Universidad de

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ISSN – p: 1390 –9428 / ISSN – e: 3028-8533 / INQUIDE / Vol. 07 / Nº 01

Facultad de

Ingeniería Química

Ingeniería Química y Desarrollo

Universidad de Guayaquil | Facultad de Ingeniería Química | Telf. +593 4229 2949 | Guayaquil – Ecuador

https://revistas.ug.edu.ec/index.php/iqd

Email: inquide@ug.edu.ec | francisco.duquea@ug.edu.ec

Pag. 1

Thermo-hydraulic design of a horizontal shell and tube condenser for

methanol condensation

Diseño térmico-hidráulico de un condensador de tubo y coraza horizontal para la condensación

de metanol

Amaury Pérez Sánchez

* ; Yerelis Pons García

, Daynel Basulto Pita

, Elizabeth Ranero González

& Eddy Javier

Pérez Sánchez

Received: 08/08/2024 – Accepted: 23/11/2024 – Published: 01/01/2025

Research

Articles

Review

Articles

Essay

Articles

* Author for correspondence.

Abstract.

Shell and tube condensers are considered to be a significant part of refrigeration and power plants systems, as well as other petrochemical applications where

the heat exchangers are usually employed. This article presents the thermo-hydraulic design of a one shell pass: two tube passes (1-2) horizontal shell and tube

heat exchanger to condense a stream of pure methanol vapours using chilled water as a coolant. Several design parameters were calculated such as the heat

transfer area (119.33 m2); the number of tubes (285); the shell internal diameter (800.56 mm) and the overall heat transfer coefficient (618.47 W/m2.K). The

calculated heat duty for this heat exchange service was of 8,272.5 kW, while the required flowrate for chilled water was of 151.59 kg/s. The calculated pressure

drops for the condensing methanol and chilled water streams were 7,372.55 Pa and 84,289.69 Pa respectively, which are below the maximum limit values set

by the process. The designed 1-2 shell and tube condenser was of pull-through floating head type, with a baffle cut of 45% and a baffle spacing equal to the shell

internal diameter.

Keywords.

Heat Exchanger Area, Methanol Condensation, Pressure Drop, Shell And Tube Condenser, Thermo-Hydraulic Design

Resumen.

Los condensadores de tubo y coraza están considerados ser una parte significativa de los sistemas de refrigeración y plantas de potencia, así como de otras

aplicaciones petroquímicas donde los intercambiadores de calor son usualmente empleados. Este artículo presenta el diseño térmico-hidráulico de un

intercambiador de calor de tubo y coraza horizontal de un paso por la coraza y dos pasos por los tubos (1-2) para condensar una corriente de vapores de metanol

puro usando agua fría como agente de enfriamiento. Varios parámetros de diseño fueron calculados tales como el área de transferencia de calor (119,33 m2); el

número de tubos (285); el diámetro interno de la coraza (800,56 mm) y el coeficiente global de transferencia de calor (618,47 W/m2.K). La carga de calor

calculada para este servicio de transferencia de calor fue de 8 272,5 kW mientras que el caudal requerido de agua fría fue de 151,59 kg/s. Las caídas de presión

calculadas para las corrientes de metanol condensante y agua fría fueron 7 372,55 Pa y 84 289,69 Pa respectivamente, las cuales están por debajo de los valores

límites máximos fijados por el proceso. El condensador de tubo y coraza 1-2 diseñado fue del tipo cabezal flotante, con un corte del deflector de 45 % y un

espaciado del deflector igual al diámetro interno de la coraza.

Palabras clave

Área De Intercambio De Calor, Condensación De Metanol, Caída De Presión, Condensador De Tubo Y Coraza, Diseño Térmico-Hidráulico

1. Introduction

Transfer of heat from one fluid to another is an important

operation for most of the chemical industries. The most

frequent application of heat transfer is the design of heat

transfer equipment to exchange heat from one fluid to

another fluid. Such devices for effective transfer of heat are

commonly named heat exchanger.

Heat exchangers are based on the principle of heat transfer

taking place between the higher temperature fluid and the

lower temperature fluid. Heat exchangers operate by

permitting the first fluid at a higher temperature to interact

with the second fluid either directly or indirectly at a lower

temperature. This allows heat to transfer from the first to the

second fluid, resulting in a reduction in the second fluid’s

University of Camagüey; Faculty of Applied Sciences; amaury.perez84@gmail.com; https://orcid.org/0000-0002-0819-6760; Camagüey; Cuba.

University of Camagüey; Faculty of Applied Sciences; yerelis.pons@reduc.edu.cu; https://orcid.org/0009-0003-0440-1784; Camagüey; Cuba.

Center of Genetic Engineering and Biotechnology of Camagüey; Department of Production; daynel.basulto@cigb.edu.cu; https://orcid.org/0009-0005-1629-

8846, Camagüey; Cuba.

University of Camagüey; Faculty of Applied Sciences; elizabeth.ranero@reduc.edu.cu; https://orcid.org/0000-0001-9755-0276; Camagüey; Cuba.

Company of Automotive Services S.A.; Commercial Department; eddy.perez@reduc.edu.cu; https://orcid.org/0000-0003-4481-1262; Ciego de Ávila, Cuba.

temperature and an increase in the first fluid’s temperature.

Depending on whether heating or cooling is necessary, heat

is transferred towards or away from the given system [1].

Typical applications include heating or cooling of a fluid

stream of interest and evaporation or condensation of

single- or multicomponent fluid streams. In other

applications, the objective may be to recover or reject heat,

or sterilize, pasteurize, fractionate, distill, concentrate,

crystallize, or control a process fluid. In most heat

exchangers, heat transfer between fluids takes place through

a separating wall or into and out of a wall in a transient

manner. In various heat exchangers, the fluids are separated

by a heat transfer surface, and ideally, they do not mix or

leak [2].

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Pag. 2

The heat exchangers are classified according to the transfer

process, number of fluids, degree of surface compactness,

construction features, flow arrangements and heat transfer

mechanisms. Heat exchangers are extensively used in many

engineering applications such as power engineering,

petroleum refining, refrigeration, air conditioning, food

industry, and biotechnological and chemical process

industries [3].

