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Химикотехнологичният и
металургичен университет
Universidade de Vigo
Departamento de Ingeniería mecánica
3D DESIGN AND ANALISYS OF STANDARD
SHELL-TUBE HEAT EXCHANGER
Student: Brais Carballedo Sánchez
Tutor: Assoc. Prof. Veselin Iliev
Assist. Prof. Iliyan M.Lessev
Academic year 2014-2015
INDEX
I. INTRODUCTION 1. Introduction 1
2. Goal of the project 2
II. REVIEW OF THE LITERATURE
3. General Aspects of heat exchangers 2
3.1 Standards 3
3.2.Configuration 3
3.3.Components 4
3. Input data 8
III. EXPERIMENTAL PART
5. Design 9
5.1. Webbusterz Engineering Software 9
5.2. Algorithm 20
5.2.1. Calculating the initial geometry according Russian and
Bulgarian standards. 22
IV. DISCUSSION OF EXPERIMENTAL RESULTS
6. CFX Simulation in Ansys Workbench 24
6.1. Design modeller 24
6.2. Meshing [ICEM FCD] 29
6.3. Setup (CFX-Pre) 33
6.4 Thermodynamic analysis 35
6.4.1. Influence of temperature 36
6.4.2. Transient analysis 39
6.4.3. Environment influence 46
7. Autodesk Inventor 3D model 47
7.1. Components 48
7.2. Assembly 54
8. Thermal and structural analysis with Autodesk 3D model in
Ansys Workbench 64
8.1. Material properties 65
8.2. Meshing 67
8.3. Steady state thermal analysis 68
8.4. Static structural analysis 71
8.4.1. Temperature load 72
8.4.2. Pressure load 75
8.4.3. Temperature and pressure load 77
8.5. Linear buckling analysis 81
8.6. Modal analysis 88
8.7. Earthquake analysis 94
CONCLUSION 109
REFERENCES 110
ANNEX I 111
ANNEX II 113
ANNEX III 114
ANNEX IV 120
ANNEX V 121
INDEX OF FIGURES
Figure. 1. Main components of a shell-tube heat exchanger 7
Figure 2. Input data 8
Figure 3. Baffle cut 9
Figure 4. Triangular layout, pitch 9
Figure 5. Basic summary of proposed design (Webbusterz Engineering Software) 19
Figure 6. Main dimensions of the shell-tube heat exchanger 23
Figure 7. Channel cover (Design modeller of Ansys Workbench) 25
Figure 8. Tubes (Design modeller of Ansys Workbench) 25
Figure 9. Bundle of tubes (Design modeller of Ansys Workbench) 26
Figure 10. Tubes domain with channel covers (Design modeller of Ansys Workbench) 26
Figure 11. Shell (Design modeller of Ansys Workbench) 27
Figure 12. Shell´s nozzles (Design modeller of Ansys Workbench) 27
Figure 13. Detailed view of the intersection between the shell and the tubes after
subtract operation (Design modeller of Ansys Workbench) 28
Figure 14. Baffles and their respective sketch (Design modeller of Ansys Workbench) 28
Figure 15. Baffles (Design modeller of Ansys Workbench) 29
Figure 16. Ansys´ model of shell and tube heat exchanger, half view (Design modeller
of Ansys Workbench) 29
Figure 17. Name selections 30
Figure 18. Detailed view of the mesh 31
Figure 19. Temperature distribution of naphtha- diesel mix in the midplane 35
Figure 20. Temperature distribution of water liquid in the midplane 36
Figure 21. Heat transfer process between the shell and the environment 46
Figure 22. Tube (Autodesk Inventor) 48
Figure 23. Rectangular pattern detail direction 1 (Autodesk Inventor) 49
Figure 24. Rectangular pattern detail direction 2 (Autodesk Inventor) 49
Figure 25. Tubesheet (Autodesk Inventor) 49
Figure 26. Channel cover (Autodesk Inventor) 50
Figure 27. Tubes´ nozzles (Autodesk Inventor) 50
Figure 28. Shell´s nozzles (Autodesk Inventor) 51
Figure 29. Flange (Autodesk Inventor) 51
Figure 30. Seal (Autodesk Inventor) 52
Figure 31. Support (Autodesk Inventor) 53
Figure 32. Baffle (Autodesk Inventor) 53
Figure 33. Tubesheet (Autodesk Inventor; assembly) 54
Figure 34. Positioning of the tubesheet (Autodesk Inventor; assembly) 54
Figure 35. Insertion of the tubes (Autodesk Inventor; assembly) 55
Figure 36. Insertion of the tubes, both sides (Autodesk Inventor; assembly) 55
Figure 37. Geometry projected to extrude the shell (Autodesk Inventor; assembly) 56
Figure 38. Shell (Autodesk Inventor; assembly) 57
Figure 39. Rotational joint between the first flange and the shell (Autodesk Inventor;
assembly) 57
Figure 40. Rotational joint between the seal and the first flange (Autodesk Inventor;
assembly) 58
Figure 41. Rotational joint between flanges (Autodesk Inventor; assembly) 58
Figure 42. End view after placing the flanges in both sides (Autodesk Inventor;
assembly) 59
Figure 43. Detailed half vies of the rotational joint between the channel cover and the
second flange (Autodesk Inventor; assembly) 59
Figure 44. End view after placing the channel covers in both sides (Autodesk Inventor;
assembly) 60
Figure 45. Detailed view of bolts, nuts and washers used to fix the flanges (Autodesk
Inventor; assembly) 61
Figure 46. Rows of baffles (Autodesk Inventor; assembly) 62
Figure 47. Placing of tubes´nozzles (Autodesk Inventor; assembly) 62
Figure 48. Placing of shell´s nozzles (Autodesk Inventor; assembly) 63
Figure 49. 3D model performed in Autodesk Inventor (end section view) 64
Figure 50. 3D model performed in Autodesk Inventor (half view) 64
Figure 51. Mesh for steady thermal analysis static structural analysis 67
Figure 52. Steady state thermal analysis_1; boundary conditions 69
Figure 53. Steady state thermal analysis_1; Temperature distribution 69
Figure 54. Steady state thermal analysis_2; boundary conditions 70
Figure 55. Steady state thermal analysis_2; Temperature distribution 70
Figure 56. Static structural analysis; thermodynamic load; loads and supports 73
Figure 57. Static structural analysis; thermodynamic load; Total deformation 73
Figure 58. Static structural analysis; thermodynamic load, equivalent elastic strain 74
Figure 59. Static structural analysis; thermodynamic load, equivalent stress 74
Figure 60. Static structural analysis; pressure load; loads and supports 76
Figure 61. Static structural analysis; pressure load; equivalent stress 76
Figure 62. Static structural analysis; pressure load; detailed view of equivalent stress 77
Figure 63. Static structural analysis; pressure load; total deformation 77
Figure 64. Static structural analysis; temperature and pressure load; loads and 78
supports
Figure 65. Static structural analysis; temperature and pressure load; total
deformation 79
Figure 66. Static structural analysis; temperature and pressure load; equivalent elastic
strain 79
Figure 67. Static structural analysis; temperature and pressure load; directional
deformation Z axis 80
Figure 68. Static structural analysis; temperature and pressure load; directional
deformation X axis 80
Figure 69. Static structural analysis; temperature and pressure load; directional
deformation Y axis 80
Figure 70. Linear buckling; support´s mesh 82
Figure 71. Linear buckling; loads and supports 83
Figure 72. Buckling mode1; Total deformation 84
Figure 73. Buckling mode2; Total deformation 84
Figure 74. Buckling mode3; Total deformation 85
Figure 75. Buckling mode4; Total deformation 85
Figure 76. Buckling mode5; Total deformation 86
Figure 77. Buckling mode6; Total deformation 86
Figure 78. Modal analysis; mesh 90
Figure 79. Mode 1; Total deformation 91
Figure 80. Mode 2; Total deformation 91
Figure 81. Mode 3; Total deformation 92
Figure 82. Mode 4; Total deformation 92
Figure 83. Mode 5; Total deformation 93
Figure 84. Mode 6; Total deformation 93
Figure 85. Earthquake analysis; ground type A; X axis direction; Equivalent stress 102
Figure 86. Earthquake analysis; Ground type A; X axis direction; Directional
deformation 102
Figure 87. Earthquake analysis; Ground type A; Y axis direction; Equivalent stress 103
Figure 88. Earthquake analysis; Ground type A; Y axis direction; Directional
deformation 103
Figure 89. Earthquake analysis; Ground type B; X axis direction; Equivalent stress 103
Figure 90. Earthquake analysis; Ground type B; X axis direction; Directional
deformation 104
Figure 91. Earthquake analysis; Ground type B; Y axis direction; Equivalent stress 104
Figure 92. Earthquake analysis; Ground type B; Y axis direction; Directional
deformation 104
Figure 93. Earthquake analysis; Ground type C; X axis direction; Equivalent stress 105
Figure 94. Earthquake analysis; Ground type C; X axis direction; Directional
deformation 105
Figure 95. Earthquake analysis; Ground type C; Y axis direction; Equivalent stress 105
Figure 96. Earthquake analysis; Ground type C; Y axis direction; Directional
deformation 106
Figure 97. Earthquake analysis; Ground type D; X axis direction; Equivalent stress 106
Figure 98. . Earthquake analysis; Ground type D; X axis direction; Directional
deformation 106
Figure 99. Earthquake analysis; Ground type D; Y axis direction; Equivalent stress 107
Figure 100.Earthquake analysis; Ground type D; Y axis direction; Directional
deformation 107
Figure 101. Earthquake analysis; Ground type E; X axis direction; Equivalent stress 107
Figure 102. Earthquake analysis; Ground type A; X axis direction; Directional
deformation 108
Figure 103. Earthquake analysis; Ground type E; Y axis direction; Equivalent stress 108
Figure 104. . Earthquake analysis; Ground type E; Y axis direction; Directional
deformation 108
INDEX OF TABLES
Table 1. TEMA designation for shell-tube heat exchangers 4
Table 2. Physical properties of naphtha-diesel 11
Table 3. Physical properties of water 11
Table 4. Typical Baffle Clearance and Tolerance 15
Table 5. Initial parameters of the Shell tube heat exchanger according to the
necessary heat transfer surface (Russian standard) 22
Table 6. Main dimensions of the shell tube heat exchanger 23
Table 7. Relationship between the size of the mesh, the number of elements and the
outlet temperatures 31
Table 8. Temperature values of cold fluid at outlet for a constant hot fluid
temperature at inlet of 260℃ and with the cold fluid temperature at inlet
increasing linearly from 15 ℃ to 45℃ at intervals of 10℃ 36
Table 9. Temperature values of hot fluid at outlet for a constant cold fluid
temperature at inlet of 15℃ and with the hot fluid temperature at inlet
increasing linearly from 260 ℃ to 340℃ at intervals of 40℃ 37
Table 10. Temperature values of hot fluid at outlet for a constant hot fluid
temperature at inlet of 260℃ and with the cold fluid temperature at inlet
increasing linearly from 15 ℃ to 45℃ at intervals of 10℃ 38
Table 11.Temperature values of the hot fluid temperature at outlet for each
timestep with a constant cold fluid temperature at inlet of 15℃ 40
Table 12. Temperature values of the hot fluid at outlet for each timestep with
the cold fluid temperature at inlet increasing linearly 41
Table 13. Temperature values of hot fluid at outlet for each timestep with the
cold fluid temperature varying step by step 44
Table 14. Temperatures values at inlets and outlets for different environment temperatures 47
Table 15. Structural Steel properties; Engineering Data of Ansys Workbench 66
Table 16. Water liquid properties; Engineering Data of Ansys Workbench 66
Table 17. Naphtha-diesel properties; Engineering Data of Ansys Workbench 67
Table 18. Linear buckling analysis; List of modes and frequencies 87
Table 19. Modal analysis; List of modes and frequencies 94
Table 20. Earthquake analysis; Types of ground 96
INDEX OF GRAPHS
Graph 1. Tube side friction factor 16
Graph 2. Shell side friction factor 17
Graph 3. Fluids´ temperatures (cross flow) 21
Graph 4. Outlet cold fluid temperature depending on number of elements
of the mesh 32
Graph 5. Outlet hot fluid temperature depending on number of elements
of the mesh 32
Graph 6. Quality of the mesh´s elements 33
Graph 7. Temperature of cold fluid at outlet in function of the cold fluid temperature
at inlet which is increasing linearly from 15℃ to 45 ℃ at intervals of 10℃
and for a constant temperature of hot fluid at inlet of 260 ℃ 37
Graph 8. Temperature of hot fluid at outlet in function of the hot fluid temperature
at inlet which is increasing linearly from 260℃ to 340 ℃ at intervals of 10℃
and for a constant temperature of cold fluid at inlet of 15 ℃ 38
Graph 9. Temperature of hot fluid at outlet in function of the cold fluid temperature
at inlet which is increasing linearly from 15℃ to 45 ℃ at intervals of 10℃
and for a constant temperature of hot fluid at inlet of 260 ℃ 39
Graph 10. Inlet cold fluid temperature for each timestep; constant over time at
15℃ 40
Graph 11. IHot fluid temperature at outlet for each timestep with a constant
cold fluid temperature at inlet of 15℃ 41
Graph 12. Inlet cold fluid temperature for each timestep increasing linearly 42
Graph 13. Outlet hot fluid temperature for each timestep with the cold fluid
temperature at inlet increasing linearly 42
Graph 14. Comparison between the hot fluid temperature at outlet with a constant
cold fluid temperature at inlet of 15℃ (in red in the graph) and the hot fluid
temperature at outlet with cold fluid temperature at inlet increasing linearly
(in orange in the graph) 43
Graph 15. Inlet cold fluid temperature at inlet for each timestep 45
Graph 16. Outlet hot fluid at outlet for each timestep 45
Graph 17. Outlet hot fluid at outlet for each timestep 46
Graph 18. Linear buckling: Modes and frequencies 87
Graph 19. Modal analysis;modes and frequencies 94
Graph 20. Recommended Type 1 elastic response spectra for ground types A to E
(5% damping) 100
Graph 21. Recommended Type 2 elastic response spectra for ground types A to E
(5(% damping) 101
Graph 22. Data earthquake; frequency [Hz] vs. acceleration [𝑚/𝑠^2] 101
ABBREVIATIONS AND SYMBOLS
�̇�: flow rate
𝑻ª: temperature
𝑷𝒓: Prantd number
𝑸: heat transfer
𝑶𝑫: Outside diameter
𝑷𝒕: pitch
𝑳: length of the shell and tube heat exchanger
𝑫𝒊: inside diameter
∆𝑻𝒍𝒎: log mean temperatura difference
𝑹𝒆: Reynolds number
𝑵𝒖: Nusselt number
𝒉: heat transfer coeficient
𝑨𝒔: area for cross flow
𝑫𝒆: equivalent diameter:
𝒋𝒇: friction factor
𝑲: correction parameter
𝑹: crown radius
𝒓: knuckle radius
𝑫: diameter
𝒗: speed
𝑪: heat capacity
𝑭: surface
∆𝑻𝒂𝒗𝒆𝒓𝒂𝒈𝒆: average temperature
𝑲: correction parameter
𝒍: length of tubes
𝒍𝟐: distance between baffles
𝜹: thickness
𝒌: resistance
𝑼: overall heat transfer coefficient
𝑭: force
𝝀: multiplication factor
𝒘: frecuenquies
𝒇: mode shapes
𝑺𝒅(𝑻): design spectrum
T: vibration period of a linear single-degree-of-freedom system
𝒂𝒈: design ground acceleration on type A ground (𝑎𝑔 = 𝛾1 ∗ 𝑎𝑔𝑅)
S: soil factor;
𝑻𝑩: lower limit of the period of the constant spectral acceleration branch
𝑻𝑪 : upper limit of the period of the constant spectral acceleration branch;
𝑻𝑫: value defining the beginning of the constant displacement response range
of the spectrum;
q: behaviour factor;
b: lower bound factor for the horizontal design spectrum.
