surge prevention cent r if compr sys

7/27/2019 Surge Prevention Cent r if Compr Sys

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SURGE PREVENTION IN CENTRIFUGAL COMPRESSOR SYSTEMS

Rainer Kurz and Robert C. WhiteSolar Turbines Incorporated

San Diego, California

Abstract

Centrifugal compressor surge and its prevention have drawn significant attention in the literature.An important aspect of surge avoidance lies in the design of the compressor station and, in par-ticular, the piping upstream and downstream of the compressor. Most anti-surge systems are per-fectly capable of avoiding surge during normal operating conditions. However, unplanned emer-gency shutdowns present a significant challenge, and surge avoidance in these cases depends to alarge degree on the station layout, in particular the volume of the piping system downstream ofthe compressor. Furthermore, the concepts used in the anti surge system (valves, piping, coolers)also impact the start-up of the station, or of individual units of the station. Start-up considerations

for stations with and without cooled recycle loops are discussed.

Nomenclature

A Flow area

cv Flow coefficient (cv=Q/SG/p)C Compressible valve coefficientFp Piping geometry factorh HeadHcooler Gas cooler heat transferJ Inertiak Isentropic exponentk ConstantKv Valve coefficientL Pipe lengthN Speed (1/s)p PressureQ Volumetric flowSG Specific gravitySM Surge margin (%)T Temperature

t TimeV VolumeY Expansion FactorZ Compressibility factor

,, Constants Density


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Subscriptsavail Availablecompr Compressorop Operating pointsurge At surgestd At standard conditionsss Steady statev Valve1 Compressor inlet2 Compressor discharge

Introduction

Recently, there have been attempts to increase the stability margin of a compressor by active(Epstein et. al. [1], Blanchini et. al. [2]) or passive means (Arnulfi et. al. [3]), or by increasing theaccuracy of determining the surge margin (McKee and Deffenbaugh [4]). It is often overlookedthat meaningful gains can be made by better understanding the interaction between the compres-

sor, the anti-surge devices (control system, valves) and the station piping layout (coolers, scrub-bers, check valves). This study focuses on centrifugal compressors driven by two-shaft gas tur-bines (Figure 1).

The possible operating points of a centrifugal gas compressor are limited by maximum andminimum operating speed, maximum available power, choke flow, and stability (surge) limit(Figure 2). Surge, which is the flow reversal within the compressor, accompanied by high fluctu-ating load on the compressor bearings, has to be avoided to protect the compressor. The usualmethod for surge avoidance (anti-surge-control) consists of a recycle loop that can be activatedby a fast acting valve (anti-surge valve) when the control system detects that the compressorapproaches its surge limit. Typical control systems use suction and discharge pressures and tem-peratures, together with the flow through the compressor to calculate the relative distance(surge margin) of the present operating point to the predicted or measured surge line of thecompressor (Figure 2). The surge margin is defined by:

constN

op

surgeop

Q

QQSM =

= (1)

If the surge margin reaches a preset value (often 10%), the anti-surge valve starts to open,thereby reducing the pressure ratio of the compressor and increasing the flow through the com-pressor. The situation is complicated by the fact that the surge valve also has to be capable ofprecisely controlling flow. Additionally, some manufacturers place limits on how far into choke(or overload) they allow their compressors to operate.


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Figure 1: Compressor station

Figure 2: Typical compressor map

A very critical situation arises upon emergency shutdown (ESD). Here, the fuel supply to the gas

turbine driver is cut off instantly, thus eliminating the power to the driven compressor1

. The in-ertia of compressor, coupling and power turbine have to balance the compressor absorbed power,causing a rapid deceleration. Because the head-making capability of the compressor is reducedby the square of its running speed, while the pressure ratio across the machine is imposed by theupstream and downstream piping system, the compressor will surge if the surge valve cannotprovide fast relief of the pressure. The deceleration of the compressor as a result of inertia anddissipation are decisive factors. The speed at which the pressure can be relieved not only de-

1 Some installations maintain fuel flow to the turbine for 1 to 2 seconds while the recycle valve opens. However, this can generate a safety hazard.


