individual report (final)

Process Design A 2014-2015:

Input Gathering Lines and

Storage TankGdansk Oil Terminal

Brendan [email protected]

Student ID: H00155203

AbstractIn this design project, Group 12 has been tasked with designing the new oil terminal in Gdansk, Poland. The refinery at Gdansk is supplied with oil from both the Gdansk port

and the Druzba Oil Pipeline from Russia. For this section, the design of the gathering pipeline will considered and the construction of the storage tank with safety considerations will be

explored in depth.

Artists Depiction of finished Oil Terminal (sourced from Gazoprojekt)

Table of ContentsIntroduction.................................................................................................................4

Process Overview....................................................................................................5

Overview of Process Employed...............................................................................5

Overall Mass Balance for the Process.....................................................................6

Overall Energy Balance...........................................................................................7

Tank Design............................................................................................................7

Type of Tank........................................................................................................8

Tank Dimensions.................................................................................................9

Maintaining Tank Conditions..............................................................................10

Temperature Control..........................................................................................11

Materials of Construction...................................................................................13

Further Safety Concerns....................................................................................13

Gathering Pipeline Design.....................................................................................15

Introduction........................................................................................................15

Pipeline Parameters...........................................................................................16

Suction Side Factors..........................................................................................17

Discharge Side Factors......................................................................................19

Pump Selection..................................................................................................21

Preventing Heat Loss.........................................................................................23

Heat Exchanger Design.........................................................................................25

Introduction........................................................................................................25

Type of Shell......................................................................................................25

Fluid Properties..................................................................................................26

Type of Head.....................................................................................................27

Fluid Allocation...................................................................................................27

Calculation- Tube Side.......................................................................................28

Calculation- Shell Side.......................................................................................31

Overall Heat Transfer Coefficient (Uo)................................................................32

Pressure Drop Tube Side...................................................................................33

Pressure Drop Shell Side...................................................................................33

Design Justifications..........................................................................................34

Report Conclusion.................................................................................................34

Equipment Specification Sheets............................................................................34

...............................................................................................................................35

Introduction

For the oil terminal a number of requirements have been set to achieve in this

project. The design as shown in the cover image, is to have 6 crude oil storage tanks

with each a capacity of 375,000m3. This means around 318,750 tonnes of oil can be

stored per tank (using the average Alvheim field blend density (Norway) which will be

the standard used during this project). This oil is to be either stored to be sent for

refining at the Gdansk refinery or to be further transported by tanker as raw crude oil.

For this project, the process from delivery of crude oil from the Naftoport to the

storage tanks; the vapour recovery and the returning of the oil to either the Naftoport

or sending the oil to the Gdansk refinery will be explored and detailed designs

provided. For the purpose of this technical section, the design of the tank will be

explored, including type of tank, tank parameters, maintaining storage conditions,

materials of construction, and safety concerns associated with the tank design

chosen. The fluid mechanics will be explored in the moving of oil to the storage tank

along the gathering lines associated with the process. This includes type of piping,

piping parameters, pump design, control valve choice and the factors that must be

overcome such as friction and safety concerns. Finally this report will look at a heat

exchanger design that is required to prepare oil for the required storage conditions.

This will involve the type of exchanger, specification of the exchanger and its

justification.

Process Overview

The drawing presented below is the area of each individual tank process associated

with this report:

Overview of Process Employed

In the above drawing, the lines L3-1; L4-1; L5-1 are included in the design report with

the design of Pump PU2-1; Tank T1-1; and the Shell and Tube Heat Exchanger

HX4-1 also included. It can be seen that the main line splits into two. With one line

as part of the above process and the other goes to an identical line. Therefore this

design is a standard used for all the 6 oil storage tanks in the facility. The main line

that splits has a flowrate of 5000m3/h and splits into half between each line to the

tanks. Giving a flowrate of 2500m3/h to each tank at peak times.

Overall Mass Balance for the Process

A spreadsheet showing the mass balance is shown below. This gives the mass

flowrate and gives the compositions expected for individual components in the oil.

This helps to predict the flash off expected and therefore the load the VRU (vapour

recovery system) can expect. Whilst the mass going in is shown, the expected mass

going out via the VPU is not. This will be detailed in another section in this overall

report.

Mass Balance

Volumetric Flowrate 2500 m^3/h

Overall (Kg/s) 595.00

Component Mass % Per Tank Input Feed (Kg/s)

Methane 6 35.7

Ethane 4 23.8

Propane 5 29.75

Butane 9 53.55

Heptane 11 65.45

Hexane 17 101.15

Heavy Oil 39 232.05

Salt Water 5 29.75

Nitrogen 1.1 6.545

CO2 1.4 8.33

H2S 1.5 8.925

Total 100 595.00

Overall Energy Balance

In this process the overall energy balance is found from the gain in energy from the

heat recovery process and the energy loss as a result of pumping fluid through the

gathering lines. This is outlined below in a spreadsheet:

This shows an overall positive gain of energy because of the energy recovery

system used.

Tank Design

For the design of the tank a number of factors had to be considered. These include:

- Type of Tank

- Tank Dimensions

- Maintaining Tank Conditions

- Materials of Construction

- Safety Considerations

In this order the tank design will be looked at in brief sections.

Energy Balance

Equiptment Energy Consumption/Gain (per unit, J/kg)) Total Energy from Process (J/kg)Pump -418.2 -836.4Heat Exchanger 53,400 3,204,000

Total: 3,203,164

where consumption makes for a negative figure

Number of Heat Exchangers 60Number of Pumps 2

Type of TankThere are two types of tanks used in industry to store liquids:

- Fixed Head Tank

- Floating Roof Tank

Each have their pros and cons and therefore the liquid behaviour and local seasonal

variables must be explored.

