ML20063P818
| ML20063P818 | |
| Person / Time | |
|---|---|
| Site: | Fermi  |
| Issue date: | 09/03/1975 |
| From: | MCDONALD ENGINEERING ANALYSIS CO., INC. |
| To: | |
| Shared Package | |
| ML20063P810 | List: |
| References | |
| ME-229, NUDOCS 8210150343 | |
| Download: ML20063P818 (54) | |
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o CERTIFICATION STATEMENT T
This Goulds Pumps, Inc. Model VIT - 8 x 14 JMC - 2 Stage Emergency Equipment Service Water Pump has been analyzed in accordance with Sargent and Lundy Specification No. 3071-134 Form 350A, Appendix A and B, including Attachment A, the ASME Section III, Class 3 Code, ASME Code Case 1636, and accepted good practice in seismic-stress ana-lysis.
The pump meets all requirements of the specification and the ASME Code and will perform its intended function during both OBE and DBE seismic loads combined with maximum nozzle and normal loads.
C')
C.K. Mcdonald, Ph.D., P.E.
Alabama Registration No. 9586 June 21, 1975
)
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/ An -/ 43 TABLE OF CONTENTS 1.
INTRODUCTION 1
2.
SUMMARY
OF RESULTS 3
2.1 Normal Case 3
4 2.2 Upset Case 5
2.3 Faulted Case 6
3.
FREQUENCY ANALYSIS 6
3.1 Dynamic Model 8
3.2 Computer Input 8
3.3 Frequency Results 13 4.
LOADING CRITERIA AND ANALYSIS 13 4.1 Seismic Loadings and Analysis 14 4.2 Nozzle Loads 14 4.3 Internal Pressure 14 4.4 Other Pump Normal Loads 15 5.
STRESS AND DEFLECTION ANALYSIS 15 5.1 Stresses in Column 16 5.2 Stresses in Column Flange and Bolting 18 5.3 Pump Casing Flanges and Bolting 21 5.4 Pump Casing 21 5.5 Stresses in Shaft 23 5.6 Motor Hold Down Bolts 24 5.7 Stresses in Shaft Key 24 5.8 Shaft. Deflections 25 5.9 Impeller Deflection and Clearance 25 5.10 Nozzle Analysis 26 5.11 Nozzle Flange Analysis Continued........
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TABLE OF CONTENTS (Cont.)
26 5.12 Analysis of Coupling 26 5.13 Analysis of Anchor Bolts 29 5.14 Analysis of Discharge Head Hold Down Bolts 30 5.15 Analysis of Discharge Head 31 5.16 Analysis of Discharge Head Flange 32 5.17 Analysis of Discharge Head Weld APPENDIX A - STRESS AND DEFLECTION COMPUTER OUTPUT APPENDIX B - ORTHOGONALITY CHECKS APPENDIX C - HODE SHAPES Appendix D - ANAI,YSIS OF FLAT FACEDiFLANGES FOR EXTERNAL FORCES
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- 1. INTRODUCTION This report covers the seismic, stress, and deflection analysis of Goulds Pumps. Inc. Model VIT - 8 x 14 JMC - 2 Stage pump for Detroit Edison Company. The analysis is directed toward verifying both the structural adequacy and functional espability of the pump.
The natural frequencies of the pump are determined by developing a lumped mass model of the pump and motor assembly.
This model is prepared for analysis by computer.
The computer code ICES-STRUDL is used to perform the frequency analysis.
Once the frequencies are determined, the seismic loads are obtained f rom the response spectra curve Jurnished for this job.
The maximum seismic loads within plus or minus.'0% of the frequencies are used for the seismic analysis. Damping of 1% OBE and 2% DBE is used.
A dynamic modal analysis of the pump and motor assembly is made, using the same model as used for the frequency analysis.
The computer code ICES-STRUDL is also used for the dynamic model analysis. The computer output l
1 includes all forces, moments, and deflections at all joints for each member of the model.
The nozzle loads and other normal operating loads are combined with the seismic loads for the vertical and both lateral directions.
This combination is made for both the OBE and DBE loadings.
The resulting stresses are com-pared to the allowable stresses given by the specification.
The resulting deflections are compared to operating clearances or other operating criteria.
A summary of the deflections and stresses is given in Section 2 and the detailed calculations are contained in Section 5 of this report.
ASME Code Case 1636 is used for the allowable stresses for the pressure parts.
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- 2.
