HYDROSTATIC TEST AND STRESS ANALYSISON SHELL AND TUBE HEAT EXCHANGER
Deepak SharmaB Tech , Punjab Technical University , India 2007
PROJECT
Submitted in partial satisfaction ofthe requirements for the degree of
MASTER OF SCIENCE
in
MECHANICAL ENGINEERING
at
CALIFORNIA STATE UNIVERSITY, SACRAMENTO
SUMMER2011
HYDROSTATIC TEST AND STRESS ANALYSISON SHELL AND TUBE HEAT EXCHANGER
A Project
by
Deepak Sharma
Approved by:
__________________________________, Committee ChairAkihiko Kumagai, Ph.D.
____________________________Date
ii
Student: Deepak Sharma
I certify that this student has met the requirements for format contained in the University
format manual, and that this project is suitable for shelving in the Library and credit is to be
awarded for the Project.
__________________________, Department Chair ________________Susan L. Holl, Ph.D Date
Department of Mechanical Engineering
iii
Abstract
of
THE HYDROSTATIC TEST AND THE STRESS ANALYSISON SHELL AND TUBE HEAT EXCHANGER
by
Deepak Sharma
The hydrostatic test is performed on a computer model of a heat exchanger to see if there
is any leakage in the heat exchanger. This test is performed on different parts separately based on
the calculations specified in the book of ASME SEC VIII DIV-1 2007 ED. Maximum allowable
pressure (MAWP) for different parts is determined and then is multiplied by a factor of 1.3. The
test is then performed on different parts based on those values continuously for 30 minutes. This
test easily tells where the leakage is which usually occurs in the joints which are welded. This test
is performed along with FEA before installing the heat exchanger. Once the CAD model is
developed, FEA is done to analyze the design thus saving money and time by reducing number of
prototypes. Through FEA we can optimize the size so as to reduce the weight of heat exchanger.
Through FEA we take a look at the entire design and understands how load is transferred between
parts and also where critical areas of design are located. Thus we can test a design before actually
building it, therefore saving the time and cost.
_______________________, Committee ChairAkihiko Kumagai, Ph.D.
_______________________Date
iv
ACKNOWLEDGMENTS
It is my distinct honor and proud privilege to acknowledge with gratitude to keen interest taken
by Professor Akihiko Kumagai, his ever-inspiring suggestions; constant supervision and
encouragement that made it possible to pursue and complete this project efficiently. Here I also
thank to the department of Mechanical Engineering and Department Chair Professor Susan L.
Holl who always guide me on proper way.
Finally I thank all the people who extended their support directly or indirectly to make this
project a complete success. In addition, it is a great pleasure to acknowledge the help of many
individuals without whom this project would not have been possible.
v
TABLE OF CONTENTS
Page
Acknowledgement............................................................................................................................v
List of Tables...................................................................................................................................ix
List of Figures..................................................................................................................................xi
1. INTRODUCTION........................................................................................................................1
2. HYDROSTATIC TEST................................................................................................................3
2.1 Test pressure determination for Tube side chamber...............................................................4
2.2 Test pressure determination for Shell side chamber...............................................................5
2.3 Determination of MAWP for Front Channel.........................................................................7
2.4 Determination of MAWP for Shell........................................................................................8
2.5 Determination of MAWP for Front Head............................................................................10
2.6 Determination of MAWP for Straight Flange on Front Head..............................................11
2.7 Determination of MAWP for Straight Flange on Rear Head...............................................12
2.8 Determination of MAWP for Rear Head..............................................................................13
2.9 Tube Side Nozzle (N1)........................................................................................................14
2.10 Shell side inlet Nozzle (N10).............................................................................................17
vi
2.11 Shell side inlet Nozzle (N11).............................................................................................19
2.12 Shell side outlet Nozzle (N12)...........................................................................................20
2.13 Shell side outlet Nozzle (N13)...........................................................................................23
2.14 Shell side outlet Nozzle (N14)...........................................................................................24
2.15 Tube side inlet Nozzle (N2)...............................................................................................26
2.16 Tube side inlet Nozzle (N3)...............................................................................................27
2.17 Tube side outlet Nozzle (N4).............................................................................................29
2.18 Tube side outlet Nozzle (N5).............................................................................................32
2.19 Tube side outlet Nozzle (N6).............................................................................................33
2.20 Shell side inlet Nozzle (N9)...............................................................................................35
2.21 Shell Side Flange................................................................................................................39
2.22 Shell Side Flange (front) - Flange hub...............................................................................43
2.23 Saddle.................................................................................................................................45
2.24 Pass Partition Plate.............................................................................................................52
2.25 Tubes..................................................................................................................................53
3. STRESS ANALYSIS.................................................................................................................56
3.1 Description of FEA on nozzle..............................................................................................56
3.2 Description of FEA on saddle..............................................................................................62
4. RESULTS...................................................................................................................................68
vii
4.1 Pressure summary for tube and shell....................................................................................68
4.2 Thickness on different parts of heat exchanger....................................................................69
4.3 Weight of heat exchanger.....................................................................................................70
4.4 Baffle....................................................................................................................................71
4.5 Study results of nozzle.........................................................................................................72
4.6 Study results of saddle..........................................................................................................73
5. CONCLUSION AND FUTURE SCOPE OF WORK................................................................74
5.1 Conclusion............................................................................................................................74
5.2 Future scope of work............................................................................................................74
References.......................................................................................................................................75
viii
LIST OF TABLES
Page
1. Table 2.1 Pressure on Tube side..........................................................................................4
2. Table 2.2 Pressure on Shell side..........................................................................................5
3. Table 3.1 Study properties of nozzle.................................................................................57
4. Table 3.2 Units...................................................................................................................57
5. Table 3.3 Material properties.............................................................................................58
6. Table 3.4 Structural Properties..........................................................................................58
7. Table 3.5 Load on nozzle...................................................................................................59
8. Table 3.6 Mesh information..............................................................................................59
9. Table 3.7 Study properties.................................................................................................62
10. Table 3.8 Units...................................................................................................................63
11. Table 3.9 Material Properties............................................................................................63
12. Table 3.10 Structural Properties........................................................................................64
13. Table 3.11 Load.................................................................................................................64
14. Table 3.12 Mesh Properties……………………………………………………………...65
15. Table 4.1 Pressure Summary for tube side........................................................................68
16. Table 4.2 Pressure Summary for Shell Side......................................................................69
17. Table 4.3 Thickness Summary..........................................................................................69
18. Table 4.4 Weight Summary of vessel................................................................................70
19. Table 4.5 Weight Summary of attachments......................................................................70
ix
20. Table 4.6 Baffle Summary.................................................................................................71
21. Table 4.7 Summary of FEA on nozzle..............................................................................72
22. Table 4.8 Summary of FEA on Saddle..............................................................................73
x
LIST OF FIGURES
Page1. Figure 2.1 Tube Side Inlet 1..............................................................................................14
2. Figure 2.2 Shell side Inlet N10..........................................................................................17
3. Figure 2.3 Shell side inlet N11..........................................................................................19
4. Figure 2.4 Shell side outlet N12........................................................................................20
5. Figure 2.5 Shell side outlet N13........................................................................................23
6. Figure 2.6 Shell side Outlet N14.......................................................................................24
7. Figure 2.7 Tube side inlet N2............................................................................................26
8. Figure 2.8 Tube side inlet N3............................................................................................27
9. Figure 2.9 Tube side outlet N4..........................................................................................29
10. Figure 2.10 Tube side outlet N5........................................................................................32
11. Figure 2.11 Tube side outlet N6........................................................................................33
12. Figure 2.12 Shell side inlet N9..........................................................................................35
13. Figure 2.13 Detail view of shell and tube heat exchanger.................................................38
14. Figure 2.14 Shell side flange.............................................................................................39
15. Figure 2.15 Closed view of shell and tube heat exchanger...............................................55
16. Figure 3.1 Stress in nozzle during FEA.............................................................................60
17. Figure 3.2 Displacement in nozzle during FEA................................................................61
18. Figure 3.3 Stress on Saddle during FEA...........................................................................66
19. Figure 3.4 Displacement in Saddle during FEA................................................................66
xi
xii
1
Chapter 1
INTRODUCTION
A heat exchanger is a device in which heat is transferred from one fluid to another. The
most commonly type used heat exchanger is shell and tube heat exchanger. One fluid runs inside
the tubes and the other one runs over it and thus heat is transferred from one fluid to another.
These types of heat exchangers are mostly used in oil refineries, refrigeration, power generation
etc. The main purpose in the heat exchanger design is to determine the overall cost of heat
exchanger.
The shell and tube heat exchanger was introduced in 1900s to meet the ever increasing
demand of the industry. During the course of the time it proved to be the best type of heat
exchanger used in the industry. Lots of improvements were done and various progresses has
been made especially in the calculation of true mean temperature difference in the tubes.
The objective of the project here is to design a shell and tube heat exchanger using the
given values which were calculated using specified pressure drops. So in this project I have
continued that work and designed the heat exchanger. Then certain tests are performed that is the
hydrostatic test and stress analysis. These tests are performed before the installation of heat
exchanger thus helping in reducing the time and money. Hydrostatic test is performed to check
the leakage in the system. It involves more calculations and some new factors like nozzle
schedule etc are used which are commonly used in the modern industries. Stress analysis or finite
element analysis is performed to determine the strength of the material involved in the heat
exchanger.
The first chapter deals with the brief introduction of the heat exchanger. The second
chapter deals with the hydrostatic test. In this chapter various calculations have been made taking
2
each part separately. In the third chapter stress analysis on the saddle is done keeping in mind
given load conditions. The fourth chapter deals with stress analysis on the nozzle showing various
displacement that occur on it when the given load is applied on it.
