GasTurbine Engineering HandbookSecond Edition phần 4 doc

From Figure 6-14, the prewhirl distributionshould be made not only from the relative Mach number at the inducer inletshroud radius, but also from Euler work distribution at the impeller

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//INTEGRA/B&H/GTE/FINAL (26-10-01)/CHAPTER 6.3D ± 231 ± [219±274/56] 29.10.2001 3:59PM

close to the sonic velocity or greater than it, a shock wave takes place in theinducer section A shock wave produces shock loss and chokes the inducer.Figure 6-12 shows the effect of inlet prewhirl on compressor efficiency.There are three kinds of prewhirl:

with respect to the inducer inlet radius This prewhirl distribution

shroud radius Therefore, it is not effective in decreasing the relativeMach number in this manner

a maximum at the inducer inlet shroud radius, contributing to adecrease in the inlet relative Mach number

Figure 6-13 Prewhirl distribution patterns

Figure 6-14 Euler work distribution at an impeller exit

Centrifugal Compressors 231

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3 Control-vortex prewhirl This type is represented by V1 AR1

with A 0, B 6 0, and the second type with B 0, A 6 0

Euler work distributions at an impeller exit, with respect to the impellerwidth, are shown in Figure 6-14 From Figure 6-14, the prewhirl distributionshould be made not only from the relative Mach number at the inducer inletshroud radius, but also from Euler work distribution at the impeller exit.Uniform impeller exit flow conditions, considering the impeller losses, areimportant factors in obtaining good compressor performance

Impeller

An impeller in a centrifugal compressor imparts energy to a fluid Theimpeller consists of two basic components: (1) an inducer like an axial-flowrotor, and (2) the radial blades where energy is imparted by centrifugal force.Flow enters the impeller in the axial direction and leaves in the radial direc-tion The velocity variations from hub to shroud resulting from these changes

in flow directions complicate the design procedure for centrifugal sors C.H Wu has presented the three-dimensional theory in an impeller, but

compres-it is difficult to solve for the flow in an impeller using the previous theorywithout certain simplified conditions Others have dealt with it as a quasi-three-dimensional solution It is composed of two solutions, one in themeridional surface (hub-to-shroud), and the other in the stream surface ofrevolution (blade-to-blade) These surfaces are illustrated in Figure 6-15

By the application of the previous method using a numerical solution tothe complex flow equations, it is possible to achieve impeller efficiencies ofmore than 90% The actual flow phenomenon in an impeller is more com-plicated than the one calculated One example of this complicated flow isshown in Figure 6-16 The stream lines observed in Figure 6-16 do not cross,but are actually in different planes observed near the shroud Figure 6-17shows the flow in the meridional plane with separation regions at the inducersection and at the exit

Experimental studies of the flow within impeller passages have shown thatthe distribution of velocities on the blade surfaces are different from thedistributions predicted theoretically It is likely that the discrepanciesbetween theoretical and experimental results are due to secondary flowsfrom pressure losses and boundary-layer separation in the blade passages.High-performance impellers should be designed, when possible, with the aid

of theoretical methods for determining the velocity distributions on theblade surfaces

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Examples of the theoretical velocity distributions in the impeller blades

of a centrifugal compressor are shown in Figure 6-18 The blades should

be designed to eliminate large decelerations or accelerations of flow in theimpeller that lead to high losses and separation of the flow Potential flowsolutions predict the flow well in regions away from the blades where

Figure 6-15 Two-dimensional surface for a flow analysis

Figure 6-16 Flow map of impeller plane

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boundary-layer effects are negligible In a centrifugal impeller the viscousshearing forces create a boundary layer with reduced kinetic energy If thekinetic energy is reduced below a certain limit, the flow in this layer becomesstagnant, then it reverses.

Figure 6-17 Flow map as seen in meridional plane

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Figure 6-18 Velocity profiles through a centrifugal compressor

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The function of an inducer is to increase the fluid's angular momentumwithout increasing its radius of rotation In an inducer section the bladesbend toward the direction of rotation as shown in Figure 6-19 The inducer

is an axial rotor and changes the flow direction from the inlet flow angle tothe axial direction It has the largest relative velocity in the impeller and, ifnot properly designed, can lead to choking conditions at its throat as shown

in Figure 6-19

There are three forms of inducer camber lines in the axial direction Theseare circular arc, parabolic arc, and elliptical arc Circular arc camber linesare used in compressors with low pressure ratios, while the elliptical arcproduces good performance at high pressure ratios where the flow hastransonic mach numbers

