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INTRODUCTION
Definition:
A pressure vessel is a closed container designed to hold gases or liquids at a pressure different from the ambient pressure. The end caps fitted to the cylindrical body are called heads.
The legal definition of pressure vessel varies from country to country, but often involves the maximum safe pressure (may need to be above half a bar) that the vessel is designed for and the pressure-volume product, particularly of the gaseous part (in some cases an incompressible liquid portion can be excluded as it does not contribute to the potential energy stored in the vessel.) In the United States, the rules for pressure vessels are contained in the American society of mechanical engineers Boiler and Pressure Vessel Code.
Uses
A pressure tank connected to a water well and domestic hot water system Pressure vessels are used in a variety of applications. These include the industry and the private sector. They appear in these sectors respectively as industrial compressed air receivers and domestic hot water storage tanks, other examples of pressure vessels are: diving cylinder, recompression chamber, distillation towers, autoclaves and many other vessels in mining or oil refineries and petrochemical plants, nuclear reactor vessel, habitat of a space ship, habitat of a submarine, pneumatic reservoir, hydraulic reservoir under pressure, rail vehicle airbrake reservoir, road vehicle airbrake reservoir and storage vessels for liquefied gases such as ammonia, chlorine, propane, butane and LPG.
In the industrial sector, pressure vessels are designed to operate safely at a specific pressure and temperature technically referred to as the "Design Pressure" and "Design Temperature". A vessel that is inadequately designed to handle a high pressure constitutes a very significant safety hazard. Because of that, the design and certification of pressure vessels is governed by design codes such as the ASME Boiler and Pressure Vessel Code in North America, the Pressure Equipment Directive of the EU (PED), Japanese Industrial Standard (JIS), CSA B51 in Canada, AS1210 in Australia and other international standards like Lloyd's, Germanischer Lloyd, Det Norske Veritas, Stoomwezen etc.
Shape of a pressure vessel
Pressure vessels can theoretically be almost any shape, but shapes made of sections of spheres, cylinders and cones are usually employed. More complicated shapes have historically been much harder to analyse for safe operation and are usually far harder to construct.
Theoretically a sphere would be the optimal shape of a pressure vessel. Unfortunately the sphere shape is difficult to manufacture, therefore more expensive, so most of the pressure vessels are cylindrical shape with 2:1 semi elliptical heads or end caps on each end. Smaller pressure vessels are arranged from a pipe and two covers. Disadvantage of these vessels is the fact that larger diameters make them relatively more expensive, so that for example the most economic shape of a 1000 litres, 250 bar (25,000 kPa) pressure vessel might be a diameter of 914.4 mm and a length of 1701.8 mm including the 2:1 semi elliptical domed end caps.
Construction materials
Generally, almost any material with good tensile properties that is chemically stable in the chosen application can be employed.
Many pressure vessels are made of steel. To manufacture a spherical pressure vessel, forged parts would have to be welded together. Some mechanical properties of steel are increased by forging, but welding can sometimes reduce these desirable properties. In case of welding, in order to make the pressure vessel meet international safety standards, carefully selected steel with a high impact resistance & corrosion resistant material should also be used.
Some pressure vessels are made of wound carbon fibre held in place with a polymer. Due to the very high tensile strength of carbon fibre these vessels can be very light, but are much trickier to manufacture.
Other very common materials include polymers such as PET in fizzy drinks containers and copper in plumbing.
Design Standards
Â¢ BS 4994
Â¢ ASME Code Section VIII Division 1
Â¢ ASME Code Section VIII Division 2 Alternative Rule
Â¢ ASME Code Section VIII Division 3 Alternative Rule for Construction of High Pressure Vessel
Â¢ ASME PVHO (Safety Standard for Pressure Vessels for Human Occupancy)
Â¢ PD 5500
Â¢ Stoomwezen
Â¢ AD MerkblÃƒÂ¤tter
Â¢ CODAP
Â¢ AS 1210
ISO 11439
Pressure Vessel has the following basic elements.
Â¢ Cylindrical Shells
Â¢ Elliptical, spherical, Torispherical, conical and flat heads
Â¢ Conical sections (including knuckles)
Â¢ Body flanges
Â¢ Skirts and lug supports
Â¢ Nozzles
Â¢ Dynamic wind analysis
Â¢ Seismic analysis
3.Design Data Sheet
Design code : ASME section VIII div-1, Ed2004up to &add2006
Type of vessel : horizontal
Inside diameter : 610.000mm
W.L to W.L height : 1000.000mm
Design temperature : 45 Ã‚Â°C
MDMT : 0 Ã‚Â°C
Working temp. (min.) : 7 Ã‚Â°C
Working temp. (Max) : 45 Ã‚Â°C
Working pressure : 0.490 Mpa (5.000 kg/cm2)
Design pressure : 0785 Mpa (8.000 kg/cm2)
MAWP : 0.785 Mpa (8.000 kg/cm2)
Product stored : compressed Air (Non Lethal)
Sp. Gravity of product : 0.0013
Max. Liquid level : NOT APPLICABLE
Corrosion allowance : 3.00MM
Heat treatment : YES (dished end)
P.W.H.T. : Nil
Radiography : spot
Hydro test pressure : 1.021 Mpa (10.40 kg/cmÃ‚Â².)
Operating weight : 400.000 kgs (approximate)
Hydro test weight : 600.00 kgs (approximate)
LOADINGS (UG-22):
Sr.NO DESCRIPTION APPLICABILITY DOCUMENTED IN
1 Internal pressure Yes
0.785 Mpa Page 6-1
2 External pressure Nil
3 Static head
(i)normal operating condition
(ii)for hydrostatic test condition
Negligible
5.98E-07 Mpa
4 Weight of vessel in normal/
Hydrostatic test condition Yes Page 20-1
5 Super imposed static reaction
From weight of attachment Nil
6 Internal attachment loading Nil
7 Cyclic & dynamic reaction Nil
8 Wind loading Nil
9 Seismic Yes
(IS1893 ZONE 3) Page 24-1 &
Page 24-2
10 Impact reaction Nil
11 Temperature gradient Nil
12 Deflagration Nil
Note:
Static head of fluids on normal operating condition is negligible.
4.EVALUATION OF MATERIALS
(ASME SEC II,PART D):
COMPONENTS SHELL, HEADS, PADS,
SADDLES & SUPPORTS NOZZLE NECK
(SMILES PIPE)
Full specification SA 516 M Gr.485 SA 106 GrB
p-no 1 1
Group no 2 1
Whether allowed by sec viii div 1,
Table ucs-23(yes/no) Yes Yes
Reference in sec ii part D page no 14 10
Reference in sec ii part D line no 20 5
Design temp. 45 45
Allowable stress at design temp. (Mpa) 138 118
Atmospheric temp. 45 45
Allowable stress at atm. Temp. (Mpa) 138 118
Permissible temp.© 550 550
Sp.cautionary notes G10,S1,T2 G10,S1,T2
Applicable cautionary notes None None
Components Flanges&
Pipe fittings Bolting
Full specification SA 516 M Gr.485 SA 193M Gr.B7
p-no 1 ---
Group no 2 ---
Whether allowed by sec viii div 1,
Table ucs-23(yes/no) Yes Yes
Reference in sec ii part D page no 14 382
Reference in sec ii part D line no 6 33
Design temp. 45 45
Allowable stress at design temp. (Mpa) 138 172
Atmospheric temp. 45 45
Allowable stress at atm. temp. (Mpa) 138 172
Permissible temp.© 550 550
Sp.cautionary notes G10,G35,S1,T2 ----
Applicable cautionary notes None None
Gaskets : Compressed Asbestos (CAF)
Conversion Factors
1 MPa = 10.197 kg/cm2
1 kg/cm2 = 0.0981 MPa
1 N/mm2 = 1 MPa
5. EVALUATION OF DESIGN PRESSURE FOR OPERATING CONDITION ( UG-21 )
Inside diameter of vessel = 610.000 mm
Maximum liquid level = 0.00 mm
Internal Pressure:
Design pressure at top = 0.785 MPa (8.00 kg/cm Ã‚Â²)
Pressure due to static head = 0.000 MPa (0.00 kg/cm Ã‚Â²)
Design pressure at bottom p = 0.785 MPa (8.00 kg/cm Ã‚Â²)
External Pressure:
External Design Pressure = 0.000 MPa (0.000 kg/cm2)
Component 1 2 3 4
MAWP,Mpa Pressure due to static fluid MPa Vacuum correction MPa Design pressure (1+2+3). MPa
Shell 0.785 0.000 0.000 0.785
Top head 0.785 0.000 0.000 0.785
Bottom head 0.785 0.000 0.000 0.785
Nozzles 0.785 0.000 0.000 0.785
All the components have been designed to = MPa
Note: evaluation of loafing during hydrostatic test condition is documented in Pg.NO: 20-1
6. EVALUATION OF JOINT EFFIENCY (BUTT WELDS) (UW-12):
FIG UW-3: Illustration of welded joint locations typical of categories A,B,C & D
Description of joints Joint category Proposed Type No. Proposed NDE Joint Efficiency table UW-12
1.shell longitudinal joints A (1) Spot Ec = 0.85
2. shell to ellipsoidal joints B (2) Spot Ec = 0.85
Ec = joint efficiency to be used in calculation of circumferential stress
El = joint efficiency to be used in calculation of longitudinal stress
7. THICKNESS OF THE SHELL UNDER INTERNAL PRESSURE (UG-27)
The minimum thickness of shells and heads used in compressed air service, steam service, & water service made from materials listed in table ucs-23 shell be 3/32 in. (2.4mm) exclusive of any corrosion allowance.
