3. Equipment and Machinery department.pdf

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3.0 EQUIPMENT & MACHINERY DEPARTMENT

Introduction

The department of APS that attends to engineering of equipment and machines is divided in two sections: the first on deals with static equipment, the second one deals with dynamic machines.

Static Equipment 3.1 Vessels

The design and fabrication of the vessels of different size, shape and type, required in process plants, represent a complex task that needs, in an engineering company, the participation of the process engineer, the project engineer, the mechanical engineer and of course the manufacturer.

A vessel is mainly constituted by a container holding a fluid (liquid or gaseous) having pressure greater or lower than the ambient pressure, with a temperature usually different from the ambient one.

More in detail, the container is constituted by a shell, usually of cylindrical shape and by the heads which can be of different shapes.

Bending rolls are used to form the cylinder. The shell can be carried out by bending rolls pieces of different diameter.

The bending rolls are usually fabricated starting from plates which are calendared, in order to obtain the cylindrical shape, and then longitudinally welded. Shielded arc welding is maybe the most common method used in vessel fabrication.

Acetylene gas welding is common, especially for welding thin plates and small attachments.

Welding of vessels usually requires the application of several layers of weld material. It is necessary that great care shall be exercised in making such “multipass” welds and it shall be care of the engineering company’s inspectors to ascertain, among the other things, the proper application of the welding procedures.

Plates for the fabrication of vessels are usually ordered by thickness.

Plates made by a sheared-plate mill are preferred because they are of better quality ad are available in a greater number of size.

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The shell plates are carried out in a wide variety of thickness; lengths up to 800 inches are produced in the thinner plates. If both the circumference and length of a shell are greater than 195 inches, it becomes necessary to employ two or more plates.

Ideally the spherical-shaped pressure vessel should be the more suitable since it withstands higher pressures for a given metal thickness, but this solution is very expensive and it is used only in particular cases (spherical vessels are used in the storage of volatile liquids and gases, but they are not suitable for the construction of the usual process vessels). The cylindrical-shaped vessel is the next best design and it is used most extensively.

The heads generally have a rounded shape, obtained from plates by means of forging. The heads for such vessel may be flat, ellipsoidal (elliptical dished), dished, hemispherical or conical. Examples of such design are given in the following figure.

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The most used curvilinear bottom is the elliptic one with the semi axes with a 1/2 rate.

All kind of bottoms have a reinforced collar so that the welding to the shell is displaced compared to the tangent line; the height of the reinforced collar changes following the used norms. The presence of the reinforced collar avoid that the welding could be located in a strongly stressed area.

All pressure vessels require nozzles for the entering and leaving fluids, for drains, vents and manholes. These nozzles can be fabricated by means of pipe, pipe couplings, forged steel nozzles, cast steel, fabricated plates, or other suitable material in accordance with the codes requirements.

The use of threaded connections is generally avoided.

Cutting of holes into the shell of the vessel for the installation on nozzles weakens the vessel; therefore reinforcement around the nozzle has to be provided. Forged steel welding collar of lengths sufficient to protrude beyond vessel insulation, if any, are mass-produced by the manufacturers. The lengths of such nozzles have been standardized so that the vessel designer has, at his availability, various lengths according to the exigencies.

Welding type nozzles are also available on the market.

Connections for small drains and vents of one inch size and under are usually made with pipe couplings.

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3.1.1 Vessels Internals

Many process vessels require certain internal elements to effect modification in the fluids passing through the vessel. They include:

Agitators, to homogenize the fluid

Baffles to deviate the fluids

Distributors to convey the fluids into a certain area of the vessel

Demisters to separate from gas the water little drops

Grids, fixed on support rings, to contain catalyst and inhibitors, rashig rings and other packing

Bubble trays, used in the distillation towers, including, among the other accessories, downcomer clamping bars, adjustable weirs, removable cup and riser assembly

Bubble trays are a typical example of vessel internals which are purchased by manufacturer specialised in the fabrication of these items, while the vessel supplier provides the trays supports.

3.1.2 External accessories

External appurtenances are usually required. The main ones, are:

Angles for the support of insulation

Lugs and brackets to support platforms and ladders

Skirt to support vessels and towers

Nameplate in which the main data of the vessels are indicated

These items are supplied by the vessel manufacturer and welded to the vessels before shipping.

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3.1.3 Calculation Formulas

According to the various codes there are various theoretical formulas used for the calculation of the thickness of shell and of the heads.

The API ASME formula for the shell is the following:

s = P Ri / (100 S Z – 0,60 P) + c

Where:

P = design pressure in Kg/cm2

Ri = internal radius in mm

S = allowable working stress in Kg/mm2

Z = efficiency of longitudinal joints

c = corrosion allowance in mm

s = thickness in mm

for the elliptical heads, the formula is:

s = P Di / (200 S Z – 0,20 P) + c

3.1.4 Wind Action

The wind action is calculated as a distributed charge that produces a flexion to the equipment. This force exercised by the wind is proportional to the exposed section and to the velocity square.

Obviously, the wind action changes depending on the vessels characteristic; more the vessel is high and thin, more dangerous will be the wind action.

The solicitations are calculated following the construction criteria considering the equipment as a console table embedded to the base, and taking into account the combination, and wind concomitance, vessel weight, radial solicitations and vertical derived from pressures.