Figure 6-19 Inducer centrifugal compressor

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Centrifugal Section of an Impeller

The flow in this section of the impeller enters from the inducer section andleaves the impeller in the radial direction The flow in this section is not com-pletely guided by the blades, and hence the effective fluid outlet angle doesnot equal the blade outlet angle

To account for flow deviation (which is similar to the effect accounted for

by the deviation angle in axial-flow machines), the slip factor is used:

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where V2 is the tangential component of the absolute exit velocity with a

absolute exit velocity, if the impeller were to have an infinite number ofblades (no slipping back of the relative velocity at outlet)

With radial blades at the exit,

Inertia and centrifugal forces cause the fluid elements to move closer toand along the leading surface of the blade toward the exit Once out of theblade passage, where there is no positive impelling action present, these fluidelements slow down

Causes of Slip in an Impeller

The definite cause of the slip phenomenon that occurs within an impeller

is not known However, some general reasons can be used to explain why theflow is changed

Figure 6-21 Forces and flow characteristics in a centrifugal compressor

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of two adjacent blades, the Coriolis forces, the centrifugal forces, and thefluid follow the Helmholtz vorticity law The combined gradient that resultscauses a fluid movement from one wall to the other and vice versa Thismovement sets up circulation within the passage as seen in Figure 6-22.Because of this circulation, a velocity gradient results at the impeller exitwith a net change in the exit angle

within an impeller passage causes the flowing fluid to experience a smallerexit area as shown in Figure 6-23 This smaller exit is due to small flow(if any) within the boundary layer For the fluid to exit this smaller area,its velocity must increase This increase gives a higher relative exit velocity

Figure 6-22 Coriolis circulation

Figure 6-23 Boundary-layer development

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Since the meridional velocity remains constant, the increase in relativevelocity must be accompanied with a decrease in absolute velocity.

Although it is not a new approach, boundary-layer control is being usedmore than ever before It has been used with success on airfoil designs when

it has delayed separation, thus giving a larger usable angle of attack Control

of the flow over an airfoil has been accomplished in two ways: by using slotsthrough the airfoil and by injecting a stream of fast-moving air

Separation regions are also encountered in the centrifugal impeller asshown previously Applying the same concept (separation causes a loss inefficiency and power) reduces and delays their formation Diverting the slow-moving fluid away lets the separation regions be occupied by a faster stream

of fluid, which reduces boundary-layer build-up and thus decreases separation

To control the boundary layer in the centrifugal impeller, slots in theimpeller blading at the point of separation are used To realize the fullcapability of this system, these slots should be directional and converging

in a cross-sectional area from the pressure to the suction sides as seen inFigure 6-24 The fluid diverted by these slots increases in velocity andattaches itself to the suction sides of the blades This results in moving theseparation region closer to the tip of the impeller, thus reducing slip andlosses encountered by the formation of large boundary-layer regions Theslots must be located at the point of flow separation from the blades Experi-mental results indicate improvement in the pressure ratio, efficiency, andsurge characteristics of the impeller as seen in Figure 6-24

to as leakage Leakage reduces the energy transfer from impeller to fluid anddecreases the exit velocity angle

loading, and the closer the fluid follows the vanes With higher vane ings, the flow tends to group up on the pressure surfaces and introduces avelocity gradient at the exit

necessity, impeller vanes are thick When fluid exits the impeller, the vanes

no longer contain the flow, and the velocity is immediately slowed Because

it is the meridional velocity that decreases, both the relative and absolutevelocities decrease, changing the exit angle of the fluid

A backward-curved impeller blade combines all these effects The exitvelocity triangle for this impeller with the different slip phenomenon changes

is shown in Figure 6-25 This triangle shows that actual operating conditionsare far removed from the projected design condition

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Several empirical equations have been derived for the slip factor (seeFigure 6-26) These empirical equations are limited Two of the morecommon slip factors are presented here

Stodola Slip Factor

The second Helmholtz law states that the vorticity of a frictionless fluiddoes not change with time Hence, if the flow at the inlet to an impeller isirrotational, the absolute flow must remain irrotational throughout theimpeller As the impeller has an angular velocity !, the fluid must have anangular velocityÐ! relative to the impeller This fluid motion is called therelative eddy If there were no flow through the impeller, the fluid in the

Figure 6-24 Percent design flowÐlaminar flow control in a centrifugal pressor

com-Centrifugal Compressors 241

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Figure 6-25 Effect on exit velocity triangles by various parameters.