SUMMARY
OF RESULTS 2.1 Normal Case ACTUAL _
ALLOWABLE 1,896 15,000 Maximum Column Stress, PSI 2,943 17,500 Max. Column Flange Stress, PSI 6,817 25,000 Bolt Stress, PSI 8,158 14,000 Max. Pump Flange Stress, PSI Bolt Stress, PSI 23,985 25,000 2,333 14,000 Max. Pump Casing Stress, PSI Max. Shaft Stress, PSI 8,431 10,000 Motor Hold Down Bolt Stress, PSI 0
25,000 Tensile 532 12,500 Shear 7,945 10,000 Shaft Key Stress, PSI 0
.05 Shaft Deflection, Inches 0
.009 Impeller Deflection (clearance), Inches 14,091 15,000 Nozzle Stress, PSI 16,532 26,250 Nozzle Flange Stress, PSI 14,090 15,000 Bolt Stress, PSI Anchor Bolt Stress, PSI 1,997 20,000 Tensile 1,747 10,000 Shear Discharge Head Hold Down Bolts 2,077 25,000 Tensile 1,614 12,500 Shear 125 15,000 Discharge Head Stress, PSI Discharge Head Flange Stress, PSI 3,129 26,250 Discharge Head Weld Stress, PSI 2,774 15,000
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e 2.2 OBE AC5fAL ALLOWABLE Maximum Columm Stress, PSI 8,027 24,750 Max. Column Flange Stress, PSI 14,162 28,875 Bolt Stress, PSI 18,725 27,500 Max. Pump Flange Stress, PSI 9,323 23,100 Bolt Stress, PSI 24,861 27,500 Max. Pump Casing Stress, PSI 2,333 14,000 Max. Shaft Stress, PSI 9,276 10,000 Motor Hold Down Bolt Stress. PSI Tensile 20,066 27,500 Shear 4,586 13,750 Shaft Key Stress, PSI 7,945 10,000 Shaft Deflection, Inches
.023
.05 Impeller Deflection (clearance), Inches 0
.009 Nozzle Stress, PSI 14,091 24,750 Nozzle Flange Stress, PSI 16,532 28,875 Bolt Stress, PSI 14,080 16,500 Anchor Bolt Stress, PSI Tensile 20,510 22,000 Shear 3,977 11,000 Discharge Head Hold Down Bolts Tensile Stress 21,333 27,500 Shear Stress 3,314 13,750 Discharge Head Btress, PSI 20,866 24,750 Discharge Head Flange Stress, PSI 14,062 28,875 Discharge Head Weld Stress, PSI
.2.003 16,500
l}!/-Nj e 2.3 DBE ACTUAL ALLOWABLE Maximum Column Stress, PSI 10,280 36,000 Max. Column Flange Stress, PSI 17,951 42,000 Bolt Stress, PSI 22,761 37,500 600 Max. Pump Flange Stress, PSI 9,630 33,500 Bolt Stress, PSI 25,091 37, Max. Pump Casing Stress, PSI 2,333 14,000 Max. Shaft Stress, PSI 9,574 10,000 Motor Hold Down Bolt Stress, PSI Tensile 14,336 37,500 Shear 3,479 18,750 Shaft Key Stress, PSI 7,945 10,000 Shaft Deflection, Inches
.035
.05 Impeller Deflection (clearance) Inches
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.009 Nozzle Stress, PSI 14,091 36,000 Nozzle Flange Stress, PSI 16,532 42,000 Bolt Stress, PSI 14,080 22,500 Anchor Bolt Stress, PSI Tensile 20,285 30,000 Shear 3,530 15,000 Discharge Head Hold Down Bolts Tensile Stress 21,099 37,500 Shear Stress 2,974 18,750 Discharge Head Stress, PSI 15.441 36,000 Discharge Head Flange Stress, PSI 13,591 42,000 Discharge Head Weld Stress, PSI 11.551 22,500
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- 3. FREQUENCY ANALYSIS A dynamic model is developed and a dynamic frequency analysis is performed using the computer code ICES-STRUDL, a large general purpose code developed by the Civil Engineering Department of the Massachusetts Institute of 'echnology. The capability of the code to perform frequency analysis (and modal analysis) has been well verified. These verification checks are contained in a set of class notes for a course entitled
" Dynamic Seismic Analysis of Nuclear Power Plant Components" taught by C. K. Mcdonald at the University of Alabama in Birmingham. These verificat-ion checks may be obtained by writing or calling T. F. Talbot at the University of Alabama in Birmingham.
3.1 Dynamic Model The dynamic model used to obtain the frequencies for both the vertical and lateral directions is given in Figure 2, page 7. This model is a typical lumped mass model where the masses are connected by massless beams.
The mass of the motor is lumped at the motor center of gravity. The mass of the pump support above the floor is lumped at the joint indicated on the sketch. The mass of the lower column is lumped at the column flanges.
The mass of the pump bowl assembly is lumped as indicated on the sketch.
The joints and members of the model are numbered to facilitate the 1
computer input. The joint numbers are enclosed in circles.
The properties of each member and the numerical values of the masses are obtained from the detail drawings. These data are included in the computer input.
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20 location of Shaft Bearing Figure 2 - Dynamic Model of Pump Y,5~(,-l(L 3.2 Computer Input The computer input data are given on pages 9 and 10. The data are believed to be self explanatory for those familiar with this type of computer code. However, a brief description of the input data is given here for the benefit of those not familiar with the format of STRUDL.
The first number in the joint coordinates is the joint number, the and the last is the Y coordinate. The first second is the X coordinate, number in the member incidences is the member number, the second is the start joint of the member and the last is the end joint of the member.
In the member properties, AX is the cross-sectional area of the member, IZ is the area moment of inertia, and SZ is the section modulus.
The dynamic degrees of freedom statement limits the degrees of freedom of each joint to the degree of freedcm specified. The dump orthog-onality statement gives a computational check on the accuracy of the frequency calculations. The off diagonal terms in the matrices given in Appendix B should be zero. The residuals in the terms give an indication of the accuracy of the computations.
3.3 Frequency Results i
The f requencies for both the lateral and vertical directions are F ven below.
VERTICAL FREOUENCIES LATERAL FREOUENCIES
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- 4. LOADING CRITERIA AND ANALYSIS i
4.1 Seismic Loadings and Analysis (a) Lateral Case The seismic loadings for the lateral case are obtained from the response spectra curve, using 1% damping.for OBE and 2% for DBE. These loadings were obtained using the frequencies calculated in Section 3 and adding plus or minus 10% to these frequencies for the worst case seismic loads. The loads used for this analysis are:
MODE FREQUENCY, CPS SEISMIC LOADING IN C's OBE DBE N-S E-W N-S E-W 1
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The N-S and E-W load cases for OBE and DBE are then combined by hand.
