Engineering Encyclopedia

Note: The source of the technical material in this volume is the ProfessionalEngineering Development Program (PEDP) of Engineering Services.

Warning: The material contained in this document was developed for SaudiAramco and is intended for the exclusive use of Saudi Aramco’s employees.Any material contained in this document which is not already in the publicdomain may not be copied, reproduced, sold, given, or disclosed to thirdparties, or otherwise used in whole, or in part, without the written permissionof the Vice President, Engineering Services, Saudi Aramco.

Chapter : Instrumentation For additional information on this subject, contactFile Reference: PCI10403 E.W. REAH on 875-0426

Engineering EncyclopediaSaudi Aramco DeskTop Standards

PUMPS AND COMPRESSORS

Engineering Encyclopedia Instrumentation

Pumps and Compressors

Saudi Aramco DeskTop Standards

CONTENT PAGE

PUMPS ................................................................................................................. 1

Centrifugal Pumps ..................................................................................... 1

Centrifugal Pump Specific Speed And Pump Profiles................... 1

Centrifugal Pump Principle Of Operation...................................... 2

Centrifugal Pump Head.................................................................. 3

Centrifugal Pump Applications...................................................... 5

Centrifugal Pump Mechanical Components .................................. 5

Centrifugal Pump Head Vs. Flow Characteristic ........................... 9

Centrifugal Pump Parallel/Series Operation ................................ 10

Reciprocating Pumps............................................................................... 12

Piston Pumps................................................................................ 12

Plunger Pumps ............................................................................. 15

Metering Pumps ........................................................................... 16

Process Calculations For Reciprocating Pumps........................... 18

Rotary Pumps .......................................................................................... 19

Gear Pumps.................................................................................. 19

Screw Pumps................................................................................ 20

Pump Horsepower (All Pumps)............................................................... 21

Net Positive Suction Head (All Pumps) .................................................. 22

NPSH Available Versus NPSH Required .................................... 23

Cavitation..................................................................................... 26

Dissolved Gases ............................................................... 26

Centrifugal Pump Calculations................................................................ 27

Performance Curves..................................................................... 27

Viscosity ...................................................................................... 28

Affinity Laws ............................................................................... 30

Example Problem 1 .......................................................... 32

Centrifugal Pump Control Systems ......................................................... 33




Typical Centrifugal Pump Installation..................................................... 36

Starting A Centrifugal Pump........................................................ 38

Operating Problems With Centrifugal Pumps ......................................... 39

Control Methods For Positive Displacement Pumps............................... 40

Recycle Control ....................................................................................... 41

Variable Speed Motor Drive.................................................................... 41

Variable Piston Stroke Length................................................................. 41

CENTRIFUGAL COMPRESSORS.................................................................... 42

Centrifugal Compressor Components And Functions ............................. 42

Thermodynamic Equations For Gas Compression .................................. 45

Centrifugal Compressor Head...................................................... 46

Centrifugal Compressors Are Polytropic ..................................... 49

Polytropic Efficiency ................................................................... 49

Compressor Discharge Temperature............................................ 50

Power Requirements .................................................................... 50

Casing Arrangements .............................................................................. 51

Intercooling .................................................................................. 51

Sidestreams .................................................................................. 52

Performance Curves ................................................................................ 53

Actual Volume ............................................................................. 55

Centrifugal Compressor Fan Laws .......................................................... 56

Surge........................................................................................................ 57

Stonewall ................................................................................................. 57

Efficiency Of An Operating Machine...................................................... 58

Procedures.................................................................................... 58

Method D - Computer Program Compress .................................. 59

Control Schemes For Centrifugal Compressors ...................................... 60

Variable Speed ............................................................................. 60




Suction Throttling ........................................................................ 61

Discharge Throttling .................................................................... 62

Antisurge Control......................................................................... 63

Combined Controls ...................................................................... 64

Common Process Problems With Centrifugal Compressors ................... 64

POSITIVE DISPLACEMENT COMPRESSORS .............................................. 65

Reciprocating Compressors Principle Of Operation................................ 65

Reciprocating Compressors Mechanical Components ............................ 66

Reciprocating Compressor Lubrication........................................ 67

Reciprocating Compressors Performance Calculations........................... 68

Volumetric Efficiency.................................................................. 70

Reciprocating Compressors Intercoolers ................................................. 72

Reciprocating Compressors Control Systems.......................................... 73

Suction Valve Lifters ................................................................... 73

Clearance Pockets ........................................................................ 73

Recycle......................................................................................... 74

Variable Speed ............................................................................. 74

Reciprocating Compressor Process Problems ......................................... 74

Liquid In Suction ......................................................................... 74

Vibration Of Piping...................................................................... 75

Leakage Of Valves And Piston Rings.......................................... 75

Detecting Valve Leakage ............................................................. 75

Mechanical Problems .............................................................................. 75

STEAM TURBINES........................................................................................... 76

Steam Turbine Introduction..................................................................... 76

Steam Turbine Principle Of Operation .................................................... 76

Curtis Stage.................................................................................. 77

Other Types Of Stages ................................................................. 78

Turbine And Cycle Efficiency................................................................. 79




Mechanical Components ......................................................................... 80

Trip And Throttle Valve............................................................... 80

Governor Valve............................................................................ 81

Steam Chest ................................................................................. 81

Hand Valves................................................................................. 81

Calculations ............................................................................................. 81

Example Problem 2 - Theoretical & Actual SteamRate, And Outlet Temperature ..................................................... 81

Example Problem 3 - Efficiency Of An OperatingTurbine......................................................................................... 85

Efficiencies Of Steam Turbines For Use In Calculations ........................ 88

Use Of Hand Valves To Maximize TurbineEfficiency..................................................................................... 89

THEORETICAL STEAM RATE TABLES........................................................ 90

Performance Curves ................................................................................ 90

Common Operating Problems ................................................................. 91

COMBUSTION GAS TURBINES ..................................................................... 92

Combustion Gas Turbines Introduction................................................... 92

How A Gas Turbine Works ..................................................................... 92

Gas Turbine Types .................................................................................. 94

Heavy Duty .................................................................................. 94

Aircraft Derivative ....................................................................... 94

Gas Turbine ............................................................................................. 95

Gas Turbine Configurations .................................................................... 96

Single-Shaft.................................................................................. 96

Dual-Shaft .................................................................................... 96

Available Models Of Gas Turbines.............................................. 98

Fuels For Gas Turbines ........................................................................... 98

Gas Turbine Efficiencies ......................................................................... 99

Cycle Efficiencies ........................................................................ 99




Site Ratings .................................................................................. 99

Thermal Efficiency Of Gas Turbine........................................... 100

Gas Turbine Performance Curves.......................................................... 100

Auxiliary Equipment ............................................................................. 101

Control Systems..................................................................................... 101

Example Problem 4 - Performance Curves For Gas TurbineCalculations ........................................................................................... 103

REFERENCES.................................................................................................. 105

WORK AID ...................................................................................................... 106

Work Aid 1 Centrifugal Pump Equations........................................ 106

Work Aid 2: NPSHr Vs. % Capacity................................................ 107

Work Aid 3: Centrifugal Pump Viscosity Correction Factors .......... 108

Work Aid 4: Affinity Laws............................................................... 109

Work Aid 5: Typical Centrifugal Pump Installation......................... 110

Work Aid 6: General Characteristics Of Centrifugal Pumps ............ 111

Work Aid 7: Performance Characteristics Of Centrifugal Pumps .... 112

Work Aid 8: Centrifugal Pump Selection Charts.............................. 113

Work Aid 9: Centrifugal Pump Troubles And Causes...................... 114

Work Aid 10: Pump Head And Horsepower ...................................... 117

Work Aid 11: Efficiency Of Gear Pumps........................................... 121

Work Aid 12: Efficiency Of Screw Pumps......................................... 122

Work Aid 13: Centrifugal Compressor - Calculation Form................ 123

Work Aid 14: Common Operating Problems For CentrifugalCompressors ................................................................ 126

Work Aid 15: Reciprocating Compressor - Calculation Form ........... 127

Work Aid 16: Typical Isentropic Efficiency Of ReciprocatingCompressors ................................................................ 129




Work Aid 17: Reciprocating Compressor - VolumetricEfficiency Calculation Form........................................ 130

Work Aid 18: Loss Correction For ReciprocatingCompressor.................................................................. 134

Work Aid 19A: Calculation Of Theoretical And Actual SteamRates, And Outlet Temperature ................................... 135

Work Aid 19B: Calculation Of Turbine Efficiency AndHorsepower From Steam Conditions........................... 137

Work Aid 19C: Calculation Of Steam Turbine Efficiency FromInlet Steam Condition And Brake Horsepower ........... 138

Work Aid 20: Typical Single Stage Steam Turbine Efficiency .......... 139

Work Aid 21: Steam Turbine Troubleshooting .................................. 140

Work Aid 22: Gas Turbine Altitude Correction Factor ForOutput And Heating Consumption And AlsoAltitude Vs. Atmospheric Pressure.............................. 141

Work Aid 23: General Electric Model M5382(C) GasTurbine -- Effect Of Compressor InletTemperature On Maximum Output, Heat Rate,And Air Flow............................................................... 142

Work Aid 24: General Electric Model M5382(C) *38,000 HPGas Turbine Performance - Inlet 120°F....................... 143

Work Aid 25: General Electric Model M5382(C) *38,000 HPGas Turbine Performance - Inlet 90°F......................... 144

Work Aid 26: General Electric Model M5382(C) *38,000 HPGas Turbine Performance - Inlet 30°F......................... 146

GLOSSARY...................................................................................................... 146

ADDENDUM (CENTRIFUGAL PUMPS - HYDRAULICINSTITUTE STANDARDS) .................................................. 161

K-502, SADP-K-502, API-616, API-679, and GPSA Engineering Data Book.



Saudi Aramco DeskTop Standards 1

PUMPS

There are two major classification of pumps: centrifugal and positive displacement.

Centrifugal Pumps

Centrifugal pumps are the most common type of pump used by Saudi Aramco and otherpetroleum companies. Almost all processes contain several centrifugal pump installations.

Calculating the process parameters for these pumps is a significant part of every processengineer's job. A basic understanding of mechanical components, control systems andinstallation details is also important.

Specification of a new pumping service is a cooperative effort between a process engineer anda mechanical specialist. This is also true for most troubleshooting assignments.

Centrifugal Pump Specific Speed and Pump Profiles

Centrifugal pumps can vary in type from radial-vane to axial flow as shown in Figure 1.Specific speed, which is the correlating parameter for the pump profiles in Figure 1 is usuallyexpressed as:

NS =N QH3/4

where:Ns = Pump specific speed,N = Rotational speed, RPM (revolution per minute)Q = Flow, gpmH = Pump head, ft/stage

The specific speed is defined as the RPM at which a geometrically similar impeller would runif it were of such a size as to discharge one gpm at a head of one foot.

Specific speed is indicative of the shape and characteristics of an impeller. Specific speed isuseful to the designer in predicting proportions required and to the applications engineer inchecking suction limitations of pumps.

Impeller form and proportions vary with specific speed as shown in Figure 1.

We will be only discussing radial flow centrifugal pumps since these are the most generallyused pumps in the refineries.




Comparison of Pump ProfilesFigure 1

Centrifugal Pump Principle of Operation

A pump converts mechanical energy into pressure in a flowing liquid. A centrifugal pumpdoes this by centrifugal action. Refer to Figure 2. A centrifugal pump has two majorcomponents: the internal impeller and the outer casing. The liquid enters the suction of thepump at A. It then flows to B and outward through the channels of the impeller marked C. Asthe liquid flows outward in the impeller, the impeller imparts a very high spinning or tangentialvelocity to the liquid. The liquid then enters the volute of the pump, area D. Here the velocityenergy is converted to pressure.

Centrifugal PumpFigure 2




Centrifugal Pump Head

Head is the term used to describe the energy imparted to the liquid. The units of head are foot-pounds (ft-lb) of force per pound of mass.

Head produced ft - lb

lb=

V2

2gwhere:

V = Velocity of impeller tip, ft/secg = gravitational constant, 32.2 ft/sec2

Note that the important velocity is the tangential velocity at the tip of the impeller. Thisvelocity is proportional to the diameter of the impeller and the rotational speed. Therefore, theequation for head can be written in terms of pump characteristics as follows:

Head (ft) = DN

1840

2

where:D = Impeller diameter, inchesN = Pump speed, RPM

As shown in Figure 3, the precise units of head are ft-lb (force) per lb (mass). However, it isconventional practice to cancel the lb units and to speak of head in terms of feet. Note that thepump vendor designs the impeller to produce the head required at the design point.

The pressure differential produced by a pump is equivalent to a column of the pumped liquid,where the height of the column is equal to the head produced by the pump. For a given flowand speed, a given centrifugal pump produces a constant head with any fluid, assuming nowear and fouling.

∆∆ H ==100 Ft

∆∆p ==43 . 3 psi

∆∆p psi( ) == Head ( Feet ) ×× 0. 433 ×× S. G.

43 . 3 psig

0psig

0psig

Head and Differential Pressure are EquivalentFigure 3




The S.G. is the specific gravity at flowing conditions.




Centrifugal Pump Applications

Centrifugal pumps are the most common type of pump and the first choice of industry with theexception of the following:

1. Very high differential pressure (over 2000 psi)

2. High viscosity fluids (over 500 centistokes)

3. Very low flow rates (less than 10 gpm)

Centrifugal Pump Mechanical Components

Figure 4 illustrates the major components of a centrifugal pump. This is a diagram of ahorizontal single-stage, overhung pump which is the most common type. Horizontal refers tothe orientation of the shaft; single-stage means there is one impeller, and overhung means thatthe impeller is outside of the two supporting bearings, not between the bearings.

The shaft runs through the center of the pump and holds the impeller at the left end. The drivemotor is connected to the right end of the shaft through a flexible coupling. The liquid entersthe suction nozzle, passes through the enclosed sections of the spinning impeller, and exitsthrough the discharge nozzle at the top of the pump. The right end of the pump is the bearinghousing. This housing contains two sets of ball bearings that support the weight of the shaft.They also absorb the axial thrust on the shaft.

Casing Rings Frontand Back Sideof Impeller

ShaftSleeve

Lantern RingConnection

Stuffing Box forMechanical Seal orPacking (Packing Illustrated)

Ball BearingThrust Bearing

Cantilevered orOverhung TypeShaft Support

Guide Bracket(Not for StructuralSupport)

Oil Lubrication System

Quenching TypePacking Gland

ImpellerBalance Port

Close Tube Impeller

End Suction Casing Circular CasingJoints withConfined Gasket

Horizontal, Single Stage, Overhung Centrifugal PumpFigure 4




The casing contains the liquid under pressure. A seal is required where the rotating shaft entersthe casing. This area is called the stuffing box and may actually contain a stuffing or packing.However, most modern pumps have mechanical seals at this point. Sealing the shaft is veryimportant to prevent leakage of the pumped fluid, which is frequently hazardous, flammable, ortoxic. Therefore, careful attention must be paid to the design, installation, and maintenance ofthe seals. Many different types of seals are available for different process conditions.

Heat is generated by friction in seal area of the shaft, and sometimes cooling is required. Achannel called the flushing connection (lantern ring connection) is available for this purpose.

Impellers may be the open, semi-closed, or closed. These are shown in Figure 5. In thepetroleum and gas process plants, most impellers are the closed type. Closed impellers cangenerate higher heads at greater efficiencies. Open and semi-closed impellers are used forliquids that contain solids. They will not clog as easily as closed impellers.

Partially Open(Semi-Closed)

Open Enclosed

Basic Types of ImpellersFigure 5

Figure 6 shows other common types of centrifugal pumps. The amount of head that can begenerated by a single impeller is limited to a maximum value. If more head is required, pumpdesigns incorporate two or more impellers. These may be arranged in a horizontal multistageconfiguration or a vertical multistage configuration.




Types of Centrifugal PumpsFigure 6A




• Horizontal-Single Stage

- The most common type- Used for moderate head, <500 ft- End suction, top discharge

or top suction, top discharge

• Vertical In-line

- Supported by piping or small foundation- Motor is supported by pump; piping forces do not affect

alignment- Lower cost, simpler maintenance- Slightly higher NPSHR than horizontal pump

• Horizontal Multistage

- Up to 8 impellers for higher head- Shaft supported between bearings

• Vertical Can

- Used when low NPSHR is needed

• Vertical-Submerged Suction

- Like vertical can type, without the can- Used in sumps or shallow wells- Used to pump water from the sea, or from reservoirs

• Submersible

- Used in oil production wells

Types of Centrifugal PumpsFigure 6B




Centrifugal Pump Head Vs. Flow Characteristic

The process performance of a centrifugal pump is described by a curve called the head versusflow characteristic. See Figure 7. Centrifugal pumps are constant-head devices. This meansthat they provide a nearly constant head, or pressure differential, even though the flow ratechanges. As Figure 7 shows, the head produced by the pump does increase somewhat as theflow rate decreases from the design point. Conversely, the head decreases at flow rates abovethe design point. However, over the normal operating range of the pump, the head is relativelyconstant or, as we say, the curve is relatively flat. Normally, the head developed at zero flow isno more than 110 to 120% of the head at the design point. This is called the shutoff point, orshutoff head.

Head Vs Flow CharacteristicFigure 7

Note that shutoff means that the flow is shut off, for example by closing a valve at thedischarge of the pump. The pump itself continues to rotate and develop differential pressure.However, a pump should not be operated this way except for a short period. After a minute ortwo, the pump may overheat and damage will occur.




A system curve can be constructed by calculating the system pressure drop with a fully opencontrol valve at several rates and plotting them on the pump curve as shown in Figure 8. Thiswill define the available control valve pressure drop and the maximum flow for the system.

Flowgpm Design

FlowMaxFlow

Head

∆∆p( )

Control Valve ²p

Pump and System CurveFigure 8

Centrifugal Pump Parallel/Series Operation

When two pumps are operated in series, as shown in Figure 9, the net pump curve can beconstructed by adding the head for each pump at several flow rates and plotting the resultinghead curve. Pumps in series may generate too much discharge pressure for the system designpressure. The control and relief system must be designed for safe operation. One remedy is toinstall a safety valve at the discharge of the second pump as shown below.

Series Pump ControlFigure 9




When two pumps are operated in parallel, the net pump curve can be constructed by adding theflow rates at several heads and plotting the resulting curve (Figure 10).

A problem that can occur with pumps operating in parallel is shown in Figure 10. Two pumpsare never exactly like. If two pumps are installed in parallel, one pump may take more thanhalf of the total flow and the other pump less than half. The pump with the lower flow ratemay be operating below its minimum acceptable flow rate. The head produced by the twopumps will be identical because they are connected to the same process. If the head curveproduced by pump B is lower than the head curve produced by pump A, the situation shown inFigure 10 will occur. Pump B will decrease its flow rate until it can produce the same head asPump A.