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3.1.5 Data Sheets

An example of vessel data sheet is given here below.

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3.2 Heat Exchangers

All the industrial plants require the supply or the removal of heat.

Therefore careful design and application of heat exchanger has to be given in the plant design.

The design of heat exchangers is seldom job of the process engineer of the project engineer and of the mechanical specialist: these three people have to work in strictly connection in order to supply to the exchanger manufacturer all the technical data, all the specifications, all the required exigencies to get final equipment in accordance with all presented requirements.

Many different types of heat exchangers are manufactured.

Specials designs can always be devised, which may be advantageous for a particular application. It is, however, preferable to use standard design or so-called “stocks” items wherever possible. Exchange manufacturers are able to produce certain exchanger types and sizes on an assembly-line production basis. Any deviation, of consequence, from these stock designs will require special operations and, therefore, increases in cost.

3.2.1 Double - Pipe Heat Exchanger

As the name implies, the double-pipe heat exchanger consists of two concentric pipes. One fluid flows in the internal pipe and the other in the annulus between the inside and outside pipes.

Such exchangers are most conveniently arranged in the form of hairpins.

The usual practical length of these hairpins is about twenty ft. Such exchangers are very easily made in practically any shop and continue to prove useful where very small surface requirements exist (approximately 100 sq ft).

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3.2.2 Double – Pipe Extended Surface Exchanger

One of the advantages of the double pipe exchanger is its low cost.

A desire to benefit from this low cost and at the same time to have the advantages of larger surfaces prompted the development of the extended surface tube.

Typical example of this exchanger is shown below.

Transverse fins are employed principally for cross-flow arrangements in either extended surface tube-and-shell exchangers or in air-cooled exchangers which are growing more popular in areas where water scarcity is a problem.

Use of fin tubes is particularly desiderable for gases, viscous liquids, or steams of small flow rate. Such fluids produce high resistances to heat flow, which are partially overcome by the larger effective area introduced by the fins.

For surface requirements below 1.000 sq ft the use of extended surface double-pipe exchangers often effects a considerable saving. Particular consideration should be given to such units for surfaces below 500 sq ft. The double-pipe sections can be arranged in the rows and connected in series.

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3.2.3 Shell and Tube Heat Exchanger

The most widely employed type of heat exchanger is the so called shell and tube.

When the exchange surface has to be enlarged for process reasons, the double-pipe exchangers could require excessive installation area; on the contrary the shell and tube design, provides an extended heat exchanger per unit volume of the occupied space.

Even if this type of heat exchanger differs in some details, there are many standard shell and tube exchangers on the industrial market.

3.2.3.1 Counterflow 1-1 exchangers

A counterflow exchanger type, where one fluid flows in a apposite direction from the other fluid, provides the most efficient exchanger of heat.

The so called 1-1 exchanger consists of one shell pass and one tube pass.

The 1-1 counterflow exchanger is used when the friction leakages, tubeside, must be kept to a minimum value and when the requests of temperature are such that real countercurrent flows have to be successfully used.

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3.2.3.2 Counterflow 1-2 and 2-4 exchangers

The 1-2 exchangers, presents an arrangement of one shell pass and two or more tube passes.

This type is the most common of all shell and tube exchangers.

The exchanger 2-4, two shell passes and four tube passes, is based on a configuration as shown in the figure.

Instead of 2-4 exchangers, we can use 1-2 exchangers connected in series, so avoiding the installation of the longitudinal baffle, necessary to separate such two shell passes, the design of which is rather complex; also the maintenance is facilitated using an exchanger in series instead of 2-4 exchanger that has to be used in any case for clean fluids.

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Classification and nomenclature of shell and tube exchangers

From the figures of various shell and tube exchangers previously shown, we can note that the main components of this equipment are:

The shell

The tube bundle

The buffles

The rear read of the shell

The distributors

The exchangers of shell and tube type are of three types:

Floating head

Fixed tubesheet

Tubes, U type

Floating Head

This type of shell and tube heat exchanger is suitable for all the services. The distributing box consists in a spherical cap bolted to the tubesheet. The assembly of the cup and of the tubesheet can float since they are free to move inside the exchanger.

Floating Head Exchanger

Fixed Tubesheet Exchanger

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Fixed Tubesheet

Both tubesheets are welded to the shell.

In case of notable dilation, a dilation joint has to be provided to absorb it.

U, tubes

This type of exchanger is used when the fluid, tube side, is not fouling, it is more simple and economic.

U, Tubes

The TEMA codes individualize the exchangers by means of three letters, the first points out the distributor, the second one the shell type, the third one indicates if we have U tubes or floating/fixed head.

AES

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• A = removable channel and cover

• E = one pass shell

• S = floating head

BKU

• B = bonnet integral cover

• K = kettle reboiler

• U = U, bundle

BEM


• E = one pass shell

• M = fixed tubesheet head

BFT


• F = two pass shell

• T = full through float head

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3.2.4 Heat Exchanger Design

The thermal calculation to determine the exchange surface is based on the following equation:

Q = S U ∆Tm

S = Q / U ∆Tm

Where

Q = heat transferred

S = required real surface based in the outside surface area of the tubes

∆t = indicates temperature difference between the hot fluid and the cold one

U = global coefficient of heat transfer

Now we will see the calculation of the temperature medium difference between two fluids with variable temperature.