The computer input data are given on page 11 and 12 and the computer output data are given in Appendix A.
(b) Vertical Case The frequencies given on page 8 show that none of the frequencies for the vertical direction is below the rigid cutoff. Thus, a dynamic modal analysis is not necessary for the vertical case. The loads used are:
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.18 g The vertical seismic calculations are performed by hand and are included in the detailed calculations given in Section 5 of this report.
4.2 Nozzle Loadings In accordance with Goulds Drawing D 1691 the nozzle loads for this pump are:
MAXIMUM RESULTANTS M, Bending Moment, In-lbs.
18,000 x
25,456 M, Bending Moment, In-lbs.
18,000 2
M, Torsional Moment, In-lbs.
36,000 y
V, V, Shear, Lbs.
600 x
y 849 V, shear, lbs.
600 z
P, axial load, lbs.
1,200 The loads are used in the detailed calculations given in Section 5.
4.3 Internal Pressure Loads The internal pressure for design is 125 psig at 100 F.
4.4 Other Pump Normal Loads The motor horsepower is 100 at 1775 R.P.M.
The torque is:
3549 in-lbs.
63000(100)
T
=
=
1775 I
1
_. )Nh-/lf
- 5. STRESS AND DEFLECTION ANALYSIS The detailed stress and deflection calculations are contained in this section.
5.1 Stresses in Column The maximum stresses in the column occur at the support, member 4 at joint 4 (a) Normal Loads The stress urder normal loads is due to internal pressure.
This stress is (using the formula given in ND-3324.3):
S PR
+
.6P 125(4.036)
+.6(125)
=
1,896 psi
=
=
t
.277 The allowable stress for SA 106 Gr B pipe is 15,000 psi Normal.
(b) OBE Loads The maximum stress under this condition is the longitudinal bending and tension due to seismic plus deadweight plus pressure.
The stresses are:
= /(66336)2
+ (71002)2 M
97168
=
Bending due to seismic Si= 97168 6,615 psi (see App. A)
=
14.69 Axial due to seismic + dead S2=
1.12(3250)=
501 psi 7.26 Long. due to pressure S3 = 125(4.036) =
911 psi 2(.277)
Total Long. Stress, psi 8,027 Memb. + bending The allowable stress is 1.65(15000) = 24,750 psi (c) DBE Loads M = h (88032)2 + (95483)2 129,872 in-lbs.
=
.~
/ 2 CG -/Il6 8,841 psi (See App. A) 129872 Bending due to seismic S
=
=
g 14.69 528 psi Axial due to seismic + dead S2 1.18(3250)
=
=
7.26 911 psi Long. due to pressure S3
=
Total longitudinal stress, PSI 10,280 psi Memb. + bending The allowable stress is 2.4(15000) = 36,000 psi.
5.2 Stresses in Column Flange and Bolting (a) Normal Loading The design pressure is 125 psig.
The design is per Appendix XI, part B.
The nomenclature is per the ASME Code.
See Appendix D for a discussion of the 4(3250)
= 50 P
= 125 + 50 = 175 method used here.
Peq
=
FD 3.1416(9.079)
.785(8.071)2(175) 8,949 lbs.
=
Hp
=
.785(9.079)2(175) 11,324 lbs.
=
H
=
2,375 lbs.
11324 - 8949
=
HT
=
1.275"; hT =.638"; hC =.690" 0
(0 ring gasket);
hp
=
Hg
=
12,925 in-lbs.
8949(1.275) + 2375(.638)
=
M
=
p 2,943 psi Max. Flange Stress 6(12925)
=
SR
=
8(.875))
(1.00)2(3.14(10.62)
The flange is an optional type and is treated as a loose type.
The maximum stress for loose type is the radial stress.
The allowable stress for the SA 515 Gr 70 flange is 17,500 psi.
Bolt
.785(9.079)2(125) = 8,08S
.785(8.071)2(125) = 6.392 H
=
H
=
p 1696( 638) + 6392(1.275) = 9,232 HT 1,696 M
=
=
p 13,380 H' =.785(9.079)2(50) = 3,235 lbs.
9232 HC
=
=
.690 24703 lbs.
Sb= 24703_
= 6,817 psi k'm1' = 3235 + 13380 + 8088
=
12 (.302 )
/
[ c-k The allowable stress for the 3/4" diameter SA 193 B 7 bolts is 25,000 psi.
(b) OBE Loading See Appendix A for the values of the moments on the flange at joint 4, member 4.
(71002)2 M =Y(66336)2 97,168
+
717 psig 16(97168)
+
4(1.12)(3250)
=
P,q 3.1416(9.079)1 3.1416(9.079)2
=
PFD 717 + 125 = 842 psig
=
.785(8.071)2(842) 43056 lbs.
Hp
=
=
.785(9.079)2(842) 54483 lbs.
H
=
=
11427 lbs.
H
=
T 62,187 in-lbs.
11427(.638) + 43056(1.275)
M
=
=
p 14,162 psi 6(62187)
=
SR
=
(1.00)4(3.14(10.62) - 8(.875))
28,875 psi The allowable stress is 1.65(17500)
=
Bolt
.785(8.071)2(125) = 6,392 H
=
D
.785(9.079)2(125) 8,088 H
=
=
1,696 HT
=
1696(.638) + 6392(1.275) = 9,232 M
=
p 13,379 9232 HC
=
=
.69
.785(9.079)2(717) 46,394 lbs.
