100m3 Design Calculation

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 1/23

ISISAN ISI SAN. VE TİC. A.Ş.

DESIGN CALCULATIONS

100 m3 BUTANE TANK

(B+F MODULES)

PREPARED BY

Yavuz Talaslıoğlu Önder ÜLKER

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 2/23

CONTENT PAGE

1) DESIGN DATA FOR 100 m³ BUTANE TANK 3

2) THICKNESS CALCULATION FOR SHELL SUBJECTED TO INTERNAL PRESSURE 4

3) THICKNESS CALCULATION FOR HEAD SUBJECTED TO INTERNAL PRESSURE 4

4) THICKNESS CONTROL UNDER EXTERNAL PRESSURE 6

5) OPENING CALCULATIONS OF THE NOZZLES 8

5) LIFTING LUG STRESS CALCULATIONS 14

6) SADDLE STRESS CALCULATIONS 15

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 3/23

1. DESIGN DATA FOR 100 m³ BUTANE TANK

DESIGN CODE CODAP 2005 Div2

VOLUME 100 m³

CONSTRUCTION CATEGORY A

SERVICE Butane

DESIGN PRESSURE (INTERNAL) P = 8 bar (0,8 MPa )

HYDRO TEST PRESSURE Pt = 11,44 bar (1,144 MPa)

EXTERNAL PRESSURE Pe = 0.05 MPa (0,5 bar )

MATERIAL P 355

MATERIAL STANDARD NF EN 10028-3

ALLOWABLE STRESS, f = 204,17 MPa

4,2,

5,1min 2.0 RmRpf

t

CORROSION ALLOWANCE, c c = 1 mm

DESIGN TEMPERATURE -10°C / 50°C

JOINT EFFICIENCY, z z =1

RADIOGRAPHIC EXAMINATION 100 %

NUMBER OF TANK 4 pieces

OUTSIDE DIAMETER De= 2.900 mm

SHELL LENGTH Lt= 14.350 mm

DEPTH OF HEAD h2 =.721 mm

EMPTY WEIGHT - 12,500 kg (approx.)

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 4/23

2. THICKNESS CALCULATION FOR SHELL SUBJECTED TO INTERNAL

PRESSURE

4,2490,

5,1355minf = min 16,204;67,236 = 204,16 N/mm²

16,204f N/mm²

Shell thickness:

cPzf

DePe

2

18,0116,2042

900.28,0

e

e. = 6,67 mm

3. THICKNESS CALCULATION FOR HEAD SUBJECTED TO INTERNAL

PRESSURE

22hDi =1,9

iD 2.884 mm

08,0

22

1

i

ii

hD

Dr 529,24 mm (C3.1.4.1)

02,0

244,0

i

ii h

DDR 2.466,11 mm (C3.1.4.2)

bys eeeMAXe ;; (C3.1.5.1a)

1) cPzf

RPes

5,02

18,05,0117,2042

11,466.28,0

se se = 5,84 mm (C3.1.5.1b)

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 5/23

2) iD

r 0,18

2,01,0 iD

r :

0032,004,0;

ReMINY

44 0065,0902,6

1006,1902,6

1006,1Y

N 0,845 (C3.1.5.1.c8)

YZ log 2,501 (C3.1.5.1.c7)

2,01,0 1,02,010

ii Dr

Dr (C3.1.5.1.c4)

8370,02943,10383,11833,0 231,0 ZZZN = 1,037 (C3.1.5.1.c3)

5,0;375,78843,1532,0 22,0 YYMAX =0,525 (C3.1.5.1.c5)

2,01,0 1,02,010

ii Dr

Dr 0,61

cf

DRe iy 2,075,0 (C3.1.5.1c)

ye = 6,8 mm

3) cfP

rDDRe i

ib

667,055,0

2,075,00433,0 (C3.1.5.1d)

be = 7,78 mm

bys eeeMAXe ;; [ (5,84 mm) ; (6,80 mm) ; (7,78 mm) ]

emin = 7,78 mm for head thickness

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 6/23

4. THICKNESS CONTROL UNDER EXTERNAL PRESSURE

Shell thickness ts = 8 mm is taken from internal pressure calculation

1 st Trial

Find A and B Factor;

hD 2/ 1,9

K = 1

)3/2( 2hLL t 14.831 mm

900.2831.14

oDL

5,1

18t 7 mm

7900.2

tDo 414

A 0,00003 from chart C4.9.1.