We suppose that the two fluids are moving in counterflow.

T1

t2

t1

T2

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If we put in the formula Q = S U ∆T

∆Ta = T1 – t2 the resulting surface

S = Q / U ∆T a

should be smaller than the one obtained using ∆Tb = T2 – t1; it is clear that in the first case the exchanger should be under sized, due to the fact that in each section of the equipment, the temperature difference is lower than the one used in the surface calculation and therefore the required performances could not be reached; in the second case, on the contrary, the exchanger should result oversized, with the increasing of costs.

So we have to assume a medium value; man can demonstrate that this medium value is supplied from the logarithmic average of the temperature difference.

LMTD = ∆T a - ∆T b / ln (∆Ta : ∆T b)

Example: ∆T a = 60° C, ∆T b = 25° C

LMTD = 60 -25 /ln (60 : 25) = 39,98 ° C

This formula is valid only for the cases that the fluids are moving in equi flow and counter flow, in a double pipe exchanger; in a shell and tube exchanger the conditions are different in fact while a fluid passes in the shell only one time, the second fluid passes in the tubes at least two times. The problem is solved making use of a corrective coefficient Ft that depends from two factors:

R = T1- T2 / t2 – t1

P = t2 – t1/T2 - t1

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Where

t1 = temperature at the tubes inlet

t2 = temperature at the tubes outlet

T1 = temperature at shell inlet

T2 = temperature at shell outlet

The LMTD correction factors given by specific diagrams:

The design global coefficient of heat transfer based on outside surface area, U, is given by the formula:

U = 1/S:hiSi + rfi S:Si + St:SmK + rfo + 1:ho

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Where

S = outside surface area of tube

hi = inside coefficient

Si = inside surface area of tube

rfi = inside fouling factor

t = thickness of tube

Sm = surfaces average between S and Si

K = conductivity of tube

rfo = outside fouling factor

ho = outside film coefficient

Apart the calculation programs, the global coefficient of heat transfer can be given, in the first approximation from the following figures:

Without change of state

Hydrocarbon liq. – Hydrocarbon liq. 150-500 Kcal/hm2 ° C

Hydrocarbon liq. – Water. 250-500 Kcal/hm2 ° C

Water – Water 500-1000 Kcal/hm2 ° C

Hydrocarbon vap. – Hydrocarbon vap. 100-200 Kcal/hm2 ° C

Hydrocarbon vap. – Water 200-350 Kcal/hm2 ° C

Gas – Gas 200 – 300 Kcal/hm2 ° C

Condenser

NH3 Vap. – water 500 -600 Kcal/hm2 ° C

Hydrocarbon vap. – Hydrocarbon liq. 250 -350 Kcal/hm2 ° C

Hydrocarbon vap. – Water 350 -600 Kcal/hm2 ° C

Gas – Water 400 -500 Kcal/hm2 ° C

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Dynamic Machines

3.3 Pumps

All the industrial plants, would be inoperative were it not for the pumps which maintain the steady fluids flow trough the plant operation and activities.

Very different types and sizes of pumps are installed in an industrial plant, depending by the requirements (head and flowrate) of the process fluids they will treat.

An unfit, or quite wrong, selection of the pumps could cause serious problems during the plant operation.

Therefore the process engineer, the project engineer and the mechanical engineer, the specialist of the rotating machine, have to strictly cooperate in order to supply to the pump manufacturer all the data, all the information, all the specifications, which allow them the selection of the size and type, among the ones of their standard production, which most nearly fits the service in question. It is also very important that, before placing the order, the manufacturer’s recommendations are properly considered.

Basing a first classification of the pumps on the physical principles their working is based on, we can consider:

Centrifugal pumps;

Volumetric pumps, that includes reciprocating and screw pumps.

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3.3.1 Centrifugal Pumps

The centrifugal pumps are the most employed because of their adaptability to operative conditions as well as for their relative constructive simplicity and for the lack of pulsations in the flux.

Deferring from the others, the centrifugal pumps can be employed both for big and very big flow rates.

From a service point of view, centrifugal pumps can be divided in:

Process Pumps

These pumps handle process liquids of Refinery: as hydrocarbons, chemical compounds, solutions of compounds. The process pumps have to be usually designed in accordance with API 610 codes; these are particularly strict codes that impose hard construction with relevant increasing of the costs.

General Service Pumps

These pumps are employed and designed for services which do not require the special alloys and mechanical design features needed by high temperature or corrosive conditions.

They treat mainly water of different type as drinking waters, river waters, dirty waters and demineralised waters.

Chemical Pumps

These pumps treat the process fluids of chemical end petrochemical plants; pumps in this category are constructed of corrosion resistant materials, such as alloy steels, rubber-lined steel, but even plastic and glass.

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3.3.2 Characteristics of a Centrifugal Pump

A centrifugal pump working is based on increased static fluid pressure.

This action is based on the Bernoulli's principle.

The rotation of the pump impeller (produced by an electric motor or turbine) provides kinetic energy to the fluid as it is drawn in, from the impeller eye (centre) to the periphery, being forced outward the impeller vanes.

As the fluid exits the impeller with increased speed, it passes through a volute or a diffuser: these are two possible arrangements in the design of the centrifugal pumps.

In both arrangements, the flow passes through an increasing area that slows it causing the conversion of kinetic energy in potential pressure energy.

