CH3080_pressure_vessels.ppt

Chemical Engineering Design
© 2012 G.P. Towler / UOP. For educational use in conjunction with
Towler & SinnottChemical Engineering Design only. Do not copy
Pressure Vessel Design

• A pressure vessel is any vessel
that falls under the definition laid
down in the ASME Boiler and
Pressure Vessel Code, Section
VIII, Rules for the Construction
of Pressure Vessels (ASME
BPV Code Sec. VIII)
• The definition applies to most
process reactors, distillation
columns, separators (flashes
and decanters), pressurized
storage vessels and heat
exchangers
Source: UOP
© 2012 G.P. Towler / UOP. For educational use in conjunction with
Towler & Sinnott Chemical Engineering Design only. Do not copy

Isn’t This Something to Leave to the
Mechanical Engineers?
• Chemical engineers are usually not properly trained or qualified to carry
out detailed mechanical design of vessels. Most mechanical designs are
completed by specialists in later phases of design
But
• The process design engineer needs to understand pressure vessel
design in order to generate good cost estimates (e.g. in Aspen ICARUS)
• Costs can vary discontinuously with vessel design
• A 10C change in temperature could double the vessel cost if it causes a change in code!
• Adding a component could cause a change in metallurgy that would mean moving to a more
expensive code design
• The process engineer will end up specifying the main constraints on the
vessel design: if you don’t know how to do this properly, you can’t really
design anything
©2012 G.P. Towler / UOP. For educationaluse in conjunction with

• Pressure Vessel Design Codes
• Vessel Geometry & Construction
• Strength of Materials
• Vessel Specifications
• Materials of Construction
• Pressure Vessel Design Rules
• Fabrication, Inspection and Testing

ASME Boiler and Pressure Vessel Code
• ASME BPV Code is the legally required standard for
pressure vessel design, fabrication, inspection and
testing in North America
Section
I Rules for construction of power boilers
II Materials
III Nuclear power plant components
IV Rules for construction of heating boilers
V Nondestructive examination
VI Recommended rules for the care and operation of heating boilers
VII Recommended guidelines for the care of power boilers
VIII Rules for the construction of pressure vessels
Division 1
Division 2 Alternative rules
Division 3 Alternative rules for the construction of high pressure vessels
IX Welding and brazing qualifications
X Fiber-reinforced plastic vessels
XI Rules for in service inspection of nuclear power plant components
XII Rules for construction and continued service of transport tanks
Most chemical plant vessels
fall under Sec. VIII D.1 or D.2
Allowable stresses are
given in Sec. II
Often used for bio-reactors

Advantages of Designing to Code
• The Code is a consensus best practice
• It is usually required by law
– Local requirements may vary (particularly overseas), but ASME
code is usually recognized as acceptable
– Always check for local regulations that may require stricter
standards
• Code rules are often applied even for vessels that don’t
require construction to code
– Savings of not following code rules are negligible as vessel
shops are set up to do everything to code

ASME BPV Code Sec. VIII Divisions
Division 1
• Rigorous analysis of local thermal
and fatigue stresses not required
• Safety factor of 3.5 against tensile
failure and 1.25 for 100,000 hour
creep rupture
• Limited to design pressures below
3000 psi (but usually costs more
than Div.2 above about 1500 psi)
Division 2
• Requires more analysis than Div.1,
and more inspection, but allows
thinner walled vessels
• Safety factor of 3.0 against tensile
failure
• Limited to design temperatures less
than 900F (outside creep range)
• More economical for high pressure
vessels, but fewer fabricators
available
• Either Division of the Code is acceptable, but provisions
cannot be mixed and matched