B = 2

EA =2

900.19900003,0 = 3

4143134

)/(34

tDKBP

oa 0,01 MPa

PPa 0.01 MPa < 0. 05 MPa therefore ring is required.

2 nd Trial

Find A and B Factor;

hD 2/ 1,9

K = 1

4/2 LL 3.708 mm

900.2708.3

oDL

1,3

18t 7 mm

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CALCULATIONS

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7900.2

tDo 414

A 0,00011 from chart C4.9.1.

B = 20 from chart C4.9.2.

4143

1204)/(3

4tD

KBPo

a 0,064 MPa

PPa 0.064 MPa > 0. 05 MPa

Therefore 2 number rings are required.

100 m³ BUTANE TANK

CALCULATIONS

Date : 13.06.2011 Page No : 8/23

d=diL

ll,

Ri

et

et,

l, t

St

G

SrS

5. OPENING CALCULATIONS OF THE NOZZLES

5.1.FORMULAS

eRD im 2

4,2,

5,1min 2.0 RmRpf

t

tim ed 2

'' 2 tim ed

For ko Graph C5.1.3 on pg.772 shall be used

eDkL m 0

)(, ttm ledMINl

)(,5.0 ''''ttm ledMINl

The following condition shall be fulfilled

GPPfSPfSPfS rrtt 5.05.05.0

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CALCULATIONS

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5.2.K5 NOZZLE

P 0,8 Design pressure (N/mm2)

Dm 2886,78 Main diameter of the shell

Ri 1440 Inside radius of the shell

R el. head 2468,1969 Radius of the head

e 7,78 The min. head thickness

et 5,56 Thickness assumed for the branch

e't 5,56 Thickness assumed for the internal protruding branch portion

(σtensile)shell 490 Tensile stress (N/mm2)

(σyield)shell 355 Yield stress (N/mm2)

f 163,3 Allowable stress for the shell (N/mm2)

(σtensile)branch 415 Tensile stress (N/mm2)

(σyield)branch 240 Yield stress (N/mm2)

f t 138,3 Allowable stress for the branch (N/mm2)

lt 200 Available length of branch (project length)

l't 0 Length of internal protruding branch portion (project length)

d 60,3 Nozzle outer diameter

di 51,18 Insider diameter of the branch

dm 55,74 Mean diameter of the branch

d'm 55,74 Mean diameter of the internal protruding branch portion

ko 1 Coefficient derived from graph C5.1.3 pg 772

δ 0,37 Value for ko (untile 4 ko is 1)

c 1 Corrosion allowance

L 129,4 Length of the shell contributing to the strength of the opening

l 15,9

l' 0

S 877,1

Ч° 3,7 Projection angle

St 126,3

G 198343,7

160.330,4 ≥ 158.675,0 opening is adequate

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5.3.MANHOLE

CODAP Division 1, Volume 2, Part C


Dm 2889 Main diameter of the shell


e 10 The shell thickness

et 20 Thickness assumed for the branch

e't 20 Thickness assumed for the internal protruding branch portion

















l 105,9

l' 0

S 1451,2

St 2298,6

G 703956,8

603.314,5 ≥ 563.165,4

opening is

adequate

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5.4.N1-N2-N4 NOZZLES
























l 28,4

l' 0

S 1451,2

St 320,2

G 316348,5


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5.5.N3-N5-N7-K4 NOZZLES
























l 15,9

l' 0

S 1451,2

St 138,7

G 276251,8


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5.6.N6 NOZZLE
























l 23,3

l' 0

S 1451,2

St 246,4

G 297428,9


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6. LIFTING LUG STRESS CALCULATIONS

Material : P355NL1

Pad : P355NL1

Lug Thickness : 17 mm

Pad Thickness : 10 mm

Total Tank Weight : 15,000 kg (100 m3 Approx. Tank Weight)

Numbers of Lifting Lugs : 1

Note: It is known that there is 3 lifting lugs on the drawing however only one of them is used

in Isısan Factory.

Design Load (F) : 15,000 kg ≈ 147.2 kN

Distance, centerline of hole

to component (h) : 70 mm

Diameter of hole (d) : 90 mm

Radius of lug (r) : 95 mm

Corrosion (c ) : 1

223050.1

345 mmNSK

em (Sb: allowable bending stress of lug)

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CALCULATIONS

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21152

mmNemem

(S: allowable shear stress of lug)

Required Thickness For Shear

110)2/9095()10115(2

2.14710

)2/()(26

36

c

drSFt = 14.2 mm

Required Thickness For Bending

110)190(10230

70)2.147(610)(

6 623

62

cLS

hFtb

= 8.6 mm

17 mm thickness is ok.

