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Research Papers: Design and Analysis

Failure Modes of American Petroleum Institute 12F Tanks With a Rectangular Cleanout and Stepped Shell Design

[+] Author and Article Information
Eyas Azzuni

Lyles School of Civil Engineering,
Purdue University,
West Lafayette, IN 47907
e-mail: eazzuni@purdue.edu

Sukru Guzey

Lyles School of Civil Engineering,
Purdue University,
West Lafayette, IN 47907
e-mail: guzey@purdue.edu

1Corresponding author.

Contributed by the Pressure Vessel and Piping Division of ASME for publication in the JOURNAL OF PRESSURE VESSEL TECHNOLOGY. Manuscript received December 5, 2017; final manuscript received August 24, 2018; published online November 12, 2018. Assoc. Editor: Steve J. Hensel.

J. Pressure Vessel Technol 140(6), 061203 (Nov 12, 2018) (18 pages) Paper No: PVT-17-1248; doi: 10.1115/1.4041340 History: Received December 05, 2017; Revised August 24, 2018

The design and fabrication of shop-welded and prefabricated relatively small tanks, when compared to field-welded tanks, used in the upstream segment of the oil and gas industry is governed by the American Petroleum Institute specification 12F (API 12F). This study explores the changing designs of API 12F tanks to include a new rectangular cleanout design with reinforcement as shell extension internally of cleanout frame and a stepped shell design. This study also investigated the introduction of two additional tank sizes in addition to existing eleven tank sizes in the current 12th edition of API 12F. The adequacy of the new design changes and proposed tank designs were verified by elastic stress analysis with nonlinear geometry, elastic–plastic stress analysis with nonlinear geometry, and elastic buckling analysis to verify their ability to operate at a design internal pressure of 16 oz/in2 (6.9 kPa) and maximum pressure during emergency venting of 24 oz/in2 (10.3 kPa). A vacuum pressure of 1.5 oz/in2 (0.43 kPa) was also investigated using the elastic buckling analysis. The stress levels and uplift of the tanks are reported in this report to provide insights into the behavior of proposed API 12F tanks exposed to higher internal pressure and vacuum pressure.

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References

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Figures

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Fig. 1

Proposed tank design with a rectangular cleanout and a stepped shell design where the bottom 4 ft (1.2 m) high shell is 1/4 in (6.4 mm) thick and the top shell is 3/16 in (4.8 mm) thick

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Fig. 2

Typical configuration of rafters in tanks with diameters equal to or greater than 15.5 ft (4.7 m). Shown in the figure tank case 10 with a diameter of 15.5 ft (4.7 m) and shell height of 16 ft (4.9 m).

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Fig. 3

New proposed cleanout opening where it is reinforced with a shell extension internally of frame (to convert in to mm multiply values by 25.4)

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Fig. 4

A Rectangular cleanout design (a) the actual cleanout geometry and dimension as proposed in the draft version 13th edition API 12F dated Nov/9/2017 [29] (b) schematic view of the modeled cleanout in FEA without the bolts, bolting flange, or gasket (to convert in to mm multiply values by 25.4, to convert ft to m multiply values by 0.3)

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Fig. 5

Top and bottom junctions with a chime projection of 3/8 in (9.5 mm)

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Fig. 6

Shell-to-bottom plate fillet weld geometry and FEA idealization

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Fig. 7

True stress–strain relationship for mild steel ASTM A36 based on ASME VIII-2 (multiply by conversion factor of 6.89 to convert ksi results to MPa)

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Fig. 8

Tank case 3 with a finer mesh near the top, bottom, and cleanout junctions

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Fig. 9

Legend convention used in plots provided in this study

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Fig. 10

Finite element analysis stress plot for a tank case 3 with shell and roof thickness of 3/16 in (4.8 mm). The legend shows units of psi. The bottom plate is 1/4 in (6.4 mm) thick with a product level of 18 in (0.46 m). The internal pressure is 114 oz/in2 (49.1 kPa). The localized stresses near the cleanout are ignored (multiply by conversion factor of 6.89 to convert psi results to kPa).