H'
=
=
67861 lbs.
46394 + 13379 + 8088 Wmi'
=
=
67861
= 18,725 psi Sb
=
12(.302) 27,500 psi The allowable stress for SA 193 B 7 bolts is 1.1(25000)
=
[M b- - C 1,
(c) DBE Loads
~
M = /(8803')2 (95483)2 129,872 in-lbs.
+
943 psi' 16(129872)
_ +
4(1.18)(3250)_
P
=
=
eq 3.1416(9.079)J 3.1416(9.079)4 943 + 125 1068 psig P
=
=
pp Flange
.785(8.071)2(1068) 54,613 lbs.
H
=
=
D
.785(9.079)2(1068) 69,106 lbs.
H
=
=
14,493 lbs.
HT
=
14493(.638) + 54613(1.275) 78,878 in-lbs.
M
=
=
p SR 6(78878) 17,951 psi
=
=
(1.00)2(3.14(10.62) - 8(.875)
The allowable stress is 2.4(17500) 42,000 psi
=
Bolt
.785(9.079)2(943)
H' 61,018 lbs.
=
=
Wmi' 82,485 61018 + 13379 + 8088
=
=
22,761 psi Sb 82485
=
12(.302)
The allowable stress is 1.5(25000) = 37,500 psi.
5.3 Pump Casing Flanges and Bolting The worst stressed flange is located at the upper end of the casing where the casing connects to the column, joint 12 on the computer model.
(a) Normal Loading The design pressure is 125 psig. The design is per Appendix XI, part B.
The nomenclature is per the ASME Code.
Peq. = (4)(920) 8 psig
=
3.1416(11.91)2
. l
/(((
r.V 133 psig 125 + 8 Ppp
=
=
.785(8.071)2(133) 6,,801 lbs.
Hp
=
=
.785(11.91)2(133) 14,810 lbs.
H
=
=
6801 8,009 14810 H
=
=
T
= 1. 200') hC
.565" 2.400" hT 0
(0 ring gasket) ; h
=
H
=
=
g D
25,933 in-lbs.
6801(2.400) + 8009(1.200)
M
=
=
p 8,158 psi Max. Flange Stress 6(25933)
S
=
=
R
(.75)d(3.14(12.87) - 8(.813))
Bolt
.785(8.071)2(125) 6,392 H
=
=
D
.785(11.91)2(125) 13,919 H
=
=
7,527 HT
=
24,373 7527(1.200) + 6392(2.400) h
=
=
p 43,138 24?73 H
=
=
C
.565
.785(11.91)2(8) 891 lbs.
H'
=
=
57,948 891 + 43138 + 13919 Wmi
=
23,985 57948 S
=
=
b 8(.302)
The flange is an optional type and is treated as loose type. The maximum stress for loose type is the radial stress.
The allowable stress for the 3/4" diameter SA 193 B 7 bolts is 25,000 psi and for the SA 216 WCB flange (using a.8 casing factor) is 14,000 psi.
(b)
OBE Loads (4959)2
=f(3338)2 3,97g
+
M 16(5978)
+
4(1.12)(920) 27 psig
=
P
=
'9 3.1416(11.91) 3.1416(11.91) 152 psig 27 + 125 P
=
=
)2SS-/2f Flange
.785(8.071)2(152) 7773 H
=
=
D
.785(11.91)2(152) 16,925 H
=
=
9,152 HT
=
29,638 9152(1.200) + 7773(2.400)
M
=
=
p 9,323 6(29638)
=
SR
=
(.75)Z(3.14(12.87) - 8(.813))
The allowable stress is 1.65(14000) 23,100 psi
=
Bolt
.785(8.071)2(125) 6,392 H
=
=
D
.785(11.91) (125) 13,919 H
=
=
7,527 HT
=
6392(2.400) + 7527(1.200) 24,373 M
=
=
p 43,138 24373 HC 2
=
.565
.785(11.91)2(27) 3006 H'
=
=
60063 lbs.
3006 + 43138 + 13919 Wmi'
=
=
24,861 60063_
Sb
=
=
8(.302) 27,500 psi The allowable stress is 1.1(25000)
=
(c) DBE Loads (5917)2 f(4390)2 7368
+
M
=
=
psig 16(7368)
P,q 3.1416(11.91)3 3.1416(11.91)2
=
157 psig 32 + 125 Pyp
=
=
Flange 9,630 psi SR 157(9323)
=
=
152 33,600 psi The allowable stress is 2.4(14000)
=
/N
- C[
f Bolt
.785(11.91)2(32)
'3,563 H'
=
=
60,620
~
Wal' 3563 + 43138 + 13919
=
=
Sb 60620 25,091 psi
=
=
8(.302)
The allowable stress is 1.5(25000) 37,500 psi.
=
5.4 Pump Casing Stress The pump casing is of complex geometry. An approximate estimate of the membrane stress can be obtained by taking a 1" strip at the maximum diameter.
Thus, the stress is:
.375" R
7" t
=
=
2,333 psi Normal, OBE, and DBE (seismic is negligible) 125(7 )
S
=
=
.375 The allowable stress for the SA 216 WCB casing is 14,000 psi including the casting factor.
5.5 Stresses in Shaft The normal operating torque is 3549 in-lbs.
The maximum stress occurs at the coupling which is located near the pump support (joint 4 on the model).
(a) Normal 3.1416(1.4975)4 1.761 in.
.494 in.