This situation is most dangerous when one pump is driven by a motor and the other by aturbine. It is impossible to set the two speeds exactly equal, and the difference in speed willcause a difference in head produced.

If two pumps are nominally identical and both driven by motors, the two head curves can beassumed to be within 3% of each other. If so, you can make the worst assumption, that is, thehead of pump B is 3% lower than the head of pump A. Then, using the system operatingconditions, plot the flow through both pumps. Make sure that the lowest flow rate is not belowthe pump minimum allowable flow rate.

Pumps in ParallelFigure 10




Reciprocating Pumps

Positive displacement pumps move a fixed volume of liquid per unit of time. Although lesscommon than centrifugal pumps, they are used for certain special applications within SaudiAramco. They are used for services with relatively low flow rates, services with highdifferential pressures, or for liquids that have high viscosities.

The two major types of positive displacement pumps used by Saudi Aramco are reciprocatingpumps and rotary pumps.

In a reciprocating pump, a piston moves back and forth within a cylinder. Liquid enters thecylinder during the suction or intake stroke. Liquid is forced out of the cylinder during thedischarge stroke. Refer to Figure 11.

Reciprocating PumpsFigure 11

Piston Pumps

Piston pumps are used for moderately high pressures. They are used in some special services,such as emergency and shutdown services, where pipelines must be completely emptied.Mixtures of liquid and vapor are encountered in these services, and the reciprocating pump isthe most reliable.

Figure 11 shows the major components of a piston pump: the cylinder, the piston that movesback and forth, and the suction and discharge valves in the ends of the cylinder. The left sideof Figure 11 shows the intake stroke. During the intake stroke, the piston moves away from theend of the cylinder, increasing the volume inside the cylinder. The decreasing pressure insidethe cylinder causes the suction valve to open and liquid to flow into the cylinder. The higherpressure on the discharge side keeps the discharge valve closed and this is no flow out of thepump.




The right side of Figure 11 shows the discharge stroke. Now the piston moves toward the endof the cylinder, forcing the liquid out of the cylinder. The higher pressure inside the cylinderpushes open the discharge valve but keeps the suction valve shut.

Some piston pumps are double acting in order to increase capacity and reduce pulsations.Figure 12 shows a double acting piston pump. There are cylinder chambers on both sides ofthe piston. While the left end of the cylinder is going through the suction stroke, the right endof the cylinder is discharging. The opposite happens when the piston moves back in the otherdirection. There is flow out of the pump in both directions of piston travel.Other important components of a piston pump are:

• Piston rod - transfers the energy from the drive mechanism to the piston.• Packing - surrounds the rod at the point where it enters the cylinder to prevent liquid

leakage.• Drive mechanism - includes a motor and a mechanism for changing the rotating

motion to a reciprocating motion. See Figure 13.

Double Acting Piston PumpFigure 12

aa

Pump CrankshaftMain Gear(Speed Reducer)

PistonCrosshead

Plunger

Connecting Rod Connecting Stub

Drive Mechanism - Reciprocating Pumps(Piston and Plunger)




Figure 13




Plunger Pumps

Plunger pumps are used for very high pressure applications, 1000 psig or greater. Examplesare glycol circulation and amine circulation in gas plants. See Figure 14. A plunger ofconstant diameter replaces the piston and piston rod. This provides the mechanical strengthneeded for high pressure.

Plunger pumps are single acting, that is, they pump on the forward stroke only. A pump thathas a single cylinder is called simplex. A pump with two cylinders operating in parallel is aduplex, and three cylinders in parallel form a triplex.

Multiple cylinders in parallel have two advantages:

• Higher capacities are possible without large diameter cylinders, an advantage at highpressures.

• The magnitude of flow rate pulsations is reduced because discharge strokes arestaggered.

Cylinder Head

Wing GuidedPoppet Valve

Suction

Plunger

With Permission from Exxon Company U.S.A.

Plunger PumpFigure 14




Metering Pumps

A metering pump is a special class of reciprocating pump. It is used for injecting very small,precisely controlled volumes of liquids into products or processes. Some of the liquids that arehandled by metering pumps are:

• Product additives.• Corrosion inhibitors.• Antifoam agents.• pH control agents; acids or alkalis.

The rates pumped by metering pumps are usually below 1 gpm. However, pumps are availablewith capacities up to about 10 gpm.

A metering pump is shown schematically in Figure 15. The cylinder has a single acting piston.There are ball type check valves at the suction and the discharge ends of the cylinder. In mostmetering pumps, a diaphragm isolates the piston from the pumped liquid. A hydraulic fluid iscontained between the diaphragm and the piston. The pumped liquid is on the other side of thediaphragm. This configuration isolates the piston and other parts of the mechanism from thepumped fluid. Such protection is required when the pumped liquid is corrosive, abrasive, toxic,or of low viscosity.

In the case of extremely toxic or flammable applications, a double diaphragm, with a leakdetector between diaphragms, is used.

Lubricating OilRelief Valve

Relief Valve

AutomaticAir-Bleed Valve

Discharge ValveBall Check

Diaphragm

SuctionBall Check

Valve

RefillValve

Hydraulic Fluid

LubricatingOil

PressureLubricated Drive

MicrometerStroke

Adjustment

High SpeedWorm

CrossHead

ConnectingRod

High SpeedWorm Worm

Gear

RotatingCrank

Zero Stroke

Discharge

Suction

DischargePosition

StrokeLength

Full Stroke

SuctionPosition

With Permission from Milton Roy.Metering Pump

Figure 15




Metering systems are usually supplied as a package on a skid. Process engineers frequentlyhave to troubleshoot these systems. Figure 16 shows a typical installation. The liquid is storedin an additive tank. The capacity of the tank is normally about one week's supply. Theadditive is delivered by a tank truck or in drums.

The liquid flows from the additive tank through stainless steel tubing to the metering pump andthen to the process. A strainer is installed just upstream of the pump to keep solids out of thepump.

Block valves allow the pump to be removed from service without disturbing the process.Normally, the spare for a metering pump is not installed in the line but rather is stored in thewarehouse. A malfunctioning pump can be removed from the line and the spare pump can beinstalled quickly.

One optional feature is a diluent added to the line downstream of the pump. Diluent increasesthe flow rate and reduces the residence time between the pump and the process. It provides avolume flow rate that is large enough to make a spray nozzle operate properly.

A relief valve or safety valve is required. It may be external as shown in Figure 16, or it maybe internal and supplied by the pump manufacturer. If the pump contains an internal relief, anexternal valve is not required unless the liquid is very flammable or toxic.

The calibration cylinder is used to verify that the proper flow rate of liquid is being deliveredby the pump. To use the cylinder, close the valve between the additive tank and the calibrationcylinder. Now, the pump will take suction from the cylinder alone. Time the rate at which thelevel falls in the calibration cylinder. The cross-sectional area of the calibration cylinder isknown. Therefore, the delivery rate can be calculated.

AdditiveTank

CalibrationCylinder

StrainerPump

Diluent (Optional)

Rotameter

ReliefValve To

Process

Metering Pump - Typical InstallationFigure 16




Process Calculations for Reciprocating Pumps

Two important calculations are displacement and volumetric efficiency.

Displacement is the volume of liquid that should theoretically be delivered by the positivedisplacement action in one minute. The equations are as follows:

Single acting pumps: Double acting pumps:

D =A( ) m( ) L S( ) n( )

231

A = π / 4( )d C2

D =2A - a( ) m( ) L S( ) n( )

231

a = π / 4( )d R2

where:D = Displacement, gpmA = Cross-sectional area of cylinder or plunger, sq. in.a = Cross-sectional area of piston rod, sq. in.m = Number of cylindersLS = Length of stroke, inchesn = Number of complete strokes/minute, RPM of the crankshaftdC = Diameter of cylinder, in.dR = Diameter of rod, in.

Because of leakage through the discharge and suction valves or around the piston, the actualvolume of liquid delivered may not equal the displacement. Volumetric efficiency is the actualrate of liquid flow divided by the displacement.

Volumetric Efficiency = Actual Flow Rate,gpm

Displacement,gpm

Reciprocating pumps are not used where it is possible to use a centrifugal pump. Centrifugalpumps are cheaper and more reliable. Reciprocating pumps are also a second choice to rotarypositive displacement pumps, described next, because of higher cost of installation andmaintenance.




Rotary Pumps

Rotary pumps are used for high-viscosity liquids. Two types of rotary pumps are used in SaudiAramco, gear pumps and screw pumps.

Gear Pumps

Gear pumps contain two intermeshing gears inside a casing. See Figure 17. Liquid enters thesuction side of the casing and flows into the spaces between the gears. Liquid cannot flowthrough the center of the pump because of the intermeshing gears. The rotating action of thegears carries the liquid around the outside of the gears in the spaces between the gear teeth andthe casing. On the discharge side, the liquid cannot flow back through the center because ofthe intermeshing gears. Therefore, it is forced out through the discharge nozzle.

External-Gear Three-Lobe

Gear PumpsFigure 17

A variation of the gear pump is the lobe pump, also shown in Figure 17. The principle ofoperation is the same. Only the shape of the rotating elements is different.

Gear pumps are used for high viscosity fluids at moderately low flow rates. They are oftenused when the viscosity is too high for good operation of a centrifugal pump. The upperviscosity limit for centrifugal pumps is about 500 cSt.

High viscosity actually helps the function of a gear pump. It minimizes backward slippagethrough the intermeshing gears. If a gear pump is used with a liquid with a low viscosity,slippage will increase and the volumetric efficiency will decrease.

Gear pumps are only suitable for moderate differential pressures up to about 500 psig. If ahigher differential pressure is required, a reciprocating pump should be used.

Typical gear pump efficiencies are shown in Work Aid 11 as a function of viscosity and head.




Screw Pumps

The screw pump contains two intermeshing screws inside a casing. One screw is rotated by amotor. Timing gears transmit this motion to the other shaft.

The principle of operation is as follows (see Figure 18). There is a space filled with liquidbetween each two threads of the screw. This space is an annular ring between the shaft and thecasing, but the ring is not complete. At the center where the screws intermesh, the space isblocked by the thread of the other screw.

As the screws rotate, these liquid-filled spaces move axially along the shaft. The liquid isforced from the suction nozzle at the end to the discharge nozzle at the center. The flow issmooth and nonpulsating.

Screw pumps are used for high-viscosity liquids. They are available for higher capacities thangear pumps. They can handle small amounts of solids without damage, if the solids are notabrasive. They can also tolerate some gas in the liquid. Because screw pumps are moreexpensive than gear pumps, they are used only for applications where they are particularlysuitable, for example, for asphalt shipping pumps.

A three screw pump (one drive screw and two idler screws), without a timing gear, is used forapplications with clean viscous fluids. They are frequently used in lube oil and seal oil servicesfor turbo compressor units.

Typical screw pump efficiencies are shown in Work Aid 12 as a function of viscosity anddifferential pressure with a correction for flow rate.

Discharge Flange

Bearing

Rotary ScrewsCoarse Pitch for LowDifferential Pressure

Suction Flange Packing Bearing

Timing Gears

Screw PumpsFigure 18




Pump Horsepower (All Pumps)

Brake horsepower is the power applied to the shaft between the pump and its driver. It iscalculated as follows for all pumps:

bhp =gpm( ) ∆P( )

1715 x Pump Eff.( )=

gpm( ) ∆H( ) S.G( )3960 x Pump Eff.( )

where:bhp = Brake horsepowergpm = Pump flow rate, actual gallons per minute.∆P = Differential pressure, psiPump Eff. = Hydraulic or volumetric efficiency of the pump, decimal fraction.∆H = Pump head, ft.S.G. = Specific gravity at flowing conditions, dimensionless

Pump efficiency is a characteristic of the pump. Typical values are 0.50 to 0.85 for centrifugalpumps and are 0.90 to 0.95 for reciprocating pumps. You can read the efficiency from themanufacturer's performance curve, at operating flow rate and head for centrifugal pumps. Youwill recall that driver power is :

Power kW( ) =bhp x 0.746Motor Eff.

where:kW = Power input to motor, kilowattsMotor eff. = Efficiency of the electric motor, as a decimal fraction. Typical values are

0.85 to 0.95. See Module 1, Work Aid 7.

The head produced is a characteristic of a centrifugal pump. It is a constant value. However,the ∆P produced is not constant. The ∆P varies directly with the specific gravity of the pumpedfluid. Also, if the specific gravity increases, the brake horsepower increases. Therefore, apump and driver set that has been designed for a liquid with a low specific gravity, such as alight hydrocarbon, may not have sufficient drive horsepower to pump water at the same flowrate. Because of the higher specific gravity, the horsepower requirement is greater and thedriver may be overloaded. Large changes in temperature can also make significant changes inthe flowing specific gravity due to expansion or contraction.




Net Positive Suction Head (All Pumps)

It is important that the fluid flowing through the centrifugal pump remain liquid at all points inthe flow path. If even a small portion of the liquid vaporizes, two problems result. First, thedensity of the fluid in the pump decreases, and the pressure differential developed willdecrease. Second, the presence of vapor bubbles in the pump can cause mechanical damage.

Frequently, a centrifugal pump is handling a liquid that is at its boiling point at the surface ofthe suction drum. This pressure is also called the "vapor pressure". It is necessary duringdesign to ensure that the actual pressure remains above the vapor pressure, at every pointthrough the flow path. The mathematical term used to cover this procedure is called NetPositive Suction Head or NPSH.

NPSH is the actual pressure of the liquid at the suction flange of the pump minus the vaporpressure of the liquid. In other words, it is the positive pressure above boiling pressure (vaporpressure). This pressure difference is expressed in feet of the liquid being pumped.

NPSH = (Actual Pressure) - (Vapor Pressure) + V2/2g

where:actual pressure = absolute pressure at pump suction, feetvapor pressure = vapor pressure at suction temperature, feetV2

/2g = velocity head term which is normally ignored, feet

A positive NPSH is required by all centrifugal pumps. The reason is that as the liquid entersthe pump, it is subjected to rapid acceleration by the spinning impeller. This accelerationdecreases the static pressure of the liquid. Vaporization occurs if the static pressure dropsbelow the vapor pressure.

Positive displacement pumps have NPSH requirements also. Available NPSH is calculated thesame way as for centrifugal pumps. For reciprocating pumps, a correction is necessary toavailable NPSH to account for pulsating flow. On each suction stroke, the liquid in the suctionline must be accelerated to a speed higher than average. The method for this correction isgiven in Engineering Data Book, Gas Processors Suppliers Association (GPSA), Pumps andHydraulic Turbines, Section 12. The amount of NPSH required is supplied by the pumpmanufacturer.




NPSH Available Versus NPSH Required

There are two kinds of NPSH. One is Net Positive Suction Head available (NPSHA), whichdepends on the design of the system, particularly on the elevation of the suction vessel abovethe pump and the friction drop in the suction line. The other is NPSH required (NPSHR), whichis the amount of net positive head required by the design of the pump to prevent vaporization.

The design and operation of the suction system to a centrifugal pump should be arranged sothat NPSHA is always greater than NPSHR. If it is not, cavitation damage to the pump or loss ofhead and capacity may occur. Saudi Aramco standards require a safety margin of 3 ft betweenavailable and required NPSH. The calculation procedure for NPSHA is shown in the followingexample, Figure 19. If the liquid is in equilibrium with the vapor in the section vessel, PS

equals PV which implifies the NPSHA calculation.

NPSHA = h + Ps - ∆Pf - PV[ ] 2.31S.G.

NPSHA = Available NPSH, ft.h = Height of Minimum Level Above Suction Flange of Pump, ft.Ps = Pressure in Vapor Space of Suction Vessel, psia∆Pf = Friction Loss in Suction, Including Contraction, psiPv = Vapor Pressure of Pumped Fluid at Flowing Conditions, psiaS.G. = Specific Gravity of Liquid at Flowing Conditions

Note: Elevation head “h” is negative when the liquid level is below thecenterline of the pump.

Ps = Pv when vapor and liquid in equilibrium at suction vesselsurface.

Calculation Procedure for NPSHA

Figure 19




NPSHR is a function of pump design and flow rate through the pump. NPSHR is always shownon the manufacturer's performance curve. For any pump, NPSHR increases as flow rateincreases. A typical centrifugal pump relationship is shown in Figure 20. Note that the NPSHR

can rise steeply at flow rates higher than design. Pumps have experienced cavitation by runningthem at too high a rate because the NPSHR increases with rate.

BEP = Best Efficiency Point (usual design point) for Centrifugal Pump

Average NPSHR as a Function of Pump Capacity at Constant SpeedFigure 20




The NPSHR for a centrifugal pump is defined by the specific speed relationship as follows:

NPSHR =

n Q0.5( )S

4 / 3

where:NPSHR = required NPSH, feetn = rotational speed, RPMQ = flow rate, gpm (to pump suction)S = suction specific speed

The suction specific speed is a property of the pump that is primarily dependent on the impellerinlet and suction inlet design. Normal suction specific speed design values range from 6,000 to12,000. Inducers and other special designs can obtain lower NPSHR. An inducer is a smallaxial pump that adds pressure at the suction of the pump.

Since NPSHR is a function of rotational speed and flow rate, changes can be expressed as:

NPSHR2 = NPSHR1 n2n1

4 / 3

at given Q

NDSHR2 = NPSHR1 Q2Q1

2 / 3

at given n

To reduce NPSHR for centrifugal pumps, manufacturers have designed impellers with largeeyes which reduce the suction velocity. However, at low flows dynamic flow problems incentrifugal pumps are magnified in large eye impeller pumps. At 60% of the best efficiencyflow rate, fluid can recirculate through the large eye accompanied by low flow hydraulicpulsatins of about on to eight hertz. This results in low flow in large eye centrifugal pumps thathave a higher NPSHR. The type of cavitation in large eye centrifugal pumps at low flow cannotbe readily detected because it does not have the typical cavitation symptom of reduceddischarge head. Cavitation damage from low flow in large eye pumps usually shows up on theback side of the impeller eye blades (invisible side).




Cavitation

Cavitation occurs when the NPSH available is less than that required. As the liquid flowsthrough the pump and decreases in pressure, small bubbles of vapor form in the suctionpassages. As soon as these bubbles reach a higher pressure in the impeller, they canrecondense and collapse so quickly that a violent force is imposed on the impeller. Thepressure generated by the collapsing bubbles can be over 10,000 psi and can physically removemetal from the impeller and in some cases the case of the pump. This makes a distinctive noisethat sounds like the rattling of stones in the pump. If cavitation continues, the damage can besevere.

Cavitation damage is most likely with single-component liquids such as water. Single-component liquids tend to recondense very suddenly. Multi-component liquids recondensemore gradually and therefore cause less damage. However, even with multi-componentliquids, the presence of vapor in the impeller can decrease the head or flow capacity.