Among the different types of heat exchangers, shell and

tube heat exchangers (STHE) are the most common used in

industry. They are relatively easy to manufacture and have

multipurpose application potentials for gaseous and liquid

media in large temperature and pressure ranges [3], with

operating temperatures of - 20 ºC to over 500 ºC, maximum

operating pressure of 600 bar and diameters ranging from

60 to over 2,000 mm [4]. According to [5] the shell and tube

heat exchanger is used when a process requires large

amounts of fluids to be heated (or vaporized) or cooled (or

condensed), while due to their design, they offer a large heat

transfer area and provide high heat transfer efficiency.

The main components of a STHE are identified in Figure 1,

showing a 1-2 heat exchanger, that is, a heat exchanger with

one shell-side pass and two tube-side passes. The tube-side

fluid enters the heat exchanger from the tube fluid inlet (1)

into front-end head (11) and from there into the tubes (4).

The tube fluid exits the tubes of the first pass into the rear-

end head (6), continues back into the second tube-side pass

from which it exits back into the front-end head, and finally

out of the heat exchanger from the tube fluid outlet (9). The

shell-side fluid enters the heat exchanger from the shell inlet

(2), flows on the shell side in a cross-parallel flow pattern in

respect to the first pass, and cross-counter flow in respect to

the tube pass, guided by baffle plates (8), and eventually

exits from shell outlet (7). The tube bundle is held in place

by the tube sheets (5), and inside the shell supported by the

baffle plates [6].

Fig 1. A shell and tube heat exchanger type 1-2.

Source: [6]

Condensers are used extensively in chemical and petroleum

processing for distillation, refrigeration, and power

generation systems. Condensers are an integral part of

almost all the operations in the process industry.

Condensers are basically two-phase flow heat exchangers in

which the heat is generated when vapours are converted to

liquid. The heat generated is rejected to a cooling medium,

which acts as a heat sink. In condensers, the latent heat is

given up by the process fluid and transferred to the cooling

medium. Cooling water or surrounding air is generally used

as cooling medium in most of the operations in the industry.

Usually, the vapors that enter a condenser are saturated,

although in many operations, the process fluid may enter the

condenser in the form of superheated vapors. This normally

happens when the vapors are not obtained as distillate of a

distillation column. The superheated vapors are first

desuperheated until saturation and then condensation takes

place. The process fluid then leaves the condenser as a

saturated or a sub-cooled liquid, depending upon the

temperature of the cooling medium and the condenser

design [7].

Condensers can also be classified as partial or total. When a

pure component is considered, it will condense isothermally

and will be totally condensed. If a mixture of components is

considered, then it can be either be condensed totally or

partially based on process conditions [7]. Condensers can

be horizontally or vertically oriented with the condensation

on the tube-side or the shell-side. The magnitude of the

condensing film coefficient for a given quantity of vapor

condensation on a given surface is notably different

depending on the orientation of the condenser. The

condensation normally takes place on the shell-side of

horizontal exchangers and the tube-side of vertical

exchangers. Horizontal shell-side condensation is normally

chosen because condensing film transfer coefficients are

usually higher [8].

Shell and tube condensers are the most commonly used heat

exchangers in process industries because of their relatively

simple manufacturing and their adaptability to different

operating conditions. Although this condenser type

differentiates itself by low-pressure drops with high flow

velocities, its capital requirement, as well as the combined

power and capital cost requirement due to pressure drops of

the pumped and compressed streams in a unit, can be very

expensive [9].

Shell-and-tube condensers with condensation on the shell-

side are widely used in both process and refrigeration

industries. This usually means that the condensing medium

is fed to the top of a shell-and-tube heat exchanger, and then

while flowing on the outside of the tubes it condenses,

leaving its latent heat to the cooling medium flowing inside

the tubes. The condensed liquid is collected at the bottom of

the shell where it leaves the condenser. The heat transfer in

a shell-and-tube condenser is difficult to predict. Factors

such as the complex geometry of the tube bank, effect of the

tube surface geometry, vapour shear effects and condensate

inundation from the tubes above all have an effect on heat

transfer [10].

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Pag. 3

In a broad sense, the design of a new heat exchanger means

the determination/selection of an exchanger construction

type, flow arrangement, tube material, and the physical size

of an exchanger to meet the specified heat transfer and

pressure drops within all specified constraints. For a STHE

a sizing problem generally refers to the determination of

shell type, diameter and length; number of tubes; tube

layout; pass arrangement; heat exchange area; overall heat

transfer coefficient; pressure drop, fluids velocities and

Reynolds number of both fluids; as well as many other

parameters. Inputs to the sizing problem are fluid flow rates,

inlet and outlet fluid temperatures, fouling factors;

diameters, length and material of tubes, and maximum

possible pressure drop of the two fluids involved [2].

There are several studies reported where a shell and tube

condenser is designed, sized or studied for a particular heat

exchange service. In this sense, in [11] a shell and tube heat

exchanger was designed to condense 0.03 kg/s of a

superheated steam stream from 320 ºC (2 bar) to 120 ºC,

using water as coolant. Also, [7] designed a shell and tube

desuperheater-condenser for ammonia-water system by

using standard correlations based on Kern’s method, while

the results were compared with HTRI Xchanger Suite

Educational 6.0 software. Likewise, [2] carried out the

geometry optimization of a shell and tube heat exchanger

working as a condenser in a brewing company, where 6,000

kg/h of steam enters at 160 ºC on the shell side to heat

cooled water to an outlet temperature of 35.5 ºC. Several

parameters were calculated such as heat load, pressure drop,

optimum insulation cost, tube length, thickness and tube

patterns, among others, thus selecting the allowable design

from the thermal design point of view. In [12] a conceptual

technological design aspect of a super vacuum hybrid

surface steam condenser was theoretically analyzed.