Vessels Specifically Excluded
by ASME BPV Code Sec. VIII Div 1
• Vessels within the scope of other sections of the BPV code. For example, power boilers
(Sec. I), fiber-reinforced plastic vessels (Sec. X) and transport tanks (Sec. XIII).
• Fired process tubular heaters.
• Pressure containers that are integral parts of rotating or reciprocating devices such as
pumps, compressors, turbines or engines.
• Piping systems (which are covered by ASME B31.3 – see Chapter 5).
• Piping components and accessories such as valves, strainers, in-line mixers and spargers.
• Vessels containing water at less than 300 psi (2 MPa) and less than 210ºF (99ºC).
• Hot water storage tanks heated by steam with heat rate less than 0.2 MMBTU/hr (58.6 kW),
water temperature less than 210ºF (99ºC) and volume less than 120 gal (450 liters).
• Vessels having internal pressure less than 15 psi (100 kPa) or greater than 3000 psi (20
MPa).
• Vessels of internal diameter or height less than 6 inches (152 mm).
• Pressure vessels for human occupancy.

ASME Code Stamp Name Plate
• Can only be used if vessel is designed, inspected and tested under the
supervision of a Certified Individual employed by the manufacturer
• The code stamp must be clearly visible on the vessel
W (if arc or gas welded)
RT (if Radio graphed)
HT (if Postweld heat treated)
Name of Manufacturer
psi at °F
Max. Allowable Working Pressure
Min. Design Metal Temperature
Manufacturer’s Serial Number
Year Built
°F at psi
W (if arc or gas welded)
RT (if Radio graphed)
HT (if Postweld heat treated)
psi at °F
Year Built
°F at psi
psi at °F
Year Built
°F at psi

Other Related Codes
• Storage tanks are usually not designed to BPV Code
– API Standard 620, Large low pressure storage tanks, Pressure 0.5 to
15 psig
– API Standard 650, Welded storage tanks, Pressures up to 0.5 psig
• Fittings are covered by other ASME codes
– ASME B16.5, Pipe flanges and flanged fittings
– ASME B16.9, Factory-made wrought buttwelding fittings
– ASME B16.11 Forged fittings, socket welding and threaded
– ASME B16.47, Large diameter steel flanges NPS26 Through NPS60
• Piping is covered by a different ASME code
– ASME B16.3, Process piping
• Heat exchangers have additional codes set by TEMA

Use of Design Codes & Standards
• The latest version of the design code should always be
consulted as regulations change
– Example: new version of ASME BPV Code Sec. VIII Div. 2 will
allow for thinner walls on high pressure vessels
• All the information given in this presentation is from the
2004 edition

Pressure Vessel Shape
• What shape of pressure vessel uses the least amount of
metal to contain a given volume, pressure?
A sphere!
• Why is this shape not more widely used?
– Usually need to have an extended section of constant cross-
section to provide support for vessel internals, trays, distributors,
etc.
– It is much easier to obtain and maintain uniform flow in a
cylindrical bed of catalyst or packing than it is in a non-uniform
cross-section
– A cylinder takes up a lot less plot space for the same volume
– A sphere is more expensive to fabricate

Pressure Vessel Shape
• Most pressure vessels are at least 2:1 cylinders: 3:1 or 4:1
are most common:
• Distillation columns are obviously an exception: diameter
is set by flooding correlations and height by number of
trays
2:1 3:1 4:1
(To scale)

Vessel Size Restrictions
• Diameter gets very expensive if > 13.5 ft. Why?
• Height (length) gets very expensive if > 180 ft. Why?
Roughly 50 cranes can lift > 180 ft
Only 14 can lift > 240 ft
• Vessels that can’t be transported have to be fabricated on site

Vessel Orientation
• Usually vertical
– Easier to distribute fluids across a smaller cross section
– Smaller plot space
• Reasons for using horizontal vessels
– To promote phase separation
• Increased cross section = lower vertical velocity = less entrainment
• Decanters, settling tanks, separators, flash vessels
– To allow internals to be pulled for cleaning
• Heat exchangers