Dissolved Gases - In addition to vaporization of the major component of the pumped liquid,dissolved gases can also vaporize, for example, air in water or nitrogen in hydrocarbons. Asthe pressure drops in the suction passages, small bubbles of dissolved gas can form. However,these gases do not condense and collapse suddenly. They redissolve quite slowly. Becausesudden collapse does not occur, the impeller damage does not occur. Furthermore, since theamount of gas released is small, the head produced by the pump is usually not affectedsignificantly. Therefore, when you calculate the vapor pressure of a liquid to be pumped, youcan usually ignore these dissolved components such as air, nitrogen, and hydrogen.

Sometimes, dissolved gases can even be beneficial. For example, if a pump operating on waterhas severe cavitation, one remedy is to inject a small amount of nitrogen or air into the pumpsuction. This gas remains as bubbles as the pressure increases. The bubbles cushion theimploding force of the condensing bubbles of water vapor.




Centrifugal Pump Calculations

Performance Curves

Analysis of an existing pump and prediction of its performance are done by means of themanufacturers performance curves. For a typical example of this curve see Figure 21. Themost important curve is head versus capacity. If you know the head that a pump will produce,you can calculate the differential pressure that it will develop.

Note that the head is shown for a range of impeller diameters. Most centrifugal pumps can befitted with impellers of different diameter in the same casing. This flexibility is a way to adaptthe pump to a changed future service. Pumps are normally purchased with an impellersomewhere near the middle of the possible size range of impellers. Therefore, if a headincrease is required by changed operating conditions, a larger impeller can be installed.

Typical Performance CurveFigure 21

Curves do not normally show the effect of a change in speed because most pumps are driven byconstant-speed motors. If a pump is purchased with a turbine driver, a family of speed curveswill be provided.

A curve of horsepower versus capacity is also shown for the range of possible impellerdiameters. Note that this horsepower is valid only for the rated specific gravity. If the liquidbeing pumped has a different specific gravity, the horsepower will have to be corrected sincebrake horsepower is directly proportional to specific gravity.




The third major characteristic shown on the performance curves is NPSH required versus flowrate. This characteristic is independent of specific gravity, operating pressure, and impellerdiameter. Impeller diameter changes do not affect the geometry on the suction side of theimpeller.

A pump curve also shows the hydraulic efficiency of a pump for various flow rates andimpeller diameters. The point of maximum efficiency is called the Best Efficiency Point(BEP). It should be somewhere near the design operating point for the pump but depends onhow the pump was selected. Remember, pumps are not generally custom designed!

Viscosity

Centrifugal pump performance curves are based on tests performed with water. When viscousfluids are pumped, head, capacity, and efficiency are all reduced. This effect becomessignificant at about 5 cSt. Correction factors for the affected variables are shown in Figure 22which is in a larger working size in Work Aid 3 and Addendum, page 180. Viscositycorrections can be significant.

1. Enter Chart at Design Capacity and Move Up To Design Head (For Multi-Stage Pumps, Use Head PerStage).

2. Move Horizontally To The Fluid Viscosity And Vertically To The Correction Curves.

Centrifugal Pump Viscosity Correction FactorsFigure 22




The Figure 23 pump curve illustrates the effect of viscosity. The original pump curve was forwater which has a viscosity of 30 SSU. Increasing the viscosity to 1000 SSU decreases thehead, decreases the efficiency, and increases the horsepower requirement. The Addendum (pp. 181-183) explains how this adjustment is made using Figure 22, Work Aid 3 or Addendum,page 180.

Pump CurveFigure 23




Affinity Laws

The relationships of impeller diameter and speed to flow rate, head, and horsepower arecommonly called affinity laws and also hold for centrifugal fans.

The effect of a change in impeller diameter of a pump can be estimated by the following:

Q2 = Q1D2

D1

H2 = H1D2

D1

2

HP2 = HP1D2

D1

3

where:Q = flow rate, gpmH = pump head, feet of fluidHP = horsepower requirement of pump, bhpD = diameter of impeller, inches1 = initial condition2 = final condition

Care must be taken when increasing the diameter of an impeller to make sure that the motor isadequate. Estimated performance changes can be made using these relationships, but rememberthey are approximate and should be verified by the manufacturer’s performance curveswhenever possible.

The effect of a change in rotational speed can be estimated by the following:

Q2 = Q1n2

n1

H2 = H1n2

n1

2

HP2 = HP1n2

n1

3

where:Q = flow rate, gpmH = pump head, feet of fluidHP = horsepower requirement of pump, bhpn = rotational speed of pump, RPM1 = initial condition




2 = final condition




Example Problem 1 - Use affinity laws to estimate centrifugal pump performance.

The 14.75" diameter pump impeller was tested at 1800 RPM as follows:

Q(gpm)

H(ft)

HP(ft)

Efficiency(%)

4000 157.0 189.5 83.73500 183.5 185.0 87.63000 200.5 174.5 87.02000 227.0 142.3 78.41000 228.5 107.0 54.0

0 230.0 76.5 0

A. Find the performance at 1600 RPM.

Q2 = (1600/1800) 4000 = 3,556 gpmH2 = (0.8889)2 157 = 124.0 ftHP2 = (0.8889)3 189.5 = 133.1 bhp

Q(gpm)

H(ft)

HP(ft)

Efficiency(%)

3556 124.0 133.1 83.73111 145.0 129.9 87.62667 158.4 122.6 87.01778 179.6 99.9 78.4

889 180.5 75.2 54.00 181.7 53.7 0




B. Find performance at the following conditions:

1800 RPM with 14" impeller 2000 RPM with 13" impellerQ2 = (14 / 14.75) 4000

= 3,797 gpmH2 = (0.9492)2 157 = 141.5 ftHP2 = (0.9492)3 189.5 = 162.1bhp

Q2 = (2000 / 1800) (13/14.75) 4000= 3,917 gpm

H2 = (0.9793)2 157 = 150.6 ftHP2 = (0.9793)3 189.5 = 178.0bhp

Q(gpm)

H(ft)

HP(ft)

Q(gpm)

H(ft)

HP(ft)

3797 141.5 162.1 3917 150.6 178.03322 165.3 158.2 3427 176.0 173.72847 180.6 149.2 2938 192.3 163.91898 204.5 121.6 1959 217.7 133.6949 205.8 91.5 979 219.1 100.5

0 207.2 65.4 0 220.6 71.8Centrifugal Pump Control Systems

The most common control device for a centrifugal pump is a control valve in the discharge line.This valve controls the amount of liquid delivered to the process. This valve takes a pressuredrop equal to the difference between the pressure supplied by the pump and the pressurerequired by the process.

A control valve is almost never used in the suction line of a pump. A pressure drop in thesuction line reduces NPSHA and could cause vapor to form, which is always harmful tocentrifugal pump operation.

Variable speed is an alternative method for controlling centrifugal pumps. The rotating speedis changed until the head generated by the pump exactly equals the head required. If the driveris a steam turbine or gas turbine, speed control is normally used. This is the case in manypipeline and production services in Saudi Aramco. It is always more efficient to controlproduced head than to control required head by throttling.

It is also necessary to control the minimum flow through a centrifugal pump. The minimumflow that can be tolerated is normally 25 to 30% of design flow to the pump. However, thisvalue can be considerately higher for pumps with double suction impellers (40 to 60% ofdesign flow). Below this flow rate, unstable operation can cause mechanical damage to thepump. If the flow rate required by the process is less than this minimum value, some excessflow is recycled from the discharge of the pump to the suction vessel. Recycle directly to thepump suction is normally not employed. This would increase the temperature of therecirculating fluid, leading to possible vaporization and pump damage.

Recycle can be controlled in the three ways shown in Figures 24 and 25:




• Manually controlled recycle

• Automatic recycle control with a control valve in the recycle line.

• An automatic minimum flow controller installed in the pump discharge line. Itsenses the net flow rate through the pump and opens a path to the recycle line whenflow drops below a preset value.




Centrifugal Pump Control Systems, Cont’d

Methods to Protect Against Low-FlowFigure 24

If natural flow balancing cannot be guaranteed, use separate flow controllers.

Or separate minimum flow recycle controls.

Controls - Pumps in ParallelFigure 25




Typical Centrifugal Pump Installation

Figure 26 shows a typical installation. Its elements are as follows:

• Normal operating pump

• The spare pump. Pumps are normally spared so that the process can operatecontinuously even if maintenance is required on one pump.

• Suction line with block valve for isolation.

• Discharge line with block valve for isolation.

• Check valve or non-return valve in the discharge line. This valve prevents reverseflow through the pump. Reverse flow would cause the impeller to spin backwards,which would damage the pump.

• Pressure gauge, PI, in the discharge line. This is to monitor the performance of thepump.

• Flushing connection to the seal. Normally a liquid is circulated through the seal tokeep it clean and cool.

• Casing vent. Before a centrifugal pump is started, be sure to vent vapors from thecasing. A centrifugal pump containing vapor will not develop differential pressure.The vent may be on the casing itself or on the discharge line.

• Kickback line or recycle line. This is the line used to keep the flow rate through thepump above the minimum value.

• Suction strainer. A suction strainer is installed upstream of the pump. It preventssolid material from entering the pump. Solid material could cause mechanicaldamage. Normally, the suction strainer is in place only during startup and is removedafter an initial period that flushes construction debris from the suction system.

Note: If the strainer is not to be removed, a differential pressure gauge should beinstalled around the suction screen.




Vent to Suct. Vess.

(If Pump Self-Venting)

~ ~~Suction Recycle (Kickback)to Suction Vessel

Discharge

Flow ControllerSet for

MinimumPump Rate

Pl

Pl

Casing Vent

MAIN(Operating)

SPARE(Standby)

Spool Piece for Suction Strainer(Strainer Installed During Startup)

Flushto

Seal

Drain

Typical Centrifugal Pump InstallationFigure 26




Starting a Centrifugal Pump

The normal method for starting a centrifugal pump is as follows. Before startup, close both thedischarge and suction block valves. Close the casing vent. Open the valve in the line to theseal.

1. Open the suction block valve to allow liquid to enter the pump.

2. Open the casing vent to release trapped gases or vapors.

3. Close the casing vent.

4. Start the pump motor; observe the pressure rise in the discharge line as indicated by thePI.

5. When the discharge pressure reaches the normal value, start to open the discharge blockvalve.

6. Gradually open the discharge block valve until it is fully open. If the discharge pressurestarts to fall, close the block valve a small amount to reestablish discharge pressure.(Reduce rate to prevent vaporization in pump suction.)




Operating Problems With Centrifugal Pumps

A list of the most common process problems is given below. For a more complete list,including mechanical problems, see GPSA Fig. 12-9.

SYMPTOM CAUSE CURE

Pump loses suction High-point pockets Modify piping so flow when flow rate in suction line. is continuously increases. horizontal or downward.

Low head, motor High viscosity. Heat fluid.Replace overload. pump and motor. Run

two pumps in parallel.

Pump loses suction Insufficient Vent casingbefore at start. venting of vapor. starting.

Cavitation noise or Insufficient NPSH. Raise suction liquid loss of capacity at level, reduce rate, high flows. new impeller.

Failure of mechanical Low flow of seal Adequate cooling seal; leakage. flush liquid. In- and flush. Proper

sufficient cooling stuffing box pressureof seals. and temperature.

High head, motor High S.G. Throttle flow. overload.




Control Methods For Positive Displacement Pumps

Positive displacement pumps are constant volume devices. (Remember that centrifugal pumpsare constant head devices.) See Figure 27.

Control of Positive Displacement PumpsFigure 27

A safety valve is always required for positive displacement pumps. Its discharge should besent back to the suction vessel to avoid overheating the pump fluid when the safety valve isoperating.

With a positive displacement pump, the head produced by the pump will rise to meet therequirements of the system. If the discharge flow is gradually restricted, the pump willcontinue to deliver the same volume. As the restriction increases, the discharge pressure willrise until one of the following happens:

• The motor overloads and stops.

• A relief valve discharges.

• A pipe or the pump casing ruptures.




Recycle Control

The most common control method for a positive displacement pump is to recycle a portion ofthe pump flow back to the suction vessel. See Figure 28. The pump is sized for approximately120% of the highest flow expected. The excess flow is recycled to the suction vessel by thethree way control valve.

By-Pass ControlFigure 28

For pumps with high differential pressure, greater than 300 psi, recycle control has twoproblems:

• There is a danger of backflow from the high-pressure system to the low-pressuresystem through the recycle port of the control valve. A check valve is provided toprevent backflow, but the check valve may stick open.

• The bypass port of the control valve operates with a very high pressure drop. Thebypass port, therefore, is a very small opening that can become plugged by smallsolid particles.

Variable Speed Motor Drive

This method of control for positive displacement pumps uses no bypass. This method is usuallybetter for services with high differential pressure.

Variable Piston Stroke Length

This method of controlling volume flow rate is used only for metering pumps.




CENTRIFUGAL COMPRESSORS

Centrifugal compressors are the basic compressor type used in Saudi Aramco. They cover awide range of capacity and head requirements. They can run for long periods of time betweenshutdowns for maintenance.

Centrifugal Compressor Components and Functions

Figure 29 shows the basic components of a centrifugal compressor. Impellers are mounted ona horizontal shaft. They are the primary rotating elements that impart velocity to the gas.Impellers are also called wheels. Diffusers are stationary elements mounted in the compressorcasing similar to a volute in a pump. There is one diffuser downstream of each impeller. Thediffuser converts velocity to pressure. Each diffuser is contained in a removable section of thecasing called a diaphragm. Each diaphragm also has a passage that directs the gas to thesuction of the next impeller. Each impeller and diffuser assembly is a stage of compression.

The shaft is supported at both ends by journal bearings. These are normally tilt-pad typebearings. Another bearing mounted on the shaft is a thrust bearing. The thrust bearing absorbsthe axial or horizontal force generated by unequal pressures on the impellers. A balance pistonmounted on the shaft neutralizes as much thrust as possible. This neutralization isaccomplished by connecting a high-pressure zone to one side of the piston and a low- pressurezone to the other side of the piston. The residual thrust is absorbed by the thrust bearing on theend of the shaft. This value changes as a function of compressor differential pressure(Discharge-Suction).

Case seals are located at each place where a shaft enters the casing. Normally there are twoseals for each casing. These seals usually contain pressurized oil to prevent the leakage of anygas from the inside of the compressor to the atmosphere. However, gas seals can also be used.These seals direct small amounts of leakage gas to flare (1 SCFM or less).

Internally, labyrinth seals minimize recirculation of gas from high-pressure zones to lowerpressure zones.




Centrifugal Compressor Components and Functions, Cont’d

With Permission from Mitsui Engineering Company

Thrust Bearing

Radial BearingDiaphragm

Casing

ImpellerSeal Ring

Shaft

Suction

Discharge

Basic Components - Centrifugal CompressorFigure 29

The casing of a centrifugal compressor is divided, or split, into halves that are held together bybolts. See Figure 30. This division permits access to the internal parts without disconnectingthe suction or discharge piping if the nozzles are mounted on the lower half of the casing. Thecasing may be split horizontally into an upper and lower half or it may be split vertically so thatone end of the compressor is removable. The vertical split is called a barrel compressor.




Centrifugal Compressor Components and Functions, Cont’d

Casing Designs for a Centrifugal CompressorFigure 30




Thermodynamic Equations For Gas Compression

Gas compression can take place in one of three separate paths. See Figure 31. The first suchmode is isothermal compression, compression taking place at constant temperature.

Actual Compression Path

Exp = I (Isothermal)

Exp = k (Isentropic)

Exp = n > k (Polytropic)

V1

V

P

P

P

2

1

Compression PathsFigure 31

Isothermal compression is not common in actual machinery because large amounts of heattransfer area must be supplied to keep the temperature constant. However, one can see that ifthe temperature were maintained constant, then pressure times volume would be a constantvalue at all points along the compression path.

PV = Constant (Isothermal Compression)

Isentropic compression is sometimes also called adiabatic compression, but its proper name isisentropic. As the name isentropic implies, this compression follows a path of constantentropy. It is, therefore, an ideal thermodynamic process. In this case, temperature is notconstant. It increases as the pressure increases because of the work of compression which isadded to the gas. The shape of the curve shown in Figure 32 is determined by the relationship:

PVk = Constant (Isentropic Compression)

The exponent k is equal to Cp/Cv, a common thermodynamic property of gases.

Figure 31 shows that the isentropic path results in a larger volume (higher temperatures) ascompression proceeds, compared to the isothermal path. This is because the greater rise intemperature results in a larger volume. Therefore, the exponent k is always larger than 1 fornonisothermal compression.




Polytropic compression is the compression path that occurs in a real centrifugal compressor.Centrifugal compression is not an ideal thermodynamic process. The inefficiency of thecompression process results in excess heat input to the gas. Therefore, temperature rises fasterthan it does in isentropic compression. The volume at the end of compression is again higherthan it was at the end of an isentropic path, due to the increased temperature of the gas.

Polytropic compression follows a path described by:

PVn = constant

The exponent n is always larger than the isentropic exponent k.

The actual compression path is the path plotted by P1T1, P2T2. The actual work can beexpressed by:

Actual Work = Isothermal WorkIsothermal Eff.

=Isentropic WorkIsentropic Eff.

=Polytropic WorkPolytropic Eff.

Centrifugal Compressor Head

The primary variable to be calculated for a compression service is the work or powerrequirement of the compressor. The equation for work is developed from three fundamentalthermodynamic relationships.

For isentropic compression:PVk = ConstantPV = ZRT

Work = P2

P1

VdP

If the proper substitution and integration are performed, the resulting equation for each stage ofcompression is:

Work =Z1RT1

MW(k −1)

k

P2

P1

k −1

k

−1

where:Z1 = Compressibility factor, at suctionR = Gas constant, 1545 ft-lb/lb mol -°FT1 = Suction temperature, °RMW = Gas molecular weightP1 = Suction absolute pressureP2 = Discharge absolute pressurek = Cp/Cv, average




The units of work in this equation are foot-pounds (force) per pound (mass). These units arecommonly simplified. The pound terms are implied and the resulting unit is feet. This workterm is then called head.

Head is energy, even though the common units for it are feet. Head is work per unit of mass.It is the work, or energy, needed to lift a unit of mass to a height which is equivalent to thehead.

The head developed by a centrifugal compressor is analogous to the head developed by apump. It can be compared to a column of fluid at the discharge of the compressor. Refer toFigure 32. Visualize a column of gas with the discharge pressure P2 at the bottom and thesuction pressure P1 at the top. The height of this column corresponds to the head required togenerate this differential pressure. The following relationship applies, which is the same as forcentrifugal pump head.

Head =∆P 2.31( )

S.G. relative to water

where:∆P = differential pressure (P2 - P1), psiS.G. = Specific gravity

The temperature and specific gravity vary along the height of the theoretical column, matchingthe temperatures along the compression path from suction to discharge. This is the reason whypolytropic head is greater than isentropic head for the same terminal pressures. Sincetemperatures are higher during polytropic compression, specific gravities are lower and ahigher column is required to achieve the same differential pressure.

• Head is Analogous to Centrifugal Pump Head- Column of Fluid at Discharge

Note: Temperatures along the theoretical column are those which occur during compression. Temperatures arehigher during polytropic compression, therefore Polytropic Head > Isentropic Head.