Similarly, in [13] the redesign and construction of a shell

and tube condenser intended for the study of forced

convection, with and without changing phase, was

developed with the purpose of providing a laboratory of an

equipment to perform practical experiences that could be

used to analyze such phenomena. In this study, the research

was conducted in three phases: 1) diagnosis, carried out

through survey techniques; 2) the techno-economic

feasibility, which involved the redesign, material selection,

and cost determination; and 3) construction and operational

testing of the redesigned condenser. In [14] an optimization

of a shell and tube condenser was performed for a low

temperature thermal desalination plant. Also, in [15] a

procedure was proposed for the design of the components

of a heat exchanger network, including the condensers of

the network, using pinch analysis to maximize heat

recovery for a given minimum temperature difference, and

using genetic algorithm in order to minimize its total annual

cost. Other authors [9] studied the effect of baffle spacing

on heat transfer area and pressure drop for shell side

condensation in the most common types of segmentally

baffled shell and tube condensers (TEMA E and J types with

conventional tube bundles), while a set of correlations were

also presented to calculate the optimum baffle spacing. In

[16] a shell and tube condenser was designed from both the

thermal and mechanical points of view, for its application in

a steam turbine of a thermal power plant. In [10] a detailed

2D Computational Fluid Dynamics (CFD) calculation for

vapour flow field and condensation rate was carried out for

geometry similar to a real full-scale shell-and-tube

condenser with 100 tubes, with condensation occurring on

the shell-side. The differences in vapour flow behaviour

were investigated for pure R22 and for a binary mixture of

R32 and R134a. In [17] a dynamic mathematical model of

a shell-and-tube condenser operating in a vapour

compression refrigeration plant was presented and

validated. The model was formulated from mass continuity,

energy conservation and heat transfer physical

fundamentals by using a lumped-parameter formulation for

the condenser that is similar to the ones presented in

previous studies, but with some differences in the selection

of control volumes and including the refrigerant dynamics

in a simplified way. In [18] a thermo-economic

optimization of a shell and tube condenser was presented,

based on two new optimization methods, namely genetic

and particle swarm (PS) algorithms. The procedure was

selected to find the optimal total cost including investment

and operation cost of the condenser. Initial cost included

condenser surface area and operational cost involved pump

output power to overcome the pressure loss. In [19] a

moving-boundary lumped parameter (MB) dynamic model

of a shell-and-tube condenser was proposed, where the

mean void fraction (MVF) correlation used can be changed

in order to analyze the influence of the MVF correlation on

the model performance, comparing the predictions obtained

with experimental data using MVF correlations frequently

mentioned in the literature. The experimental data used

consisted of transients obtained varying the working

conditions in a wide range, and being the MVF correlations

used derived from different local void fraction expressions

which allowed a relatively easy analytical or numerical

integration. In [20] the energy, exergy and economic

assessments of a shell and tube condenser of 580 MW

nuclear power plant was carried out using different water-

based hybrid nanofluids as coolant. The effect of

nanoparticle concentration on reductions of coolant

requirement, pumping power and operating cost was also

investigated. Similarly, in [21] the constructal design of a

shell-and-tube condenser with ammonia-water as the

working fluid was studied based on constructal theory. A

complex function (CF) formed by entropy generation rate

(EGR) and total pumping power (TPP) was minimized, and

the tube diameter was optimized, while the parameter that

influences on the optimal results were researched. In [22] a

horizontal shell and tube condenser was designed from the

thermo-hydraulic point of view using the calculation

approach reported by [23], in roder to condense ethanol

vapours usign chilled water as coolant. Finally, in [24] a

linear design optimization approach was presented for the

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Pag. 4

first time for horizontal shell and tube condensers that

guarantees global optimality. The numerical results

obtained in this study indicated that the proposed approach

can present solutions with lower heat transfer areas than two

design procedures presented in chemical process design

textbooks.

There are textbooks [8,25,26,27] that present several design

methodologies related with the design of condensers using

heuristics based on choices made by the designer. The main

concept is to find one viable exchanger, not necessarily the

optimal one [24].

In certain chemical processing plant, a vapour stream of

pure methanol is obtained at the overhead of a distillation

column, and it is desired to condense it using a shell and

tube heat exchanger. In this context, in the present paper a

1-2 shell and tube condenser was designed from the thermo-

hydraulic point of view, in order to condense methanol

vapours by using chilled water as coolant. To accomplish

this task, the calculation approach described in [23] was

applied, where several important design parameters were

calculated such as the number of tubes, heat exchange area,

shell internal diameter, overall heat transfer coefficient and

the pressure drops of both fluids.

2. Materials y methods.

2.1. Problem statement.

It’s desired to condense 30,000 kg/h of methanol vapours

that comes from the overhead of a distillation column at a

temperature of 120 ºC. Chilled water is available at 2 ºC for

this service, while the temperature rise is to be limited to 13

ºC for this coolant. According to plant standards, stainless

steel tubes with a nominal diameter of ¾ inch 40ST are

required, with a length of 5.0 m. The condenser is to operate

at 5.0 bar, the vapours are to be totally condensed and no

subcooling is required. The pressure drop of the chilled

water and methanol should not exceed 100,000 Pa and

10,000 Pa, respectively. Design a suitable horizontal one

shell, two tube pass (1-2) shell and tube heat exchanger for

this condensation service, using a flow arrangement of

countercurrent type.

2.2. Calculation methodology.

2.2.1. Heat exchange area of the condenser.

The calculation steps included in the calculation procedure

to design the horizontal shell and tube heat condenser are

presented below, specifically to determine the heat

exchange area.

Step 1. Definition of the initial parameters:

• Vapour flowrate (



): [kg/h].

• Operating pressure of the condenser (



): [bar].

• Inlet temperature of vapour (



): [ºC].

• Condensation temperature of vapour at the

operating pressure of condenser (



): [ºC].

• Molecular weight of methanol (



): [kg/kmol].

• Enthalpy of vapour methanol at 



and 



(



[kJ/kg].

• Enthalpy of liquid (condensate) methanol at 



and





(



): [kJ/kg].

• Inlet temperature of chilled water (



): [ºC].

• Outlet temperature of chilled water (



): [ºC].

• Inside diameter of tubes (



): [m].

• Outside diameter of tubes (



): [m].

• Length of tubes (



): [m].

• Thermal conductivity of the tubes material (



[W/m.K].

• Fouling factor of water (



): [K.m

/W].

• Fouling factor of condensing methanol (



[K.m

/W].

• Number of tube side passes (



Step 2. Heat duty ():











󰇛









󰇜

(1)

Step 3. Assumption of the overall heat transfer coefficient

(



Step 4. Location of fluids inside the shell and tube heat

exchanger.