3
Chapter 2
HYDROSTATIC TEST
To start with first we have to decide the NPS (nominal pipe size). NPS is a set of standard
pipe sizes used for low or high pressure and temperature. It was set up by American Standards
Association in 1927.Pipe size is specified according to two variables
NPS for diameter
Schedule for wall thickness
Now for the given schedule, the wall thickness remains the same but the outside diameter
of the pipe increases. And for the given NPS , the outside diameter remains the same but the wall
thickness increases with the schedule. The pipe outside diameter and wall thickness is obtained
according to the NPS and schedule of the pipe. Based on this theory the nozzle summary,
pressure summary, thickness summary is calculated.
A hydrostatic test is a test in which leaks are found in the pipelines of the pressure vessel
such as heat exchanger. Hydro testing of pipes are done to ensure there is no leaks and to expose
defective materials such as corrosive materials which are not visible to the naked eye. Hydrostatic
test is very important to ensure the proper functioning of the heat exchanger under the industrial
conditions.
ASME (American Society of Mechanical Engineers) requires hydrostatic test to ensure
that the heat exchanger is intact and all the parts are tight enough to withstand the high pressure
and temperature. According to the ASME the test should be performed at 1.3 times the
MAWP(maximum allowable working pressure). According to ASME the test should be
performed at higher pressure but mostly tests are performed at 1.3 times the MAWP. The test is
4
performed for 30 minutes to ensure there is no leakage. The test is performed on tubes and shell
sides separately in such a way that leaks can be determined easily.
After the test is over, inspection is done to ensure there is no leakage in any part of the
heat exchanger. This inspection is done at a pressure equal to test pressure divided by1.3.
hydrostatic test is performed every 2 years for high pressure heat exchanger and every 5 years for
low pressure heat exchanger.
Calculations
2.1 Test pressure determination for Tube side chamber
Shop hydrostatic test gauge pressure is 845 psi at 70 °F (the chamber MAWP = 650 psi)
The shop test is performed with the vessel in the horizontal position.
Table 2.1 Pressure on Tube side
Identifier
Local test
pres(psi)
Test liquid
static head
psi
UG-99
stress
ratio
UG-99
pressure
factor
Stress during
test psi
Allowable
test stress
psi
Stress
excessive?
Front Head
(1)
846.358 1.358 1 1.30 18,891 34,200 No
Straight
Flange on
846.358 1.358 1 1.30 21,410 34,200 No
Front
Channel
846.358 1.358 1 1.30 21,410 34,200 No
Tubes 846.272 1.272 1 1.30 3,707 23,400 No
Front
Tubesheet
846.358 1.358 1 1.30 See tubesheet report
TS Inlet
(N1)
846.597 1.597 1 1.30 23,719 48,600 No
TS Inlet
Aux1 (N2)
846.428 1.428 1 1.30 3,764 48,600 No
TS Inlet
Aux2 (N3)
846.428 1.428 1 1.30 3,764 48,600 No
TS Outlet
(N4)
845.217 0.217 1 1.30 23,681 48,600 No
TS Outlet
Aux1 (N5)
845.196 0.196 1 1.30 3,758 48,600 No
TS Outlet
Aux2 (N6)
845.196 0.196 1 1.30 3,758 48,600 No
Identifier
Local test
pressure
psi
Test
liquid
static
head
UG-99
stress
ratio
UG-99
pressure
factor
Stress
during test
psi
Allowab
le test
stress
psi
Stress
excess
ive?
5
Notes:
Front Head limits the UG-99 stress ratio.
PL stresses at nozzle openings have been estimated using the method described
in PVP-Vol. 399, pages 77-82. (3) VIII-2, AD-151.1(b) used as the basis for
nozzle allowable test stress.
The zero degree angular position is assumed to be up, and the test liquid height is
assumed to the top-most flange.
The test temperature of 70 °F is warmer than the minimum recommended
temperature of 10 °F so the brittle fracture provision of UG-99(h) has been met.
2.2 Test pressure determination for Shell side chamber
Hydrostatic test gauge pressure is 845 psi at 70 °F (the chamber MAWP = 650 psi) The
shop test is performed with the vessel in the horizontal position.
Table 2.2 Pressure on Shell side
6
Tubes 846.272 1.272 N/A 1.30 NI NI NI
Front
Tubesheet
846.358 1.358 1 1.30 See tubesheet report
Shell Side
Flange (front)
846.358 1.358 1 1.30 42,6
82
51,300 No
SS Inlet
(N9)
845.217 0.217 1 1.30 23,6
81
48,600 No
SS Inlet
Aux1 (N10)
845.196 0.196 1 1.30 3,75
8
48,600 No
SS Inlet
Aux2 (N11)
845.196 0.196 1 1.30 3,75
8
48,600 No
SS Outlet
(N12)
846.597 1.597 1 1.30 23,7
19
48,600 No
SS Outlet
Aux1 (N13)
846.428 1.428 1 1.30 3,76
4
48,600 No
SS Outlet
Aux2 (N14)
846.428 1.428 1 1.30 3,76
4
48,600 No
Notes:
Shell limits the UG-99 stress ratio.
NI indicates that test stress was not investigated.
PL stresses at nozzle openings have been estimated using the method described
in PVP-Vol. 399, pages 77-82. (4) VIII-2, AD-151.1(b) used as the basis for
nozzle allowable test stress.
The zero degree angular position is assumed to be up, and the test liquid height
is assumed to the top-most flange.
The test temperature of 70 °F is warmer than the minimum recommended
temperature of 10 °F so the brittle fracture provision of UG-99(h) has been
met.
7
2.3 Determination of MAWP for Front Channel
Component: Cylinder
Material specification: SA-516 70 (II-D p. 18, ln. 22)
Material is impact test exempt per UG-20(f)
UCS-66 governing thickness = 0.625 in
Internal design pressure: P = 650 psi @ 400°F
Static liquid head:
Pth = 1.3582 psi (SG=1.0000, Hs=37.6250", Horizontal test head)
Corrosion allowance: Inner C = 0.0000"
Outer C = 0.0000" Design MDMT = -20.00°F
No impact test performed
Rated MDMT = -20.00°F
Material is not normalized
Material is not produced to Fine Grain Practice
PWHT is not performed
Radiography: Longitudinal joint - Full UW-11(a) Type 1
Left circumferential joint - Full UW-11(a) Type 1
Right circumferential joint - Full UW-11(a) ype 1
Estimated weight: New = 421.7529 lb corr = 421.7529 lb
Capacity: New = 78.4174 gal corr = 78.4174 gal
ID = 31.0000"
Length Lc = 24.0000" t = 0.6250"
8
Design thickness, (at 400.00°F) UG-27(c)(1)
t = P*R/(S*E - 0.60*P) + Corrosion
= 650.00*15.5000/(20000*1.00 - 0.60*650.00) + 0.0000
= 0.5138" (2.1)
Maximum allowable working pressure, (at 400.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t) – Ps
=
=
20000*1.00*0.6250 / (15.5000 + 0.60*0.6250) - 0.0000
787.4016 psi (2.2)
Maximum allowable pressure, (at 70.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t)
= 20000*1.00*0.6250 / (15.5000 + 0.60*0.6250)
= 787.4016 psi (2.3)
% Extreme fiber elongation - UCS-79(d)
= (50 * t / Rf) * (1 - Rf / Ro)
= (50 * 0.6250 / 15.8125) * (1 - 15.8125 / ∞)
= 1.9763 % (2.4)
2.4 Determination of MAWP for Shell
Component: Cylinder
Material specification: SA-516 70 (II-D p. 18, ln. 22)
Material is impact test exempt per UG-20(f)
UCS-66 governing thickness = 0.625 in
Internal design pressure: P = 650 psi @ 400°F
9
Static liquid head:
Pth = 1.3582 psi (SG=1.0000, Hs=37.6250", Horizontal test head)
Corrosion allowance:
Inner C = 0.0000"
Outer C = 0.0000"
Design MDMT = -20.00°F No impact test performed
Rated MDMT = -20.00°F Material is not normalized
Material is not produced to Fine Grain Practice
PWHT is not performed
Radiography: Longitudinal joint - Full UW-11(a) Type 1
Left circumferential joint - Full UW-11(a) Type 1
Right circumferential joint - Full UW-11(a) Type 1
Estimated weight: New = 1687.0117 lb corr = 1687.0117 lb
Capacity: New = 223.4131 gal corr = 223.4131 gal
ID = 31.0000"
Length Lc = 96.0000" t = 0.6250"
Design thickness, (at 400.00°F) UG-27(c)(1)
t = P*R/(S*E - 0.60*P) + Corrosion
= 650.00*15.5000/(20000*1.00 - 0.60*650.00) + 0.0000
= 0.5138" (2.5)
10
Maximum allowable working pressure, (at 400.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t) - Ps
=
=
20000*1.00*0.6250 / (15.5000 + 0.60*0.6250) - 0.0000
787.4016 psi (2.6)
Maximum allowable pressure, (at 70.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t)
= 20000*1.00*0.6250 / (15.5000 + 0.60*0.6250)
= 787.4016 psi (2.7)
% Extreme fiber elongation - UCS-79(d)
= (50 * t / Rf) * (1 - Rf / Ro)
= (50 * 0.6250 / 15.8125) * (1 - 15.8125 / ∞)
= 1.9763 % (2.8)
2.5 Determination of MAWP for Front Head
Component: Ellipsoidal Head
Material Specification: SA-516 70 (II-D p.18, ln. 22)
Straight Flange governs MDMT
Internal design pressure: P = 650 psi @ 400 °F
Static liquid head:
Ps= 0 psi (SG=1, Hs=0" Operating head)
Pth= 1.3582 psi (SG=1, Hs=37.625" Horizontal test head)
Corrosion allowance: Inner C = 0" Outer C = 0"
11
Design MDMT = -20°F
No impact test performed
Rated MDMT = -20°F
Material is not normalized
Material is not produced to fine grain practice
PWHT is not performed
Do not Optimize MDMT / Find MAWP
Result Summary
The governing factor is internal pressure
Minimum thickness = 0.0625” + 0” = 0.0625”
Design thickness due to internal pressure = 0.5054”
Maximum allowable working pressure (MAWP) = 803.21 psi
Maximum allowable pressure (MAP) = 803.21 psi
The head internal pressure design thickness is 0.5054".