𝜸𝟏=importance factor
𝒂𝒈𝑹 =reference peak ground acceleration on type A ground
1
I. INTRODUCTION
1. Introduction
In recent years, the demand for energy resources has grown exponentially, as the same time
that the global development, but not with the amount of these resources. This has resulted in
an overall effort of the industries in optimizing their processes and the maximization of the
energy consumed. Often the plant productivity operation and treatment is related with the
effectiveness with which it is used and / or recovers the heat in determinate process points. It
is in this area, heat exchangers play an essential role.
Generally heat exchangers are designed for a specific service, depending on variables such as
process conditions, cost, space, etc. It is for this reason that the design work is an activity that
is performed frequently. For given conditions, there may be more than one design that meets
the requirements. When this occurs, the basis of selection generally is the cost.
The design procedure for most exchangers is based on trial and error. A preliminary
arrangement is assumed and then verifies, this makes the design work is long and slow. With
the advancement of computational capabilities, the iterative calculation has been simplified,
and should continue to improve gradually.
Heat exchangers are devices widely used in different types industries, particularly in the oil and
petrochemical industry. In recent years the international oil industry anticipates an aggressive
expansion plan for all its installations in order to increase production significantly in the
medium term, to satisfy the growing demand worldwide.
However, the oil business is highly competitive, the investments needed to build or extend any
installation are high and the profitability in these projects is very sensitive to such investment.
In this regard it is important to make a proper assessment and / or design of all teams that
form a plant in the oil industry. In this industry is very frequent to use heat exchangers in
various production processes.
2
2. Goal of the project
The aim of this project is to analyse, model and simulate a shell and tube heat exchanger
which is part of a crude oil refining installation.
With certain initial conditions and input data we will make several calculations to obtain the
necessary surface for the heat transfer, then we will collate on standards the approximate
geometrical dimensions required and we will try to build an initial prototype of our exchanger,
first in Ansys Workbench to start performing various analyses, mostly thermodynamic, and
after that the definitive model in Autodesk Inventor, which we will import to Ansys to do all the
final necessary analyses, thermodynamic and structural, to verify the good performance of the
exchanger.
II. REVIEW OF THE LITERATURE
3. General aspects of heat exchangers
A heat exchanger is a mechanism for transferring thermal energy between two or more fluids
through a solid surface or by direct contact of the fluids, without the use of external heat or
work. The fluids may be simple substances or mixtures.
The most common applications involve cooling, heating, evaporation or condensation of a fluid
stream, and recovery or re-injection system heat, distilled, fractionated or control process
fluids, among other.
In some heat exchangers, the fluids involved are in direct contact, in other heat transfer takes
place through a wall separating fluid, called heat transfer surface. The former are called heat
exchangers direct contact, and the following exchangers indirect contact. Among these
exchangers shell and tube heat exchangers are most commonly used in the process industries.
This is the type of heat exchanger which is frequently used in refineries.
Shell and tube heat exchanger provides a relationship between the area of heat transfer and
weight-volume rather big. Despite this, it is relatively easy to build in a variety of sizes, and
their mechanical properties allow it to withstand severe operating conditions.
It is one of the least expensive exchangers, maintenance is comparatively simple and in
addition can be easily cleaned and those failure prone components (gaskets and tubes) can be
easily replaced.
3
3.1. Standards
Nowadays exist different official standards which we must take into account.
The mechanical design of presure vessels, as the most maiority the equipment for Industrial
processes, are governed by different rules and codes. For all countries of Europe the PED
97/23/EC (pressure equipment directive) is the current rule for pressure vessels like shell and
tube heat exchangers. Some countries have furthermore her own rules like CODAP 2000
(France), PD 5500 (British) or AD 2000 (Germany). The European government allows all
companies in Europe to use his own rule for delivering to other European countries, too. This
country rules will be replaced with a global European rule. Its name is DIN EN 13445. This rule
is new for all countries in Europe. It is the counterpart to the American ASME Code and TEMA
Standard [1], which is present in many continents and sometimes in Europe, too.
3.2. Configuration
A wide variety of configurations used in the heat exchangers of shell and tube, depending on
the desired performance of heat transfer, pressure drop and the methods used to reduce
thermal stresses, prevent leakage, easy maintenance, withstand the pressures and
temperatures operation, and corrosion.
Although as we have seen above, there are different standards according to which build and
choose the right configuration for our exchanger but is relevant to quote the terminology that
TEMA has developed for the basic types of shell and tube heat exchangers. In this system, each
exchanger is designated with three letters, the first indicating the headstock, the second the
type of shell, and the third the rear head. We can see this classification in the next Table.
4
Table 1: TEMA designation for shell-tube heat exchangers
3.3. Components
As its name implies, shell and tube heat exchangers consists of a shell (a large pressure vessel)
with a bundle of tubes inside it. One fluid runs through the tubes, and another fluid flows over
the tubes (through the shell) to transfer heat between the two fluids. The set of tubes is called
a tube bundle, and may be composed of several types of tubes: plain, longitudinally finned,
etc.
5
Although there are a huge variety of specific designs, the basic components are as follows:
Tubes
The tubes are the fundamental components, providing the surface of heat transfer between
the fluid flowing inside the tubes and the shell. The tubes may be complete or soldiers and
generally are made from copper or steel alloys. Other alloys of nickel, titanium or aluminium
may be required for specific applications. In our case they are made from steel.
The tubes may be bare or finned. The extended surfaces are used when one of the fluids has a
transfer coefficient much lower than the other fluid heat. The double finned tubes can further
improve efficiency. The fins provide two to four times the area of heat transfer tube would
provide bare. In our case fins are not necessary.
The number of tube passes through the shell depends on the pressure drop available. At
higher speeds, increase the heat transfer coefficients, but also the friction losses and erosion
materials. Therefore, if the pressure loss is acceptable, it is advisable to have less number of
tubes, but of greater length in a small area. Generally the tube passes between 1 and 8. The
design standards have one, two or four tube passes. In multiple designs even numbers of steps
are used. Steps odd numbers are not common, and result in thermal and mechanical problems
in the manufacture and operation.
Selecting the spacing between tubes is a balance between a short distance to increase heat
transfer coefficient on the side of the shell, and the space required for cleaning. In most
exchangers, the relationship between the spacing tubes and the tube outer diameter varies
between 1.25 and 2. The minimum value is limited to 1.25 because at lower values, the union
between the tube and the tubesheet is becomes very weak, and can cause leaks at the joints.
Tubesheet
The tubes are held in place by being inserted into holes in the tube plate, to be fixed by
welding or expanding. The tubular plate is usually a simple metal plate that has been drilled to
accommodate the tubes (in the desired pattern), gaskets and bolts. In the event that an extra
leakage protection is required can be used a double tubesheet.
The space between the tubesheets must be open to the atmosphere so that any leaks can be
detected quickly. For most dangerous applications a triple tubesheet seals and even gaseous
recirculation system leakage can be used.
6
Tubesheet besides its mechanical requirements must be capable of withstanding the corrosive
attack of both exchanger fluids and should be electrochemically compatible with the material
of the tubes. Sometimes are constructed of steel covered under a metallurgical corrosion
resistant alloy carbon.
Baffles
There are two types of baffles, transverse and longitudinal. The purpose of the longitudinal
baffles is to control the general direction of flow of the shell side. For example, Type F, G and H
shells have longitudinal baffles. The transverse baffles have two functions, the most important
is to keep the tubes in position during the operation and prevent vibration caused by flow
induced vortices. Secondly they guide the flow of the shell to as close as possible to the
characteristics of the cross flow.
The most common type of baffle is simply segmented. The cut segment must be less than half
the diameter to ensure that adjacent baffles overlap in at least a complete row of tubes. For
liquid flow in the shell side the baffle cut is generally 20 to 25 percent.
Shell
The shell is the envelope of the second fluid. It is generally circular section and is made of a
steel plate formed into a cylindrical shape and welded longitudinally. Shells with small
diameters (up to 24 inches) can be made by cutting a desired diameter tube to the correct
length (pipe shells). The spherical shape of the shell is important in determining the diameter
of the baffles that can be inserted and the effect of leakage between the baffle and the shell.
Tube shells are usually rounder than the shell shifty.
In large exchangers the shell is made of low carbon steel whenever possible for reasons of
economy although other alloys can also be used where corrosion or high temperatures require
it.
The inlet nozzle plate usually has a plaque under it to prevent current high speed impinge
directly on top of the tube bundle. That impact can cause erosion, cavitation, and vibration. In
order to place this plaque and leave enough space between it and the shell so that the
pressure drop is not excessive it may be necessary to omit some tubes of circular pattern
completely. In this project we will not consider this plaque.
7
Nozzles
The nozzles are input and output ports, they simply direct the fluid flow to the inlet and outlet
of tubes and shell.
As the fluid side of the tubes is generally the most corrosive, the nozzles of the tubes are
typically made of alloy materials (compatible with the tubesheet). Should be coated instead of
solid alloys.
Channels (Heads)/Front channel and rear channel Channels or heads are required for shell-and-tube heat exchangers to contain the tube side
fluid and to provide the desired flow path.
As we have seen previously many types of channels are available.
The channel type is selected based on the application. Most channels can be removed for
access to the tubes to inspect them without disturbing the arrangement.
The rear channel is often selected to match the front channel. However, there can be
circumstances where they are different such as when removable bundles are used.
In the following figure we can see these main components of a shell-tube heat exchanger.
(Fig.1)
Fig. 1: Main components of a shell-tube heat exchanger [2]
8
4. Input data
As we have already mentioned our heat exchanger is part of a of a crude oil refining
installation and is entirely made of 2mm thick steel.
The heat exchanger has a function to cool the liquid mixture of 62,5% naphtha and 37,5 %
diesel with a flow rate of 200000 kg/hr from a temperature of 260C to 240C. The cooler
agent is water at 15C with a flow rate of 500000 kg/hr.
Fig. 2: Input data
To start performing the design of the heat exchanger we will also consider as starting
conditions the following characteristics:
-Shell: one-pass with 25% cut baffles (Fig.3)
-Tubes: The tubes are fixed, one-pass, with next configuration:
-Length: 5m
-Outer diameter= 50,8 mm
-Triangular layout; pitch=90mm (Fig.4)
9
Fig. 3: Baffle cut Fig.4: Triangular layout, pitch
Some of these parameters can be modified later if it is necessary to satisfy the requirements;
we just take them as a reference to start our calculation, the important thing is that our heat
exchanger design meets the objective of cooling the mixture with water in the corresponding
terms.
The fluids are not corrosive, but the mixture is severely fouling.
III. EXPERIMENTAL PART
5. Design
Based on the above characteristics and initial data to be met by our exchanger we must obtain
a geometric features that allow us to design a first prototype with which we can perform
various analyses to study and verify its operation.
5.1. Webbusterz Engineering Software
Nowadays we have several softwares to perform this first approximation. There are many
companies that offer their own program for customers based on certain initial conditions can
get an idea of what they need.
10
In our case, in the first instance, we chose to use the official version of the program Shell
&Tube Heat Exchanger Design Software provided by Webbusterz Engineering Software [3].
Step by step different data were introduced depending on the starting conditions and helping
with the databases that the program gives us.
Because of the mixture is severely fouling it must go through the inside of the tubes which will
facilitate cleaning. And therefore the cold fluid, water, will be inside the shell.
Step 1: Calculation of average temperature
In this first step we will indicate the input and output temperature of fluids and the mass flow
corresponding to each one.
Shell Side Data: cold-fluid
Liquid: Water
Temperature: From 15℃ to 18,72℃ (the final temperature is calculated by the
program)
Flow Rate: 500.000 𝑘𝑔/ℎ
Mean temperature: 16,86℃
Tube Side Data: hot-fluid
Liquid: Naphtha-Diesel
Temperature: From 260℃ to 240℃
Flow rate∶ 200000 𝑘𝑔/ℎ
Mean Temperature: 250℃
Step 2: Physical Properties
In this step the physical properties of the fluids were defined.
The properties of the mixture were calculated depending on the properties of naphtha and
diesel components [4] and their respective percentage.
Hot fluid: mixture of naphtha and diesel
11
Cold fluid: water
Physical properties
Water
Density [𝑘𝑔/𝑚^3 ] 999
Viscosity [𝑘𝑔/(𝑚 ∗ 𝑠)] 0,001002
Thermal Conductivity [𝑊/(𝑚 ∗ 𝐾)] 0,602
Heat capacity [𝐾𝐽/(𝑘𝑔 ∗ 𝐾)] 4,182
*Prand number (Pr) 13,98
**Maximum pressure drop [bar] 1
Table 3: Physical properties of water
* Prand number is calculated by the program.
**The value of the pressure drop if not so relevant for this first calculation.
Step 3: Heat Duty
The amount of heat required to perform the operation was denoted by the program.