This conversion results in an increased pressure of the fluid downstream the pump.

It is important to fix that the centrifugal pump supplies a head to the fluid, and not a pressure. This head, measured in meters of water column, does not depend by the fluid; the final pressure depends by the specific gravity of the treated fluid and they are connected by the following:

Where: Δp is the differential pressure between suction and discharge;

γ is the specific gravity; H is the Head.

The effective energy the pump gives to the fluid depends by the flowrate, the head and the specific gravity of the fluid:

This is the Hydraulic Power.

Δp = γ Η

Ph = Q γ Η g

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The power needed by the pump to supply the hydraulic power to the fluid is the Pump Shaft Power: it takes into account all the losses due to friction, leakages an warming up of the fluid, resumed in the efficiency [η] of the pump. It can be obtained dividing the hydraulic power by the efficiency:

This power term is used to choose the nominal power of the driver; it shall be higher than the Psp.

A safety increase is adopted, according to API requirements (+10%, +15% or +25%, depending by the value of required power).

E.g.: if [Q]=mc/h; head expressed as ∆p (= γH) in kPa, and to

obtain the power in kW, the formula is:

10036,1270197,1

⋅⋅⋅⋅Δ⋅

=η

pQPsp

Differently, if [Q]=mc/s and [H]=m, with [γ]=kg/mc the power in kW

shall be:

ηγ⋅⋅⋅

=97,101HQPsp

The head a centrifugal pump can supply to the fluid is higher as the impeller diameter increases. On the same pump, the impeller can have a diameter variable between a minimum and a maximum value.

Due to high head requirements, that would lead to increase too much the diameter of the impeller (and consequently the exit speed of the fluid, with less efficiency), it is preferred to use two or more impellers (and volutes/diffusers): these are the so-called “multistages centrifugal pumps”. The head enhancement is equally divided between the stages that treat the same flowrate, in series.

Psp = (Q γ Η g) / η

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For high flowrates requirements, or when the flowrate vary greatly, 2 or more pumps can be arranged for parallel operation. The flowrate is equally divided between the pumps, each of them gives the same head to the treated part of fluid.

NPHS of a pump

The acronym NPSH means ”net positive suction head”, but apart from the definition that let always engineering students perplexed, we try now to catch the essence.

The NPHS of a pump practically is the resistance that pump offer to the fluid flux; it is an intrinsic pump characteristic that only the manufacturer can knows. This is the “requested” NPSH.

A more rigorous definition can be given only once we have described the cavitation phenomena.

At the impeller inlet, the fluid is accelerated, consequently it undergoes through a local pressure decrease: if the pressure goes under the vapour pressure value (proper of the fluid and varying with temperature only) the fluid starts flashing. Passing trough the impeller vanes, the pressure rises up and the bubbles previously generated in the fluid suddenly collapse, causing a mechanical pit on the impeller surface. If this working condition for the pump goes on, in not too much time the impeller of the pump will definitively break down. This phenomenon is well known as “Cavitation”.

NPSH Required can be defined as the minimum fluid pressure required at the pump suction so to avoid any risk of cavitation.

This is an intrinsic characteristic, proper of considered pump: it cannot be calculated, but only obtained as information from the pump manufacturer.

Different is the NPSH available.

The NPSH available is a characteristic depending by the hydraulic plant and completely independent by the pump itself. It can be calculated as the difference between the liquid pressure at inlet flange of the pump (Hs) and the vapour pressure of the liquid at operative temperature (Pv).

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e.g.:

Hs = Hb + H1 – Hd

Hb is the hydrostatic pressure at the top of liquid level;

H1 is the static head, measured from the centerline of the pump suction to the top of the liquid level: if the level is below the centerline of the pump it will be a negative number.

Hd is the pressure drop in the piping, fittings and valves, measured from the pipe inlet, to the pump inlet flange.

NPSHa = [Hb - (H1 + Hd)] – Pv = Hs - Pv

To assure a correct operating for the centrifugal pump it is necessary that:

NPSHa > NPSHr

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Example for calculation.

Normally, an available NPSH 0, 5 m bigger than the one requested is accepted.

The ESSO accepts also 0,3 m if the NPSH test, performed at manufacturer workshop, has been attended by the client.

Example: water pumping from an atmospheric tank located at a lower level than the pump one.

Example:

atmospheric pressure in loco = 9, 78 m

suction ground water level = 4 m

load loss between suction and foot valve = 0, 75 m

vapour pressure, at reference temperature = 0, 0143 m

The pump manufacturer requests NPSH equal to 3 ,5 m.

We shall therefore have:

NPHS available = 9,78 – (4 + 0,75 + 0,0143) ≡5 m

Being NPHS requested equal to 3, 5 we should not be afraid of cavitation phenomena.

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3.3.3 Specific Speed and Suction Specific Speed

Pump Specific Speed (ns) is a dimensionless numeric value which roughly defines the pump geometry and the shape of the pump characteristics. It is calculated basing on the flowrate, head and rotating speed values measured at the best efficiency point with the maximum possible diameter impeller for given pump:

As the specific speed increases, the ratio of the impeller outlet diameter to the inlet of the eye diameter decreases. This ration becomes 1.0 for an axial flow pump. Radial flow impellers develop head through centrifugal force, and are characterized by low flow and high head designs. Pump of higher specific speeds develop head partly by centrifugal force and partly by axial force. A higher specific speed indicates a pump design with head generation more by axial forces and less by centrifugal forces. An axial flow or propeller pump with a specific speed of 10,000 or greater generates its head exclusively through axial forces. Axial flow impellers are high flow and low head designs.

