PART – II

ROTARY EQUIPMENT

CHAPTER - 1

PROCESS PUMPS

1.1 Introduction

To enable the procurement of a pump, it is necessary to convey to the vendor all pertinent

process information. This process information is conveyed to the Specialist/Mechanical

group of TPPL in the form of Process group's standard Pump Data Sheet. Based on this

information, the Mechanical group prepares the pump specification sheet to be sent to the

vendors. The pump calculation sheet helps the process engineer in assembling the required

data in the pump data sheet. Following guidelines should be used in preparing pump data

sheets.

1.2 Types of Pumps

Pumps generally used in process industries are of centrifugal, reciprocating (piston, plunger

or diaphgram type) or rotary (gear, screw, lobe or others) types. A proper selection of the

type of pump to be used in a particular service has to be made.

1.3 Selection of Pump Type

Except for special purpose pumps for specific services, most of the process pumps are

made in standard sizes, and hence, the main problem is to select the size and type that most

nearly fits the service in question. Though the final selection of the pump will be made

close cooperation between the vendor and the Mechanical group, a preliminary selection of

the type of pump required is made by the process engineer. Selection of thc pump type can

be made on the basis of capacity-head requirements or fluid properties e.g. viscosity, solid

content and corrosive or erosive nature. Use Figure 1.1 for pump type selection based on

head-capacity requirements.

Centrifugal type pumps will handle liquids having viscosity up to about 200 cst. For higher

viscosities, specify screw, gear or reciprocating types. If solids are present in the stream to

be handled, the choice of pumps is further restricted. When solids are present, all internal

passages should have adequate dimensions. If solids are abrasive, close internal tolerances

between stationary and moving parts are undesirable. Generally, centrifugal type pump

(especially open impeller type) is the choice for handling liquids present in them.

1.4 Fluid Properties

Properties of the fluid to be pumped require careful attention of the process engineer. The

process engineer should give all the pertinent information asked for in the process data

sheet regarding the fluid characteristics. When solids or abrasive material are present,

specify the amount and size of particles. Similarly, if corrosive and toxic materials are

involved, specify the exact nature and concentration. Specify the pour point, congealing

point etc. accurately to determine the type of seal flush arrangement and any special

jacketing or heating arrangement required.

1.5 Suction Conditions and NPSH

Improper suction conditions are the largest source of pump troubles. Careful attention

should be given to NPSH (Net positive suction head). NPSH is the net remaining pressure

at the suction flange of the pump after all negative forces that restrict liquid from getting

into the pump are subtracted from all the positive forces that assist liquid in getting into the

pump. Two terms of NPSH are referred to:

NPSHA = Net Positive Suction Head available in the system expressed as meters of

liquid.

NPSHR = Net Positive Suction Head required by the pump expressed as meters of

liquid.

Calculate NPSHA in the system as follows:

NPSHA = Terminal Pressure in the Vessel + Height of Fluid above Pump Centre Line

(see note) + Atmospheric Pressure – (Vapour Pressure of Liquid + Friction

Loss in Suction Piping up to Pump Centre Line + Entrance and Exit Losses

From Vessel + Loss in Suction Filter + Loss in Control Vale, Exhaner etc.,

if any)

Note :

1. The height of liquid in the vessel should be taken to be at the vessel bottom

tangent line.

2. Pump centre line should be considered at 3/4 meter above the ground level.

Calculate NPSHA carefully considering all conditions i.e. start-up, original fill of lines,

winter operation, all control valve pressure drops, summer operation, piping conditions,

exit and entrance losses. Provide NPSHA at 1 meter over worse conceived NPSHR curve

by the manufacturer. In general, NPSHR of pumps should be considered as 3.0 meters,

though the exact requirement will be specified by the pump manufacturer. In case of

submersible pumps, the available submergence should be specified.

NPSH is a function of flow. It should, therfore, always be determined at design capacity

regardless of the total head required. It. is not uncommon for NPSHR curve to turn upward

at low flows. Low flow effects can often be amplified by selecting too large a pump which

forces recirculation within the pump.

When liquids at their bubble points are pumped from closed vessels, NPSHA is only the

static liquid head above the pump centre minus the friction losses in the suction piping. In

such cases, it is usual to elevate the vessels suitably (or sometimes cool the liquid before it

enters the pump) to get a margin of NPSHA over NPSHR. Sometimes, booster pumps (low

head, high capacity pumps which require low NPSH) are used ahead of main line pumps to

improve the NPSHA for the main pump which has high capacity high head requirements.