Duty 𝑄 = 2160,70𝑊
Physical properties
Diesel 37,5%
Naphtha 62,5%
Mixture Naphtha-diesel
Density [𝑘𝑔/𝑚^3 ] 830 751 780,625
Viscosity [𝑘𝑔/(𝑚 ∗ 𝑠)] 0,002473 0,000000529 0,000928
Thermal Conductivity [𝑊/(𝑚 ∗ 𝐾)] 0,15 0,1164 0,129
Heat capacity[𝐾𝐽/(𝑘𝑔 ∗ 𝐾)] 1,75 2,06 1,94375
*Prand number(Pr) 6,96
**Maximum pressure drop [bar] 1
Table 2: Physical properties of naphtha-diesel
12
The overall heat transfer coefficient was calculated once the corresponding fluids were
selected in the database, in our case:
-Hot fluid: refinery hydrocarbons
-Cold fluid: Water
Assumed Overall Heat Transfer Coefficient = (400 + 550)/2 = 475 𝑊/(𝑚2 ℃)
Step 4: Dimensional characteristics
The different features of the exchanger were defined as follow:
Exchanger type: Fixed-tube plate
Exchanger Layout:
-Number of Shell Passes=1
-Number of Tube Passes=1
The diameter required for the shell was estimated by the software.
-Shell diameter: 339𝑚𝑚
Now the dimensional data corresponding to the bundle of tubes.
Tube size
-Outside Diameter (OD) = 50,8𝑚𝑚
-Pitch (Pt) = 90𝑚𝑚
-Tube Length (L) = 5𝑚
-Inside Diameter (Di) = 19,86𝑚𝑚
-BWG = 12
-Tube thickness= 2,769
-Tube arrangement=Triangular
13
The relation between Tube size/BWG/Thickness was found in the database of Standard Pipe
Sizes provided by the software.
Step 5: Calculation of Log mean temperature difference and True temperature
difference
The fluids are in countercurrent because this configuration provides better performance. The
logarithmic temperature difference was calculated as well as the true difference temperature
through the R and S parameters and temperature correction coefficient.
-Log Mean Temperature Difference
∆𝑇𝑙𝑚 = 233,0452℃
-True Temperature Difference
𝑅 = 5,3763
𝑆 = 0,0152
Estimate Temperature Correction Factor= 0,9997711
True Temperature Difference= 232,9986℃
Step 6: Heat Transfer Area
The program provides the heat transfer area required in m^2:
-Heat transfer area = 19,523𝑚2
Step 7: Number of tubes
The number of tubes required and other parameters was calculated; we must adjust the
number to an integer.
Area of one tube= 0,239𝑚2
Number of tubes= 81,55 Adjust= 82
14
-Number of tubes per pass = 82
-Tube cross sectional area = 0,00031𝑚2
-Area per Pass = 0,0254𝑚2
-Tube side volumetric flow = 0,0712 𝑚^3/𝑠
-Tube side velocity = 2,80 𝑚/𝑠
Step 8: Tube Side Heat transfer Coefficient and Baffles
-Reynolds Number (Re)= 46800
-Prandt number (Pr)= 13,98
-𝐿/𝐷𝑖 = 151
-Nusselt Number (Nu), we must choose the right correlation for calculate it according with the
Reynolds number calculated above, in our case:
Gnielinski:
𝑁𝑢 =𝑓
2∗ [𝑅𝑒 − 1000] ∗
𝑃𝑟
1 + 12,7∗ [(
𝑓
2∗ 0,5] ∗ [(𝑃𝑟 ∗ 0,67) − 1] = 313,38
-Heat Transfer Coefficient (hi) = 2035,34 𝑊/(𝑚^2 ∗ ℃)
-Baffle Spacing = 67,8𝑚𝑚
-Number of baffles = 43
Step 9: Baffles cut and type/tolerance
Information about the baffles: segmental baffles with 25%cut
We must choose the most adequate clearance according with the diameter of our shell.
15
Diameter _shell Clearance Tolerance
Plate shells 6-25 in(152-635mm)
1/8in (3,2mm) -1/32in(-0,8mm)
Table 4: Typical Baffle Clearance and Tolerance
Baffle diameter = 335,8𝑚𝑚
Step 10: Shell Side Heat Transfer Coefficient
-Area for Cross Flow (As) = 0,0165𝑚2
-Equivalent Diameter (De)= 325,17𝑚𝑚
-Shell Side Volumetric Flow= 0,139 𝑚^3/𝑠
-Shell Side Velocity= 8,43 𝑚/𝑠
-Mass Velocity= 8417,51 𝑘𝑔/(𝑚^2 ∗ 𝑠)
-Reynolds Number= 2730000
-Prandt Number= 6,96
-Nusselt number, according with Reynolds number we should choose the next correlation:
𝑁𝑢 = 𝐽ℎ ∗ 𝑅𝑒 ∗ 𝑃𝑟0,33 ∗ [𝑢
𝑢𝑤𝑎𝑙𝑙]
0,14
= 36251,15
With our number of Reynolds for the shell we choose the parameter Jh in the chart: Jh=0,007
-Heat Transfer Coefficient (ℎ𝑜) = 67113,18 𝑊/(𝑚^2 ∗ ℃)
Step 11: Material of construction
Our heat exchanger is made with steel, which has a thermal conductivity (K) of 43 𝑊/(𝑚 ∗ ℃)
Overall heat transfer coefficient
Design Uo,calc= 744,3596 𝑊/(𝑚^2 ∗ ℃)
Clean Uo,calc= 1975,4311 𝑊/(𝑚^2 ∗ ℃)
16
Assume Uo,ass= 475 𝑊/(𝑚^2 ∗ ℃)
Effectiveness and Number of Transfer Units
Effectiveness= 0,125
Number of Transfer Units= 0,14
Thermal Capacity Ratio= 0,186
Step 12: Pressure Drop for inlet and outlets nozzles
Tube side Pressure Drop
Friction factor (𝑗𝑓) = 0,0035 in function of Reynolds number for tube side and according with
next chart:
Graph 1: Tube side friction factor
Tube side pressure drop= 0,205916314𝑏𝑎𝑟
Maximum Pressure Drop= 1𝑏𝑎𝑟
Shell side Pressure Drop
Friction factor (jf) = 0,025 in function of Reynolds number for the shell and according with the
next chart:
17
Graph 2: Shell side friction factor
Shell side pressure drop= 3,274925505𝑏𝑎𝑟
Maximum Pressure Drop= 1𝑏𝑎𝑟
Step 13
-Shell side
Material of construction: Carbon Steel Permissible stress= 145𝑁/𝑚𝑚^2
Corrosion allowance= 1
Working pressure = N/(mm)^2 (the working pressure is not a key parameter now, it will be
evaluated later)
Working temperature= 16
Head
Crown radius 𝑅 = 0,8 ∗ 𝐷 = 0,8 ∗ 339 = 271,2𝑚𝑚
Knuckle radius 𝑟 = 0,154 ∗ 𝐷 = 0,154 ∗ 339 = 52,206𝑚𝑚
18
Nozzles
Inlet/outlet= 245𝑚𝑚
The procedure followed to calculate the diameter of the inlet and outlet nozzles was as
follows:
1º- A fluid speed of 3 𝑚/𝑠 was set, this value depends on the installation where will place our
heat exchanger.
2º- With this speed and the corresponding rate flow of each fluid the required diameter for the
nozzles was calculated according with the next equation:
𝑄 = 𝜋 ∗𝐷2
4∗ 𝑣
So the outer diameter required:
𝐷 = √4 ∗ 𝑄[
𝑚3
𝑠 ]
𝜋 ∗ 𝑣[𝑚𝑠 ]
Nozzle of the shell:
𝐷𝑛𝑜𝑧𝑧𝑙𝑒_𝑠ℎ𝑒𝑙𝑙 = √4 ∗ 500000
𝑘𝑔ℎ
∗1
999𝑚3
𝑘𝑔∗
1ℎ3600𝑠
𝜋 ∗ 3𝑚𝑠
≅ 243𝑚𝑚
Nozzles of the tubes:
𝐷𝑛𝑜𝑧𝑧𝑙𝑒_𝑡𝑢𝑏𝑒𝑠 = √4 ∗ 200000
𝑘𝑔ℎ
∗1
780,625𝑚3
𝑘𝑔∗
1ℎ3600𝑠
𝜋 ∗ 3𝑚𝑠
173,8𝑚𝑚
Once known the required diameter we have checked in the following website
http://www.husltd.com/en/products/41873-seamless-pipes [5] from a reputable
manufacturer of pipes to select the diameter on the market that comes closest to ours, for it
must also select a certain thickness. Then we conclude:
19
Shell:
Outer diameter required= 243𝑚𝑚
Official diameter= 245𝑚𝑚
Tubes:
Outer diameter required= 173,8𝑚𝑚
Official diameter= 178𝑚𝑚
Finally, the program give us a design proposal, we can check it in the next figure (Fig.5):
Fig. 5: Basic summary of proposed design (Webbusterz Engineering Software)
The program also offers us a data sheet where all the characteristics of the exchanger are
collected as the fluid properties, the desing data, the thermal design and the exchanger
configuration. We can review this data sheet in the ANNEX I.
With these results we have tried to do a first approximation model but we have noticed some
geometric inconsistencies regarding the area and the number of tubes and also with the
20
nozzles, therefore we can conclude that this program is not a very reliable source to
successfully perform our modelling and subsequent simulation.
5.2. Algorithm
As the program does not provide us completely correct results we opt for performing the
design by hand with the help of a calculation algorithm:
The first step was calculating the heat transfer necessary to carry out the operation:
𝑄1 = 𝑄2 = 𝑄
𝑚1 ∗ 𝐶1 ∗ (𝑡1𝑜𝑢𝑡 − 𝑡1𝑖𝑛) = 𝑚2 ∗ 𝐶2 ∗ (𝑡2𝑜𝑢𝑡 − 𝑡2𝑖𝑛)
500000𝑘𝑔
ℎ∗ 4182
𝐽
𝑘𝑔º𝐶∗ (𝑡1𝑜𝑢𝑡 − 15)º𝐶 = 200000
𝑘𝑔
ℎ∗ 1943,75
𝐽
𝑘𝑔ℎ(240 − 260)º𝐶
= 7775.000.000𝐽
ℎ= 21597222,222𝑊 = 𝑄
Then, clearing the above equation the water outlet temperature was calculated:
𝑡1𝑜𝑢𝑡 = 18,7183℃
And the area necessary for the heat transfer was calculated applying the following formula.
𝐹 =𝑄
𝐾 ∗ ∆𝑡𝑎𝑣𝑒𝑟𝑎𝑔𝑒
To do this we must first calculate the average working temperature, as we are working with
counter flow fluids (Graph 3) the procedure was as follows:
21
Graph 3: Fluids´ temperatures (cross flow)
∆𝑡𝑖𝑛 = 260℃ − 18,7183℃ = 241,2817℃
∆𝑡𝑜𝑢𝑡 = 240℃ − 15℃ = 225℃
∆𝑡𝑖𝑛
∆𝑡𝑜𝑢𝑡=
241,2817
225= 1.0724 ≤ 2
∆𝑡𝑎𝑣𝑒𝑔𝑎𝑟𝑒 =241,2817℃ + 225℃
2= 233,14085℃ = 233,14085°𝐾
The K correction parameter is defined according with the next equation but checking in the
corresponding bibliography the following approximated value was taken:
𝐾 =1
1ℎ𝑤𝑎𝑡𝑒𝑟
+ 𝜏𝑟𝑒𝑠𝑠𝑖𝑠𝑡𝑒𝑛𝑐𝑒 + 𝜏𝑑𝑖𝑟𝑡𝑦 +1
ℎ𝑛𝑎𝑝ℎ𝑡𝑎_𝑑𝑖𝑒𝑠𝑒𝑙
=120 + 270
2= 175
𝑊
𝑚2𝐾
Substituting the respective values in the formula we can conclude that the area required for
the heat transfer is:
𝐹 =2159722,222𝑊
195𝑤
𝑚^2𝐾 ∗ 233,14085°𝐾= 𝟐𝟏, 𝟖𝟖𝟐𝟑𝒎𝟐
With this value we can check in the official standards tables which dimensions are
recommended for our heat exchanger.
22
5.2.1 . Calculating the initial geometry according Russian
and Bulgarian standards.
Once the necessary surface for heat transfer occurs is known we check in the official standards
to find the approximate parameters that allow us to start building an initial prototype of our
heat exchanger.
Specifically we will use the Russian standards GOST 15118-79, GOST-15120-79 and 15122-79
[6]. These standards provide us an approximate initial relationship between the necessary heat
transfer area in 𝑚2, the diameter of the shell, the number of tubes and the distance between
the baffles, in this case for single-pass exchangers and for a tubes diameter of 25mm with
2mm thick.
The required surface must be at least 22𝑚2 to ensure the correct heat transfer. As we can see
in the corresponding standard table attached in ANNEX II the nearest surface that meets this
requirement is 26𝑚2 which correspond to the following parameters (Table 5):
Shell diameter [mm] Number of tubes Heat transfer
surface[𝒎𝟐]
Distance between
bafles[mm]
400 111 26 250
Table 5: Initial parameters of the Shell tube heat exchanger according to the necessary heat transfer surface (Russian standard)
After these initial parameters were obtained we turn to other standards that provide us more
detailed information about the geometrical dimensions of the different parts forming the heat
exchanger.
We have used the Bulgarian standard BDS_EN_ISO_16812 [7], which specifies the main
parameters for horizontal heat exchanger with stable tubes, with diameter under 600mm and
with basic pressure until 1,6 MPa. To check some small details in some parts of the geometry
we also will get help with the following Bulgarians standards:
-BDS EN 1092-1:2008: Specifies requirements for circular steel flanges.
-BDS EN 5643:1984: Glossary of refrigeration, heating, ventilation and air conditioning terms.
-BDS 11767:1974: Chemical equipment and oil refining. Chandeliers vertical supports of vessels
and equipment. Design and basic dimensions.
23
-BDS EN 1514-1:1997: Flanges and their joints. Dimensions of gaskets for PN-designated
flanges.
In the ANNEX III we can review these standards.
In Fig.6 can be seen the dimensions, the main of them are written in the table 6.
Fig. 6: Main dimensions of the shell-tube heat exchanger
Table 6: Main dimensions of the shell tube heat exchanger
Diameter of the shell 400mm
Diameter of tubes 25mm
Length of tubes( l ) 3000mm
Triangular layout; pitch 32mm
Number of tubes 104
Number of baffles 20
Distance between baffles (l2) 250mm
Length of heat exchanger (L) 3600mm
24
As we can see in the above table after collating in Russian and Bulgarian standards the
recommended geometric characteristics consistent with the necessary surface for heat
transfer occurs we see that some of the parameters as the diameter of the tubes ,the pitch ot
the length of the heat exchanger differ from their initial values. However this design is
guaranteed with our exchanger fulfill its task of cooling the mixture of naphtha-diesel in the
corresponding terms.