Provided positive tolerance = 6.00mm
Possible inside diameter (corroded)
= 622.00mm
Inside radius (corroded) R = 311.000mm
C.A = 3.000mm
Internal design pressure at bottom P = 0.785 Mpa
Max. allowable stress S = 138 Mpa
Longitudinal joint efficiency EL = 0.85
Circumferential joint efficiency Ec =0.85
1) Circumferential stress (Longitudinal joint)
0.385SEL = 45.161 Mpa
Design pressure P = 0.785 Mpa
Since P<0.385SEL UG57© (1) can be applied
Min. design thickness of shell,
t = [PR/(SEc-0.6P)]+C.A
= {[0.785*311]/[(138*0.85)-(0.6*0.785)]}+3
= 5.089 mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦.(a)
(2) longitudinal stress (Circumferential joint
1.25 SEc = 146.625 Mpa
Design pressure P = 0.785 Mpa
Since P < 1.25 SEc UG27© (2) can be applied
Min. design thickness of shell,
t = [PR/(2SEL+0.4P)]+C.A
= {[0.785*311]/[(2*138*0.85) +(0.4*0.785)]}+3
t = 4.039 mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦ (b)
(3 )Min. Thickness as per UG 16 (b) (4);
Min. thickness of shell in compressed air service,
= 2.5 mm + C.A
= 2.5 + 3
=5.500 mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦..©
Design Thickness = max. Of (a, b, c)
Design Thickness = 5.500 mm
Provided thickness = 10.00 mm
tr for compensation calculation: (using E=1 as per UG-37)
Minimum required thickness
tr = PR/{SE-0.6P}
= [0.785*311]/[(138*1)-(0.6*0.785)]
tr = 1.774 mm
8 . FORMED HEADS PRESSURE ON CONCAVE SIDE : (UG - 32)
Type of head = 2:1 ELLIPSOIDAL
Provided positive tolerance = 7.000 mm
Positive inside diameter (corroded) D = 623.00 mm
C.A = 3.00mm
Internal design pressure at bottom P = 0.785 Mpa
Max. allowable stress S = 138 Mpa
Joint efficiency E = 0.85
Inside depth measured from T.L h = 155.750 mm
Ratio of the major axis to the minor axis D/2h = 2.000
For D/2h=2.00 from table 1-4.1 K = 1.000
Spherical radius factor K1 = 0.90 (from table UG-37)
D/2h 3.0 2.8 2.6 2.4 2.2 2.0 1.8 1.6 1.4 1.2 1.0
K1 1.36 1.27 1.18 1.08 0.99 0.90 0.81 0.73 0.65 0.57 0.50
Equivalent Spherical radius L = K1*D
= 0.90*623
L = 560.700 mm
(a) As per appendix 1 ,clause 1-4©
Min. design thickness
t = [(P*D*K)/(2*S*E-0.2P)] + C.A.
= {[0.785*623*1]/(2*138*0.85)-(0.2*0.785)]} + 3
t = 5.085 mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦.. (a)
(b) Min. thickness as per UG 16 (b) (4) :
Min. thickness of shell in compressed air service,
= 2.5 mm + C.A.
= 2.5 mm + 3
= 5.500mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦.(b)
© Thickness calculation as per UG -32(b),
0.665 SE P = 78.0045 Mpa
Design pressure P = 0.785 Mpa
Since P < 0.665 SE UG 32(f) can be applied.
When the thickness of a hemispherical head does not exceed 0.665E the following formula shall apply.
Min. design thickness of head,
t = [PL/ (2SE-0.2P)] + C.A.
= {[00.785*560.7]/(2*138**0.85)-(0.2*0.785)]} + 3
t = 4.877 mm Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦Â¦.©
Design thickness = max. Of (a,b,cs)
= 5.500 mm
Thinning allowan = 2.00 mm
Design thickness including
Thinning allowance = 7.500 mm
Provided nominal thickness = 10.000 mm
Provided min. thickness = 8.000 mm
tr for compensation calculation : (using E=1 as per UG-37)
(i) As per UG â€œ 37
tr = PDK1/[2SE-0.2P]
= [0.785*623*0.9]/[(2*138*1)-(0.2*0.785)]
tr =1.595 mm
(ii) For nozzle reinforcement limit outside the spherical portion:
Not applicable for the present case since all nozzles on dish end comply with rule (I) describe above.
(d) STRAIGHT FACE REQUIREMENT:
Length of straight face:
As per the UW-13 (b)(3)&fig UW-13.1(d)
The min. length of the straight face shall not be less then 3th but need not exceed 38 mm.