The vapours will be located in the shell side, while the

chilled water will flow inside the tubes.

Step 5. Log mean temperature difference ():



󰇛









󰇜



󰇛









󰇜



󰇛









󰇜

󰇛









󰇜

(2)

Step 6. Factor R:



















(3)

Step 7. Factor S:



















(4)

Step 8. Temperature correction factor (



For a 1 shell: 2 tube pass heat exchanger, the correction

factor is given by the following equation:









󰇛







󰇜



󰇛



󰇜

󰇛



󰇜



󰇛



󰇜

󰇯

󰇣

󰇛







󰇜

󰇤

󰇣

󰇛







󰇜

󰇤

󰇰

(5)

Step 9. True temperature difference ():









(6)

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Pag. 5

Step 10. Trial heat exchange area (



















(7)

Step 11. Surface area of one tube (



Ignoring the tube sheet thickness [23]:













(8)

Where 



and 



are given in m.

Step 12. Number of tubes (

















(9)

Step 13. Tube pitch (



A square pitch is selected. Thus:









(10)

Where 



is given in mm.

Step 14. Tube bundle diameter (

























(11)

Where 



is given in mm, while 



and  are constants that

depend on the type of pitch selected and the number of tube

passes [23].

Step 15. Number of tubes in centre row (

















(12)

Where both 



and 



are given in mm.

Step 16. Assume a condensation film coefficient (

󰇛󰇜

Step 17. Mean temperature of both the shell-side (

󰆽

) and

tube-side (



) fluids:

• Shell side (methanol vapours):



󰆽













(13)

• Tube side (chilled water):

















(14)

Step 18. Tube wall temperature (









󰆽



󰇛



󰆽





󰇜







󰇛󰇜

(15)

Step 19. Mean temperature condensate (





󰆽







󰆽







(16)

Step 20. Physical properties of the liquid methanol at the

mean temperature condensate (



• Viscosity (



) [Pa.s].

• Density (



) [kg/m

• Thermal conductivity (



) [W/m.K].

Step 21. Density of vapour methanol at mean vapour

temperature (





















󰆽











(17)

Where 



is 1.0 bar.

Step 22. Condensate loading on a horizontal tube (





















(18)

Step 23. Average number of tubes in a vertical tube row

(

















(19)

Step 24. Calculated mean condensation film coefficient for

a tube bundle (

󰇛󰇜



󰇛󰇜





󰇩







󰇛









󰇜











󰇪









(20)

Where  is the gravitational acceleration = 9.81 m/s

Step 25. Verification of the mean condensation film

coefficient calculated in step 24 [

󰇛󰇜

] with the mean

condensation coefficient assumed in step 16 [

󰇛󰇜

In this approach, if the value of 

󰇛󰇜

is close enough to the

assumed value [

󰇛󰇜

] then no correction of 



is necessary.

Step 25. Tube cross-sectional area (





























(21)

Where 



is given in m.

Step 26. Density of tube-side fluid (chilled water) at 



(



Step 27. Heat capacity of the tube-side fluid (chilled water)

at 



(



Step 28. Required flowrate of the tube-side fluid (chilled

water) (











󰇛









󰇜





(22)

Where  is given in kW.

Step 29. Velocity of the tube-side fluid (





















(23)

Step 30. Film coefficient for the tube side fluid (



• Specifically for water:









󰇛





󰇜













(24)

Where 



is given in mm.

Step 31. Overall heat transfer coefficient calculated (



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Pag. 6











󰇛󰇜











󰇡









󰇢





































(24)

Where 



and 



are given in m.

Step 32. Compare the calculated overall heat transfer

coefficient (



) with the overall heat transfer coefficient

assumed in step 3 (



If the values are significantly different, then the calculation

procedure should be repeated using a new trial assumed

value for 



which could be close enough to 



2.3. Pressure drops.

2.3.1. Shell side pressure drop.

Step 33. Selection of the heat exchanger.

Step 34. Selection of the baffle spacing (



) and cut.

Step 35. Clearance (



Step 36. Shell internal diameter (















(25)

Where both 



and 



are given in mm.

Step 37. Cross-flow area (









󰇛









󰇜













(26)

Where all the parameters are given in m.

Step 38. Mass velocity of the vapour (

















(27)

Step 39. Viscosity of the methanol vapour (



) at the mean

temperature 

󰆽

Step 40. Linear velocity of the vapour (

















(28)

Step 41. Shell-side equivalent diameter (



• For square pitch arrangement:















󰇛













󰇜

(29)

Where both 



and 



are given in m.

Step 42. Reynolds number of the vapour (





















(30)

Step 43. Shell-side friction factor (



)

Step 44. Pressure drop of the shell-side fluid (vapour) (



The pressure drop of the shell side fluid can be assumed as

50% of that calculated using the inlet flow. Thus:











󰇩



󰇧









󰇨

















󰇛





󰇜



















󰇪

(31)

Where 



and 



are given in mm, while 



and 



are given

in m. The term















, which is the viscosity

correction factor, could be neglected [23].

2.3.2. Tube side pressure drop.

Step 45. Viscosity of tube-side fluid (chilled water) at 



(



Step 46. Reynolds number of tube-side fluid (chilled water)

(

























(32)

Where 



is given in m.

Step 47. Tube-side friction factor (



Step 48. Pressure drop of the tube-side fluid (chilled water)

(











󰇩



























󰇪















(33)

Where the term 









 



can be neglected [23] and 



and 



are given in m.

3. Results

3.1. Heat exchange area of the condenser

Step 1. Definition of the initial parameters

Table 1 shows the initial parameters required to design the

horizontal shell and tube condenser.

Table 1. Initial parameters required to design the horizontal

shell and tube condenser.

Parameter

Symbol

Value

Unit

Vapour flowrate





30,000

kg/h

Operating pressure of

the condenser





5.0

bar

Inlet temperature of

vapour





120

ºC

Condensation

temperature of vapour

at the operating

pressure of condenser





110.73

ºC

Molecular weight of

methanol





32.04

kg/kmol

Enthalpy of vapour

methanol at 



and 







1,134.12

kJ/kg

Enthalpy of liquid

methanol at 



and 







141.42

kJ/kg

Inlet temperature of

chilled water





ºC

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Pag. 7

Outlet temperature of

chilled water





ºC

Inside diameter of

tubes





20.93

Outside diameter of

tubes





26.67

Length of tubes





5.0

Thermal conductivity

of the tubes material





W/m.K

Fouling factor of water





0.00035

K.m

Fouling factor of

condensing methanol





0.00020

K.m

Number of tube side

passes





As reported by [28].