IV. DISCUSSION OF EXPERIMENTAL RESULTS
6. CFX analysis in Ansys Workbench
After knowing the geometry a first prototype of our shell and tube heat exchanger was built to
later analyse his behaviour under different conditions. To do that we have helped with ANSYS
Workbench platform[8], an engineering simulation program which is used to model
engineering projects and evaluate the risks and benefits in a virtual environment. It has
integrated different modules to let you work with different kinds of analysis. A CFX project was
created, a computational fluid dynamics code which is a branch of fluid mechanics that uses
numerical methods and algorithms to solve and analyse problems that involve fluid flows.
Finally we will discuss about the different results of our thermodynamic analysis to guarantee
the correct performance of the heat exchanger in stationary and transient conditions and
under different changes of temperature. After obtaining these results we will make other kind
of analysis to evaluate other magnitudes as the stress or strain.
6.1. Design modeller
The first step of the analysis is to do the heat exchanger geometry in Design Modeller. First the
channel cover was made revolutionizing the sketch created in plane XY (in yellow on Fig.7). To
do that in addition to the aforementioned standards we have also helped with the German
standard DIN 28013 for semi ellipsoidal head [9].
25
Fig. 7: Channel cover (Design modeller of Ansys Workbench)
To create the tubes, this cover shell was frozen to avoid the mixing of bodies. The tubes were
extruded from a sketch in a new plane on the bottom of the cover (in yellow on Fig.8). After
that with the help of the command “pattern” all the tubes were created according with the
pitch established ,first a quarter was performed and then with “mirror” body operation the
bundle of tubes was completed (Fig.9):
Fig. 8: Tubes (Design modeller of Ansys Workbench)
26
Fig. 9: Bundle of tubes (Design modeller of Ansys Workbench)
Then the cover was frozen to obtain only one body and with “mirror” body operation again the
whole domain of the tubes with the front l and rear channel cover was completed (Fig.10).
Fig. 10: Tubes domain with channel covers (Design modeller of Ansys Workbench)
The tubes were frozen to make another independent body, the shell. A plane separate 20 mm
from the cover was created leaving the space suitable for the tubesheet and the circle sketch
was extruded (in yellow on Fig.11):
27
Fig. 11: Shell (Design modeller of Ansys Workbench)
The next step was to create two new planes in one and another side of the shell separated
from it by a distance corresponding to the height of the nozzles. Then the cylinders with the
suitable diameter were extruded (in yellow on Fig.12).
Fig. 12: Shell´s nozzles (Design modeller of Ansys Workbench)
After with “subtract” body operation the intersection between tubes and shell was performed
to obtain the corresponding holes for tubes in the shell (Fig.13):
28
Fig. 13: Detailed view of the intersection between the shell and the tubes after subtract operation (Design modeller of Ansys Workbench)
The last step is to model and place the baffles. The two bodies were frozen to create the
baffles as independents bodies. A new sketch was made in a new plane located in a
determinate distance from the start of the shell. It was extruded and with the help of
command “patter” the first row of the baffles was created. To make the other row the mirror
function was used to copy all of the baffles and then moving them (Fig.14, Fig.15):
Fig. 14: Baffles and their respective sketch (Design modeller of Ansys Workbench)
29
Fig. 15: Baffles (Design modeller of Ansys Workbench)
Finally symmetry tool was applied to see better all the details of our heat exchanger (Fig.16).
The two bodies which forming our heat exchanger were grouped in only one part to work
easier later.
Fig. 16: Ansys´ model of shell tube heat exchanger, half view (Design modeller of Ansys Workbench)
6.2. Meshing [ANSYS ICEM FCD]
To solve the CFD equations by finite volume method we need an appropriate mesh for our
heat exchanger. The meshing tools of ANSYS workbench have the benefit of being highly
automated so this will simplify the mesh generation process. When the ANSYS Meshing
application is launched from the ANSYS Workbench Project Schematic, the physics preference
will be set based on the type of system being edited; in our case for a Mechanical Model
30
system the Mechanical physics preference is used. It is interesting for us to group some
geometric faces and regions of mesh together and assigned names in order to located them
and work easier and also to be available in CFX-Pre to define the boundary conditions.
For our heat exchanger we should define the following name selections (Fig.17)
Inlet cold fluid
Outlet cold fluid
Inlet hot fluid
Outlet hot fluid
Symmetry
Fig. 17: Name selections
After that we should check and define the mesh attributes starting for the sizing of the mesh.
Once the values of these parameters were defined we can generate the mesh and visualize the
result checking the size and the number of nodes and elements (Fig.18). The mesh must be the
appropriate to guarantee reliable results. To be sure of that several simulations were
performed with different minimum mesh size or different number of elements to later
compare the results between the temperature in the outlet of cold and hot fluid in each case.
31
Fig. 18: Detailed view of the mesh
The results were exported to Excel to visualize them better and see the tendency in a graphic
(Table 7; Graph 4; Graph 5).
MeshMinSize[mm] 5,5 6 7 8 9
Nº Elements 8168778 6516817 4220451 2991855 2132173
OuletColdFluidTemp[C] 21,927 21,924 21,919 21,946 22,056
OutletHotFluidTemp[C] 184,873 185,288 185,990 186,370 186,967
Table 7: Relationship between the size of the mesh, the number of elements and the outlet temperatures
32
Graph 4: Outlet cold fluid temperature depending on number of elements of the mesh
Graph 5: Outlet hot fluid temperature depending on number of elements of the mesh
We could prove that it was not necessary more than four millions of elements, after that point
the difference between the results for the outlet temperature of cold fluid is insignificant, the
final temperatures are practically the same (fig. 13), in the results of the outlet temperature of
hot fluid we can see a slight difference but this would be approximately only one degree so we
could conclude that 7mm of mesh sizing is enough to make our simulation.
21.850
21.900
21.950
22.000
22.050
22.100
8168778 6516817 4220451 2991855 2132173
Ou
tle
tCo
ldFl
uid
Tem
p [
℃]
Nº Elements
183.500
184.000
184.500
185.000
185.500
186.000
186.500
187.000
187.500
8168778 6516817 4220451 2991855 2132173
Ou
tle
tHo
tFlu
idTe
mp
[℃
]
Nº Elements
33
Although with greater number of elements we obtain best results, this translates into many
equations and from the computational point of view is impractical excessive refinement if this,
like in our case, is not strictly necessary.
On the other hand, the meshing tools of Ansys Workbench offer the possibility of visualizing
various characteristics of the mesh graphically. If we choose in the metric mesh the element
quality option we can see in a graphic the quality of the different numbers of elements, we
notice that the most of them have a quality close to 90%, with which we can verify the good
quality of our mesh. (Graph 6)
Graph 6: Quality of the mesh´s elements
6.3. Setup [CFX-Pre]
After meshing the physics-definition pre-processor was loaded, CFX-Pre [10], to define all the
necessary to make a correct analysis. Our fist analysis was a steady analysis, which it is to say
the magnitudes are constants with the time. Automatically the setup after importing the mesh
creates the domains and the respective interface between them. We have two domains one
for the shell and the another one for the tubes and the fluid interface where the fluids will be
exchange the heat transfer. In addition we must define the boundaries according with the
regions that we have defined before creating the mesh: the inlets, the outlets and the
34
symmetry for shell and tubes. After that we could define the boundary conditions and the
other parameters in relation with our heat transfer between naphtha-diesel and water.
1º Domain: shell
In the shell circulate the cold fluid, water, according with the following terms:
Reference pressure = 1 atm
Non buoyant and stationary model
Homogeneous model
Wall roughness: smooth wall
Heat transfer: conservative interface flux
- Inlet cold fluid
Mass flow rate: 138, 88 kg/s
Flow direction: normal to boundary condition
Turbulence: medium intensity 5%
Heat transfer: static temperature: 15C
Volume fraction: water=1; naphtha-diesel=0
- Outlet cold fluid
Flow regime: subsonic
Static pressure: 506625 Pa
- Shell default
Mass and momentum=no slip wall
Smooth wall
Heat transfer= adiabatic
- Symmetry
2º Domain: Tubes
As we have already explained previously because of the mixture is severely fouling it must go
through the inside of the tubes which will facilitate cleaning, so in the tubes circulate the hot
fluid, naphtha-diesel, according with the following terms, which are the same of the shell.
Wall roughness: smooth wall
Heat transfer: conservative interface flux
35
- Inlet cold fluid
Mass flow rate: 55.5555 kg/s
Flow direction: normal to boundary condition
Turbulence: medium intensity 5%
Heat transfer: static temperature: 240C
Volume fraction: water=0; naphtha-diesel=1
- Outlet cold fluid
Flow regime: subsonic
Static pressure: 506625 Pa
-Tubes default
No slip wall
Smooth wall
Heat transfer adiabatic
-Symmetry
6.4. Thermodynamic analysis
After finishing with the geometry, generating the appropriate mesh and checking and defining
all the setup parameters we could start doing some simulations to check the behaviour of our
model with respect some important parameters as the temperature.
First of all we had a look of the variation of the temperature of both fluids in the interface (Fig.
19; Fig. 20):
Fig. 19: Temperature distribution of naphtha-diesel mix in the midplane
36
Fig. 20: Temperature distribution of the water liquid in the midplane
We can see that the distribution of the temperature is symmetric as expected, occurring
temperatures higher in the central tubes.
6.4.1. Influence of temperature
To complete our steady thermodynamic analysis a simulation was done to see the influence of
changing the values of temperature in the inlet cold/hot fluid with a constant temperature in
the inlet hot/cold fluid.
In the first one the temperature of inlet cold fluid was increased from 15C till 45C at intervals
of 10C for a constant temperature of the inlet hot fluid of 260C (Table 8).
InletColdFluidTemp[℃] InletHotFluidTemp [℃] OutletColdFluidTemp[℃]
15 260 22,51
25 260 32,23
35 260 41,92
45 260 51,824
Table 8. Temperature values of cold fluid at outlet for a constant hot fluid temperature at inlet of 260℃ and with the cold fluid temperature at inlet increasing linearly from 15 ℃ to
45℃ at intervals of 10℃
37
Graph 7: Temperature of cold fluid at outlet in function of the cold fluid temperature at inlet which is increasing linearly from 15℃ to 45 ℃ at intervals of 10℃ and for a constant
temperature of hot fluid at inlet of 260 ℃
As the graphic show us, if we increase the inlet temperature of the water the outlet
temperature also increase linearly.
Then we have done the same maintaining constant the inlet cold temperature and changing
the values of the inlet hot temperature fluid, naphtha-diesel.
For a constant temperature of 15℃ for the cold fluid the same happens to hot fluid, the
variation of the outlet temperature respect on inlet temperature is linearly (Table 9).
InletColdFluidTemp[℃] InletHotFluidTemp[℃] OutletHotFluidTemp[℃]
15 260 192,86
15 300 231,588
15 340 270,323
Table 9: Temperature values of hot fluid at outlet for a constant cold fluid temperature at inlet of 15℃ and with the hot fluid temperature at inlet increasing linearly from 260 ℃ to
340℃ at intervals of 40℃
0
10
20
30
40
50
60
15 25 35 45
Ou
tle
tCo
ldFl
uid
Tem
p [
℃]
InletColdFluidTemp [℃]
38
Graph 8. Temperature of hot fluid at outlet in function of the hot fluid temperature at inlet which is increasing linearly from 260℃ to 340 ℃ at intervals of 10℃ and for a constant
temperature of cold fluid at inlet of 15 ℃
Now we check the variation of the outlet temperature of hot fluid in function of the inlet
temperature of cold fluid with a constant temperature in the inlet hot fluid of 260℃ (Table
10).
150
170
190
210
230
250
270
290
260 300 340
Ou
letH
otF
luid
Tem
p [
℃]
InletHotFluidTemp [℃]
InletColdFluidTemp[℃] InletHotFluidTemp[℃] OutletHotFluidTemp[℃]
15 260 192,86
25 260 203,59
35 260 205,99
45 260 219,313
Table 10. Temperature values of hot fluid at outlet for a constant hot fluid temperature at inlet of 260℃ and with the cold fluid temperature at inlet increasing linearly from 15 ℃ to
45℃ at intervals of 10℃
39
We can notice that if the temperature of water is between 25C and 35C the difference is not
so important and it not happen the same with the extreme temperatures of cold fluid (15C
and 45C) where the change is more significant, so the behaviour in this case is not linearly.
NOTE: To ratify and properly completed this analysis we must also check the variation of the
outlet hot fluid temperature in function of the inlet cold fluid temperature not only for a
constant temperature of 260C in the inlet hot fluid also for 300C and 400C.
6.4.2. Transient analysis
Until now all of our simulations were steady state but our next analysis was transient to check
the effects and changings of the temperature with the time.
In a first analysis we have maintained the inlet temperature of the cold fluid, water, at a
constant temperature of 15℃ to see how the hot fluid temperature at outlet varies.
For the purpose, we have defined six time steps of 10 seconds each one cycling 60 seconds in
total.
Graph 9: Temperature of hot fluid at outlet in function of the cold fluid temperature at inlet which is increasing linearly from 15℃ to 45 ℃ at intervals of 10℃ and for a constant
temperature of hot fluid at inlet of 260 ℃
175
180
185
190
195
200
205
210
215
220
225
15 25 35 45
Ou
letH
otF
luid
Tem
p [
℃]
InletColdFluidTemp [℃]
40
We can see the results in the following table where are specified the values of the cold fluid
temperature at inlet and the values of the hot fluid temperature at outlet for each time step.
Likewise we have created a chart in Excel to visualize better these results.
Timesteps [s] InletColdFluidTemp [℃] OutletHotFluidTemp [℃]
0 15 260.000
10 15 218.247
20 15 216.461
30 15 216.668
40 15 216.987
50 15 216.905
60 15 216.674
Table 11: Temperature values of the hot fluid temperature at outlet for each timestep with a constant cold fluid temperature at inlet of 15℃
Graph 60: Inlet cold fluid temperature for each timestep; constant over time at 15℃
0
2
4
6
8
10
12
14
16
0 10 20 30 40 50 60
Inle
tCo
ldFl
uid
tem
p [
C]
Timesteps [s]
InletColdFluidTemp
41
Graph 11. Hot fluid temperature at outlet for each timestep with a constant cold fluid temperature at inlet of 15℃
We note that the hot fluid temperature at outlet passes during the first seconds from the
initial condition of 260℃ to approximately its final temperature, approximately 217℃,
oscillating later around this value.