Raw Crude Oil in any case will experience flash off of low carbon number

hydrocarbons which form a vapour above the liquid surface. This means the vapour

pressure of these hydrocarbons is much higher and (and associated gases also

entrained in the oil) will begin to form and typically grow the vapour film unless high

pressures/low temperatures are used. Using either of these options would increase

safety concerns and wouldn’t be applicable for oil storage (crude oil is usually stored

at atmospheric pressure). The best way to prevent this using floating roof tanks as

the weight of the roof will stop any vapour forming and sit on top of the liquid. This is

an advantage over the fixed roof tank as pressure relief valves and inert gas supply

would be necessary to prevent pressure build-up which adds more complications to

the design. (Engineering, 2014, pp. 4-5)

Another concern in the design is the seasonal weather conditions of Gdansk.

Gdansk (which is located in the north of Poland on the Baltic coast) is subject to a

temperate climate which means warm summers and very cold winters. This creates

further concern particularly in winter. Winters are subject to seasonal snowfalls which

can result in large build-up of snow depth. This would risk becoming a hazard with

the floating roof tank as the weight could cause the roof to sink below the stored

crude oil allowing for a spill if this snow is not removed. This would require a lot of

man hours to remove and it is very hard to predict when this work would be needed.

Whilst a floating roof is better for dealing with flash off it makes more sense in this

location for fixed roof tanks to be employed. For this reason the fixed roof tank is the

preferred option for the design in this report as it is much easier to control and

predict the behaviour of stored liquid and its vapour output than seasonal weather

forecasting. (Engineering, 2014, p. 5)

Tank Dimensions The aim during the building of tanks is to a) ensure they are safe to use (looking at

the local area and what events happen) and b) to build the most economical tank.

The best way to do this is using mathematical optimisation taking the known

requirement of the volume being 375,000m3. Optimisation can be used to find the

smallest surface area to fulfil the set volume. This has numerous advantages such

as reduced construction cost (less material required); reduces the environmental

effect (less surface area for the wind loading factor); and also reduces the area for

force from the liquid-vapour pressure in the tank to affect. The optimisation method

uses differentiation to find the minimum turning point (at which the least surface

exists at). The following method was employed:

First the equations for both surface area and volume are obtained:

Volume (Cylinder )=π r2h

Surface Area (Cylinder )=2 πrh+2π r2

The volume is equal to 375,000m3 so:

π r2h=375,000

So it can be assumed:

h=375,000π r2

This can be substituted into the Surface Area equation to remove the height variable

making for a one variable equation. Hence this can be solved to find the radius:

Surface Area=2πr (375,000π r2 )+2π r2

Simplification finds:

Surface Area=(750,000r )+2π r2

Beginning the differentiation leads to the following:

S A '=4πr−( 1,500,000r2 )Making SA’ equal to zero:

4 πr=( 1,500,000r2 )Multiplying throughout by r2:

4 π r3=1,500,000

Making the radius (at the minimum turning point):

r=49.24m

Substituting this back into the volume formula to get height:

h= 375,000π (49.24 )2

Getting a height of 49.23m

Surface Area is therefore:

Surface Area=2π (49.24 ) (49.23 )+2 π (49.24 )2

Giving a surface area of 30,465.05m2

Maintaining Tank ConditionsIt has been proposed that the following conditions are maintained. The tank is to be

kept at atmospheric pressure and 10°C. Atmospheric pressure means there is no

pressurised vessel explosion potential, which in turn improves safety. The extra

pressure would require an increased vessel wall thickness which would incur

increased building cost. Keeping the tank at 10°C reduces the vapour pressure of

the crude blend, which in turn reduces the amount of flash off of light hydrocarbons

and keeps the oil in mix to be sent to its final destination as a raw material.

To maintain pressure and temperature:

- Making use of a pressure relief valve (which sends vapour to a flare or a

Vapour Recovery System (VPU))

- Inert gas supply (here nitrogen will be used)

- Using a temperature control system

Temperature ControlThere are a few designs that could be employed for this requirement which include:

- External Heat Exchange System

- Internal Tank Coils

- Jacketed Vessel

An external heat exchange system would require both piping/pumps and a heat

exchanger such as a shell and tube. There are many options on using this design

which as a result could make it rather complicated. The required pumping power and

pipes would have an extra cost making this non-economical in both the short and

long term. The small temperature difference would make the exchanger inefficient

and with a high flowrate this would make it even more difficult to design (the tank can

contain large amounts of fluid so the required volumetric flowrate to maintain this

temperature would be very high). The other two design options will be considered.

(Richardson, 1999, p. 505)

Internal Coils are another option. For heating, it would be required that the coils be

situated at the bottom (much like a kettle) and for cooling at the top (which is all to do

with heat rising and cold sinking so convection will occur). As both heating and

cooling abilities will be needed, this would increase the area of coils required, with

the overall cost to install and maintain these coils from factors such as corrosion;

increasing the overall cost. However the heat transfer is very good with coils (as

there is direct contact between liquid and heat transfer surface) so it would not be

much trouble to ensure conditions are maintained. There are many options available

for choosing the fluids inside the coil (whether they are heating or cooling) and coils

are widely used in industry so sourcing need not be too much hassle. (Perry, 2008,

p. Section 11 p21)

The other option is the Jacketed Vessel design. Jacketed vessels are simply vessels

where within the walls of the vessel are heating/cooling mediums which can reduce

the effect of local temperatures on tank fluid and change temperature in the tank.

This makes the jacketed vessel option very desirable in dealing with the safety

concerns brought about by Gdansk’s local seasons. However jacketed vessels are

not very effective with vessels on the scale that is considered in this design as heat

transfer area would be very limited (to the surface area of the inside of the tank).

(Perry, 2008, p. Section 11 p22)

For this design internal coils have been chosen to perform the maintaining of

temperature in the tank. Below is further design info:

- For large tanks, the hairpin design is typically used as it is easily erected in

the field where it is more economical to construct a storage tank of the scale

considered in this project.

- For heating this is built at the bottom of the tank and for cooling the top due to

heat rising and cold sinking.