Compressor HeadFigure 32




Centrifugal Compressors Are Polytropic

Centrifugal compressors operate in a polytropic manner. The work input is greater than theideal amount. The temperature rise occurs at a faster rate than it does during isentropiccompression. This is accounted for mathematically by substituting the polytropic exponent nfor the isentropic exponent k. The following equation results:

Head =Z1RT1

MW(n − 1)

n

P2

P1

n−1

n− 1

A fixed relationship exists between n and k as shown in the following equation:

n −1n

=

k −1k

Polytropic Efficiency

Polytropic Efficiency

Polytropic efficiency is a characteristic of each compressor. Polytropic efficiency is equal toreversible work divided by total work applied to the gas. Reversible work and total work aredifferent because of the friction losses caused by the gas passing through the impellers and thediffusers at high velocity. For a centrifugal compressor, the polytropic efficiency is between60% and 85%.

Polytropic efficiency is shown on the manufacturer's performance curve. It varies with volumeflow rate and compressor speed. The manufacturer's curve is the best place to find thepolytropic efficiency to make calculations. If this is not possible, a reasonable approximationcan be made using the following formula.

Polytropic Efficiency = 0.0109 logbase e (ACFM) + 0.643

where:ACFM = Actual cubic feet per minute at suction condition

Note that above efficiency equation will give the efficiency at the machine's Best EfficiencyPoint (BEP). At speeds and flow rates above or below BEP, the efficiency will be lower like acentrifugal pump.




Compressor Discharge Temperature

The discharge temperature of a centrifugal compressor can be estimated using the followingequation.

T2 = T1P2

P1

n −1

n

where:T1 = Suction Temperature, °RT2 = Discharge Temperature, °RP2/ P1 = Compression ratio (absolute pressure), psia/psian = Polytropic exponent, dimensionless

This calculation of discharge temperature is approximate unless the compressibility factor (z) is1.0, because gas compressibility has an effect on temperature rise. If the compressibility is lessthan 1.0, the temperature calculated will be lower than the actual temperature.

Power Requirements

The energy that is imparted to the gas is called gas horsepower. Head is energy per unit ofmass flow assuming 100% efficiency. Horsepower is obtained by multiplying head times theweight flow and dividing by efficiency to obtain the actual energy imparted to the gas. Theproper conversion factor must also be included.

Gas Horsepower (ghp) = (Hpoly )

lbMASSmin

Poly. Eff. (33,000)where:

Hpoly = Polytropic Head, FT - LBF

LBH

or FT

Poly Eff. = Polytropic Efficiency, decimal fractionlbMASS

min.= Gas flow rate, pounds per minute

33,000 = Conversion Factor FT −− LB F

Horsepower - min.

Brake horsepower is the total horsepower required at the shaft of the compressor. This is equalto gas horsepower plus mechanical losses. Mechanical losses are caused by friction betweenthe rotating surfaces. To estimate mechanical losses see GPSA Engineering Data Book Figure13-38. To estimate total mechanical losses, add bearing friction losses to oil seal frictionlosses.

Work Aid 13 is a calculation form to facilitate the calculation of head, discharge temperature,and brake horsepower.




Casing Arrangements

Intercooling

Frequently, a compressor service requires two or more casings. The gas is cooled in betweencasings. The reasons for intercooling can be any of the following:

• To avoid exceeding a maximum temperature limit set by the mechanical parts or bythe seal oil.

• To reduce power requirements.

• Intercooling is convenient if additional casings are necessary because many impellersare required.

For calculations, each casing is treated as a separate compressor. Each casing is often referredto as a stage. This stage is a process stage and should not be confused with theimpeller/diffuser assembly stage discussed earlier. See Figure 33.

Casing Arrangements - IntercoolersFigure 33




Sidestreams

Sometimes additional gas is added to or extracted from a compressor casing between wheels(impellers). This is common practice with refrigeration compressors, where some gas isavailable at higher pressure. This gas is called a sidestream.. Sidestreams may also be takenout before discharge pressure is reached.

These sidestreams divide the compressor into sections. Each section must be calculated as aseparate compressor and has its own performance curve. See Figure 34.

Casing Arrangements - SidestreamsFigure 34




Performance Curves

Figure 35 shows a generalized performance curve. Performance curves contain the followinginformation:

• Head versus flow characteristic at several speeds

• Horsepower versus flow rate and speed

• The surge limit

a

Compressor PerformanceLow Compression Ratio

20 40 60 80 100 120

130

120

110

100

90

80

70

60

120

110

100

90

80

70

60

50

40

30

Percent of Design Inlet acfm

Generalized Performance CurveFigure 35




Manufacturers plot performance curves in several ways. The x axis may show actual cubic feetper minute or volume flow at standard conditions. The y axis may show polytropic head,pressure ratio for a particular gas, or discharge pressure for a particular gas and a particularsuction pressure. The most useful parameters on a performance curve are head and efficiencyvs. actual flow since they are relatively unaffected by gas composition or inlet temperaturechanges. Figure 36 shows a typical manufacturer's performance curve for a specificcompressor.

a

55000

50000

45000

40000

35000

30000

25000

20000

15000

10000

5000500 1000 1500 2000 2500 3000

Inlet V olume, CFM

76%

72%

72%

75%

60%

Ratio of Specific Heats 1.255Specific Gravity 0.700Suction T emperature 100ÞFSuction Pressure 500.0 psia

15700 rpm

15000 rpm

14000 rpm

68%13000 rpm

12000 rpm

11000 rpm

10000 rpm

75%

With Permission from Exxon Company, U.S.A.

Typical Manufacturer's Performance Curve - Head and EfficiencyFigure 36




Remember that the compressor always produces the same polytropic head at a given speed andactual volume flow.* If the gas composition or the suction temperature changes, then thepressure ratio and the discharge pressure will change. If the molecular weight of the gasincreases, the pressure ratio will increase. The horsepower required will also increase.

The polytropic efficiency for a machine is also constant at a given actual volume flow rate andspeed.

Manufacturer's performance curves are used for the following purposes.

• To determine whether a particular operation will be within the limits of the machine.The curve will tell you if an operating condition such as flow, gas composition,suction pressure or discharge pressure is feasible.

• To determine the correct speed for a set of process conditions such as suction ACFMand head.

• To determine the brake horsepower required for an operation, so that you can see ifthe driver will have enough power.

• To compare actual operating head and efficiency with the predicted values. Thisdetermines whether the machine is performing normally or whether it needsmaintenance.

Actual Volume

Manufacturer's curves and the machine's performance are based on actual volume flow at thesuction of the compressor. The units are actual cubic feet per minute. Process data is given instandard cubic feet per minute. To convert from standard cubic feet per minute to actual cubicfeet per minute, use the following equation:

ACFM = SCFM x 14.7P1

xT1

520x Z

where:ACFM = Actual cubic feet per minute at suction conditionsSCFM = Standard cubic feet per minute (60°F, 1 Atm)

SCFM =lb. molminute

x 379 or SCFM = lb./hr.

60 MW( ) x 379 MW = molecular weight( )

P1 = Suction pressure, psiaT1 = Suction temperature, °RZ = Compressibility factor, at suction conditions.

* This assumption is valid for gas density changes of 20%. Greater changes affect the head produced. In theseinstances, a new performance must be supplied by the original equipment manufacturer (OEM).




Centrifugal Compressor Fan Laws

Fan laws for centrifugal compressors are the same as the affinity laws for centrifugal pumps.The equations show the relationship between volume flow rate, head, horsepower, andcompressor speed. Impeller (wheel) diameters are not changed in centrifugal compressors.They can be used to predict performance at one speed if the performance at another speed isalready known. The equations are as follows:

Q2 = Q1

N2

N1

H2 = H1

N2

N1

2

, ∆T2 = ∆T1

N2

N1

2

bhp2 = bhp1

N2

N1

3

where:Q = Suction flow, actualH = Polytropic headbhp = Brake horsepowerN = Speed, RPM

These relationships are used to draw head and horsepower curves at speed N2, if the curve atspeed N1 is known. Start with any point on the head curve at speed N1. Calculate both H2 andQ2. This gives an equivalent operating point on the curve for speed N2. A series of thesepoints defines the curve for N2. Similarly, for the horsepower curve, calculate bhp2 and Q2 toobtain equivalent operating points.

Similar relationships exist for impellers of different diameters. However, centrifugalcompressor impeller diameters are very seldom changed in the field. Speed changes are muchmore common for centrifugal compressors.

It should be noted that the fan laws are reasonable approximations and do not include theeffects of gas density and multistage compressor performance. They can be used for estimatingpurposes only.




Surge

One important characteristic of a centrifugal compressor is its surge point. Surge is a conditionat which flow through the compressor becomes unstable. This condition must be avoided toprevent damage to the machine.

Surge occurs as follows: As the system resistance increases, a centrifugal compressor reacts bybacking up on its curve. That is, the flow decreases so that the head produced can rise to matchthe system demand. When the highest point on the compressor curve is reached, thecompressor cannot increase the discharge pressure further. At this point, the system dischargepressure is higher than the maximum possible discharge pressure of the compressor. The flowin the impellers becomes unstable and reverses, causing the discharge pressure to collapse.After a few seconds forward flow resumes. The discharge pressure rises again and the cyclerepeats every few seconds.

Surge occurs at a predictable flow rate. This flow rate is shown on the manufacturer's curve.In practice, controls are provided to keep the actual flow rate above this minimum value.

It is normal practice to take careful precautions to prevent surge. Surge disrupts the processand it can damage the compressor. As a result of the reversing flow, the direction of shaftthrust reverses. The temperature rises because the gas is internally recycled and recompressed.Compressor vibration and speed fluctuations are quite common. The reversing axial motion,high temperatures, and fluctuating pressure can also damage the compressor seals. In a severecase, failure of the seal or the thrust bearing, or even the impellers, can occur.

External piping can also be damaged. A check valve is normally installed at the discharge of acentrifugal compressor. During surge, this check valve can slam shut many times. This causesloud noise, pipe vibrations, and possible leaks at piping flanges.

Stonewall

Another phenomenon encountered in centrifugal compressors is stonewall. As the flow ratethrough the compressor increases beyond the design value, the amount of head developeddecreases. The greater the flow rate, the faster the developed head decreases. At a certainpoint the head developed drops to zero. This is called the stonewall condition. Stonewall is theresult of reaching sonic velocity in some part of the compression path, often in an impeller or adiffuser. Once sonic velocity is reached, the velocity cannot increase further and the headdrops to zero.




Efficiency of an Operating Machine

A process engineer is frequently asked to calculate the efficiency of an operating centrifugalcompressor. This actual efficiency can be compared with the predicted efficiency. If the actualefficiency is deficient, compressor maintenance is required to remove deposits in thecompressor or to replace damaged impellers, labyrinth seals or diffusers.

The definitions of efficiency are as follows:

Efficiency = Theoretical hp

Actual hp

Note that gas horsepower (ghp) is used, not brake horsepower (bhp). Mechanical losses are notincluded in efficiency, by convention.

Isentropic Efficiency = Minimum adiabatic work

Actual work, excluding mechanical losses

Polytropic Efficiency = Minimum work along polytropic path

Actual work, excluding mechanical losses

Procedures

There are four different ways to calculate operating efficiency to monitor performance. Theprocess engineer wants to know whether the compressor is fouling, or if there is mechanicaldeterioration due to erosion or corrosion. Method D is the best method.

Method A. Compare driver power output to compressor power input.Method B. Compare compressor's actual temperature rise to isentropic temperature rise.Method C. Using a Mollier chart, compare actual ∆h to isentropic ∆h.Method D. A computer program, such as COMPRESS.




Method D - Computer Program COMPRESS

The fourth method for calculating compressor efficiency is by using a computer program suchas COMPRESS. The input data are

• Compressor T1, T2, P1, P2

• Gas composition

• Gas flow rate information

The program calculates:

• Polytropic efficiency

• Gas horsepower

• Polytropic exponent n

• Polytropic head

The program uses an equation of state to calculate enthalpies and entropies at inlet and outlet.The COMPRESS program is the most accurate of the four methods. It is also the mostconvenient, when a PC is available. Other computer programs such as PRO/IITM may also beused.

The greatest source of potential error with a computer program is in the accuracy of input data.For critical calculations, calculate the power output of the driver as a check on the COMPRESScalculation. After accounting for mechanical losses in the compressor and for gear efficiency,the power output of the driver should match the power input of the compressor.




Control Schemes For Centrifugal Compressors

Variable Speed

A control system must match the performance curve of the compressor to the systemrequirements. One way to match the compressor performance and system requirements is touse a variable speed driver. Steam turbines and combustion gas turbines are usually capable ofspeed control. The range of control is normally from 80% to 105% of rated speed. Motorsnormally have a fixed speed, but they can be converted into variable speed devices by changingtheir electrical input frequency.

Figure 37 illustrates the principle of speed control. The solid line shows the head capacitycurve at design speed. The design point is on this curve. The desired operating point is at alower flow rate and a lower head. The objective is to find the operating curve shown by thedashed line which passes through that operating point. This operating curve will be at a newspeed N2, which is lower than the design speed N1.

Variable Speed ControlFigure 37




Suction Throttling

If a fixed speed driver is used, suction throttling is an alternative method to control compressorflow. Throttling the suction increases the actual volume of the gas and moves the operatingpoint away from the surge point. Suction throttling utilizes a butterfly control valve in thesuction line upstream of the compressor.

Figure 38 shows the principle of suction throttling control. The speed of the compressor isconstant; therefore, there is only one operating curve, shown by the solid line in the diagram.The operating point is matched to the operating curve by a different method. As the throttlevalve in the suction closes, the pressure downstream of the valve decreases. As the pressuredecreases, the volume of suction gas increases. At the same time, the compression ratiorequired by the machine is increasing because the discharge pressure remains constant whilethe suction pressure is dropping. This causes the actual operating point to move from point Ato point B.

Suction ThrottlingFigure 38




Discharge Throttling

Discharge throttling of a centrifugal compressor is not used as the primary control because itincreases the horsepower required from the driver and moves the compressor towards the surgepoint. However, discharge throttling is often used as a secondary control to prevent stonewall.See Figure 39.

System Head

Compressor Head

P1 P3 P2

Q ²P

²P

P3

P2

ACFM

Head

Discharge ThrottlingFigure 39




Antisurge Control

In addition to matching process flow to compressor capacity, the flow rate must be kept higherthan the surge point. This higher flow rate is accomplished by recycling a portion of thecompressor discharge flow back to the suction vessel. This practice keeps the flow through thecompressor above the minimum flow required to keep the compressor out of surge. Refer toFigure 40. A flow transmitter is located in the discharge line from the compressor. A signalfrom this flow transmitter controls the control valve in the compressor recycle line. If thedischarge flow falls below the minimum safe value, the recycle valve opens and maintains theminimum flow rate. The circuit must be arranged so that the recycle flow always flows througha cooler. Otherwise, the recycling gas would continue to be heated and exceed the temperaturelimits of the compressor.

Discharge

Recycle

Suction

Antisurge ControlFigure 40

In many installations a small computer is added to the controls. The computer calculates theactual surge flow at any moment. This flow rate is not a constant value, but can change withprocess conditions and gas composition.

Another requirement is that the recycle controller must respond quickly when the flow dropsbelow the minimum. Normal flow controllers experience reset windup. With reset windup, itcan take up to one minute before the control valve opens. A compressor recycle controllermust have special features to eliminate reset windup. In addition, the instrumentation usedmust have adequate accuracy and the control valve must open quickly (1-2 seconds).

With air compressors the recycle does not have to be returned to suction and can be vented tothe atmosphere.




Combined Controls

Figure 41 shows a combined control scheme for a typical refrigeration circuit. The controls areshown in a simplified manner, but they illustrate all of the principles mentioned so far.

This compressor has a side inlet or a second suction nozzle operating at a pressure higher thanthe first suction pressure. This is a common feature of refrigeration machines and divides thecompressor into two sections. The first section is between the first suction and the sidestreaminlet. The second section is between the sidestream inlet and the discharge. The flow rate forthe second section is different from the flow rate for the first section. Both flow rates must becontrolled to keep the two sections out of surge. Therefore, two flow sensors and two recycleloops are used. This compressor is assumed to have a constant-speed driver. Therefore, apressure controller in the suction line is the primary flow-control device. This pressurecontroller matches the compressor head to the head required by the system.

PC PC Recycle

SuctionKnockout Recycle 3 250 psig psig Accumulator EconomizerCooling Load 30 psig

Cooling Load

Combined Control Scheme - Refrigeration CircuitFigure 41

In the circuit shown, a second pressure controller is included in the discharge line. This issometimes necessary with multisuction machines. If this pressure controller is not included, thedischarge pressure will drop during cold weather, a condition that might be undesirable for thecompressor or the process. If the pressure drops very far, it may not be possible to keep all ofthe compressor wheels out of surge. Secondly, the operation of the external circuit (forexample, the economizer) would be affected when the discharge pressure drops. All thepressures in the machine drop, including all the pressure at the sidestream inlet. This couldhave an adverse effect on the process. Note that discharge pressure control is a secondarycontrol. The primary control for matching compressor and process is suction throttling.

Common Process Problems With Centrifugal Compressors

Work Aid 14 lists the most common problems encountered with centrifugal compressors. Thepossible causes of these problems are also listed.




POSITIVE DISPLACEMENT COMPRESSORS

Positive displacement compressors include reciprocating, lobe, screw, and sliding vane. Onlyreciprocating compressors will be discussed since these are the most common positivedisplacement compressors.

Reciprocating compressors are used in the following circumstances where centrifugalcompressors--the most common type--cannot be used or are not appropriate:

• Low flow rates, where centrifugal compressors are impractical or not economic.

• Very high discharge pressures, over 1000 psig.

• Gases with low molecular weight, below 15.

• Services where the molecular weight of the gas can vary greatly.

Reciprocating Compressors Principle Of Operation

Figure 42 shows the principle of operation for a reciprocating compressor. It is very similar toa reciprocating pump. A piston moves back and forth within a cylinder. Valves on the suctionand discharge sides of the cylinder open at the appropriate time to admit gas to the cylinder orto expel it through the discharge line. The valves are spring loaded and act like check valves.The springs help to make sure that the valves seat positively at the proper time. The valvesopen and close automatically as the gas pressures change.

a

PSuct

Suction V alve

Cylinder

Piston

Piston Rod

PDisch

Fs

Fs

DischargeValve

F = Spring Force on V alves

Reciprocating Compressor - Principle of OperationFigure 42




Reciprocating Compressors Mechanical Components

Figure 43 is a cross-sectional drawing of a reciprocating compressor showing the majormechanical components. The cylinder is the chamber in which the piston moves back andforth. The spring-loaded valves are mounted in the ends of the cylinder. Note that this cylinderis double-acting. Compression takes place on both the forward stroke and the back stroke ofthe piston. The piston is fitted with piston rings that provide a close fit between the piston andthe cylinder. The cylinder and cylinder heads contain cooling jackets. Cooling water circulatesthrough these spaces. The piston rod moves back and forth to drive the piston. The piston rodis the mechanical part that has the most stress. Packing seals the point where the piston rodenters the cylinder, to prevent loss of gas to the atmosphere.