Furthermore another analysis was made to evaluate also the variation of the hot fluid
temperature at outlet but this time increasing linearly the cold fluid temperature at inlet
according to the following equation:
15 +1
3∗ (
𝑡
1[𝑠])
We can see the results in the following table and graphs.
Timesteps [s] InletColdFluidTemp [℃] OutletHotFluidTemp [℃]
0 15 260
10 18,333 218,628
20 21,667 217,514
30 25 218,312
40 28,333 219,215
50 31,667 219,71
60 35 220,075
Table 12. Temperature values of the hot fluid at outlet for each timestep with the cold fluid
temperature at inlet increasing linearly
200.000
220.000
240.000
260.000
280.000
0 10 20 30 40 50 60
Ou
tle
tHo
tFlu
idte
mp
[C
]
Timesteps [s]
OutletHotFluidTemp
42
Graph 12. Inlet cold fluid temperature for each timestep increasing linearly according with
the equation 𝟏𝟓 +𝟏
𝟑∗ (
𝒕
𝟏[𝒔])
Graph 13. Outlet hot fluid temperature for each timestep with the cold fluid temperature at
inlet increasing linearly according with the equation 𝟏𝟓 +𝟏
𝟑∗ (
𝒕
𝟏[𝒔])
We note that as in the previous case the outlet temperature of the mixture away from the
initial temperature condition during the first seconds reaching approximately its final value
and varying in subsequent timesteps around this value.
The results of the analysis above were exported to a single chart to more clearly compare the
temperature variation of output in both cases.
0.0000
5.0000
10.0000
15.0000
20.0000
25.0000
30.0000
35.0000
40.0000
0 10 20 30 40 50 60
Inle
tCo
ldFl
uid
tem
p[C
]
Timesteps [s]
InletColdFluidTemp
190.0000
200.0000
210.0000
220.0000
230.0000
240.0000
250.0000
260.0000
270.0000
0 10 20 30 40 50 60
Ou
tle
tHo
tFlu
idte
mp
[C
]
Timesteps {s}
OuletHotFluidTemp
43
The first time step was not taken into account to visualize better the results once the
temperature dropped from initial condition.
Graph 14. Comparison between the hot fluid temperature at outlet with a constant cold fluid temperature at inlet of 15℃ (in red in the graph) and the hot fluid temperature at outlet
with cold fluid temperature at inlet increasing linearly (in orange in the graph)
We note that roughly temperatures follow a similar pattern but in the first case the
temperature reaches values lower reaching almost 216℃ because the water kept at a constant
temperature allows the mixture to cool further.
These analyses can be of great importance to assess for example the consequences in the
event of an accident at the installation that supplies water to 15℃. If for some reason this
service is damaged and water undergoes a temperature rise have observed how would be
influenced in this case the hot fluid temperature at outlet.
Now we are going to do another kind of transient analysis oscillating the value of the water
temperature at inlet between 15ºC and 30ºC depends of the time according with the following
equation:
15 ∗ 𝑠𝑡𝑒𝑝 (60 −𝑡
1[𝑠]) + 30 ∗ 𝑠𝑡𝑒𝑝 (
𝑡
1[𝑠]− 60) ∗ 𝑠𝑡𝑒𝑝 (120 −
𝑡
1[𝑠]) + 15
∗ 𝑠𝑡𝑒𝑝 (𝑡
1[𝑠]− 120) ∗ 𝑠𝑡𝑒𝑝 (180 −
𝑡
1[𝑠]) + 30 ∗ 𝑠𝑡𝑒𝑝 (
𝑡
1[𝑠]− 180)
∗ 𝑠𝑡𝑒𝑝 (240 −𝑡
1[𝑠]) + 15 ∗ 𝑠𝑡𝑒𝑝 (
𝑡
1[𝑠]− 240) ∗ 𝑠𝑡𝑒𝑝 (300 −
𝑡
1[𝑠]) + 15
∗ 𝑠𝑡𝑒𝑝 (𝑡
1[𝑠]− 300) ∗ 𝑠𝑡𝑒𝑝 (60 −
𝑡
1[𝑠])
214
215
216
217
218
219
220
221
10 20 30 40 50 60
Ou
tle
tHo
tFlu
idte
mp
[C
]
Timesteps [s]
OuletHotFluidTemp
44
To perform this analysis 30 time steps were defined of 10 seconds each one cycling 300
seconds in total.
As a result the different values of the temperatures were obtained for each time step, we can
check them in the next table and charts.
Timesteps [s] InletColdFluidTemp; water [℃] OutletHotFluidTemp; naphtha-diesel [℃]
0 15 260
10 15 220,923
20 15 216,891
30 15 216,502
40 15 216,565
50 15 216,717
60 22,5 217,948
70 30 219,349
80 30 219,667
90 30 219,723
100 30 219,717
110 30 219,396
120 22,5 218,346
130 15 217,124
140 15 216,977
150 15 216,819
160 15 216,593
170 15 216,953
180 22,5 217,989
190 30 219,178
200 30 219,465
210 30 219,57
220 30 219,601
230 30 219,468
240 22,5 218,215
250 15 217,211
260 15 216,92
270 15 216,786
280 15 216,815
290 15 216,911
300 7,5 215,883
Table 13. Temperature values of hot fluid at outlet for each timestep with the cold fluid temperature varying step by step
45
Graph 15. Inlet cold fluid temperature at inlet for each timestep
Graph 16. Outlet hot fluid at outlet for each timestep
0
5
10
15
20
25
30
35
0
10
20
30
40
50
60
70
80
90
10
0
11
0
12
0
13
0
14
0
15
0
16
0
17
0
18
0
19
0
20
0
21
0
22
0
23
0
24
0
25
0
26
0
27
0
28
0
29
0
30
0
Inle
tCo
ldFl
uid
Tem
p;w
ate
r [C
]
Timesteps [s]
InletColdFluidTemp;water [C]
210
220
230
240
250
260
270
0
10
20
30
40
50
60
70
80
90
10
0
11
0
12
0
13
0
14
0
15
0
16
0
17
0
18
0
19
0
20
0
21
0
22
0
23
0
24
0
25
0
26
0
27
0
28
0
29
0
30
0
Ou
tle
tHo
tFlu
idTe
mp
; nap
hat
a-d
iese
l [C
]
Timesteps [s]
OutletHotFluidTemp;naphata-diesel [C]
46
Graph 17. Outlet hot fluid at outlet for each timestep
We found that with the variation interval of the cold fluid temperature at inlet the hot fluid
temperature at outlet varies in sinusoidal shape with an amplitude of about 3 ° C.
6.4.3. Environment influence
Other important aspect is to know how our heat exchanger will behave under the influence of
the environment temperature. We should notice that our heat exchanger must be work
correctly under different extremis conditions, the performance in the winter and summer
should be suitable.
We can see in the Fig. 21 below the heat transfer process that takes place on both sides of the
shell.
Fig. 21. Heat transfer process between the shell and the environment
213
214
215
216
217
218
219
220
221
222
10
20
30
40
50
60
70
80
90
10
0
11
0
12
0
13
0
14
0
15
0
16
0
17
0
18
0
19
0
20
0
21
0
22
0
23
0
24
0
25
0
26
0
27
0
28
0
29
0
30
0
Ou
tle
tHo
tFlu
idTe
mp
; n
aph
ata
-die
sel
[C]
Timesteps [s]
OutletHotFluidTemp;naphata-diesel [C]
47
To do this simulation we must calculate the heat transfer coefficient for the process, we must
take into account the phenomenon of convection and conduction according with the next
equation:
𝑈 =1
1ℎ1
+𝛿𝑘
+1
ℎ2
[𝑊 ∗ 𝑚2
𝐾]
ℎ1: transfer coefficient, its value is smaller compared with the other factors of the equation so
we can omit it.
𝛿, thickness = 0,008𝑚
k: steel resistance = 50 ÷ 60 𝑊
𝑚∗𝐾, we have taken an intermediate value,
55𝑊
𝑚∗𝐾
ℎ2: transfer coefficient= 35 ÷ 60 𝑊
𝑚2∗𝐾 , we have taken an intermediate value,
50𝑊
𝑚∗𝐾
Replacing the respective values in the equation above we obtain the heat transfer coefficient:
𝑈 =1
0,00855
+1
50
= 49,639 [𝑊 ∗ 𝑚2
𝐾]
Several analyses were performed in order to check the temperature variation at inlets and
outlets of our heat exchanger under different environment temperature. First a analysis were
performed defining an average environment temperature of 20℃ and then another two with
the lowest and highest environment temperature registered in Bulgaria, -38,5℃ and 45,2℃,
respectively.[11]
We can observe in the table that the temperatures at inlets and outlets don’t suffer practically
any change with the environment temperature.
7. Autodesk Inventor 3D model
Our definitive 3D model was built with Autodesk Inventor 3D CAD software, which is first and
foremost 3D parametric modelling software it has capabilities reaching far beyond the task of
creating 3D models[12].
Tªenvironment [℃]
InletColdFluidTª [℃]
InletHotFluidTª [℃]
OutletColdFluidTª [℃]
OuletHotFluidTª [℃]
20 15 260 22,532 201,126
-38,5 15 260 22,518 201,125
45,2 15 260 22,538 201,126
Table 14.Temperatures values at inlets and outlets for different environment temperatures
48
This software offers an easy-to-use set of tools for 3D mechanical design, documentation, and
product simulation. Digital Prototyping with Inventor helps you design and validate your
products before they are built to deliver better products, reduce development costs, and get
to market faster.
7.1. Components
The procedure to build a 3D model of our heat exchanger consists in first create all the
components that we need to then join all of them in an assembly model:
Remind that to create each component we have helped with the Bulgarian standards
aforementioned and attached in the ANNEX III.
Tubes
To create the tubes a circle with 25mm of diameter was made in a new sketch and it was
extruded with a depth of 3000mm. Then with the shell tool the material from the interior part
was removed creating a hollow cavity with 2mm of thick. (Fig.22)
Fig. 22: Tube (Autodesk Inventor)
Tubesheet
To build it a circle with 402mm of diameter and 20mm of thick was extruded, to make the
holes two rectangular patterns were created according with the pitch between tubes, and then
the unnecessary holes were supressed. A quarter was performed and after that with the
mirror feature all the holes were completed. We can check in the following figures the
procedure performed to create the rectangular pattern (Fig.23; Fig.24) and the final body of
the tubesheet (Fig.25)
49
Fig. 23: Rectangular pattern detail direction 1 Fig. 24: Rectangular pattern detail direction 2
Fig. 25: Tubesheet (Autodesk Inventor)
Channel cover
To build it the follow sketch (in blue Fig.26) was revolved and then the shell tool was applied to
remove the material getting a wall of 8mm thick. The large face was also removed with the
shell tool but the small one was necessary to do it in another way, making a hole, the shell tool
didn’t let us remove this face because of possible topology changes. (Fig.26)
50
Fig. 26: Channel cover (Autodesk Inventor)
Nozzles
Cylindrical part through which the fluids enter and leave. Its size and construction depends on
the others pipes of the installation, through which the fluid reaches. To build it first the follow
sketch (in blue in Fig.27) was revolved and then the interior hole was made. Finally with a
circular pattern the around holes were done, which serve to set the inlet or outlet pipe to the
nozzle with the respective bolts. (Fig.27)
Fig. 27: Tubes´ Nozzles (Autodesk Inventor)
51
To build the nozzles of the shell a small modification was made to avoid possible problems
when exporting the geometric model to Ansys Workbench. The bottom side that is in contact
with the shell was bended extruding a circumference with the diameter of the shell cutting the
underside of the nozzle thus guaranteeing the union at the same level without any gap.
(Fig.28)
Fig. 28. Shell´s nozzles (Autodesk Inventor)
Flange
They are used to join the channel cover with the shell of the heat exchanger. Also they allow
disarmament and removal or cleaning of internal parts. To make it the next sketch was
revolved and then the interior hole was performed. Finally with a circular pattern the around
holes for bolts were made. (Fig.29)
Fig. 29: Flange (Autodesk Inventor)
52
Seal
His function is seal the join between flanges. His construction is easy, just was needed to
extrude the follow profile. (Fig.30)
Fig. 30: Seal (Autodesk Inventor)
Supports
They are responsible for supporting the weight of the heat exchanger and allow it can be fixed
to a stable surface.
They should have the necessary height to allow the connection of the respective elbows joints
of the installation´s pipes with nozzles.
To build it first the tubular region which will hold the shell was created in the XY plane, to do
that the corresponding sketch of a circumference portion was extruded with its respective
thickness. Then a new plane with an offset respect to XZ plane was created, this offset
corresponds with the separation distance between the base and the tangent ZX Plane to the
tubular region. In this plane the base was created with all its details to secure the support to a
stable surface. The last step was to create in this same plane the two rectangular sketches to
extrude them till the tubular region and in that way get the walls forming all the structure.
(Fig.31)
53
Fig. 31: Support (Autodesk Inventor)
Baffles
Taking advantage of the tubesheet that we have previously built the baffle was built simply
cutting part of the piece and suppressing the holes. (Fig.32)
Fig. 32: Baffle (Autodesk Inventor)
54
7.2. Assembly
After creating all the necessary parts to build our heat exchanger we have proceeded to
assemble them.
We must create the suitable functional assembly relationships between all of elements using
the constraint and joint tools.
Assembly relationships are the glue and nails of construction when it comes to building your
assemblies. Properly using assembly relationships will permit the construction of stable
assemblies, assist in developing stack-up tolerances, and allow parts to be driven to show the
animation of a process.
We have started placing in the assembly model the tubesheet which was created previously. It
is recommended focus the tubesheet in the origin, the centre point corresponding with the
intersection of the main auxiliary planes, this will let us to work easier during all the process,
we can check this in the following figures (Fig.33;Fig.34)
Fig. 33: Tubesheet Fig. 34: Positioning of the tubesheet
Then we must insert the tubes in the tubesheet, to do that we have used the insert constraint
which is a combination of a face-to-face mate constraint between planar faces and a mate
constraint between the axes of the two components. An offset of 4mm was established from
the extreme face of the tubes with respect to the face of the tubesheet. First a quarter part
55
was performed and then with the assembly feature command “mirror “all the inserts were
completed. (Fig.35)
Then we must do the same in the other extreme of the tubes but in this case is enough with
only stablish the insert constraint between two tubes and the tubesheet. (Fig.36)
Fig. 35: Insertion of the tubes (Autodesk Inventor; assembly)
Fig. 36: Insertion of the tubes, both sides (Autodesk Inventor; assembly)
56
After that a new component was created to make the shell. Be aware that sometimes is better
to create the element during the assembly because you can help with the geometry of other
components which are already assembled as in this case.