- For the size of vessel, it is best to agitate the fluid to improve heat transfer

(note: on the downside, this will encourage more flash off)

- When heating with steam on the coil side, transfer coefficients range between

200 and 400 W/m2 K (Towler, 2009, p. 820)

- When cooling with water, transfer coefficients are between 60-300W/m2 K

(Towler, 2009, p. 819)

- A drawing of the design is shown below:

(Perry, 2008, p. Section 11 p21)

Materials of ConstructionThe choice of building material is important as it must withstand a) pressure of tank

contents and b) local environment variables. For the tank type chosen (fixed roof

tank) there is a chance pressure could build, such as the event of a pressure relief

valve failing. Because the tanks are on the scale they are, it can take hours to empty

the tanks contents so it is important this material can withstand any increased

pressure from within. Otherwise the pressure will be atmospheric so should not

present too many problems but consideration is needed in case of the above event.

Gdansk being close to the coast is subject to storms and high winds. This creates a

factor known as wind loading. From earlier the surface area of the tank is minimised

seriously limiting this effect but this factor is still worth being respectful of. Asides

from this not much else affects Gdansk that would present construction issues.

(Anon., n.d.)

For this design, steel-reinforced concrete will be considered due to its availability,

cost effectiveness and strength. Specifically deformed steel bars will be used. These

are by far the most commonly used and will be post-tensioned. Post-tensioning

means that the concrete will be tensioned on site and this is simply because the tank

is such a large size that it would not be economical to do so in the factory and then

transport such large and numerous pieces to the site. Whilst this requires some

special hardware at the end of each steel tendon to anchor them to the concrete, this

is not a big enough issue to consider using pre-tensioned concrete instead. (Perry,

2008, pp. 10-140) (Suchorski, 2000 (Reapproved 2006))

Further Safety ConcernsFurther safety concerns are standard when it comes to designing storage vessels to

contain hydrocarbons:

- Risk of Spills

- Risk of Fire

- Risk of Explosion

- Risk of Corrosion

Risk of spills are a danger to both the installation, the environment and company

reputation. Spills can cost billions in damages as seen with the recent spill in the Gulf

of Mexico by British Petroleum. Whilst the spill by BP is a slightly different scenario it

does demonstrate the affect oil can have if not under close control. In the event of an

oil tank spill the liquid needs to be contained to minimise its affects to the

surrounding area. A good method to do this is designing a bund area or a pool in

which the spilled oil may collect in. It is a simple idea but very effective on the basis it

is designed well and proper response instructions are followed. The Australian

government environmental department suggests the bund area should be able to

contain 133% of the total volume of the vessel requiring bunding (SA, n.d., p. 4). As

vapour is likely in the event of a spill, quick action is required to reduce

environmental impact of escaping gases. This is because methane in a well-known

greenhouse gas which is twenty-times worse than its counterpart carbon dioxide

(US, n.d.). A potential solution could be for the spilled oil to be collected in drains

surrounding to tank and stored underground, this in turn will reduce escaping gases.

The risk of fire is a big problem with crude oil. When burned a thick black smoke

cloud forms which in large quantities can cause health concerns in the local area.

Due to the size of the installation at Gdansk this becomes very important. A similar

accident was seen in London at the Buncefield oil terminal which produces aviation

fuel for neighbouring Gatwick airport (UK, n.d.). The fire caused a cloud of soot

particles to form over London which could be seen from space. Therefore it essential

that freighting capabilities are available to the installation. One of the most important

parts of any hydrocarbon combusting is the requirement of oxygen which if cut off,

stops the reactions occurring. In this case a foam injection system can be employed

to help fight crude oil fires. The foam helps cut oxygen off from the oil and being

flame retardant forms an unreactive layer on top of the oil killing and preventing any

flames. (Engineering, 2014, p. 5)

Risk of explosion is another issue requiring consideration. As a fixed roof tank is the

design of choice it comes with an added concern, hydrocarbon vapour formation.

Light hydrocarbons are very susceptible to sudden explosions in the right conditions.

These conditions include being in the vapour phase, in which more surface area in

available for reactions to occur. This results in unstoppable chain reactions. Not only

are these explosions catastrophic to any installation (as fire fighting systems such as

foam injection can be destroyed) but the following fires can add to problems. Whilst

explosions are dangerous, they are easily prevented. Purging the tanks with inert

gases to mix in with hydrocarbons prevents oxygen creating an explosive

atmosphere. Also ignition sources can be eliminated. As hydrocarbons are well-

known insulators, electrical potential can grow inside a tank which with time can

create a spark. By earthing a tank, this potential can be removed. Strict rules are

also needed to prevent workers bringing any potential ignition sources near the tank

area during cleaning for example. (Engineering, 2014, pp. 3-5)

Corrosion is the final risk that should be considered. Since internal heating coils are

being considered for this design it is important to look at what effect the fluid may

have. Since raw crude oil is being stored, this brings the issue of brine and its

corrosive properties. Water is known to corrode metals but this is generally a very

slow process. However alkali metallic salt solutions speed this process up by a large

factor. It has been proposed that water drains are used at the bottom of the vessels.

Water is much denser than oil so sinks and forms at the bottom of tanks when

stationary. This water or brine requires removal to prevent corrosion to the coils

situated at the bottom of the tank. Using drains with level control systems (which can

detect how high water level is due to differing electrical conductivities) can eliminate

this and also create a more desirable crude which has reduced water content

therefore requiring less treating further downstream. (Engineering, 2014, pp. 3-5)

Gathering Pipeline Design

IntroductionThe naftoport in Gdansk can have up to 4 tankers docked at any one time. From this

it is estimated that a maximum flowrate of 15,000 m3/h can be a possibility in supply

for the storage installation. This is split into three transmission line (each carrying

5,000 m3/h) and each transmission line feeds two tanks (as there are 6 crude

storage tanks) so a flowrate of 2,500 m3/h will be considered in this design.