As reported by [23].

Source: Own elaboration.

Table 2 presents the results of the parameters calculated in

steps 2-15.

Table 2. Results of the parameters calculated in steps 2-15.

Step

Parameter

Symbol

Value

Units

Heat duty



8,272.5

Assumption of

the overall heat

transfer

coefficient





650

W/m

Log mean

temperature

difference



106.87

ºC

Factor R



0.713

Factor S



0.110

Temperature

correction factor





0.998

True

temperature

difference



106.65

ºC

Trial heat

exchange area





119.33

Surface area of

one tube





0.419

Number of tubes





 285

Tube pitch





33.34

Tube bundle

diameter





707.56

Number of tubes

in centre row





 22

As reported by [26] in the range of 550-1,100 W/m

.K for

the condensation of organic solvents in shell and tube,

water-cooled condensers.

The condensation range is small and the change in

saturation temperature will be linear, so the log mean

temperature difference can be used [23].

The constants 



and  have values of 0.156 and 2.291

respectively, for a square pitch and two tube passes [23].

Source: Own elaboration.

Step 16. Assume a condensation film coefficient (

󰇛󰇜

According to the range reported by [8] of 1,000-2,500

W/m

.K for the condensation of low viscosity organic

compounds, a value of 1,600 W/m

.K was assumed for the

condensation film coefficient of methanol on the shell side.

Table 3 displays the results of the parameters calculated in

steps 17-24.

Table 3. Results of the parameters calculated in steps 17-24.

Step

Parameter

Symbol

Value

Units

Mean

temperature of

methanol



󰆽

115.36

ºC

Mean

temperature of

chilled water





8.50

ºC

Tube wall

temperature





71.95

ºC

Mean

temperature

condensate





93.66

ºC

Viscosity of

liquid

methanol





0.000269

Pa.s

Density of

liquid

methanol





717.38

kg/m

Thermal

conductivity of

liquid

methanol





0.1806

W/m.K

Density of

vapour

methanol at 

󰆽





5.027

kg/m

Condensate

loading on a

horizontal tube





0.0058

kg/s.m

Average

number of tubes

in a vertical

tube row





 15

Calculated

mean

condensation

film coefficient

for a tube

bundle



󰇛󰇜

1,611.97

W/m

As reported by [28].

Source: Own elaboration.

Step 25. Verification of the mean condensation film

coefficient calculated in step 24 [

󰇛󰇜

] with the mean

condensation coefficient assumed in step 16 [

󰇛󰇜

Since the calculated mean condensation film coefficient

(

󰇛󰇜

= 1,611.97 W/m

.K) is close enough to the assumed

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Pag. 8

value in step 16 (

󰇛󰇜

= 1,600 W/m

.K), then no correction

of the tube wall temperature (



) is needed.

Table 4 describes the results of the parameters calculated in

steps 25-31.

Table 4. Results of the parameters calculated in steps 25-

31.

Step

Parameter

Symbol

Value

Units

Tube cross-

sectional area





0.049

Density of

chilled water at









999.818

kg/m

Heat capacity of

chilled water at









4.1977

kJ/kg.K

Required

flowrate of

chilled water





151.59

kg/s

Velocity of

chilled water





3.094

m/s

Film coefficient

for chilled water





8,576.87

W/m

Calculated

overall heat

transfer

coefficient





618.47

W/m

Source: Own elaboration.

Step 32. Comparison of the calculated overall heat transfer

coefficient (



) with the overall heat transfer coefficient

assumed in step 3 (



The value of the overall heat transfer coefficient calculated

in step 31 (



= 618.47 W/m

.K) is close enough to the

overall heat transfer coefficient assumed in step 3 (



= 650

W/m

.K), so there is no need to repeat the calculation

procedure and carry out a new trial, and the heat exchanger

area of the condenser will be 119.33 m

3.2. Pressure drops.

3.2.1. Shell side pressure drop

Step 33. Selection of the heat exchanger.

In this study we select a pull-through floating head heat

exchanger, with no need for close clearance. The reported

advantages of these heat exchangers include that they are

more versatile than fixed head and U-tube exchangers, are

suitable for high temperature differentials and, as the tubes

can be rodded from end to end and the bundle removed, are

easier to clean and can be used for fouling liquids [23].

Step 34. Selection of the baffle spacing (



) and cut.

As specified by [23], the baffle spacings used in commercial

shell and tube heat exchangers range from 0.2 to 1.0 shell

internal diameter (



). Close baffle spacing will give higher

heat-transfer coefficients, but at the expense of higher

pressure drops. The optimum spacing will usually be

between 0.3 to 0.5 times the shell diameter. However,

although the construction of a condenser is similar to other

shell and tube exchangers, wider baffle spacings are

applied, typically 



=



[23]. In this study we obeyed this

last rule-of-thumb, that is, we selected a value for the baffle

spacing equal to the shell internal diameter.

The term “baffle cut” is used to specify the dimensions of a

segmental baffle. The baffle cut is the height of the segment

removed to form the baffle, expressed as a percentage of the

baffle disc diameter. Baffle cuts from 15 to 45 % are

commonly used [23]. In this study, we chose a baffle cut of

45%.

Step 35. Clearance (



According to [23], for a value of the tube bundle diameter

(



) of 707.56 mm (0.707 m), the clearance will be 93 mm

for a pull-through floating head heat exchanger.

Step 36. Shell internal diameter (















󰨙

(25)

Thus, the baffle spacing (



) will be 800.56 mm.

Table 5 exhibits the results of the parameters calculated in

steps 37-44.

Table 5. Results of the parameters calculated in steps 37-44.

Step

Parameter

Symbol

Value

Units

Cross-flow

area





0.128

Mass velocity

of the vapour





65.10

kg/m

Viscosity of

the methanol

vapour at the

mean

temperature 

󰆽





0.0000128

Pa.s

Linear

velocity of the

vapour





12.95

m/s

Shell-side

equivalent

diameter





0.0263

Reynolds

number of the

vapour





133,760.15

Shell-side

friction factor





0.023

Pressure drop

of methanol

(vapour)





7,372.55

For a baffle cut of 45% and a Reynolds number of 1.34x10

[23].