Specific speed (ns) identifies the approximate acceptable ratio of the impeller eye diameter (D1) to the impeller maximum diameter (D2) in designing an impeller:

• ns = 500 to 5000 D1/D2 > 1.5 radial flow pump

• ns = 5000 to 10000 D1/D2 < 1.5 mixed flow pump

• ns = 10000 to 15000 D1/D2 = 1 axial flow pump

nS = n (Q)1/2 / (Η)3/4

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These are the reference values when ns is calculated using US system of measure:

n = rpm; H = ft; Q = gpm.

If SI (n = rpm; H = m; Q = m3/s) is used, result ns can be converted in US system multiplying by 51.64.

Remark:

For double-suction pumps the Q value used in ns calculation shall be the total flowrate divided by 2. In analogy, for multistage pumps the H value shall be the total Head supplied divided by the number of the pump stages.

Suction Specific Speed (nss) may be used to determine what pump geometry - radial, mixed flow or axial - to use for maximum efficiency and prevent cavitation.

Suction Specific Speed is commonly used as a basis for estimating the safe operating range of a pump.

Suction Specific Speed is dimensionless and are expressed as:

where:

nss = Suction Specific Speed;

n = rpm;

Q = flowrate capacity (m3/h, l/s, m3/min, US gpm, British gpm) at Best Efficiency Point BEP;

NPSHa = available Net Positive Suction Head (m, ft).

As a rule of thumb the Specific Suction Speed should be below 8500 if calculated with US measure system (165 if calculated with SI) to avoid cavitation.

nss = n (Q)1/2 / (NPSHa)3/4

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As for Specific Speed, the conversion factor for the suction specific speed is 51,64.

3.3.4 Selection of Centrifugal Pump

Task of process engineer, or project engineer or mechanical specialist, belonging to an Engineering Company, is the one to select, and subsequently to verify, a proper pump which satisfies the requirements of the process and of the mechanical data sheet; the design and the construction of the pump is task of one of the many manufacturer Firms, operating on the market.

A centrifugal pump which operates, at a certain speed, provides for a certain flow rate with a certain head.

These parameters are linked each to other: their functional relationship is graphically represented by the characteristic curve given by the pump manufacturer, drawn on a coordinate axis system, where you can find the flowrate on the X axis (abscissa) and the head on the Y axis (ordinate). On the same coordinate system we can draw the curve representing the resistant characteristic of the hydraulic circuit, typically a parabolic shaped curve (remember that the hydraulic resistance of the circuit has a quadratic relationship with the flowrate).

The pump regulates itself so to work at the equilibrium condition, that is the intersection point between these two curves (the working point [P] of the centrifugal pump).

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If a decrease of the hydraulic resistance downstream the pump happens, related curve tends to be less slanting and the working point (intersection) moves along the curve towards the growing flowrates direction (point P1), viceversa, in case of hydraulic resistance increasing happens, the working point moves towards decreasing flowrates direction (point P2).

The indicated point P3, is the operating point at shut-off condition: zero flow for closed discharge valve.

It’s important to fix that on a same pumps several impellers with different diameters can be mounted: the manufacturer shall choose the diameter so that the working point shall be as near as possible to the maximum reachable efficiency (BEP: Best Efficiency Point).

The graph can be completed with the indication of efficiency and absorbed power curves:

You can see that for centrifugal pumps the absorbed power curve is continuously rising with growing flowrates: this means that we have the minimum absorbed power at shut-off and the maximum at the end of the curve.

Due to process requirements, there are two possible scenarios:

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pump starting with the discharge valve closed: in this case at the start up the pump absorbs the power measured at shut-off: less than power absorbed at the working point;

pump starting with the discharge valve open: in this case at the start up the pump absorbs the maximum power, since the starting point is at the end of the characteristic curve.

This means that the electric motor chosen as driver shall have a nominal power not only major than absorbed power at working point, but major than absorbed power at the end of the curve: otherwise driver shall not be enough powerful to assure the start up of the pump.

3.3.5 Centrifugal Pump Characteristic Curve.

Typical characteristic curves for a centrifugal pump: the design impeller has 407,5 mm diameter, intermediate between the minimum (369,6 mm) and the maximum (425,5 mm) diameter available for the considered pump.

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Mixed flow centrifugal pumps and axial flow or propeller pumps have considerably different characteristics as shown in figures below.

The head curve for a mixed flow pump is steeper than for a radial flow pump. The shut-off head is usually 150% to 200% of the design head, the brake horsepower remains fairly constant over the flow range.

For a typical axial flow pump, the head and brake horsepower both increase drastically near shutoff as shown below

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The distinction between the above three classes is not absolute, and there are many pumps with characteristics falling somewhere between the three.

3.3.6 Centrifugal Pumps: Classification

In following table are indicated the most important types of centrifugal pumps, usually installed in a refinery plant and described in the International reference most used standard (API Std. 610):

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Hereafter, you will find the schematic representation of the various types of pumps described in the table.