Furthermore to this the crude oil must be taken to a height of 32.2m so extra head

will be required of the pump. At this stage it’s clear to see a large centrifugal pump

will be required. A balance of both flowrate and head needs to be found so to make

this more economical the line will be further split in two giving two input gathering

lines per tank. This means flowrate in each line will be a maximum of 1250m3/h

reducing the velocity and allowing for a more economical pump to be employed.

As part of the design the following factors will be looked at:

- Pipeline Parameters

- Suction Side Factors (Friction and NPSH)

- Discharge Side Factors (Friction and Head)

- Pump Selection

- Preventing Heat Loss

Pipeline ParametersA rule of thumb used by engineers estimates the optimum inside diameter of

pipeline. On top of this gathering lines for oil range from 8-12” in diameter. The

equation used for finding the optimum inside diameter is shown:

di=0.33(MFρ )

0.5

Where;

- MF is the mass flowrate in kg/s

- ρ is the density of the fluid in kg/m3

(Towler, 2009, p. 260)

From this equation the optimum inside diameter was found to be about 195mm.

From this the standard pipe DN300 schedule 20 was chosen (Blue Scope Pipeline

Supplies, 2008). The chosen diameter is larger than the optimum however this was

to keep fluid velocity low enough not to cause significant pressure drops (as the

crude is rather viscous so incurs more friction). This has a wall thickness of 6.4mm.

However it should be considered that crude oil is particularly corrosive so for this

application the inside of the tube will be lined with rubber (with the same thickness as

the tube wall). The outside diameter is 323.9mm. The inside diameter is found by

taking away the wall thickness multiplied by four when the rubber lining is included.

This is demonstrated below:

di=323.9−4 (6.4 )

This gives an inside diameter of 298.3mm. From this pipe calculations can be made.

Furthermore to this schedule 20 piping for this pipe gives a pressure rating of 49.95

barg which is more than sufficient as transmission lines are expected to pump fluid at

a maintained ten bar giving a safety factor of just under 5.

It is important to calculate factors such as volumetric flowrate for later calculations on

the suction and discharge sides of the pump. So using the below formula we get:

Q=MFρ

Where:

- Q is volumetric flowrate

Volumetric flowrate for the whole line will be a maximum of 0.35m3/s.

Suction Side FactorsIn terms of the suction side of the pump it must be ensured that pressure drop is not

greater than the pressure supplied by the previous pump or vessel. If so it risks fluid

not making it to the suction point of the pump and the NPSH (Net Positive Suction

Head) being very low or negative. In this design, it is known that pressure in the

transmission line was held at 10 bar. However the last pump had to provide pressure

to overcome heat exchangers in the start of the gathering line and pump along a

relatively long pipeline to reach the gathering lines. The following factors are

considered:

- Pressure Drop on first heat exchanger is equal to 0.3 bar

- The length of pipeline from the end of the gathering line start to the pump is

57m

Pressure at the start of the gathering line is 4.625 bar. Once the gathering line

begins, 57m of pipeline must be crossed with a 90 degree junction and gate valve to

also overcome until the suction side of the pump is reached. Following this a

pressure drop of 1.07 bar was calculated allowing for a NPSH of 39.1m at maximum

flowrate. This is sufficient for the pump and eliminates the risk of cavitation.

To calculate pipeline friction losses first fluid velocity; Reynolds number; and relative

roughness must be calculated. They can be found with the following equations:

u=QA

Where:

- u is the velocity

- A is the cross sectional area of the pipe (A=π4

(di))

ℜ= ρudiμ

(Towler, 2009, p. 240)

Where:

- ρ is the density of the fluid (850kg/m3)

- Re is the Reynolds number (which is a dimensionless unit)

- μ is the viscosity of the fluid

RR= edi

(Towler, 2009, p. 240)

Where:

- RR is relative roughness

- e is absolute roughness of pipe material (rubber- 0.00015m)

So velocity is 5m/s; Reynolds number is 143645; and relative roughness is 0.0005.

Also pipe fittings and equipment drops must be accounted for. Along the suction side

there is a heat exchanger (with drop 0.3 bar); a 90° bend; a sudden constriction (as

the pipe reduces in diameter from the transmission line diameter); and a gate valve.

For the fittings, the equivalent length method will be used. For this data was used

from Coulson and Richardson’s Volume 6 for the values for pipe fittings (Towler,

2009, p. 243) . The table on the following page shows the fittings equivalent length:

So 55.5 multiplied by the inside diameter will give the equivalent length. This added

to the length of 57m of pipeline that already exists gives a total length of 73.56m.

This can now be applied into the pipe pressure loss equation.

Now the pipe pressure loss can be found:

∆ Pf=8∅ ( Ldi )( ρ u2

2 ) (Towler, 2009, p. 239)

Where:

- ∅ is the Stanton-Pannell friction factor found from the chart with Reynolds number and relative roughness (0.0024)The total pressure drop on this side plus the heat exchanger pressure drop was

found to be 1.07 bar. Now NPSH (Net Positive Head Suction) should be found for

pump specification, found by:

NPSH=( Pρg )+H−( Pfρg )−(Pvρg )

(Towler, 2009, p. 251)

Where:

- P is the pressure at the start of the gathering line

- g is the gravitational acceleration constant (9.81m2/s)

- H is the height of the fluid above the pump centreline (in this case zero)

- Pv is the vapour pressure of the liquid (found to be 29.6kPa)

This gives a NPSH value of 39.1m. This gives plenty of pressure at the suction

nozzle of the pump to prevent cavitation.

Discharge Side FactorsEverything on this side of the pump is working against the energy provided by the

pump. Performing an energy balance finds that without a pump present the fluid

would lack energy by 418 joules for every kilogram of fluid. From this energy balance

it is found that the pump must have a power rating of roughly 165kW (once taking

into account efficiency). To summarise the discharge pipe is a length of 40.2m with a

globe valve; three 900 bends; a set of parallel heat exchangers and pressure drop

from the tank inlet nozzle. This pipe also must climb vertically by 33.2m to reach the

tank inlet. The following page shows the pump data sheet produced for the design:

Like the suction side the pressure drop along the discharge is calculated in the

same way. First the pipe fittings must be accounted for. The table on the following

page shows this:

This gives a total length (equivalent + actual pipeline length) of 250.2m long.