Figure 6-26 Various slip factors as a function of the flow coefficient

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impeller channels would rotate with an angular velocity equal and opposite

to the impeller's angular velocity

To approximate the flow, Stodola's theory assumes that the slip is due tothe relative eddy The relative eddy is considered as a rotation of a cylinder

of fluid at the end of the blade passage at an angular velocity ofÐ! about itsown axis The Stodola slip factor is given by

U2

264

37

where:

Z the number of blades

Calculations using this equation have been found to be lower than mental values

experi-Stanitz Slip Factor

Stanitz calculated blade-to-blade solutions for eight impellers andconcluded that for the range of conditions covered by the solutions, U is

approximately the same whether the flow is compressible or incompressible

3

with experimental results for radial or near-radial blades

Diffusers

Diffusing passages have always played a vital role in obtaining goodperformance from turbomachines Their role is to recover the maximumpossible kinetic energy leaving the impeller with a minimum loss in totalpressure The efficiency of centrifugal compressor components has beensteadily improved by advancing their performance However, significant

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further improvement in efficiency will be gained only by improving thepressure recovery characteristics of the diffusing elements of these machines,since these elements have the lowest efficiency.

The performance characteristics of a diffuser are complicated functions ofdiffuser geometry, inlet flow conditions, and exit flow conditions Figure 6-27

Figure 6-27 Geometric classification of diffusers

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shows typical diffusers classified by their geometry The selection of anoptimum channel diffuser for a particular task is difficult, since it must bechosen from an almost infinite number of cross-sectional shapes and wallconfigurations In radial and mixed-flow compressors the requirement ofhigh performance and compactness leads to the use of vaned diffusers asshown in Figure 6-28 Figure 6-28 also shows the flow regime of a vane-island diffuser

Matching the flow between the impeller and the diffuser is complexbecause the flow path changes from a rotating system into a stationaryone This complex, unsteady flow is strongly affected by the jet-wake ofthe flow leaving the impeller, as seen in Figure 6-29 The three-dimensionalboundary layers, the secondary flows in the vaneless region, and the flowseparation at the blades also affects the overall flow in the diffuser

The flow in the diffuser is usually assumed to be of a steady nature toobtain the overall geometric configuration of the diffuser In a channel-typediffuser the viscous shearing forces create a boundary layer with reducedkinetic energy If the kinetic energy is reduced below a certain limit, the flow

in this layer becomes stagnant and then reverses This flow reversal causes

Figure 6-28 Flow regions of the vaned diffuser

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separation in a diffuser passage, which results in eddy losses, mixing losses,and changed-flow angles Separation should be avoided or delayed toimprove compressor performance.

The high-pressure-ratio centrifugal compressor has a narrow yet stableoperating range This operating range is due to the close proximity of thesurge and choke flow limits The word ``surge'' is widely used to expressunstable operation of a compressor Surge is the flow breakdown periodduring unstable operation The unsteady flow phenomena during the onset

of surge in a high-pressure-ratio centrifugal compressor causes the mass flowthroughout the compressor to oscillate during supposedly ``stable'' operations.The throat pressure in the diffuser increases during the precursor period

(except plenum pressure) suddenly drop at the surge point The suddenchange of pressure can be explained by the measured occurrence of backflowfrom the collector through the impeller during the period between the twosudden changes

Figure 6-29 Jet-wake flow distribution from an impeller

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Scroll or Volute

The purpose of the volute is to collect the fluid leaving the impeller ordiffuser, and deliver it to the compressor outlet pipe The volute has animportant effect on the overall efficiency of the compressor Volute designembraces two schools of thought First, the angular momentum of the flow

in the volute is constant, neglecting any friction effects The tangential

equa-tion shows the relaequa-tionship if the angular momentum is held constant

Assuming no leakage past the tongue and a constant pressure around the

flow in the impeller Q is given by

of the double-vortex in the symmetrical volute Where the impeller is charging directly into the volute, it is better to have the volute width largerthan the impeller width This enlargement results in the flow from the

dis-Centrifugal Compressors 247

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impeller being bounded by the vortex generated from the gap between theimpeller and the casing.