1.6 Pumping Capacity

Proper care should be taken in establishing the rated capacity of the pump. Normal flow

rate required by the process to be pumped should be estimated. A margin over this normal

should be added to take care against pump runout, internal leakage/ slippage in the pump

and contingencies. Following guidelines should be useful for specifying the capacity:

Type of Pump % margin to be added to normal flow

requirement

Centrifugal

3. Single stage

4. Multi stage

10

10

Reciprocating 20

Rotary 25

Pumps should not be oversized beyond the above recommended margins. Operating at

reduced capacities can result in increased bearing loads, reduction in pump efficiency and

sometimes increased NPSH requirement. When a future capacity increase in anticipated,

find out whether it can be accommodated by adding one more pump in future or using a

new wide impeller instead of specifying the pump for the future requirements and running

it at reduced flows now.

1.7 Discharge Conditions

The suction pressure and discharge pressure at the pump suction and discharge flanges

respectively should be estimated by drawing the system sketch in the pump calculation

sheet. All pressure drops in the system should be considered for various conditions – start

up, shutdown, fouled condition of piping/ equipment, start of run, end of run etc. The

difference between the discharge pressure and suction pressure so calculated is the

differential pressure or differential head when expressed in meters of liquid. The

differential head required for different conditions should be estimated and specified in the

pump process datasheet after taking about 10% margin over the normal differential head

requirement.

A centrifugal pump will not produce a higher pressure than its shut-off pressure even if the

discharge line is completely blocked. On the average, shut-off pressure is the maximum

suction pressure plus 1.2 times the differential pressure depending on the characteristic

curve of the pump. To avoid changes in the design at a later stage, the pump shut-off

pressure should be considered as maximum suction pressure plus 1.25 times the differential

pressure. All downstream equipment in the system e.g. piping, valves, exchangers, vessels

etc. should be designed for this shut-off pressure. During detailed engineering, this shut-off

pressure should be checked against the actual shut-off pressure specified by the vendor.

1.8 No. of Pumps and Sparing Philosophy

1.8.1 Pumps in Regular Use

Normally, only one pump is provided for regular use. In some cases, however, when

pumping capacities are very large and a single pump is not available, more than one

operating pump may be specified. Where substantial future capacity has to be catered for, it

is advisable to put additional pump in future (for which space should be kept in the system)

than to oversize the present pump.

1.8.2 Sparing Philosophy

A combination of factors including desired process reliability, service factor, service

factor, operating conditions, customers maintenance philosophy and cost etc. should be

considered to establish pump sparing philosophy. In general, 100% spare should be

provided for essential services and hot (above 200°C), high pressure, dirty, clogging and

polymerising service. For clean fluids and non-essential services, 50% spares can be

specified. If more than two pumps are running in parallel for the same service, a common

standby pump can be specified. Sometimes a common spare can be provided for two

compatible services.

Following criteria can be used for identifying the various service conditions :

•••• Essential Services

∼∼∼∼ Furnace and reactor charge pumps

∼∼∼∼ Product and reflux pumps

•••• Non-essential Services

∼∼∼∼ Product transfer pumps

∼∼∼∼ Chemical and additive injection pumps

∼∼∼∼ Product blending and circulation pumps

•••• Dirty/Clogging Services

∼∼∼∼ Pump feeding from tank bottoms

∼∼∼∼ Pumps taking suction from tower

∼∼∼∼ Bottoms slurry service

•••• Pump Reliability

∼∼∼∼ Pumps having mechanical seals are generally more reliable than pumps with

stuffing boxes.

∼∼∼∼ Rotary and positive displacement pumps are less reliable than centrifugal

pumps.

1.8.3 Warm-up Connection for Standby Pumps

Standby pumps under hot services (temperature above 150°C) should be always kept

heated up by providing a warm-up connection. Refer to Chapter 7 of Volume-IV of the

Process Engineering Manual titled P&ID Development for typical sketches of these warm-

up connections.

1.8.4 Pumps Running in Parallel

If two identical pumps are simultaneously working in parallel, they will deliver the sum of

the capacities to the same total head. The reasons for parallel pump operation can be:

1. Economy - two pumps and one standby can be more economical than one larger pump

and a standby

2. Two or three flow rates

3. Flow rates too large for a single pump to handle

Some cooling water and fire water pumps are typical examples.

1.9 Characteristic Curves of Centrifugal Pumps

Figure 1.2 is a typical characteristic curve of a centrifugal pump. The nature of head

capacity curve, efficiency and NPSH curves can be different depending upon the type of

pump. For a given system, the head-capacity curve of the system is superimposed on the

head-capacity curve of the pump and the point of intersection of the two curves is the duty

point at which the pump will operate at its best efficiency. The system head curve is a

function of the system static head and pressure head which are constant and the friction

head which varies with the flow.

Pump vendors will supply the characteristic curve for each pump. Following points should

be considered while finally accepting the pump:

• The slope of the head capacity curve should not be too steep if the pump delivers into a

distribution system, since a small change in flow will cause a large change in delivery

pressure.