To do that the geometry of the tubesheet was projected in a new plane at the beginning of the
tubes, 4mm before the tubesheet. It was extruded with the same length that the tubes,
3000mm. Then the shell tool was applied to hollow the solid body out and create the thin wall
feature of 8mm. (Fig.37; Fig.38)
Fig. 37: Geometry projected to extrude the shell (Autodesk Inventor; assembly)
57
Fig. 38: Shell (Autodesk Inventor; assembly)
We can see two holes in the shell, these holes were made later projecting the lower profile of
the nozzles so that they could be inserted into the shell.
Then the first flange was placed and a rotational joint between it and the shell was established,
rotational joint position a component in place and create one rotational degree of freedom.
(Fig.39)
Fig. 39: Rotational joint between the first flange and the shell (Autodesk Inventor; assembly)
Then the seal was placed in the flange with also a rotational relationship between them.
(Fig.40)
58
Fig. 40: Rotational joint between the seal and the first flange (Autodesk Inventor; assembly)
After that the other flange was placed with a rotational degree of freedom respect to the first
flange. (Fig.41)
Fig. 41: Rotational joint between flanges (Autodesk Inventor; assembly)
The same procedure was done in another side. (Fig.42)
59
Fig. 42. End view after placing the flanges in both sides (Autodesk Inventor; assembly)
In the next step the channel cover was placed with a rotational joint respect to the flange and
also with 19mm de gap respect to it. We can see it the following half view. (Fig.43)
Fig. 43: Detailed half view of the rotational joint between the channel cover and the second flange (Autodesk Inventor; assembly)
The same procedure must be followed to place the other channel cover in the other side.
60
Fig. 44. End view after placing the channel covers in both sides (Autodesk Inventor; assembly)
After that we must fix the flanges with the respective group of bolts, nuts and washers.
The diameter of each hole is 30mm so according with this we must take the corresponding bolt
with metric 30, the same for nuts and washers. We take them from the content center of
Autodesk, they are the following:
-Bolt: ISO 4017M30x100:1
-Nut: ISO 4035M30:1
-Washer: ISO 7092 ST 30-140 HV:1
The bolt was positioned applying two constraints, one mate constraint between the face of the
flange and the interior face of the head of the bolt and other one between the axe of the
corresponding hole and the axe of the bolt.
To position the washer two constraints were established: mate between one face of the
washer and the face of the flange and mate between the axe of the washer and the axe of the
hole.
Finally the nut was set with a mate constraint to join its face to the face of the washer and with
other mate between the axe of the nut and the axe of the bolt.
We can see all the groups of bolt, nut and washer in the next figure. (Fig.45)
61
Fig. 45: Detailed view of bolts, nuts and washers used to fix the flanges (Autodesk Inventor; assembly)
Only one bolt was assembled with its respective nut and washer and then with the help of
circular pattern tool the rest of them were placed, this tool does just what you’d expect it to: It
patterns a feature or set of features around an axis.
We must do the same operation twice for each joint in each side between the flanges.
The next step was to create the baffles. To place them mate constraints were established
between the axes of two holes of the baffle and two of the axes of the corresponding tubes,
two tubes are enough to position the baffle. Also another mate constraint was established
between the baffle and the face of the flange with an offset of 298mm to positioning the first
baffle.
To complete the rest of the first row of baffles the rectangular pattern tool was used, the
feature of the first baffle was selected and the straight edge of one tube was used to establish
the pattern direction, the distance between the baffles was defined as 420mm.
To do the second row we have preceded in the same way, first the relationship between the
baffle and the tubes was set and then the first baffle was placed with a mate constraint
respect the first baffle of the first row with an offset of 180mm.
In the following picture (Fig.46) the shell was hidden to see better the rows of baffles.
62
Fig. 46: Rows of baffles (Autodesk Inventor; assembly)
Our heat exchanger is almost finished we just have to assemble the nozzles and the supports.
To integrate the nozzle in the assembly model through which enter and leave the hot fluid we
have fixed them with a rotational joint between the nozzle and the channel cover with gap of
20mm. We must do the same in the other side. (Fig.47)
Fig. 47: Placing of tubes´ nozzles (Autodesk Inventor; assembly)
63
To place the nozzles of the shell first a plane with an offset from the flange of 200mm was
created. In the plane a new sketch and a new part were created to make a new axe. Then a
mate constraint between this axe and the axe of the nozzle was established. Last another mate
constraint between the face of the nozzle and the tangent plane of the shell was defined. We
must do the same for the other nozzle. (Fig.48)
Fig. 48: Placing of shell´s nozzles (Autodesk Inventor; assembly)
The last step is to place the supports of our heat exchanger. They were placed stabilising the
following relationships: mate constraint between the axe of the support and the axe of the
shell, angle de 0º between any three parallel planes and finally a mate constraint respect to
the flange with an offset of 610mm. The same procedure to fix the other support
Finally the whole structure was built, we can see it in the end section view and also in the half
view to see better the details (Fig. 49; Fig. 50)
64
Fig. 49: 3D model performed in Autodesk Inventor (end section view)
Fig. 50: 3D model performed in Autodesk Inventor (half view)
NOTE: although the unions would be made by welding to avoid problems while working with
Autodesk Inventor we have simple grounded in place all the components so that them cannot
move or rotate unintentionally. A grounded component is fully constrained and has 0 degrees
of freedom]
In ANNEX V attached at the end of this project we can see the drawings of each component
and also the drawing of the whole shell and tube heat exchanger with their respective
dimensions.
8. Thermal and structural analysis with Autodesk 3D model in Ansys Workbench
Once that the definitive 3D model of our heat exchanger was created in Autodesk Inventor it
was imported to Ansys Workbench, we could do it directly because Autodesk has an option to
open automatically Ansys Workbench with our geometry model or we could just import the
geometry file “ipt” from Ansys Workbench.
65
Several analyses were made with this model to ensure that our exchanger endure all loads to
which it was subjected and to ensure the viability of the structure to any possible incident.
First of all, three different geometrical solids were differentiated in the Desing Modeler of
Ansys, one for the metal part and the others for the corresponding fluids. These correspond
with three different domains.
To get the three bodies we have proceeded as follow.
A circle was created in the plane of the tubesheet with the corresponding diameter of the
shell. Then it was extruded in two asymmetric directions to cover all the heat exchanger and
finally a boolean operation was performed to subtract from this cylinder the rest of the solids.
Five bodies were obtained, the three aforementioned and other two which were unnecessary,
so they were suppressed.
Then the symmetry tool was applied in the middle plane of the structure to obtain only half
part of the heat exchanger to work easier and visualize better the results later.
8.1. Material properties
The next step was to assign the correspond material to each body. In the Engineering Data of
Ansys Workbench we could find different materials as the structural steel and the water liquid,
however for mixture of naphtha and diesel a new material with its corresponding properties
had to be created, these properties were already calculated and defined in the beginning .We
can check the different materials and their properties in the following tables. (Table.13;
Table.14; Table.15)
66
Table. 15: Structural Steel properties; Engineering Data of Ansys Workbench
Table 16: Water liquid properties; Engineering Data of Ansys Workbench
67
Table 17: Naphtha-diesel properties; Engineering Data of Ansys Workbench
8.2. Meshing
After that we must create the mesh.
Meshing is the process in which your geometry is spatially discretized into elements and
nodes. This mesh along with material properties is used to mathematically represent the
stiffness and mass distribution of our structure.
Our model was automatically meshed at solve time. The default element size is determined
based on a number of factors including the overall model size, the proximity of other
topologies, body curvature, and the complexity of the feature. If necessary, the fineness of the
mesh is adjusted up to four times (eight times for an assembly) to achieve a successful mesh.
As a result finally we got the next mesh with 1908917 nodes and 1002355 elements Fig.51.
Fig. 51: Mesh for steady thermal analysis and static structural analysis
68
8.3. Steady state thermal analysis
On the one hand, the structure will be affected by various thermal stresses due to
temperature changes. By the nature of its operation the shell will be at a temperature
significantly different from the tubes, so expansions or contractions will be different, resulting
in the occurrence of stresses in both components, transmitted through the tubesheet. The
effects of thermal stress vary according to circumstances.
The fixed tubesheet exchanger is especially vulnerable to this condition because there is no
way to deal with this difference in expansions.
Therefore, a thermodynamic analysis was performed to study the behaviour of our heat
exchanger under the influence of these changes in temperature and to calculate the
thermodynamic load to which the structure will be subjected.
For a steady-state (static) thermal analysis in Mechanical, the temperatures {T} are solved for
in the matrix below:
[𝐾(𝑇){𝑇}] = {𝑄(𝑇)}
Assumptions:
No transient effects are considered in a steady state analysis
[K] can be constant or a function of temperature
{Q} can be constant or a function of temperature
Fixed temperatures represent constraints {T} on the system (like fixed displacements
on structures).
Shells: temperatures may vary over the surface (no through-thickness temperature
variation)
The only required material property for steady state is thermal conductivity, which is input in
the Engineering Data application. The value of this property for each material is in the tables
above.
As with structural analyses, contact regions are automatically created to enable heat transfer
between parts in assemblies.
By default, perfect thermal contact is assumed, meaning no temperature drop occurs at the
interface.
69
Two different analyses were performed, in the first one only the temperatures in the inlets and
outlets were stablished as boundary conditions as we can see in the next figure (Fig.52)
Fig. 52: Steady state thermal analysis_1; boundary conditions
As a result after solving this first steady state analysis the temperature distribution in the heat
exchanger was obtained (Fig.53)
Fig. 53: Steady state thermal analysis_1; Temperature distribution
We have observed as expected that in areas close to the nozzles the temperatures are
approaching to the boundary conditions while the rest of the structure is at temperature
around 100℃.
70
To perform the second analysis another boundary condition was added, the convective heat
transfer between the shell and the environment, which assume that is at average temperature
of 22 °C (Fig.54)
Fig. 54: Steady state thermal analysis_2; boundary conditions
The second analysis was solved, we can see the result in the next picture (Fig.55):
Fig. 55: Steady state thermal analysis_2; Temperature distribution
In this case was noticed that most of the structure reaches a temperature close to the
environment temperature as we could anticipate although in areas close to the nozzles which
provide the input and output of naphtha-diesel mix temperatures are higher approaching to
the boundary conditions. Presumably, after an infinite time the structure would reach the
environment temperature.
71
NOTE: However, we must remember that this is a steady state analysis which gives us a rough
idea of the thermodynamic behaviour of the structure but to make a more specific and closer
to reality study we should perform a dynamic analysis to take into account changes in currents
and temperature with the time.
8.4. Static structural analysis
To check the influence in our structure of the temperature and pressure we must perform a
static structural analysis, this analysis is also a first necessary step to perform other analysis
later.
For a linear static structural analysis, the global displacement vector {x} is solved for in the
matrix equation below:
[𝐾]{𝑥} = {𝐹}
Assumptions made for linear static structural analysis are:
[K] , which is the global stiffness matrix, is constant
– Linear elastic material behaviour is assumed
– Small deflection theory is used
{F} , which is the global load vector, is statically applied
– No time-varying forces are considered
– No damping effects
In structural analyses, all types of bodies supported by Mechanical may be used.
Young’s Modulus and Poisson’s Ratio are always required for linear static structural analyses.
-Density is required if any inertial loads are present.
-Thermal expansion coefficient is required if a temperature load is applied.
-Stress Limits are needed if a Stress Tool result is present.
- Fatigue Properties are needed if Fatigue Tool result is present.
72
8.4.1. Temperature load
After calculating the thermodynamic load to which the structure is subjected by a structural
analysis we will see the influence of that load on the structure considering parameters as
strain or stress.
A new static structural analysis was created, in which the loads and secondly the “supports” or
boundary conditions were established.
Loads and supports respond in terms of the degrees of freedom (DOF) available for the
elements used.
-Loads:
Imported Load Body temperature: thermodynamic load calculated in the previous
steady state thermal analysis.
-Supports:
Fixed support: the undersides of the two brackets which hold the structure were
established as fixed supports restricting all the degrees of freedom.
Displacement: as we are working only with half of the model we must restrict the
movement of the structure in the median plane in the Y axis direction.
Displacement2: in the tubes´ nozzles the movement in the X axis direction was
restricted.
Displacement3: for the nozzles of the shell in this case the movement along the axis Z
was restricted.
We can see these loads and supports in the figure below (Fig. 56):
73
Fig. 56: Static structural analysis; thermodynamic load; loads and supports
After that the analysis was run. To evaluate the effects of this thermodynamic load on the
structure we check different magnitudes as the total deformation, the equivalent elastic strain
or the equivalent stress. The results are the following. (Fig.57; Fig.58; Fig.59)
Fig. 57: Static structural analysis; thermodynamic load; Total deformation
In the previous image we can see the deformation in mm experienced by the structure with
the temperature. The deformation is symmetric relative to the Z axis and does not reach large
values although it is going to rise to the sides of the structure peaking at approximately 3 mm
maximum at the edge of the nozzle which gives access to the naphtha-diesel mixture where
the highest temperature is.
74
Fig. 58: Static structural analysis; thermodynamic load; equivalent elastic strain
With regard to the equivalent elastic strain we note that most of the structure hardly suffers
stress due to the thermodynamic load although again in areas near the nozzles of the tubes it
is where we find small stresses because the temperature in these zones is higher, in any case
they are very small values.
Fig. 59: Static structural analysis; thermodynamic load; equivalent stress
The equivalent stress we see that also occurs symmetrically relative to the axis Z. The
maximum stress is about 514MPa and is located in the central part of the shell, in the supports
and around the joint edges between the nozzles and the shell.
75
8.4.2. Pressure load
Furthermore besides the temperature difference, there are other sources of mechanical stress.
Some are the result of the construction methods, for example, the stresses in the tube and the
tubesheet produced when tubes are rolling or welding. To these are added the processes
caused by the behaviour of currents, especially during the operation.
To protect exchangers of the permanent deformation and fatigue is necessary to make a
design for these conditions and to ensure not exceeding the values of design stresses.