Head (Closure) Designs
• Hemispherical
– Good for high pressures
– Higher internal volume
– Most expensive to form & join to shell
– Half the thickness of the shell
• Ellipsoidal
– Cheaper than hemispherical and less
internal volume
– Depth is half diameter
– Same thickness as shell
– Most common type > 15 bar
• Torispherical
– Part torus, part sphere
– Similar to elliptical, but cheaper to fabricate
– Cheapest for pressures less than 15 bar

Tangent and Weld Lines
• Tangent line is where
curvature begins
• Weld line is where
weld is located
• Usually they are not
the same, as the
head is fabricated to
allow a weld away
from the geometrical
joint

Welded Joints
• Some weld types are not
permitted by ASME BPV Code
• Many other possible variations,
including use of backing strips
and joint reinforcement
• Sec. VIII Div. 1 Part UW has
details of permissible joints,
corners, etc.
• Welds are usually ground
smooth and inspected
– Type of inspection depends on
Code Division
Butt weld
Single fillet
lap weld
Double fillet
lap weld
Double fillet
corner joint
Double welded
butt weld

Gasketed Joints
(a) Full face gasket
(b) Gasket within bolt circle
(c) Spigot and socket
(d) O-ring
• Used when vessel must be opened
frequently for cleaning, inspection, etc.
• Also used for instrument connections
• Not used at high temperatures or
pressures (gaskets fail)
• Higher fugitive emissions than welded
joints

Nozzles
• Vessel needs nozzles for
– Feeds, Products
– Hot &/or cold utilities
– Manways, bursting disks, relief valves
– Instruments
• Pressure, Level, Thermowells
• Sample points
• More nozzles = more cost
• Nozzles are usually on side of vessel, away
from weld lines, usually perpendicular to
shell
• Nozzles may or may not be flanged (as
shown) depending on joint type
• The number & location of nozzles are
usually specified by the process engineer

Nozzle Reinforcement
• Shell is weakened around nozzles, and must also support eccentric
loads from pipes
• Usually weld reinforcing pads to thicken the shell near the nozzle. Area
of reinforcement = or > area of nozzle: see Code requirements

Swaged Vessels
• Vessel does not have to be
constant diameter
• It is sometimes cheaper to make a
vessel with several sections of
different diameter
• Smaller diameters are usually at
the top, for structural reasons
• ASME BPV Code gives rules for
tapered sections

Vessel Supports
• Supports must allow for
thermal expansion in
operation
• Smaller vessels are usually
supported on beams – a
support ring or brackets are
welded to the vessel
• Horizontal vessels often
rest on saddles
• Tall vertical vessels are
often supported using a
skirt rather than legs. Can
you think why?

Vessel Supports
• Note that if the vessel rests on a
beam then the part of the vessel
below the support ring is hanging
and the wall is in tension from the
weight of material in the vessel,
the dead weight of the vessel itself
and the internal pressure
• The part of the vessel above the
support ring is supported and the
wall is in compression from the
dead weight (but probably in
tension from internal pressure)

Jacketed Vessels
• Heating or cooling jackets are
often used for smaller vessels
such as stirred tank reactors
• If the jacket can have higher
pressure than the vessel then
the vessel walls must be
designed for compressive
stresses
– Internal stiffening rings are often
used for vessels subject to
external pressure
– For small vessels the walls are just
made thicker

Vessel Internals
Source: UOP
• Most vessels have at least
some internals
– Distillation trays
– Packing supports
– Distribution grids
– Heating or cooling coils
• These may require support
rings welded to the inside of
the vessel
• The internals & support rings
need to be considered when
calculating vessel weights for
stress analysis

Stress and Strain
• Stress  = force divided by area over which it is applied
– Area = original cross section in a tensile test
– Stress can be applied directly or can result from an applied strain
– Examples: dead weight, internal or external pressure, etc.
• Strain ε = distortion per unit length
– Strain = elongation divided by original length in tensile test
– Strain can be applied directly or can result from an applied stress
– Example: thermal movement relative to fixed supports
F F
Cross-sectional area A
L0
 = F / A
ε = (L – L0)/L0