A J
=
=
=
32 5381 psi 3549(.749)
V
=
=
.494 Downthrust + deadweight of shaft and impellers = 8800 lbs.
4997 psi 8800 S
=
=
1.761 The combined stress is:
Vc4997)2 2
. c33,13
,,431,,,
S
=
c 4.
2 f( < / A$
The allowable stress is 10,000 psi.
(b) OBE The bending stress is (assuming the shaft bends with the column):
.749(97168) 1,132 psi S
=
=
b 4.375(14.69)
The stress due to vertical seismic is:
.12(375) 26 psi S
=
=
y 1.761 The total axial stress is:
4997 + 1132 +
26 6155 psi Sa
=
The combined stress is:
S
- Y (6155)2 (5381)2 9,276 psi
+ 76155
+
=
e 4
2 The allowable stress is 10,000 psi.
(c) DBE The bending stress is (assuming the shaf t bends with the column):
b
.747(129872) 1,514 psi S
=
=
(4.375)14.69 The stress due to vertical seismic is:
.18(375) 38 psi S
=
=
y 1.761 The total axial stress is:
6549 psi 4997 + 1514 + 38 Sa
=
=
The combined stress is:
Y(6549)2 (5381)2
+
6549 9574 psi S
+
=
e 4
4 The allowable stress is 10,000 psi.
/.L.f $ #/2 )
5.6 Motor Hold Down Bolts The motor is fastened to the mount by 4 - 5/8" hold down bolts on a 14.75" bolt circle.
~
(a) Normal A
532 psi shear due to torque 3549 V
=
=
4(7.375)(.2260)
The allowable stresses for SA-193-B7 bolts are 25,000 tensile, 12500 shear, PSI.
(b) OBE
=kk13676)
+ (69615) 2 M
70946 in-lbs.
=
= \\[(706 )2 + (3596 )2 = 3,665 lbs.
V The tensile bolt load is obtained from the equation for anchor bolt stress given by Steel Plate Engineering Data Vol. 2, Useful Information on the Design of Plate Structures, Published by A.I.S.C.
The weight is 1250 lbs.
(1
.12)1250 = 4535 lbs.
4(70946)
P
=
4(14.75) 4 20066 psi tensile V
4535 S
4,054 psi shear 3665
=
=
=
=
i
.2260 4(.226)
The total shear is:
532
+
4054 4586 psi.
=
The allowable stresses are 27,500 tensile and 13,750 shear, PSI.
(c)
= f(13677)2 + (49727)2 51574 M
=
V =Y(707)2+(2568)2 2664
=
(1
.18)(1250) 3240 lbs.
4(51574)
P
=
=
4(14.75) 4 14,336 psi tensile Vi 2664 2,947 psi shear 3240 S
=
=
=
=
.2260 4(.226)
The total shear is:
532 + 2947 3479 psi.
=
The allowable stresses are 37,500 tensile, and 18,750 shear, PSI.
/ A $~4 - / A f 5.7 Stresses in Shaft Key The stresses in the shaft key are due to operating torque and are:
--*l h
- t
_F
+
t=.375" I
d = 1.4975" 7,945 psi shearing stresss, Normal, OBE, DBE 3549(2) (1.1)
V
=
=
1.4975(.375)(1.75)
The allowable stress for the ASTM 582 Tp 416 key is 10,000 psi Normal 10,000 psi OBE, and 10,000 psi DBE.
5.8 Shaft Deflections Good practice in transmission shaft design limits the maximum deflection of a shaft relative to its bearings to.01 inches per foot.
The shaft is supported such that it is rigid between supports and will thus experience the same acceleration as the supports.
This accleration will clearly be less than 2 g OBE or 3 g DBE.
Thus, the maximum deflection is:
1-1/2" Shaft wt. of shaft x Ig =.500 lbs/in l' l l l l l l l 1 l l l l 11111111 l l l l l l l I o
n l,
60.0" max. span l'
5(.500)(2)(60)4
.023" OBE
=
D
=
'84(29000000)(.249) 4
.035" DBE
_5 (. 50_0 ) (3) (60)
=
D
=
384(29000000)(.249)
These deflections are well within the allowable, f
/ A S / * /.2 y
~25-5.9 Impeller Deflections and Clearance The impeller deflections are obtained from the computer putput, Appendix A.
The clearance is.009".
The deflections are:
Upper Impeller Jt. 14 - Jt. 15 = 0" Normal D
=
2 (2.09524)- (2.09524))
-((2.02265) - (2.02265)) = 0.00000" OBE
=
= d(2.75709)- (2. 75709)) 2 -((2.75804)' (2.75804))2 = 0.00000" DBE Lower Impeller D = Jt. 17 - Jt. 18 = 0" Normal
=Y((2.18862)- (2.18862)) 2 - ((2.11343) - (2.11343))2 = 0.00000" OBE
=d(2.88004)-(2.88004))2 -((3.02162)2 - (3.02160)2 = 0.00002" DBE 5.10 Nozzle Analysis The nozzle / casing intersection is a fabricated tee and is analyzed per ND-3652.
The nomenclature is that of the Code.
5.95
.0588 i=
h
.3
=
=
=
88)2/3 3.1416(4.174) (.307) 16.80 Z
=
=
The nozzle loads are given in Section 4.2.
Normal, OBE and DBE The bending moment resolved to the centerline of the tee is:
25456 + 849(15.5) = 38,616 in-lbs.
M
=
b M =/(38616)2+ (36000)2 52,794 in-Ibs.