Reciprocating Compressors - Components and FunctionsFigure 43

The connecting rod transmits motion from the crankshaft to the piston rod, through thecrosshead. The connecting rod converts rotary motion to reciprocating motion. The crossheadabsorbs the nonaxial forces from the connecting rod and transmits only axial forces to thepiston rod. The crankshaft is driven by a mechanical engine, which may be an electric motor,steam turbine, or a gas turbine. It may also be driven by a gas engine with reciprocatingpistons. In this case, a single crankshaft is connected to the pistons of both the engine and thecompressor. A single driver and crankshaft may be attached to several cylinders. Thecylinders may be successive stages of the same process gas service or they may be separateservices.




Reciprocating Compressor Lubrication

Compressor cylinders may be lubricated or nonlubricated. In the lubricated type, oil isinjected:

• Between the piston and cylinder

• Between the piston rod and packing

Lubrication reduces power requirements and temperature, which decreases the amount ofmaintenance required. However, some of the lubricating oil is always entrained in the gasstream. Oil separators at the discharge can remove most of this oil, but not all of it.

Nonlubricated cylinders may be used in services where oil contamination of the gas isundesirable. In this case, the piston rings and packing are made of low-friction materials suchas Teflon, and piston maximum speeds are less than lubricated applications (approximately75%). Examples of services where nonlubricated compressors may be used are:

• Instrument air, because oil in instrument air supplies can clog instruments.

• Oxygen, because mixtures of oil and oxygen are explosive.

• Refrigeration, because oil in refrigerant can freeze in low-temperature sections of theprocess.




Reciprocating Compressors Performance Calculations

The manufacturer's performance predictions should always be used as a first choice. If they arenot available, you can use the equations in the following sections to make reasonableapproximations. The equations for calculating gas horsepower and brake horsepower (whichare the same for any compressor -- positive displacement or centrifugal) are as follows:

Gas horsepower (ghp) = Head x flow rate lb./min.( )

Isentropic Eff. x 33,000

Brake horsepower = Gas horsepower

Mechanical efficiency

Typical isentropic and mechanical efficiencies for reciprocating compressors are given in WorkAid 16.

The isentropic formula for head is used. In this form, the gas property k is used directly ratherthan the exponent n, which is used in polytropic compression.

Head = Z1RT1

MW(k −1)

k

(r)k −1

k −1

r = P2

P1

where:Z1 = Compressibility factor, suction conditions.

R = Gas constant = 1545 ft.lb.

lb. mol- °FT1 = Suction temperature, °RMW = Molecular weightk = Cp/Cv, average of suction and discharge.

(Determine k from GPSA Figure 13-8 or 13-6 and 13-7)r = Compression ratio (absolute pressure)P1 = Suction pressure, psiaP2 = Discharge pressure, psia




Reciprocating Compressors Performance Calculations, Cont’d

The isentropic formula for head is used because reciprocating compression follows a path thatis very close to a true isentropic path. There are two reasons for this:

• The efficiency of compression is very high, approximately 90%. Therefore, theexcess heat resulting from inefficiency is small.

• The jackets surrounding the cylinder are cooled with cooling water. This providessome cooling for the gas during the compression stroke. This cooling offsets the heatgain from inefficiency. The net effect is a temperature rise that is almost the same asin isentropic compression.

The equation for discharge temperature is also of the isentropic form.

T2 = T1P2

P1

k−1

k

where:T = Temperature °R,P = Pressure, psiak = cp/cv

Subscripts1 = inlet2 = outlet

This equation gives the correct discharge temperature for most situations. However, in somecases, the efficiency of a reciprocating compressor can be somewhat lower, and the dischargetemperature will be higher than predicted by this equation. The conditions that might causelower efficiency are:

• Very low low-compression ratios, below 2 to 1.

• High-speed machines, which have proportionately higher valve losses.




Volumetric Efficiency

Figure 44 illustrates of the action of the suction and discharge valves during one completecycle of a reciprocating compressor. The volumetric efficiency of an operating compressor iscalculated to determine whether the valves and the piston are operating properly. An actualvolumetric efficiency significantly less than the theoretical value indicates that the valves or thepiston rings are leaking and that maintenance is required. The method is as follows:

(1) To calculate the actual volumetric efficiency from plant data, divide the suction flow rateby the displacement volume. The displacement volume is obtained from the following:

Single acting: D = A x m x Ls x n

1728

Double acting (without tail rod): D = 2A - a( ) m x Ls x n

1728

Double acting (with tail rod): D = 2 A - a( ) m x Ls x n

1728where:

D = Displacement, ACFM (actual cubic feet/minute)A = Cross-sectional area of cylinder, sq. in.a = Cross-sectional area of piston rod, sq. in.m = Number of cylindersLs = Length of stroke, in.n = Speed, strokes/minute, or RPM of crankshaft.

Action of Suction and Discharge ValvesFigure 44




A tail rod is an extra piston rod included in some compressors, on the side opposite the mainpiston rod. It helps to stabilize piston motion, and to reduce peak stress on the piston rod.

(2) Obtain the theoretical value for volumetric efficiency from the equation below. If theoperating conditions are those originally specified, the theoretical volumetric efficiencymay be available from the vendor specifications.

Volumetric efficiency is the volume flow rate of suction gas divided by the displacement. Fora reciprocating compressor, the theoretical volumetric efficiency is considerably less than100% because of clearance. Clearance is that portion of the cylinder not swept by the piston.Clearance includes volume at the end of the cylinder and underneath the valve chambers. Atthe end of a discharge stroke, the clearance volume is filled with gas at discharge pressure.During the subsequent suction stroke, this gas begins to expand. The suction valve does notopen until the gas in the clearance volume expands from discharge pressure to suction pressure.After that point, gas is admitted to the cylinder until the end of the suction stroke. However,gas is admitted during only 70 to 80% of the total suction stroke. The amount of lost suctionvolume depends on the compression ratio, the properties of the gas, and the amount ofclearance volume. The equation for calculating theoretical volumetric efficiency is as follows:

VE = 1.00− L − CZs

Zd

r

1

k −1

where:VE = Volumetric efficiencyC = Clearance volume fraction of displacement volumeZs = Compressibility factor, suction conditionsZd = Compressibility factor, dischargeL = Loss factor, for losses resulting from backflow through valves and around the

piston. See Work Aid 18r = Compression ratio (Pd/Ps)Pd = Discharge pressure, psiaPs = Suction pressure, psia




Reciprocating Compressors Intercoolers

An intercooler reduces the gas temperature between stages of compression. For mostreciprocating compressors, allowable discharge temperatures are limited to approximately350°F. Above this temperature, degradation of lube oil can occur. In nonlubricatedcompressors, damage to the piston rings and packing can occur even at temperatures less than300°F. The discharge temperature restriction places limits on the compression ratio that maybe used; if higher compression ratios are required, two or more stages of compression are used.

Figure 45 illustrates a common intercooler application, a utility air compressor. Utility air istypically compressed to about 100 psig. This requires a compression ratio of about 8. Withoutintercooling, the discharge temperature would be 550°F. In Figure 45, the total compressionratio is 9 due to the intercooler pressure drop and a discharge pressure of 105 psig.

IntercoolersFigure 45

When the air is cooled after the first compression stage, water condenses from the air and mustbe removed in a knockout drum.

NOTE: When you calculate head and power for two-stage or multistage compressors,calculate each stage separately. Remember that the intercooler and its piping will takesome pressure drop.




Reciprocating Compressors Control Systems

Suction Valve Lifters

Suction valve lifters are one technique for controlling volume flow. If the suction valve is heldopen continuously, gas will pass back and forth through the suction valve without passingthrough to the discharge line. This technique is used commonly to reduce the starting torque ofthe machine and for capacity control. However, it has three drawbacks:

• Capacity control is limited to 0 or 100% for each cylinder end.

• Valves may overheat because the same gas is continuously passing back and forthacross them.

• Forces on the crankshaft become unbalanced.

Clearance Pockets

Clearance pockets are another way to control compressor capacity. A clearance pocket is asmall volume just outside the cylinder. If this volume is open to the cylinder, it increases theclearance because this volume will not be swept by the piston. The increased clearance reducesthroughput. Saudi Aramco Design Practices SADP-K-403 permits the use of clearance pocketswith fixed volume but not variable volume. There may be one to four clearance pockets percylinder. Finally, clearance pockets may be controlled manually or automatically by a flowcontroller. See Figure 46.

Clearance PocketsFigure 46




Recycle

The third method for controlling reciprocating compressors is to recycle some of the dischargegas back to the suction. This is the least efficient control method because the compressor isalways operating at full capacity and consuming full power. However, it is often the mostreliable method from a mechanical point of view, because valve lifters and clearance pocketscan be causes of frequent maintenance. They tend to result in unbalanced loads on themechanical components. With gases containing dirty or fouling components, the mechanismsthat operate lifters and pockets can become fouled. It also is the method that can significantlyextend the turndown ratio of a compressor.

If recycled gas is employed, it is important to have a cooler in the loop, so that the same gas isnot recycled continuously which will result in increasing suction and discharge temperatures.

Variable Speed

Variable speed is sometimes used for capacity control, but not often. Saudi Aramco SADP-K-403 discourages the use of variable speed control.

Suction throttling, which is common for centrifugal compressors, is not used for reciprocatingmachines because it increases piston rod loads.

Reciprocating Compressor Process Problems

Liquid in Suction

Reciprocating compressors cannot tolerate liquid in the suction gas. The process must becarefully designed to prevent this condition. Small amounts of liquid continually entering thecylinders will damage the valves. Liquid will also wash lubricant away from cylinder walls,increasing wear. Large slugs of liquid can cause serious damage. Pistons or piston rods canbreak. In extreme cases, the cylinder head will blow off the compressor, and a fire will result.

To prevent liquid formation in suction lines:

• Locate the suction knockout drum close to the compressor.

• Avoid low points in the line between the drum and the compressor.

• Steam trace the suction line from the drum to the compressor to preventcondensation. Steam tracing is a small pipe carrying steam, located outside the gaspipe, under the insulation.

• Control cooling water temperature to avoid condensation within the cylinders.




Vibration of Piping

Vibration of the suction or discharge piping sympathetically with the movement of the pistonscan also be a problem. Two solutions are possible:

• Larger pulsation bottles can be used. Every reciprocating compressor has large,cylindrical pieces of piping called pulsation bottles or pulsation dampeners, at thesuction and the discharge. They smooth out the pulsations of the flow.

• The vibration may be due to a resonance between the piping and the compressors. Inthis case, the piping configuration can be changed or orifices added in order tochange the natural frequency of the piping system. Such action requires a pulsationanalysis to be performed to predict the action required.

Leakage of Valves and Piston Rings

The most common cause of low volumetric efficiency is leakage or back flow in the valves.Leaking piston rings are another cause. Valves and rings are normally replaced duringscheduled maintenance periods. Six months is a typical interval between planned maintenanceoperations. Significant wear and leakage can occur in less than six months. The most commoncause of rapid wear is the presence of liquid and/or solids in the gas.

Detecting Valve Leakage

A leaking suction valve is usually hotter than normal. This is caused by the back flow of hotgas through the valve during the discharge stroke. Discharge valves are always hot. So thismeans of detection cannot be used. On two-stage compressors, the interstage pressure is agood indication of valve leakage. If the interstage pressure is higher than normal, a valve orpiston ring in the second stage is leaking. If the interstage pressure is lower than normal, avalve or piston ring in the first stage is leaking.

Mechanical Problems

For a checklist of mechanical troubles, see GPSA Engineering Data Book, Figure 13 - 27.




STEAM TURBINES

Steam Turbine Introduction

Steam turbines are used by Saudi Aramco to drive electric generators, gas compressors, andcertain critical pumps. Steam turbines are prime movers. They convert the pressure and heatenergy of steam to mechanical energy (work). Turbines perform the opposite function of acentrifugal compressor.

Steam Turbine Principle Of Operation

The two major components of a steam turbine are nozzles and blades. The blades aresometimes called buckets. Nozzles are stationary; blades rotate. Steam contains energy in theform of pressure and temperature. Nozzles convert this energy into velocity energy. In anozzle, the pressure drops and the velocity increases (Figure 47).

Nozzle Bucket(Blade)

V1

V2

F

Steam Turbine - Principle of OperationFigure 47

The high-velocity jets from the nozzles strike the blades and cause them to move. In themoving blades, velocity energy is converted to mechanical work, or power.

Blades are located in rows on rotating wheels. Nozzles are arranged on stationary wheels,between the rotating wheels.

A stage contains one row of nozzles, followed by one row of blades (Figure 48). Turbinesmay be single-stage or multistage.

Nozzles and BladesFigure 48




Curtis Stage

A Curtis stage is a special kind of stage that takes a relatively high pressure drop. It is used forsingle-stage turbines and as the first stage in most older design multistage turbines. Present dayturbine design uses a Rateau stage since material and blade attachment methods allow higherblade operating stresses.

A Curtis stage has one row of nozzles, followed by three rows of buckets. The sequence is asfollows:

1. Nozzles2. Rotating buckets that develop power3. Fixed buckets that turn the direction of the steam4. A second row of rotating buckets, that develop more power.

All of the pressure drop takes place in the nozzles. Only velocity changes in the three rows ofbuckets.

Blades

BearingGovernor

NozzleGovernor

Valve

Seals

SteamInlet

SteamChest

With Permission From Elliot Company

Impulse Turbine With Single Wheel Curtis StagingFigure 48A




Other Types of Stages

In a multistage turbine, each stage after the first one has one row of nozzles (stationary) andone row of blades (rotating). These stages may be the "Rateau" type or the "reaction" type.

Model 2EPG4up to 4000 hp 4800 rpm Governor

GovernorValveSteam

Chest

BearingsJournalThrust

With Permission From Elliot Company

Multistage Mechanical Drive TurbineFigure 48B




Turbine and Cycle Efficiency

Turbine Isentropic Efficiency is the efficiency of the turbine. It is the actual work produced bythe turbine divided by the ideal or isentropic work that is expected for the given steamconditions.

Turbine Isentropic Efficiency = Actual Work

Ideal (Isentropic) Work

For Given Steam Conditions, P1, T1, and P2

where:T = Temperature °R,P = Pressure, psia

Subscripts1 = inlet2 = outlet




Turbine and Cycle Efficiency, Cont’d

Cycle Efficiency or thermal efficiency involves more than the steam turbine. A steam cycleincludes a steam generator, the turbine, and a means of disposing of the exhaust steam.Exhaust steam that will be used by another process is useful heat output. If exhaust steam iscondensed, the heat of condensation is lost. Cycle efficiency is defined as work output plusany useful heat output divided by the fuel fired in the steam generator.

Cycle Efficiency = Work + Useful Heat

Fuel Fired

Mechanical Components

The steam path through a turbine is outlined in Figure 49.

Steam TurbineSupply Exhaust Trip and Governor Steam Nozzles Throttle Valve Chest and Valve Blades

Hand Nozzles Valve and Blades

(Single Valve Turbines Only)

Steam Path through TurbineFigure 49

Trip and Throttle Valve

The trip and throttle valve is a manual (start up) valve and a safety device that shuts off thesupply of steam in case of a malfunction. The usual malfunctions are:

• Overspeed of the machine

• Loss of oil pressure

• High vibration

• An abnormal process condition

The trip valve takes a minimum pressure drop when it is open. The trip valve is sometimescombined with the governor valve.




Governor Valve

The governor valve is the main control for the rate of steam flow into the turbine. It acts withthe governor to maintain the speed of the turbine. The governor valve may be a single valve,or for more complex machines, it may be multiple valves. It may be operated mechanically orhydraulically.

Steam Chest

The steam chest is a chamber between the governor valve and the nozzles; in the steam chest,the steam pressure and temperature are at their highest values in the turbine.

Hand Valves

Hand valves may be provided on turbines that have a single governor valve. When the turbineis not operating at full load, the efficiency will be improved if some of the nozzles are closedoff. Hand valves are used for this purpose.

Calculations

Example Problem 2 - Theoretical & Actual Steam Rate, and Outlet Temperature

The method used for predicting turbine conditions uses the Mollier Chart for steam. Thefollowing example illustrates the calculation.

Given:Inlet steam pressure 600 psiaInlet steam temperature 700°FOutlet steam pressure 2 psiaTurbine efficiency 75%Brake horsepower required 1000 hp

Calculate:• Theoretical steam rate• Actual steam rate (water rate)• Steam outlet condition

+ temperature+ % moisture

See Work Aid 19A.




Answer:

Use the Mollier chart for steam (Figure 50) for a graphic illustration of this problem.

1. Locate the Inlet Steam Temperature and Pressure on the Mollier diagram.

Read inlet enthalpy, h1 1350 Btu/lb

2. Move vertically downward, along a line of constant entropy, to the outlet pressure of 2psia.

Read the outlet enthalpy, h2 923 Btu/lb

3. Calculate the isentropic (ideal) ∆h.

∆his = h1 - h2= 1350 - 923= 427 Btu/lb




Use of Mollier Diagram for Steam Turbine CalculationsFigure 50




4. The conversion factor from heat to work is:

2545Btu

hp − hrTherefore,

Theoretical Steam Rate, TSR =2545

Isentropic ∆h

=2545 Btu

hp − hr

427 Btulb

= 5.96 lbhp - hr

5. Actual Steam Rate, ASR (Water Rate)

ASR = TSRTurbine Efficiency

=5.96 lb/hp- hr

0.75

= 7.95 lb/hp/hr6. Calculate Steam Flow Rate.

Steam Flow Rate = hp x Actual Steam Rate

= 1000 hp x 7.95 lbhp - hr

= 7.950 lbhr

7. Outlet Steam Condition:Calculate actual outlet enthalpy

Actual ∆h = ∆his x Turbine Efficiency= 427 Btu/lb x 0.75= 320 Btu/lb

Actual h2 = h1 - Actual ∆h= 1350 -320= 1030 Btu/lbLocate the outlet steam condition onthe Mollier chart, at:

h = 1030 Btu/lb and 2 psiaRead Outlet Moisture Content = 8.4%Read Outlet Temperature = 130°F




Example Problem 3 - Efficiency of an Operating Turbine

The calculation of an operating turbine efficiency is also solved using the Mollier Chart (Figure51). The method is illustrated by the following:

Given, from plant operating data:Inlet Steam Pressure 400 psiaOutlet Steam Pressure 40 psiaInlet Steam Temperature 650°FOutlet Steam Temperature 320°FSteam Flow Rate 28,000 lb/hr

Calculate:Turbine Thermodynamic EfficiencyBrake Horsepower, bhp

Use Work Aid 19B.

Procedure:See Figure 51 for a graphic illustration of this problem.