Typical Stress-Strain Curve
for a Mild Steel

Creep
• At high temperatures, strain can continue to increase
over time under constant load or displacement
– Creep strain = increase in strain at constant load
– Creep relaxation = reduction in stress at constant displacement
• Accumulated creep strain can lead to failure: creep
rupture
Time
Stress
or
Strain
Stress
Strain
Time
Stress
or
Strain
Stress
Strain
Fracture
Low Temp High Temp

Principle Stresses & Maximum Shear
Stress
• For a two-dimensional system the
principal stresses at any point
are:
• The maximum shear stress is half
the algebraic difference between
the principal stresses:
• Compressive stresses are taken
as negative, tensile as positive
Normal stresses x, y
Shear stress τxy
τxy
x
y
x
y
τxy
1, 2 = ½(x+ y) ± ½[(y - x)2 + 4τxy
2]
Maximum shear stress = ½(1 - 2)
For design purposes, often just use 1 - 2

Failure of Materials
Failure of materials under combined tensile and shear stresses is not simple to
predict. Several theories have been proposed:
• Maximum Principal Stress Theory
– Component fails when one of the principal stresses exceeds the value that causes
failure in simple tension
• Maximum Shear Stress Theory
– Component fails when maximum shear stress exceeds the shear stress that causes
• Maximum Strain Energy Theory
– Component fails when strain energy per unit volume exceeds the value that causes
• BPV Code gives values for maximum allowable stress for different materials as a
function of temperature, incorporating a safety factor relative to the stress that
causes failure (ASME BPV Code Sec. II)
• Failure in compression is by buckling, which is much harder to predict than
tensile failure. The procedure in the Code is iterative. This should definitely be
left to a specialist

Loads Causing Stresses on Pressure
Vessel Walls
• Internal or external pressure
• Dead weight of vessel
• Weight of contents under normal
or upset conditions
• Weight of contents during
hydraulic testing
• Weight of internals
• Weight of attached equipment
(piping, decks, ladders, etc)
• Stresses at geometric
discontinuities
• Bending moments due to supports
• Thermal expansion, differential
thermal expansion
• Cyclic loads due to pressure or
temperature changes
• Wind & snow loads
• Seismic loads
• Residual stresses from
manufacture
• Loads due to friction (solids flow)
All these must be combined to determine principal stresses

Example: Wind Load
• Wind exerts a pressure on one side
of the vessel
• Resulting force acts like a uniform
beam load and exerts a bending
moment on vessel
• Windward wall is placed in tension,
leeward in compression
• Vortex shedding can cause vibration
– Hence spirals on chimneys
– Usually not needed for columns due to
ladders, pipes, decks, etc.
Wind
Bending
moment

Thin Cylinder Subject to Internal
Pressure
• Forces due to internal pressure are balanced by
shear stresses in wall
• Horizontal section:
• Vertical section:
• Similar equations can be derived for other
geometries such as heads (see Ch 13)
H
L
Longitudinal stress, L
Hoop stress, H
Inside diameter, D
Wall thickness, t
t
D
P
t
D
D
P
L
L
4
4
2






t
D
P
t
h
D
h
P
H
H
2
2




Height,
h

Vessel Specifications Set By the
Process Engineer
• The process engineer will usually specify the following
parameters based on process requirements:
– Vessel size and shape (volume, L and D)
– Vessel orientation and elevation
– Maximum and minimum design pressure
– Maximum and minimum design temperature
– Number of nozzles needed (& location)
– Vessel internals
And often also:
– Material of construction
– Corrosion allowance
• There is often a lot of dialogue with the mechanical
engineer to set the final specifications

Design Pressure
• Normal operating pressure
• The pressure at which you expect the process to usually be operated
• Maximum operating pressure
• The highest pressure expected including upset conditions such as startup,
shutdown, emergency shutdown
• Design pressure
• Maximum operating pressure plus a safety margin
• Margin is typically 10% of maximum operating pressure or 25 psi, whichever is
greater
• Usually specify pressure at top of vessel, where relief valve is located
• The BPV Code Sec. VIII Div. 1 doesn’t say much on how
to set the design pressure
• “..a pressure vessel shall be designed for at least the most severe condition of
coincident pressure and temperature expected in normal operation.”