At flows different from design conditions, there exists a circumferentialpressure gradient at the impeller tip and in the volute at a given radius

Figure 6-30 Flow patterns in volute

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At low flows, the pressure rises with the peripheral distance from the volutetongue At high flows, the pressure falls with distance from the tongue Thiscondition results because near the tongue the flow is guided by the outer wall

of the passage The circumferential pressure gradients reduce efficiency awayfrom the design point Nonuniform pressure at the impeller discharge results

in unsteady flows in the impeller passage, causing flow reversal and ation in the impeller

separ-CentrifugalCompressor PerformanceCalculating the performance of a centrifugal compressor in both design andoff-design conditions requires a knowledge of various losses encountered

in a centrifugal compressor

The accurate calculation and proper evaluation of losses within a fugal compressor is as important as the calculation of the blade-loading para-meters If the proper parameters are not controlled, efficiency decreases Theevaluation of various losses is a combination of experimental results and theory.The losses are divided into two groups: (1) losses encountered in the rotor, and(2) losses encountered in the stator

centri-A loss is usually expressed as a loss of heat or enthalpy centri-A convenient way

to express them is in a nondimensional manner with reference to the exit

available from the energy equation

2

The adiabatic head that is actually available at the rotor discharge is equal

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Therefore, the adiabatic efficiency in the impeller is

imp qqia

The calculation of the overall stage efficiency must also include lossesencountered in the diffuser Thus, the overall actual adiabatic head attainedwill be the actual adiabatic head of the impeller minus the head losses

vaneless diffuser space

Rotor losses are divided into the following categories:

inlet The inlet of the rotor blades should be wedgelike to sustain a weakoblique shock, and then gradually expanded to the blade thickness to avoidanother shock If the blades are blunt, a bow shock will result, causing theflow to detach from the blade wall and the loss to be higher

incidence angle that is either positive or negative, as shown in Figure 6-31

A positive incidence angle causes a reduction in flow Fluid approaching

a blade at an incidence angle suffers an instantaneous change of velocity atthe blade inlet to comply with the blade inlet angle Separation of the bladecan create a loss associated with this phenomenon

surface of the rotor as seen in Figure 6-32 This loss is the same for a given

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size disc whether it is used for a radial-inflow compressor or a radial-inflowturbine Losses in the seals, bearings, and gear box are also lumped in withthis loss, and the entire loss can be called an external loss Unless the gap is

of the magnitude of the boundary layer, the effect of the gap size is gible The disc friction in a housing is less than that on a free disc due to theexistence of a ``core,'' which rotates at half the angular velocity

gradients in the boundary layer Deceleration of the flow increases theboundary layer and gives rise to separation of the flow The adverse pressuregradient that a compressor normally works against increases the chances ofseparation and causes significant loss

Figure 6-31 Inlet velocity triangles at nonzero incidents

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Clearance loss When a fluid particle has a translatory motion relative

to a noninertial rotating coordinate system, it experiences the Coriolis force

A pressure difference exists between the driving and trailing faces of animpeller blade caused by Coriolis acceleration The shortest and least resis-tant path for the fluid to flow and neutralize this pressure differential isprovided by the clearance between the rotating impeller and the stationarycasing With shrouded impellers, such a leakage from the pressure side to thesuction side of an impeller blade is not possible Instead, the existence of apressure gradient in the clearance between the casing and the impellershrouds, predominant along the direction shown in Figure 6-33, accountsfor the clearance loss Tip seals at the impeller eye can reduce this lossconsiderably

This loss may be quite substantial The leaking flow undergoes a largeexpansion and contraction caused by temperature variation across the clear-ance gap that affects both the leaking flow and the stream into which itdischarges

on the impeller wall caused by turbulent friction This loss is determined byconsidering the flow as an equivalent circular cross section with a hydraulicdiameter The loss is then computed based on well-known pipe flow pressureloss equations

Figure 6-32 Secondary flow at the back of an impeller

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Stator Losses

impel-ler exit of a compressor and is a direct function of the air exit angle As theflow through the compressor decreases, there is an increase in the absoluteflow angle at the exit of the impeller as seen in Figure 6-34 Part of the fluid isrecirculated from the diffuser to the impeller, and its energy is returned tothe impeller

a wake in the vaneless space behind the rotor It is minimized in a diffuser,which is symmetric around the axis of rotation

Figure 6-33 Leakage affecting clearance loss

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Vaneless diffuser loss This loss is experienced in the vaneless diffuserand results from friction and the absolute flow angle.