• The slope should not be too flat if the pump capacity is to be controlled by throttling

discharge (manually or by a control valve).

• When pumps are to run in parallel, the head-capacity curve should be stable i.e. a curve

with the head constantly increasing as one approaches zero capacity. A dropping type

curve (where the shut-off head is less than the maximum head) gives an unstable

operation. This requirement is also valid for all pumps within battery limits having

standby.

• Manufacturers supply characteristic curves of pumps based on water as test fluid used in

the testing shop. While handling viscous fluids in actual practice, certain corrections

need to be made for viscosity to get the actual head, capacity and efficiency. Use Figure

1.3 for correction charts for pumping viscous fluids.

1.10 Choice of Driver

Electric motor is by far the most common drive for pumps in the process industry.

Occasionally, special considerations such as reliability of power, safety considerations and

criticality of service require turbine drives to be used. Sometimes, the plant utility balance

makes turbine drives necessary on large pumps. Steam turbines are effective for standby

pumps for a few larger, vital pumps such as charge pumps, cooling water pumps, unit

pump out pumps or flushing out pumps. These considerations have to be firmed up during

the design basis stage.

Steam turbines can be condensing or non-condensing type. The choice of non-condensing

or condensing operation is affected by the demand for the exhaust steam. Generally, pump

drives are not made condensing type without a rigorous review as the increased complexity

or condensing is rarely worth the small savings achieved in the utilities.

In some cases like offshore platforms, oil and gas processing terminals etc., gas turbines are

used as drivers for large pumps.

1.11 Utilities for Pumps

All pumps will need some utilities – electrical power for motor driven pumps ; steam for

steam-turbine driven pumps; cooling water for condensers of steam turbines; cooling water

for bearings, casing pedestals and packing; steam for seal quench, jacketing etc. Though

the exact requirements of different utilities needed by a pump will be specified by the pump

vendor, it is essential for a process engineer to have a fairly good idea about these

requirements so that they can be included in the utility summary sheet of the plant. These

requirements can later be firmed up after getting the vendor information. Following

guidelines can be used for estimating utilities required by the pump.

1.11.1 Power for Pump Driver

Compute the power consumption for the pump as shown in the pump calculation sheet.

Refer Figures 1.4, 1.5 and 1.6 for efficiencies of different types of pumps and Table 1.1 for

motor efficiencies. Take 120% of the power so computed for utility estimation.

1.11.2 Steam for Turbines

Steam requirements of a steam turbine can be estimated by using Figures 1.7 and 1.8.

Following instructions apply for the use of these figures.

1. Compute the delivered horse power (BHP) as shown in the pump calculation sheet.

2. Determine turbine efficiency from table below :

KW Overall Efficiency % 150 35

From 151 – 500 45 From 501 – 2000 55 From 2001 – 6000 65

Above 6000 70

3. Determine steam rate from Figures 1.7 and 1.8.

Table 1.1 : Electric Motors – Recommended Sizes and Efficiencies

Pump Requirement at Design Conditions

Probable Motor Rating

Motor Efficiency @ % of Full Load

0 – 0.5 1 81 82 82.5 0.51 - 0.75 1.5 67 73 75 0.75 – 1.00 2 75 78 80 1.01 – 2.00 3 75 79 80 2.01 – 4.00 5 81 83 84 4.01 – 6.00 7.5 75 80 81.5 6.01 – 8.00 10 80 84 85 8.01 – 12.0 15 81 85 86.5 12.1 – 16.0 20 80 83 86 16.1 – 20.0 25 83 86.5 88 20.1 – 26.1 30 83 86.5 88.5 26.2 – 34.8 40 85 88 88.5 34.9 – 43.5 50 80 85 87.5 43.6 – 52.2 60 84 88 89.5 52.3 – 65.2 75 87 89.5 90.5 65.3 – 87.0 100 84 89 91 87.1 – 114 125 85 89.5 91.5 115 – 136 150 86 89 91 137 – 182 200 88 91 92.5 183 – 227 250 90 92.5 93.5 228 – 273 300 90.5 93 94 274 – 318 350 91 93 94 319 – 364 400 91 93 93 365 – 409 450 91 93 93 410 – 455 500 91.5 93 93.5 456 – 545 600 93 94 94.5

Note : This table applies to totally enclosed motors (i.e. explosion proof)

• For non-condensing turbines and inlet steam pressure up to 18.5 kg/cm2a use Figure 1.7.

For all condensing turbines and non-condensing turbines with inlet system pressure up

to 43.5 kg/cm2a, use Figure 1.8.

• Determine heat theoretically available from saturated steam of the anticipated inlet and

exhaust pressure from graph 1 of Figure 1.7.

• Correct heat available for unsaturation or superheat as indicated on the figures.

• Determine steam rate from efficiency and heat available by use of graph 2 of Figure 1.7.