Turbine Efficiency =Actual Enthalpy Drop

Isentropic Enthalpy Drop1. Locate steam inlet condition on the

Mollier chart at 400 psia and 650°F.Read h1. 1335 Btu/lb

2. Calculate isentropic ∆h from P1 to P2.Follow a constant entropy line verticallydownward to P2, 40 psia. Read h2is. 1127 Btu/lb

∆his = h1 - h2is

= 1335 - 1127= 208 Btu/lb

3. Calculate Actual ∆hact. On theMollier chart, locate the actual outletcondition at P2 = 40 psia, T2 = 320°F.Read h2 act. 1196 Btu/lb

Calculate ∆hact:∆hact = h1 - h2act

= 1335 - 1196= 139 Btu/lb




4. Calculate Turbine Efficiency.

Efficiency = ∆hact

∆his

= 139208

= 0.67 or 67%




aa

1390

1380

1370

1360

1350

1340

1330

1320

1310

1300

1290

1280

1270

1260

1250

1240

1230

1220

1210

1200

1190

1180

1170

1160

1150

1140

1130

1120

1100

1100

1090

1080

1070

440 °F

400 °F

360 °F

320 °F

280 °F

240 °F

Efficiency of an Operating TurbineFigure 51




5. Calculate Brake Horsepower

Water Rate =2545∆hact

=2545

Btuhp− hr

139 Btulb

= 18.3 lb/hp-hr

6. bhp =Steam Flow Rate

Water Rate

=28,000 lb./hr

18.3 lb/hp- hr

= 1530 hp

Efficiencies of Steam Turbines for Use in Calculations

Use efficiency curves provided by manufacturers, if they are available. If they are notavailable, estimates can be made from references. For multistage turbines, GPSA Figures 15-12, 15-13, and 15-17 give the basic efficiency of turbines operating at full power.

For factors to estimate the efficiency at less than full power, use GPSA Figure 15-11. Forsingle-stage noncondensing turbines, efficiencies can be obtained from Work Aid 20.




Use of Hand Valves to Maximize Turbine Efficiency

Turbines should always be operated at maximum efficiency to reduce the amount of steamrequired. If a turbine is operating at full power, the efficiency is determined only by the turbinedesign. At reduced power, the main steam valve, called the governor valve, is partially closed.This means that a pressure drop will be taken through the governor valve. This pressure dropdoes not provide any power; therefore, it reduces the efficiency of the turbine.

Downstream of the governor, the steam passes through several nozzles. Most turbines have ameans of closing off some nozzles when the turbine is not fully loaded. This increases thesteam flow through the nozzles that remain open. The result is a higher pressure drop throughthe nozzles. The steam governor valve will then have to open. The net result is a shift ofpressure drop from the governor valve to nozzles, which increases efficiency.

On large turbines, the multiple valves that block off some of the nozzles are all controlled bythe governor mechanism. They open and close automatically at the proper time. On smallerturbines, only the main valve is controlled by the governor. The others are hand valves. Thehand valve must be operated manually in order to achieve maximum efficiency. It must beeither full open or full closed; it is not designed to be a throttle valve.




THEORETICAL STEAM RATE TABLES

Theoretical steam rate tables are an alternative method to calculate steam rates without usingthe Mollier diagram. The GPSA Manual Figure 15-15 is such a table. Using a table of thistype, you can calculate the steam rate required for a given horsepower, if the steam inletconditions and the outlet pressure are known. Interpolation is required.

With steam rate tables, you cannot calculate the outlet steam temperature or steam quality.Also, it is not possible to calculate the efficiency of an operating turbine from plant data.Mollier charts must be used for these two calculations.

Performance Curves

Figure 52 is a typical performance curve for a condensing turbine. The curves for backpressureturbines are similar in format.

Steam Conditions:

Inlet 600 psig, 750°FExhaust 4 in Hg Absolute

Typical Turbine Performance CurveFigure 52




Performance Curves, Cont’d

Figure 53 is a performance curve for an extraction steam turbine. This is a family of curves tocover the variable of extraction rate. Curves for extraction turbines are usually plotted for onespeed only.

Steam Conditions:Inlet 600 psig, 750°FExhaust 4 in Hg AbsoluteExtraction 250 psig

Typical Performance Curve - Extraction Steam TurbineFigure 53

Common Operating Problems

Work Aid 21 summarizes the major process operating problems for steam turbines such as:

• Insufficient power developed• Low efficiency• Erosion of blades• Exhaust too hot• Vibration• Failure to start quickly on automatic cut-in




COMBUSTION GAS TURBINES

Combustion Gas Turbines Introduction

Gas turbines are used to drive process equipment and electric generators. Because they requirefew utilities, they are suitable for installation in remote locations. Models are available with awide range of horsepower. Gas turbines, unlike steam turbines, are not custom designed foreach application.

How a Gas Turbine Works

A gas turbine has three major components:

• Air compressor• Combustion• Power turbine

In Figure 54, the air compressor and the power turbine are mounted on the same shaft. Thetemperatures and pressures shown are typical values; however, there is a considerable range inthese values, depending on air compressor and combustor design.

Air from the atmosphere enters the inlet of the air compressor. The air compressor is usuallyan axial bladed compressor. At the outlet of the air compressor, the pressure is about 100 psigand the temperature has risen to about 400°F. The air compressor is usually an axialcompressor with 8 to 20 rows of blades. In small gas turbines (below 2000 bhp), the aircompressor can be a centrifugal compressor. Compression ratios vary from 5 to 30, though acompression ratio of 10 to 18 is most common. The air flows from the compressor to thecombustion.

In the combustion, fuel is added and combustion raises the temperature of the air toapproximately 1800°F. The temperature rise increases the volume of the air significantly,which greatly increases the amount of energy available in the air. Only a small part of theavailable oxygen is consumed in the combustion, because there is a limit on the temperaturewhich can be reached. The higher the temperature, the higher the efficiency and power output,but nozzle and blade materials limit the practical temperature to about 2300°F. A compressormay be required for a gas fuel.




How a Gas Turbine Works, Cont’d

The heated air fuel mixture flows to the power turbine. The power turbine is also an axialdevice, somewhat like a steam turbine. In the power turbine, the pressure is reduced from 100psig to near atmospheric pressure. Work is extracted from the air as it flows through the powerturbine.

Because the gas flowing through the turbine has been heated in the combustion, the energyavailable to the turbine is greater than the energy consumed by the air compressor. The netdifference between these two energies is available as shaft work to drive a machine.Approximately 60% of the total power produced by the gas turbine is required to drive the aircompressor. The power turbine is usually an axial flow turbine, with 2 to 6 rows of blades.

AtmosphericAir

Fuel

Air

Compressor

Combustor

400ÞF 1800ÞF

0 psig900ÞF

PowerTurbine

Temperature

Entropy

100 psig

Work

Note: Temperatures and Pressures are Typical

How a Gas Turbine WorksFigure 54




Gas Turbine Types

Heavy Duty

Heavy duty gas turbines are designed to run approximately three years continuously without ashutdown for maintenance. To achieve this goal, heavy duty turbines are conservativelydesigned. They operate with relatively low combustion temperatures and therefore have lowerefficiencies (20-30%). They are available in a wide range of sizes including very large modelsproducing over 200,000 bhp.

Aircraft Derivative

Another type of gas turbine is similar to aircraft jet engines. It is lightweight and compact. Forthis reason, it is frequently used on offshore platforms. These machines are designed tooperate with high temperatures to achieve high efficiency (30-40%). As a result they haveshorter run lengths between overhauls.




Gas Turbine

With Permission from Solar Turbines Inc., a Division of Caterpillar

Gas Turbine InternalsFigure 55




Gas Turbine Configurations

Single-Shaft

A single-shaft gas turbine has the air compressor and the power turbine on the same shaft,running at the same speed (Figure 56). This type is best for constant speed applications.Therefore, it is the type commonly used to generate electric power. It is not usually used formechanical drive (pump or compressor) applications since the starting power is much greaterthan a generator.

PowerTurbine

LoadAirCompressor

Single Shaft Gas TurbineFigure 56

On most modern combustion gas turbines the air compressor is designed for maximumefficiency at design speed. As a result, it will surge if the speed is lowered significantly. Thisis why a two-shaft design is needed for variable speed drive.

Dual-Shaft

A dual-shaft gas turbine has the air compressor and the high pressure turbine that drives itmounted on one shaft. See Figure 57. A second low pressure turbine, commonly called thepower turbine, and the load are connected to a second shaft. Because there are two shafts, thecompressor and the power turbine can operate at different speeds. This makes the turbinesuitable for variable speed applications. It is used to drive process equipment such as pumpsand compressors. Since the high pressure turbine is not connected directly to the load, thestarting horsepower requirement is considerably less than a single shaft turbine.

Compressor Comp.Turbine

PowerTurbine

Load

Dual Shaft Gas Turbine




Figure 57




Available Models of Gas Turbines

A wide range of gas turbines is available to the industry with horsepowers ranging from 700 to200,000. Approximately 20 different manufacturers make gas turbines. See the GPSA ManualFigure 15-33 for a partial list of available models.

Saudi Aramco uses combustion gas turbines in the following services:

• Electric power generators

• Pipeline pumps

• Water injection pumps

• Offshore platforms

Fuels For Gas Turbines

Gas turbines can operate with a wide variety of fuels, both gases and liquids. The mostcommon are:

• Natural gas

• Mixed refinery gases, H2 and C1 to C5

• Kerosene

• Diesel fuel

It is also possible to burn heavier liquids, such as crude oil and heavy fuel. the combustorsmust be designed for the actual fuel which is used. Fuel pressure must be high enough to passthrough a control valve and then enter the combustion. The combustion operates at thedischarge pressure of the air compressor. For liquid fuels, the gas turbine installation caninclude a fuel pump. Gas fuels must be supplied at the required pressure or a fuel gas boostercompressor must be used.

Combustion takes place in the very high excess air to reduce power turbine blade temperatures.The exhaust gas is usually about 16% oxygen.




Gas Turbine Efficiencies

Cycle Efficiencies

Gas turbine thermal efficiencies are highly dependent on the heat recovery. Types of heatrecovery include:

• Regenerative cycle The effluent is exchanged with the compressed air to thecombustion which reduces the fuel required.

• Exhaust heat recovery Effluent heat recovery such as a waste heat boiler is used.

• Combined cycle The effluent is used to generate steam which is used topower a steam turbine.

• Supplemental firing The effluent is used as preheated air to a heater since theoxygen content is typically 16%.

Site Ratings

Other site specific factors can also effect gas turbine performance. In general lower air densitywill reduce the power output. The standard ratings used by manufactures are at theseconditions.

• Ambient air temperature - 59°F.

• Altitude - sea level.

• Ambient air pressure - 29.92 in Hg.

• Inlet and exhaust pressure losses - none.

• Natural gas fuel with a specific heating value

To calculate the Site Rated Power, one starts with Standard Rated Power and makes correctionsfor site conditions. Manufacturers will usually supply curves to make these corrections fortheir turbine. GPSA has general curves in Figures 15-29 to 15-32 which can be used to makeapproximations.




Thermal Efficiency of Gas Turbine

Thermal efficiency = 2544

Heat rate, Btu/hp- hr

= 3414

Heat rate, Btu/kW - hr

Heat rate is affected by:

• Inlet and outlet pressure losses

• Ambient air temperature

• Percentage of rated load

Note that the heat rate is not affected by altitude (inlet pressure).

Gas Turbine Performance Curves

The information normally provided on a manufacturer's performance curve is as follows:

• Effect of altitude on maximum power output.

• Effect of inlet air temperature on maximum power output, heat rate, and air flow rate.

• Effect of percentage load and speed on the heat rate and exhaust temperature.

Work Aids 22-26 are the manufacturer's curves for a General Electric Frame 5 turbine, dualshaft. They can be used in the Exercises.




Auxiliary Equipment

In addition to the major components, there are a number of auxiliary items in a gas turbineinstallation.

• A common lubrication system for all of the rotating components. The system willcontain a reservoir, circulating pumps, coolers, and piping to the various bearings.For aeroderivative gas turbines, a separate lube oil reservoir is required since the oilused is different than the driven load lubricating oil (synthetic vs. mineral).

• Air filter. It is very important that the air to a gas turbine be clean. Therefore, amajor component, particularly in desert environments, is the air filter, which removesairborne solid particles. The primary air filter is often an inertial device to removelarge particles. If a significant number of small particles are present there will be asecond stage containing a fabric filter medium.

• A new type of filter called pulse clean is becoming popular. In this system a pulse ofair is blown backward through one section, while the other sections are operatingnormally.

• Noise suppression. Gas turbines are inherently very noisy. Therefore, silencers areusually included to control the noise. There may be a silencer on both the inlet andthe exhaust. If low noise levels are important, the turbine may also have a cocoon, oracoustic enclosure around the casing.

• Starting systems. An auxiliary starting device is needed to get the air compressor upto minimum speed before fuel can be introduced. The starting device may be anelectric motor or a small turbine. A starting turbine can be driven by steam,compressed air, or natural gas. The starting motor may also be a diesel engine or agasoline engine.

Sometimes the starting turbine is a steam turbine that is also used during normaloperation. This turbine is called a "helper" and is used to increase the power outputof the installation.

Control Systems

There are two basic control systems. The first is the speed control during operation. If theturbine is variable speed, this controller is a speed governor. If the turbine drives an electricpower generator, the speed is fixed by its connection to the grid. Therefore, the primarycontroller determines the amount of load or the amount of power generated by the turbine.

The other control system is an automatic sequence controller. This controls the steps takenduring startup and shutdown to assure equal thermal growth of components. During startup,this system increases the speed and the load gradually through a programmed sequence.




Shutdown sequence is normally controlled only on large turbines. It decreases the loadgradually.




Example Problem 4 - Performance Curves for Gas Turbine Calculations

Basis

Use the information below to perform the required calculations.

A gas turbine has the following ISO ratings:hp 38,000Heat rate 8700 Btu/hp-hr

Inlet air filter takes a pressure drop of 6 inches water.

There is no pressure drop on the exhaust side.

The fuel is 100% methane gas (977 Btu/SCF, LHV).

Site Factors - 800 ft elevation and temperature 110°F.

The manufacturer's curves are (Work Aids 22, 23 and 24)

Determine:

1. What will be the maximum continuous power available?28,780 hp (See A next page)

2. What is the thermal efficiency of the turbine?0.276 (See A next page)

3. What will the fuel consumption be at maximum continuous power?271,500 SCF/hr (See D next page)




Example Problem 4, Cont’d

Correction factorsPower Heat Rate

Altitude correction (Work Aid 22) (800 ft) 0.97 1.00

Inlet Pressure Drop Effects (Work Aid 24)(6″ H20)

Power: 1.00 - 64

x 0.016 = 0.976

Heat rate: 1.00 + 64

x 0.006 = 1.009

Temperature factor (Work Aid 23) (110°F) 0.80 1.05

A. Site available power = 38,000 x (0.97) x (0.976) x (0.80)= 28,780 hp

B. Heat rate at SAP = 8700 x (1.00) x (1.009) x (1.05)= 9217 Btu/hp-hr

C. Thermal efficiency = 25449217

= 0.276

D. Fuel consumption = 28,780 hp x 9217 Btu

hp - hr x

1977 Btu/SCF

= 271,500 SCF/hr




REFERENCES

Saudi Armco Standards• SAES-G-005 Centrifugal Pump• SAES-G-006 Positive Displacement Pumps• SAES-K-402 Centrifugal Compressors• SAES-K-403 Reciprocating Compressors• SAES-K-501 Steam Turbines• SAES-K-502 Combustion Gas Turbines

Saudi Armco Design Practices• SADP-G-005 Centrifugal Pump• SADP-G-006 Positive Displacement Pumps• SADP-K-402 Centrifugal Compressors• SADP-K-403 Reciprocating Compressors• SADP-K-501 Steam Turbines• SADP-K-502 Combustion Gas Turbines

API Standards• API-610 Centrifugal Pumps• API-675 Positive Displacement Pumps• API-611 General Purpose Steam Turbines for Refinery Service• API-612 Special Purpose Turbines for Refinery Services• API-617 Centrifugal Compressors for General Refinery Service• API-618 Reciprocating Compressors for General Refinery Service• API-616 Type H Industrial Combustion Gas Turbines for Refinery Service (Heavy

duty)• API-679 Type G Aeroderivative (Lightweight) Combustion Gas Turbines For

Refinery Service

Vendor Bulletins• Elliot Bulletin H-31K Single Stage Turbines• Elliot Bulletin H-37B Multivalve Turbines

Engineering Data Book, Gas Processors Suppliers Association (GPSA),• Section 12 Pumps and Hydraulic Turbines• Section 13 Compressors• Section 15 Steam and Combustion Gas Turbines

Hydraulic Institute Standards




WORK AID 1: CENTRIFUGAL PUMP EQUATIONS

EQUATIONS FOR CALCULATION OF CENTRIFUGAL PUMP HEAD

• Pump ∆P = P2 - P1 (psig or psia)P1 = Suction pressureP2 = Discharge pressure

• Head (feet) = ∆P psi( )x2.31

sp.gr.sp. gr. = Specific gravity relative to water

• ∆P = Head (ft) x 0.433 x sp. gr.

• Density of water at standard temperature (60°F)

sp. gr. = 1.0Density = 8.33 lb/gal

= 62.4 lb/ft3= 350 lb/barrel= 2205 lb/metric ton

EQUATIONS FOR CALCULATING POWER

• bhp =gpm( ) x ∆P

1715 x Pump Eff.( )∆P = Differential Pressure, psi

• kW = bhp x 0.746Motor Eff.

kW = Operating Load of Motor, kilowatts




WORK AID 2: NPSHR VS. % CAPACITY

BEP is Best Efficiency Point.

Average NPSHR as a Function of Centrifugal Pump Capacity at Constant SpeedFigure 62




WORK AID 3: CENTRIFUGAL PUMP VISCOSITY CORRECTION FACTORS

1. Enter chart at design capacity and move up to design head (for multi-stage pumps, use headper stage).

2. Move horizontally to the fluid viscosity and vertically to the correction curves.

Centrifugal Pump Viscosity Correction FactorsFigure 63




WORK AID 4: AFFINITY LAWS

Q2 = Q1 x D2

D1

Q2 = Q1 x N2

N1

H 2 = H1 x D2

D1

2

H 2 = H1 x N 2

N1

2

bhp2 = bhp1D2

D1

3

bhp2 = bhp1N2

N1

3

where:Q = Flow Rate

H = Head Developed

bhp = Power Required

D = Impeller Diameter

N = Rotating Speed of Impeller

Subscripts1 = Before2 = After




WORK AID 5: TYPICAL CENTRIFUGAL PUMP INSTALLATION

aa

Vent to Suct. Vess.