Applying this to the pressure drop formula used in the suction side section and

adding on the pressure drop over the heat exchangers (60kPa), the total comes to

Fitting Equivalent Length90 degree bend 23d x3Globe Valve 450dTank Inlet 50dTotal 569d

3.22 bar. Now the head required of the pump must be found to define the duty. This

is found by using the equation:

Head Required=( Pfρg )−( Pρg )+Z

(Towler, 2009, p. 245)

Where:

- Z is the height that the fluid must overcome

This gives a head requirement of 40.38m.

Also the energy balance should be found at this point as this can be used later to

find the power rating of the pump. The equation is shown below:

W=g ∆ z+ ∆Pρ

−∆ Pfρ

Where:

- ∆z is the change in height (m)

- g is the gravitational constant (m2/s)

- ∆P is the difference in system pressures (N/m2)

- ∆Pf is the pressure loss due to friction (N/m2)

(Towler, 2009, p. 245)

This came out at a value of -418J/kg. The negative suggests that a pump is required.

Pump SelectionThe previous page shows the pump data sheet with the type of pump. This section

aims to justify the choice of pump. The pump chosen is a centrifugal magnetic drive

sealless pump. One of its main applications is in the petrochemical industry so its

design makes it ideal for the fluid being handled. It is capable of handling high

viscosity liquids and can provide the very high flowrate required in this design. More

importantly, it can provide the head needed to lift the crude oil into the tank inlet. The

suction specific speed shows a value greater than 9000 which according to Perry’s

Chemical Engineering Handbook makes this particular pump sitting at the average

for pump design. The sealless design is excellent in preventing spills and leaks

which is important in the transporting of any hydrocarbon. Also the magnetic drive

prevents any heat gain by the liquid as it travels through the pump casing reducing

the requirement of the following heat exchangers. This also in turn increases the

efficiency of the pump as less energy is lost as heat. Being centrifugal makes

maintenance of this pump relatively easy and the manufacturer has designed this

pump to require less maintenance over the course of time. Allowing for more

continuous operation. Its limitations are of course the inability to dry start so fluid is

required to be in the casing prior to start up. There is enough NPSH to allow for this

but in the event of pressure loss upstream this pump must be switched off or risk

breaking down. Finally extra power must be allowed for as if extra head is required

(say if the liquid temperature is lower, making the liquid more viscous) then it should

be available to ensure the process is continuous (M-Pumps, n.d.). More info on this

pump can be found at:

http://www.mpumps.it/?plg_cat_view=c&id=3&g=2

The performance curves of this model of pump defines what size the pump must be.

The following page shows the performance curves:

22

(M-Pumps, n.d.)

23

The performance curve shows the ideal pump is in the 300-500 type. This pump can handle the

flowrate and head plus a little extra if any complications present themselves. Now that the type of

pump is chosen, the power rating can be found. Taking large centrifugal pumps to have an

efficiency between 75-93% (Evans, n.d.)The lowest of this range shall be selected as the viscosity

of the oil is likely to lower the efficiency anyway. The equation to find the power rating is shown:

PR=W (MF) /efficiency

(Towler, 2009, p. 245)

Where:

- PR is power rating

The power rating was found to be 165kW.

Preventing Heat LossHeat Loss is considered here as winter temperatures in Gdansk can be particularly low. The

temperature of the fluid is important as the lower the fluid is, the more viscous it becomes and

therefore the more resistant to flow it becomes. With oil this is of concern. The piping is already

lined with a thick layer of rubber. Rubber is a well-known insulator so will reduce the heat transfer

to the steel pipe wall. Steel of course is a good conductor so would pass heat easily. A calculation

has been made to show heat loss at average winter night time temperatures and calm conditions

(i.e. low or no wind where only natural convection occurs). Heat Loss can be found from the below

equation:

Q=UAo (Tco−Ta )

Where:

- Q is the energy of heat transferred in Watts

- U is the heat transfer coefficient

- Ao is the outside area of the pipe

- Tco is the temperature of the crude oil (30°C/303K)

- Ta is the temperature of the surrounding air

(Chopey, n.d.)

0

Average night time temperatures in Gdansk are -2.7°C (Ta= 270.3K). The outside area is found

with a simple calculation:

Ao=2πrL

Where:

- r is the radius of the pipe

- L is the length

In the calculation the aim was to find heat loss per meter of pipeline so L in this case was taken as

1. For these calculations the imperial system was used so there was unit conversion at the

beginning and end.

Now U is to be found:

U= 1

r 3r 1hi

+r 3 ln (r 2

r 1)

k 1+r 3 ln ¿¿¿

Where:

- r3 is the outside radius

- r2 is the radius to the steel wall

- r1 is the radius to the rubber wall (inside radius)

- hi is the inside heat transfer coefficient

- ho is the outside heat transfer coefficient

- k1/k2 are thermal conductivities with respect to their materials

K1 (natural rubber) was found to be 0.13W/m K (Toolbox, n.d.) and K2 (steel) 51W/m K

(Richerson, 2006)

(Chopey, n.d.)

1

A pipe drawing is shown below:

hi was found to be 150W/m k and ho was calculated finding both the heat transfer coefficients due

to convection and radiation. Overall this came to ho being 54BTUhf t 2

.

This gave a U value of 4.944BTU

hf t 2℉

After calculating Q and converting it to standard units it gave an answer of 71.36W per m of

pipeline. Since the pipeline prior to the heat exchanger is 60m this loss can become significant.

For this reason the pipeline will require insulation to reduce heat loss to the surroundings. It should

also be noted that this is assuming calm conditions which with a costal location it is unlikely the

loss of heat will be as low as calculated. Types of insulation for outdoor purposes include foam

rubbers, asbestos (not really used however), and polymers.