% Extreme fiber elongation
= (75*t / Rf)*(1 - Rf / Ro)
= (75*0.625 / 5.5825)*(1 - 5.5825 / ∞)
= 8.3968% (2.9)
The extreme fiber elongation exceeds 5 percent. Heat treatment may be required.
2.6 Determination of MAWP for Straight Flange on Front Head
Design thickness, (at 400.00°F)
t = P*R/(S*E - 0.60*P) + Corrosion
= 650.00*15.5000/(20000*1.00 - 0.60*650.00) + 0.0000
12
= 0.5138" (2.10)
Maximum allowable working pressure, (at 400.00°F)
P = S*E*t/(R + 0.60*t) - Ps
= 20000*1.00*0.6250 / (15.5000 + 0.60*0.6250) - 0.0000
Maximum allowable pressure, (at 70.00°F)
P = S*E*t/(R + 0.60*t)
= 20000*1.00*0.6250 / (15.5000 + 0.60*0.6250)
= 787.4016 psi (2.12)
% Extreme fiber elongation
= (50 * t / Rf) * (1 - Rf / Ro)
= (50 * 0.6250 / 15.8125) * (1 - 15.8125 / ∞)
= 1.9763 % (2.13)
2.7 Determination of MAWP for Straight Flange on Rear Head
Design thickness, (at 400.00°F) UG-27(c)(1)
t = P*R/(S*E - 0.60*P) + Corrosion
= 650.00*15.6158/(20000*1.00 - 0.60*650.00) + 0.0000
= 0.5177" (2.14)
Maximum allowable working pressure, (at 400.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t) – Ps
13
=
=
20000*1.00*0.5177 / (15.6158 + 0.60*0.5177) - 0.0000
650.1148 psi (2.15)
Maximum allowable pressure, (at 70.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t)
= 20000*1.00*0.5177 / (15.6158 + 0.60*0.5177)
= 650.1148 psi (2.16)
% Extreme fiber elongation - UCS-79(d)
= (50 * t / Rf) * (1 - Rf / Ro)
= (50 * 0.5177 / 15.8746) * (1 - 15.8746 / ∞)
= 1.6306 % (2.17)
2.8 Determination of MAWP for Rear Head
Design thickness for internal pressure, (Corroded at 400 °F) UG-32(d)(1)
t = P*D/(2*S*E - 0.2*P) + Corrosion
= 650*31.2316/(2*20,000*1 - 0.2*650) + 0
= 0.5092" (2.18)
The head internal pressure design thickness is 0.5092".
Maximum allowable working pressure, (Corroded at 400 °F) UG-32(d)(1)
P = 2*S*E*t/(D + 0.2*t) - Ps
=
=
2*20,000*1*0.5092/(31.2316 +0.2*0.5092) - 0
650.04 psi (2.19)
14
The maximum allowable working pressure (MAWP) is 650.04 psi.
Maximum allowable pressure, (New at 70 °F) UG-32(d)(1)
P = 2*S*E*t/(D + 0.2*t) - Ps
=
=
2*20,000*1*0.5092/(31.2316 +0.2*0.5092) - 0
650.04 psi (2.20)
The maximum allowable pressure (MAP) is 650.04 psi.
% Extreme fiber elongation - UCS-79(d)
= (75*t / Rf)*(1 - Rf / Ro)
= (75*0.5177 / 5.5682)*(1 - 5.5682 / ∞)
= 6.9731%
The extreme fiber elongation exceeds 5 percent. Heat treatment may be required.
Results
The governing condition is internal pressure
Minimum thickness = 0.0625” + 0” = 0.0625”
Design thickness due to internal pressure = 0.5092”
Maximum allowable working pressure(MAWP) = 650.04”
Maximum allowable pressure (MAP) = 650.04 psi
15
2.9 Tube Side Nozzle (N1)
Figure 2.1 Tube Side Inlet 1
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.12371).
Nozzle UCS-66 governing thk: 0.625 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*2.6125/(20,000*1 - 0.6*650)
= 0.0866 in (2.21)
Required thickness tr from UG-37(a)
tr = P*R/(S*E - 0.6*P)
= 650*15.5/(20,000*1 - 0.6*650)
= 0.5138 in (2.22)
Area required per UG-37(c)
Allowable stresses: Sn = 20,000, Sv = 20,000 psi fr1 = lesser of 1 or Sn/Sv = 1
16
fr2 = lesser of 1 or Sn/Sv = 1
A = d*tr*F + 2*tn*tr*F*(1 - fr1)
=
=
5.225*0.5138*1 + 2*0.7*0.5138*1*(1 - 1)
2.6844 in2 (2.23)
Area available from FIG. UG-37.1
A1 = larger of the following= 0.5812 in2
= d*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 5.225*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.5812 in2
= 2*(t + tn)*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 2*(0.625 + 0.7)*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.2948 in2 (2.24)
A2 = smaller of the following= 1.9169 in2
= 5*(tn - trn)*fr2*t
= 5*(0.7 - 0.0866)*1*0.625
= 1.9169 in2
= 5*(tn - trn)*fr2*tn
= 5*(0.7 - 0.0866)*1*0.7
= 2.1469 in2 (2.25)
17
A41 = Leg2*fr2
= 0.8752*1
= 0.7656 in2
Area = A1 + A2 + A41
= 0.5812 + 1.9169 + 0.7656
= 3.2637 in2 (2.26)
As Area >= A the reinforcement is adequate.
Allowable stresses in joints UG-45(c) and UW-15(c)
Groove weld in tension: 0.74*20,000 = 14,800 psi
Nozzle wall in shear: 0.7*20,000 = 14,000 psi
Inner fillet weld in shear: 0.49*20,000 = 9,800 psi
2.10 Shell side inlet Nozzle (N10)
Figure 2.2 Shell side Inlet N10
Calculations for internal pressure 650 psi @ 400 °F
18
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.27)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.28)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
19
The combined weld sizes for t1 and t2 are satisfactory.
UG-45 Nozzle Neck Thickness Check
Wall thickness per UG-45(a): tr1 = 0.0319 in (E =1)
Wall thickness per UG-45(b)(1):tr2 = 0.1253 in
Wall thickness per UG-16(b): tr3 = 0.0815 in
Standard wall pipe per UG-45(b)(4): tr4 = 0.1179 in
The greater of tr2 or tr3: tr5 = 0.1253 in
The lesser of tr4 or tr5: tr6 = 0.1179 in
Required per UG-45 is the larger of tr1 or tr6 = 0.1179 in
Available nozzle wall thickness new, tn = 0.154 in
The nozzle neck thickness is adequate.
20
2.11 Shell side inlet Nozzle (N11)
Figure 2.3 Shell side inlet N11
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Limits of reinforcement per UG-40
Parallel to the vessel wall: (Rn + tn + t )= 1.225 in
Normal to the vessel wall outside: 2.5*(tn - Cn) + te = 0.3375 in (2.29)
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.30)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
21
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.31)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
2.12 Shell side outlet Nozzle (N12)
Figure 2.4 Shell side outlet N12
22
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.12371).
Nozzle UCS-66 governing thk: 0.625 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*2.6125/(20,000*1 - 0.6*650)
= 0.0866 in (2.32)
Required thickness tr from UG-37(a)
tr = P*R/(S*E – 0.6*P)
= 650*15.5/(20,000*1 – 0.6*650)
= 0.5138 in (2.33)
Area required per UG-37(c)
Allowable stresses: Sn = 20,000, Sv = 20,000 psi fr1 = lesser of 1 or Sn/Sv = 1
fr2 = lesser of 1 or Sn/Sv = 1
A = d*tr*F + 2*tn*tr*F*(1 - fr1)
=
=
5.225*0.5138*1 + 2*0.7*0.5138*1*(1 - 1)
2.6844 in2 (2.34)
Area available from FIG. UG-37.1
23
A1 = larger of the following= 0.5812 in2
= d*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 5.225*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.5812 in2
= 2*(t + tn)*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 2*(0.625 + 0.7)*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.2948 in2 (2.35)
A2 = smaller of the following= 1.9169 in2
= 5*(tn - trn)*fr2*t
= 5*(0.7 - 0.0866)*1*0.625
= 1.9169 in2
= 5*(tn - trn)*fr2*tn
= 5*(0.7 - 0.0866)*1*0.7
= 2.1469 in2 (2.36)
A41 = Leg2*fr2
= 0.8752*1
= 0.7656 in2 (2.37)
Area = A1 + A2 + A41
= 0.5812 + 1.9169 + 0.7656
24
= 3.2637 in2 (2.38)
As Area >= A the reinforcement is adequate.
