Torispherical Head Thickness (ASME)
Calculate the minimum thickness of a torispherical (standard flanged-and-dished) pressure-vessel head, t = (0.885·P·L) ÷ (S·E − 0.1·P), from the internal pressure P (MPa), the spherical crown radius L (mm), the allowable stress S (MPa) and the joint efficiency E. The TORISPHERICAL head is the most COMMON and economical head type in medium-pressure vessels (and universal in shallow tanks): it combines a central spherical crown (radius L) with a toroidal knuckle transition at the edge, joining the cylindrical shell — a form easier and cheaper to stamp than the hemispherical, and more compact (lower height). The 0.885 factor and formula hold for the standard ASME geometry with L ≈ D (crown radius equal to diameter) and the knuckle radius of 6% of the diameter. The price of the economy is a GREATER thickness than the hemispherical (the toroidal transition concentrates stress) and a critical knuckle region, where high bending stresses can arise. The torispherical head is the practical 'middle ground' between the costly hemispherical and the flat (which needs enormous thicknesses). This formula is essential in designing vessels with this head type. Enter the pressure, crown radius, allowable stress and joint efficiency.
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Torispherical head thickness (ASME)
The minimum thickness of a torispherical head (the standard flanged-and-dished type) on a pressure vessel is t = (0.885·P·L) ÷ (S·E − 0.1·P), computed from the internal pressure P, the crown radius of the spherical cap L, the allowable stress S and the joint efficiency E. The torispherical head is the most common and economical head type on medium-pressure vessels: it blends a central spherical crown (of radius L) with a toroidal knuckle at the rim that joins it to the shell — a shape that is easier and cheaper to make by pressing than a hemispherical head, and more compact as well. The 0.885 factor and the formula hold for the standard ASME geometry with L ≈ D and a knuckle radius equal to 6% of the diameter. The price of that economy is a greater thickness than the hemispherical head calls for (the toroidal knuckle concentrates stress) plus a critical region at the knuckle itself, where high bending stresses can build up. The torispherical head is the practical 'middle ground' between the expensive hemispherical head and the flat one (which demands enormous thicknesses). This formula is essential when designing vessels that use this head type. Enter the pressure, the crown radius, the allowable stress and the joint efficiency.
Related Tools
Hemispherical Head Thickness (ASME)
Calculate the minimum thickness of a pressure-vessel hemispherical head by the ASME Section VIII formula, t = (P·r) ÷ (2·S·E − 0.2·P), from the internal pressure P (MPa), internal radius r (mm), allowable stress S (MPa) and joint efficiency E. Heads close the ends of a pressure vessel's cylindrical shell, and their shape is decisive for structural efficiency. The HEMISPHERICAL (half-sphere) head is the MOST EFFICIENT of all: since the sphere distributes pressure equally in all directions (uniform membrane stress), the hemispherical head needs only about HALF the thickness of the cylindrical shell of the same radius and pressure (compare the '2·S·E' in the denominator with the shell's 'S·E'). So it is the choice for high-pressure vessels. The drawbacks are costlier fabrication and greater height (more space). For moderate pressures and costs, elliptical (2:1) or torispherical heads, intermediate, are used. The head-type choice is a trade-off among thickness/material (cost), space and fabrication ease. This formula is fundamental in the complete vessel design, combining shell and heads. Enter the pressure, internal radius, allowable stress and joint efficiency.
Cylindrical Shell Thickness (ASME)
Calculate the minimum wall thickness of a pressure-vessel cylindrical shell by the ASME Section VIII Division 1 formula, t = (P·r) ÷ (S·E − 0.6·P), from the internal design pressure P (MPa), the internal radius r (mm), the material allowable stress S (MPa) and the welded-joint efficiency E (0-1). The pressure vessel — used in boilers, chemical reactors, heat exchangers, compressed-air and LPG tanks, autoclaves — is a CRITICAL safety component: a failure under pressure can be explosive and catastrophic. So its design is rigorously codified, the ASME BPVC (Boiler and Pressure Vessel Code) being the world's most used. This formula gives the minimum cylindrical-shell thickness to safely resist the circumferential (hoop) stress. The '−0.6·P' term refines the thin-wall formula for moderately thick walls. The joint efficiency E (0.70 to 1.0, per weld type and radiographic-inspection degree) penalizes strength at the welded region — fully radiographed welds have E=1.0, uninspected welds lower E. The corrosion allowance is added to the calculated thickness. This is the central pressure-vessel design calculation, and underestimating is inadmissible. Enter the design pressure, internal radius, allowable stress and joint efficiency.