To evaluate these effects, another static structural analysis was performed to see the
behaviour of the structure under different pressure loads to which it is subject.
The following loads and supports were established:
-Loads:
Pressure: pressure load of 5Mpa acting in the normal direction to the shell and
nozzles. This value was taken from the results of the calculation by the software of
Webbusterz Engineering.
Hydrostatic Pressure: hydrostatic pressure due to fluids
Supports:
Fixed support: the undersides of the two brackets which hold the structure were
established as fixed supports restricting all the degrees of freedom.
Displacement: as we are working only with half of the model we must restrict the
movement in the Y axis of the structure in the median plane.
We can see the loads and support established in the following figure (Fig.60)
76
Fig. 60: Static structural analysis; pressure load; loads and supports
Once these boundary conditions were stablished the analysis was solved. As a result we have
checked the total deformation and the equivalent stress of the structure under the influence
of pressure load (Fig.61; Fig. 62;Fig.63)
Fig. 61: Static structural analysis; pressure load; equivalent stress
If we look at the results of the equivalent stress we note that due to the pressure load a
bulging of the structure occurs which mostly suffers tensions between 8 and 12 Mpa reaching
a critical point of maximum stress (37,3MPa) at the junction of the nozzle to the shell. It could
be minimized making a rounding. In the next figure (Fig.61) we can see this critical point in a
detailed view.
77
Fig. 62: Static structural analysis; pressure load; detailed view of equivalent stress
.
Fig. 63: Static structural analysis; pressure load; total deformation
Based on the results of the total deformation of the structure, sum of the different
deformations in each axis, we can observe a slight instability which carries a maximum
deformation in the right nozzle. Anyway they are very small deformation values.
8.4.3. Temperature and pressure load
Finally a structural analysis was performed to assess the effect of the simultaneous action of
the thermodynamic and pressure loads.
In a new static structural analysis the following loads and supports were established.
78
Loads:
-Pressure: pressure load of 5Mpa acting in the normal direction to the shell and nozzles. This
value was taken from the results of the calculation by the software of Webbusterz Engineering.
-Hydrostatic pressure: hydrostatic pressure due to fluids.
-Imported Load temperature
-Supports:
-Fixed support: the undersides of the two brackets which hold the structure were established
as fixed supports restricting all the degrees of freedom.
-Displacement: as we are working only with half of the model we must restrict the movement
in the Y axis of the structure in the median plane.
Fig. 64: Static structural analysis; temperature and pressure load; loads and supports
We have solved it and we have exanimated different parameters as the total deformation, the
equivalent elastic strain and the directional deformation with respect to the different axes
(Fig.65; Fig.66; Fig.67; Fig.68; Fig.69)
79
Fig. 65: Static structural analysis; temperature and pressure load; total deformation
We observe a nearly symmetrical deformation around the Z axis with a maximum value of 2.3
mm in the zone of maximum temperature in the nozzle through which the hot fluid enters.
Fig. 66: Static structural analysis; temperature and pressure load; equivalent elastic strain
If we look at the equivalent stress we can conclude that it is virtually null.
Let’s check the directional deformation:
80
Fig. 67: Static structural analysis; temperature and pressure load; directional deformation Z
axis
Fig. 68: Static structural analysis; temperature and pressure load; directional deformation X
axis
Fig. 69: Static structural analysis; temperature and pressure load; Directional deformation Y
axis
81
In all three cases we observe a symmetrical deformation with respect to the corresponding
axes reaching the highest value of 2 mm deformation in the axis X direction at the left nozzle.
On the other hand we see that the structure suffers less deformation relative to the axis Y.
8.5. Linear buckling
As the most of the structures our heat exchanger requires an evaluation of its structural
stability.
At the onset of instability, buckling, a structure will have a very large change in displacement
under essentially no change in the load (beyond a small load perturbation)
Linear buckling analysis predicts the theoretical buckling strength of an ideal linear elastic
structure.
Imperfections and nonlinear behaviours prevent most real world structures from achieving
their theoretical elastic buckling strength.
Linear buckling generally yields unconservative results by not accounting for these effects.
Although unconservative, linear buckling has the advantage of being computationally cheap
compared to nonlinear buckling solutions.
In this case to verify the stability of our structure a linear buckling analysis was simulated in
order to check whether the supports give in to the weight of the exchanger.
We must pay special attention to the thin walls forming the support since beforehand is to
presuppose that will be our critical point.
For the purpose, before performing this buckling analysis a static structural analysis was
needed. The support´s geometry file “ipt” from Autodesk Inventor was imported and the mesh
was created automatically. We can see it in the next figure (Fig.70)
82
Fig. 70: support´s mesh
In this case because the geometry is not so complex the process was easier. The mesh has
9193 nodes and 4424 elements.
After meshing process the force acting on each support was calculated. Without any extra
load, the only strength was the sheer weight of the structure, expected to be evenly
distributed between the two supports.
The mass of the exchanger was calculated as the sum of the different bodies that make up it:
-Steel metal part: mass=517,64 kg
-Shell: mass of water=102,04 kg
-Tubes: mass of naphtha-diesel= 66,294 kg
Once that value of the different masses is known the weight of the exchanger and in
consequence the force that have to endure each support was calculated as follow:
𝐹 =(517,64 + 102,04 + 66,294)[𝑘𝑔] ∗ 9,81 [
𝑚𝑠2]
2= 3364,70247[𝑁]
This force was established in the static structural analysis, it was applied in the tubular bases of
the supports in contact with the shell, towards the negative direction of the Y axis.
83
Also a fixed support was established in the lower face of the bracket which will be fixed to the
surface at which the exchanger will be set. (Fig.71)
Fig.71: Linear buckling; loads and supports
For material properties, Young’s Modulus and Poisson’s Ratio are required as a minimum. We
can check the values of these properties in the tables above.
Then the linear buckling analysis was created and solved.
In a buckling analysis all applied loads (F) are scaled by a multiplication factor (λ) until the
critical (buckling) load is reached:
𝐹 ∗ 𝜆 = 𝑏𝑢𝑐𝑘𝑙𝑖𝑛𝑔 𝑙𝑜𝑎𝑑
The “Solution Information” branch provides detailed solution output.
As a result, the different deformations of each mode and the corresponding load multiplier
were obtained.
84
Fig. 72: Buckling mode1; Total deformation
Fig. 73: Buckling mode2; Total deformation
85
Fig.74: Buckling mode3; Total deformation
Fig. 75: Buckling mode4; Total deformation
86
Fig. 76: Buckling mode5; Total deformation
Fig. 77: Buckling mode6; Total deformation
87
The Load Multiplier for each buckling mode is shown in the Details view as well as the graph
and chart areas. The load multiplier times the applied loads represent the predicted buckling
load.
𝐹𝑏𝑢𝑐𝑘𝑙𝑖𝑛𝑔 = (𝐹𝑎𝑝𝑝𝑙𝑖𝑒𝑑 ∗ 𝜆)
Graph 18: Linear buckling; modes and frequencies
Table 18: Linear buckling; list of modes and frequencies
The multiplication factor is much higher than one which verified that our structure will not
yield buckling.
If we test and apply a force of 1N we see that the multiplication factor is equal to 1.32 or what
is the same force that will produce buckling will be 1,32MN.
88
A more accurate approach to predicting instability is to perform a nonlinear buckling analysis.
This involves a static structural analysis with large deflection effects turned on. A gradually
increasing load is applied in this analysis to seek the load level at which your structure
becomes unstable. Using the nonlinear technique, your model can include features such as
initial imperfections, plastic behaviour, gaps, and large-deflection response. In addition, using
deflection-controlled loading, you can even track the post-buckled performance of your
structure (which can be useful in cases where the structure buckles into a stable configuration,
such as "snap-through" buckling of a shallow dome).
8.6. Modal analysis
Another big problem in the shell and tube heat exchangers are the vibrations induced by the
flow. The tubes may vibrate and be forced against the baffles, and even crash into other tubes,
which can cause severe deformation and wear. The continuous flexing can cause fatigue.
Most of these vibrations come from the vortices, formed due to the flow conditions. They are
usually small, but very numerous, and with very high frequencies, worsen this condition at
higher fluid velocities.
The damage caused by the tube vibration has become a growing phenomenon when the
dimensions of the heat exchangers and flow quantities are increased. Among these causes are:
-Vortex shedding: shedding frequency of the fluid in systems cross flow over the tubes may
coincide with a natural frequency of the tubes and cause resonant vibrations over a wide
range.
-Flexible coupling fluid: the fluid flowing over the tubes causes them shaped vibration swirling
motion. The elastic coupling mechanism occurs when the speed "review" is exceeded and is
self-exciting vibration and grows in amplitude.
This mechanism occurs very frequently in process heat exchangers that have been damaged by
vibration.
-Pressure fluctuation: Pressure fluctuations due to turbulence developed in the body of a
cylinder, or those who come to it from the current to enter the system may cause a potential
mechanism vibration of the tubes. The tubes correspond to the portion of the spectrum close
to its natural frequency energy.
89
-Acoustic coupling: acoustic coupling or resonance develops when standing waves of the fluid
side of the shell are in phase with the shedding vortex tubes. The standing waves are
perpendicular to the axes of the tubes and the cross flow direction. Only occasionally tubes are
damaged; however, the noise caused by this can be very annoying.
To evaluate these different influences which may cause unwanted vibrations in our structure
we should do specific analysis isolating various parts of the exchanger, for example studying
the flow around a single tube.
Due to the complicity these analyses require we will not do them but a modal analysis was
performed in order to obtain the main natural frequencies of the structure and vibrational
modes associated, in this way the natural critical system frequencies that could produce
resonance are known.
For a free vibration analysis, the natural circular frequencies 𝑤𝑖 and mode shapes 𝑓𝑖 are
calculated from:
([𝐾] − 𝑤2 ∗ [𝑀]){∅𝑖} = 0
Assumptions:
- [K] and [M] are constant:
- Linear elastic material behaviour is assumed
- Small deflection theory is used, and no nonlinearities included
- [C] is not present, so damping is not included
- {F} is not present, so no excitation of the structure is assumed
- The structure can be constrained or unconstrained
- Mode shapes {f} are relative values, not absolute
Modal analysis can employ any type of geometry
The critical requirement is to define stiffness as well as mass in some form. Stiffness may be
specified using isotropic and orthotropic elastic material models (for example, Young's
modulus and Poisson's ratio).Mass may derive from material density or from remote masses.
Structural and thermal loads are not available in free vibration.
Contact regions are available in free vibration analyses however contact behaviour will differ
for the nonlinear contact types.
All contact will behave as bonded or no separation in a modal analysis.
Before performing the modal analysis we must create a static structural analysis with the
corresponding geometry, in this case for obvious reasons due to the characteristics of the
modal analysis we must use the entire body of our heat exchanger which was built previously
90
in Autodesk Inventor. After that the mesh was created automatically, in this case because the
geometry is more complex as a result we have not obtained a mesh of high quality but anyway
suitable for the analysis. We can see in detail the mesh in the next figure (Fig.78)
Fig. 78: Modal analysis; mesh
In the static structural analysis the load of the standard earth gravity and also the fixed
supports for the brackets were defined.
Then the modal analysis was added and solved, notice that in this case we are working with
the whole of the structure and not with only half part to reach correct results.
After solving, a list of the six first natural frequencies of the structure and a representation of
the deformed modes associated with each vibration frequency were obtained.
91
Fig. 79: Mode 1; Total deformation
Fig. 80: Mode 2; Total deformation
92
Fig. 81: Mode 3; Total deformation
Fig. 82: Mode 4; Total deformation
93
Fig. 83: Mode 5; Total deformation
Fig. 84: Mode 6; Total deformation
94
Graph19:Modal analysis;modes and frequencies
Table 19: Modal analysis; list of modes and frequencies
Because there is no excitation applied to the structure the mode shapes are relative values not
actual ones.
Mode shape results are mass normalized.
The same is true for other results (stress, strain, etc.)
Because a modal result is based on the model’s properties and not a particular input, we can
interpret where the maximum or minimum results will occur for a particular mode shape but
not the actual value.
8.7. Earthquake analysis
As a final test another analysis was performed in order to check the response of our structure
in the event of a possible earthquake.
Earthquake analyses can be performed by applying different procedures.
95
The most popular procedure is the Response Spectrum analysis (RS-analysis). The RS-analysis is
cheap to use in terms of numerical costs as it is based on modal results. However, the
spectrum solution can only show positive results, i.e. positive stresses and strains, as it only
records the maximum amplitudes for each mode and the superposition of these results in turn
will give the positive results.
Another procedure is to perform a full transient analysis of the earthquake. Such analyses are
computational expensive. However, they will give results based on the dynamic equation of
equilibrium and hence both positive (tensile) and negative (compressive) stress results will be
reported for the full length of the earthquake.
We have opted for the first option for being computationally easier and provide a good
approximation.
There are two steps in running a response spectrum analysis in ANSYS.
First we need to run a modal analysis which will give use the modes/eigenvalues of the
structure, we already did it, we can see the results above. Secondly we run the Response
Spectrum analysis which does the following:
-Calculates the participation factor for each of the structures frequencies.
-Find the maximum accelerations from the given Response Spectre (for each mode)
-Scales the modal displacements found in the modal analysis to physical mode shapes based
on acceleration, participation factors and circular frequencies.
-Finally superpose these modal results to the final result using i.e. the SRSS method.
Step 1 - Modal analysis: The Modal analysis gave us the eigenfrequencies of the structure. We
are going to run a response spectrum analysis where the ground acceleration is applied in the
y-direction. Hence, we need to make sure that the effective mass in the y-direction is higher
than 90 % of the total mass as most codes use this as a requirement for the RS analysis. RS-
analysis we will use the first 6 modes as input.
Step 2: RS analysis: The RS-analysis uses the modal results obtained above as input for
calculation of the earthquake response. To include the response spectre data containing the
relation between structural acceleration and the structures frequencies we insert the tool ‘RS
Acceleration’ and include earthquake data as a table (frequency [Hz] vs. acceleration [𝑚/
𝑠^2 ]).
96
To calculate this earthquake data and gather the information necessary to perform the
analysis we have helped with the next European Standard: “Eurocode 8: Design of structures
for earthquake resistance -Part 1: General rules, seismic actions and rules for buildings”[13]
The first aspect to consider is to distinguish the different types of ground in which it could be
installed our heat exchanger. (Table 20)
- (1) Ground types A, B, C, D, and E, described by the stratigraphic profiles and parameters
given in Table 20 and described hereafter, may be used to account for the influence of local
ground conditions on the seismic action. This may also be done by additionally taking into
account the influence of deep geology on the seismic action.