Where
Nominal thickness of head th = 10.000 mm
Nominal thickness of shell ts = 10.000 mm
Min. length of the straight face 3th = 30.000 mm
Therefore provided straight face length = 50.000 mm
Taper transition requirement:
Offset between shell and head th-ts = 0.000 mm
1/4th thickness of thinner section = 2.000 mm
1/8th = 3.000 mm
Since th-ts is not exceeding min. of the 1/4th thickness of thinner section and 3 mm,
However based on actual thickness of formed head the 1: 3 taper may be provided
9(A) . NOZZLE NECK THICKNESS CALCULATION. (UG - 45)
From pipe schedules 4 inch outer diameter = 114.3 mm &
Schedule 80 & nominal thickness = 11.13
corrosion allowance = 3 mm
radial neck = (114.3+2*3-2*11.13)/2
= 49.02
from pipe schedule
Thickness of standard pipe wall = 6.01
t2 = ( pi*Rn )/(S.E-0.6pÃ‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬i ) + (corrosion allowance)
= (0.785*49.02)/(118-0.6*0.785) + 3
= 3.327 mm
t3 = ( pi*Rn )/(S.E-0.6pÃ‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬Ã‚Â¬i) + (corrosion allowance)
= (0.785*311)/(0.6*0.785) + 3
= 4.774 mm
Description Formula/symbol Units Hand hole
N5 A (DN 100)
Location Shell
Type of attachment Fig UW-16.1(a)
Internal design pressure Pi (from page 5-1) Mpa 0.785
External design pressure Pe (from page 5-1) Mpa 0.00
Radius neck (corroded) Rn mm 49.020
Radius shell (corroded) R mm 311.00
Allowable stress for shell material Sshell (from page 4-1) Mpa 138.000
Allowable stress for head material Shead (from page 4-1) Mpa 138.000
Allowable stress for neck material Sneck (from page 4-1) Mpa 118.00
Thickness of standard pipe wall t1 mm 6.020
Corrosion allowance C.A (page from 3-1) mm 3.000
Joint efficiency E 1
Design thickness as per UG 45 (a) t2=PiRn/(SE-0.6Pi)+C.A mm 3.327
Design thickness as per UG 45 (b)(1)
(i) shell with E=1 t3=PiR/ (SE-0.6Pi)+C.A mm 4.774
(ii) shell with E=1 Not applicable mm Not applicable
Design thickness as per UG 45 (b)(2) t4 mm Not applicable
Design thickness as per UG 45 (b)(3) t5=greater (t3,t4) mm Not applicable
Design thickness as per UG 45 (b)(4) t6=0.875*t1+C.A mm 8.268
Design thickness as per UG 45 (b)(1)-(4) t7=smaller(t3,t4,t5&t6) mm 4.774
Design thickness as per UG 16 (b) t8=larger(t7&t8) mm 5.500
Design thickness as per UG 45 (b)(1) tn=greater of (t2&t9) mm 5.500
Design thickness as per UG 45 mm 5.500
Nominal thickness 11.130
DN 100 [NPS 4] 11.130THK{sch.120} pipe
FIG (16.1A): FULL PENETRATION WELD WITH INTEGRAL REINFORCEMENT
9(B) . NOZZLE NECK THICKNESS CALCULATION (UG - 45)
Description Formula/symbol Units Air outlet
N2 (DN 50)
Location Head
Type of attachment Fig UW-16.1(a)
Internal design pressure Pi (from page 5-1) Mpa 0.785
External design pressure Pe (from page 5-1) Mpa 0.000
Radius neck (corroded) Rn mm 22.080
Radius shell (corroded) R mm 311.00
Allowable stress for shell material Sshell (from page 4-1) Mpa 138.000
Allowable stress for head material Shead (from page 4-1) Mpa 138.000
Allowable stress for neck material Sneck (from page 4-1) Mpa 118.00
Thickness of standard pipe wall t1 mm 3.910
Corrosion allowance C.A (page from 3-1) mm 3.000
Joint efficiency E 1
Design thickness as per UG 45 (a) t2=[(Pi*D*K)/(2*SE-0.2Pi)] + C.A mm 3.147
Design thickness as per UG 45 (b)(1)
(i) shell with E=1 Not applicable mm Not applicable
(ii) shell with E=1 t3=[(Pi*D*K)/(2*SE-0.2Pi)] + C.A mm 4.769
Design thickness as per UG 45 (b)(2) t4 mm Not applicable
Design thickness as per UG 45 (b)(3) t5=greater (t3,t4) mm Not applicable
Design thickness as per UG 45 (b)(4) t6=0.875*t1+C.A mm 6.421
Design thickness as per UG 45 (b)(1)-(4) t7=smaller(t3,t4,t5&t6) mm 4.769
Design thickness as per UG 16 (b) t8=2.5 mm +C.A. mm 5.500
Design thickness as per UG 45 (b) t9=larger (t7&t8) mm 5.500
Design thickness as per UG 45 tn=greater of (t2&t9) mm 5.500
Nominal thickness mm 11.070
DN 50 [NPS 2] 11.07THK{sch.XXS} pipe
9 © . NOZZLE NECK THICKNESS CALCULATION (UG - 45)
Description Formula/symbol Units PL & drain
N3&N4 (DN 20)
Location Shell
Type of attachment Fig UW-16.1bb)
Internal design pressure Pi (from page 5-1) Mpa 0.785
External design pressure Pe (from page 5-1) Mpa 0.000
Radius neck (corroded) Rn mm 17.180
Radius shell (corroded) R mm 311.00
Allowable stress for shell material Sshell (from page 4-1) Mpa 138.000
Allowable stress for head material Shead (from page 4-1) Mpa 138.000
Allowable stress for neck material Sneck (from page 4-1) Mpa 138.00
Thickness of standard pipe wall t1 mm 2.870
Corrosion allowance C.A (page from 3-1) mm 3.000
Joint efficiency E 1
Design thickness as per UG 45 (a) t2=PiRn/(SE-0.6Pi)+C.A mm 3.098
Design thickness as per UG 45 (b)(1)
(i) shell with E=1 t3=PiR/ (SE-0.6Pi)+C.A mm 4.774
(ii) shell with E=1 Not applicable mm Not applicable
Design thickness as per UG 45 (b)(2) t4 mm Not applicable
Design thickness as per UG 45 (b)(3) t5=greater (t3,t4) mm Not applicable
Design thickness as per UG 45 (b)(4) t6=0.875*t1+C.A mm 5.511
Design thickness as per UG 45 (b)(1)-(4) t7=smaller(t3,t4,t5&t6) mm 4.774
Design thickness as per UG 16 (b) t8=2.5 mm +C.A. mm 5.500
Design thickness as per UG 45 (b) t9=larger (t7&t8) mm 5.500
Design thickness as per UG 45 tn=greater of (t2&t9) mm 5.500
Nominal thickness 7.820
DN 20 [NPS 3/4] NPT half coupling C.I.6000
FIG: 16.1(b): weld reinforcement
9 (D) . NOZZLE NECK THICKNESS CALCULATION (UG - 45)
Description Formula/symbol Units Air inlet /SV
N1 N6 (DN 50)
Location Shell
Type of attachment Fig UW-16.1(a)
Internal design pressure Pi (from page 5-1) Mpa 0.785
External design pressure Pe (from page 5-1) Mpa 0.00
Radius neck (corroded) Rn mm 22.080
Radius shell (corroded) R mm 311.00
Allowable stress for shell material Sshell (from page 4-1) Mpa 138.000
Allowable stress for head material Shead (from page 4-1) Mpa 138.000
Allowable stress for neck material Sneck (from page 4-1) Mpa 118.00
Thickness of standard pipe wall t1 mm 3.910
Corrosion allowance C.A (page from 3-1) mm 3.000
Joint efficiency E 1
Design thickness as per UG 45 (a) t2=PiRn/(SE-0.6Pi)+C.A mm 3.147
Design thickness as per UG 45 (b)(1)
(i) shell with E=1 t3=PiR/ (SE-0.6Pi)+C.A mm 4.774
(ii) shell with E=1 Not applicable mm Not applicable
Design thickness as per UG 45 (b)(2) t4 mm Not applicable
Design thickness as per UG 45 (b)(3) t5=greater (t3,t4) mm Not applicable
Design thickness as per UG 45 (b)(4) t6=0.875*t1+C.A mm 6.421
Design thickness as per UG 45 (b)(1)-(4) t7=smaller(t3,t4,t5&t6) mm 4.774
Design thickness as per UG 16 (b) t8=2.5 mm+C.A. mm 5.500
Design thickness as per UG 45 (b) t9=larger (t7&t8) mm 5.500
Design thickness as per UG 45 tn= greater of (t7&t8) mm 5.500
Nominal thickness mm 11.070
DN 50 [NPS 2] 11.07 THK {sch.XXS} pipe
9 (E) . NOZZLE NECK THICKNESS CALCULATION (UG - 45)
Description Formula/symbol Units Hand hole
N5 A (DN 100)
Location Head
Type of attachment Fig UW-16.1(a)
Internal design pressure Pi (from page 5-1) Mpa 0.785
External design pressure Pe (from page 5-1) Mpa 0.00
Radius neck (corroded) Rn mm 49.020
Radius shell (corroded) R mm 311.00
Allowable stress for shell material Sshell (from page 4-1) Mpa 138.000
Allowable stress for head material Shead (from page 4-1) Mpa 138.000
Allowable stress for neck material Sneck (from page 4-1) Mpa 118.00
Thickness of standard pipe wall t1 mm 6.020
Corrosion allowance C.A (page from 3-1) mm 3.000
Joint efficiency E 1
Design thickness as per UG 45 (a) t2=[(Pi*D*K)/(2*SE-0.2Pi)] + C.A mm 3.326
Design thickness as per UG 45 (b)(1)
(i) shell with E=1 Not applicable mm Not applicable
(ii) shell with E=1 t3=[(Pi*D*K)/(2*SE-0.2Pi)] + C.A mm 4.770
Design thickness as per UG 45 (b)(2) t4 mm Not applicable
Design thickness as per UG 45 (b)(3) t5=greater (t3,t4) mm Not applicable
Design thickness as per UG 45 (b)(4) t6=0.875*t1+C.A mm 8.268
Design thickness as per UG 45 (b)(1)-(4) t7=smaller(t3,t4,t5&t6) mm 4.770
Design thickness as per UG 16 (b) t8=2.5 mm + C.A mm 5.500
Design thickness as per UG 45 (b) t9=larger (t7&t8) mm 5.500
Design thickness as per UG 45 tn=greater of (t2&t9) mm 5.500
Nominal thickness mm 11.130
DN 100 [NPS 4] 11.130 THK {sch.120} pipe
10A EVELUATION OF FILET SIZE OF ATTACHMENT : (UW16)
Nomenclature:
(i) t = corroded vessel wall thickness = 7.00 mm
(ii) tn = corroded nozzle neck thickness
(iii) te = reinforcing pad thickness =0.000mm
(iv) tmin = smaller of 19 mm or thickness of thinner of part joined
(v) tc = smaller of 6 mm or 0.7 tmin.