(If Pump Self-Venting)

~ ~~Suction Recycle (Kickback)to Suction Vessel

Discharge

Flow ControllerSet for

MinimumPump Rate

Pl

Pl

Casing Vent

MAIN(Operating)

SPARE(Standby)

Spool Piece for Suction Strainer(Strainer Installed During Startup)

Flushto

Seal

Drain

Typical Centrifugal Pump InstallationFigure 64




WORK AID 6: GENERAL CHARACTERISTICS OF CENTRIFUGAL PUMPS

Pump Type andConstruction Style

DistinguishingConstruction

Characteristics

UsualOrienta-

tion

UsualNo. ofStages

RelativeMaint-enance

Require-ment

Comments

Capacity varies with headCENTRIFUGAL Low to Medium Specific

SpeedHorizontal

Single Stage Overhung,Process Type

Impeller CantileveredBeyond Bearings

Horiz. 1 Low Most Common Style Usedin Process Service

Two Stage Overhung,Process Type

2 Impellers CantileveredBeyond Bearings

Horiz. 2 Low For Heads Above SingleCapacity

Single Stage ImpellerBetween Bearings

Impeller BetweenBearings; CasingRadially or Axially Split

Horiz. 1 Low For High Flows to 1100Feet Head

Slurry Large Flow Passages,Erosion Control Features

Horiz. 1 High Low Speed; AdjustableAxial Clearance

Canned Pump and Motor Enclosedin Pressure Shell; noStuffing Box

Horiz. 1 Low Low-Head CapacityLimits for Models Usedin Chemical Services

Multistaged,Horizontally SplitCasing

Nozzles Usually inBottom Half of Casing

Horiz. Multi Low For ModerateTemperature-PressureRatings

Multistage Barrel Type Outer Casing ConfinesInner Stack ofDiaphragms

Horiz. Multi Low For High Temperature-Pressure Ratings

Vertical

Single Stage ProcessType

Vertical Orientation Vert. 1 Low Style Used Primarily toExploit Low NPSHRequirement

Multistage Process Type Many Stages, LowHead/Stage

Vert. Multi Medium High Head Capability,Low NPSH Requirement

High Speed Speeds to 23,000 RPM,Head to 5800 Feet

Vert. 1 Medium Attractive Cost for HighHead/Low Flow

Sump Casing immersed in Sumpfor InstallationConvenience and PrimingEase

Vert. 1 Low Low Cost Installation

Multistage Deep Well Very Long Shafts Vert. Multi Medium Water Well Service withDriver at Grade

Comparison of Pump Types and Construction Styles:General Characteristics

Figure 65




WORK AID 7: PERFORMANCE CHARACTERISTICS OF CENTRIFUGAL PUMPS

Pump Type andConstruction Style

CapacityGPM

Max.Head

Ft

Max.P2Psi

TypicalNPSH/Req. Ft.

Max.ViscosSSU

Effic-iency

%

SolidsToler-ance

Max.PumpingTemp. °F

CENTRIFUGALHorizontal

Single Stage Overhung 15-5,000 800 600 6-20 3000 20-80 Mod.High

850

Two Stage Overhung 15-1,200 1400 600 6-22 2000 20-75 Mod.High

850

Single Stage ImpellerBetween Bearings

15-40,000 1100 980 6-25 3000 30-90 Mod.High

400-500

Slurry 1000 400 600 5-25 3000 20-80 High 850

Canned 1-20,000 5000 10,000 6-20 2000 20-70 Low 1000

Multistaged, Horiz. Split 20-11,000 5500 3000 6-20 2000 65-90 Medium 400-500

Multistage, Barrel Type 20-9,000 5500 6000 6-20 2000 40-75 Medium 850

Vertical

Single Stage ProcessType

20-10,000 800 600 1-20 3000 20-85 Medium 650

Multistage 20-80,000 6000 700 1-20 2000 25-90 Medium 500

In-Line 20-12,000 700 500 6-20 2000 20-80 Medium 500

High Speed 5-400 5800 2000 4-40 500 10-65 Low 500

Sump 10-700 200 200 1-22 2000 40-75 Mod.High

Multistage Deep Well 5-400 6000 2000 1-20 2000 30-75 Medium 400

Note: These data are typical only. Many exceptional cases exist.

Comparison of Pump Types and Construction Styles:Performance Characteristics

Figure 66




WORK AID 8: CENTRIFUGAL PUMP SELECTION CHARTS

Figure 67




WORK AID 9: CENTRIFUGAL PUMP TROUBLES AND CAUSES

Trouble: Possible Causes: Trouble: Possible Causes:

1. Failure todeliverliquid:

a. Wrong direction of rotation.b. Pump not primed.c. Suction line not filled with liquid.d. Air or vapor pocket in suction

line.e. Inlet to suction pipe not

sufficiently submerged.f. Available NPSH not sufficient.g. Pump not up to rated speed.h. Total head required greater than

head for which pump is capable ofdelivering.

2. Pump doesnot deliverratedcapacity:

a. Wrong direction of rotation.b. Suction line not filled with liquid.c. Air or vapor pocket in suction

line.d. Air leaks in suction line or

stuffing boxes.e. Inlet to suction pipe not

sufficiently submerged.f. Available NPSH not sufficient.g. Pump not up to rated speed.h. Total head greater than head for

which pump designed.j. Foot valve too small.k. Foot valve clogged with trash.m. Viscosity of liquid greater than

that which pump designedn. Mechanical defects ...

(1) Wearing rings worn(2) Impeller damaged(3) Internal leakage resulting

from defective gaskets.o. Discharge valve not fully opened.

3. Pump doesnot developrateddischargepressure

a. Gas or vapor in liquidb. Pump not up to rated speedc. Discharge pressure greater than

pressure for which pump designedd. Viscosity of liquid greater than

that for which pump designede. Wrong rotationf. Mechanical defects…

(1) Wearing rings worn(2) Impeller damaged(3) Internal leakage resulting

from defective gaskets.

4. Pump losesliquid afterstarting:

a. Suction line not filled with liquid.b. Air leaks in suction line or

stuffing boxes.c. Gas or vapor in liquid.d. Air or vapor pockets in suction

line.e. Inlet to suction line not sufficient.f. Available NPSH not sufficient.g. Liquid seal piping to lantern ring

plugged.h. Lantern ring not properly located

in stuffing box.





5. Pumpoverloadsdriver:

a. Speed too high.b. Total head lower than rated head.c. Either or both the specific gravity

and viscosity of liquid differentfrom that for which pump is rated.

d. Mechanical defects ...(1) Misalignment(2) Shaft bent(3) Rotating element dragging(4) Packing too tight

6. Vibration: a. Starved suction.(1) Gas or vapor in liquid(2) Available NPSH not

sufficient(3) Inlet to suction line not

sufficiently submerged(4) Gas or vapor pockets in

suction lineb. Misalignment.c. Worn or loose bearings.d. Rotor out of balance.

(1) Impeller plugged(2) Impeller damaged

e. Shaft bent.f. Improper location of control valve

in discharge line.g. Foundation not rigid.

Check List for Centrifugal Pump Troubles and CausesFigure 68




WORK AID 9, CONT’D


7. Stuffingboxesoverheat:

a. Packing too tight.b. Packing not lubricated.c. Wrong grade of packing.d. Insufficient cooling water to

jackets.e. Box improperly packed.

8. Bearingsoverheat:

a. Oil level too low.b. Improper or poor grade of oil.c. Dirt in bearingsd. Dirt in oil.e. Moisture in oil.f. Oil cooler clogged or scaled.g. Failure of oiling system.h. Insufficient cooling water

circulation.j Bearings too tight.k. Oil seals too close fit on shaft.m. Misalignment.

9. Bearingswearrapidly:

a. Misalignment.b. Shaft bent.c. Vibration.d. Excessive thrust resulting from

mechanical failure inside thepump.

e. Lack of lubrication.f. Bearings improperly installed.g. Dirt in bearings.h. Moisture in oil.j. Excessive cooling of bearings.

Check List for Centrifugal Pump Troubles and CausesFigure 68, Cont’d




WORK AID 10: PUMP HEAD AND HORSEPOWER

50psigNLL

LLLEL=20 Ft.

DrumD-1

EL=100 Ft. 160psig

ColumnC-1

3 Ft.

Grade

E-1 E-2

600 gpmS.G. = 0.72Pump Eff. = 0.69

Figure 69

Line Lengths:Suction - 100 equivalent ftDischarge - 500 equivalent ft

Pressure drops:Suction line - 0.2 psi/100 ftDischarge line - 2.2 psi/100 ftE-1 - 22 psiE-2 - 17 psiControl valve - 20 psi minimumorifice - 1 psi

Liquid in Drum D-1 in equilibrium with vapor.




Calculate pump head, brake horsepower (required), and NPSHA.





Answer:

1. Calculate P2 Discharge pressure (Max.).a. Vessel pressure 160 psigb. Static head(100-3) ft x 0.433 x 0.72 + 30.2 psic. Friction pressure dropsE-2 + 17 psiE-1 + 22 psiControl valve + 20Flow orifice + 1 psi

Line = 500 x 2.2100

+ 11 psi

P2 = 261.2 psi

2. Calculate P1, suction pressure (Min.).a. Vessel pressure 50 psigb. Static head(20-3) ft x 0.433 x 0.72 +5.3 psic. Friction dropLine = 100 x 0.2/100 -0.2

P1 = 55.1 psig

3. Calculate ∆P.∆P = P2 - P1

= 261.2 - 55.1= 206.1 psi

4. Calculate head required.

Head =∆P 2.31( )

sp. gr.

=206.1 2.31( )

0.72

= 661 ft





5. Calculate brake horsepower.

bhp =gpm x ∆P

1715 x Eff.

=600 x 206.11715 x Eff.

= 104.5 hp

6. Check pump headTotal heat = ∆ Vessel pressure

+ ∆ Elevation+ Total friction drop

a. ∆ Vessel pressure(160-50) x 2.31 = 353 ft 0.72b. ∆ Elevation100 - 20 = 80 ft

c. Total friction drop

Line 2.2 x 500100

+ 0.2 x 100100

= 11.2 psi = 11.2 x 2.310.72

= 36 ft

Exch. etc. 22 + 17 + 20 + 1 = 60 psi = 60 x 2.310.72

= 192 ft

Total head required = 353 + 80 + 36 + 192 = 661 ft

7. NPSHA Calculation

NPSHA = h + Ps - DPf − Pv[ ] 2.31S.G.

Ps = Pv (Liquid in equilibrium with vapor)

NPSHA = 20 − 3( )− 0.2[ ] 2.310.72

= 17 - 0.64 = 16.36 ft




WORK AID 11: EFFICIENCY OF GEAR PUMPS

Efficiency of Gear PumpsFigure 70




WORK AID 12: EFFICIENCY OF SCREW PUMPS

Efficiency of Screw PumpsFigure 71




WORK AID 13: CENTRIFUGAL COMPRESSOR - CALCULATION FORM

Gas

MW

Suction Flow Rate: SCFM, lb/min., ACFM

P1 psia

P2 psia

r = P2/ P1 = =

T1 °F, °R

Polytropic Efficiency:

From Manufacturer's Specification

or: 0.0109 ln (Suction ACFM) + 0.643 = 0.0109 ln ( ) + 0.643 =

lst Trial 2nd Trial 3rd Trial

T2, °R assumed

k1 (GPSA Figure 13-8 or 13-6)

k2

k avg.

(k-1)/k

n − 1( )/n =k -1( )/k

Poly. Eff.

T2 = T1(r)(n-1)/n

T2 calculated




Z1 (GPSA 23-3)





Polytropic Head:

H poly = Z1(1544)T1

MW(n −1)

n

(r)n−1

n −1

H poly =( )(1544)( )

( )( )( )( ) −1[ ]

Hpoly = feet

Gas Horsepower:

ghp = Hpoly x lb/min

Poly Eff. x 33,000

ghp = ( ) ( )

( ) 33,000( )

ghp =

Mechanical Losses (GPSA Figure 13-38) hp

bhp = ghp + Mechanical Losses

= +

= hp




WORK AID 14: COMMON OPERATING PROBLEMS FOR CENTRIFUGALCOMPRESSORS

Common Problems Possible Cause(s)

Surge • Improper setting of recycle flow control

• Slow response of recycle controller and/or valve

• Deposits in rotor or diffuser

• Blockage in discharge line or recycle line

• Low gas density (Low MW or high suction temperature)

Driver Overload • High suction pressure

• High molecular weight of gas

• Low inlet temperature of gas

Vibration • Liquid in suction

• Deposits on rotor

• Rotor erosion/corrosion

• Mechanical problems (GPSA p. 13-39)

Tripout or • Liquid in suction knockout drumAutomatic Shutdown

• High discharge temperature

• Loss of lube or seal oil

• Loss of buffer gas

• High axial displacement

• Instrument malfunction, false trip

• High thrust bearing pad temperature

• Overspeed of driver (steam or gas turbine)




WORK AID 15: RECIPROCATING COMPRESSOR - CALCULATION FORM

(Page 1 of 2)

Gas

MW

Suction Flow Rate: SCFM, lb/min., ACFM

P1 psia

P2 psia

r = P2/P1 = =

T1 °F, °R

Isentropic Efficiency:

(From Manufacturer Specor from Work Aid 16)

lst Trial 2nd Trial 3rd Trial

T2, °F/°R assumed / / /

Tavg =T1 + T2

2,°F

k at Tavg. (GPSA Figure 13-8 or 13-6)

(k-1)/k

T2 = T1(r)(k-1)/k, °R

T2, °F calculated

Z1 (GPSA 23-3)





Isentropic Head:

H is = Z1(1545)T1

MW(k −1)

k

(r)k−1

k −1

H is = ( )(1545)( )

( )( )( )( ) −1[ ]

His = feet

Gas Horsepower:

ghp = His x

lbmin

Is. Eff. x 33,000

ghp = ( ) x ( )

( ) x 33,000

ghp =

Mechanical Efficiency (Work Aid 16)

Brake Horsepower. =ghp

Mech. Eff.

= ( ) ( )

=




WORK AID 16: TYPICAL ISENTROPIC EFFICIENCY OF RECIPROCATINGCOMPRESSORS

BHP ==W His

33,000 ηη is

ηηm

Typical Isentropic Efficiency of Reciprocating CompressorsFigure 72




WORK AID 17: RECIPROCATING COMPRESSOR - VOLUMETRIC EFFICIENCYCALCULATION FORM (PAGE 1 OF 3)

EXPECTED VOLUMETRIC EFFICIENCY:

P1 Suction pressure, psia

P2 Discharge pressure, psia

r =P2

P1

= ( )

( ) =

k, average

Zs Z at suction

Zd Z at discharge

C = Clearance volumeDisplacement volume

L Loss correction factor = (Work Aid 18)

VE = 1.00 − L − CZs

Zd

(r)1

k −1

= 1.00 − ( ) − ( )( )( )

( )1

( ) −1

=




WORK AID 17 (PAGE 2 OF 3)

ACTUAL VOLUMETRIC EFFICIENCY

Actual Suction Flow Rate:

Compressor Suction Flow: SCFM lb/min.

MW

Zs

T1 Temp. °F, °R

P1 Press. psig psia

Suction SCFM = lbmin

x379MW

= ( ) x 379( )

=

Suction ACFM = SCFM s14.7P1

xT1

520xZs

= ( ) x 14.7

( )x

( )520

x ( )

= ACFM




WORK AID 17 (PAGE 3 OF 3)

Displacement:

A, Cross-sectional area of cylinder sq. in.a, Cross-sectional area of piston rod sq. in.m, Number of cylinders Ls, Length of stroke in.

n, strokes/minute (RPM)

Single Acting

D =A( ) n( ) Ls( ) n( )

1728

D = ( ) ( ) ( ) ( )

1728= ACFM

Double Acting without Tail Rod

D =2A − a( ) m( ) LS( ) n( )

1728

D =2 ( ) − ( )[ ] ( ) ( ) ( )

1728= ACFM

Double Acting with Tail Rod

D =2A − a( ) m( ) LS( ) n( )

1728

D =2 ( ) − ( )[ ] ( ) ( ) ( )

1728= ACFM

Volumetric Efficiency:

VE =Suction,ACFM

D, ACFM

= ( ) ( )

=







WORK AID 18: LOSS CORRECTION FOR RECIPROCATING COMPRESSOR

0.15

0.13

0.11

0.09

0.07

0.05

0.031.0 2.0 3.0 4.0 5.0 6.0

Compression Ratio r

Vo

lum

etri

c E

ffic

ien

cy L

oss

Co

rrec

tio

n, L

Lin

es o

f C

on

stan

t In

let

Pre

ssu

re, p

sia

1000800600

400

200

100

50

15andlower

Source: Natural Gas Processors Suppliers Association Engineering Data Book, 1966

Loss Correction for Reciprocating Compressor Volumetric Efficiency CalculationFigure 73




WORK AID 19A: CALCULATION OF THEORETICAL AND ACTUAL STEAMRATES, AND OUTLET TEMPERATURE

Steam Conditions:P1: psiaT1: °FP2: psiaTurbine Efficiency: (from manufacturer’s curve or

GPSA, Figures 15-11, 12, 13, and 17)bhp required: Btu/lb

1. h1 (from Mollier) Btu/lb

2. Move vertically (isentropicaly) on Mollier from P1T1 to P2

h2 isentropic = Btu/lb

3. ∆his = h1 - h2

= (____) - (____)=

4. Theoretical Steam Rate:

TSR = 2545∆h is

= 2545( ) = lb

hp− hr

5. Actual Steam Rate:

ASR = TSRTurbine Eff. =

( )( ) = lb

hp − hr

6. Steam Flow Rate = hp x ASR= (____) x (____)= hp

Outlet Steam Conditions:

Actual ∆h = ∆his x Turbine Eff.= (____) x (____)= Btu/lb

Actual h2 = h1 - Actual ∆h= (____) x (____)= Btu/lb

On Mollier, locate P2, Actual h2, and read

T2 = __________ °F




% Moisture = __________




WORK AID 19B: CALCULATION OF TURBINE EFFICIENCY ANDHORSEPOWER FROM STEAM CONDITIONS

P1: psiaT1: °FP2: psiaT2: °FSteam Flow Rate: lb/hr

1. h1 (from Mollier P1T1) Btu/lb

2. Move vertically (isentropicaly) on Mollier from P1T1 to P2


3. h2 actual (from Mollier at P2T2) = Btu/lb

4. ∆h is = h1 - h2 is= (____) - (____)= Btu/lb

5. ∆h act = h1 - h2 act= (____) - (____)

= Btu/lb

6. Turbine Efficiency = ∆h act∆h is

= ( )( )

=

7. Water Rate = 2545∆h is

= 2545( ) = lb

hp− hr

8. bhp = Steam Flow RateWater Rate

= ( ) lb/hr ( ) lb/hp- hr

= hp




WORK AID 19C: CALCULATION OF STEAM TURBINE EFFICIENCY FROMINLET STEAM CONDITION AND BRAKE HORSEPOWER

P1: psiaT1: °FP2: psiabhp: hpSteam Flow Rate: lb/hr

1. h1 (from Mollier) Btu/lb

2. Move vertically on Mollier from P1T1 to P2


3. Æh isentropic = h1 - h2 isentropic = (____) - (____)

= _____ Btu lb

4. TSR = 2545 = _2545_ = _____ __lb_Æhis ( ) hp-hr

5. ASR = Steam Flow Rate bhp

= ( ) lb/hr = ________ lb _( ) hp hp-hr

6. Turbine Efficiency = TSR = _( ) = _________ASR ( )

7. Outlet Steam Condition:

ÆhActual = ÆhIsentropic x Turbine Efficiency

= ( ) x ( ) = ________ Btu/lbh2Actual = h1 - Æh actual

= ( ) - ( ) = ________ Btu/lb

8. On Mollier, locate point at P2,h2 actual.

Read % Moisture ______________ %T2 ______________ °F




WORK AID 20: TYPICAL SINGLE STAGE STEAM TURBINE EFFICIENCY

Correction Factors

EfficiencyCondition Multiplier

P1 = 600 psig = 0.80P2 = 50 psig = 1.12P2 = 0 psig = 0.90N = 1,800 RPM = 0.68

Typical Efficiencies, Single Stage Turbines, Non-Condensing"Normal Efficiency" Type

Figure 74




WORK AID 21: STEAM TURBINE TROUBLESHOOTING

STEAM TURBINESCOMMON OPERATING PROBLEMS

Problem Possible Cause

Insufficient Power Developed • Steam pressure too low.• Backpressure too high.• Supply temperature too low.• Deposits in steam path.

