer curvature, and compressive stresses at the inner curvature. Only compressive stresses have a tendency to cause buckling.
If resulting compressive strains due to primary and secondary stresses are less than the compressive strain capacity of the pipes, then there is no risk of buckling.
Among the loading scenarios, the installation case produces the highest compressive stress of 37,085 psi (bending stress at node 70 with SIF = 1). Compressive strain = compressive stress ÷ modulus of elasticity, and is shown in Eq. 3:
⑀cf = σ ÷ E = 37,085 ÷ 28,000,000 = 0.0013 = 0.13% (3) The ultimate compressive strain (Eq. C-13 in CSA Z662) is: ⑀c = 0.5 × (t ÷ D) − 0.0025 + 3,000 × [(Pi − Pe) × D ÷ 2tE)]2
= 0.5 × (0.375 ÷ 16) − 0.0025 + 0 = 0.0092 (4) The allowable compressive strain is Фec × ⑀c (Eq. C-12 in CSA Z662), where Фec is the resistance factor for compressive strain and is equal to 0.8. The allowable compressive strain is Фec × ⑀c, or 0.8 × 0.0092 = 0.0074 = 0.74%. The calculated compressive strain of 0.13% is less than the allowable compres- sive strain of 0.74%. Therefore, buckling is unlikely.
Fatigue failure. Only cyclic loads contribute to fatigue. The displacement applied to align the flange is not cyclic in nature; therefore, no specific fatigue evaluation is required. The code expansion case is primarily used as a check against fatigue failure:
• Expansion stress SE = 22,400 psi • Code allowable stress SA = 37,489 psi • Since SE < SA, it passes the design.
Brittle fracture. This evaluation is carried out using pro- prietary engineering assessment software, as per API 579-1/ ASME FFS-1.2 Brittle fracture assessment requires an estima- tion of the stress intensity factor driving the crack. The prima- ry and secondary stresses have different impacts on the stress intensity factor. For this reason, these components should be individually estimated.
The operating stress is calculated in Eq. 5:
σ = [PD ÷ (4t) + MA ÷ Z + MC ÷ Z], when SIF = 1 (5) Therefore, the operating stress at −49°F = 49,941 psi. [PD ÷ (4t) + MA ÷ Z] is sustained stress, which is the primary stress.
The primary stress equals 8,060 psi.
The secondary stress component is obtained by subtract- ing the primary stress from the operating stress; i.e., 49,941 − 8,060 = 41,881 psi. The residual stress at the weld is also a sec- ondary stress, which is separately estimated by the engineering assessment software using rules in API 579-1, Appendix E. Since the pipe welds are not heat treated, weld residual stresses are as high as yield strength.
The fracture toughness required for the assessment is esti- mated from Charpy impact test results and the yield strength of the material. At −49°F, the material met a minimum Charpy V-notch number of 15 ft-lb, and the minimum specified yield strength σy = 52,000 psi.
Lower-bound fracture toughness was calculated using the Welding Research Council 265 correlation:
Klc = (2 σy) √ (CVN ÷ σy − 0.01) = 2 ×
52 √ (15 ÷ 52 − 0.01) = 45 ksi √in. (6) where Klc = fracture toughness.
FIG. 2. Piping section B.
0.0
0.0 0.1 0.2 0.3 0.4 0.5 0.6 Load ratio, Lr
Failure assessment diagram (FAD): DPHI = 90°
(0.16, 1.05) 0.7 0.8 0.9 1.0 1.1 1.2 1.3 Toughness r atio, Kr 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3
Hydrocarbon Processing | MAY 201469
Maintenance and Reliability
A crack-like flaw was assumed, as per the rules in ASME Section 8, Division 2/API 579-1 for MDMT determination using a fracture mechanics approach. The crack location, ori- entation and size are as follows:
• Location: A surface-breaking flaw originating from the OD surface
• Orientation: Parallel to the circumferential weld • Flaw depth: 0.09375 in. (flaw depth is 25% of
the thickness)
• Flaw length 2C: 0.5625 in. (six times the depth).