• Estimate steam requirements by multiplying the brake horse power and the steam rate

calculated from Figures 1.7 and 1.8.

1.11.3 Cooling Water Pumps

Cooling water may be required for a pump for cooling its bearing, pedestral, gland packing,

etc. This water can be fresh raw water, circulating cooling water or sea water. Generally,

the water after cooling the various parts of the pumps is routed to drain. Use the following

general guidelines for estimating water requirement for pump cooling :

Up to 120°C 0.5 m3/hr Up to 250°C 1.5 m3/hr Above 250°C 2.0 m3/hr

Some minor leakage of oil from the pump can be expected which will also be routed to

drain along with the above water. Take about 200 ppm of oil content in the outlet water to

drain for the purpose of effluent summary of the plant.

1.11.4 Steam for Quenching and Jacketing

Pumps handling high pour point products like Bitumen, LSHS, DMT, Phthalic Anhydride

and other such products are often steam jacketed. Specify steam conditions available in the

datasheet. The approximate requirement of steam for each such pump may be taken as 200

kg/hr.

1.12 Pump Sealing System

Process data sheets for pumps should indicate the type of sealing to be provided. Use the

following guidelines for selection of type of sealing:

Service Type of Sealing External Flushing Clean hydrocarbons and non-corrosive chemicals which will not solidify at ambient conditions at temperatures upto 300°C and pressure upto 40 kg/cm2a

Single Mechanical Seal

Required

Clean liquids which will solidify at ambient conditions, temperature upto 300°C and pressures upto 40 kg/cm2a.

Single Mechanical Seal

Required

Very corrosive, dirty or high pressure service

Double Mechanical Seal

Required

Slurries Packed Box Not Applicable Water clean hydrocarbons, non-corrosive chemicals at moderate pressures

Packed Box Not Applicable

Pumps taking suction from pumps where leakage can be redirected to the sumps

Packed Box Not Applicable

Typical mechanical seal arrangements for centrifugal pumps are shown in Figure 1.9.

Where external seal flushing is required, the process engineer should make suitable

provision to supply the flush liquid. API-610 provides recommendations on the seal

flushing plan for centrifugal pumps for different services (refer Figures 1.10 and 1.11)

The bubble temperature of the external flush fluid should be about 10°C higher than the

maximum operating temperature at operating pressure to avoid vapourisation of the fluid

and consequent vapour locking.

1.13 Material of Construction

In the petroleum and petrochemical industries, the selection of material of construction of a

pump is usually dictated by considerations of corrosion, erosion, personal safety and liquid

contamination. Tables 1.2 and 1.3 based on the recommendations of API-610 can serve as a

useful guide to the process engineer for selection of material of construction of pumps.

Obtain the recommendations of Specialist/ Mechanical group to verify these materials of

construction and specify these recommendations in the process data sheets.

Table 1.2: Material Class and Material Class Abbreviation

Table 1.3 : Material Recommendations for Different Services

Service Temperature Range (°F)

Material Class

See Note

Fresh water, condensate, cooling-tower water <212 I-1 or I-2 <250 I-1 or I-2 4

250 – 350 S-5 4 Boiling water and process water > 350 C-6 4

>200 C-6

Boiler feed water - Axially split - Double casing (barrel) >200 S-6 Boiler circulator >200 C-6 Foul water, reflux drum water, water draw and hydrocarbons containing these waters, including reflux streams

<350 >350

S-3 or S-6 C6

5

Propane, butane, liquefied petroleum gas and ammonia (NH3) <450 S-1 Diesel oil; gasoline; naphtha; kerosene; gas oils; light, medium, and heavy lube oils; fuel oil; residuum; crude oil; asphalt; synthetic crude bottoms

<450 450-700

>700

S-1 S-6 C-6

5,6 5

Noncorrosive hydrocarbons, e.g. catalytic reformate, isomaxate, desulfurized oils

450-700 S-4 6

Xylene, toluene, acetone, benzene, furfural, MEK, cumene <450 S-1 Sodium carbonate, doctor solution <350 I-1 Caustic (sodium hydroxide), concentration of ≤ 20% <210

≥210 S-1 7

8 Seawater <200 9 MEA, DEA, TEA – stock solutions <250 S-1 DEA, TEA – lean solutions <250 S-1 7 MEA – lean solution (CO2 only) 175-300 S-9 7 MEA – lean solution (CO2 and H2S) 175-300 7,10 MEA, DEA, TEA – rich solutions <175 S-1 7 Sulfuric acid concentration

85% <100 S-1 5 85% - 15% <100 A-8 5 15% - 1% <100 A-8 5

1% <450 A-8 5 Hydrofluoric acid concentration of >96% <100 S-9 5

Notes :

1. Separate materials recommendations should be obtained for services not clearly identified by the service descriptions listed in this table.