Allowable stresses in joints UG-45(c) and UW-15(c)
Groove weld in tension: 0.74*20,000 = 14,800 psi
Nozzle wall in shear: 0.7*20,000 = 14,000 psi
Inner fillet weld in shear: 0.49*20,000 = 9,800 psi
2.13 Shell side outlet Nozzle (N13)
Figure 2.5 Shell side outlet N13
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.39)
Required thickness tr from UG-37(a)
25
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
2.14 Shell side outlet Nozzle (N14)
Figure 2.6 Shell side Outlet N14
26
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.41)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.42)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
27
The combined weld sizes for t1 and t2 are satisfactory.
28
2.15 Tube side inlet Nozzle (N2)
Figure 2.7 Tube side inlet N2
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.43)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.44)
This opening does not require reinforcement per UG-36(c)(3)(a)
29
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
2.16 Tube side inlet Nozzle (N3)
Figure 2.8 Tube side inlet N3
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
30
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.45)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.46)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
31
2.17 Tube side outlet Nozzle (N4)
Figure 2.9 Tube side outlet N4
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.12371).
Nozzle UCS-66 governing thk: 0.625 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*2.6125/(20,000*1 - 0.6*650)
= 0.0866 in (2.47)
Required thickness tr from UG-37(a)
tr = P*R/(S*E - 0.6*P)
= 650*15.5/(20,000*1 - 0.6*650)
= 0.5138 in (2.48)
32
Area required per UG-37(c)
Allowable stresses: Sn = 20,000, Sv = 20,000 psi
fr1 = lesser of 1 or Sn/Sv = 1
fr2 = lesser of 1 or Sn/Sv = 1
A = d*tr*F + 2*tn*tr*F*(1 - fr1)
=
=
5.225*0.5138*1 + 2*0.7*0.5138*1*(1 - 1)
2.6844 in2 (2.49)
Area available from FIG. UG-37.1
A1 = larger of the following= 0.5812 in2
= d*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 5.225*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.5812 in2
= 2*(t + tn)*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 2*(0.625 + 0.7)*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.2948 in2 (2.50)
A2 = smaller of the following= 1.9169 in2
= 5*(tn - trn)*fr2*t
= 5*(0.7 - 0.0866)*1*0.625
= 1.9169 in2
33
= 5*(tn - trn)*fr2*tn
= 5*(0.7 - 0.0866)*1*0.7
= 2.1469 in2 (2.51)
A41 = Leg2*fr2
= 0.8752*1
= 0.7656 in2 (2.52)
Area = A1 + A2 + A41
= 0.5812 + 1.9169 + 0.7656
= 3.2637 in2 (2.53)
As Area >= A the reinforcement is adequate.
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.625 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.25 in
t1(actual) = 0.7*Leg = 0.7*0.875 = 0.6125 in
The weld size t1 is satisfactory. t2(actual) = 0.5625 in
The weld size t2 is satisfactory. t1 + t2 = 1.175 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
Allowable stresses in joints UG-45(c) and UW-15(c)
Groove weld in tension: 0.74*20,000 = 14,800 psi
Nozzle wall in shear: 0.7*20,000 = 14,000 psi
34
Inner fillet weld in shear: 0.49*20,000 = 9,800 psi
2.18 Tube side outlet Nozzle (N5)
Figure 2.10 Tube side outlet N5
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Limits of reinforcement per UG-40
Parallel to the vessel wall: (Rn + tn + t )= 1.225 in
Normal to the vessel wall outside: 2.5*(tn - Cn) + te = 0.3375 in
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
35
= 0.0129 in (2.54)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.55)
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.135 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
36
2.19 Tube side outlet Nozzle (N6)
Figure 2.11 Tube side outlet N6
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.09576).
Nozzle UCS-66 governing thk: 0.154 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*0.39/(20,000*1 - 0.6*650)
= 0.0129 in (2.56)
Required thickness tr from UG-37(a)
tr = P*Ro/(S*E + 0.4*P)
=
=
650*3.3125/(20,000*1 + 0.4*650)
0.1063 in (2.57)
37
This opening does not require reinforcement per UG-36(c)(3)(a)
UW-16(d) Weld Check
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.0945 in
t1(actual) = 0.7*Leg = 0.7*0.1875 = 0.1312 in
The weld size t1 is satisfactory.
t2(actual) = 0.106 in
The weld size t2 is satisfactory.
t1 + t2 = 0.2372 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
38
2.20 Shell side inlet Nozzle (N9)
Figure 2.12 Shell side inlet N9
Calculations for internal pressure 650 psi @ 400 °F
Nozzle is impact test exempt to -155 °F per UCS-66(b)(3) (coincident ratio = 0.12371).
Nozzle UCS-66 governing thk: 0.625 in
Nozzle rated MDMT: -155 °F
Nozzle required thickness per UG-27(c)(1)
trn = P*Rn/(Sn*E - 0.6*P)
= 650*2.6125/(20,000*1 - 0.6*650)
= 0.0866 in (2.58)
Required thickness tr from UG-37(a)
Tr = P*R/(S*E – 0.6*P)
= 650*15.5/(20,000*1 – 0.6*650)
= 0.5138 in (2.59)
Area required per UG-37(c)
Allowable stresses: Sn = 20,000, Sv = 20,000 psi
39
fr1 = lesser of 1 or Sn/Sv = 1
fr2 = lesser of 1 or Sn/Sv = 1
A = d*tr*F + 2*tn*tr*F*(1 - fr1)
=
=
5.225*0.5138*1 + 2*0.7*0.5138*1*(1 - 1)
2.6844 in2 (2.60)
Area available from FIG. UG-37.1
A1 = larger of the following= 0.5812 in2
= d*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 5.225*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.5812 in2
= 2*(t + tn)*(E1*t - F*tr) - 2*tn*(E1*t - F*tr)*(1 - fr1)
= 2*(0.625 + 0.7)*(1*0.625 - 1*0.5138) - 2*0.7*(1*0.625 - 1*0.5138)*(1 - 1)
= 0.2948 in2 (2.61)
A2 = smaller of the following= 1.9169 in2
= 5*(tn - trn)*fr2*t
= 5*(0.7 - 0.0866)*1*0.625
= 1.9169 in2
= 5*(tn - trn)*fr2*tn
= 5*(0.7 - 0.0866)*1*0.7
40
= 2.1469 in2 (2.62)
A41 = Leg2*fr2
= 0.8752*1
= 0.7656 in2 (2.63)
Area = A1 + A2 + A41
= 0.5812 + 1.9169 + 0.7656
= 3.2637 in2 (2.64)
As Area >= A the reinforcement is adequate.
UW-16(d) Weld Check
tmin = lesser of 0.75 or tn or t = 0.625 in
t1(min) or t2(min) = lesser of 0.25 or 0.7*tmin = 0.25 in
t1(actual) = 0.7*Leg = 0.7*0.875 = 0.6125 in
The weld size t1 is satisfactory. t2(actual) = 0.5625 in
The weld size t2 is satisfactory. t1 + t2 = 1.175 >= 1.25*tmin
The combined weld sizes for t1 and t2 are satisfactory.
The nozzle neck thickness is adequate.
41
Figure 2.13 Detail view of shell and tube heat exchanger
Here is the detail view of shell and tube heat exchanger. This heat exchanger is designed
based on the heat exchanger designed earlier using specified pressure drop. Total number of
nozzles in this heat exchanger is 10. Numbers of Baffles used are 4.this heat exchanger designed
covers less space on floor and is more cost efficient.
42
2.21 Shell Side Flange
Figure 2.14 Shell side flange
Flange calculations for Internal Pressure + Weight Only
Gasket details from facing sketch 1(a) or (b), Column II
Gasket width N = 1.0497 in b0 = N/2 = 0.5249 in
Effective gasket seating width, b = 0.5*b01/2 = 0.3622 in
G = (OD of contact face) - 2b = 32.375 in
hG = (C - G)/2 = (35 - 32.375)/2 = 1.3125 in hD = R + g1/2 = 1 + 1/2 = 1.5 in
43
hT = (R + g1 + hG)/2 = (1 + 1 + 1.3125)/2 = 1.6563 in
Hp = 2*b*3.14*G*m*P
= 2*0.3622*3.14*32.375*1.5603*650
= 74,686.02 lbf
H = 0.785*G2*P
= 0.785*32.3752*650
= 534,813.75 lbf
HD = 0.785*B2*P
= 0.785*312*650
= 490,350.25 lbf
HT = H - HD
= 534,813.8 - 490,350.3
= 44,463.5 lbf (2.65)
Wm1 = H + Hp
= 534,813.8 + 74,686.02
= 609,499.75 lbf (2.66)
Wm2 = 3.14*b*G*y
= 3.14*0.3622*32.375*5,500
= 202,511.91 lbf (2.67)
Required bolt area, Am = greater of Am1, Am2 = 24.37999 in2
44
Am1 = Wm1/Sb = 609,499.8/25,000 = 24.38 in2
Am2 = Wm2/Sa = 202,511.9/25,000 = 8.1005 in2
Total area for 20- 1.5 in dia bolts, corroded, Ab = 28.1 in2
W = (Am + Ab)*Sa/2
= (24.38 + 28.1)*25,000/2
= 655,999.88 lbf (2.68)
MD = HD*hD = 490,350.3*1.5 = 735,525.4 lb-in
MT = HT*hT = 44,463.5*1.6563 = 73,642.7 lb-in
HG = Wm1 - H = 609,499.8 - 534,813.8 = 74,686 lbf
MG = HG*hG = 74,686*1.3125 = 98,025.4 lb-in
Mo = MD + MT + MG = 735,525.4 + 73,642.7 + 98,025.4 = 907,193.4 lb-in
Mg = W*hG = 655,999.9*1.3125 = 860,999.8 lb-in (2.69)
The bolts are adequately spaced so the TEMA RCB-11.23 load concentration factor does not
apply.