(vi) =leg size of outer element fillet weld =(1/2 tmin) /0.707
Required fillet weld size :
Nozzle designation Neck thk corroded
tn (mm) t min
(from iv )
(mm) tc
(from v)
(mm) Size of outer
Element fillet weld (from vi) (mm) Nozzle attachment weld=tc/0.707
(mm)
N5 A 8.130 7.000 4.900 NOT APPLICABLE 6.93
N2 8.070 7.000 4.900 NOT APPLICABLE 6.93
N3 &N4 4.820 4.820 3.374 NOT APPLICABLE 4.77
N1,N6 8.070 7.000 4.900 NOT APPLICABLE 6.93
N5 B 8.130 7.000 4.900 NOT APPLICABLE 6.93
( I ) Neck thickness corroded = nominal thickness of neck â€œ corrosion allowance
= 11.13 â€œ 3
= 8.130 mm
(II) t min = 10 â€œ 3 { we are getting t shell = 10 mm & t nozzle = 11.13 mm which ever less is used}
(III) tc = smaller of 6 mm or 0.7 t min
= 0.7*7
= 4.900 mm which is less than 6 mm
(IV) nozzle attachment = tc/0.707
= 4.900/0.707
= 6.93 mm
For safety we get 8 mm, so fillet sizes are safe.
Provided fillet size weld:
Nozzle designation Size of outer
Element fillet weld (mm) Nozzle attachment weld
(mm)
N5 A - 8.00
N2 - 8.00
N3 &N4 - 8.00
N1,N6 - 8.00
N5 B - 8.00
Therefore provided fillet sizes are safe &hence satisfactory
10 OPENIGS & REINOFRECE MENT CALCULATIONS
CHECK FOR THE LARGE OPENINGS [UG 36 B (1)]
Corroded inside diameter of vessel = 616.000 mm
One half of the vessel diameter = 308.000mm
As per UG 36 (b) (1),
Limiting size for large opening = smaller of (One half of the vessel diameter or 500 mm)
Nozzle designation Finished diameter of the
Opening (corroded) mm Whether less than one half of the
Vessel dia. Or 500 mm
N5 A 98.040 YES
N2 44.160 YES
N3 &N4 34.360 YES
N1,N6 44.160 YES
N5 B 98.040 YES
Since corroded diameter of the finished opening of above described nozzle doesnâ„¢t exceed the prescribed limitin UG-36(b)(1),none of the opening needs to be analyzed as a large openings.
Check for :
(i) Reinforcement requirement for small opening in pairs [UG(36)©(3)]
(ii) Analysis of multiple openings (UG 42)
(iii) Opening requiring Reinforcement (UG 37)
Row Nozzle designation N5 B N2 N3
N4
1 Corroded finished diameter of opening,d1 (mm) 98.040 44.160 34.360 34.360
location Orientation Refer sketch
Elevation Head Head Shell Shell
2 Possible pair of nozzles from sketch-page 11-8 - - N6 N1
Corroded finished diameter of opening d2(mm) - - 44.160 44.160
3 d1+d2 (mm) - - 78.52 78.52
4 Actual distance between nozzles of row 1&2(mm) - - 350.00 520.880
5 Actual distance between nozzles is greater than(d1+d2 )mm - - Yes Yes
6 Check for UG 36 ©(3)(d) - - No No
7 Two times average diameter of d1 and d2(mm) - - 78.520 78.520
8 Is actual distance between nozzles greater than the value in row 7 - - Yes Yes
9 Is analysis required as per UG 42 for multiple openings - - No No
10 Is calculation required for reinforcement Yes No No No
Row Nozzle designation N5 A N1 N6
1 Corroded finished diameter of opening,d1 (mm) 98.040 44.160
location Orientation
Elevation Shell Shell Shell
2 Possible pair of nozzles from sketch-page 11-8 N3 N6 N3
Corroded finished diameter of opening d2(mm) 34.36 44.60 44.160
3 d1+d2 (mm) 132.40 88.32 78.52
4 Actual distance between nozzles of row 1&2(mm) 502.71 492.67 350.00
5 Actual distance between nozzles is greater than(d1+d2 )mm Yes Yes Yes
6 Check for UG 36 ©(3)(d) No No No
7 Two times average diameter of d1 and d2(mm) 132.40 88.32 78.520
8 Is actual distance between nozzles greater than the value in row 7 Yes Yes Yes
9 Is analysis required as per UG 42 for multiple openings No No No
10 Is calculation required for reinforcement Yes No No
P=spacing ,center to center between openings
U1,U2,Â¦=Ligament width
(d1+d2)2=average diameter of pair of opening
FIG UG39: Multiple opening in rim of head with large central opening
Â¢ Actual distances between all nozzles are greater than their ave diameter and no cluster of are formed.
Â¢ Nozzle N5 A & N5B requires reinforce cement as per UG36 © (3) (A).
Welded brazed & fluid connect meeting the applicable rules and with a finished opening not larger than 89 mm or 31/2inch. In vessel shell or heads with a required minimum thickness of 10 mm or 3/8 inch or less 23/8 (10 mm) diameter in vessel shell or heads required minimum thickness 10 mm(3/8 inch).
Â¢ No pair needs to be analyzed as a pair of opening as per UG 42 (A)
When only two opening are spaced at greater then two times of average diameter the limit of reinforcement will not overlap.