2

Heat Exchanger Design

IntroductionFor the design proposed, it is required that the feed be cooled down to ten degrees centigrade for

storage. However for transportation it is better for the oil to be heated for more laminar flow and

reduced resistance. So a heat exchanger has been designed to bring crude oil temperature from

30°C down to storage temperature (10°C). For designing the exchanger, a number of factors had

to be considered before getting into calculations:

- The type of shell

- The heat load and fluid properties

- The type of head to use

- Fluid Allocation (to which side fluids are on)

Type of ShellIn this design, the most economical design was chosen, this being the TEMA E-type shell 1:1

pass. This means both shell and tube side have one pass. This design is the most widely used

with heat exchangers in industry. Its downside is the pressure drop is typically larger than some

comparable designs. (Towler, 2009, p. 826)

Fluid PropertiesIn order to carry out the design calculations, the properties of both fluids should be collected at

both the inlet and outlet temperatures of the fluid respectively. This includes;

- Specific Heat Capacity (J. Burger, 1985)

- Viscosity (Jiskoot, n.d., p. For viscosity)

- Thermal Conductivity (S. K. Elam, 1989)

- Density (Jiskoot, n.d., p. For density)

- Fouling factors (Towler, 2009, p. 822)

The table on the following page shows the properties collected:

3

Proposed valuesT1 (°C) 30 t1 (°C) 5T2 (°C) 10 t2 (°C) 12.22927631Mass FR (kg/s) 5 6.5

Fluid PropertiesAlveheim Crude Oil (Shell Side- T1/T2) Inlet Mean Outlet Cooling Water (Tube Side- t1/t2) Inlet Mean OutletCp (kJ/kg k) 1.9200 4.1009 4.0860 4.0710(J/kg k) 1920.0000 4085.9500K (W/m k) 0.1250 0.5795 0.5978 0.6160p (kg/m^3) 839.3400 846.4850 853.6300 1000.0000 997.9000 995.8000v (m^2/s) 0.0000 0.0000 0.0000 0.0000 0.0000 0.0000µ (Ns/m^2) 0.00614 0.00978 0.01343 0.00146 0.00556 0.00965Ro/Ri (m^2 k/W) 0.0005 0.0010

Heat Load Calculating t2Q= MCp(T1-T2) Rearranging the heat load equation

Q (kW) 192 t2 12.2292763W 192000

This also shows the heat load and exit temperature of the cooling water stream (t2). It should be

noted that it was difficult to obtain a lot of the properties of the Alvheim Crude Oil because of the

number of components present. However the change in density and viscosity were possible to

estimate. Using “IP Part VII Density, Sediment and Water; Section 2 Continuous Density

Measurement September 1997; Formulae from Annex F "Correlation Equations"; API-ASTM-IP

(ISO 91-1) Petroleum Measurement; Tables\database\tech\notes\spreadsheet\Density Correction”

helped to find the density compensation for temperature for the crude oil. Also using the ASTM

D341 standard to estimate the kinematic viscosity at varying temperatures, helped find the values

of viscosity. These values are fortunately widely available for water. (Laboratory, n.d.)

Type of HeadThere are two head types typically used. This is the fixed head and floating head types. For this

design the log mean temperature difference was found to be relatively low, therefore there is not

much risk of thermal stress on the tubes. It is more economical to use the fixed head in this case

as it is much cheaper than the floating head design. (Towler, 2009, p. 826)

Fluid Allocation

4

Which side the fluid is allocated to can improve the overall economics of the design. The tube side

fluid is typically:

- The lowest flowrate

- The most corrosive liquid

- The most fouling liquid

Tube side typically is the most difficult at keeping pressure drops low. However tubes are much

more easily cleaned and replaced making the above ideals. Shell side fluid typically has the

following characteristics:

- Highest viscosity

- Highest flowrate

It is better to have the viscous liquid in the shell side as it tends to work better with lower flowrates

and the viscosity also improves the heat transfer of the shell side. As the diameter of the shell is

always greater than tube diameter it is easier to achieve a high flowrate without too much pressure

drop. From the fluid properties it is easy to see the cooling water is much more fouling so passing

this through the tubes is the best option. Also the water should be pumped at a flowrate close to

but not exceeding 4m/s to reduce the fouling. (Towler, 2009, p. 843)

Following this discussion the design can now be calculated and specified following any required

optimisation. For the calculation values of the overall heat transfer coefficient are typically between

60-300W/m2 K however as discovered after numerous trials the coefficient was found to be only

28.5W/m2 K. This may be due to the high fouling of the water and the low flowrates used.

Furthermore the log mean temperature for the optimised design was found to be 10.07°C using

the equation shown:

∆Tm=(T 1−t 2 )−(T 2−t 1)

ln (T 1−t 2)(T 2−t 1)

(Towler, 2009, pp. 838-839)

Where:

- T1 is the entry temperature of Hot Side (Crude Oil)

- T2 is the exit temp. of hot side

- t1 is the entry temp. of the cold side (Cooling Water)

- t2 is the exit temp. of the cold side

Applying these values the following formula can be used to find the trial area required:

5

Q=UA ∆Tm

Rearranged to:

A= QU ∆Tm

Giving an area of 1069m2. (Towler, 2009, p. 817)

Calculation- Tube SideFor tube side calculations, the parameters of the tube had to be found. From optimising the

design, it was found that a 10m length (exchangers are more efficient when longer), DN50 tube (2”

outside diameter), which gave a tube area of 1.57m2 (per tube, denoted as At). From this the

number of tubes can be found using:

Nt= AAt

Here Nt was found to be 426 tubes.

Now the cross sectional area if the tube inside is found so that fluid velocity and the factors

affecting heat transfer can be found (these factors include Reynolds number and finally Nusselts

number).