Stresses at operating conditions - VIII-1, Appendix 2-7
f = 1
L = (t*e + 1)/T + t3/d
= (3*0.1931 + 1)/1.838313 + 33/67.032
= 1.261815 (2.70)
45
SH = f*Mo/(L*g12*B)
= 1*907,193.4/(1.261815*12*31)
= 23,192 psi (2.71)
SR = (1.33*t*e + 1)*Mo/(L*t2*B)
= (1.33*3*0.1931 + 1)*907,193.4/(1.261815*32*31)
= 4,562 psi (2.72)
ST = Y*Mo/(t2*B) - Z*SR
= 10.6729*907,193.4/(32*31) - 5.5058*4,562
= 9,587 psi (2.73)
Allowable stress Sfo = 20,000 psi
Allowable stress Sno = 20,000 psi
ST does not exceed Sfo
SH does not exceed Min[ 1.5*Sfo, 2.5*Sno ] = 30,000 psi
SR does not exceed Sfo
0.5(SH + SR) = 13,877 psi does not exceed Sfo
0.5(SH + ST) = 16,390 psi does not exceed Sfo
Stresses at gasket seating - VIII-1, Appendix 2-7
SH = f*Mg/(L*g12*B)
= 1*860,999.8/(1.261815*12*31)
46
= 22,011 psi (2.74)
SR = (1.33*t*e + 1)*Mg/(L*t2*B)
= (1.33*3*0.1931 + 1)*860,999.8/(1.261815*32*31)
= 4,330 psi (2.75)
ST = Y*Mg/(t2*B) - Z*SR
= 10.6729*860,999.8/(32*31) - 5.5058*4,330
= 9,099 psi (2.76)
Allowable stress Sfa = 20,000 psi
Allowable stress Sna = 20,000 psi
ST does not exceed Sfa
SH does not exceed Min[ 1.5*Sfa, 2.5*Sna ] = 30,000 psi
SR does not exceed Sfa
0.5(SH + SR) = 13,170 psi does not exceed Sfa
0.5(SH + ST) = 15,555 psi does not exceed Sfa
Flange rigidity per VIII-1, Appendix 2-14
J = 52.14*V*Mo/(L*E*g02*KI*h0)
= 52.14*0.3008*907,193.4/(1.2618*27,900,000*0.6252*0.3*4.4017)
= 0.7836226 (2.77)
The flange rigidity index J does not exceed 1; satisfactory.
47
2.22 Shell Side Flange (front) - Flange hub
Component: Flange hub
Material specification: SA-516 70 (II-D p. 18, ln. 22)
Material impact test exemption temperature from Fig UCS-66 Curve D = -48 °F
Fig UCS-66.1 MDMT reduction = 17.8 °F, (coincident ratio = 0.8220296)
Rated MDMT is governed by UCS-66(b)(2)
UCS-66 governing thickness = 0.625 in
Internal design pressure: P = 650 psi @ 400°F
Static liquid head:
Not Considered
Corrosion allowance: Inner C = 0.0000" Outer C = 0.0000"
Design MDMT = -20.00°F No impact test performed
Rated MDMT = -55.00°F Material is normalized
Material is produced to Fine Grain Practice
PWHT is not performed
Radiography: Longitudinal joint - Seamless No RT
Left circumferential joint - N/A
Right circumferential joint - Full UW-11(a) Type 1
Estimated weight: New = 34.4515 lb corr = 34.4515 lb
Capacity: New = 6.5348 gal corr = 6.5348 gal
ID = 31.0000"
Length Lc = 2.0000"
t = 0.6250"
Design thickness, (at 400.00°F) UG-27(c)(1)
48
t = P*R/(S*E - 0.60*P) + Corrosion
= 650.00*15.5000/(20000*1.00 - 0.60*650.00) + 0.0000
= 0.5138" (2.78)
Maximum allowable working pressure, (at 400.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t) - Ps
=
=
20000*1.00*0.6250 / (15.5000 + 0.60*0.6250) - 0.0000
787.4016 psi (2.79)
Maximum allowable pressure, (at 70.00°F) UG-27(c)(1)
P = S*E*t/(R + 0.60*t)
= 20000*1.00*0.6250 / (15.5000 + 0.60*0.6250)
= 787.4016 psi (2.80)
% Extreme fiber elongation - UCS-79(d)
= (50 * t / Rf) * (1 - Rf / Ro)
= (50 * 0.6250 / 15.8125) * (1 - 15.8125 / ∞)
= 1.9763 % (2.81)
2.23 Saddle
Saddle material:A36
Saddle construction is: Web at edge of rib
Saddle allowable stress: Ss = 24,000 psi
Saddle yield stress: Sy = 36,000 psi
49
Saddle distance to datum: 17.0625 in
Tangent to tangent length: L = 133.2 in
Saddle separation: Ls = 70.6875 in
Vessel radius: R = 16.125 in
Tangent distance left: Al = 43.45 in
Tangent distance right: Ar = 19.0625 in
Tubesheet distance left: Atsl = 15.6 in
Saddle height: Hs = 28.125 in
Saddle contact angle: θ = 120°
Wear plate thickness: tp = 0.25 in
Wear plate width: Wp = 7 in
Wear plate contact angle: θw = 132°
Web plate thickness: ts = 0.25 in
Base plate length: E = 32 in
Base plate width: F = 6 in
Base plate thickness: tb = 0.375 in
Number of stiffener ribs: n = 3
Largest stiffener rib spacing: di = 15.375 in
Stiffener rib thickness: tw = 0.25 in
Saddle width: B = 5 in
Anchor bolt size & type: 1 inch series 8 threaded
50
Anchor bolt material: SA-193 B8
Anchor bolt allowable shear: 18,800 psi
Anchor bolt corrosion allowance: 0 in
Anchor bolts per saddle:4
Base coefficient of friction: µ = 0.45
Weight on left saddle: operating corr = 5,104 lb, test new = 7,568 lb
Weight on right saddle: operating corr = 1,646 lb, test new = 2,731 lb
Weight of saddle pair = 180 lb
51
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.
(2) If CL of tube sheet is located within a distance of Ro/2 to CL of saddle, the shell is assumed
stiffened as if tube sheet is a bulk head.
Longitudinal stress between saddles (Weight ,Operating, right saddle loading and geometry
govern)
S1 = +- 3*K1*Q*(L/12) / (π*R2*t)
= 3*0.3706*1,646*(133.2/12) / (π*15.81252*0.625)
= 41 psi
Sp = P*R/(2*t)
= 650*15.5/(2*0.625)
= 8,060 psi (2.82)
Maximum tensile stress S1t = S1 + Sp = 8,101 psi
Maximum compressive stress (shut down) S1c = S1 = 41 psi
Tensile stress is acceptable (<=1*S*E = 20,000 psi)
Compressive stress is acceptable (<=1*Sc = 15,071 psi)
Longitudinal stress at the left saddle (Weight ,Operating)
Le = 2*(Left head depth)/3 + L + 2*(Right head depth)/3
= 2*8.375/3 + 133.2 + 2*8.3171/3
= 144.3281 in (2.83)
52
w = Wt/Le = 6,750/144.3281 = 46.77 lb/in
Bending moment at the left saddle:
Mq = w*(2*H*Al/3 + Al2/2 - (R2 - H2)/4)
= 46.77*(2*8.375*43.45/3 + 43.452/2 - (16.1252 - 8.3752)/4)
= 53,272.9 lb-in (2.84)
S2 = +- Mq*K1'/ (π*R2*t)
= 53,272.9*9.3799/ (π*15.81252*0.625)
= 1,018 psi (2.85)
Sp = P*R/(2*t)
= 650*15.5/(2*0.625)
= 8,060 psi (2.86)
Maximum tensile stress S2t = S2 + Sp = 9,078 psi
Maximum compressive stress (shut down) S2c = S2 = 1,018 psi
Tensile stress is acceptable (<=1*S = 20,000 psi)
Compressive stress is acceptable (<=1*Sc = 15,071 psi)
Tangential shear stress in the shell (left saddle, Weight ,Operating)
Qshear = Q - w*(a + 2*H/3)
= 5,104 - 46.77*(43.45 + 2*8.375/3)
= 2,810.79 lbf (2.87)
53
S3 = K2.2*Qshear/(R*t)
= 1.1707*2,810.79/(15.8125*0.625)
= 333 psi (2.88)
Tangential shear stress is acceptable (<= 0.8*S = 16,000 psi)
Circumferential stress at the left saddle horns (Weight ,Operating)
S4 = -Q/(4*t*(b+1.56*Sqr(Ro*t))) - 3*K3*Q/(2*t2)
= -5,104/(4*0.625*(5+1.56*Sqr(16.125*0.625))) - 3*0.0503*5,104/(2*0.6252)
= -1,191 psi (2.89)
Circumferential stress at saddle horns is acceptable (<=1.5*Sa = 30,000 psi)
The wear plate was not considered in the calculation of S4 because the wear plate width is not at
least {B +1.56*(Rotc)0.5} =9.9524 in
Ring compression in shell over left saddle (Weight ,Operating)
S5 = K5*Q/((t + tp)*(ts + 1.56*Sqr(Ro*tc)))
= 0.7603*5,104/((0.625 + 0.25)*(0.25 + 1.56*Sqr(16.125*0.875)))
= 726 psi (2.90)
Ring compression in shell is acceptable (<= 0.5*Sy = 16,250 psi)
Saddle splitting load (left, Weight ,Operating)
Area resisting splitting force = Web area + wear plate area
Ae = Heff*ts + tp*Wp
= 5.375*0.25 + 0.25*7
= 3.0938 in2 (2.91)
54
S6 = K8*Q / Ae
= 0.2035*5,104 / 3.0938
= 336 psi (2.92)
Stress in saddle is acceptable (<= (2/3)*Ss = 16,000 psi)
Shear stress in anchor bolting, one end slotted
Maximum seismic or wind base shear = 0 lbf
Thermal expansion base shear = W*µ = 5,194 * 0.45= 2,337.3 lbf
Corroded root area for a 1 inch series 8 threaded bolt = 0.551 in2 ( 4 per saddle )
Bolt shear stress = 2,337.3/(0.551*4) = 1,060 psi
Anchor bolt stress is acceptable (<= 18,800 psi)
Web plate buckling check (Escoe pg 251)
Allowable compressive stress Sc is the lesser of 24,000 or 8,870 psi: (8,870)
Sc = Ki*π2*E/(12*(1 - 0.32)*(di/tw)2)
= 1.28*π2*29E+06/(12*(1 - 0.32)*(15.375/0.25)2)
= 8,870 psi (2.93)
Allowable compressive load on the saddle
be = di*ts/(di*ts + 2*tw*(b - 1))
= 15.375*0.25/(15.375*0.25 + 2*0.25*(5 - 1))
= 0.6578 (2.94)
Fb = n*(As + 2*be*tw)*Sc
55
= 3*(1.1875 + 2*0.6578*0.25)*8,870
= 40,351.84 lbf (2.95)
Saddle loading of 7,658 lbf is <= Fb; satisfactory.