Equation (8) of ND-3652 gives:
878 + 14023 14091 psi < 15,000 125(8.625)
+
.75(5.95)(52794)
=
=
4(.307) 16.80 The nozzle / casing intersection does not require reinforcing due to internal pressure.
L
/b d6 5.11 Nozzle Flange Analysis The nozzle flange must be checked for the external loads, The flange calculation for the design pre:sure is shown on page 27.
~
The equivalent pressure is calculated per ND-3647.
P,q 16(25456)
+
4(1200) 162 psig Maximum
=
=
3.1416(9.625) 3.1416(9.625) 125 +
162 287 psig Maximum P
=
=
FD 5.12 Analysis of Coupling The coupling which has the maximum bending stress is located near the support baseplate. The coupling has the following dimensions.
0.D. = 2" I.D. = 1.5" J = 3.1416((2)' - (1.50)')
= 1.074 in.
32
.537 in.
I
=
The coupling is much stronger than the shaft, thus it is adequate.
5.13 Analysis of Anchor Bolts The anchor bolts must withstand the overturning moments due to the upper portion of the pump and motor, the lower column portion, and the nozzle loads.
The seismic overturning moments are obtained from Appendix A as the sum of the load on member 3 at joint 4 plus the load on member 4 at joint 4.
The nozzle loads are assumed to act in the worst direction for overturning, as shown below.
The nozzle loads are assumed to act parallel to the worst seismic loads.
/[f[, f)[ SHEET A WELDING NECK FLANGE DESIGN
- 2 7--
DESIGN COldDiflON3 GA5KET cnd BOLTitJG CALCUL ATIONS
(,',,", $*,*[*,*3 D Jgn Pressure,7 : 12s' P s ;> = 2 F 7 G a 'k ['ai'$
Fa'iao De'ad' N=
.roo 8 eign Temperos.sve
/O'o 'p""
b I'D*
l6O b=
, g $""
E #'
.e M areetal SA - los to #/g '* o, p.
y=
4 goo stliag m aie.;oi sA 6% (5 g.
SP}3t. A L WouuD
- 5. S.
m=
- 3. o o Corredon Allo-oace W.: = a,1Gy 3 /, ~2. l A. = $*7 W.3/3. or W.i/5. =
g,, 3 o g
==
DeJon Temp.,5,.
i ~7 6'00 He = 2bvCm!==
l3 O} @
As x=
Q, & I~le ig Atm. Temp. 5,.
3 p.$ o o H
= C'rr/4 = gg g g 1 W==.$(A. + A.)S.
Deden Temp., S.
$ Ocp O W.s '== H, + H = gzqcg W.,==
" ~
"3 Atm. Te mp., S.
I5Ooo G**kd *~'d'h Chk ("*' sed Face ONLY): N== A 5./2rrG =
6 CONDITION LOAD X
LEVER ARM
=
FAOME NT Ho =,s>r/4 =
i 714.p ho =.5( C - B T =
l.515 Mo = Hoho =
2 5 e) G (,
Mc =
H, =
i3Ol2 Ac =.5tC - G) u== Nche =
,3 g 5g 3, o g,3
=
e gy,g g,,
Mr== N - Ho =
3 *) Q hr m.S'[h p hc') =
- j. 2. M Mr = Hrhr =
f.Q4 M.
44h 28
'=
Griket Sess.ng Mc = W kc = sic - G)
=
M.
=
=
"f0' STRESS CALCUL AIs -Operating Conditions (use M)
$HAPE CONSTANTS N'.*.7,. *c'd*.Y *.'.*,' N 1.5 5,.
Long. Hub. 5== f M/Agi', %, l(c 731 K== A/8
/.f(g !h/h.
=,M*
==
5,,
Radiol Fig, Se== $ M/At'
$$T T
=
f,M C F
=,p-5,,
Tong. Fig., Sr = { MY/f') - Z$e ' g'.g g $~
Z
==
.4S y, 4. :, 'd. Y
=
s,.
- T.sts + 5 lor.5 (s + sr1 l Y
4, (, f ;
i
=
f,o h ""*j' STRESS CALCULATION-Gsshet Seating (use M) 47 U
== K. 0 7 '$ e== F/h.
=
1.1 ?9 d=U h. g.' = ~5.4-3 91/9.
=
a.
Long. Hub. Sm um fM/Agi'
{
5,,
Rodial Fig., Se== $M/At'
- h.== V85== l.FR4-V Se, -
Tong. Fig Sr = (MY/f') - ZSe OTHER STRESS FORMULA F ACTOR 5
!*7 5(Sn + 5.) or.5(sn + Sr) 7 (c'$umed) s,.
i
/. /2 Y l FLes. ALL c W.' l.t.5'(s7 soo ') s 2. 8 3 Wy oeE a = re + I g. S 5 '3 gj(Igoo 1000 PsL n ac;#SO iJ gg.
t g, af3,,,,
,,.7 ; f 14011 r = a/T
.R 01
_ s 14oro p s6 p1
- 2. 48 h 4 = e'/d
.4o1 S QL.T ALLiN. e 6 fog) 1.I s M00 sfg A,,y
,$gg i S'* OP'l * ^L M = M./n rsl8 4Po ra%
1L
- f. T(i n'D i20h2 'L'2.1:7D0p5I bSC M s. M./s h =.c.,2S*
I
" s c '" " e ~ ~ ' ' ' " + ' ~ ' ' ] * "
M. ond M. in obove eqwoi;ons by:
2e + t l
J.o +-
w v
c e,-
- o TAYLOR FORGE & PIPE WORKS a
h, g u
B r.72.
r.o. so. 4e5. Ch:coeo vo. lit I.