A=π4di

Where:

- di is the inside diameter (here is 45.8mm)

This gives an area of 0.00165m2. From this the velocity can be found but first the volumetric

flowrate needs to be derived. This is simply the mass flowrate (6.5kg/s) divided by the density

(average density is 846.5kg/m3). This gives a volumetric flowrate of 0.0065m3/s. Velocity is found

with the equation:

V=VFA

Where:

6

- VF is volumetric flowrate

- A is the cross sectional area (found in the previous calculation)

- V is the velocity of the fluid

Here the velocity of the water was found to be 3.95m/s which is close to but not exceeding 4m/s

as required.

Reynolds Number is found now by the equation:

ℜ=Vρdiμ

(Sieder, 1936)

Where μ is the viscosity and ρ is the density.

Re was found to be 32451 which gives turbulent flow.

Prandtl’s number is found from:

Pr=μCpK

(Towler, 2009, p. 846)

Where:

- Cp is the specific heat capacity

- K is the thermal conductivity

Here the solution was 38.

Before calculating Nusselt’s number the viscosity correction factor is calculated. This involves first

finding the mean temperature of the tube wall. The solution to this is the average of the two mean

bulk fluid values found by taking the average of the inlet and outlet temperatures of each fluid

respectively. The mean temperature of the wall was found to be 15°C. The viscosity correction

factor is given by:

VC= μμw

(Towler, 2009, p. 846)

Where μw is the viscosity at the mean tube wall temperature for water.

7

Nusselt’s number is found from the solution of:

Nu=0.0023 (ℜ )0.8 (Pr❑)0.33( μμw )

0.14

(Dittus, 1930)

Here the solution was Nu = 391.5

Now the tube side heat transfer coefficient can be found:

hi= K (Nu )di

Here, hi was 4954W/m2 K

Calculation- Shell SideFirst thing to find is the tube pitch (Pt) and the bundle diameter (Db):

Pt=1.25do

Db=do( Nt0.249 )

0.453

(Kern, 1950)

Here, Pt= 62.5mm and Db= 1.8m

The pitch was chosen with 1.25do spacing as this is the recommended amount for triangular pitch

design. The bundle diameter equation is given by Sinnott as a general correlation he found. To

find the shell diameter the chart given by Coulson and Richardson’s volume 6 gives typical values

for bundle clearances for fixed head exchangers. In this case the value of 26mm was taken.

Ds=Db+0.026

Whilst 1.8m bundle diameter wasn’t on the chart, an assumption was made to 26mm being the

clearance (Towler, 2009, p. 831). This gives a shell diameter of 1.83m. Now the estimated area for

cross flow can be found:

as=DsB ¿)

(Towler, 2009, p. 855)

8

This gives a value of 0.134m2. Where:

- B is the baffle spacing (which is in this case is equal to Ds/5)

Now equivalent diameter must be found:

(Kern, 1950)

This gives a value of 0.036m. Also the volumetric flowrate is equal to 0.006m3/s. The mean

velocity is therefore:

u=VFas

Which gives a value of 0.045m/s

The modified Reynolds number must be found. It is modified as the velocity which is normally

used is only an average (as velocity varies throughout the shell cross-sectional area). Now the

value G replaces u where G is the product of volumetric flowrate and density and viscosity it

multiplied by the cross sectional area. This is shown below

ℜ=Gdeasµ

(Kern, 1950)

This gives a value of 210. It should be noted that this is laminar flow so the Nusselt’s number is

adjusted for the shell side to account for this. Prandtl’s number is as before giving a value of 150

for oil. Nusselt’s number is found by the equation below (under laminar conditions):

Nu=1.86 (RePr )0.33( deL )0.33

( μμw )

0.14

(Towler, 2009, p. 847)

9

Where:

- L is the tube length (10m)

This gave Nu=8.83

The outside heat transfer coefficient is found the same way as the inside except it is denoted ho.

The value obtained was 29.66W/m2 K where μw was 0.01017Ns/m2

The calculated overall heat transfer coefficient can now be found.

Overall Heat Transfer Coefficient (Uo)Found by the following equation:

1Uo

= 1hi ( dodi )+ Xw

Kw ( doDw )+ 1ho +Ri( dodi )+Ro

(Towler, 2009, p. 817)

Where:

- Xw is the tube wall thickness (3.38mm)

- Kw is the thermal conductivity of steel (51W/m K) (Richerson, 2006)

- dw is the tube wall mean diameter (0.0479m)

Ri/Ro are the fouling factors shown in the fluid properties table. This gave a Uo value of

28.87W/m2K.

It can be seen that the error between trial Uo and calculated Uo is very small and the trial is smaller

than the calculated making this an acceptable value. Now pressure drops must be found for both

sides of the exchanger.

Pressure Drop Tube SideFound by finding the solution to:

∆ Pf=Np(8Ф [ Ldi ][ μμw ]

0.14

+4) [ ρ u22 ] (Moody, 1944)

10

Where:

- Np is the number of tube passes (1)

- Ф is the friction factor found from the Stanton-Pannell friction chart

Ф is found by finding the relative roughness (which is the absolute roughness divided by the inside

diameter of the tube) against the Re number. Relative roughness is 0.001 (as absolute roughness

of steel is 0.000045m). This gives a Ф of 0.0025. The pressure drop is therefore 0.55 bar

(550kPa).

Pressure Drop Shell SideThis is more important as this determines the drop that has to be compensated for in the pump

design. It is found by using:

∆ Pf=8Ф (Nb+1 )[Dsde ] [ pu22 ] [ μμw ]

0.14

(Kern, 1950)

Where:

- Nb is the number of baffles (found by L/B) which is 27

As the flow is laminar this time the friction factor is independent of relative roughness. Therefore Ф

is 0.08. The pressure drop calculated is 0.0077 bar (77kPa).

Design JustificationsIt was found during the design of this exchanger that it was very difficult to have high flowrates

with acceptable pressure drops and number of tubes. Despite using the largest tubing size

available it simply wasn’t acceptable to flowrates any higher than stated in the specification sheet.