Primary bending + axial stress in the saddle due to end loads (assumes one saddle slotted)
σb = V * (Hs - xo)* y / I + Q / A
= 0 * (28.125 - 13.3353)* 3.2917 / 14.87 + 5,104 / 10.5448
= 484 psi (2.96)
The primary bending + axial stress in the saddle <= 24,000 psi; satisfactory.
Secondary bending + axial stress in the saddle due to end loads (includes thermal
expansion, assumes one saddle slotted)
σb = V * (Hs - xo)* y / I + Q / A
= 2,337.3 * (28.125 - 13.3353)* 3.2917 / 14.87 + 5,104 / 10.5448
= 8,135 psi (2.97)
The secondary bending + axial stress in the saddle < 2*Sy= 72,000 psi; satisfactory.
Saddle base plate thickness check (Roark sixth edition, Table 26, case 7a)
where a = 15.375, b = 5.75 in
tb = (β1*q*b2/(1.5*Sa))0.5
= (2.2393*40*5.752/(1.5*24,000))0.5
= 0.2864 in (2.98)
The base plate thickness of 0.375 in is adequate.
Foundation bearing check
56
Sf = Qmax / (F*E)
= 7,658 / (6*32)
= 40 psi (2.99)
Concrete bearing stress < 750 psi ; satisfactory.
57
2.24 Pass Partition Plate
Minimum Front Pass Partition Plate Thickness
Front tube side pressure drop: q = 1 psi
Front pass plate material: SA-516 70 (II-D p. 18, ln. 22)
Front pass plate allowable stress: S = 20,000 psi
Front pass plate dimension: a = 30.9375 in
Front pass plate dimension: b = 33.75 in
Front pass plate thickness, new: T = 0.5625 in
Front pass plate corrosion allowance: C = 0.0625 in
Front pass plate fillet weld leg size, new 0.25 in
From TABLE RCB-9.131 t = 0.5
From TABLE RCB-9.132, three sides fixed, a/b = 0.9167, B = 0.2624
t = b*(q*B/(1.5*S))1/2 + C
= 33.75*(1*0.2624/(1.5*20,000.00))1/2 + 0.0625
= 0.1623 in (2.100)
The pass partition plate thickness of 0.5625 in is adequate.
Pass partition minimum weld size = 0.75*t + C/0.7= 0.1641 in
The pass partition fillet weld size of 0.25 in is adequate.
58
2.25 Tubes
Component: Tubes
Material specification: SA-179 Smls tube (II-D p. 6, ln. 11)
Material is impact test exempt per UCS-66(d) (NPS 4 or smaller pipe)
Internal design pressure: P = 650 psi @ 400°F
External design pressure: Pe = 650 psi @ 400°F
Static liquid head:
Pth = 1.2724 psi (SG=1.0000, Hs=35.2500", Horizontal test head)
Corrosion allowance: Inner C = 0.0000" Outer C = 0.0000"
Design MDMT = -20.00°F No impact test performed
Rated MDMT = -155.00°F Material is not normalized
Material is not produced to Fine Grain Practice
PWHT is not performed
Estimated weight: New = 5.2439 lb corr = 5.2439 lb
Capacity: New = 0.1193 gal corr = 0.1193 gal OD = 0.7500"
Length Lc = 80.0000" t = 0.0850"
Design thickness, (at 400.00°F) Appendix 1-1
t = P*Ro/(S*E + 0.40*P) + Corrosion
=
=
650.00*0.3750/(13400*1.00 + 0.40*650.00) + 0.0000
0.0179" (2.101)
59
Maximum allowable working pressure, (at 400.00°F) Appendix 1-1
P = S*E*t/(Ro - 0.40*t) - Ps
=
=
13400*1.00*0.0744 / (0.3750 - 0.40*0.0744) - 0.0000
2886.6765 psi (2.102)
Maximum allowable pressure, (at 70.00°F) Appendix 1-1
P = S*E*t/(Ro - 0.40*t)
=
=
13400*1.00*0.0744 / (0.3750 - 0.40*0.0744)
2886.6765 psi (2.103)
External Pressure, (Corroded & at 400.00°F) UG-28(c)
L/Do = 80.0000/0.7500 = 50.0000
Do/t = 0.7500/0.033147 = 22.6263
From table G: A = 0.002230
From table CS-1: B = 11030.3105 psi
Pa = 4*B/(3*(Do/t))
= 4*11030.3105/(3*(0.7500/0.033147))
= 650.0004 psi
Design thickness for external pressure Pa = 650.0004 psi
= t + Corrosion = 0.033147 + 0.0000 = 0.0331"
Maximum Allowable External Pressure, (Corroded & at 400.00°F) UG-28(c)
L/Do = 80.0000/0.7500 = 50.0000
Do/t = 0.7500/0.0744 = 10.0840 (2.104)
From table G: A = 0.010996
60
From table CS-1: B = 12939.5146
psi
Pa = 4*B/(3*(Do/t))
= 4*12939.5146/(3*(0.7500/0.0744))
= 1710.8915 psi (2.105)
Figure 2.15 Closed view of shell and tube heat exchanger
61
Chapter 3
STRESS ANALYSIS
Stress analysis is a part of engineering that decides the stress in the given part when it is
subjected to a particular type of load. That means it tells us whether that particular part can
withstand that particular or different types of load or not. Stress analysis is required for different
types of materials like those involved in heat exchangers, tunnels etc. it is a very important factor
considering the design of the various mechanical parts.
The main purpose of the stress analysis is to determine the strength of the material or a
collection of materials that in turn are used in different bodies like heat exchangers automobiles
etc . The material’s maximum tensile strength, fatigue and other factors are noted down and then
the given force is applied to check whether the tensile strength, fatigue are less than that of the
material.
There is one more important factor considered and that one is factor of safety. It defines
the capacity of the system beyond the actual given load that is how much the system is stronger
than it usually is for a given particular load. In that case factor of safety comes into place.
Numerically it is defined as
Factor of safety = Material strength Design Load
Factor of safety is different for different mechanical components. For example in case of
heat exchangers it is taken as 3.5 – 4. Now here we are doing the FEA analysis on different parts
of the heat exchanger separately. First we are performing on the test on nozzle.
62
3.1 Description of FEA on nozzle
Nozzle Finite element analysis (FEA) is done to calculate the stresses and flexibility of
the meshing of the nozzle and shell. This also calculates the given number of loads on the nozzle
and estimates its functioning over wide range of operating conditions. This also is done to
increase the safety of the equipment. Below are the steps which we do while doing the stress
analysis:-
Step 1
Table 3.3 Study properties of nozzle
Study name Study 1
Analysis type Static
Mesh Type: Solid Mesh
Solver type FFE Plus
Inplane Effect: Off
Soft Spring: Off
Inertial Relief: Off
Thermal Effect: Input Temperature
Zero strain temperature 298.000000
Units Kelvin
This step includes the general properties of the material chosen. These are the steps that are
performed while doing the finite element analysis of any part. Each value has to be filled out
according to its characteristics.
Step 2
63
Table 3.4 Units
Unit system: SI
Length/Displacement mm
Temperature Kelvin
Angular velocity rad/s
Stress/Pressure N/m^2
This table tells us the system of units used. The system used here is SI system and the units are
chosen accordingly.
Step 3
Table 3.5 Material properties
No. Body Name Material Mass Volume
1 Solid Body 1(CirPattern1)
SA516 Steel 26.937 kg 0.00343146 m^3
2 SolidBody 1(Cut-Extrude2)
SA516 Steel 193.166 kg 0.0246071 m^3
The material chosen here is SA 516 steel. This is chosen because it is the most cost
effective and it withstands the given loads without any failure.