Reinforcement limit calculation for nozzles N5 A DN 100[NPS-4] [UG40]
1 Designation Formula/symbol
2 Location Shell
3 Shape of opening Circular
4 Size of opening 114.30 mm
5 Required min. Neck thickness tm 5.500 mm
6 Corrosion allowance C 3.000 mm
7 Provided neck thickness(corroded) tn 8.130 mm
8 Diameter of finish opening (corroded) d 98.040 mm
9 Neck radius(corroded) Rn 49.020 mm
10 Nominal shell thickness 10.00 mm
11 Shell thickness (corroded) t 7.00 mm
12 Provided rein forcing element thickness te 0.00 mm
Limits of reinforcement
13 Parallel to vessel wall d
Rn+tn+t
Greater of (d,Rn+tn+t) 98.040 mm
64.150 mm
98.040 mm
14 Normal to vessel wall 2.5t
2.5tn+te
Smaller of (2.5t 2.5tn+te)
17.500 mm
20.325 mm
17.500 mm
15 Ã‚Â¾ of 13 for app.1-7 rules
Reinforcement limit calculation for nozzles N5 B DN 100[NPS-4] [UG40]
1 Designation Formula/symbol
2 Location Head
3 Shape of opening Circular
4 Size of opening 114.30 mm
5 Required min. Neck thickness tm 5.500 mm
6 Corrosion allowance C 3.000 mm
7 Provided neck thickness(corroded) tn 8.130 mm
8 Diameter of finish opening (corroded) d 98.040 mm
9 Neck radius(corroded) Rn 49.020 mm
10 Nominal shell thickness 10.00 mm
11 Shell thickness (corroded) t 7.00 mm
12 Provided rein forcing element thickness te 0.00 mm
Limits of reinforcement
13 Parallel to vessel wall d
Rn+tn+t
Greater of (d,Rn+tn+t) 98.040 mm
64.150 mm
98.040 mm
14 Normal to vessel wall 2.5t
2.5tn+te
Smaller of (2.5t 2.5tn+te) 17.500 mm
20.325 mm
17.500 mm
15 Ã‚Â¾ of 13 for app.1-7 rules
Reinforcement limit calculation for nozzles N5 A DN 100[NPS-4] [UG40]
NO Descriptation Formula / symbol value
1 Required thickness of seamless shell on circumferential stress tr from page7-2 1.774 mm
2 Required thickness of seamless nozzle wall tm 5.500 mm
3 Nozzle wall thickness tn 8.130 mm
4 Thickness of reinforcing element te 0.000 mm
5 Provided shell thickness t 7.000 mm
6 Corrosion allowance C.A 3.000 mm
7 Finished diameter of opening( corroded) d 98.040 mm
8 Distance of nozzle projects beyond the inner surface of the vessel wall h 0.000 mm
9 Nominal thickness of internal projection of nozzle wall ti 0.000 mm
10 Size of outward nozzle fillet weld Leg 41 8.000 mm
11 Size of RF pad fillet weld Leg 42 0.000 mm
12 Size of inward nozzle fillet weld Leg 43 0.000 mm
13 Allowable stress in nozzle Sn 118.00 MPa
14 Allowable stress in vessel Sv 138.00 MPa
15 Allowable stress in reinforcing element Deleted - MPa
16 Correction factor F 1.000
17 Stress reduction factor
(inserted throâ„¢ wall) fr1=Sn/Sv 0.855
18 Stress reduction factor fr2=Sn/Sv 0.855
19 Stress reduction factor fr3=lesser of (Sn or Sp)/Sv 0.855
20 Stress reduction factor fr4=Sp/Sv 0.000
21 Area required for reinforcement A=dtrF+2tntrF(1-fr1) 178.15 mm2
22 Area in excess thickness in the vessel wall available for reinforcement a)d(E1t-ftr)-2tn(E1t-Ftr)(1-fr1)
b)2(t+tn)(E1t-Ftr)-2tn(E1t-Ftr)- (1-fr1)
A1=larger of a & b 503.98 mm2 145.81 mm2
503.98 mm2
23 Area in excess thickness in the nozzle wall available for reinforcement a)5(tn-tm)fr2t
b)2(tn-tm)fr2t(2.5tn+te)
A2=smaller of a & b 78.709 mm2
639.91 mm2
78.709 mm2
24 Area available for reinforcement when nozzle extend inside the vessel a)5 t ti fr2
b)5 t ti fr2
c)2 h ti fr2
A3=smaller of a,b&c 0.000 mm2
0.000 mm2
0.000 mm2
0.000 mm2
25 Area available in outward nozzle weld A41=(leg 41)2 fr3 54.725 mm2
26 Area available in outer pad weld A42=(leg 42)2fr4 0.000 mm2
27 Area available in inward nozzle weld A43=(leg 43)2fr2 0.000 mm2
28 Total area available for reinforcement A1+A2+A3+A41+A42+A43 637.42 mm2
29 Is area provided is greater than required Action code YES
Reinforcement limit calculation for nozzles N5 B DN 100[NPS-4] [UG40]
NO Descriptation Formula / symbol value
1 Required thickness of seamless shell on circumferential stress tr from page7-2 1.595 mm
2 Required thickness of seamless nozzle wall tm 5.500 mm
3 Nozzle wall thickness tn 8.130 mm
4 Thickness of reinforcing element te 0.000 mm
5 Provided shell thickness t 7.000 mm
6 Corrosion allowance C.A 3.000 mm
7 Finished diameter of opening( corroded) d 98.040 mm
8 Distance of nozzle projects beyond the inner surface of the vessel wall h 0.000 mm
9 Nominal thickness of internal projection of nozzle wall ti 0.000 mm
10 Size of outward nozzle fillet weld Leg 41 8.000 mm
11 Size of RF pad fillet weld Leg 42 0.000 mm
12 Size of inward nozzle fillet weld Leg 43 0.000 mm
13 Allowable stress in nozzle Sn 118.00 MPa
14 Allowable stress in vessel Sv 138.00 MPa
15 Allowable stress in reinforcing element Deleted - MPa
16 Correction factor F 1.000
17 Stress reduction factor
(inserted throâ„¢ wall) fr1=Sn/Sv 0.855
18 Stress reduction factor fr2=Sn/Sv 0.855
19 Stress reduction factor fr3=lesser of (Sn or Sp)/Sv 0.855
20 Stress reduction factor fr4=Sp/Sv 0.000
21 Area required for reinforcement A=dtrF+2tntrF(1-fr1) 160.13 mm2
22 Area in excess thickness in the vessel wall available for reinforcement a)d(E1t-ftr)-2tn(E1t-Ftr)(1-fr1)
b)2(t+tn)(E1t-Ftr)-2tn(E1t-Ftr)- (1-fr1)
A1=larger of a & b 521.29 mm2 150.82 mm2
521.29 mm2
23 Area in excess thickness in the nozzle wall available for reinforcement a)5(tn-tm)fr2t
b)2(tn-tm)fr2t(2.5tn+te)
A2=smaller of a & b 78.709 mm2
639.91 mm2
78.709 mm2
24 Area available for reinforcement when nozzle extend inside the vessel a)5 t ti fr2
b)5 t ti fr2
c)2 h ti fr2
A3=smaller of a,b&c 0.000 mm2
0.000 mm2
0.000 mm2
0.000 mm2
25 Area available in outward nozzle weld A41=(leg 41)2 fr3 54.725 mm2
26 Area available in outer pad weld A42=(leg 42)2fr4 0.000 mm2
27 Area available in inward nozzle weld A43=(leg 43)2fr2 0.000 mm2
28 Total area available for reinforcement A1+A2+A3+A41+A42+A43 654.72 mm2
29 Is area provided is greater than required Action code YES
Action code:
Yes: No further calculation is required & the nozzle is self reinforced
NO: provide external elements for compensation
Conclusion: Provided reinforcement is adequate
12. STRENGTH OF WELD PATH (UG-41 / U-15)
AND MAXIMUM SHEAR STRESS DUE TO IMPOSED LOAD (UG-45C)
For nozzles N1, N2, N3, N4 & N6 :
As these nozzles are exempted from reinforcement by UG-36 C (3), strength calculation for these welds are exempted from strength path calculation. (UW 15 (b) (2) )
For nozzles N5A & N5B :
The type of weld [Figure UW â€œ 16.1 (a)] provided is exempted by UW-15(b)(1) for strength calculation for nozzle attachment weld for pressure loading.
MAXIMUM SHEAR DUE TO IMPOSED LOAD (UG-45C)
As per Appendix L-7.5, 3.4,
Shear stress due to shear load on
circumference of Handhole neck = W/ r tn
= 0.232 MPa
Where,
Shear load W = Blind flange wt. + Bolting flange wt.
Blind flange = 7.7 kg
Bolting flange weight = 6.8 kg
W = 14.5 kgs
= 20 kgs
Inside radius of handhole neck r = 49.020 mm
Minimum neck thickness tn = 5.500 mm
Shear stress due to torsion load
on circumference of Handhole neck = W/2 r 2 tn
= 0 MPa
Where,
Torsion load = 0 N
Combined shear stress = Shear stress due to shear load + Shear stress due to torsion load
= 0.232 MPa
As per UG-45 ©
Allowance shear stress = 70% of the allowable tesile stress for the Handhole neck Material.
= 83 MPa
Where,
Allowable tensile stress for neck material = 118 MPa
Since combined shear stress is less than allowable shear the design is safe.
13. CHECK FOR FLANGE RATINGS
Ref: (i) UG-11 & UG-44
(iii) ASME B 16.5 Table 2-1.1
Designation Nozzles
1 Nozzle designation Ni, N2, N5A/B & N6
(Bolting Flange)
2 Flange material SA 105M
3 Material Group number 1
4 Design Temperature (From Page 3-1) 45 C.
5 Coincident design pressure (From Page 3-1) 0.785 MPa
6 MAWP (From Page 3-1) 0.785 MPa
Selection of rating of flange based on MAWP
7 Class 150 Class
8 Working pressure permitted by Table 2-1.1 1.940 MPa
9 Whether MAWP < Working pressure permitted YES
10 Fastener material SA 193M Gr. B7
SA 194M Gr. 2H
Designation Nozzles
1 Nozzle designation Ni, N2, N5A/B & N6
(Bolting Flange)
2 Flange material SA 105M
3 Material Group number 1
4 Design Temperature (From Page 3-1) 45 C.
5 Hydrostatic test pressure (From Page 3-1) 0.785 MPa
Selection of rating of flange based on MAWP
6 Class 150 Class
7 Working pressure permitted by Table 2-1.1 1.940 MPa
8 Whether Hydrostatic test pressure < Working pressure permitted YES
9 Fastener material SA 193M Gr. B7
SA 194M Gr. 2H
Hence selected flange rating of Class 150 is safe.