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pends on the reaction time of the valve, but also on the time constants imposed by the piping sys-tem. The transient behavior of the piping system depends largely on the volumes of gas enclosedby the various components of the piping system, which may include, besides the piping itself,various scrubbers, knockout drums, and coolers. The system boundaries for this study are thefirst downstream check valve, while the upstream boundary may be eithera check valve or an infinite plenum (at constant pressure, Figure 3).

The requirements of the anti-surge system for such situations as ESD or other massive systemdisturbances are distinctly different from the usual process control case.

TT FT PT

COMPRESSORENGINE

PT TT

ANTI-SURGE

CONTROLLER

SOLENOID

ENABLE

24VDC

4 - 20mA

LIMIT

SWITCH

POSITION

TRANSMITTER

4 - 20mA

ANTI-SURGE

CONTROL VALVE

FAIL OPEN

SV = SUCTION VALVE

LV = LOADING VALVE

VV = VENT VALVE

DV = DISCHARGE VALVE

TT = TEMPERATURE

TRANSMITTER

FT = FLOW TRANSMITTER

PT = PRESSURE

TRANSMITTER

AFTERCOOLERSV

LV

VV

DV

SCRUBBER

Figure 3: Anti-surge and recycle system.

The former is a massive, fast change in conditions that first and foremost requires fast systemreaction. This requires, among others, extremely fast opening valves with sufficiently large flowareas. The latter requires the capability to precisely control slow changes in the process, such that

no oscillations occur, which can be accomplished with precisely positioned valves.Surge control valves are primarily sized to fit the compressor. During steady-state recycling, therequired capacity of the recycle valve can be directly derived from the compressor map. To han-dle transient conditions, the required capacity must be greater to allow for the volumes on eitherside of the compressor. With the initiation of a shutdown, the compressor can be expected todecelerate approximately 30% in the first second. With a 30% loss in speed, the head the com-pressor can develop at its surge limit will drop by approximately 50%. The recycle control valve


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must, therefore, reduce the pressure ratio across the compressor by one-half in that first second toavoid surge.

The following guidelines pertain to a typical one valve, one compressor arrangement. Morecomplex systems of cascaded valves or valves around multiple compressors require a more de-tailed analysis.

To facilitate both precise throttling at partial recycle and the need to reduce the pressure differ-ence across the compressor quickly during a shutdown, surge control valves with an equal per-centage characteristic are advantageous (Figure 4). The equal percentage characteristic spreadsthe first half of the valve's fully open capacity over the first 2/3 of the valves travel for a globevalve, and about one third of the valve's fully open capacity over the first 2/3 of the valvestravel for a ball valve. This greatly improves controllability at partial recycle throttling. In orderto avoid surge during a shutdown, the valve must open to the required capacity in significantlyless than one second.

Figure 4: Typical valve characteristics for globe and ball valves.

Surge control systems must be sized to meet two diverse objectives: During steady-state recy-cling, the required capacity of the recycle valve can be directly derived from the compressormap: the smaller the valve, the smoother the control. During transient conditions, the requiredcapacity increases due to the volumes on either side of the compressor. Therefore, to avoid surge


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during a shutdown, the bigger the valve, the better. To facilitate both smooth throttling at partialrecycle andthe need to reduce the pressure differential (DP) across the compressor quickly dur-ing a shutdown, control valves with an equal percentage characteristic are recommended. Withan equal percentage characteristic, the more the valve is opened, the greater the increase in flowfor the same travel. We recommend two types of valves for surge control: globe valves andnoise-attenuating ball valves.