Low Efficiency • Deposits in steam path.• Erosion of nozzles or blades.• Hand valves open at reduced power.

Erosion of Blades • Too much moisture in turbine; inlet temperaturetoo low or outlet pressure too low.

Exhaust Too Hot • Low efficiency• Low steam flow rate

Vibration • Deposits• Erosion• Broken blades• Damaged bearings• Misalignment of piping

Failure to Start Quickly • Water in supply line; traps not working. on Automatic Cut-In




WORK AID 22: GAS TURBINE ALTITUDE CORRECTION FACTOR FOROUTPUT AND HEATING CONSUMPTION AND ALSO ALTITUDEVS. ATMOSPHERIC PRESSURE

a

1.00

0.90

0.80

0.70

14.7 14.0 13.0 12.0 11.00

2

1

3

4

5

6

7

8

Atmospheric Pressure - psiaGas T urbine Dept.

With Permission from General Electric Company, K.D. Knapp Sept. 21, 1970

Notes: 1. Altitude Pressure Calculated by Methods of NACA Report No. 218.2. Heat Rate and Thermal Efficiency Unaffected by Altitude.

Figure 75




WORK AID 23: GENERAL ELECTRIC MODEL M5382(C) GAS TURBINE --EFFECT OF COMPRESSOR INLET TEMPERATURE ONMAXIMUM OUTPUT, HEAT RATE, AND AIR FLOW

With Permission from General Electric Company V. Poua, Rev A Feb. 26, 1987

Notes: 1. Compressor Speed - 5100 RPM; 100% Speed2. Load Turbine Design Speed - 4670 RPM

Figure 76




WORK AID 24: GENERAL ELECTRIC MODEL M5382(C) *38,000 HP GASTURBINE PERFORMANCE - INLET 120°°F

aa

106 %

110 %

115 %

120 %

125 %130 %

Percent Design Heat Rate

989 Deg F950 Deg F900 Deg F

850 Deg F

100

90

80

70

60

50

40

3040 50 60 70 80 90 100 110

800 Deg F

Load T urbine Shaft Speed - Percent

Estimated PerformanceCompressor Inlet T emperature 120 F (48.9 C)

Compressor Inlet Pressure 14.7 psia (1.0133 bar)FuelDesign OutputDesign Heat Rate (LHV)Design Heat Consumption (LHV) 10Design Air FlowDesign Shaft Speed

hpbtu/hp-hr

btu/hrlbs/hrrpm

*Natural Gas380008700330.6

978,0004670

o o

-6

Notes:1. Altitude Correction on W ork Aid 222. Ambient T emperature Correction on W ork Aid 233. Pressure Drop Effects:

4" H **htr= horsepower= heat rateO Exhaust

% Effects hp* htr**- 1 . 6 + 0 . 6- 0 . 6 + 0 . 6

PPBO62786

* hp4" H 2O Inlet2

4. For Additional 4" H 20 Pressure Drop Increase Exhaust T emp. by 2°F5. Operation at Constant High Pressure Set Speed


Figure 77





aa


Estimated PerformanceCompressor Inlet T emperature 90 F (32.2 C)

Compressor Inlet Pressure 14.7 psia (1.0133 bar)FuelDesign OutputDesign Heat Rate (LHV)Design Heat Consumption (LHV) 10Design Air FlowDesign Shaft Speed

hpbtu/hp-hr

btu/hrlbs/hrrpm

*Natural Gas380008700330.6

978,0004670


103%

110%

115%

120%

125%130%

135%Percent Design Heat Rate

110

100

90

80

70

60

50

40

3040 50 60 70 80 90 100 110

975 Deg F

900 Deg F

850 Deg F

800 Deg F

750 Deg F

o

-6

o

105%

% Effects hp* htr**- 1 . 6 + 0 . 6- 0 . 6 + 0 . 6

PPBO62786

* hp = horsepower= heat rate

4" H 2 O Inlet

2O Exhaust 4" H **htr


With Permission from General Electric Company V. Poua, Rev A Dec. 2, 1986




Figure 78





aa


Estimated PerformanceCompressor Inlet T emperature 30 F (-1.1 C)

Compressor Inlet Pressure 14.7 psia (1.0133 bar)

FuelDesign OutputDesign Heat Rate (LHV)Design Heat Consumption (LHV) 10Design Air FlowDesign Shaft Speed

hpbtu/hp-hr

btu/hrlbs/hrrpm

*Natural Gas380008700330.6

978,0004670


110%

115%

120%

130%

Percent Design Heat Rate

110

100

90

80

70

60

50

4040 50 60 70 80 90 100 110

900 Deg F

800 Deg F

750 Deg F

o

-6

120

945 Deg F

850 Deg F

98%

125%

105%

o

100%

4" H **htr= horsepower= heat rateO Exhaust

% Effects hp* htr**- 1 . 6 + 0 . 6- 0 . 6 + 0 . 6

PPBO62786

* hp4" H 2O Inlet2



Figure 79




actual steam rate (ASR) See Water Rate.

actual volume The volume of a given mass of liquid at actualtemperature in a process.

additive A component of a product mixture that is a very smallpercentage of the total.

adiabatic compression A compression process in which no heat is added orremoved.

antifoam agent A chemical that is added in very small quantities to aprocess to reduce the formation of foam.

backpressure turbine A steam turbine that does not exhaust into a condenser.The exhaust pressure will typically be 15 psig orhigher.

balance piston A device installed on the shaft of a centrifugalcompressor. It balances the thrust forces of theimpellers.

bearings The parts that support the rotating shaft.

best efficiency point (BEP) The point on the performance curve of a centrifugalcompressor where the efficiency is at a maximum.

blade A component of a steam turbine that converts steamenergy to mechanical energy. Blades are mounted onrotating wheels.

brake horsepower The quantity of power required to turn the shaft of apump or compressor. The power loading on the shaftbetween the pump and its driver.

buckets Another name for blades. Usually, the blades of aCurtis stage.

casing The outer housing of a centrifugal pump or compressor.The pressure-containing component.

cavitation The implosion of vapor bubbles in a liquid inside apump on the impeller vanes. Potentially damaging.




circuit A section of plant containing a pump, piping, and heatexchangers. A flow path between two points.




clearance A volume in a cylinder that is not swept by the piston.

clearance pocket A chamber attached to a cylinder that may be open tothe cylinder or closed off. It is used to control capacityof a reciprocating compressor.

cocoon An acoustical enclosure surrounding a gas turbine. It isused to reduce noise emission.

combined cycle A cycle that includes a gas turbine to generate power, awaste heat boiler to recover heat from the gas turbineexhaust, and a steam turbine that consumes steam fromthe waste heat boiler and generates power.

combustion The component of a gas turbine between the aircompressor and the power expander. It is the placewhere fuel is burned in the compressed air.

compressibility factor, Z Correction factor for ideal gas law.

Z =PV

nRT

compressor The first component of a gas turbine, which compressesambient air.

connecting rod The rod that connects the camshaft to the crosshead. Itchanges circular motion to reciprocating motion.

corrosion inhibitor A chemical that is added in small quantities to aprocess to prevent corrosion of the internal surfaces ofthe equipment.

crankshaft The rotating element that transmits power from thedriver to the connecting rods.

crosshead The mechanical element between the connecting rodand the piston rod. It absorbs the nonaxial forces fromthe connecting rod and transmits only axial forces tothe piston rod.

curtis stage A type of steam turbine stage with one row of nozzlesand one or more rows of buckets. The usual sequenceof components is: nozzles, rotating buckets, stationaryturning buckets, rotating buckets.




cycle efficiency In a steam turbine cycle, the sum of power output plususeful heat divided by fuel input.

cylinder The principal component of a reciprocatingcompressor. It contains the piston, suction anddischarge valves, and packing around the piston rod.

delta pressure (∆∆P) The pressure difference from pump suction to pumpdischarge.

design point The specified condition of volume and head forselection of a pump. Also called "rated point."

diaphragm A removable section inside of a casing. It contains thediffuser and a return passage, which directs the gas tothe suction of the next impeller.

diffuser A component of a centrifugal pump or compressorlocated after an impeller. The diffuser convertsvelocity head to pressure head and directs the flow tothe next impeller.

discharge Pump outlet.

displacement (pump) The volume of liquid that is theoretically pumped by apositive displacement pump without any back leakageor slip.

displacement (compressor) The volume of the space swept by the piston(s). Thetheoretically maximum capacity of a reciprocatingcompressor.

double acting Pumping on both the forward and back strokes(Reciprocating pump).

driver A motor or turbine which provides the power for thepump.

dual-shaft gas turbine A gas turbine having two shafts. This permits the aircompressor and the load turbine to run at differentspeeds. It also reduces the load on the starting device.

duplex Containing two cylinders.




efficiency The hydraulic (pressure) energy added to the liquid,divided by the power input to the shaft.




enthalpy It is also called Heat Content.

entropy A quantity that is the measure of the amount of energyin a system not available for doing work; a smallchange in entropy, ∆S, is equal to ∆Q/T, where ∆Q is asmall increment of heat added or removed and T is theabsolute temperature.

erosion Damage to the internal parts of a compressor caused byabrasion by solid particles.

extraction The process of removing medium pressure steambetween the wheels of a steam turbine.

eye The center of the impeller where liquid enters theimpeller.

flushing A small flow of liquid which keeps solids away fromthe seal and also cools the seal.

fouling The deposition of solid material on the internalpassages of a compressor.

gas horsepower The total energy imparted to gas in a compressor. Itincludes the losses due to gas friction, but does notinclude mechanical friction losses.

gear pump A positive displacement pump containing twointermeshing gears inside a casing.

governor A device that regulates the speed of a steam turbine. Itmay be mechanical or electronic.

governor A device that regulates the speed of a gas turbine.

governor valve The primary valve controlling the steam flow to aturbine.

gpm Flow rate in gallons per minute.

hand valve A valve used to shut off the steam supply to a portionof the inlet nozzles.

head The energy added to a liquid by a pump, ft-lb force/lb.mass. Also referred to as simply "feet."




heat rate A measure of fuel consumption in a gas turbine. It isthe fuel fired divided by the power output, in Btu/hp-hr.

impeller The rotating element of a centrifugal compressor thatdevelops velocity head. Also called a Wheel.

intercooler A gas cooler located between two casings of acompressor.

intercooler A gas cooler located between compressor stages.

interstage pressure On a two-stage or multistage compressor, the pressureexisting between stages.

isentropic compression Ideal compression along a path of constant entropyunder adiabatic conditions.

isentropic efficiency For a compression process, the ideal work requireddivided by the actual work imparted to the gas.

isentropic efficiency The ideal work for a compression service divided bythe actual work applied to the gas. The extra, nonidealwork is converted to heat that raises the gastemperature or is removed by jacket cooling.

isentropic head The energy per unit weight of gas applied during anideal compression process.

isothermal compression Compression at constant temperature. Heat must beremoved during the compression process.

jacket A chamber that surrounds the cylinder. Cooling watercirculates through the jacket.

journal bearing A bearing that supports the weight of the shaft of acentrifugal compressor.

kickback A recycle stream that increases the flow rate through apump, independent of process requirements.

labyrinth seal A seal made of several rings in series that fit veryclosely to a shaft and impeller eye. A labyrinth sealminimizes leakage but cannot stop it completely.




mechanical efficiency The hydraulic power divided by the total powersupplied to the pump shaft. Total power = hydraulicpower + mechanical friction losses.

mechanical efficiency The work applied to the gas divided by the driver brakehorsepower. The difference, or loss, is due tomechanical friction.

metering pump A kind of reciprocating pump that delivers small,measured amounts of liquid to a product or process.

mollier diagram A diagram that shows the relationship betweenenthalpy, entropy, temperature, and pressure for aparticular gas.

net positive suction Actual pressure at the pump suction minus vaporhead available (NPSH)A pressure of the liquid. The amount of pressure drop

that can occur before vaporization begins.

nozzle The component of a steam turbine that convertspressure energy to velocity energy.

oil seal A seal at the end of a shaft. It is lubricated with oilwhich positively prevents leakage of gas from thecasing of a compressor.

packing A seal around the plunger or rod of a reciprocatingpump to prevent leakage.

packing A flexible material that seals the space between themoving piston rod and the cylinder.

performance curve Graphs that show head produced, power required,NPSH required, and efficiency; all as functions of flowrate.

performance curve A curve supplied by the manufacturer which shows therelationship between capacity, head, horsepower, andefficiency.

piston The element that moves pack and forth in the cylinderof some reciprocating pumps.

piston The component that moves back and forth in thecylinder. It compresses the gas.




piston pump A type of reciprocating pump that contains a piston anda piston rod with a diameter smaller than the piston.

piston ring A ring surrounding the piston providing a close fit withthe cylinder. It minimizes gas leakage past the piston.

piston rod The rod that pushes and pulls the piston.

piston rod A rod that transmits force to move the piston.

plunger pump A high-pressure, single acting, reciprocating pump. Aplunger, which is the same diameter as the cylinder,takes the place of piston and piston rod.

polytropic compression The type of compression which takes place in a realcentrifugal compressor as compared to ideal isentropiccompression.

polytropic efficiency For a compression process, the minimum work along apolytropic path divided by the actual work imparted tothe gas.

polytropic head The energy per unit weight of gas applied duringpolytropic compression.

positive displacement pump A pump that operates on the principle of a constantvolume of liquid being delivered.

power turbine An expansion turbine that converts the energy of a hotcompressed gas to shaft power. Same as expander.

pulsation bottle or dampener A large chamber made of piping componentsimmediately upstream or downstream of a cylinder. Itsmoothes out the pulsations caused by the piston.

rateau stage A steam turbine stage with one row of nozzles and onerow of blades. A relatively small pressure drop is takenin the rotating blade of a Rateau stage.

reciprocating Exhibiting back and forth motion.

recycle A return flow of some liquid from the discharge side tothe suction side. Also called "kickback."




regenerative cycle A gas turbine cycle that includes a heat exchanger. Theheat exchanger transfers heat from the exhaust gas tothe compressed air before the combustion.

rotary Exhibiting rotating motion.

safety valve A valve that protects a pipe or vessel fromoverpressure. It opens automatically at a set pressure.

seal A device that prevents leakage at the point where therotating shaft enters the casing.

sequence controller An instrument that controls the startup or shutdownsequence of a gas turbine.

screw pump A positive displacement pump containing twointermeshing screws inside a casing.

shutoff The condition when a pump is rotating but flow isblocked at the discharge. (i.e., pump is acting as amixer.)

shutoff head The head produced by a pump when the discharge isblocked and flow is zero. Usually maximum headproduced.

sidestream A stream of gas that is introduced into a casing afterone or more wheels. It can also be removed from acasing before the final discharge nozzle.

simplex Containing one cylinder.

single acting Pumping on the forward stroke only. Typical ofplunger pumps.

single-shaft gas turbine A gas turbine in which the air compressor, the powerturbine, and the load are all connected to the same shaftand therefore run at the same speed.

specific gravity The density of a liquid divided by the density of waterat 60°F.

stage A section of a pump containing one impeller and onediffuser. Pumps may have one or more stages.




stage A section of a steam turbine containing one set ofnozzles and one or more row of blades.

starting device Device that provides power to accelerate thecompressor to sufficient speed so that the air flowthrough the compressor is sufficient to sustaincombustion (commonly called “lite-off”).

steam chest A chamber upstream of the first stage nozzles of asteam turbine. The area of highest steam temperatureand pressure in a turbine.

stonewall The maximum flow condition for a centrifugalcompressor. Stonewall is reached when the velocitybecomes sonic at some point in the compression path.

suction Pump inlet.

suction specific speed suction specific speed is a property of the pump that isprimarily dependent on the impeller inlet and suctioninlet design.

surge An unstable operating condition in a centrifugalcompressor. It is caused by process conditions thatresult in flow rate being too low or the requireddischarge pressure being too high.

tail rod An extra piston rod in some compressors on the sideopposite the main piston rod. It helps to stabilize pistonmotion and reduces peak stress on the piston rod.

theoretical steam rate (TSR) The flow rate of steam in pounds per hour required toproduce 1 horsepower in an ideal turbine. TSR isdetermined by steam inlet temperature and pressure andoutlet pressure.

thermal efficiency For a gas turbine cycle, the sum of power output plususeful heat output divided by the fuel consumed.

throttling Restricting the flow of a fluid, usually by means of acontrol valve in the suction or discharge processsystem.




thrust An axial force on the shaft of a compressor. It iscaused by unequal pressures on the sides of theimpellers.




thrust bearing A bearing located on the shaft of a centrifugalcompressor that absorbs the axial force on the shaft.

triplex Containing three cylinders.

turbine efficiency The theoretical steam rate divided by the actual steamrate. Also, the actual work output divided by thetheoretical work output for a given pressure range.

valve lifter A device that holds a valve open, normally used onsuction valves. When a valve is held open, gas doesnot flow through the cylinder. Used to reduce motortorque during starting. Also used for compressorcapacity control. However, open period is usuallyautomatically controlled to avoid overheating and/orlubricant accumulation in the cylinder.

volumetric efficiency The volume of liquid actually pumped, divided by thepump's theoretical displacement.

volumetric efficiency The actual volume of suction gas compressed, dividedby the theoretical displacement.

volute The annular area between the impeller and casing. Theplace where liquid velocity energy is converted topressure.

water rate The actual steam rate required per unit of power.(Pounds per horsepower-hour.)

wheel Another name for impeller.












































































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Engineering Encyclopedia