OVERHUNG:

OH1 type OH2 type

Foot-mounted single stage overhung Centerline-mounted single stage overhung

OH3 type OH4 type

Vertical in/line single Stage overhung (Flexible coupling)

Vertical in/line single stage overhung

(Rigid coupling)

OH5 type OH6 type

Vertical in/line single stage overhung High speed integral gear-driven

single stage overhung

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BETWEEN BEARINGS:

BB1 type

Axially split one/two-stage between bearings

BB2 type

Radially split one/two-stage between bearings

BB3 type

Axially split multistage

Between bearings

BB4 type

Single casing radially split multistage Between bearings

BB5 type

Double casing radially split multistage Between bearings (Barrel Pumps)

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VERTICAL:

VS1 type

Wet pit, vertically suspended, single-casing diffuser with discharge through the column

VS2 type

Wet pit, vertically suspended, single-casing volute with discharge through the column

VS3 type VS4 type

Wet pit, vertically suspended, single-casing axial flow with discharge through the column

Vertically suspended, single-casing volute line-shaft driven with separate discharge

VS5 type VS6 type

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Vertically suspended cantilever sump pumps Double casing diffuser vertically suspended

VS7 type

Double casing volute vertically suspended

3.3.7 Example of a Mechanical Data Sheet of a Centrifugal Pump

Here below is given an example of a centrifugal pump data sheet.

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3.3.8 Positive Displacement Pumps (Volumetric Pumps)

Positive displacement pumps working principle is based on the transfer of finite volumes of fluid from the suction side to the discharge side of the pump.

The flowrate is proportional to the pump speed (as for dynamic pumps), while the head does not depend by the flowrate and by the pump speed, but is highly variable, depending by the pressure in the hydraulic circuit downstream the pump.

A Positive Displacement Pump, unlike a Centrifugal Pump, will produce the same flow at a given rotation speed, no matter what the discharge pressure is: as consequence it cannot be operated against a closed discharge valve and, unlike centrifugal pumps, it does not have a shut-off referred head.

If a Positive Displacement Pump is allowed to operate against a closed valve, placed somewhere in the circuit downstream the pump, it will continue to produce flow which will increase the pressure in the discharge line as long as the power of the pump driver is enough to guarantee the pump working. This leads to the real risk of damage for the part of circuit downstream the pump and there installed equipment, or for the pump itself, or both.

As consequence, a relief valve shall always be installed on the discharge side of a positive displacement pump, so to prevent such dangers.

Following is a representation of typical characteristic curves for positive displacement

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pumps.

Q [mc/h]

H[m]

rated working point

resistant characteristiccurve

Q rated

H

Slippage

IdealReal

characteristic curve

You can see that over a certain head, the characteristic curve has a deviation form the ideal vertical curve. This is due to the fact that, as the discharge pressure increases for the higher resistance of external circuit, some amount of liquid will leak from the discharge of the pump back to the pump suction, reducing the effective flow rate of the pump.

The rate at which liquid leaks from the pump discharge to its suction is called slippage, and of course it implies a dramatic decrease of pump efficiency.

The positive displacement pumps can be divided in two main classes

reciprocating

rotary

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Typical application of positive displacement design are in the following types of pump:

rotary lobe pump;

progressing cavity pump (spiral pump);

gear pump;

piston pump;

diaphragm pump;

screw pump;

vane pump;

peristaltic.

Gear pump Internal gear pump Lobe pump

Diaphragm pump sectional drawing

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3.4 Compressors

The compressors are another very important category of machinery, always present in an oil refinery; they are operating machines directed to increase the pressure of gases and they are very similar to pumps: the main and evident difference is that pumps treat liquids, while compressors treat gases.

As for the pumps, different types of compressors are produced, depending by the physical principles their working is based on. A first basic classification is between:

Dynamic compressors (centrifugal and axial);

Volumetric or positive displacement compressors (reciprocating and rotary).

Since the specific volume of a gas varies greatly with its pressure and temperature, always the flow rate is indicated by 3 terms:

Volumetric flow rate at operating temperature;

Weight flow rate, not depending by pressure and temperature variations;

Volumetric flow rate @ normal conditions (atmospheric pressure and 0 ° C temperature).

The specific volume is linked to gas conditions (pressure and temperature) by the following:

PTRZVsp ⋅

⋅⋅= 510

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⎥⎥⎥

⎦

⎤

⎢⎢⎢

⎣

⎡

⎟⎟⎠

⎞⎜⎜⎝

⎛⋅=

−n

n

PPTT

1

1

212

Where: Z is the compressibility index of the gas; R is the gas constant (=8314/gas molecular weight).

When compressed, a gas naturally increases its temperature, the final temperature (T2) can be calculated by the formula:

Where:

temperatures are expressed in Kelvin degrees;

P2/P1 = β is the compression ratio (P1 and P2 are respectively the suction and discharge pressures);

pkk

nn η⋅−

=− 11 ; k is the specific heats ratio = X ⎟⎟

⎠

⎞⎜⎜⎝

⎛

v

p

cc

;

ηp is the polytropic efficiency of the compression.

The polytrophic efficiency is unitary when the compression can be considered as adiabatic (e.g. reciprocating compressors), otherwise, for polytropic compression the efficiency is less than 1 (e.g. in dynamic compressors).