The globe valves capacity (Cv) varies with the square of the percentage travel. The noise-attenuating ball valves capacity (Cv) varies with the cube of the percentage travel. Both valvesare sized to be throttling at about two-thirds open at surge conditions. As such, the noise-attenuating ball valve will have 50% more capacity to depressurize the discharge volume thanthe globe valve. This additional capacity makes the noise-attenuating ball valve the better choicein installations where there is a single surge control valve (i.e. there is no hot bypass) and the dis-charge volumes are large (e.g., the discharge system includes an aftercooler). The valve ismatched to the compressor (Figure 5).

Figure 5: Matching of Valve and Compressor. The valve characteristic for a number of openingpositions (60%,70% and 100%) is superimposed to the compressor performance map.

Surge Phenomenon

Day [5] has provided very detailed measurements of compressor surge cycles, pointing out thedominating influence of the discharge volume.Figure 6 shows the head-versus-flow characteristic of a typical centrifugal compressor at con-stant speed, including the areas of unstable operation. At flows lower than the stability limit


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(surge line), the compressor initially shows a reduced capability to generate head with reducedflow, until it experiences reverse flow, that is, the gas now flows from the discharge to the suc-tion side of the compressor. Once flow reversal occurs, the amount of flow depends on the pres-sure ratio across the compressor, since in this situation the compressors acts more or less like anorifice. The flow reversal means that the pressure downstream of the compressor is graduallyreduced. The speed of pressure reduction depends largely on the size of the volume downstreamof the compressor. Once the pressure is reduced sufficiently, the compressor will recover andflow gas again from the suction to the discharge side. Unless action is taken, the events repeatagain. Ongoing surge can damage thrust bearings (due to the massive change of thrust loads),seals, and eventually overheat the compressor.

Figure 6: Simplified surge cycle

Modeling the Piping Surge Control Interaction

Design of the piping and valves, together with the selection and the placement of instrumentswill significantly affect the performance of an anti-surge control system. This should be a majorissue during the planning stage because the correction of design flaws can be very costly oncethe equipment is in operation. Typical configurations for recycle systems are outlined in Figure3. In its simplest form, the system includes a flow-measuring element in the compressor suction,

instruments to measure pressures and temperatures at suction and discharge, the compressor, anaftercooler and a discharge check-valve, as well as a recycle line with a control valve, connectedupstream of the discharge check valve and compressor flow-measuring device.

The control system monitors the compressor operating parameters, compares them to the surgelimit, and opens the recycle valve as necessary to maintain the flow through the compressor at adesired margin from surge. In the event of an ESD, where the fuel to the gas turbine is shut off


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instantly, the surge valve opens immediately, essentially at the same time the fuel valve is clos-ing.

In a simple system, the boundaries for the gas volume (V) on the discharge side are establishedby the discharge check valve, compressor, and recycle valve (Figure 3). The volume on the suc-tion side is usually orders of magnitude larger than the discharge volume and, therefore, can beconsidered infinite. Thus, for the following considerations, the suction pressure remains constant.This is not a general rule, but is used to simplify the following considerations. This yields thesimplified system, consisting of a volume filled by a compressor and emptied through a valve(Figure 7).

Figure 7: Simplified system and transient characteristic.

The basic dynamic behavior of the system in Figure 7 is that of a fixed volume where the flowthrough the valve is a function of the pressure differential over the valve. In a surge avoidancesystem, a certain amount of the valves flow capacity will be consumed to recycle the flow

through the compressor. Only the remaining capacity is available for de-pressurizing the dis-charge volume.

The worst case scenario for a surge control system is an ESD, particularly if the compressor isalready operating close to surge when the engine shutdown occurs

2. With the initiation of shut-

down, the compressor will decelerate rapidly under the influence of the fluid forces counteractedby the inertia of the rotor system. A 30% loss in speed equates to approximately a loss in head ofabout 50%. The valve must, therefore, reduce the pressure across the compressor by about half inthe same time as the compressor loses 30% of its speed. This speed loss is very rapid.