9
., f. j8 oi...so s.i.. o n....
c.
4,s.
mi.det ke P
ll.75
- C,hi....
s.". r..' e '.';...
P a b-v A L I
i
.. Ae.gelo.
I*I'*
Ne w.e k b
l y,
I)
- h, -
% - Y $ halts g
Compute d Do
1 A - /.y Checked 4
I*yl.efe*00 & Pap. W.*bs.1941 F8 '*d " L
lEb -/ !2,
- W = WEIGHT
=
L P = N0ZZLE AXIAL LOAD A y V = N0ZZLE SHEAR LOAD k
p M = N0ZZLE BENDING 4
A>2 g
T = N0ZZLE TORSION 10.5" i
U II
~
20.5" The overturning moments are:
NORMAL 600(10.5) + 36,000 - 439 (10.25)' =
37,800 in-lbs.
M
=
x 600(15.5) + 18,000 = 27,300 in-lbs.
M
=
y 18000 + 1200(10,5) + 600(25.75) - 4493(10.25) = 0 in-lbs.
M
=
z (600)2 (1200)2 1342 lbs.
+
R
=
=
0
+ 37.800 1997 psi tensile S
=
=
2(.4617)(20.5) 27300
+
1342
=
1,747 psi shear V
=
4(1.414)(10.25)(.4617)
_4(.4617)
The allcwable stress for SA-307 bolts is 10,000 psi shear and 20,000 psi tensile.
(
[
OBE i
37800 +.12(439) ' + 248627 - 286,480 in-lbs.
M
=
x 27,300 in-lbs.
M y
l 101,773~ in-lbs.
M
=.12(4493) + 101,234
=
l z
l
=!(600+1002)
+ (1200 + 4022) 5,462 lbs.
R
=
l 286480 + 101773
= 20,510 psi tensile l
S
=
l 2(.4617)(20.5) 3,977 psi shear 27300
+
5462 V
=
=
4(1.414)(10.25)(.4617) 4(.4617)
M.
The allowable stresses are 11,000 psi shear and 22,000 psi tensile.
DBE 37800 +.18(439) + 222,365 = 260,244 M
=
x 27,300 in-lbs.
M y
123,748 in-lbs.
.18(4493) + 122,939 M
=
z
=f(600 1109)2 4,636
+ (1200
+ 3110)
R
+
=
20,285 psi tensile 260244 + 123748 S
=
=
2(.4617)(20.5) 3,530 psi shear 27300
+
4636 V
=
=
4(1.414)(10.25)(.4617) 4(.4617)
The allowable stresses are 15,000 psi shear and 30,000 psi tensile.
5.14 Analysis of Discharge Head Hold Down Bolts The moments acting on these bolts are the same as those acting on the anchor bolts.
There are four 1" diameter bolts on a 21.25" bolt circle.
Normal 2,077 psi tensile S = 37800 +
0 2(21.25)(.707)(.6057) 1,614 psi shear 27300
+
1342 V
=
=
4(10.63)(.6057) 4(.6057)
The allowable stresses for A-193-B7 bolts are 12,500 psi shear and 25,000 psi tensile.
OBE 21,333 psi tensile S = 286480 +
101773
=
2(21.25)(.707)(.6057) 3,314 psi shear 27300
+
5462 V
=
=
4(10.63)(.6057) 4(.6057)
The allowable stresses are 13,750 psi shear and 27,500 psi tensile.
DBE 21,099 psi tensile 260244 +
123748 S
=
=
2(21.25)(.707)(.6057) l
/ d[6 - /
2,974 psi shear 27300
+
4636 V
=
=
4(10.63)(.6057) 4(.6057)
The allowable stresses are 18,750 psi shear and 37,500 psi tensile.
5.15 Analysis of Discharge Head N
The maximum stresses in the discharge head occur at the lower end of the cutout.- This occurs, for practical purposes, at joint 3.
The weight of the motor and support above this location is 1475 pounds.
The* member and it's properties are shown below:
AY
.375" x 2.875" plate Typ.
IY-total = 13.84 in X-total = 59.0 in' 2.5 I
T
\\/
yp.
/\\
3 " Typ.
10.75"O.D. Pipe x.307" Wall Normal 15,000 125 psi Compressive Stress 1475 S
=
=
11.83 OBE 23565 (3.88)
+
119554(2.22)
+ 1.12(1475)
= 20,866 psi Bend.
S
=
59.0 13.84 11.83
< 24,750 psi DBE 15441 psi Bend.
23566(3.88)
+
85683(2.22)
+
1.18(1475)
=
S
=
59.0 13.84 11.83
< 36,000 psi j,1 TC
/]'s 5.16 Analysis of Discharge Head Flange The lower flange of the discharge head is subjected to the same moments and forces as the anchor bolts. This flange is analyzed by use of Cases 10 and 20 of Roark's Formulas for Stress and Strain, pages 241 and 242.
The parameters are:
1.62" 10.625" r, = b = 5.375" t =
a =
The stress coeffecients are (from Roark):
.40 1.14 Sp Sg
=
=
Normal
= \\ ((600)(10,5) + 36000)2 + (18000 + 1200(10,5) + 600(15.5))2 = 58,149
/
MT 3,129 psi
+
.40(4932)_
1.14(58149) 2
=
S
=
10.625(1.62)
(1.62)'
The allowable stress is 26,250 psi for the SA 515 Gr 70 plate.