Brittle fracture assessment requires a failure assessment diagram (FAD), which depicts the interaction between two failure modes, namely plastic collapse (represented by Lr) and brittle fracture (represented by Kr). If (Lr, Kr) falls below the FAD, it passes assessment.
Running engineering-assessment software for an assessment temperature of −49°F yielded a (Lr, Kr) that fell above the FAD (FIG. 3). A value of Kr > 1 clearly indicated a risk of brittle fracture.
The exercise was repeated with an assessment temperature of −40°F, with primary and secondary stresses of 8,060 psi and 41,261 psi, respectively. The fracture toughness at −40°F was calculated using Eqs. 7–10:
Klc = 33.2 + 2.806 exp × [0.02 × (T − Tref + 100)] (7)
Tref = −28.7°F when solved using Klc =
45 ksi √in. for T = −49°F (8)
T = −40°F and Tref = −28.7°F (9)
Klc = 49.7 ksi√in. (10)
The fracture mechanics assessment resulted in (Lr, Kr) = (0.16, 0.88), which fell below the FAD. Therefore, −40°F is ac- ceptable as the MDMT.
Local failure. When tensile strain exceeds allowable limits, and it manifests as a rupture, local failure occurs. CSA Z662 establishes strain limits per principles of fracture mechanics, which is a rigorous procedure. Instead of utilizing the limits outlined in CSA Z662, local failure was ruled out using alterna- tive reasoning. CSA Z662 allows an installation strain of 2.5% (primary and secondary strains). The primary and secondary tensile stresses are the highest for the operating case at −40°F, which is 49,321 psi. This translates to a strain of just 0.18%. Flange leakage. Several methods exist for calculating the risk of flange leakage.
Equivalent pressure method. The axial force and moment at the flange at operating conditions are converted to an equiva- lent pressure in this method. Stress-analysis software performs this calculation, and its computation is shown in Eq. 11:
Equivalent pressure Pe = (16M ÷ 3.14G3) +
(4Fa ÷ 3.14G2) + P (11)
where M is the resultant moment in in.-lb, Fa is the axial force in lb, P is the internal pressure and G is the effective gasket di- ameter. Therefore:
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Maintenance and Reliability
Pe = 16 × 173,674 × 12 ÷ (3.14 × 17.443) + 4 ×23,463 ÷ (3.14 × 17.442) + 500 = 2,600 psi (12)
This result is higher than the 716 psig allowed at the design temperature of 140°F for a Class 300 flange. Therefore, the equivalent pressure method indicates a risk of flange leakage.
ASME B31.8 flange leakage calculation. In this calcula- tion, the allowable moment for flange leakage is ML = (C ÷ 4) × (SbAb − PAp). The Class 300, 16-in. flange has 20 bolts of size 1¼ in. The tensile area of the bolt is 1 in. In this calculation:
• Ab = Total area of flange bolts, in.2 = 20 × 1 in.2 = 20 in.2 • Ap = Area to outside of gasket contact = π ÷ 4 × (18.25)2
= 261.5 in.2
• C = Bolt circle, in. = 21.25 in. • P = Internal pressure = 500 psi
• Bolt stress Sb = 50,000 × 0.7 = 35,000 psi, where 50,000 psi is the initial bolt stress, and the factor 0.7 is used to account for a joint relaxation of 30%.
Therefore, ML = (21.25 ÷ 4) × (35,000 × 20 – 500 × 261.5) = 3,024,140 in.-lb = 252,012 ft-lb.
The maximum bending moment in the flange from the stress-analysis software’s operating case is 173,674 ft-lb. This is less than the allowable flange moment of 252,012 ft-lb. There- fore, this method predicts that flange leakage is unlikely.