test results They are a function of the impeller blade loading and the vanelessspace radius ratio They also take into account the blade incidence angle andskin friction from the vanes

leav-ing the vaned diffuser is lost

Losses are complex phenomena and as discussed here are a function ofmany factors, including inlet conditions, pressure ratios, blade angles, andflow Figure 6-35 shows the losses distributed in a typical centrifugal stage ofpressure ratio below 2:1 with backward-curved blades This figure is only aguideline

Compressor Surge

A plot showing the variation of total pressure ratio across a compressor as

a function of the mass flow rate through it at various speeds is known as aperformance map Figure 6-36 shows such a plot

temperature and pressure The surge line joins the different speed lines where

Figure 6-34 Recirculating loss

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the compressor's operation becomes unstable A compressor is in ``surge''when the main flow through the compressor reverses its direction and flowsfrom the exit to the inlet for short time intervals If allowed to persist, thisunsteady process may result in irreparable damage to the machine Lines

of constant adiabatic efficiency (sometimes called efficiency islands) are alsoplotted on the compressor map A condition known as ``choke'' or ``stonewalling'' is indicated on the map, showing the maximum mass flow ratepossible through the compressor at that operating speed

Compressor surge is a phenomenon of considerable interest, yet it is notfully understood It is a form of unstable operation and should be avoided inboth design and operation Surge has been traditionally defined as the lowerlimit of stable operation in a compressor and involves the reversal of flow.This reversal of flow occurs because of some kind of aerodynamic instabilitywithin the system Usually a part of the compressor is the cause of theaerodynamic instability, although it is possible that the system arrangementcould be capable of augmenting this instability Figure 6-36 shows a typical

Figure 6-35 Losses in a centrifugal compressor

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performance map for a centrifugal compressor with efficiency islands andconstant aerodynamic speed lines The total pressure ratio can be seen tochange with flow and speed Compressors are usually operated at a workingline separated by some safety margin from the surge line.

Surge is often symptomized by excessive vibration and an audible sound;however, there have been cases in which surge problems that were not audiblehave caused failures Extensive investigations have been conducted on surge.Poor quantitative universality of aerodynamic loading capacities of differentdiffusers and impellers, and an inexact knowledge of boundary-layer behaviormake the exact prediction of flow in turbomachines at the design stagedifficult However, it is quite evident that the underlying cause of surge isaerodynamic stall The stall may occur in either the impeller or the diffuser

Figure 6-36 Typical compressor performance map

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When the impeller seems to be the cause of surge, the inducer section

is where the flow separation begins A decrease in the mass flow rate, anincrease in the rotational speed of the impeller, or both can cause the com-pressor to surge

Surge can be initiated in the diffuser by flow separation occurring atthe diffuser entrance A diffuser usually consists of a vaneless space withthe prediffuser section before the throat containing the initial portion of thevanes in a vaned diffuser The vaneless space accepts the velocity generated

by the centrifugal impeller and diffuses the flow so that it enters the vaneddiffuser passage at a lower velocity, avoiding any shock losses and resultantseparation of the flow When the vaneless diffuser stalls, the flow will notenter the throat A separation occurs, causing the flow to finally reverse andsurge the compressor Stalling of the vaneless diffuser can be accomplished

in two waysÐby increasing impeller speed or decreasing the flow rate.Whether surge is caused by a decrease in flow velocity or an increase inrotational speeds, either the inducer or vaneless diffuser can stall Whichstalls first is difficult to determine, but considerable testing has shown thatfor a low-pressure-ratio compressor, the surge initiates in the diffuser sec-tion For units with single-stage pressure ratios above 3:1, surge is probablyinitiated in the inducer

Most centrifugal compressors have for the most part impellers with ward leaning impeller blades Figure 6-37 depicts the effects of impellerblade angle on the stable range and shows the variance in steepness of theslope of the head-flow curve

back-The three curves are based on the same speed and show actual head back-Therelationship of ideal or theoretical head to inlet flow for different bladeangles would be represented by straight lines For backward leaning blades,the slope of the line would be negative The line for radial blades would

be horizontal Forward leaning blades would have a positively sloped line.For the average petrochemical process plant application, the compressor

it provides a wider stable range and a steeper slope in the operating range.This impeller design has proven to be about the best compromise betweenpressure delivered, efficiency, and stability Forward leaning blades are notcommonly used in compressor design, since the high exit velocities lead tolarge diffuser losses A plant air compressor operating at steady conditionsfrom day to day would not require a wide stable range, but a machine in aprocessing plant can be the victim of many variables and upsets So morestability is highly desirable Actually, the lower curve in Figure 6-37 appears