The larger the volumes are in the system, the longer it will take to equalize the pressures. Obvi-ously, the larger the valve, the better its potential to avoid surge. However, the larger the valve,the poorer its controllability at partial recycle. The faster the valve can be opened, the more flowcan pass through it. There are, however, limits to the valve opening speed, dictated by the needto control intermediate positions of the valve, as well as by practical limits to the power of theactuator. The situation may be improved by using a valve that is only boosted to open, thus com-bining high opening speed for surge avoidance with the capability to avoid oscillations by slowclosing.

2 Similar considerations are to be made for the trip of an electric motor driver. The main difference is the different inertia of the motor (and thegearbox).


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If the discharge volume is too large and the recycle valve cannot be designed to avoid surge, ashort recycle loop (hot recycle valve) may be considered, where the recycle loop does not in-clude the aftercooler.

While the behavior of the piping system can thus be predicted quite accurately, the questionabout the rate of deceleration for the compressor remains. It is possible to calculate the powerconsumption for a number of potential steady-state operating points. The operating points areimposed by the pressure in the discharge volume, which dictates the head of the compressor. Fora given speed, this determines the flow that the compressor feeds into the discharge.

In a simple system as described above, mass and momentum balance have to be maintained(Sentz [6], Kurz and White [8]). From this complete model, some simplifications can be derived,based on the type of questions that need to be answered. Obviously, for relatively short pipes,with limited volume (such as the systems desired for recycle lines), the pressure at the valve andthe pressure at compressor discharge will not be considerably different. For situations like this,the heat transfer can also be neglected. The set of equations then is reduced to:

[ ]vQQV

pk

dt

dp

= 22 (2)

The rate of flow through the valve is calculated with the standard ISA method [7]3:

5.0

222

12 11360

=

ZTSGp

ppYcFQ vpstd (3)

and

),,( 222 ZTpQQ stdstdv

= (4)

The compressibility Z2 is calculated with the Redlich-Kwong equation of state. Equations 2 , 3and 4 mean that the discharge pressure change depends on the capability of the valve to releaseflow at a higher rate than the flow coming from the compressor. It also shows that the pressurereduction for a given valve will be slower for larger pipe volumes (V). Kurz and White [8] haveshown the validity of the simplified model.

The discharge pressure p2 in Eq. 2 is a function of the compressor operating point, expressed by:

++

=

+=

N

Q

N

Q

N

h

ZT

SGNQh

k

k

p

p kk

2

2

1

11

2

287

),(11

(5 a, b)

Alternatively, a lookup table, showing the head-flow relationship for the compressor, can beused.

3 Qstd is the standard flow. Fp is the piping geometry factor. It is usually not known and can be assumed to be 1. The pressure is assumed to beconstant in the entire pipe volume. It is thus the same just upstream of the valve and at the discharge pressure of the compressor.


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The above relationship can be used for any positive flow. If the compressor exhibits reverseflow, it can be modeled as an increasing flow backwards through the compressor with timeacross the surge limit, and a 1% hysteresis in recovery head (head must decrease 1% below thesurge limit before the compressor will recover from surge). When recovery head is reached flowthrough the compressor begins decreasing, through zero and then flow begins increasing in theforward direction. This is purely for esthetics. No attempt is made to accurately represent whathappens after the compressor stalls. It should be noted that this somewhat crude formulation suf-fices for the present study because we want to determine whether the compressor will go intosurge at ESD or not. The post-surge behavior is, thus, not important and is only introduced tokeep the numerical model stable.The behavior of the compressor during ESD is governed by two effects. The inertia of the systemconsisting of the compressor, coupling and power turbine (and gearbox where applicable) iscounteracted by the torque (T) transferred into the fluid by the compressor (mechanical losses areneglected). The balance of forces thus yields:

dt

dNJT = 2 (6)