It is recommended that only Group No. 1 gaskets be used for Class 150 flanged joints.
Hence Compressed Asbestos (CAF) Gasket is selected.
16. EVALUATION OF HEAT TREATMENT REQUIREMENTS
FOR COLD FORMED COMPONENTS (UCS-79)
Action Code :-
NO : Post Forming heat treatment is not required. When percentage elongation < 5%
YES : Post Forming heat treatment Is required. When either of the following conditions exists.
(a) When percentage elongation > 40%.
(b) When percentage elongation > 5% but <40% & any of the conditions 7 to 11 exists.
CHECK : Proceed further to check whether conditions 7 to 11 exists.
SHELL HEAD
1 Forming temperature Ft Atmospheric Atmospheric
2 Final center line radaius (Infinity for flat plate), Rf 311.000 mm 105.91 mm*
3 Original center line radius (Infinity for flat plate), R0 Infinity mm Infinity mm
4 Nominal thickness, t 10.000 mm 10.000 mm
5 Thickness after forming, tmin 10.000 mm 8.000 mm
6 % extreme fiber elongation Applicable formulae Result [50t/Rf] [1-(Rf/R0)]
1.61 % [75t/Rf] [1-(Rf/R0)]
7.08%
Is % extreme fiber elongation exceeding 5% NO YES
Action code NO YES
7 Is lethal service NO NO
8 Is impact test required Not required Not required
9 Is thickness exceeding 16 mm before cold forming NO NO
10 % Reduction in thickness Applicable formulae Result (t-tmin) (100/t)
0 (t-tmin) (100/t)
20%
11 During forming 120 C to 480 C N.A. N.A.
12 Action Code YES
* Based on 17% of D, as per provided in UG-32 (d)
Conclusion :
Head shall be heat treated after forming. Specified heat treatment : Stress Relieving.
Shell is not required to be heat treated after forming.
17. EVALUATION OF POST WELD HEAT TREATMENT REQUIREMENT (UCS-56)
Weld Type L/S Butt weld Butt weld Set in groove weld Fillet weld
Location Shell Shell to Head Nozzle neck to shell Pad plate to Shell
MOC SA 516M Gr. 485 SA 516M Gr. 485 SA 516M Gr. 485 SA 516M Gr. 485
Nominal thk. As per UW-40 (f) (mm) 10.00 10.00 10.00 10.00
Limit thick as per UCS-56 (mm) 38.00 38.00 38.00 38.00
Is Nominal thk. > Limit thickness NO NO NO NO
PWHT Required as per table UCS-56 NO NO NO NO
MDMT < - 48 C (Check for UCS-68) NO NO NO NO
PWHT required as service requirement as per U2 (a) (3) NO NO NO NO
CONCLUSION :
PWHT is Not Mandatory, since all checks have resulted negative.
18. EVALUATION OF FULL RADIOGRAPHY REQUIREMENTS (UW-11)
Uw-11 (A) Check Description Result Action Code
1 Whether the vessel contains lethal substance NO NONE
2 Whether all butt welds in vessel exceeds 32 mm (UCS-57) NO NONE
3 Whether vessel is an unfired stream boiler having design pressure > 350 kpa NO NONE
4 Is there any category-C joint in nozzle that either exceed DN 250 or 29 mm vessel wall thickness for 1,2 & 3 above NO NONE
5 Are the category A & D butt welds of vessel designed with joint efficiency of 1 NO NONE
6 Whether electro gas / electroslag welding is used NO NONE
Action code:
YES : Full radiography is required as per UCS-57
NONE : Full radiography of joint is not required.
Conclusion:
Full RT is not mandatory for fulfilling code requirements.
Specified NDE is documented on page no. 6-1 of this calculation.
19. EVALUATION OF IMPACT EST REQUIREMENTS OF MATERIALS
MDMT = 0 Ã‚Â°C
Component Shell Head Pipe
Type of Weld Butt Weld Butt Weld Fillet Weld/Groove
Material of Construction SA 516M Gr. 485 SA 516m gr. 485 SA 106 Gr. B
P. No. 1 1 1
Gr. No. 2 2 1
Applicable Curve D D B
Governing thickness Fig UCS-66-3 (mm) 10.000 8.000 8.000
Exemption thick. Per UG-20 (f) (mm)
For curve B,D material 1 1 1
Component Flange Fasteners Pipe Fittings
Type of Weld Butt Weld - Fillet Weld/ Groove
Material of Construction SA 105M SA 193M Gr. B7 SA 105 M
SA 194M Gr. 2H
P. No. 1 - 1
Gr. No. 2 - 2
Applicable Curve B - B
Governing thickness Fig UCS-66-3 (mm) 11.130 - 7.82
Exemption thick per UG-20 (f) (mm)
For curve B material 1 - 1
A) Check for UG-20 (f) exemptions :
1. Whether all the materials are P-No. 1, Gr. No. 1 or 2 : Yes
Whether thick. Exceeds 25 mm for materials listed in Curve B of Fig. UCS-66 : No
2. Is the vessel hydro tested as per UG-99 (b) : Yes
3. Is design temperature between -29 C and 345C : Yes
4. Whether thermal or mechanical loadings are a controlling design requirement : No
5. Whether cyclic loading is a controlling design requirement : No
As per UG-20 9f), all components are exempted from impact testing.
B) As per UCS-66-C Exemption temperature for ASME B 16.5 flanges is -29 C
C) For bolting, note (e) of General notes of Fig. UCS-66 is applied.
Exemption temperature for SA 193 M Gr. B7 & SA 194M Gr. 2H 64 mm dia and under is â€œ 48 C
There for all components are exempted from impact testing.
D) Since all base metals are exempted from impact test, weld metal & HAZ also exempted from impact testing as per UCS-67.
20. EVALUATION OF HYDROSTATIC TEST PRESSURE (UG-99(B))
MAWP = 0.785 MPa. Aat 0 Ã‚Â°C
Material Allowable stress at Design temperature of 45 Ã‚Â°C (MPa) Allowable stress at Test temperature of 45 Ã‚Â° C.
SA 516M Gr. 485 138 138
SA 106 Gr. B 118 118
SA 105M 138 138
SA 193M Gr. B7 172 172
Applicable Formulae :
(i) Stress Ratio = Allowable Stress at DesignTemp. 45 Ã‚Â°C
Allowable Stress Test Temperature at 45 Ã‚Â°C
(ii) Hydrotest pressure = (1.3 * MAWP *Lowest Stress Ratio ® ), MPa
Material From Formula (i) Hydrotest Pr. From Formula (ii) in MPa
SA 516 M Gr. 485 1 .021
SA 106 Gr. B 1 1.021
SA 105M 1 1.021
SA 193M Gr. B7 1 1.021
Therefore Hydrostatic test Pressure at the top of the Vessel
= 1.021 MPa
= 10.40 kg/cm2
Say = 11.00 kg/cm2
21. TOLERANCES (UG 80 & UG-81)
Shell Tolerances (UG-80)
Inside diameter of vessel = 610.00 mm
Corrosion allowance = 3.00 mm
Design pressure = 0.785MPa
Allowable stress for shell material, S = 138 MPa
Nominal shell thickness = 10.000 mm
As per UG-80 (a) (1), the difference between the maximum and minimum inside diameters at any cross section shall not exceed 1% of nominal diameter at the cross section under consideration.
Corroded inside diameter of vessel = 616 mm
Ovality (Dmax-Dmin) = 1% of I.D. (Corroded)
= 6.16 mm
Provided tolerance = 6.00 mm
Check for adequacy of shell thickness considering maximum possible diameter.
Possible inside diameter of vessel = 622 mm (in corroded condition)
Radius of vessel = 311.000 mm
Design thickness of shell t = PR/(SE-0.6P) + CA
= 4.775 mm
Required design thickness is less than the specified thickness considering maximum possible diameter.
Possible inside diameter of vessel = 622 mm (in corroded condition)
Radius of vessel = 311.000 mm
Design thickness of shell t = PR/(SE-0.6P) + CA
= 7.775 mm
Required design thickness is less than the specified thickness 10mm, hence safe.
