Pressure Vessel Thickness Calculator (ASME VIII Div.1)

They solve different problems. A pipe schedule (ASME B36.10M) is a pre-selected commercial wall thickness standardized per nominal size — it knows nothing about your pressure, temperature, material allowable stress, or weld quality. UG-27 vessel thickness is an engineering calculation: t = P·R/(S·E − 0.6P), where S comes from the ASME II-D allowable stress table at your design temperature and E reflects the radiography of the seam. A SCH 40 pipe may be grossly over- or under-designed for a given vessel duty — the vessel formula sizes the wall for the actual load case.

UG-27(c)(1) is not thin-wall theory with a fudge factor — it is the exact Lamé (thick-wall) solution restated. For a cylinder, the maximum hoop stress at the inside surface is σ = P·(R² + Ri²... )/... which the code rewrites as t = P·R/(S·E − 0.6P). The 0.6P in the denominator and the +0.4P in the longitudinal formula are the exact Lamé coefficients for the inside surface, so the formula stays accurate well beyond the t ≤ R/2 limit. That is why UG-27 is valid up to P = 0.385·S·E, far past where a simple 'hoop stress = PR/t' thin-wall formula would be dangerously unconservative.

E comes from Table UW-12 and depends on joint type and radiography, not on engineering judgment. For a Type (1) double-welded butt joint: E = 1.00 with full radiography (RT-1, 100% per UW-11), E = 0.85 with spot radiography (RT-2, UW-52), and E = 0.70 with no radiography. Fully radiographed longitudinal seams are the norm for most ASME-stamped process vessels; 0.85 spot RT is common for lower-pressure storage. Vessels in lethal service require full RT by UW-2.

Yes — MAWP is computed on the corroded thickness (nominal thickness minus corrosion allowance), which is the conservative end-of-life condition. This calculator therefore reports MAWP for the selected nominal plate at your design temperature, using t(corroded) = t(nominal) − CA in the inverted UG-27/UG-32 formulas. The nameplate MAWP on a real vessel is certified by the manufacturer and may also be limited by nozzles, flanges, or the relief valve set pressure.

The allowable stress tables in this tool stop at 650°F (343°C). Above that temperature, carbon and low-alloy steels enter the time-dependent (creep) regime, where the allowable stress depends on design life and the values must be read from the latest ASME II-D edition — interpolation or extrapolation from public republished tables is not valid engineering practice. For stainless steels the table extends further; for Cr-Mo steels like SA-387 Gr.11, check II-D directly.

No — external pressure design is deliberately out of scope. Under vacuum, a thin cylinder fails by shell buckling (collapse), not yielding, so thickness follows UG-28 using geometry factors L/Do and Do/t with material curves from Section II Part D Subpart 3, plus stiffening-ring design if needed. If your vessel sees full vacuum or jacket pressure, complete a separate UG-28 calculation; a conservative first estimate is to design the shell for 15 psi external pressure.

Use it for scoping, budgeting, and checking fabricator quotes — not as the design of record. Code-stamped vessels must be designed and certified under VIII-1 (or VIII-2) by the manufacturer and reviewed by an Authorized Inspector (AI); the U-stamp is only applied after hydro/pneumatic test and data report sign-off. This tool implements UG-27/UG-32 faithfully but does not handle nozzles (UG-37), fatigue, wind/seismic, MDMT impact (UCS-66), or external pressure.

The 6% knuckle radius of a standard flanged-and-dished head concentrates bending stress at the knuckle, and the UG-32(e) stress multiplier M = ¼(3+√(L/r)) ≈ 1.77 reflects that. A 2:1 semi-elliptical head blends curvature smoothly, keeps K = 1, and typically needs only half the thickness (or less) of an F&D head for the same duty. That is why 2:1 ellipsoidal heads dominate modern vessel design despite slightly higher forming cost.