OBE 600(10,5) + 36000 + 248627 = 290,927 M
=
x 141,134 18000 + 1200(10.5) + 600(15.5) + 101234
=
M
=
z (141134)2 323353 (290927)
+
M
=
=
T 14,062 1.14( 3235335,
+.40(1.12)(4932)
=
S
=
10.625(1.62)'
(1.62)4 28,875 psi.
The allowable stress is 1.65(17500)
=
DBE 600(10.5) + 36000 + 222365 = 264,665 M
=
x 162,839 in-lbs.
18000 + 1200(10.5) + 600(15.5) + 122939
=
M
=
z (264665)2
=f(162839)2 310,748
+
M
=
7 13,591 psi 1.14(310748)_
+
.40(1.18)(4932)
=
S
=
10.625(1.62)Z (1.62)Z 42,000 psi.
The allowable stress is 2.4(17500)
=
/2 h
=l)b '
s 5.17 Analysis of Discharge Head Weld The weld at the junction of the discharge head / discharge lead flange must be checked. The dimensions are as shown below:
NMi 3
PIPE 10. 75" 0.D. X. 307" g
g* \\
WALL M 7 i
I i
The weld has a 10.75"I.D. and.309" throat thickness.
Thus:
3.1416(11.368)4 -
(10.75)')
28.90 in.3 Section Modulus 2
=
=
32(11.368) 3.1416(11.368)2 -
(10.75)2) 10.74 in.2 Area A
=
=
4 Normal M
600(7.88) + 36000 = 40,728
=
x M
27,300
=
y H
18000 + 1200(7.88) + 600(20.875) - 1682(5.375) = 30,941 in-lbs.
=
z 3.1416(10.136)2(125)
P 10,086 lbs. due to internal pressure
=
=
4 1342 lbs. shear load R
=
y =
(40728)2
+
(30941) 51,148
=
i 51148_
+
10086 V
2,709 psi
=
=
28.9 10.74 V2 27300
+
1342 598
=
=
5.375 (10.74) 10.74 V,,x
= f( 2709, 698.)2 2,774 psi
+
=
The allowable stress for the weld is 15,000 psi.
~'.
OBE 40728 + 248627 = 289,355 M
=
x 27,300 in-lbs.
~
M
=
y 18000 + 1200(7.88) + 600(20.875) + 101234 - (1
.12)(1682)(5.375)
M
=
z
= 133,259 T = (289335)2 + (133259) = 318,548 10,086 lbs. ;
M 5462 P
R
=
=
V1 318548
+
10086 11,962
=
=
28.9 10.74 27300
+ 5462 995 V
=
=
2 2(3.1416)(5.375)'(.309) 10.74 (995)2 V,,x (11962)
+
12,003
=
=
The allowable stress is 16,500 psi.
DBE 40728 + 222365 263,093 M
=
=
x 27,300 in-lbs.
M
=
y M,
18000 + 1200(7.88) + 600(20.875) + 122939 - (1 18)(1682)(5.375) -._
=
155,507
=
T = (155507)2 + (263093)2 = 305,615 10,086 lbs.
M 4,636 P
R
=
=
11,514 psi 305615
+
10086 V
=
=
y 28.9 10.74 918 psi 2
+
4636 27300 V
=
=
2(3.1416)(5.375)'(.309) 10.74
=[(11514)
(918)2 11551 psi V
+
=
max The allovable stress is 22,500 psi.
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APPEND 1X D - ANALYSIS OF FLAT FACED FIANGES FOR EXTERNAL FORCES The ASME Code does not provide rules for the analysis of Flat Faced flanges which are subjected to external forces. The Code does provide rules for Raised Face flanges subject to external forces due to thermal and dead-weight only - not seismic induced forces - see NC-3647 or ND-3647. for example.
Since no formal Code guidance is given for the analysis of Flat Faced flanges subjected to external loads, the method given by the Code for the analysis of Raised Faced flanges subject to external loads due to thermal and deadweight will be applied here. However, the Code method will be modified to include seismic loads and to be applicable to Flat Faced flanges.
The method used is outlined below.
- 1. The equivalent pressure due to external forces and moments, including seismic, is calculated by the equation given in NC-3647(or ND-3647).
- 2. This equivalent pressure is added to the design pressure (or operating pressure if allowed by the design specification).
- 3. The total pressure obtained in 2 is used for the design pressure for the flange. The flange stresses are calculated per Appendix XI.
- 4. The equivalent pressure obtained in 1 is used to calculate an equivalent i
hydrostatic end force H' by the same method as used for H in XI-3321.2.
- 5. The pump design pressure (or operating pressure if allowed by the design specification) is used to calculate Wm1 = H + He + Hg per XI-3321.2.
- 6. The total bolt load is Wml' = H' + H + He + Hc.
Thus, the only modification to the Code method is that the contact force He' due to the equivalent pressure is not included. This force was omitted
D-2
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because it is believed that the force will be very small when the flange in a is subjected to an external moment. The flange would have to distort very dif ficult to distort manner in order for this force to be developed by an external moment. The co. tact force could be developed by an axial n
force on the connecting pipe. However, the equivalent pressure due to F is always very small relative to that due to M.
In addition, it should be noted that the hydrostatic end force H' contains a great deal of conservatism. This can be seen by deriving the equation for the equivalent pressure given by the Code.
Finally, it should be noted that the use of the equivalent hydrostatic end force H' will yield a bolt stress very close to the stress obtained by use of the equation given for anchor bolt stress of circular stacks in
" Design of Plate Structures", Steel Plate Engineering Data - Volume 2, published by the American Iron and Steel Institute.
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