Step 4Table 3.6 Structural Properties
Property Name Value Units Value Type
Elastic modulus 2e+011 N/m^2 Constant
Poisson's ratio 0.26 NA Constant
Shear modulus 7.93e+010 N/m^2 Constant
Mass density 7850 kg/m^3 Constant
64
Tensile strength 4e+008 N/m^2 Constant
Yield strength 2.5e+008 N/m^2 Constant
This table tells here the general properties of the SA 516 steel like what is its yield
strength, Poisson’s ratio etc.
Step 5
Table 3.7 Load on nozzle
Load name Selection set Loading type Description
Force-1 <Shell1-1, Nozzle-1>
on 2 Face(s) apply normal force 4.48 N using uniform distribution
Sequential Loading Load acting tangentially
The loading on the nozzle is sequential loading that is it is acting uniformly from all sides
as can be seen in the figure 3.16.
Step 6Table 3.8 Mesh information
Mesh Type: Solid Mesh
Mesher Used: Standard mesh
Automatic Transition: Off
Smooth Surface: On
Jacobian Check: 4 Points
Element Size: 37.615 mm
Tolerance: 1.8807 mm
Quality: High
65
Number of elements: 9720
Number of nodes: 19469
Time to complete mesh(hh;mm;ss): 00:00:06
Computer name: MT6
This table clearly tells the type of mesh and number of nodes in nozzle while performing
the test. The time effectiveness can also be seen from the fact that after all the values were put in,
it took only 6 sec to complete the analysis. Thus this test clearly reduces the time from design to
production.
Figure 3.16 Stress in nozzle during FEA
As we can see in the figure there is no red region in any part of the nozzle. So nozzle can
withstand the given load .
66
Figure 3.17 Displacement in nozzle during FEA
There is small red portion in the bottom as can be seen in the figure. But when we see the
corresponding values on the right hand side, we find that they are vey less. It is approximately
equal to 0.0000139mm which is very less and can be neglected.
67
3.2 Description of FEA on saddle
Stress analysis on the saddle is done to predict the stress distributions in the heat
exchanger. FEA is used where number of saddles is two or more. The saddle should be designed
to meet two loading conditions:-
Primary load failures
Secondary/fatigue failures
The first one includes the excessive distortion due to pressure more than the given
pressure. This is an important factor and should be kept in mind while designing the saddle.
The second one includes the crack that can occur in the saddle due to excessive pressure
or due to ageing of the material.
Below are the steps which we do while doing the stress analysis:-
Step 1
Table 3.9 Study properties
Study name Study 1
Analysis type Static
Mesh Type: Solid Mesh
Solver type FFE Plus
Inplane Effect: Off
Soft Spring: Off
Inertial Relief: Off
68
Thermal Effect: Input Temperature
Zero strain temperature 298.000000
Units Kelvin
This step includes the general properties of the material chosen. These are the steps that are
performed while doing the finite element analysis of any part. Each value has to be filled out
according to its characteristics.
Step 2
Table 3.10 Units
Unit system: SI
Length/Displacement mm
Temperature Kelvin
Angular velocity rad/s
Stress/Pressure N/m^2
This table tells us the system of units used. The system used here is SI system and the units are
chosen accordingly.
Step 3
Table 3.11 Material Properties
No. Body Name Material Mass Volume
1 SolidBody 1(Boss-Extrude4)
SA 516Steel 39.2619 kg 0.00500151 m^3
69
The material chosen here is SA 516 steel. This is chosen because it is the most cost
effective and it withstands the given loads without any failure.
Step 4
Table 3.12 Structural Properties
Property Name Value Units Value Type
Elastic modulus 2e+011 N/m^2 Constant
Poisson's ratio 0.26 NA Constant
Shear modulus 7.93e+010 N/m^2 Constant
Mass density 7850 kg/m^3 Constant
Tensile strength 4e+008 N/m^2 Constant
Yield strength 2.5e+008 N/m^2 Constant
This table tells here the general properties of the SA 516 steel like what is its yield
strength, Poisson’s ratio etc.
Step 5
Table 3.13 Load
Load name Selection set Loading type Description
Force-1 <Bracket> on 1 Face(s) apply normal force 23314 N using uniform distribution
Sequential Loading
70
The loading on the saddle is sequential loading that is it is acting uniformly from all sides
as can be seen in the figure 3.16.
Step 6
Table 3.14 Mesh Properties
Mesh Type: Solid Mesh
Mesher Used: Standard mesh
Automatic Transition: Off
Smooth Surface: On
Jacobian Check: 4 Points
Element Size: 0.94752 in
Tolerance: 0.047376 in
Quality: High
Number of elements: 7875
Number of nodes: 16077
Time to complete mesh(hh;mm;ss): 00:00:04
Computer name: MT6
This table clearly tells the type of mesh and number of nodes in saddle while performing
the test. The time effectiveness can also be seen from the fact that after all the values were put in,
it took only 4 sec to complete the analysis. Thus this test clearly reduces the time from design to
production.
71
Figure 3.18 Stress on Saddle during FEA
The stresses during performing the FEA can be seen in the picture. It is clearly
seen that is very few red zone that is very few danger zones and that can be neglected.
Figure 3.19 Displacement in Saddle during FEA
72
The displacement analysis is done and it is clearly seen in the picture that there
are very few red zones in the saddle. When we check the values corresponding to the red
zones, they are found to be very less and can be neglected.
73
Chapter 4
RESULTS
All the results that came above after doing the hydrostatic test proves that this heat
exchanger will work very well in the given conditions. Hydrostatic test is a must for long
functioning of the heat exchanger. This test is performed on different parts separately and based
on that test full result summary is given below in the tables:-
4.1 Pressure summary for tube and shell
Table 4.1 Pressure Summary for tube side
IdentifierP
Design
( psi)
T Design(°F)
MAWP ( psi)
MAP ( psi)
MAEP ( psi)
Teexternal(°F)
MDMT (°F)
MDMT Exemption
Total Corrosio
n
Allowance
ImpactTest
Front Head 650.0 400.0 803.2 803.2 0.00 400.0 -20 Note 1 0.000 No
Straight Flange 650.0 400.0 787.4 787.4 N/A 400.0 -20 Note 2 0.000 No
Front Channel 650.0 400.0 787.4 787.4 N/A 400.0 -20 Note 2 0.000 No
Front Tubesheet 650.0 400.0 796.0 928.3 680.7 400.0 -20 Note 3 0.125 No
Tubes 650.0 400.0 2886. 2886. 1710. 400.0 -155 Note 4 0.000 No
TS Inlet (N1) 650.0 400.0 650.0 650.0 N/A 400.0 -54 Note 5 0.019 No
TS Inlet Aux1 650.0 400.0 650.0 650.0 N/A 400.0 -155 Note 6 0.019 No
TS Inlet Aux2 650.0 400.0 650.0 650.0 N/A 400.0 -155 Note 6 0.019 No
TS Outlet (N4) 650.0 400.0 650.0 650.0 N/A 400.0 -54 Note 5 0.019 No
TS Outlet Aux1 650.0 400.0 650.0 650.0 N/A 400.0 -155 Note 6 0.019 No
TS Outlet Aux2 (N6)
650.0 400.0 650.00
650.00
N/A 400.0 -155 Note 6 0.019 No
The MAWP for different parts of the heat exchanger are calculated and are as shown in
the table. It is now multiplies by 1.3 to perform the hydrostatic test.
74
Table 4.2 Pressure Summary for Shell Side
IdentifierP
Design( psi)
T Desig
n(°F)
MAWP ( psi)
MAP ( psi)
MAEP ( psi)
Teextern
al(°F
MDMT (°F)
MDMT Exemption
Total Corrosi
on
All
ImpactTest
Front 650.0 400.0 680.79 991.73 796.01 400.0 -20.0 Note 3 0.125 No
Shell 650.0 400.0 787.40 787.40 N/A 400.0 -20.0 Note 2 0.000 No
Straight 650.0 400.0 650.11 650.11 N/A 400.0 -20.0 Note 8 0.000 No
Rear Head 650.0 400.0 650.04 650.04 0.00 400.0 -20.0 Note 7 0.000 No
Tubes 650.0 400.0 1710.8 1710.8 2886.6 400.0 N/A N/A 0.000 No
Shell Side 650.0 400.0 749.18 749.18 0.00 400.0 -55.0 Note 9 0.000 No
Shell Side 650.0 400.0 787.40 787.40 N/A 400.0 -55.0 Note 10 0.000 No
Saddle 650.0 400.0 650.00 N/A N/A N/A N/A N/A N/A N/A
SS Inlet 650.0 400.0 650.00 650.00 N/A 400.0 -155 Note 6 0.019 No
SS Inlet 650.0 400.0 650.00 650.00 N/A 400.0 -155 Note 6 0.019 No
SS Outlet 650.0 400.0 650.00 650.00 N/A 400.0 -54.0 Note 5 0.019 No
SS Outlet 650.0 400.0 650.00 650.00 N/A 400.0 -155 Note 6 0.019 No
SS Outlet 650.0 400.0 650.00 650.00 N/A 400.0 -155 Note 6 0.019 No
SS Inlet 650.0 400.0 650.00 650.00 N/A 400.0 -54.0 Note 5 0.019 No
The MAWP for different parts of the heat exchanger are calculated on the shell side and
are as shown in the table. It is now multiplies by 1.3 to perform the hydrostatic test. After
performing the test it was found that there is no leakage in the system.