to have a more gentle slope than either the middle or upper curve This

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comparison is true in the overall sense, but it must be remembered that thenormal operating range lies between 100% flow (Q) and flow at surge, plus

a safety margin of, usually, about 10% The right-hand tail ends of all threecurves are not in the operating range The machine must operate with asuitable margin to the left of where these curves begin their steep decent ortail-off, and in the resultant operating range, the curve for backward leaningblades is steeper This steeper curve is desirable for control purposes Such

a curve produces a meaningful change in pressure drop across the orifice for

a small change in flow The blade angle by itself does not tell the overallperformance story The geometry of other components of a stage will con-tribute significant effects also

Most centrifugal compressors in service in petroleum or petrochemicalprocessing plants use vaneless diffusers A vaneless diffuser is generally asimple flow channel with parallel walls and does not have any elementsinside to guide the flow

Forward Leaning Blade β 2 > 90°

Radial Blade

β 2 = 90°

Backward Leaning Blade β 2 < 90°

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When the inlet flow to the impeller is reduced while the speed is heldconstant, there is a decrease in the relative velocity leaving the impeller andthe air angle associated with it As the air angle decreases, the length of theflow path spiral increases The effect is shown in Figure 6-38

If the flow path is extended enough, the flow momentum at the diffuserwalls is excessively dissipated by friction and stall With this greater loss, thediffuser becomes less efficient and converts a proportionately smaller part

of the velocity head to pressure As this condition progresses, the stage willeventually stall This could lead to a surge

Vaned diffusers are used to force the flow to take a shorter, more efficientpath through the diffuser There are many styles of vaned diffusers, withmajor differences in the types of vanes, vane angles and contouring, andvane spacing Commonly used vaned diffusers employ wedge-shaped vanes(vane islands) or thin-curved vanes In high head stages, there can be two tofour stages of diffusion These usually consist of vaneless spaces to deceleratethe flow, followed by two or three levels of vaned blades in order to preventbuild-up of boundary layer, which causes separation and surging of thecompressor Figure 6-38 indicates the flow pattern in a vaned diffuser Thevaned diffuser can increase the efficiency of a stage by two to four percen-tage points, but the price for the efficiency gain is generally a narroweroperating span on the head-flow curve with respect to both surge andstonewall Figure 6-39 also shows the effect of off-design flows

Excessive positive incidence at the leading edge of the diffuser vane occurswhen the exit flow is too small at reduced flow, and this condition brings on astall Conversely, as flow increases beyond the rated point, excessive negative

Impeller Eye Blade Diffuser O.D.

Paths of Particle in Diffuser

Normal Condition Good Flow Angle Relatively Short Flow Path.

Minimum Friction Loss

Near Surge Shallow Flow Angle.

Long Flow Path High Frictional Loss Possibility of Flow Re-entering the Impeller.

Impeller O D

Figure 6-38 Flow trajectory in a vaneless diffuser

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incidence can cause stonewall Despite its narrowing effect on the usableoperating range on the characteristic curve, the vaned diffuser has its applica-tion in situations where efficiency is of utmost importance Although seldomused, movable diffuser vanes or vane islands can be used to alleviate the shocklosses at off-design conditions However, as the adjusting mechanismsrequired are quite complicated, they generally are applied only to single-stagemachines.

It should be noted that the illustrations of the flow paths in Figures 6-37through 6-39 are somewhat simplistic Each flow path is indicated by a singlestreamline The actual flow field is far more complex, with flow separationand recalculation present Nevertheless, these figures should help with apractical understanding of the effects of changes in velocity triangles.Stationary guide vanes direct the flow to the eye of the impeller in anorderly fashion Depending upon the head requirements of an individualstage, these vanes may direct the flow in the same direction as the rotation