NOTE: The ground classification scheme accounting for deep geology for use in a country may
be specified in its National Annex, including the values of the parameters S, 𝑇𝐵, 𝑇𝐶 and 𝑇𝐷
defining the horizontal and vertical elastic response spectra.
Table 20: Types of ground
97
- (2) The site should be classified according to the value of the average shear wave velocity,
𝑣𝑆_30, if this is available. Otherwise the value of 𝑁𝑆𝑃𝑇 should be used.
- (3) The average shear wave velocity 𝑣𝑆_30 should be computed in accordance with the
following expression:
𝑣𝑆_30 =30
∑ℎ𝑖𝑣𝑖
𝑖=1,𝑁
where ℎ𝑖 and 𝑣𝑖 denote the thickness (in metres) and shear-wave velocity (at a shear strain
level of 10−5or less) of the i-th formation or layer, in a total of N, existing in the top 30cm.
(4)P For sites with ground conditions matching either one of the two special ground types 𝑆1
or𝑆2, special studies for the definition of the seismic action are required. For these types, and
particularly for 𝑆2, the possibility of soil failure under the seismic action shall be taken into
account.
NOTE: Special attention should be paid if the deposit is of ground type 𝑆1. Such soils typically
have very low values of 𝑣𝑠low internal damping and an abnormally extended range of linear
behaviour and can therefore produce anomalous seismic site amplification and soil-structure
interaction effects (see EN 1998-5:2004, Section 6). In this case, a study to define the seismic
action should be carried out, in order to establish the dependence of the response spectrum
on the thickness and 𝑣𝑠, value of the soft clay/silt layer and on the stiffness contrast between
this layer and the underlying materials.
Only the horizontal response of our shell and tube heat exchanger in the event of an
earthquake was analysed, so that the earthquake affect the structure vertically the epicentre
would have to occur just at the surface where the exchanger is anchored, which would be
highly unlikely.
According with the Eurocode to calculate the desing spectrum for elastic analysis we must bear
in mind the following:
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1-The capacity of structural systems to resist seismic actions in the non-linear range generally
permits their design for resistance to seismic forces smaller than those corresponding to a
linear elastic response.
2-To avoid explicit inelastic structural analysis in design, the capacity of the structure to
dissipate energy, through mainly ductile behaviour of its elements and/or other mechanisms,
is taken into account by performing an elastic analysis based on a response spectrum reduced
with respect to the elastic one, henceforth called a "design spectrum". This reduction is
accomplished by introducing the behaviour factor q.
3- The behaviour factor q is an approximation of the ratio of the seismic forces that the
structure would experience if its response was completely elastic with 50/0 viscous damping,
to the seismic forces that may be used in the design, with a conventional elastic analysis
model, still ensuring a satisfactory response of the structure. The values of the behaviour
factor q, which also account for the influence of the viscous damping being different from 5%,
are given for various materials and structural systems according to the relevant ductility
classes in the various Parts of EN 1998. The value of the behaviour factor q may be different in
different horizontal directions of the structure, although the ductility classification shall be the
same in all directions.
(4)P For the horizontal components of the seismic action the design spectrum, 𝑆𝑑(𝑇),
shall be defined by the following expressions:
0 ≤ 𝑇 ≤ 𝑇𝐵: 𝑆𝑑(𝑇) = 𝑎𝑔 ∗ 𝑆 ∗ [2
3+
𝑇
𝑇𝐵∗ (
2.5
𝑞−
2
3)]
𝑇𝐵 ≤ 𝑇 ≤ 𝑇𝐶 : 𝑆𝑑(𝑇) = 𝑎𝑔 ∗ 𝑆 ∗2.5
𝑞
𝑇𝐶 ≤ 𝑇 ≤ 𝑇𝐷: 𝑆𝑑(𝑇) {= 𝑎𝑔 ∗ 𝑆 ∗
2.5
𝑞∗ [
𝑇𝐶
𝑇]
≥ 𝛽 ∗ 𝑎𝑔
𝑇𝐷 ≤ 𝑇: 𝑆𝑑(𝑇) {= 𝑎𝑔 ∗ 𝑆 ∗
2.5
𝑞∗ [
𝑇𝐶∗𝑇𝐷
𝑇2 ]
≥ 𝛽 ∗ 𝑎𝑔
Where
99
T: is the vibration period of a linear single-degree-of-freedom system
𝑎𝑔: is the design ground acceleration on type A ground (𝑎𝑔 = 𝛾1 ∗ 𝑎𝑔𝑅)
S: is the soil factor;
𝑇𝐵: is the lower limit of the period of the constant spectral acceleration branch
𝑇𝐶 : is the upper limit of the period of the constant spectral acceleration branch;
𝑇𝐷: is the value defining the beginning of the constant displacement response range
of the spectrum;
q: is the behaviour factor;
b: is the lower bound factor for the horizontal design spectrum.
𝛾1=importance factor
𝑎𝑔𝑅 =reference peak ground acceleration on type A ground
NOTE: The value to be ascribed to b for use in a country can be found in its National Annex.
The recommended value for b is 0,2.
(2)P The values of the period, 𝑇𝐵and 𝑇𝐷and of the soil factor S describing the shape of the
elastic response spectrum depend upon the ground type.
NOTE: The values to be ascribed to 𝑇𝐵, 𝑇𝐶 , 𝑇𝐷and S for each ground type and type (shape) of
spectrum to be used in a country may be found in its National Annex. If deep geology is not
accounted, the recommended choice is the use of two types of spectra: Type 1 and Type 2. If
the earthquakes that contribute most to the seismic hazard defined for the site for the
purpose of probabilistic hazard assessment have a surface-wave magnitude, Ms, not greater
than 5,5, it is recommended that the Type 2 spectrum is adopted. For the five ground types A,
B, C, D and E the recommended values of the parameters S, 𝑇𝐵, 𝑇𝐶 and 𝑇𝐷 are in Table 19 for
the Type 1 Spectrum and in Table 20 for the Type 2 Spectrum. Graoh 15 and Graph 16 show
the shapes of the recommended Type 1 and Type 2 spectra, respectively, normalised by 𝑎𝑔 for
5% damping. Different spectra may be defined in the National Annex, if deep geology is
accounted for.
100
Table 19: Values of the parameters describing the recommended Type I elastic response
spectra
Table 20: Values of the parameters describing the recommended Type 2 elastic response
spectra
Graph 20: Recommended Type 1 elastic response spectra for ground types A to E
(5% damping)
101
Graph 21: Recommended Type 2 elastic response spectra for ground types A to E
(5(% damping)
In our case as our heat exchanger will be installed in Bulgaria so if we check in the national
stantards we see that we must adopt the Type1.
Knowing the values of the above parameters and applying the corresponding formulas the
different earthquake data (frequency [Hz] vs. acceleration [𝑚/𝑠^2 ]) to each kind of ground
were calculated. We can check this data in the excel sheet attached in ANNEX IV and the
corresponding chart in the next Graph 17.
Graph 22: Data earthquake; frequency [Hz] vs. acceleration [𝒎/𝒔^𝟐]
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Once the earthquake data was introduced two analyses were made for each kind of ground,
one in the X axis direction and another in the Y-axis direction.
The RS Espectrum analysis was solved for both directions and different results as equivalent
strain and stress were examined .Then we compare the results in both directions to study what
is the best option for positioning / orienting our exchanger.
Ground type A
Direction: X axis
Fig. 85: Earthquake analysis; ground type A; X axis direction; equivalent stress
Fig.86: Earthquake analysis; ground type A; X axis direction; directional deformation
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Direction: Y axis
Fig.87: Earthquake analysis; ground type A; Y axis direction; equivalent stress
Fig. 88: Earthquake analysis; ground type A; Y axis direction; directional deformation
Ground type B
Direction: X axis
Fig.89: Earthquake analysis; ground type B; X axis direction; equivalent stress
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Fig.90: Earthquake analysis; ground type B; X axis direction; directional deformation
Direction: Y axis
Fig.91: Earthquake analysis; ground type B; Y axis direction; equivalent stress
Fig.92: Earthquake analysis; ground type B; Y axis direction; directional deformation
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Ground type C
Direction: X axis
Fig. 93: Earthquake analysis; ground type C; X axis direction; equivalent stress
Fig.94: Earthquake analysis; ground type C; X axis direction; directional deformation
Direction: Y axis
Fig.95: Earthquake analysis; ground type C; Y axis direction; Equivalent stress
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Fig.96: Earthquake analysis; ground type C; Y axis direction; directional deformation
Ground type D
Direction: X axis
Fig.97: Earthquake analysis; ground type D; X axis direction; equivalent stress
Fig.98: Earthquake analysis; ground type D; X axis direction; Directional deformation
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Direction: Y axis
Fig.99: Earthquake analysis; ground type D; Y axis direction; Equivalent stress
Fig.100: Earthquake analysis; ground type D; Y axis direction; Directional deformation
Ground type E
Direction: X axis
Fig.101: Earthquake analysis; ground type E; X axis direction; Equivalent stress
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Fig.102: Earthquake analysis; ground type E; X axis direction; Directional deformation
Direction: Y axis
Fig.103: Earthquake analysis; ground type E; Y axis direction; Equivalent stress
Fig.104: Earthquake analysis; ground type E; Y axis direction; Directional deformation
After considering all the results we observe that in both directions are very small deformations
and the stress are more or less similar in no case exceeding the tensile strength of the
material. Whereupon we cannot anticipate what would be the best orientation for our heat
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exchanger, in both cases the viability and stability of the structure would ensure. However,
when our exchanger will be positioned in a predetermined area it would require a more
detailed study of the particular case following the guidelines of the Bulgarian national
standards.
CONCLUSION
After studying the results of all the analyses to which it has been subjected to our heat
exchanger we can largely ensure their proper functioning in the terms and conditions outlined
above.
However, it should be note that there has been an approximate baseline study, that should not
be taken as definitive to do a real project, it is advisable following the guidelines of this project
perform more in-depth analysis such as fluid dynamic analysis to study the potential vortices
or losses that may suffer, transient analysis and nonlinear structural analysis although more
complex help better understand the behaviour of the exchanger.
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REFERENCES
[1] TEMA STANDARDS
[2] SLAVOV V., SECTIONAL DRAWING OF THE HEAT EXCHANGER, NOTES, UCTM-SOFIA, 2014
[3] Webbusterz Engineering Software
[4] http://www.engineeringtoolbox.com/ (11/2014)
[5] http://www.husltd.com/en/products/41873-seamless-pipes (11/2014)
[6] Russian standards GOST 15118-79, GOST-15120-79 and 15122-79.
[7] BDS_EN_ISO_16812
[8] ANSYS CFX – Solver Modelling Guide release 15.0, ANSYS, Inc., November 2013
[9] DIN 28013 for semi ellipsoidal head
[10] ANSYS CFX – Solver Theory Guide release 12.1, ANSYS, Inc., November 2009
[11] http://inews.bg/ (6/2015)
[12] Mastering Autodesk Inventor 2014
[13] Eurocode 8: Design of structures for earthquake resistance -Part 1: General rules,
seismic actions and rules for buildings
- Chemical Engineering Design, Volume 6,R. K. Sinnot
-Perry´s chemical Engineering
-THOME J. R., ENGINEERING DATA BOOK. WIELAND-WERKE, GERMANY, 2015
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ANNEX I Weebbusterz engineering software results
SHELL & TUBE HEAT EXCHANGER DATA SHEET
Company: Company: UCTM-Sofia Project: Project Name: Design-Shell-tube
Heat Exchanger
Description: Exchanger: Fixed-tube plate
Engineer: Brais Item Tag:
Revision: Date: 11/03/2011
FLUID PROPERTIES
Fluid Name: Water Naphta-Diesel
Allocation: Shell Side Tube Side
Mass Flow: 500000 kg/hr 200000 kg/hr
Velocity: 8.43E+00 m/s 2.80E+00 m/s
IN OUT IN OUT
Operating Temperature:
15 'C 18.72 'C 260 'C 240 'C
Density: 999 kg/m3 780.625 kg/m3
Viscosity: 0.001002 kg/ms 0.000928 kg/ms
Specific Heat: 4182 J/kg'C 1943.75 J/kg'C
Thermal Conductivity: 0.602 W/m'C 0.129 W/m'C
Latent Heat: 0 kJ/kg 0 kJ/kg
Molecular Weight:
Allowable Calculated Allowable Calculated
Pressure Drop: 1 Bar 3.274925505
Bar 1 Bar 0.205916314 Bar
Fouling Factor: 0.00018 m2'C/W 0.00035 m2'C/W
DESIGN DATA
Design Pressure: 0 N/mm2 0 N/mm2
Working Pressure: N/mm2 N/mm2
Design Temperature: 0 'C 0 'C
Working Temperature: 16 'C 250 'C
Material of Construction:
Carbon Steel Steel
THERMAL DESIGN
Heat Duty: 2,160.70 kW
Heat Transfer Coef. 67,113.18 W/m2'C 2,035.34 W/m2'C
LMTD Corrected: 233.0452 'C
Overdesign Factor to be Applied: None
Overall Ht.Transfer Coefficient - Design: 744.3596 Clean: 1975.4311
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W/m2'C W/m2'C
EXCHANGER CONFIGURATION
No. of Shells Passes: 1 No. of Tube Passes: 1
Shell Inner Diameter: 339 mm Shell Outer Diameter: NA
Tube Inner Diameter: 19.862 mm Tube Outer Diameter: 25.4 mm
Area: 19.6308 m2 No. Of Tubes: 82
Bundle Diameter: 0 mm Tube Length: 3 m
Tube Pitch: 90 mm Layout: Triangular
No. of Baffles: 43 Type of Baffle: Segmental
Baffle Spacing: 67.8 mm % Cut: 25%
Notes:
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ANNEX II Russian standards GOST 15118-79, GOST-15120-79 and 15122-
79.
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ANNEX III Bulgarian standard BDS_EN_ISO_16812
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116
117
118
119
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ANNEX IV Earthquake data
S Tb Tc Td T q
A 1 0,15 0,4 2 4 3
B 1,2 0,15 0,5 2 4 1,5
C 1,15 0,2 0,6 2 4 D 1,35 0,2 0,8 2 4 E 1,4 0,15 0,5 2 4 vertical 1 0,05 0,15 1 4
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ANNEX V Drawings