2. Cast iron casings, where recommended for chemical services are for nonhazardous locations only. Steel casing (S-1 or I-I) should be used for pumps in services location near process plants or in any location where released valeased vapour for a failure could create a hazardous situation or where pumps could be subjected to hydraulic shock, for example, in loading services.

3. Mechanical seal materials, for streams containing chlorides, all springs and other metal parts should be Alloys 20 or better. Buna-N and Neoprene should not be used in any service containing aromatics. Viton should not be used in services containing aromatics above 200°F.

4. Oxygen content and buffering of water should be considered in the selection of material. 5. The corrosiveness of foul waters, hydrocarbons over 450°F, acids, and acid sludges may very widely. A

material recommendation should be obtained for each service. The material class indicated above will be satisfactory for many of these services but must be verified.

6. If product corrosivity is low, Class S-4 materials may be used for services at 451-700°F. A separate materials recommendation should be obtained in each instance.

7. For temperatures greater then or equal to 160°F, all welds should be stress relieved. 8. Alloy 20 or Monel pump material and double mechanical seals should be used with a pressurized seal oil

system. 9. For seawater services, the purchaser and the vendor should be agree on the construction materials the

best suit the intended use. 10. Class A-7 material should be used, except for carbon steel casings.

CHAPTER – 2

PROCESS COMPRESSORS

2.1 General

To enable the procurement of a compressor, it is necessary to convey to the vendors all

pertinent process information. This process information is conveyed to the Specialist/

Mechanical group of the company in the form of process groups standard "Compressor

Data Sheet". Based on this information, the Specialist/Mechanical group prepares the

compressor specification sheet to be sent to the vendors. Following guidelines should help

the process engineer to prepare process data sheets for compressors.

2.2 Type of Compressors

Compressors generally used in the process industry can be classified as positive

displacement, dynamic or thermal types as shown in Figure 2.1.

2.2.1 Positive Displacement Type Compressors

Positive displacement type compressors compress a constant volume of fluid in each stroke

or rotation. They include reciprocating, rotary and diaphragm compressors.

Reciprocating compressors are the oldest and still the most widely used type of

compressors. They consist of one or more cylinders each with a piston or plunger that

moves back and forth, displacing a positive volume with each stroke. They are available

from less than 1 hp to more than 3000 hp and for pressures ranging from below

atmospheric to 50,000 psig (about 3520 kg/cm2a ). They run at speeds from 150 rpm to

1000 rpm. Overall efficiency of reciprocating compressors range from 75-80%. Vacuum

pumps are a variation of the reciprocating compressor. Suction pressure of0.5 to 1.5 inches

of mercury absolute are quite common. Vacuum pumps are characterised by high

compression ratio.

Diaphragm compressors use a hydraulically pulsed flexible diaphragm to displace the

fluid.

Rotary compressors cover lobe-type, screw type, vane type and liquid ring type, each

having a casing with one or more rotating elements that either mesh with each other such as

lobes or screws, or that displace a fixed volume with each rotation. The screw type is

probably the most important for general refinery and petrochemical plant application.

Screw compressors run at speeds from 2000 to 15000 rpm. These compressors can operate

on relatively dirty gases and are by nature nonlubricated. Efficiency is somewhat lower

than those for comparable reciprocating compressors.

Figure 2.1 Types of Compressors

2.2.2 Dynamic Type Compressors

Dynamic type compressors include radial-flow (centrifugal), axial-flow and mixed flow

machines. They are rotary continuous-flow compressors in which the rotating element

(impeller or bladed rotor) accelerates the fluid as it passes through the element, converting

the velocity head into static pressure, partially in the rotating element and partially in

stationary diffusers or blades.

Centrifugal compressors have variable head-capacity characteristics. These are used for

handling large volume of gas at relatively intermediate discharge pressures. These

compressors are simple, require less maintenance and have longer reliability factors.Axial

compressors are somewhat similar to centrifugal compressors, but instead of having

impellers moving inside diffusers, they consist of a series of stator and rotory blades. These

are used for handling very large flows at relatively low heads. Generally, they do not have

a stable operating range over as wide a capacity and are therefore, considered less versatile.

2.2.3 Thermal Type Compressors

Ejectors are “thermal” compressors that use a high velocity gas or steam jet to entrain the

inflowing fluid, then convert the velocity of the mixture to pressure in a diffuser.

2.3 Selection of Compressor Type

Though the final selection of the compressor for a specific requirement will be done by

close cooperation of the vendor and the Specialist/Mechanical group, a preliminary

selection of the type of compressor required is made by the process engineer based on his

process needs. Figure 2.2 should be used for the general range of application of the various

compressor types.

Figure 2.2: General Range of Application of Various Compressor Types

The main differences between reciprocating and centrifugal compressors (the two most

commonly used type of compressors) are summarised in Figure 2.3.