Knowing the inertia (J) of the system and measuring the speed variation with time during run-down yields the torque and, thus, the power transferred to the gas:

( )dt

dNNJNTP == 222 (7)

If the rundown would follow through similar operating points, then P~N3, which would lead to a

rundown behavior of:

( ) ( )( ) =

=+==

0

2

2

2

2

2

12

1)(

22tN

tJ

k

tNcdt

J

kdNN

N

J

k

dt

dN

(8)

Regarding the proportionality factor (k) for power and speed, this factor is fairly constant, nomatter where on the operating map the rundown event starts. Thus, the rate of deceleration,which is approximately determined by the inertia and the proportionality factor, is fairly inde-pendent of the operating point of the compressor when the shutdown occurred; i.e., the time con-stant (dN/dt(t=0)) for the rundown event is proportional to k/J. However, the higher the surgemargin is at the moment of the trip, the more head increase can be achieved by the compressor atconstant speed.

The model described above, which contains and accounts for the primary physical features ofthe discharge system, can be used to determine whether the combination of discharge volumeand valve size can prevent the compressor from surge during an ESD. It allows the two impor-tant design parameters to be easily varied to avoid surge during ESD. The surge valve size andopening speed can be increased for a given discharge volume or the maximum allowable dis-charge volume for a given configuration of valves and compressor characteristic can be limited.The second method, which has the advantage of being more transparent for the station design, isused here.


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The simplified model calculates the maximum discharge volume where the head across the com-pressor can be reduced by half in one second, based on the assumption that this reflects the speeddecay during an ESD as outlined above. Therefore, the calculation of the instant compressorspeed is replaced by a fixed, presumed to be known, deceleration rate. The assumption is madethat the power turbine and compressor will lose about 20 to 30 % speed in the first second of de-celeration. This is, for example, confirmed by data from Kurz and White [8] showing a 30%speed reduction of a gas turbine driven compressor set, and Bakken et. al. [9], where the gas tur-bine driven configurations lost about 20-to-25% speed in the first second, while the electric mo-tor driven configuration lost 30% speed in the first second. As a result of the loss of 25% speed,the head the compressor can produce at the surge line is about 56% lower than at the initialspeed, if the fan law is applied. A further assumption is made about the operating point to be thedesign point at the instant of the ESD.

Any ESD is initiated by the control system. Various delays in the system are caused by the timefor the fuel valve to shut completely, the time until the hot pressurized gas supply to the powerturbine seizes, and the opening time of the recycle valve. ESD data show it is a valid assumptionthat the surge control valve reaches full open simultaneously with the beginning of deceleration

of the power turbine / compressor. This is the starting time (T0) for the model.Usually, the suction volume (no check valve) is more than three orders of magnitude greater thanthe discharge volume and is therefore considered at a constant pressure. The general idea is nowto consider only the mass flow into the piping volume (from the compressor) and the mass flowleaving this volume through the recycle valve. Since the gas mass in the piping volume deter-mines the density and, thus, the pressure in the gas, we can for any instant see whether the headrequired to deliver gas at the pressure in the pipe volume exceeds the maximum head that thecompressor can produce at this instant. Only if the compressor is always capable of making morehead than required can surge be avoided.

A further conceptual simplification can be made by splitting the flow coefficient of the recyclevalve (cv) into a part that is necessary to release the flow at the steady-state operating point of thecompressor (cv,ss) and the part that is actually available to reduce the pressure in the piping vol-ume (cv,avail).

The first stream and, thus, cv,ss of the valve necessary to cover it are known. Also known is the cvrating of the valve. Thus, the flow portion that can effectively reduce the backpressure is the de-termined by the difference:

ssvvavailv ccc ,, = (9)

The model is run at constant temperature. Most of the compressor systems modeled contain af-

tercoolers. The thermal capacity of the cooler and the piping are much larger than the thermalcapacity of the gas; thus, the gas temperature changes are negligible within the first second.