7. SADDLE STRESS CALCULATIONS

Saddle material: S355J2 Saddle construction is: Centered web Saddle allowable stress: Ss = 140 MPa Saddle yield stress: Sy = 355 MPa Saddle distance to datum: 2.300 mm Tangent to tangent length: L = 14.351,6 mm Saddle separation: Ls = 9.650 mm Vessel radius: R = 1.450 mm Tangent distance left: Al = 2.350,8 mm Tangent distance right: Ar = 2.350,8 mm Saddle height: Hs = 1.700 mm Saddle contact angle: = 120 ° Wear plate thickness: tp = 10 mm Wear plate width: Wp = 600 mm Wear plate contact angle: w = 132 ° Web plate thickness: ts = 15 mm Base plate length: E = 2.600 mm Base plate width: F = 400 mm Base plate thickness: tb = 17 mm Number of stiffener ribs: n = 5 Largest stiffener rib spacing: di = 636,83 mm Stiffener rib thickness: tw = 12,7 mm Saddle width: B = 400 mm Anchor bolt size & type: 20 mm Anchor bolt material:

100 m³ BUTANE TANK

CALCULATIONS

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Anchor bolt allowable shear: 103,421 MPa Anchor bolt corrosion allowance: 0 mm Anchor bolts per saddle: 2 Base coefficient of friction: = 0,45

Weight on left saddle: operating corr = 5.682,15 kg, test new = 55.512,9 kg Weight on right saddle: operating corr = 5.897,15 kg, test new = 55.728,36 kg Weight of saddle pair = 1.075,01 kg

Notes: (1) Saddle calculations are based on the method presented in "Stresses in Large Cylindrical Pressure Vessels on Two Saddle Supports" by L.P. Zick.

Bending + pressure between saddles (MPa)

Bending + pressure at the saddle (MPa)

Load Vessel condition

S1 (+)

allow (+)

S1 (-)

allow (-)

S2 (+)

allow (+)

S2 (-)

allow (-)

Weight Operating 58,545 153,591 0,935 71,674 61,334 153,591 3,724 71,674 Weight Test 79,133 319,5 8,836 71,674 106,076 319,5 35,779 71,674

Tangential shear (MPa)

Circumferential stress (MPa)

Stress over saddle (MPa) Splitting (MPa)

Load Vessel condition

S3 allow S4 (horns)

S4 (Wear plate)

allow (+/-) S5 allow S6 allow

Weight Operating 2,982 122,873 -24,174 -40,108 230,386 7,833 177,5 1,259 93,333 Weight Test 27,911 255,6 -228,448 -379,019 319,5 74,024 319,5 11,901 319,5

Load Case 1: Weight ,Operating

Longitudinal stress between saddles (Weight ,Operating, right saddle loading and geometry govern) S1 = +- 3*K1*Q*(L/12) / (*R2*t) = 3*0,2956*57.831,34*(14.351,6/12) / (*1.4452*10) = 0,935 MPa Sp = P*R/(2*t)

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= 0,8*1.440/(2*10) = 57,61 MPa Maximum tensile stress S1t = S1 + Sp = 58,545 MPa Maximum compressive stress (shut down) S1c = S1 = 0,935 MPa Tensile stress is acceptable (<=1*S*E = 153,591 MPa) (153,591 MPa) Compressive stress is acceptable (<=1*Sc = 71,674 MPa) Longitudinal stress at the right saddle (Weight ,Operating) Le = 2*(Left head depth)/3 + L + 2*(Right head depth)/3 = 2*728,23/3 + 14.351,6 + 2*730/3 = 15.323,76 mm w = Wt/Le = 113.554,22*10/15.323,76 = 74,1 N/cm Bending moment at the right saddle: Mq = w*(2*H*Ar/3 + Ar

2/2 - (R2 - H2)/4) = 74,1/10000*(2*730*2.350,8/3 + 2.350,82/2 - (1.4502 - 7302)/4) = 26.045,7 N-m S2 = +- Mq*K1'/ (*R2*t) = 26.045,7*1e3*9,3799/ (*1.4452*10) = 3,724 MPa Sp = P*R/(2*t) = 0,8*1.440/(2*10) = 57,61 MPa Maximum tensile stress S2t = S2 + Sp = 61,334 MPa Maximum compressive stress (shut down) S2c = S2 = 3,724 MPa Tensile stress is acceptable (<=1*S = 153,591 MPa) Compressive stress is acceptable (<=1*Sc = 71,674 MPa) Tangential shear stress in the shell (right saddle, Weight ,Operating) Qshear = Q - w*(a + 2*H/3) = 57.831,34 - 7,41*(2.350,8 + 2*730/3) = 36.804,75 N S3 = K2,2*Qshear/(R*t) = K2,2*36.804,75/(1.445*10) = 2,982 MPa Tangential shear stress is acceptable (<= 0.8*S = 122,873 MPa)