Figure 2.3: Comparison of Reciprocating and Centrifugal Compressors

2.4 Fluid Properties

All the properties of fluid to be compressed should be clearly specified in the process

datasheet of the compressor. The composition of the fluid, moisture content, molecular

weight, ratio of specific heats (K = Cp/Cv), critical temperature and pressure and the

compressibility factor (Z) should be known. Volume I of the Process Engineering Manual

titled Process Databook can be used for properties for pure gases. The GPSA Databook

can be used to supplement the Process Databook. In case of mixed gas, calculate the

properties of the mixture by using the following guidelines. Further, any impurities like

solid particles, polymers, corrosive and erosive elements in the gas should be clearly

defined.

2.4.1 Adiabatic Exponent (K = Cp/Cv)

To calculate K, the adiabatic exponent i.e. the ratio of specific heats (Cp/Cv) of a gas

mixture, first calculate MCP for the mixture based on mole fractions of individual

components:

( ) ( )iP

n

imixP MCyiMC ∑

==

1

CENTRIFUGAL vs RECIPROCATING The advantages of a centrifugal compressor over a reciprocating machine are : a. Lower installed first, cost where pressure and volume conditions are favourable. b. Lower maintenance expense. c. Greater continuity of service and dependability. d. Less operating attention. e. Greater volume capacity per unit of plot area. f. Adaptability to high-speed-low-maintenance-cost drivers. The advantages of a reciprocating compressor over a centrifugal machine are : a. Greater flexibility in capacity and pressure range. b. Higher compressor efficiency and lower power cost. c. Capability of delivering higher pressures. d. Capability of handling smaller volumes. e. Less sensitive to changes in gas composition and density.

where n is the total number of components in the mixture, yi is the mole fraction of the ith

component and (MCp)i is the molar heat capacity of the ith component. The adiabatic

exponent is then calculated as below:

( )( ) RMC

MCK

mixP

mixPmix −

=

If MCp is in Btu/ibmol °R or kcal/kgmol °C, the value of R is 1.986. Refer to Figure 2.4 for

molar heat capacities of individual hydrocarbons. Since the heat capacities vary

considerably with temperature, K is normally determined at the average of suction and

discharge temperatures. Aproximate adiabatic exponents for hydrocarbon gases based on

average molecular weights are given in Figure 2.5.

2.4.2 Average Compressibility Factor (Zavg)

Compressibility factors, which modify the ideal gas law are important primarily for

multistage reciprocating and centrifugal compressors where volumes must be calculated for

each succeeding stage. For perfect gases like H2, N2, air etc., compressibility factor is 1.0.

Calculate average compressibility factor of non-ideal gas/gas mixture as below:

221 ZZ

Zavg

+=

Z1 = Compressibility factor at suction conditions

Z2 = Compressibility factor at discharge conditions

Use the following for calculating the discharge temperature T2 based on the suction

temperature T1 and the suction and discharge pressures (P1 and P2) :

Figure 2.4: Molar Heat Capacities

Figure 2.5: Approximate Adiabatic Exponents for Hydrocarbon Gases

K

1K

1

22 P

PTT

−

=

Charts given in section 7.8 of the Process Databook can be used for calculating

compressibility factors for hydrocarbon gases of different molecular weights.

2.4.3 Composition of gas

The composition of gas to be compressed should be known and the same is to be indicated

in the data sheet of the compressor. In case of a mixture of gases, give the composition of

the individual constituents of the mixture and the molecular weight of the mixture. Also

specify the water content in the gas, if any. Moisture content is very important because

when gases are compressed, their ability to hold water or other potential condensates

decreases at a given temperature and provision must be made for separating the condensed

material between different stages of compression. For air compressors, the inlet relative

humidity is required from which the amount of water present in the air can be determined.

2.4.4 Compressor Capacity

The normal flow rate required by the process should be estimated in kg/hr.

2.5 Compressor Calculations

A process engineer is required to make certain basic thermodynamic calculations for

estimating the head and horsepower requirements so that he can make a preliminary

assessment of the various utilities needed by the compressor (as described in section 2.10).

The amount of work required in compressing a gas is dependent on the polytropic curve

involved and increases with increasing values of N. Refer Figure 2.6. The path requiring

the least amount of work is N = 1 which is equivalent to isothermal compression. For

adiabatic (isentropic) compression, N is equal to the Adiabatic Exponent (K). It is usually

impractical to build a isothermal machine. Most machines tend to operate along a

polytropic path which approaches the isentropic. Most compressor calculations are

therefore based on an efficiency applied to account for true behaviour. Certain type of

Figure 2.6: Polytropic Compression Curves

compressors, like reciprocating and single stage centrifugal, follow an adiabatic

compression cycle while others like multistage centrifugal, follow a polytropic

compression cycle. The following equations are useful. The legend for the various symbols

used is given at the end of this section.