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or tip speed of the wheel, an action known as positive pre-swirl This isusually done to reduce the relative Mach number entering the inducer, inorder to prevent shock losses This, however, reduces the head delivered butimproves the operating margin The opposite action is known as counter-rotation or negative pre-swirl This increases the head delivered but alsoincreases the inlet relative mach number Negative pre-swirl is rarely used,since it also decreases the operating range Sometimes the guide vanes are set

at zero degrees of swirl; these vanes are called radial guide vanes Movableinlet guide vanes are occasionally employed on single-stage machines, or onthe first stage of multi-stage compressors driven by electric motors at con-stant speed The guide vane angle can be manually or automatically adjustedwhile the unit is on stream to accommodate off-design operating require-ments Because of the mechanical complexity of the adjusting mechanism andphysical dimensional limitations, the variable feature can only be applied tothe first wheel in almost all machine designs Hence, the effect of changingvane angle is diluted in the stages downstream of the first Although the flow

to the entire machine is successfully adjusted by moving the first stage vanes,the remaining stages must pump the adjusted flow at a fixed guide vane angle.Incidentally, a butterfly throttle valve in the suction line to the machinewill produce nearly the same effects as moving the first stage guide vanes.However, throttling is not as efficient as moving the guide vanes, so that inmany cases, the added cost of the movable vane mechanism can be justified

by power savings

Effects of Gas Composition

Figure 6-40 shows the performance of an individual stage at a given speedfor three levels of gas molecular weight

The heavy gas class includes gases such as propane, propylene, andstandardized refrigerant mixtures Air, natural gases, and nitrogen are typ-ical of the medium class Hydrogen-rich gases found in hydrocarbon proces-sing plants are representative of the light class

The following observations can be made with respect to the curve forheavy gas:

1 The flow at surge is higher

2 The stage produces slightly more head than that corresponding tomedium gas

3 The right-hand side of the curve turns downward (approaches wall) more rapidly

stone-4 The curve is flatter in the operating stage

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It is the last point (4) that often presents a problem to the designer of theantisurge control system It should be noted that the flatness gets worse asstages are added in series Since the RTS is small, there is a large change inflow corresponding to a small change in Head The control system, there-fore, must be more responsive It should be obvious that curves for lightergases have a more desirable shape.

ExternalCauses and Effects of Surge

The following are some of the usual causes of surge that are not related tomachine design:

1 Restriction in suction or discharge of a system

2 Process changes in pressure, temperature, or gas composition

3 Internal plugging of flow passages of compressor (fouling)

4 Inadvertent loss of speed

5 Instrument or control valve malfunction

6 Malfunction of hardware such as variable inlet guide vanes

SURGE POINTS

Figure 6-40 Effect of gas composition

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The effects of the size and configuration of the connected system, as well

as different operating conditions, on the intensity of surge can be ing For example, a compressor system in a test set-up at the factory mayexhibit only a mild reaction to surge At the installation, however, the samecompressor with a different connected system may react in a tumultuousmanner Surge can often be recognized by check valve hammering, pipingvibration, noise, wriggling of pressure gauges or an ammeter on the driver,

astonish-or lateral and/astonish-or axial vibration of the compressastonish-or shaft Mild cases of surgesometimes are difficult to discern

Surge Detection and Control

Surge-detection devices may be divided into two groups: (1) staticdevices, and (2) dynamic devices To date, static surge-detection deviceshave been widely used; more research work needs to be done beforedynamic detection devices are generally used A dynamic device will prob-ably meet the requirements and hopes of many engineers for a controldevice that can anticipate stall and surge, and prevent it Obviously, detec-tion devices must be linked to a control device that can prevent the unstableoperation of a compressor

Static surge-detection devices attempt to avoid stall and surge by themeasurement of compressor conditions and ensure that a predecided value

is not exceeded When conditions meet or exceed the limit, control action

is taken A typical pressure-oriented anti-surge control system is shown inFigure 6-41 The pressure transmitter monitors the pressure and controls adevice, which opens a blowoff valve A temperature-sensing device correctsthe readings of flow and speed for the effect of temperature A typical flow-oriented device is also shown in Figure 6-42

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In all static surge-detection devices, the actual phenonemon of flow sal (surge) is not directly monitored What is monitored are other conditionsrelated to surge Control limits are set from past experience and a study ofcompressor characteristics.

rever-Dynamic surge-detection and control methods are under study Theyattempt to detect the start of a reversal of flow before it reaches the criticalsituation of surge This procedure uses a boundary-layer probe

The author has a patent for a dynamic surge-detection system, using aboundary-layer probe, presently undergoing field tests This system consists

of specially mounted probes in the compressor to detect boundary-layer flow

Figure 6-41 Pressure-oriented anti-surge control system

Figure 6-42 Flow-oriented anti-surge control system

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reversal, as shown in Figure 6-43 The concept assumes that the boundarylayer will reverse before the entire unit is in surge Since the system ismeasuring the actual onset of surge by monitoring the flow reversal, it isnot dependent on the molecular weight of the gas and is not affected by themovement of the surge line

The use of pressure transducers and casing accelerometers in the exitpiping has been instrumental in detecting compressor surge It has beenfound that as the unit approaches surge, the blade passing frequency (num-ber of blades times rpm) and its second and third harmonic become excited

In a limited number of tests it has been noted that when the second harmonic

Figure 6-43 Boundary-layer surge prediction technique

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of the blade passing frequency reaches the same order of magnitude as theblade passing frequency, the unit is very close to surge.