Source: Own elaboration.

3.2.2. Tube side pressure drop.

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Pag. 9

Table 6 shows the results of the parameters calculated in

steps 45-48.

Table 6. Results of the parameters calculated in steps 45-48.

Step

Parameter

Symbol

Value

Units

Viscosity of

chilled water at 







0.001364

Pa.s

Reynolds number

of chilled water





47,467.47

Tube-side friction

factor





0.0033

Pressure drop of

the chilled water





84,289.69

For a Reynolds number of 4,74x10

[23].

Source: Own elaboration.

4. Discussion.

The heat duty for this heat exchange service was 8,272.5

kW, with a true temperature difference of 106.65 ºC. In [7]

a shell and tube desuperheater-condenser was designed to

condense a stream of ammonia using cooling water as

coolant, and by using the standard correlations based on

Kern’s method. In this study, the calculated heat duty is

903.78 kW. In [23] the heat duty is 4.368.8 kW for a

horizontal shell and tube condenser designed to condense a

stream of mixed light hydrocarbon vapours using cooling

water. In [8] a horizontal shell and tube condenser is

designed to condense acetone vapours on the shell side by

using cooling water as coolant, thus obtaining a calculated

heat duty of 2,865 kW. In another study [22], a value for the

calculated heat duty of 6,578.47 kW was obtained for a

horizontal shell and tube condenser designed to condense a

stream of ethanol vapours.

The calculated film coefficient for chilled water (8,576.87

W/m

.K) was 5.32 times higher than the calculated mean

condensation film coefficient for the condensing methanol

vapours (1,611.97 W/m

.K). This result doesn’t agree with

the findings of [7], where the mean condensation film

coefficient of an ammonia stream condensing on the shell

side (11,836.73 W/m

.K) is 1.80 times higher than the heat

transfer coefficient of the cooling water on the tube side

(6567.32 W/m

.K). It’s worth mentioning that in this study

[7] the horizontal shell and tube heat exchanger was also

designed using the HTRI software, and the results obtained

by this software regarding the calculation of the film heat

transfer coefficients for both streams agree with those

reported on the present study, that is, the value of the film

heat transfer coefficient for cooling water (7,081.82

W/m

.K) is higher (1.41 times) than the mean condensation

film coefficient of condensing ammonia on the shell side

(4,995.21 W/m

.K). In [23] the calculated film coefficient

for cooling water flowing inside the tubes (7,097 W/m

.K)

is 4.90 times higher than the calculated mean condensation

film coefficient (1,447 W/m

.K) for the mixed light

hydrocarbon stream condensing on the shell side. In [8], the

calculated film coefficient for a cooling water stream

flowing on the tube side (7,483 W/m

.K) is 5.02 times

higher than the calculated mean condensation film

coefficient (1,491 W/m

.K) for the acetone stream

condensing on the shell side. In [22] the value of tube side

heat transfer coefficient for chilled water is 6,265.59

W/m

.K, which is 7.55 times higher than the mean

condensation film coefficient for the ethanol condensing

vapours (829.38 W/m

.K). The value of the film coefficient

for chilled water calculated in this study is higher than the

range suggested by [8] of 2,000-6,000 W/m

.K, while the

value of the mean condensation film coefficient for the

condensing methanol vapours agrees with the range

reported by the same author of 1,000–2,500 W/m

.K.

Respecting the overall heat transfer coefficient calculated in

this study (618.47 W/m2.K), its value agrees with the range

reported by [26] of 550-1,100 W/m

.K; the range reported

by [25] of 300-1,000 W/m

.K, and the range reported by

[28] of 568-1,136 W/m

.K, while it’s slightly lower than the

range reported by [23] of 700-1,000 W/m

.K.

About 151.59 kg/s of chilled water will be needed in this

study to carry out the condensation of the pure methanol

vapours. In [8], a flowrate of cooling water of 68.54 kg/s is

needed to condense 5.8 kg/s of acetone vapours. In [23] a

flowrate of cooling water of 104.5 kg/s is needed to

condense 12.5 kg/s of mixed light hydrocarbon vapours. In

[22] about 157 kg/s are needed to condense 6.94 kg/s of pure

ethanol vapours. In [7] the flowrate of cooling water

required to condense 0.6841 kg/s of ammonia vapours is

35.97 kg/s.

The shell and tube condenser designed in this paper will

have the following parameters:

• Heat exchange area: 119.33 m

• Number of tubes: 285.

• Tube bundle diameter: 707.56 mm.

• Number of tubes in centre row: 22.

• Shell internal diameter: 800.56 mm.

In [7], the horizontal desuperheater-condenser designed to

carry out the condensation of ammonia vapours, has the

following design parameters:

• Heat exchange area: 116.62 m

• Number of tubes: 784.

• Shell internal diameter: 840 mm.

In [8], the horizontal shell and tube condenser designed to

condense acetone vapours using cooling water, has the

following design results:

• Heat exchange area: 96.0 m

• Number of tubes: 488.

• Shell internal diameter: 623 mm.

In [23] the horizontal shell and tube heat exchanger

condenser designed to condense a stream of mixed light

hydrocarbon vapours has the following design parameters:

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Pag. 10

• Heat exchange area: 364 m

• Number of tubes: 1194.

• Shell internal diameter: 1130 mm.

In [22] the horizontal shell and tube heat exchanger

condenser designer to condense a stream of ethanol vapours

has the design parameters presented below:

• Heat exchange area: 223.68 m

• Number of tubes: 731.

• Tube bundle diameter: 744.57 mm.

• Shell internal diameter: 1,684 mm.