2.5.1 Adiabatic Compression

(a) Adiabatic Head:

−

−

+=

−

K

1K

1P

P

TM

1545

2

ZZH

K

1K

1

2

121

ad

(b) Adiabatic Gas Horse Power :

ad

adad 3300

WHHP

η=

Adiabatic efficiency (ηad) can be taken as 65-70% for most of the cases for preliminary

estimation.

(c) Adiabatic Discharge Temperature :

K

1K

1

21ad2 P

PT)T(

−

=

2.5.2 Polytropic Compression

(a) Polytropic Head :

−

−

+=

−

N

1N

1P

P

TM

1545

2

ZZH

N

1N

1

2

121

poly

(b) Polytropic Gas Horse Power :

poly

poly

poly 33000

WHHP

η=

Polytropic efficiency (ηpoly) can be taken as 70-75% for most of the cases for preliminary

estimation.

(c) Polytropic Discharge Temperature :

N

1N

1

21poly2 P

PT)T(

−

=

2.5.3 Relationship of Polytropic and adiabatic Exponents

The value of quantity N in polytropic relations is found as :

η

−=

−

poly

1

K

1K

N

1N

ad

ad

poly

poly HH

η=

η

Legend of symbols used

HP = Gas Horse Power

H = Head, ft lb/lb

Z1 = Compressibility Factor

M = Molecular Weight of gas

T = Temperature of gas in °R

K = Ratio of specific heats Cp/Cv, adiabatic exponent

N = Polytropic exponent

W = Weight flow, lb/min

η = Compression Efficiency

Subscripts

ad = Adiabatic Compression

poly = Polytropic Compression

1 = At suction of compressor

2 = At discharge of compressor

2.6 Compression Ratio

Compression ratio of any stage of a compressor is the ratio of absolute pressure at the

discharge of that stage to the absolute pressure at the suction of that stage. It is normal

practice to balance the cmpression stages so that each stage is designed for approximately

the same compression ratio. This is done partly for economic reason but also to limit

maximum interstage discharge temperatures. For example, discharge temperatures in

lubricated reciprocating type air compressors are limited to 160°C to avoid the potential

danger of fires from a reaction of oxygen with hot oil. In general, it is not advisable to use a

single stage compressor for compression ratio above 3.5 at relatively low pressures. For

suction pressures above 70 kg/cm2a, a compression ratio of not more than 2.5 per stage

should be considered. When multistage compressors are involved, care should be taken to

provide means for routing the interstage condensate to some appropriate location in the

process flow scheme.

2.7 Compressor Capacity

A process engineer will usually indicate his requirement of flow in the process data sheet

expressed as kg/hr. In case of recycle streams, the material balance of various recycle

streams should be indicated in the process data sheet by making a system sketch. In case

inlet flow in volumetric units is required, the following expression can be used :

W*1VQ1 =

1

111 P144

RTZV =

where,

Q = inlet flow ft3/min

W = Weighted flow lb/min

Z1 = Compressibility factor at suction temp.

R = Gas constant (1544/Mol. wt.)

T1 = Suction Temperature °R

P1 = Suction Pressure psia

2.8 Choice of Driver

Electric motors, stream turbines and in some cases gas turbines and oil/gas engines are used

to supply motive power to compressors. The selection of a particular drive is a function of

utilities available, cost and preference of owners. These alternate choices should be

reviewed at the stage of fixing up the design basis.

2.8.1 Reciprocating Compressors

In case of reciprocating compressors, oil/gas engine drives have certain advantages in cases

where gas or oil is in abundance and no other power supply is conveniently available. This

might be the case in oil or gas field installations or for pipeline pumping and compression

units. Direct steam engine drive is also used frequently for reciprocating compressors. It

should be recognised, however, that waste steam or condensate from such units contains oil

and cannot be directly returned to a condensate system. Steam or gas turbines are not very

common drives for reciprocating compressors. If used, care must be taken to provide speed

reducing gears, flywheels, torsional couplings etc. Electric motors are used quite frequently

as drives for reciprocating compressors in the process plants. Motors for big compressors

are generally synchronous type, whereas small compressors may use a V-bolt drive, gear

reducer or direct valve with an induction motor.

2.8.2 Centrifugal Compressors

Electric motors, steam turbines and in some case, gas turbines are used to supply motive

power to centrifugal compressors. Since these compressors generally operate well above

3600 rpm, speed-increasing gears are employed with motor drives. Steam turbines can

either be condensing, non-condensing or extraction type, depending on such matters as

steam cost and need for process steam at different pressure levels.

Gas turbines may be used to supply power to large (5000-7000 hp) centrifugal

compressors, for instance in ethylene plants, offshore gas compression units, etc.