Process CentrifugalCompressors

and thus large surge-to-choke margins Figure 6-44 shows a cross section of

a typical multistage centrifugal compressor used in the process industries.The common method of classifying process-type centrifugal compressorsdriven by gas turbines is based on the number of impellers and the casingdesign Table 6-2 shows three types of centrifugal compressors For eachtype of compressor, approximate maximum ratings of pressure, capacity,and brake horsepower are also shown Sectionalized casing types haveimpellers, which are usually mounted on the extended motor shaft, andsimilar sections are bolted together to obtain the desired number of stages.Casing material is either steel or cast iron These machines require minimumsupervision and maintenance, and are quite economic in their operatingrange The sectionalized casing design is used extensively in supplying airfor combustion in ovens and furnaces

The horizontally split types have casings split horizontally at the section and the top The bottom halves are bolted and doweled together asshown in Figure 6-45 This design type is preferred for large multistage units

mid-Figure 6-44 Cross section of a typical multistage centrifugal compressor.(Courtesy Elliott Company, Jeannette, PA.)

FPO

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Table 6-2 IndustrialCentrifugalCompressor Classification

Based on Casing Design

Approximate Maximum Ratings

Casing Type

Approximate Pressure psig (Bar)

Approximate Inlet Capacity cfm (cmm)

Approximate Power Horsepower (kW)

Figure 6-45 Horizontally split centrifugal compressor with shrouded rotors tesy of Elliott Company.)

(Cour-FPO

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The internal parts such as shaft, impellers, bearings, and seals are readilyaccessible for inspection and repairs by removing the top half The casingmaterial is cast iron or cast steel.

There are various types of barrel or centrifugal compressors pressure types with overhung impellers are used for combustion processes,ventilation, and conveying applications Multistage barrel casings are usedfor high-pressures in which the horizontally split joint is inadequate Figure6-46 shows the barrel compressor in the background and the inner bundlefrom the compressor in front Once the casing is removed from the barrel, it

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weight One must also know the corrosive properties of the gas so thatproper metallurgical selection can be made Gas fluctuations due to processinstabilities must be pinpointed so that the compressor can operate withoutsurging

Centrifugal compressors for industrial applications have relatively lowpressure ratios per stage This condition is necessary so that the compressorscan have a wide operating range while stress levels are kept at a minimum.Because of the low pressure ratios for each stage, a single machine may have

a number of stages in one ``barrel'' to achieve the desired overall pressureratio Figure 6-47 shows some of the many configurations Some factors to

be considered when selecting a configuration to meet plant needs are:

1 Intercooling between stages can considerably reduce the powerconsumed

2 Back-to-back impellers allow for a balanced rotor thrust and imize overloading the thrust bearings

min-3 Cold inlet or hot discharge at the middle of the case reduces oil-sealand lubrication problems

4 Single inlet or single discharge reduces external piping problems

5 Balance planes that are easily accessible in the field can appreciablyreduce field-balancing time

6 Balance piston with no external leakage will greatly reduce wear onthe thrust bearings

7 Hot and cold sections of the case that are adjacent to each other willreduce thermal gradients, and thus reduce case distortion

8 Horizontally split casings are easier to open for inspection thanvertically split ones, reducing maintenance time

9 Overhung rotors present an easier alignment problem becauseshaft-end alignment is necessary only at the coupling between thecompressor and driver

10 Smaller, high-pressure compressors that do the same job will reducefoundation problems but will have greatly reduced operationalrange

Impeller Fabrication

Centrifugal-compressor impellers are either shrouded or unshrouded.Open, shrouded impellers that are mainly used in single-stage applicationsare made by investment casting techniques or by three-dimensional milling.Such impellers are used, in most cases, for the high-pressure-ratio stages.The shrouded impeller is commonly used in the process compressor because

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Figure 6-47 Various configurations of centrifugal compressors.

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