Head Tolerance (UG-81)
As per UG-81 (a), the maximum outside deviation from the specified shape shall not exceed 1.25% of nominal ID and maximum inside deviation from specified shape shall not exceed 5/8% of nominal ID.
Corroded I.D. of head = 616.00 mm
i) Maximum outside deviation = 1.25 % of Corroded I.D.
= 7.70 mm
Provided tolerance = 7.00 mm
ii) Maximum inside deviation = 5/8% of Corroded I.D.
= 3.85 mm
Provided tolerance = 3.00 mm
22. DESIGN OF LIFTING LUGS
(1) IS 800 : 1984 Code of practice for general construction in steel.
(2) Strength of materials by Domkundwar.
DESIGN DATA
Lifting Load = 400.000 Kgs
Dynamic factor = 2.00
Design Load = 800.000 Kgs
Number of lifting lug = 2 Nos.
Material of Lifting Lug = SA 516M Gr. 485
Yield strength fy = 260.000 MPa
Allowable tensile & compressive stress = 0.60 x fy
= 171.600 MPa
(As per IS 800 cl. 4.1.1.
Allowable bending stress = 0.45 x fy = 171.600 MPa
(As per IS 800 cl.6.4.1.)
Allowable bending stress = 0.66 x fy = 171.600 MPa
(As per IS 800 cl. 6.2.1)
NOMENCLATURE
P : Load on each lug in N
: Angle made by lifting force with the axis in degree
F : Horizontal force in N
R : Lifting force in N
M1, M2 : Bending moment induced in lug N-mm
T : Thickness of lifting lug in mm
T : Breadth of lifting lug in mm
L : Distance from lug hole centre to top of lug in mm
Ss : Shear stress in MPa
St : Tensile stress in MPa
Sc : Compressive stress in MPa
Sb : Bending stress in MPa
Ssa : Allowable shear stress in MPa = 117.000 MPa
Sta : Allowable tensile stress in MPa = 156.000 MPa
Sca : Allowable compressive stress in MPa =156.000 MPa
Sba : Allowable bending stress in MPa = 171.600 MPa
DESIGN CALCULATION
P = Lifting load / 2 = 3924.000 N
= 15.00 (Assumed)
F = P x tan = 1051.433 N
R = F / sin = 4062.424 N
I = Length of lug = 10.000 mm
E = Weld joint efficiency = 0.65
L = Distance from lug hole centre to top of lug = 90.00 mm
PART A
Bending moment M1 = P x t / 2
= 19620.00 N-mm
M2 = F x L
= 94628.94 N-mm
Total bending moment M1 + M2 = 114248.9 N-mm
Section modulus Z = t2 x 1 / 6
= 2250.000 mm3
Induced Bending stress Sb = (M1 + M2) / Z
Sb = 50.777 MPa
Total tensile stress in the lug. (Sb / Sba) + (St /Sta) < 1
= 0.296 + 0.019
= 0.315 < 1
HENCE SAFE
PART B :
a) Shear stress of welding part
Length of the weld = 260.000 mm
Fillet weld = 8.000 mm
Area of weld, Aw = 0.707 x Fillet weld x length of weld
= 1470.560 mm2
Shear stress = P / Aw = 2.668 MPa (a)
b) Shear stress due to break away of welding part
Shear stress = F /Aw = 0.715 MPa (b)
Total shear stress
Total shear stress = (a) + (b) = 3.383 MPa
Allowable shear stress for the weld = Ssa x E
= 76.05 MPa
Induced shear stress < Allowable shear stress. HENCE SAFE
CHECK FOR STRESS INDUCED IN THE SHELL AT THE POINT OF ATTACHMENT OF PAD WITH THE SHELL
At the edge of pad, as per Annex G. Notation from G.2.2.
Cx = Half length of rectangular loading area in longitudinal direction, in mm.
= 45.00 mm
C f = Half length of rectangular loading area in circumferential direction, in mm.
= 4.00 mm.
D = Mean dia of vessel, in mm.
= 610 + 10
D = 620.00 mm
t = Analysis thickness of vessel shell
= 10.00 mm
t1 = Analysis thickness of pad plate
= 10.00 mm.
L = Length of cylindrical part
= 1000.00 mm
r = Mean radius of cylinder
= 310.00 mm
Cf/ Cx = 0.09
Cx / r = 45/310
= 0.15
2Cx / L = (2 x 45) / 1000
= 0.09
W = 800*9.81 / 2
W = 3924.00 N
64r/t (Cx/r)2 = 64 x 252 / 8 (45/252)2
= 41.806
M f / W = 0.150 (as per figure G.2.2.6)
Mx / W = 0.075 (as per figure G.2.2.7)
N f t/W = 0.130 (as per figure G.2.2.8)
Nx t/W = 0.130 (as per figure G.2.2.9)
M f = 0.15 W
= 588.60 N
Mx = 0.075 W
= 294.30 N
N f = - 0.128 W/t
= - 50.23 N/mm.
Nx = - 0.130 W/t
= - 51.01 N/mm.
Circumferential Stress = Nf/t + 6 Mf/ t2
= 48.903 N / mm2 (Compression)
= - 40.339 N/mm2 (Tension)
Longitudinal Stress = Nx/t + 6 Mx / t2
= 12.557 N / mm2 (Compression)
= - 22.759 N / mm2 (Tension)
Primary membrane stress
S = (PR/t + 0.6P)/E
S = 24.81 N/mm2
Total induced tensile stress = 86.27 N / mm2
Allowable tensile stress = 156.000 N/mm2
Total induced tensile stress = 86.27 N / mm2 < 156.000 N/mm2
HENCE SAFE
23.DESIGN OF SADDLE SUPPORT.
As per PD 5500: 2006, Annex G
Design pressure = 0.785 MPa 0.785 N/mm2
Pressure due to static head at mid span = 0.011 MPa 0.011 N/mm2
Density of water = 1000 kg/m3 1000.00 Kg/m3
I.D. of Vessel = 0.000 mm 0.000 mm
Inside Radius of Vessel, ri = 0.000 mm 0.000 mm
Vessel Length (W.L. to W.L.), L = 1000.00 mm 1000.00 mm
Depth of head, b = 0.850 mm 0.850 mm
Thickness of shell, t = 0.0000 mm 0.0000 mm
Thickness of head = 10.000 mm 10.000 mm
Mean radius ® = 0.000 mm 0.000 mm
A = 200.000 mm 200.000 mm
b1 = 50.000 mm 50.000 mm
b2 = 50.000 mm 50.000 mm
Distance between saddle (Ls) = 600.000 mm 600.000 mm
Distance from base to centerline of vessel (B)
= 221.000 mm 221.000 mm
Reactions per support arising from seismic load
Seismic force * (B / Ls) = 0.000 N 0.000 N
Reactions per support arising due to Vessel Weight and Seismic
W1 = Max. [Operating wt. & Hydrotest Wt.] + Seismic Load
W1 = 0.000 kgs = 0.00 N
At Mid Span from equation G.7
M3 = {W,L/4} {((1 + 2 (r2 â€œ b2)/L2)/(1 + 4b/3L)) â€œ (4A/L)}
= 0.00 Ibs-inch = 0 N-mm
At the supports from equation G.8
M4 = -W,A {((1 â€œ ((1 â€œ A/L + (r2 â€œ b2)/L2)/(2AL)) â€œ (1 + 4b/3L)}
= 0.00 Ibs-inch = 0 N-mm
i) At mid span
The stress at the highest point of cross section from equation G.9
F1 = (Pmr/2t) â€œ (M3/*r2t)
= psi N.mm
Check for condition when the vessel is full of product with zero top pressure
i.e. Pm = 0.011 N/mm2
f1 = psi N.mm2
The stress at the lowest point of cross section from equation G. 10
F2 = (Pmr/2t) â€œ (M3/*r2t)
= psi N.mm2
ii) At Supports
The stress near the equator from equation G. 11
F3 = (Pmr/2t) â€œ (M3/*r2t)
= psi N.mm2
Where, K1 = 0.107 from table G.2 For = 120
Since A > r/2 and the shell is not stiffened
The stress at the lowest point of cross-section from equation G. 12
F4 = (Pmr/2t) â€œ (M3/*r2t)
= psi N.mm2
Allowable Direct stresses
The stress intensity acting at the point should be taken as
= Max (S1-S2 ; S1 + 0.5 P ; S2 + 0.5 P)
Where S1 and S2 are principle stresses
For this the stress intensity should not exceed f i.e. maximum allowable stress
For SA 516 Gr. 70 max. allowable stress f = 45 psi
F = 0.31 N/mm2
The primary membrane circumferential stress
i) at the highest point of cross section q
S = Pr/t
=
= psi N.mm2
ii) at the lowest point of cross section
The primary membrane Stress intestates involving the longitudinal stress are (S - Sz) and (Sz + 0.5P)
A) At mid point
B) I) at highest point of cross section
(S - Sz) = #DIV/0! - #DIV/0!