Design for Vacuum
• The minimum internal pressure a vessel can experience is
full vacuum (-14.7 psig)
• Vacuum can be caused by:
– Intentional process operation under vacuum (including start-up and
shutdown)
– Cooling down a vessel that contains a condensable vapor
– Pumping out or draining contents without allowing enough vapor to
enter
– Operator error
• Vacuum puts vessel walls into compressive stress
• What happens if vessel is not designed for vacuum
conditions?

Vessel Subjected to Excess Vacuum
• Normal practice is to design for vacuum if it can be
expected to occur

Design Temperatures
• Maximum:
– Highest mean metal temperature expected in operation, including
transient conditions, plus a margin
– Margin is typically plus 50F
• Minimum
– Lowest mean metal temperature expected in operation, including
transient conditions, upsets, auto-refrigeration, climatic conditions,
anything else that could cause cooling, minus a margin
– Margin is typically -25F
– MDMT: minimum design metal temperature is important as metals
can become brittle at low temperatures
• Designer should allow for possible failure of upstream
equipment (e.g., loss of coolant on upstream cooler)

Design Temperature Considerations
• Due to creep, maximum allowable stress drops off
rapidly at higher temperatures
– Forces designer to use more expensive alloys
• BPV Code Sec. VIII Div.2 cannot be applied for design
temperatures > 900F (no creep safety factor in Div.2)
• The Code allows design of vessels with different
temperature zones
– Very useful for high temperature vessels
– Not usually applied to medium temperature vessels such as heat
exchangers, distillation columns

Design Temperature & Pressure
Exercise 1
• What is the design
pressure?
120 + 25 = 145 psig
• What is the design
temperature?
340 + 50 = 390F
100 psig
180 F
120 psig
340 F

Design Temperature & Pressure
Exercise 2
• What is the shell-side
design pressure?
588 + 58 = 646 psig
• What is the tube-side
design temperature?
482 + 50 = 532F
Oil
400 psig
120 F
390 psig
450 F
Steam
40 barg
482 F

Materials Selection Criteria
• Safety
– Material must have sufficient strength at design conditions
– Material must be able to withstand variation (or cycling) in
process conditions
– Material must have sufficient corrosion resistance to survive in
service between inspection intervals
• Ease of fabrication
• Availability in standard sizes (plates, sections, tubes)
• Cost
– Includes initial cost and cost of periodic replacement

Commonly Used Materials
• Steels
– Carbon steel, Killed carbon steel – cheap, widely available
– Low chrome alloys (<9% Cr) – better corrosion resistance than CS, KCS
– Stainless steels:
• 304 – cheapest austenitic stainless steel
• 316 – better corrosion resistance than 304, more expensive
• 410
• Nickel Alloys
– Inconel, Incolloy – high temperature oxidizing environments
– Monel, Hastelloy – expensive, but high corrosion resistance, used for
strong acids
• Other metals such as aluminum and titanium are used for special
applications. Fiber reinforced plastics are used for some low
temperature & pressure applications. See Ch 7 for more details