2.9 Number of Compressors and Sparing Philosophy

Generally, the reliability factor for reciprocating machines vary from 95% to 98% and for

centrifugal machines from 99.5% to 100%. If the client has no particular preference,

following criteria (based on economic considerations) can be used for selecting the number

of compressors. However, the actual number of spares will depend on the clients

philosophy on operation and maintenance.

Reciprocating

Up to 100 KW One running (100%) + one standby

(100%)

From 100 KW and above Min. of 33-50% as spare capacity

Centrifugal

Since centrifugal compressors are of large size and have comparitively higher service

factor, it is not conventional to have a spare unit. Also, upto 15000 hp, it is conventional to

put in one full size machine rather than two 50% size machines with adequate spare parts

including a complete rotor to reduce the downtime of the machine.

2.10 Utilities for Compressors

All compressors need utilities like electric power, steam or gas for their drivers, cooling

water for condenser, jacket, bearing, intercooler, aftercooler and oil cooler etc., fuel gas and

starting air for gas turbines. Though the exact requirement of various utilities needed by a

compressor will be specified by the compressor vendor, it is essential for a process

engineer to have a fairly good idea about these requirements at the stage of process design

work so that these can be included in the overall utility summary sheet of the plant. These

requirements can later be firmed up after getting the vendor information. The process

engineer should indicate the utility conditions in the compressor process data sheet.

Following guidelines can be used for estimating utilities required by the compressor.

2.10.1 Power for Driver

Following steps are involved in estimating the electric power required by the compressor:

(a) Calculate the gas horse power of the compressor as given in section 2.5.

(b) Calculate the BHP of the compressor by dividing the gas horse power by the

mechanical efficiency. Take 95% as mechanical efficiency for centrifugal

compressors. Use following table for mechanical efficiencies for reciprocating

compressors.

Total Compression Ratio Mechanical Efficiency 1.5 65 2 74 3 81 4 83 5 82 6 81 7 79

(c) Applying the overload capacity requirement of the compressor and the service factor

of electric motors, estimate the driver horse power as follows :

Power requirement of Motor = BHP / (0.9 x Motor efficiency)

Motor efficiencies can be taken as 95-98%

2.10.2 Steam Requirement for Turbines

(a) Reciprocating Compressors :

Steam requirement for the estimating BHP is calculated from Figure 1.7 and 1.8 given in

Chapter 1 for pumps.

(b) Centrifugal Compressors :

The turbine is specified to develop 120% of the BHP required at the design conditions. The

theoretical steam rate in kg/HP (hr) can be obtained from the expression 21 hh

643

−

where h1 – h2 is the enthalpy change of steam expressed in kcal/kg. Actual steam rate can

then be obtained by dividing the theoretical steam rate by the turbine adiabatic efficiency

which is given below:

BHP Adiabatic Efficiency 500 61 800 66 1000 67 1200 69 1500 70 2000 72 2500 73

3000 & above 74

2.10.3 Fuel Gas for Gas Engines and Turbines

Gas Engines

Use an average figure of 2400 kcal/BHP/hr. Estimate the fuel gas requirement based on

this figure and the calorific value of the fuel gas available for use. Gas engines can be run

at fuel rated power when burning any fuel gas having a heating value of 7100-23000

kcal/m3, but no more than 10 mole% butane. If it contains more than 10 mole% butane, the

engine must be derated by 10% i.e. a 500 hp engine will be suitable for 450 hp.

Gas Turbines

Use an average figure of 3000-3200 kcal/BHP/hr. Estimate the fuel gas requirement based

on this average figure and the calorific value (net) of the fuel gas available for use. Horse

power rating of gas turbines are based on 26.5°C combustion air temperature and 300 meter

altitude. They should be derated by 5% for every 5°C increase in air temperature.

2.10.4 Cooling Water Requirements

Different types of compressor and their drives require cooling water. Estimate the cooling

water as follows. Consider a 9°C rise in temperature of cooling water from inlet to outlet.

Compressor Type/ Drive Cooling Water

Requirement

Type of Water

Reciprocating

Compressors

- Compressor Jacket 125 kcal/BHP/hr Treated

- Inlet Cooler 250 kcal/BHP/hr Treated

- Aftercooler 250 kcal/BHP/hr Treated

- Gas Engine Cylinder Jacket 1000 kcal/BHP/hr Closed System

Centrifugal Compressors

- Oil Cooler 20 m3/hr (min.) Treated

-

2.10.5 Cooling Water for Turbine Condenser

Cooling water is required for condensers of stem turbine driven compressors. Considering

that condenser will operate at 750 mm Hg, calculate the change in steam enthalpy from

intel steam conditions to the condenser conditions. From the steam rate found in section

2.10.2, find the total heat to be removed. Considering a temperature rise of 9 oC in cooling

water inlet to outlet, estimate the cooling water requirements of the condenser.