Tank Shell Plate Thickness Calculator

A cylindrical shell under internal pressure carries that load mainly as hoop tension in the plate. For a thin wall the balance is simply pressure times radius equals stress times thickness, which rearranges to t = P R / S. The working form used here, t = P R / (S E - 0.6 P), adds two corrections. The joint efficiency E, typically between 0.7 and 1.0, derates the plate to account for the weld seam being the weakest line through the shell, with 1.0 permitted only when the seam is fully radiographed and lower values such as 0.85 or 0.70 applied for spot or no examination. The 0.6 P term shifts the reference from the inside surface toward the mid-wall, approximating the exact thick-cylinder solution and keeping the formula honest as walls get heavier; it is valid only in the thin-wall regime, roughly while thickness stays under half the radius or pressure under about 0.385 times S times E, beyond which a thick-wall treatment is required. Corrosion allowance, commonly 1.5 mm for clean service, 3 mm for mildly corrosive product, and 6 mm or more for aggressive duty, is added afterward because it is sacrificial metal not counted as carrying load. Units and applicability are the practical traps. Allowable stress in this calculation is in megapascals and pressure in kilopascals, so 137.9 MPa corresponds to the familiar 20,000 psi allowable used for common carbon plate grades; mixing the two scales by a factor of a thousand. More importantly, a large atmospheric storage tank is not really a pressure vessel: its shell is loaded by the hydrostatic head of the stored product, so API 650 sizes each shell course by the one-foot method using the product height above a point one foot below the course seam, which is why tanks are built with progressively thicker plate toward the bottom. API 650 also imposes minimum thicknesses independent of any calculated result, scaling with tank diameter from 5 mm for smaller tanks up to 10 mm for the largest, because thin plate cannot survive handling, erection, and wind loading regardless of what the stress equation permits. Wind girders, seismic overturning, roof loading, and foundation settlement are separate checks this formula does not address.

Pressure vessel rules were written in response to explosions. Through the nineteenth century boiler failures were routine and often catastrophic, and design was left to the builder's judgment. In the United States the turning point came on 20 March 1905, when the boiler at the Grover Shoe Factory in Brockton, Massachusetts exploded, destroying the building and killing 58 people. Massachusetts responded with a mandatory boiler law, other states followed with conflicting rules, and the resulting patchwork pushed manufacturers to seek a single national standard. The American Society of Mechanical Engineers formed a committee in 1911 and issued the first Boiler and Pressure Vessel Code in 1914 and 1915. Rules for unfired pressure vessels, which became Section VIII, followed in the 1920s, and the allowable stress plus joint efficiency framework used in this calculation descends directly from that work. The underlying mechanics was older. Peter Barlow's early nineteenth century work on the strength of cylinders gave the simple hoop stress relation still known as Barlow's formula, and Gabriel Lame published the exact elasticity solution for thick-walled cylinders in the 1830s, which is what the 0.6 P correction approximates. Large welded oil storage tanks got their own track because riveted construction gave way to welding in the interwar period: the American Petroleum Institute issued specifications for welded oil storage tanks from the 1930s, and that lineage became API Standard 650, adopted in the early 1960s and revised through many editions since. API 650 kept the shell-course logic tied to hydrostatic head rather than design pressure, preserving a clear division of labor that persists today, with ASME Section VIII governing pressure vessels and API 650 governing atmospheric and low-pressure storage tanks.