The flow calculated above in each step of the iteration is then subtracted from the gas containedin the discharge and a new pressure in the pipe volume is calculated. The calculation yields themaximum allowable piping volume for the set parameters that will not cause surge at ESD.

Application


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The simplified model lends itself to iteratively determine the maximum allowable discharge vol-ume for a given valve configuration. This is important, because the valve size can be determinedearly in the project phase, based on the considerations outlined in Figure 5. With a known valveconfiguration, the station designer can be provided with the maximum volume of piping andcoolers between the compressor and the check valve, that allows to avoid surge during ESDs.The calculation requires to specify the head-flow-speed relationship of the compressor, and thedefinition of the surge line as a function of either compressor speed, compressor head or com-pressor flow. Further, the valve needs to be described by its maximum capacity (cv), as well asby its capacity as a function of valve travel (cv~(travel)

3for a ball valve, cv~(travel)

2for a globe

valve), and the opening behavior, including the delay. The train deceleration is modeled asshown in Figure 8, based on test data and the calculations outlined by eq. 7 and 8. The dischargecheck valve is assumed to be closed as soon as the recycle flow exceeds the compressor flow, i.e.the depressurization begins. The calculation procedure is started by initiating the deceleration ofthe train and the valve opening. For each time step, the compressor head and flow (based onspeed and system pressures), and the flow through the valve (based on system pressures andvalve opening) are calculated. The mass of the gas trapped between surge valve and compressordischarge is subsequently determined, yielding a new discharge pressure. If surge occurs (i.e if

the flow drops below the flow at the surge line), the backwards flow through the compressor isassumed to increase with time in surge, with a recovery once the required head drops 1% belowthe head at surge. The modeling of the backwards flow is not critical (and it is only made toavoid numerical instabilities), because the only information that is expected from the model iswhether the compressor will surge for the given configuration, or not.

Figures 9 and 10 show typical results of these simulations. In Figure 11, the discharge volume issmall enough , and while the actual flow of the compressor approaches the minimum allowableflow (surge flow) at about 500ms after the initiation of the ESD , surge can be avoided.


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Figure 8: Valve opening characteristic and power turbine deceleration

Figure 9: Actual Flow and Flow at the surge line during ESD. Recycle valve sizing and dis-charge volume allow for ESD without surge


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Figure 10: Actual Flow and Flow at the surge line during ESD. Compressor is in surge at 0.6 safter trip.

The simulation results for another configuration (Figure 10) show a surge event at about 700msafter the initiation of the ESD. For this configuration, either the valve size has to be increased, orthe discharge volume has to be reduced, to avoid compressor surge during an emergency shut-down.

Start-up ConsiderationsThe design of the anti surge and recycle system also impacts the start-up of the station. Particularattention has to be given to the capability to start up the station without having to abort the startdue to conditions where allowable operating conditions are exceeded. Problems may arise fromthe fact that the compressor may spend a certain amount of time recycling gas, until sufficientdischarge pressure is produced to open the discharge check valve (Figure 3), and gas is flowinginto the pipeline.

Virtually all of the mechanical energy absorbed by the compressor is converted into heat in thedischarged gas. In an un-cooled recycle system, this heat is recycled into the compressor suctionand then more energy added to it. A cubic foot of natural gas at 600 psi weighs about 2 lb (de-pending on composition). The specific heat of natural gas is about 0.5 Btu/lb (again dependingon composition). 1 Btu/sec equals 1.416 hp. If the recycle system contains 1000 cubic feet, thereis a ton of gas in it. 1416 hp will raise the temperature of the gas about 1 degree per second. This

approximates what happens with 100% recycle. At 100% recycle, eventually this will lead tooverheating at the compressor discharge. The problem usually occurs when there is a long pe-riod between the initial rotation of the compressor and overcoming the pressure downstream ofthe check valve.