Modified flange leakage calculation. Since two common methods for flange leakage predicted contradicting results, an attempt was made to develop an alternate flange leakage calcula- tion. This calculation is a modification of the ASME B31.8 leak- age equation integrating principles from ASME PCC-1. When the external moment is less than ML, the gasket seating stress will be adequate. The modified equation is shown in Eq. 13:
ML = (C ÷ 4) × (SbAb – Fa – Hp − AGPG) (13)
where:
Gasket area = AG = 0.785 × (Do2 − Di2) = 0.785 × (18.252 −
16.632) = 44.36 in.2
Total bolting force = Sb × Ab = 35,000 psi × 20 = 35,000 × 20 = 700,000 lb
Hydrostatic force = Hp = 0.785 G2 × P = 0.785 × 17.442 × 500 = 119,380 lb
Axial force in the flange = Fa = 23,463 lb (from operating case)
PG = 4,000 psi, or the minimum recommended operating gas-
ket seating stress for a spiral-wound gasket. (Note: This stress is larger than the gasket factor multiplied by the design pressure, which is used for the flange design.) Gasket reaction = AGPG = 44.36 × 4,000 = 177,440 lb.
Therefore:
ML = (21.25 ÷ 4) × (700,000 − 23,463 − 119,380 −
177,440) ÷ 12 = 168,104 in.-lb (14)
The bending moment in the flange equals 173,674 ft-lb, which is 3% larger than calculated ML, indicating a slight risk
of flange leakage. The leakage risk can be minimized by retorquing the flanges to a bolt stress of 50,000 psi.
Bolt and flange stress limits. External alignment devices, such as come-alongs, will be used to align the flanges. When the alignment devices are removed af- ter tightening the nuts, some changes in the bolt loads may occur, although they will not be substantial.3, 4
Based on finite-element analysis results,4 it is known that external moments have maximum influ- ence on the longitudinal hub stress. The gap between mating flanges was measured after releasing the come-alongs to rule out flange rotation that might result from hub stress, if excessive. Results. The analyses showed that plastic collapse, buckling, local failure and fatigue failure of the piping are unlikely. A brittle fracture assessment at an MDMT of −49°F failed, but it passed for −40°F. The line must be rerated for an MDMT of −40°F. Also, the modified flange leakage calculation is ac- cepted as the final criteria for flange integrity. The flange leak- age can be addressed by retorquing the flanges to a bolt stress of 50,000 psi.
If flange alignment tolerance exceeds limits, it should be corrected. When a deviation is accepted, an engineering analy- sis should be performed to identify the possible risks and miti- gations. In this example, joint integrity was established by re- torquing. This was proven by a hydrotest.
Prior to the assessment, flange integrity was treated as the only major concern. However, the assessment revealed that brit- tle fracture is an equally important concern. Like flange integ- rity, brittle fracture is influenced by residual tensile stress that results from forces and moments applied in aligning the flanges.
The modified flange leakage appears to reasonably predict joint integrity for this case study. The calculations may be use- ful for assessments of other flange joints in the presence of ex- ternally applied forces and moments.
LITERATURE CITED
1 Canadian Standards Association, Z662-11, Oil and Gas Pipeline Systems, 6th Ed., 2011.
2 American Petroleum Institute 579-1/ASME FFS-1, Fitness-for-Service, 2nd Ed., June 2007.
3 American Society of Mechanical Engineers, Standards and Certification, “PCC-1- 2010: Guidelines for pressure boundary bolted flange joint assembly,” 2010. 4 Tagaki, Y., H. Torii, T. Sawa and Y. Omiya, “Effect of external bending moment on
sealing performance of pipe flange connection,” ASME 2010 Pressure Vessels and Piping Division/K-PVP Conference, Bellevue, Washington, July 2010.
JOHN THARAKAN is a static equipment specialist at Suncor Energy Inc.’s Maintenance and Reliability group in Canada. He has an MS degree in mechanical engineering design and more than 30 years of experience in the oil industry. Mr. Tharakan develops best practices for reliability improvement and performs failure analyses and fitness- for-service evaluations on in-service equipment.
MUHAMMAD ANISUZZAMAN is a mechanical engineer at Suncor Energy Services Inc. in Canada. He holds an MS degree in mechanical engineering from McGill University in Montreal, Canada. Mr. Anisuzzaman’s research interests include fracture mechanics and static and dynamic stress analyses of piping and mechanical systems.