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Fig. 11

Internal pressure causing yielding failure of top junction (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa)

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Fig. 12

Internal pressure causing failure of bottom junction (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa)

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Fig. 13

Internal pressure causing yielding failure of top junction with the results grouped based on nominal diameter (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa and by the conversion factor of 0.3 to convert ft to m)

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Fig. 14

Relative failure ratio of bottom junction to top junction measured at the pressure value that causes yielding for each junction

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Fig. 15

Top and bottom junctions used to extract data

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Fig. 16

Membrane stress in top junction due to internal pressure of 16 oz/in2 (6.9 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 17

Membrane stress in top junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 18

Membrane plus bending stress in top junction due to internal pressure of 16 oz/in2 (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 19

Membrane plus bending stress in top junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 20

Membrane stress in bottom junction due to internal pressure of 16 oz/in2 (6.9 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 21

Membrane stress in bottom junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 22

Membrane plus bending stress in bottom junction due to internal pressure of 16 oz/in2 (6.9 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 23

Membrane plus bending stress in bottom junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 24

Typical location of maximum membrane von Mises stress at bottom shell for case 14 with entire shell thickness of 1/4 in (6.4 mm), bottom plate thickness of 1/4 in (6.4 mm), roof thickness of 1/4 in (6.4 mm), and stored product level of 18 in (0.46 m) at internal pressure of 24 oz/in2 (10.3 kPa) with the stress units of psi (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 25

Cleanout junction from the outside and inside. The contour lines are to highlight the geometry of the cleanout and help identify geometric changes.

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Fig. 26

Typical location of maximum membrane von Mises stress near the meeting point of bottom plate and cleanout weld. The legend shows the stress magnitude in psi (multiply by conversion factor of 6.89 to convert psi results to kPa).

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Fig. 27

Typical location of maximum membrane plus bending von Mises stress near the meeting point of bottom plate and cleanout weld. The legend shows the stress magnitude in psi (multiply by conversion factor of 6.89 to convert psi results to kPa).

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Fig. 28

Membrane stress in cleanout junction due to internal pressure of 16 oz/in2 (6.9 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 29

Membrane stress in cleanout junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 30

Membrane plus bending stress in cleanout junction due to internal pressure of 16 oz/in2 (6.9 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 31

Membrane plus bending stress in cleanout junction due to internal pressure of 24 oz/in2 (10.3 kPa) (multiply by conversion factor of 6.89 to convert psi results to kPa)

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Fig. 32

Maximum uplift at tank bottom reported for internal pressure level of 16 oz/in2 (6.9 kPa) (to convert in to mm multiply values by 25.4)

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Fig. 33

Maximum uplift at tank bottom reported for internal pressure level of 24 oz/in2 (10.3 kPa) (to convert in to mm multiply values by 25.4)

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Fig. 34

Plastic strain, εpeq, as compared to the limiting triaxial strain, εL, measured near the cleanout junction, the shown dashed line represents the global minimum triaxial strain from all the tank cases

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Fig. 35

Internal pressure divided by design factor that causes global rupture failure in each tank (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa)

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Fig. 36

Tank with diameter of 9.5 ft (2.9 m) and height of 12 ft (3.7 m), case 3, behavior with nonlinear material at internal pressure of 112 oz/in2 (48.3 kPa) compared to an over pressurized tank in Wyoming [37] of most probably the same diameter and height. The von Mises stress magnitude in psi is used to contour the deformed shape of the tank. The displacement magnification factor for the deformed shape is 1.0. The bottom plate thickness of the modeled tank was 1/4 in (6.4 mm), the shell and roof plate thicknesses were 3/16 in (4.8 mm), and the product level was 18 in (0.46 m) (multiply by conversion factor of 6.89 to convert psi results to kPa).

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Fig. 37

Buckling of the tanks due to internal pressure, P (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa)

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Fig. 38

Buckling of the tanks due to vacuum pressure, −P (multiply by conversion factor of 0.43 to convert oz/in2 results to kPa)

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Fig. 39

Buckling of the tanks due to positive internal pressure

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Fig. 40

Buckling of the tanks due to vacuum pressure

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