Relative Cost of Metals
• The maximum allowable stress values are at 40ºC (100ºF) and are taken from
ASME BPV Code Sec. II Part D. The code should be consulted for values at
other temperatures. Several other grades exist for most of the materials listed.
• Finished vessel relative costs are not the same as materials relative costs as
vessel cost also includes manufacturing costs, labor and fabricator’s profit
Metal Type or grade Price Max allowable stress Relative cost rating
($/lb) (ksi = 1000 psi)
Carbon steel A-285 0.27 12.9 1
Austenitic stainless steel 304 0.90 20 2.2
316 1.64 20 4
Aluminum alloy A03560 1.27 8.6 2.4
Copper C10400 3.34 6.7 27
Nickel 99%Ni 8.75 10 48
Incoloy N08800 3.05 20 7.5
Monel N04400 6.76 18.7 20
Titanium R50250 9.62 10 27

Corrosion Allowance
• Wall thicknesses calculated using BPV Code equations
are for the fully corroded state
• Usually add a corrosion allowance of 1/16” to 3/16” (1.5
to 5 mm)
• Smaller corrosion allowances are used for heat transfer
equipment, where wall thickness can affect heat transfer

Determining Wall Thickness
• Under ASME BPV Code Sec. VIII D.1, minimum wall
thickness is 1/16” (1.5mm) with no corrosion allowance
• Most pressure vessels require much thicker walls to
withstand governing load
– High pressure vessels: internal pressure usually governs
– Thickness required to resist vacuum usually governs for lower
pressure vessels
– For vessels designed for low pressure, no vacuum, then analysis
of principal stresses may be needed
– Usual procedure is to design for internal pressure (or vacuum),
round up to nearest available standard size and then check for
other loads

Design for Internal Pressure
• ASME BPV Code Sec. VIII D.1 specifies using the larger
of the shell thicknesses calculated
– For hoop stress
– or for longitudinal stress
• Values of S are tabulated in ASME BPV Code Sec.II for
different materials as function of temperature
i
i
i
P
SE
D
P
t
2
.
1
2 

i
i
i
P
SE
D
P
t
8
.
0
4 

S is the maximum allowable stress
E is the welded joint efficiency

Some Maximum Allowable Stresses
Under ASME BPV Code Sec. VIII D.1, Taken From Sec. II Part D
Material Grade Min Tensile Min Yield Maximum Maximum allowable stress at temperature F
strength strength temperature (ksi = 1000 psi)
(ksi) (ksi) (ºF) 100 300 500 700 900
Carbon steel A285 45 24 900 12.9 12.9 12.9 11.5 5.9
Gr A
Killed carbon A515 60 32 1000 17.1 17.1 17.1 14.3 5.9
Steel Gr 60
Low alloy steel A387 60 30 1200 17.1 16.6 16.6 16.6 13.6
1 ¼ Cr, ½ Mo, Si Gr 22
Stainless steel 410 65 30 1200 18.6 17.8 17.2 16.2 12.3
13 Cr
Stainless steel 304 75 30 1500 20.0 15.0 12.9 11.7 10.8
18 Cr, 8 Ni
Stainless steel 347 75 30 1500 20.0 17.1 15.0 13.8 13.4
18 Cr, 10 Ni, Cb
Stainless steel 321 75 30 1500 20.0 16.5 14.3 13.0 12.3
18 Cr, 10 Ni, Ti
Stainless steel 316 75 30 1500 20.0 15.6 13.3 12.1 11.5
16 Cr, 12 Ni, 2 Mo

Welded Joint Categories
ASME BPV Code has four categories of welds:
A. Longitudinal or spiral welds in the main shell, necks or nozzles, or
circumferential welds connecting hemispherical heads to the main
shell, necks or nozzles.
B. Circumferential welds in the main shell, necks or nozzles or
connecting a formed head other than hemispherical.
C. Welds connecting flanges, tubesheets or flat heads to the main
shell, a formed head, neck or nozzle.
D. Welds connecting communicating chambers or nozzles to the main
shell, to heads or to necks.