Since we have a superior limit for the discharge temperature, due to commercial design requirements, the gas temperature at the end of compression is a very important reference value the manufacturer has to consider during the design phase of the equipment, so to fix the number of compression stages.

The single stage compression ratio can be calculated as nth root (for n stages) of the total.

[e.g. if β = 6, and we have n = 3 stages, then β1 = β2 = β3 = 817,163

≅ ].

Between two consecutive stages, the gas shall be cooled down by an inter-stage cooler. In case the cooling would cause the production of some condensate, it shall be necessary to separate it by a separator, installed downstream the cooler.

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3.4.1 Dynamic Compressor: Centrifugal and Axial

As for the centrifugal pumps, the working of centrifugal compressors is based on the Bernoulli principle: the rotation of an impeller provides kinetic energy to the gas, and this energy is then converted into pressure static energy in the volute.

Axial flow compressors produce a continuous flow of compressed gas, and have the benefits of high efficiencies and large mass flow capacity, particularly in relation to their cross-section.

However they require several rows of aerofoils to achieve large pressure rises, making them complex and expensive. The motion of the blades relatively to the fluid adds velocity or pressure or both to the fluid as it passes through the rotor.

The fluid velocity is increased through the rotor, and the stator converts kinetic energy to pressure energy.

Some diffusion also occurs in the rotor in most practical designs.

The increase in velocity of the fluid is primarily in the tangential direction (swirl): the stator removes this angular momentum.

Axial compressor can manage larger flowrate than centrifugal compressors, but in comparison they develop very lower head.

Typical working range in terms of flowrate (volumetric, inlet) is between 40.000 mc/h and 1 million mc/h; in terms of head, between 10 and 800 bar (as discharge pressure).

High compression ratio is available for axial compressors working as part of a gas turbine gas generator, for other aims, typically for refinery services; this requirement is preferable to be obtained by centrifugal compressors that reach same performances with smaller machine.

Both centrifugal and axial compressors are subject to a physical limit regarding the working, well known as surge phenomena: it imposes an inferior limit to the flowrate the equipment can manage safely.

As consequence, 0-100% regulation is available only by an intercooled partial recirculation of gas flowrate that of course involves a drastic downfall for the global efficiency of the machinery.

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Typical characteristic curves for centrifugal compressor:

CHARACTERISTIC CURVES FOR VARIOUS SPEEDS

WORKING POINT

The characteristic operating curve is fixed for a given compressor at constant speed. Variation of the suction pressure or system conditions will change the differential pressure developed by the machine.

The differential pressure will increase for any condition which causes increase suction inlet gas density.

The characteristic curve moves up due to any of following gas characteristic variations:

1. increased suction pressure;

2. increased molecular weight;

3. lower inlet temperature;

4. lower compressibility factor;

5. lower k [cp/cv] value.

Some process cannot fix gas composition and system conditions exactly and it is important to recognized the possible implications of changes in the suction conditions on the compressor performance, so to verify the compressor is well sized for the different possible working conditions.

EFFICIENCY CURVES FOR VARIOUS SPEEDS

RISING SPEED

SURGE LIMIT

RISING SPEED

mc/h

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Before examining it, we want to consider more deeply the surge phenomena: it is primed from a lessening of flowrate, due to compressor incorrect operating, or unstable process conditions (typical example is the restriction at the compressor inlet or discharge).

The lower inlet volume flow rate makes the pressure head decrease as consequence of turbulent dissipation of the velocity head of the gas leaving the impeller.

As flow is further reduced, the pressure developed, by the compressor tends to be lower than the pressure in the outer discharge line: this leads to a momentary flow reversal which reduces the discharge line pressure, and as it becomes less then the developed the flow takes again the proper direction, with rising discharge pressure, up to a certain point that cause a new start of the phenomena, that becomes a repeating cycle, unless proper correction action is taken.

Pressuring, surging is an unstable working condition, with rapid flow and discharge pressure pulsation, which produces high frequency reversal in axial thrust of the compressor shaft, varying its intensity from an audible rattle to a violent shock.

To avoid any risk of surging, it is important the rated working point of the compressor is enough remote form the surge line.

DEL

TA P D

SURGE REGIONC

A

VOLUME FLOW RATE

OPERATING REGION

B

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Several things can be done to prevent surge in centrifugal or axial compressors during working:

Reducing the resistance of the outlet stream (reducing the

compression ratio);

Increasing the flow through the compressor.

Both solutions have as consequence that the working point goes away from the surge limit, reaching a safety operating point.

Compressor Regulation

The operation of the compressor to meet or establish the desired point on the head-capacity system curve requires a control which can be variable speed or constant speed type.

Variable speed control

It is the most efficient method of controlling the capacity of a compressor.

Constant speed control

This system includes:

a. Inlet throttling: This is very common and simple way to vary the capacity when using constant speed driver. The gas density at the inlet of compressor is reduced by throttling action but this does not alter the system after the discharge. There is an energy loss during this operation but much less than the loss with the throttling on discharge side of the machine. The pumping capacity in terms of weight of flow is reduced in proportion to the density decrease whereas the volume pumped capacity remains the same (generally a butterfly valve is used for this service).

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b. For large machines the control is made by adjusting the orientation of the inlet distributor.

Inlet throttling with variable inlet guide vane.

This device is the most efficiency control for constant speed compressor.