Hydrostatic Test Pressure
Calculate the hydrostatic test pressure of a pressure vessel by the (simplified) ASME rule, P_test = 1.3 · MAWP, from the maximum allowable working pressure MAWP (MPa). Before entering service (and periodically, at revalidations), every pressure vessel undergoes a HYDROSTATIC TEST: it is filled with WATER (not gas!) and pressurized ABOVE the operating pressure, to verify structural integrity and tightness before entrusting it with a hazardous fluid. ASME VIII Div. 1 (rule UG-99) requires a test pressure of 1.3 times MAWP (corrected by the allowable-stress ratio at test and design temperatures, simplified here). Using WATER is a fundamental safety matter: water is practically incompressible, so it stores very little energy when pressurized — if the vessel ruptures during the test, the failure is localized and relatively safe (it leaks, not explodes); whereas a compressed gas stores enormous energy and a rupture would be EXPLOSIVE, possibly lethal. The 1.3×MAWP test subjects the vessel to higher-than-operating stresses, revealing defects (cracks, bad welds, insufficient thickness) with margin, without reaching general yielding. Passing the hydrostatic test is a condition for the vessel's certification and operation. Enter the MAWP.
Hemispherical Head MAWP
Calculate the maximum allowable working pressure (MAWP) of a pressure-vessel hemispherical head, MAWP = (2·S·E·t) ÷ (r + 0.2·t), from the allowable stress S (MPa), the joint efficiency E, the available thickness t (mm) and the internal radius r (mm). Each pressure-vessel component has its own MAWP — the maximum pressure IT withstands with its available thickness — and the WHOLE vessel's MAWP is the SMALLEST among all its components' MAWPs (shell, heads, nozzles), since the vessel is as strong as its weakest component. This formula gives the hemispherical head's MAWP, the inverse of that head's thickness calculation. The factor 2 in the numerator (versus 1 in the shell) reflects the greater efficiency of the spherical form: for the same thickness, radius and material, the hemispherical head withstands about DOUBLE the cylindrical shell's pressure. So in a well-designed vessel with hemispherical heads, the cylindrical SHELL is usually the component governing the vessel's MAWP (the weakest), and the heads have margin. Comparing the components' MAWPs identifies the weakest link and guides repairs and reinforcements. Recomputing MAWP with the remaining thickness measured at inspection is part of vessel integrity management. Enter the allowable stress, efficiency, thickness and radius.
Thickness with Corrosion Allowance
Calculate the total thickness to specify for a pressure-vessel component including the corrosion allowance, t_total = t_calculated + CA, from the minimum pressure-calculated thickness t_calculated (mm) and the corrosion allowance CA (mm). The thickness from the ASME formulas is the MINIMUM needed to resist pressure — but the vessel will operate for DECADES, and corrosion (and erosion) will consume wall material over time. If the vessel were made exactly at the minimum thickness, the first corrosion would already leave it below safe. So a CORROSION ALLOWANCE (CA) is added — a 'sacrificial' over-thickness, typically 1.5 to 6 mm, sized for the expected corrosion rate times the design life (e.g., 0.1 mm/year × 25 years = 2.5 mm). Thus the thickness specified for fabrication is the structural minimum plus the corrosion allowance. Over life, inspection (by ultrasound) measures the REMAINING thickness; when corrosion consumes the whole allowance and the thickness approaches the structural minimum, the vessel must be repaired or retired. The corrosion allowance is like a 'life reserve' built into the wall. Enter the calculated thickness and the corrosion allowance.
Thickness/Diameter Ratio (Thin Wall)
Calculate a pressure vessel's thickness/diameter ratio, t/D, from the wall thickness t and the diameter D (same unit). This ratio is the criterion deciding whether a vessel can be treated as THIN-walled or needs THICK-walled (Lamé) theory. The distinction is fundamental because the formulas change: in THIN walls (rule of thumb t/D < 0.05, or t/r < 0.1), stress is practically UNIFORM across the thickness, and the simple membrane formulas hold (σ = P·r/t for hoop) — the case of the vast majority of vessels, pipes and tanks. In THICK walls (larger t/D, as in very-high-pressure vessels — hydrogenation reactors, gun barrels, high-pressure hydraulic tubing), stress VARIES strongly across the thickness (maximum at the inner surface, decreasing outward), and the simple formulas dangerously underestimate the inner peak stress — Lamé's equations must be used. Checking the t/D ratio is thus the first step in choosing the correct calculation theory. Vessels with t/D above ~0.1 require thick-wall analysis. This simple check avoids the serious error of applying thin-wall formulas to a thick vessel. Enter the thickness and the diameter.
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