The calculated pressure drop for the chilled water was

84,289.69 Pa, which is 11.43 times higher than the

calculated pressure drop of the condensing methanol

(7,372.55 Pa). Both values of the pressure drop are below

the limits established by the heat transfer process. In [23]

the pressure drop of the cooling water flowing inside the

tubes (53,000 Pa) is 40.77 times higher than the pressure

drop of a mixed light hydrocarbon vapours condensing on

the shell side (1,300 Pa). Likewise, in [7] the pressure drop

of cooling water flowing inside the tubes (49,130 Pa) is

40.16 times higher than the pressure drop of condensing

ammonia on the shell (1,223.4 Pa). In [22], the pressure

drop of the chilled water stream located in the tubes

(42,192.63 Pa) is 4.19 times higher than the pressure drop

of the condensing ethanol on the shell side (10,069.25 Pa).

5. Conclusions.

In the present work a 1-2 horizontal shell and tube

condenser was designed from the thermo-hydraulic point of

view to condense a stream of pure methanol vapours using

chilled water as coolant. The design calculation

methodology employed to carry out the design task was the

reported in [23]. Several key design parameters were

calculated such as the heat transfer area (119.33 m

); the

number of tubes (285); the tube bundle diameter (707.56

mm) and the shell internal diameter (800.56 mm). Other

important thermal parameters were also computed such as

the heat duty (8,272.5 kW), the overall heat transfer

coefficient (618.47 W/m

.K) and the required flowrate of

chilled water (151.59 kg/s). The calculated values of the

pressure drop of both the shell-side (7,372.55 Pa) and tube-

side (84,289.69 Pa) fluids were below the maximum limit

set by the heat exchanger process. The designed 1-2 shell

and tube condenser is of pull-through floating head type,

with a tube length of 5.0 m, a baffle cut of 45% and a baffle

spacing of 800.56 mm.

6.- Author Contributions.

1. Conceptualization: Amaury Pérez Sánchez.

2. Data curation: Heily Victoria González, Elizabeth

Ranero González.

3. Formal analysis: Amaury Pérez Sánchez, Arlenis

Cristina Alfaro Martínez, Eddy Javier Pérez Sánchez.

4. Acquisition of funds: Not applicable.

5. Research: Amaury Pérez Sánchez, Heily Victoria

González, Eddy Javier Pérez Sánchez.

6. Methodology: Amaury Pérez Sánchez, Arlenis Cristina

Alfaro Martínez, Elizabeth Ranero González.

7. Project management: Not applicable.

8. Resources: Not applicable.

9. Software: Not applicable.

10. Supervision: Amaury Pérez Sánchez.

11. Validation: Amaury Pérez Sánchez.

12. Writing - original draft: Eddy Javier Pérez Sánchez,

Elizabeth Ranero González.

13. Writing - revision and editing: Amaury Pérez Sánchez,

Heily Victoria González.

7.- References

[1] E. J. Fernandes and S. H. Krishanmurthy, "Design and analysis of

shell and tube heat exchanger," Int. J. Simul. Multidisci. Des. Optim.,

vol. 13, no. 15, pp. 1-8, 2022. https://doi.org/10.1051/smdo/2022005

[2] D. Bogale, "Design and Development of Shell and Tube Heat

Exchanger for Harar Brewery Company Pasteurizer Application

(Mechanical and Thermal Design)," American Journal of

Engineering Research, vol. 03, no. 10, pp. 99-109, 2014.

[3] P. Bichkar, O. Dandgaval, P. Dalvi, R. Godase, and T. Dey, "Study

of Shell and Tube Heat Exchanger with the Effect of Types of

Baffles," Procedia Manufacturing, vol. 20, pp. 195-200, 2018.

https://doi.org/10.1016/j.promfg.2018.02.028

[4] B. M. Reyes, "Diseño de un intercambiador de calor (condensador)

para el sistema de la Bomba de Vapor Desalinizadora Solar de Alta

Potencia del proyecto Walfisch," Tesis de Diploma, Universidad

Técnica Federico Santa María, Valparaíso, Chile, 2021.

[5] I. C. Nwokedi and C. A. Igwegbe, "Design of Shell and Tube Heat

Exchanger with Double Passes," Journal of Engineering Research

and Reports, vol. 3, no. 4, pp. 1-12, 2018.

https://doi.org/10.9734/JERR/2018/v3i416883

[6] J. Saari, Heat Exchanger Dimensioning. Lappeenranta, Finland:

Lappeenranta University of Technology, 2015.

[7] S. Sahajpal and P. D. Shah, "Thermal Design of Ammonia

Desuperheater-Condenser and Comparative Study with HTRI,"

Procedia Engineering, vol. 51, pp. 375-379, 2013.

https://doi.org/10.1016/j.proeng.2013.01.052

[8] R. Smith, Chemical Process Design and Integration. West Sussex,

England: John Wiley & Sons Ltd., 2005.

[9] B. K. Soltan, M. Saffar-Avval, and E. Damangir, "Minimizing capital

and operating costs of shell and tube condensers using optimum

baffle spacing," Applied Thermal Engineering, vol. 24, pp. 2801-

2810, 2004. https://doi.org/10.1016/j.applthermaleng.2004.04.005

[10] T. Karlsson and L. Vamling, "Flow fields in shell-and-tube

condensers: comparison of a pure refrigerant and a binary mixture,"

International Journal of Refrigeration, vol. 28, pp. 706-713, 2005.

https://doi.org/10.1016/j.ijrefrig.2004.12.008

[11] D. E. Dana, "Diseño de un condensador para un loop experimental

para porbar maniobras de arranque del reactor CAREM," Proyecto

Integrador, Universidad Nacional de Cuyo, Mendoza, Argentina,

2016.

[12] R. K. Kapooria, S. Kumar, and K. S. Kasana, "Technological

investigations and efficiency analysis of a steam heat exchange

condenser: conceptual design of a hybrid steam condenser " Journal

of Energy in Southern Africa, vol. 19, no. 3, pp. 35-45, 2008.

[13] M. Lucena, C. Linárez, and E. Rodríguez, "Rediseño y construcción

de un condensador para el estudio de convección forzada," REDIP.

UNEXPO. VRB. Venezuela, vol. 5, no. 4, pp. 941-957, 2015.

[14] R. Elakkiyadasan, P. M. Kumar, M. Subramanian, N. Balaji, M.