The vanes provide inlet gas throttling and the pre-rotation of inlet gas with the consequence of a lower surge limit.

c. Discharge throttling

The absorbed power remains constant for this type of operation.

For dynamic compressors, we can define the following operating limits:

Minimum Operating Speed - the minimum speed for acceptable operation, below this value the compressor may be controlled to stop or go into an "Idle" condition.

Maximum Allowable Speed - the maximum operating speed for the compressor. Beyond this value stresses may rise above prescribed limits and rotor vibrations may increase rapidly. At speeds above this level the equipment will likely become very dangerous and shall be controlled to slower speeds.

Stonewall or Choke - occurs under one of following events. Typically for high speed equipment, as flow increases, the velocity of the gas/fluid can approach the gas/fluid's sonic speed somewhere within the compressor stage. This location may occur at the impeller inlet "throat" or at the vaned diffuser inlet "throat". In most cases, it is generally not detrimental to the compressor, but has impact on equipment efficiency. For low speed equipment, as flows increase, losses increase too, such that the pressure ratio drops to 1:1.

Different construction arrangements are available and used in compressors manufacture, varying with the required characteristics (in terms of flowrate and compression ratio) and with the gas handled. Growing up the compression ratio, the number of stages becomes higher, with the obvious need of interstage cooling features, and involves different arrangements for the compressor casing (axial split for low discharge pressure, radial split with higher values, and barrel configuration for the highest required discharge pressures).

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The applicable operative range for centrifugal compressor is very wide: the rated power can reach extremely high values (up to around 100 MW), and the type of driver changes, as well.

For power up to around 20 MW an electric motor still can be used (with inverter for speed regulation and start-up), for higher values, using of steam turbine or gas turbine as driver is preferable.

In case of electric motor drivers, it is important to consider not only the absorbed power at rated working point, but the starting torque needed at compressor start-up.

3.4.2 Reciprocating Compressor

Reciprocating Compressors

In the reciprocating compressors, the gas compression is consequent to the volume decrease due to the motion of the piston inside the cylinder.

The reciprocating compressor is mainly constituted of following components:

The crankshaft inside a frame: it receives rotating motion by an external driver;

The flywheel, placed between the driver and the crankshaft end, that regularizes the rotating motion;

The connecting rod, that converts the shaft rotating motion in alternative motion;

The crosshead that connects the connecting rod to the piston rod;

Distant piece, that houses the crosshead and the sliding piston rod;

The piston, that moves inside the Cylinder of the compressor and together with it realizes the variable volume for the gas;

The inlet and outlet valves;

Unloaders and clearance pockets (with fixed or variable volume), are special valves that control the percent of full load carried by the compressor at a given rotational speed of the driver.

Pulsation dampers (on suction and discharge), that mitigate the gas pressure pulsation consequent to the fact that the reciprocating compressor operates discrete volumes of gas.

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In a reciprocating compressor the number of cylinders is variable, depending on the flowrates and compression ratio requirements.

The position of the cylinders can be variously arranged: vertically, in “V” angle and horizontally.

Main characteristics of the reciprocating compressor is that it is a slow speed equipment: an important data to be considered with attention is the average piston linear speed.

Theoretically there is no superior limit to the pressure these type of compressor can assure, and for this reason they are typically used for extremely high pressure requirements, obtained with multistage (and multi-cylinder), intercooled, design.

Valves Clearance pocket

Cylinder Distant piece Piston

Piston Rod

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3.5 Heater and Furnaces

3.5.1 General

The primary function of a fired heater is to supply all heat required by the process stream to raise its temperature as need for the distillation, operation, catalytic reaction, etc.

A fired heater utilizes gaseous or liquid fuels often produced as a by-product.

Size and type of heaters vary considerably, depending upon the duty and required service.

3.5.2 Types

There are different types of heaters: cylindrical, single or double box with horizontal or vertical coils.

The most simple and common type is the vertical cylindrical (see fig. 1 here below), commonly used up to 40 MMkcal/h of duty.

The main parts of the furnace are:

Radiation section

This section is a vertical cylinder made by a steel plate shell, internally lined by castable refractory. Inside this cylinder the coils are vertical, located close to the refractory wall all around the internal perimeter.

The floor is made by steel plates, refractory lined, and supports for the burners.

The heat generated in the radiant tubes is mainly produced by direct radiation from the heating flame.

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Convection section

This section is located on the top of the radiant and the heat is exchanged by convection between the flue gas flowing to the stack and the horizontal tube bundle.

The first two or three rows face directly the radiant top and the burner flames, so may be considered as an extension of the radiant coil. This part is called the “shock zone” and it is always made by bare tubes.

All the other rows of convective usually finned or studded, to increase the heat flow.

Stack

The stack collects and discharges flue gases. It is made by steel plates internally lined by castable refractory.

A damper is installed to control the draft through the stack.

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Figure 1. vertical heater with radiant section

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Data and information to be supplied to the Manufacturer, to get relevant offer, are:

Physical and the chemical properties of charge and product as: density or API gravity, viscosity (at several temperatures), inlet and outlet molecular weight of vapours, coking characteristics of charge.

Operating conditions as: charge flow rate, inlet and outlet pressure and temperature, maximum allowable pressure drop, furnace duty (Kcal/h adsorbed).

Fuel: analysis of gas, heat content, viscosity (oil) density, temperature.

Desired efficiency.

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