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Circumferential stress at the right saddle horns (Weight ,Operating) S4 = -Q/(4*(t+tp)*(b+1,56*Sqr(Ro*t))) - 3*K3*Q/(2*(t2+tp

2)) = -57.831,34/(4*(10+10)*(400+1,56*Sqr(1.450*10))) - 3*0,0529*57.831,34/(2*(102+102)) = -24,174 MPa Circumferential stress at saddle horns is acceptable (<=1,5*Sa = 230,386 MPa) Circumferential stress at the right saddle wear plate horns (Weight ,Operating) S4 = -Q/(4*t*(b+1,56*Sqr(Ro*t))) - 3*K3*Q/(2*t2) = -57.831,34/(4*10*(400+1,56*Sqr(1.450*10))) - 3*0,0434*57.831,34/(2*102) = -40,108 MPa Circumferential stress at wear plate horns is acceptable (<=1,5*Sa = 230,386 MPa) Ring compression in shell over right saddle (Weight ,Operating) S5 = K5*Q/((t + tp)*(ts + 1,56*Sqr(Ro*tc))) = 0,7603*57.831,34/((10 + 10)*(15 + 1,56*Sqr(1.450*20))) = 7,833 MPa Ring compression in shell is acceptable (<= 0,5*Sy = 177,5 MPa) Saddle splitting load (right, Weight ,Operating) Area resisting splitting force = Web area + wear plate area Ae = Heff*ts + tp*Wp = 22,3*1,5 + 1*60 = 93,45 cm2 S6 = K8*Q / Ae = 0,2035*57.831,34 / 9.344,9986 = 1,259 MPa Stress in saddle is acceptable (<= (2/3)*Ss = 93,333 MPa) Load Case 2: Weight ,Test

Longitudinal stress between saddles (Weight ,Test, right saddle loading and geometry

govern)

S1 = +- 3*K1*Q*(L/12) / (*R2*t)

= 3*0,2956*546.508,57*(14.351,6/12) / (*1.4452*10)

= 8,836 MPa

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CALCULATIONS

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Sp = P*R/(2*t)

= 0,98*1.440/(2*10)

= 70,297 MPa

Maximum tensile stress S1t = S1 + Sp = 79,133 MPa

Maximum compressive stress (shut down) S1c = S1 = 8,836 MPa

Tensile stress is acceptable (<= 0,9*Sy = 319,5 MPa) (319,5 MPa)

Compressive stress is acceptable (<=1*Sc = 71,674 MPa)

Longitudinal stress at the right saddle (Weight ,Test)

Le = 2*(Left head depth)/3 + L + 2*(Right head depth)/3

= 2*728,23/3 + 14.351,6 + 2*730/3

= 15.323,76 mm

w = Wt/Le = 1.090.904,24*10/15.323,76 = 711,9 N/cm

Bending moment at the right saddle:

Mq = w*(2*H*Ar/3 + Ar2/2 - (R2 - H2)/4)

= 711,9/10000*(2*730*2.350,8/3 + 2.350,82/2 - (1.4502 - 7302)/4)

= 250.218,9 N-m

S2 = +- Mq*K1'/ (*R2*t)

= 250.218,9*1e3*9,3799/ (*1.4452*10)

= 35,779 MPa

Sp = P*R/(2*t)

= 0,98*1.440/(2*10)

= 70,297 MPa

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CALCULATIONS

Date : 13.06.2011 Page No : 20/23

Maximum tensile stress S2t = S2 + Sp = 106,076 MPa

Maximum compressive stress (shut down) S2c = S2 = 35,779 MPa

Tensile stress is acceptable (<= 0,9*Sy = 319,5 MPa)

Compressive stress is acceptable (<=1*Sc = 71,674 MPa)

Tangential shear stress in the shell (right saddle, Weight ,Test)

Qshear = Q - w*(a + 2*H/3)

= 546.508,57 - 71,19*(2.350,8 + 2*730/3)