Welded Joint Efficiencies Allowed Under
ASME BPV Code Sec. VIII D.1
Joint Description Joint Category Degree of Radiographic Examination
Full Spot None
Double-welded butt joint A, B, C, D 1.0 0.85 0.70
or equivalent
Single-welded butt joint A, B, C, D 0.9 0.8 0.65
with backing strip
Single-welded butt joint A, B, C NA NA 0.60
without backing strip
Double full fillet lap A, B, C NA NA 0.55
joint
Single full fillet lap B, C NA NA 0.50
joint with plug welds
Single full fillet lap A, B NA NA 0.45
joint without plug welds

Closures Subject to Internal Pressure
• Hemispherical heads
• Ellipsoidal heads
• Torispherical heads
i
i
i
P
SE
D
P
t
4
.
0
4 

i
i
i
P
SE
D
P
t
2
.
0
2 

i
c
i
P
SE
R
P
t
1
.
0
885
.
0


Rc is the crown radius: see Ch 13

Example
• What is the wall thickness required for a 10ft diameter 304
stainless steel vessel with design pressure 500 psi and
design temperature 700F?
• From the table, S = 11700 psi
• Assume double-welded butt joint with spot radiography, E = 0.85
• For hoop stress
• For longitudinal stress
• So hoop stress governs, choose t = 3.25 or 3.5 inches,
depending on what’s readily available as plate stock
inches
11
.
3
500
2
.
1
85
.
0
11700
2
12
10
500
2
.
1
2










i
i
i
P
SE
D
P
t
inches
49
.
1
500
8
.
0
85
.
0
11700
4
12
10
500
8
.
0
4










i
i
i
P
SE
D
P
t

Software for Pressure Vessel Design
• Rules for external pressure, combined loads are more
complex
• Design methods and maximum allowable stresses are
coded into software used by specialist designers, such as:
• COMPRESS (Codeware Inc.) has free demo version
http://www.codeware.com/support/tutorials/compress_video_tutorial.html
• Pressure Vessel Suite (Computer Engineering Inc.)
• PVElite and CodeCalc (COADE Inc.)
• Simple ASME BPV Code Sec. VIII D.1 methods are
available in Aspen ICARUS
• Good enough for an initial cost estimate if the process engineer puts in realistic
vessel specifications
• Useful for checking to see if changes to specifications give cost discontinuities
• Not good enough for detailed vessel design

Example
• What is the cost of a 10ft diameter, 100ft long 304 stainless
steel vessel with design pressure 500 psi and design
temperature 700F?

Example
In Aspen ICARUS, if we just enter the
dimensions and material:

Example
Total cost $575,500

Example: With More Complete Specifications

Example: With More Complete Specifications
Total cost is now
$1,814,400

Example: ICARUS Results
• ICARUS finds a wall thickness of 3.308”, based on 3.183”
for hoop stress and 0.125” corrosion allowance
• Vacuum design thickness is 1.05”, so internal pressure is
governing
• Not clear why the ICARUS hoop stress calculation comes
out different, but possibly due to considering combined loads
under wind or seismic conditions
• Note that proper specification of vessel design changed cost
by factor 3.15
• Results are good enough for preliminary costing, but not for
mechanical design

Vessel Manufacture
• Shell is usually made by rolling plate and then welding
along a seam:
– Difficult to form small diameters or thick shells by this method
– Long vessels are usually made in 8’ sections and butt welded
• Thicker vessels are made by more expensive drum
forging – direct from ingots
• Closures are usually forged
– Hence restricted to increments of 6” in diameter
• Nozzles, support rings etc. are welded on to shell and
heads

Post Weld Heat Treating (PWHT)
• Forming and joining (welding) can leave residual
stresses in the metal
• Post-weld heat treatment is used to relax these stresses
• Guidelines for PWHT are given in the ASME BPV Code
Sec. VIII D.1 Part UW-40
• PWHT requirements depend on material and thickness
at weld:
- Over 38mm for carbon steel
- Over 16mm for low alloy

CH3080_pressure_vessels.ppt

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