Design codes: IS 2825 and ASME Section VIII

Why it matters

A pressure vessel stores enormous energy; when one fails it can kill people far from the plant. Design codes turn a century of failures into rules on materials, thickness formulas, welding, inspection and testing, and in India and most export projects a vessel cannot be legally operated or insured unless it is built and stamped to a recognised code. A chemical engineer writing a vessel data sheet must speak the code's language: design pressure, design temperature, allowable stress, joint efficiency and test pressure.

Key ideas

IS 2825 (Code for unfired pressure vessels). The Indian Standard for fusion-welded unfired vessels. It classifies vessels into Class 1, 2 and 3 according to service (lethal or flammable contents, pressure, temperature, thickness). The class decides the extent of radiography, the post-weld heat treatment needed and therefore the weld joint efficiency that may be used in the thickness formula. It gives formulas for cylindrical and spherical shells, heads, openings and flanges, and rules for fabrication, inspection and pressure testing. Many Indian university syllabi and design textbooks use its formulas.

ASME Boiler and Pressure Vessel Code, Section VIII. The most widely used code in the world, in three divisions:

• Division 1: design by rule. Simple formulas based on membrane stress, conservative allowable stresses, the workhorse for most refinery and chemical-plant vessels.
• Division 2: alternative rules with design by analysis permitted. Allowable stresses are higher (lower safety factor) but analysis, materials testing and inspection are more demanding, so it pays for thick, expensive vessels.
• Division 3: very high pressures (roughly above 70 MPa), where thick-wall theory, fatigue and fracture mechanics govern.

Design pressure and temperature. The design pressure is set above the maximum expected operating pressure by a margin (commonly 5–10 % or a fixed amount, set by the project specification) and at least equal to the relief-valve set pressure. Add the static head of liquid for the lower parts of tall vessels. The design temperature is the highest (or lowest) metal temperature expected, plus a margin. The allowable stress is read at the design temperature.

Allowable (design) stress. Codes divide the tensile strength, the yield strength and (at high temperature) creep-rupture strength each by its own factor and take the lowest. For example ASME VIII Division 1 uses roughly the lower of tensile strength/3.5 and two-thirds of yield strength at temperature, while Division 2 uses a smaller factor on tensile strength. IS 2825 has its own factors. Never compute f yourself in a design: take it from the code's tables for the exact material grade and temperature.

Weld joint efficiency (J or E). Welds may contain defects, so the allowable stress across a weld is reduced by a factor ≤ 1 that depends on joint type and how much of it is radiographed. For a double-welded butt joint the familiar values are 1.0 for full radiography, 0.85 for spot radiography and 0.70 with none; confirm the value for your joint and class from the code book.

MAWP. The maximum allowable working pressure is the highest pressure the vessel can take at the design temperature in its corroded condition, calculated from the actual (nominal minus corrosion allowance) thickness. It is often a little above the design pressure because plates are rounded up.

Pressure testing. Every vessel is hydrostatically tested before service at a pressure above the design pressure, so that it is proven with a margin. ASME VIII Division 1 currently requires about 1.3 × MAWP × (allowable stress at test temperature / allowable stress at design temperature); IS 2825 has a similar rule. Pneumatic testing is used only where water cannot be tolerated, with extra precautions, because a gas test stores far more energy.

Formulas

t = P·Di / (2·f·J − P) (IS 2825, cylindrical shell, internal pressure)

• t = minimum thickness, excluding corrosion allowance (mm); P = design pressure (MPa, gauge); Di = inside diameter (mm); f = allowable stress at design temperature (MPa); J = weld joint efficiency (–).

t = P·R / (S·E − 0.6·P) (ASME VIII Div. 1, circumferential (hoop) stress)

• R = inside radius (mm); S = allowable stress (MPa); E = joint efficiency (–). Valid while t ≤ R/2, i.e. P ≤ 0.385·S·E.

t = P·R / (2·S·E + 0.4·P) (ASME VIII Div. 1, longitudinal stress)

• Usually about half the hoop thickness, so hoop stress governs a plain cylinder.

MAWP = S·E·t_c / (R + 0.6·t_c)

• t_c = corroded thickness = nominal thickness − corrosion allowance (mm). This is the hoop formula solved for pressure.

P_test ≈ 1.3 × MAWP × (S_test / S_design) (ASME VIII Div. 1 hydrotest; confirm in the current edition)

Worked examples

Example 1 (standard): IS 2825 shell thickness Given: P = 1.2 MPa, Di = 1600 mm, f = 130 MPa at design temperature, J = 0.85 (spot radiography), CA = 2 mm, plates in even millimetres. f is given data; normally take it from your code book.

1. t = P·Di / (2·f·J − P)
2. t = 1.2 × 1600 / (2 × 130 × 0.85 − 1.2) = 1920 / 219.8 = 8.74 mm
3. Add CA: 8.74 + 2 = 10.74 mm
4. Round up: nominal thickness = 12 mm.

Example 2 (GATE level): is an existing vessel good for a new duty? Given: an ASME VIII Div. 1 vessel, inside diameter 2000 mm (R = 1000 mm), nominal shell thickness 16 mm, CA = 3 mm, S = 138 MPa, E = 1.0. A process change needs a design pressure of 2.0 MPa.

1. Required hoop thickness: t = P·R / (S·E − 0.6·P) = 2.0 × 1000 / (138 − 1.2) = 2000 / 136.8 = 14.62 mm
2. Validity: 0.385 × S × E = 53.1 MPa ≫ 2.0 MPa, so the thin-shell formula applies.
3. Longitudinal check: t = 2.0 × 1000 / (2 × 138 + 0.8) = 7.23 mm — hoop governs.
4. Available corroded thickness: t_c = 16 − 3 = 13 mm < 14.62 mm.
5. MAWP: MAWP = 138 × 1.0 × 13 / (1000 + 0.6 × 13) = 1794 / 1007.8 = 1.78 MPa.
6. MAWP = 1.78 MPa < 2.0 MPa: the vessel cannot be re-rated without a thicker shell, a lower corrosion allowance justified by inspection data, or a lower design pressure.

Common mistakes

• Using operating pressure instead of design pressure, or forgetting the liquid static head at the bottom of a tall vessel.
• Using absolute pressure: the formulas use gauge pressure (the difference across the wall).
• Mixing radius and diameter forms: the IS form uses Di and 2fJ; the ASME form uses R and S·E. Do not combine pieces of both.
• Taking tensile strength or room-temperature allowable stress instead of the code value at design temperature.
• Calculating MAWP from the nominal thickness instead of the corroded thickness.
• Assuming J = 1 without specifying full radiography.

For GATE CH

Expect thickness or allowable-pressure numericals with a given allowable stress, joint efficiency and corrosion allowance; questions on which stress (hoop or longitudinal) governs; and conceptual items on joint efficiency, design versus operating pressure, and why hydrostatic rather than pneumatic testing is preferred. Practise rearranging the thickness formula for pressure and checking the result against the plate actually available.

Quick check

1. Which ASME Section VIII division is design by rule with the simplest formulas?
2. Why is the allowable stress multiplied by a joint efficiency?
3. A shell has R = 500 mm, S = 120 MPa, E = 1, corroded thickness 6 mm. What is its MAWP (ASME hoop formula)?
4. Why is a hydrostatic test safer than a pneumatic test at the same pressure?

Answers: 1. Division 1. 2. Welds may contain defects; the factor reflects the joint type and the extent of radiography. 3. MAWP = 120 × 6 / (500 + 3.6) = 1.43 MPa. 4. Water is nearly incompressible, so it stores very little energy and a failure leaks rather than explodes.