= #DIV/0! psi - #DIV/0! N.mm
(Sz + 0.5 P) = #DIV/0! - #DIV/0!
56.893 psi 0.39 N.mm2
B) At the supports
i) at Equator
ii)
(S-Sz) = #DIV/0! - #DIV/0!
= #DIV/0! psi -
(Sz + 0.5 P) = #DIV/0! + 0.5 x 0.785
= #DIV/0! psi -
Maximum stress induced = #DIV/0! psi -
#DIV/0!
Limits for the longitudinal compressive stress (Y.2.5.3 & A.3.5)
Calculate compressive stress should not exceed x S x f
Where,
K = Pe/Pyss
= #DIV/0!
Where,
Pe = 1.21 x Ex t2/r2
= #DIV/0! psi -
Pyss = 2 x S x f x t/r
= #DIV/0! psi -
=
#DIV/0! From fig A-2 in terms of K with S and f
for this case K = Pe/Pyss,
Using equation 3.24 and 3.25 of 3.6.4
Where,
E = modulus of elasticity
= 30457980 psi = 210000 N/mm2
S = A factor relating f to effective yield point of material, for carbon steel S = 1.4
= 1.40 (As per 3.6.1)
f = 44.97 psi = 0.31 N/mm2
From Fig. A.2
= 0.45 + 0.00625 K
= #DIV/0!
x S x f = #DIV/0! Psi = #DIV/0! N/mm2
i.e. Allowable compressive stress
as in this case there is no longitudinal compressive stress
#DIV/0!
#DIV/0!
Tangential shearing Stresses at the support (G.3.3.2.4)
A = 7.874 inch = 200.00 mm
r/2 = 0.0000 inch = 0 mm
A > r/2, Equation G. 13
Tangential shearing Stress in the Shell
q = [(K3W1) / (rt)] [(L-2A) / {(L + (4b/3)}]
= #DIV/0! Psi = #DIV/0! N/mm2
From Table G.3, for = 120
K3 = 1.171
Allowable tangential shearing stresses
= Min[0.8f, (0.06Et)/r]
= #DIV/0! Psi = #DIV/0! N/mm2
#DIV/0!
Circumferential stresses : (G.33.2.5.1 & Y.2.7)
Maximum stress
Maximum value of the circumferential stress occur in the region of the saddle support.
Therefore as per G.3.3.2.5.1,
Thickness of saddle plate (t1) = thickness of shell plate (t)
And b2 = b1 + 10 t.
t = 0.315 inch = 8.00 mm
b2 = 1.969 inch = 50.00 mm
Stress at the lowest point of cross-section
Equation G.16 In the region at the edge of saddle plate where the thickness is t + t for = 120
F5 = [-K5 W1]/[(t+t1)b2]
= 0 psi 0.00 N/mm2
Where, K5 factor from table G.5, for = 120
When saddle is welded to vessel as per G.3.3.2.5.1 divide K5 by 10
K5 = 0.076
In the region at the edge of saddle plate where the thickness is t & for = 132
F5 = [-K5 W1] [tb2]
= #DIV/0! Psi =
Where, K5 factor
From table G.5, for = 132
when saddle is welded to vessel as per G.3.3.2.5.1 divided K5 by 10
As per G.3.3.3.2.5 when the saddle is welded to the vessel. The value of f5 should not exceed f
#DIV/0!
Stress at the horn of saddles
L/r = #DIV/0!
#DIV/0!
F6 = -[w1/(4tb2)]-[(12k6w1r)/(Lt2)]
Since saddle plate has been extended by 12â„¢and has a width value b2,the stress at the edge of the extended saddle plate and the edge of the extended saddle plate should be determind.
Stress at the edge of saddle = 120
Thickness = shell thickness (t) + saddle thickness (t1)
t = 0.315 inch = 8.00 mm
b2 = 1.969 inch = 50.00 mm
= 120
A/r = #DIV/0!
K6 = #DIV/0! From table G.4
F6 = #DIV/0! Psi = #DIV/0! N/mm2
Stress at the edge of saddle = 132
Thickness = shell thickness (t)
t = 0.000 inch = 0.000 inch
b = 1.969 inch = 50.00 mm
= 132
A/r = #DIV/0!
K6 = 0.031 from table G.4
Numerical value of circumferential stress f6 should not exceed 1.25f
#DIV/0!
Check for adequacy of web plate
From pressure vessel hand book by Eugene f. megyesy
The saddle at the lowest section must be resist the horizontal force (F).the effective cross section of the saddle to resist this load is 1/3 of the vessel radius ®
F = k11w1
Where
K11 = constant depend on contact angle
= 0.204
r/3 = 0.000 inch = 0.000 mm
web plate thickness = 0.315 inch = 8.00 mm
contact angle = 120
w1 = load on each saddle
force F = 0.000 lbs
effective area of web plate = (r/3)* web plate thickness
= 0.000 inch2 = 0.000 mm2
Induced stress due to force F = effective area of web plate
= #DIV/0! Psi = #DIV/0! N/mm2
Allowable stress = 2/3 allowable tensile stress
= 29.98 psi = 0.21 N/mm2
#DIV/0!
Design of base plate
(reference pressure vessel design manually by D.R.Moss )
Length of base plate A = 18.504 inch = 470.00 mm
Width of base plate F = 1.969 inch = 50.00 mm
Thickness of base plate tb = 0.315 inch = 8.00 mm
Area of base plate Ab=A*F = 36.425 inch2 = 23500 mm2
Bearing pressure Bp = W1/Ab
= 0.000 psi = 0.00 N/mm2
Induced bending stress fb = 3W1F/(4Atb2)
= 0.000 psi = 0.00N/mm2
Allowable bending stress = 2/3*allowable tensile stress
= 29.98 psi = 0.21 N/mm2
Induced bending stress<Allowable bending stress .Hence thickness of base plate is adequate
Number of ribs
(Reference pressure vessel design manual by D.R.MOSS)
Length of the base plate A = 18.504 inch = 470.00 mm
Number of ribs n = (A/24)+1
= 1.77 Nos.
Provided no. ribs = 2 Nos.
Conversion factor:
1 N/MM2 = 145.038 psi.
1 N =0.225 lbs
24.SEISMIC ANALYSIS
Design pressure p = 8.00 psi.
Inside diameter of vessel d = 0.000 inch
Overall height of vessel H = 3.396 ft
Distance between bottom T.L to C.G
Of vessel L = 530.00 mm
= 20.866 inch
Allowable stress for shell material S = 45 psi.
Nominal shall thickness t = 0.000 inch
Thickness of shall in corroded condition t1 = 0.000 inch
Operating weight of vessel W1 = 400 lbs
Hydro test weight of vessel W2 = 600 lbs
Uniform operating weight of vessel w = 118 lbs/ft
As per IS 1893-1984
Seismic zone = III
As per seismic coefficient method,
Horizontal seismic coefficient a h = ÃƒÅ¸ l a0
= 0.072
Where,
Coefficient depending upon the soil foundation ÃƒÅ¸ = 1.2(Table 3)
Important factor l = 1.5 (Table 4)
Basic horizontal seismic coefficient a 0 = 0.04(Table 2)
For Operating Condition:
Horizontal seismic force F = W l a h
= 29 lbs
Moment induced at the bottom tangent line Mo = F L
= 601 in-lb
For Hydro test condition:
Horizontal seismic force F = w1 a h
= 43 lbs
Moment induced at the bottom
tangent line Mo = FL
= 901 in-lb
Note: Seismic load is included in saddle design calculation hence separate seismic analysis is not required