Karthick, and S. Kaliappan, "Optimization of Shell and Tube

Condenser for Low Temperature Thermal Desalination Plant," E3S

Web of Conferences, vol. 309, p. 01011, 2021.

https://doi.org/10.1051/e3sconf/202130901011

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Guayaquil

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ISSN – p: 1390 –9428 / ISSN – e: 3028-8533 / INQUIDE / Vol. 07 / Nº 01

Facultad de

Ingeniería Química

Ingeniería Química y Desarrollo

Universidad de Guayaquil | Facultad de Ingeniería Química | Telf. +593 4229 2949 | Guayaquil – Ecuador

https://revistas.ug.edu.ec/index.php/iqd

Email: inquide@ug.edu.ec | francisco.duquea@ug.edu.ec

Pag. 11

[15] B. Allen, M. Savard-Goguen, and L. Gosselin, "Optimizing heat

exchanger networks with genetic algorithms for designing each heat

exchanger including condensers," Applied Thermal Engineering, vol.

29, pp. 3437–3444, 2009.

https://doi.org/10.1016/j.applthermaleng.2009.06.006

[16] J. Canoura, "Diseño de un condensador para planta de producción

industrial," Trabajo de Fin de Grado, Universidad de La Coruña, La

Coruña, España, 2016.

[17] R. Llopis, R. Cabello, and E. Torrella, "A dynamic model of a shell-

and-tube condenser operating in a vapour compression refrigeration

plant," International Journal of Thermal Sciences, vol. 47, pp. 926-

934, 2008. https://doi.org/10.1016/j.ijthermalsci.2007.06.021

[18] H. Hajabdollahi, P. Ahmadi, and I. Dincer, "Thermoeconomic

optimization of a shell and tube condenser using both genetic

algorithm and particle swarm," International Journal of

Refrigeration, vol. 34, pp. 1066-1076, 2011.

https://doi.org/10.1016/j.ijrefrig.2011.02.014

[19] V. Milián, J. Navarro-Esbrí, D. Ginestar, F. Molés, and B. Peris,

"Dynamic model of a shell-and-tube condenser. Analysis of the mean

void fraction correlation influence on the model performance,"

Energy, vol. 59, pp. 521-533, 2013.

http://dx.doi.org/10.1016/j.energy.2013.07.053

[20] S. K. Singh and J. Sarkar, "Energy, exergy and economic assessments

of shell and tube condenser using hybrid nanofluid as coolant,"

International Communications in Heat and Mass Transfer, vol. 98,

pp. 41-48, 2018.

https://doi.org/10.1016/j.icheatmasstransfer.2018.08.005

[21] H. Feng, C. Cai, L. Chen, Z. Wu, and G. Lorenzini, "Constructal

design of a shell-and-tube condenser with ammonia-water working

fluid," International Communications in Heat and Mass Transfer,

vol. 118, p. 104867, 2020.

https://doi.org/10.1016/j.icheatmasstransfer.2020.104867

[22] A. Pérez, M. I. L. Rosa, Z. M. Sarduy, E. Ranero, and E. J. Pérez,

"Thermo-hydraulic design of a horizontal shell and tube heat

exchanger for ethanol condensation," Nexo Revista Científica, vol.

36, no. 04, pp. 551-571, 2023.

https://doi.org/10.5377/nexo.v36i04.16770

[23] R. Sinnott and G. Towler, Chemical Engineering Design, 6th ed.

(Coulson and Richardson's Chemical Engineering Series). Oxford,

United Kingdom: Butterworth-Heinemann, 2020.

[24] I. P. S. Pereira, M. J. Bagajewicz, and A. L. H. Costa, "Global

optimization of the design of horizontal shell and tube condensers,"

Chemical Engineering Science, vol. 236, p. 116474, 2021.

https://doi.org/10.1016/j.ces.2021.116474

[25] S. Kakaç, H. Liu, and A. Pramuanjaroenkij, Heat Exchangers -

Selection, Rating and Thermal Design, 3rd ed. Boca Raton, U.S.A:

CRC Press, 2012.

[26] E. Cao, Heat transfer in process engineering. New York, U.S.A: The

McGraw-Hill Companies, Inc., 2010.

[27] M. Nitsche and R. O. Gbadamosi, Heat Exchanger Design Guide - A

Practical Guide for Planning, Selecting and Designing of Shell and

Tube Exchangers. Oxford, U.K: Butterworth Heinemann, 2016.

[28] D. W. Green and M. Z. Southard, Perry's Chemical Engineers'

Handbook, 9th ed. New York, U.S.A: McGraw-Hill Education, 2019.

Nomenclature





Tube cross-sectional area





Surface area of one tube





Heat exchange area





Cross-flow area



Heat capacity

kJ/kg.K





Clearance





Inside diameter of tubes





Outside diameter of tubes





Tube bundle diameter





Shell-side equivalent diameter





Shell internal diameter





Temperature correction factor



Gravitational acceleration

m/s2



Mass velocity

kg/m2.s



Enthalpy

kJ/kg



󰇛󰇜

Assumed condensation film

coefficient

W/m2.K



󰇛󰇜

Calculated mean condensation

film coefficient for a tube bundle

W/m2.K





Shell-side friction factor





Tube-side friction factor



Thermal conductivity

W/m.K





Constant





Baffle spacing



Log mean temperature difference

ºC





Length of tubes



Mass flowrate

kg/h



Molecular weight

kg/kmol



Constant





Number of tube side passes





Number of tubes in centre row





Average number of tubes in a

vertical tube row





Number of tubes





Tube pitch





Atmospheric pressure

bar





Operating pressure of the

condenser

bar



Pressure drop



Heat duty



Factor



Reynolds number





Fouling factor of condensing

methanol

K.m2/W





Fouling factor of water

K.m2/W



Factor



Temperature of cold fluid

ºC





Mean temperature of cold fluid

ºC





Inlet temperature of vapour

ºC





Condensation temperature of

vapour at the operating pressure

of condenser

ºC



󰆽

Mean temperature of hot fluid

ºC





Mean temperature condensate

ºC





Tube wall temperature

ºC



True temperature difference

ºC



Velocity

m/s





Assumed overall heat transfer

coefficient

W/m2.K





Calculated overall heat transfer

coefficient

W/m2.K

Greek symbols



Density

kg/m



Viscosity

Pa.s





Condensate loading on a horizontal tube

kg/s.m

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Pag. 12

Subscripts



Inlet



Outlet



Liquid or condensate



Tube-side fluid



Vapour