= 344.508,2 N

S3 = K2,2*Qshear/(R*t)

= K2,2*344.508,2/(1.445*10)

= 27,911 MPa

Tangential shear stress is acceptable (<= 0.8*S = 255,6 MPa)

Circumferential stress at the right saddle horns (Weight ,Test)

S4 = -Q/(4*(t+tp)*(b+1,56*Sqr(Ro*t))) - 3*K3*Q/(2*(t2+tp2))

= -546.508,57/(4*(10+10)*(400+1,56*Sqr(1.450*10))) - 3*0,0529*546.508,57/(2*(102+102))

= -228,448 MPa

Circumferential stress at saddle horns is acceptable (<= 0,9*Sy = 319,5 MPa)

Circumferential stress at the right saddle wear plate horns (Weight ,Test)

S4 = -Q/(4*t*(b+1,56*Sqr(Ro*t))) - 3*K3*Q/(2*t2)

= -546.508,57/(4*10*(400+1,56*Sqr(1.450*10))) - 3*0,0434*546.508,57/(2*102)

= -379,019 MPa

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CALCULATIONS

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Ring compression in shell over right saddle (Weight ,Test)

S5 = K5*Q/((t + tp)*(ts + 1,56*Sqr(Ro*tc)))

= 0,7603*546.508,57/((10 + 10)*(15 + 1,56*Sqr(1.450*20)))

= 74,024 MPa

Ring compression in shell is acceptable (<= 0,5*Sy = 319,5 MPa)

Saddle splitting load (right, Weight ,Test)

Area resisting splitting force = Web area + wear plate area

Ae = Heff*ts + tp*Wp

= 22,3*1,5 + 1*60

= 93,45 cm2

S6 = K8*Q / Ae

= 0,2035*546.508,57 / 9.344,9986

= 11,901 MPa

Stress in saddle is acceptable (<= 0,9*Sy = 319,5 MPa)

Shear stress in anchor bolting, one end slotted

Maximum seismic or wind base shear = 0 N

Thermal expansion base shear = W* = 63.102,48 * 0,45= 28.396,12 N

Corroded root area for a 20 mm bolt = 2,3484 cm2 ( 2 per saddle )

Bolt shear stress = 28.396,11/(234,8382*2) = 60,459 MPa

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CALCULATIONS

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Anchor bolt stress is acceptable (<= 103,421 MPa)

Web plate buckling check (Escoe pg 251)

Allowable compressive stress Sc is the lesser of 140 or 128,335 MPa: (128,335)

Sc = Ki*2*E/(12*(1 - 0,32)*(di/tw)2)

= 1,28*2*19,99E+04/(12*(1 - 0,32)*(636,83/15)2)

= 128,335 MPa

Allowable compressive load on the saddle

be = di*ts/(di*ts + 2*tw*(b - 1))

= 25,0719*0,5906/(25,0719*0,5906 + 2*0,5*(15,748 - 1))

= 0,501

Fb = n*(As + 2*be*tw)*Sc

= 5*(4.889,5 + 2*12,73*15)*128,335

= 3.382.439,81 N

Saddle loading of 551.779,72 N is <= Fb; satisfactory.

Primary bending + axial stress in the saddle due to end loads (assumes one saddle

slotted)

b = V * (Hs - xo)* y / I + Q / A

= 0 * (1.700 - 1.199,14)* 200 / (1e4*33.887,76) + 57.831,34 / 62.119,6

= 0,931 MPa

The primary bending + axial stress in the saddle <= 140 MPa; satisfactory.

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CALCULATIONS

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Secondary bending + axial stress in the saddle due to end loads (includes thermal

expansion, assumes one saddle slotted)

b = V * (Hs - xo)* y / I + Q / A

= 28.396,11 * (1.700 - 1.199,14)* 200 / (1e4*33.887,76) + 57.831,34 / 62.119,6

= 9,325 MPa

The secondary bending + axial stress in the saddle < 2*Sy= 710 MPa; satisfactory.

Saddle base plate thickness check (Roark sixth edition, Table 26, case 7a)

where a = 636,83, b = 192,5 mm

tb = (1*q*b2/(1,5*Sa))0,5

= (3*0,531*192,52/(1,5*140))0,5

= 16,76 mm

The base plate thickness of 17 mm is adequate.

Foundation bearing check

Sf = Qmax / (F*E)

= 551.779,72 / (400*2.600)

= 0,531 MPa

Concrete bearing stress < 11,432 MPa ; satisfactory.

100m3 Design Calculation

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