4.2 Thickness on different parts of heat exchanger
Table 4.3 Thickness Summary
ComponentIdentifier
Material Diameter(in)
Length(in)
Nominal t(in)
Design t(in)
JointE
Load
Front Head SA-516 70 31.00 ID 8.38 0.6250 0.5054 1.000 Internal
Straight Flange on Front Head SA-516 70 31.00 ID 2.00 0.6250 0.5138 1.000 Internal
Front Channel SA-516 70 31.00 ID 24.00 0.6250 0.5138 1.000 Internal
Front Tubesheet SA-516 70 37.25 OD
3.70 3.7000 3.6176 1.000 Unknown
Tubes SA-179 Smls tube
0.7500 OD
80.00 0.0850 0.0331 1.000 External
75
Shell SA-516 70 31.00 ID 96.00 0.6250 0.5138 1.000 Internal
Straight Flange on Rear Head SA-516 70 31.23 ID 2.00 0.5177 0.5177 1.000 Internal
Rear Head SA-516 70 31.23 ID 8.32 0.5092 0.5092 1.00 Internal
Here is the summary of thickness of different parts of the heat exchanger. The material used here
is SA 51670. Each dimension like length, breadth etc are calculated on every part.
4.3 Weight of heat exchanger
Table 4.4 Weight Summary of vessel
ComponentWeight ( lb) Contributed by Vessel Elements
MetalNew*
MetalCorroded*
Insulation & Supports
Lining Piping+ Liquid
OperatingLiquid
TestLiquid
Front Head 240.97 240.97 0.00 0.00 0.00 0.00 195.26Front Channe l 409.56 409.56 0.00 0.00 0.00 0.00 656.52Front Tubesheet 911.67 880.87 0.00 0.00 0.00 0.00 0.00Shell 1,674.82 1,674.82 0.00 0.00 0.00 0.00 1,971.90Tubes 2,601.00 2,601.00 0.00 0.00 0.00 0.00 493.61Rear Head 197.83 197.83 0.00 0.00 0.00 0.00 199.25Saddle 180.00 180.00 0.00 0.00 0.00 0.00 0.00TOTAL: 6,215.84 6,185.04 0.00 0.00 0.00 0.00 3,516.53
This table tells us the total weight of the heat exchanger. When we add all the values it
comes out to be 6962 lb .
Table 4.5 Weight Summary of attachments
ComponentWeight ( lb) Contributed by Attachments
Body Flanges Nozzles & Flanges
PackedBeds
Trays & Supports
Rings &
VerticalLoads
76
ClipsNew Corroded New Corroded
Front Head 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00Front Channe l 0.00 0.00 134.94 134.37 0.00 0.00 0.00 0.00Front 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00Shell 339.24 339.24 134.94 134.37 0.00 137.40¹ 0.00 0.00Rear Head 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00TOTAL: 339.24 339.24 269.88 268.75 0.00 137.40 0.00 0.00
Vessel operating weight, Corroded: 6,930 lb
Vessel operating weight, New: 6,962 lb
Vessel empty weight, Corroded: 6,930 lb
Vessel empty weight, New: 6,962 lb
Vessel test weight, New: 10,479 lb
Vessel center of gravity location - from datum - lift condition
Vessel Lift Weight, New: 6,961 lb
Center of Gravity: 67.77"
4.4 Baffle
Table 4.6 Baffle Summary
Baffle
Name
Distance from
Front Tubesheet (in) Cut Direction
Cut Distance from
Center (in) Baffle Weight
(lb)
Baffle
#1
17.500 Downwards 6.200 34.349
Baffle
#2
35.000 Upwards 6.200 34.349
77
Baffle
#4
52.500 Downwards 6.200 34.349
Baffle
#3
70.000 Upwards 6.200 34.349
For better efficiency the baffles should be cut accordingly as shown in the table. The
distance of the baffles from the tubesheet is specified accordingly with their direction of cut.
4.5 Study results of nozzle
Table 4.7 Summary of FEA on nozzle
Name Type Min Location Max Location
Stress1 VON: von Mises Stress
2.12762 N/m2
Node: 8470
(-827.723 mm,
342.867 mm,
208.502 mm)
847.1 N/m2
Node: 19227
(-566.039 mm,-0.00328805 mm,-401.287
Displacement1
URES: Resultant Displacement
6.84308e-012 mm
Node: 5042
(-566.039 mm,
-6.85283e-012 mm,
401.287 mm)
1.34909e-005 mm
Node: 15863
(-326.784 mm,
-3.582e-005 mm,
-409.224 mm)
As the diagram clearly indicates that the min stress is 2.12762 N/m2 and the maximum
stress is 847.1 N/m2 and the colour in the diagram is blue which clearly shows that nozzle will
work perfectly in the given working conditions when max stress is applied.
78
4.6 Study results of saddle
Table 4.8 Summary of FEA on Saddle
Name Type Min Location Max Location
Stress1 VON: von Mises Stress
165.412 N/m2
Node: 7112
(-5.64706 in,
0 in,
-2.5 in)
3.15364e+007 N/m2
Node: 15894
(0.473022 in,
11.7122 in,
-3.5639 in)
Displacement1 URES: Resultant Displacement
0 mm
Node: 1362
(-16 in,
0 in,
-3 in)
0.255345 mm
Node: 280
(8.70051 in,
13.8879 in,
-3.56437 in)
The maximum stress in case of saddle is 3.15364 N/m2 and the minimum stress is
165.412 N/m2 and it works perfectly fine under the given load conditions.
79
Chapter 5
CONCLUSION AND FUTURE SCOPE OF WORK
5.1 Conclusion
Both Hydrostatic test and Stress analysis proves that the new shell and tube heat
exchanger which is designed using specified pressure drops can be used in the industry. The
calculations done in the hydrostatic test is based on the ASME SEC VIII DIV-1 2007 ED. The
overall cost of this heat exchanger would be approximately $ 19000. There was no specific
leakage found during the hydrostatic test performed separately on different parts. The hydrostatic
test performed should never exceed 1.3 times the MAWP. During the FEA there is a slight
displacement both in nozzle and saddle which looks alarming but actually the red zone
corresponds to a very low value which can be neglected.
5.2 Future scope of work
The hydrostatic test performed here is only limited to coolant being hot water and fluid to
be cooled is steam. New set of calculations has to be performed when using different fluids.
Other tests have to be performed before the installation of heat exchanger where leakage is not an
issue as this test is performed where leakage is a major issue. There was a fiber elongation which
exceeds 5 percent, so heat treatment might be needed. In addition to the FEA performed field test
should be done before the installation of heat exchanger.
REFERENCES
80
1. Mcadams, W.H., Heat Transmission, (McGraw-Hill, New York), pp 430-441, 1954
2. Jenssen, S.K., Heat exchanger optimization, Chemical Eng. Progress, 65(7), pp 59, 1969
3. Sidney, K., and Jones, P.R., Programs for the price optimum design of heat exchangers,
British Chemical Engineering, April,pp 195-198, 1970
4. Steinmeyer, D.E., Energy price impacts design, Hydrocarbon Process, November, pp
205, 1976.
5. Steinmeyer, D.E., Take your pick – capital or energy, CHEMTECH, March, 1982
6. Peter, M.S. and Timmerhaus, K.D., Plant design and economics for chemical engineers,
pp 678-696 3rd Ed., McGraw-Hill, New York, 1981
7. Kovarik, M., Optimal heat exchanger, Journal of Heat Transfer, Vol. 111, May, pp 287-
293, 1989
8. Polley, G.T., Panjeh Shahi, M.H. and Nunez, M.P., Rapid design algorithms for shell-
and-tube and compact heat exchangers, Trans IChemE, Vol. 69(A), November,pp 435-
444, 1991
9. Jegede, F.O. and Polley, G.T., Optimum heat exchanger design, Trans IChemE, Vol.
70(A), March, pp 133-141, 1992
10. Saffar-Avval, M. and Damangir, E., A general correlation for determining optimum
baffle spacing for all types of shell-and-tube heat exchangers, International Journal of
Heat and Mass Transfer, Vol. 38 (13), pp 2501-2506, 1995
11. Poddar, T.K. and Polley, G.T., Heat exchanger design through parameter plotting, Trans
IChemE, Vol. 74(A), November, pp 849-852, 1996
12. Steinmeyer, D.E., Understanding ∆P and ∆T in turbulent flow heat exchangers, Chemical
Engineering Progress, June,pp 49-55, 1996
81
13. Soylemez, M.S., On the optimum heat exchanger sizing for heat recovery, Energy
Conversion and Management, 41, pp1419-1427, 2000
14. Murlikrishna, K. and Shenoy, U.V., Heat exchanger design targets for minimum area and
cost, Trans IChemE, Vol. 78(A), March, pp 161-167, 2000
15. Bevevino, J.W., ET. al., Standards of tubular exchanger manufacturing association,
TEMA, New York, 6th Edition, 1988
16. Mukherjee, R., Effectively design shell-and-tube heat exchangers, Chemical Engineering
Progress, February, pp 21-37, 1998
17. Crane, R.A., Thermal Aspects of Heat Exchanger Design, University of South Florida,
Department of Mechanical Engineering
18. Gulyani, B.B. and Mohanty, B., Estimating log mean temperature difference, Chemical
Engineering, November, pp127-130, 2001
19. Sukhatme, S.P., Heat exchangers, A Text Book on Heat Transfer, University Press India
Limited, pp 201-226 1996
20. Taborek, J., Shell-and-tube heat exchangers, Heat Exchanger Data Handbook,
Hemisphere, 1983
21. Poddar, T.K. and Polley, G.T., Optimize shell-and-tube heat exchangers design,
Chemical Engineering Progress, September,pp 41- 46, 2000
22. ASME 2007
23. Stanley Yokell , P.E. member of the ASME.
24. L.P. Zick , stress in cylindrical pressure vessel on saddle supports.
82