The number of tubes for this design came to 680. Higher flowrates seen tubes in the thousands

and tens of thousands which simply wouldn’t be economical. For this reason it has been decided

that to cool all the fluid down around 60 heat exchangers in parallel would be needed. It is

suggested that perhaps lowering this number and using the coil system in the tank to cool the fluid

may be a better option however the parallel heat exchangers will be considered. On the up side,

the pressure drop achieved shell side for the oil is very good at only 0.01 bar so even a large

number of parallel exchangers would not cause to large a problem. Also the heat recovered by the

water could be used elsewhere on plant to use as hot water or steam at reduced production cost.

11

Report Conclusion

To conclude, this report has covered the relevant topics involved with the process described. A

storage tank has been designed, taking the most economical method via an optimisation

procedure which minimised cost and improved safety. Maintenance of conditions was looked at

including relief valves and immersion coils and the justification for their choice. The type of tank

used and its justification, also its setbacks and how they are handled. Further safety concerns

such as dealing with and preventing disasters like oil spills and fires was explored. The pipeline

design was explored looking at pipe parameters like size, piping losses and lengths. What type of

pump was looked at, giving the specification required of the chosen pump and the justification

given to why it was chosen. A heat exchanger design was detailed showing the most economical

approach and specifying and backing up the choice of exchanger chosen.

Equipment Specification Sheets

(On following pages…)

Works Cited

12

Anon., n.d. climatemps. [Online]

Available at: http://www.gdansk.climatemps.com/

[Accessed 29 March 2015].

Bell, K., 1963. Final Report of the Co-operative Research Program on Shell and Tube Heat

Exchangers.

Blue Scope Pipeline Supplies, 2008. Technical Data Chart. [Online]

Available at:

http://www.bluescopepipelinesupplies.com.au/sites/default/files/Revised_Flip_Chart_BSPS_May_

2008.pdf


Chopey, N. P., n.d. Heat Transfer- Heat Loss on uninsulated surfaces. In: Handbook of Chemical

Engineering Calculations (2nd Edition). s.l.:s.n., pp. 7-12 - 7-14.

Dittus, F. a. B. L., 1930. Heat Transfer in automobile radiators of the tubular type. In: University of

California at Berkeley: Publications in Engineering 2, p. 443.

Engineering, D. o. E. I. a. C., 2014. In: Processing Fluids in Process Industries. s.l.:Forth Valley

College, pp. 3-7.

Evans, J., n.d. Pump ED 101. [Online]

Available at: http://www.pumped101.com/efficiency.pdf

[Accessed 21 Feburary 2015].

J. Burger, P. S. M. C., 1985. Thermal Methods of OIl Recovery. [Online]

Available at: https://books.google.co.uk/books?id=9nWgt-

F7TNMC&pg=PA48&lpg=PA48&dq=specific+heat+capacity+of+crude+oil&source=bl&ots=hPcuuB

N4f9&sig=GUJndukm2pKzGIk_kH88h4tSTpw&hl=en&sa=X&ei=_sQbVZmcDcy17gb-

uYHwAw&ved=0CD4Q6AEwCQ#v=onepage&q=specific%20heat%20capacity%2

[Accessed 01 April 2015].

Jiskoot, n.d. Calculate Density. [Online]

Available at: http://www.jiskoot.com/Calculations/Density_VCF/index.html


Kern, D., 1950. Process Heat Transfer. s.l.:McGraw-Hill.

Laboratory, M. H. T., n.d. Fluid Properties Calculator. [Online]

Available at: http://www.mhtl.uwaterloo.ca/old/onlinetools/airprop/airprop.html

13


Moody, L., 1944. Friction factors for pipe flow, p. 671.

M-Pumps, n.d. CN MAG-M. [Online]

Available at: http://www.mpumps.it/?plg_cat_view=c&id=3&g=2


Palen, J. a. T. J., 1969. Solution of shell side flow pressure drop and heat transfer by stream

analysis method, p. 53.

Perry, D. W. G. a. R. H., 2008. Perry's Chemical Engineering Handbook. 8 ed. s.l.:McGraw Hill

Companies.

Richardson, J. M. C. a. J. F., 1999. Chemical Engineering. 6 ed. s.l.:Elsevier.

Richerson, D. W., 2006. Modern Ceramic Engineering- Properties, Processing and Use in Design,

s.l.: CRC Press.

S. K. Elam, I. T. K. S. R. A. A., 1989. Thermal Conductivities of Crude Oils. In: Experimental

Thermal and Fluid Science (Volume 2). s.l.:s.n., pp. 1-6.

SA, E., n.d. EPA South Australia. [Online]

Available at: http://www.epa.sa.gov.au/xstd_files/Waste/Guideline/guide_bunding.pdf


Sieder, E. a. T. G., 1936. Heat Transfer and pressure drop of liquids in tubes. In: s.l.:s.n., p. 1429.

Stanton, T. a. P. J., 1914. Similarity of motion in relation to the surface friction of fluids, p. 199.

Suchorski, C. K. N. a. D. M., 2000 (Reapproved 2006). Reinforcement for Concrete- Materials and

Applications, s.l.: American Concrete Institute.

Toolbox, E., n.d. Thermal Conductivities of some common materials and gases. [Online]

Available at: http://www.engineeringtoolbox.com/thermal-conductivity-d_429.html


Towler, R. S. a. G., 2009. Chemical Engineering Design. 5 ed. s.l.:Elsevier.

Tubular Heat Exchangers Manufacturers Association, 1988. Standards of the Tubular Heat

Exchanger Manufacturers Association. 7 ed. New York: s.n.

UK, H., n.d. Buncefield Incident. [Online]

Available at: http://www.hse.gov.uk/comah/buncefield/buncefield-report.pdf

14


US, E., n.d. EPA US. [Online]

Available at: http://epa.gov/climatechange/ghgemissions/gases/ch4.html


15

individual report (final)

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