Low pressure ratio compressors often do not require aftercoolers. There are three primary strate-gies that can be employed to avoid overheating the un-cooled compressor during start-up:


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-accelerate quickly-delay hot gas re-entering the compressor-throttled recycle.

Compressors without cooling must be accelerated and placed on line quickly to avoid overheat-ing. Un-cooled compressor sets cannot be started and accelerated to idle. They must be acceler-ated quickly through the point where the discharge check valve opens and the recycle valvecloses. If acceleration slows when the discharge pressure is met, and recycle valve closesslowly, a shutdown may still occur. Often standard start sequences are very conservative andcan be shortened to reduce the time it takes to get a compressor on line.

Extending the length of the recycle line downstream of the recycle valve increases the total vol-ume of gas in the recycle system. This reduces the heat buildup rate by delaying when the hotgas from the compressor discharge reaches the suction. Some heat will be radiated through thepipe walls. If the outlet is far upstream into a flowing suction header, dilution will occur.

Figure 11 outlines a solution to a rather difficult starting problem for a compressor station with-

out after cooling capacity: To start the first unit is relatively easy, because there is virtually nopressure differential across the main line check valve, and therefore the unit check valve willopen almost immediately, allowing the flow of compressed gas into the pipeline. However, ifone additional unit is to be started, the station already operates at a considerable pressure ratio,and therefore the unit check valve will not open until the pressure ratio of the starting unit ex-ceeds the station pressure ratio. Ordinarily the unit would invariably shut down on high tempera-ture before this can be achieved. By routing the recycle line into the common station header, theheat from the unit coming on line is mixed with the station suction flow. This equalizes the inlettemperature of all compressors; higher for the compressors already on line, lower for the com-pressor coming on line. With this arrangement overheating of a compressor coming on line isnearly always avoided.


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Figure 11 a and b: Original (a) and improved (b) station layout. The original layout features indi-vidual recycle lines, while the improve layout allows to feed the recycled flow into the commonstation header.


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Figure 12:Temperature build up in a recycle loop, consisting of 3000 ft of 24in pipe. Compressorshutdown levels would be reached after about 20 minutes (assuming a shut down setpoint of

350 F).

Figure 13a,b,c: Temperature rise in the recycle loop during startup at a function of (a)Power tur-bine and compressor speed, (b) gas producer speed, (c) time (in minutes). The power turbinestarts to turn at about 75% gas producer speed, at which point the temperature starts to rise. Af-ter the discharge check valve opens (at 0.2 minutes after the compressor starts to rotate), 95%gas producer speed and 70% power turbine speed), the temperature drops rapidly.


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Figure 12 shows the problem of a conventional system without after-cooling: The temperature inthe recycle line starts to rise, eventually leading to a shut down of the compressor once the settemperature limits are reached. Figure 13 outlines the start-up event with the revised system. The

power turbine and the compressor start to rotate once the gas producer provides sufficient power.Subsequently, the gas temperature rises, but, because the discharge pressure required to openthe check valve is reached fast enough, overheating can be avoided.

Further analysis of the start-up problem indicates the advantage of throttling the recycle valve,rather than starting the unit with the recycle valve fully open. Figure 5 illustrates this: At 70%open setting, the startup of the compressor is relatively closer to the surge line than at 100% opensetting. For any given speed, the power requirement of the compressor is lower when it is closerto surge than when it is farther in choke. Therefore, for a given amount of available power, thestart is quicker if the compressor operates closer to surge. If the rate of acceleration is quicker,the heat input into the system is lower. Actively modulating the surge during start-up is virtually

impossible as the parameters defining the surge limit of the compressor are too low to be practi-cally measured. Returning to Figure 5 the surge limit of a compressor matches well with a fixedtravel (constant Cv ) line for a recycle valve. As such, a compressor can be started with a fixedrecycle valve position

Conclusions and RecommendationsA model to simulate shutdown events was developed and used to define simpler rules that helpwith proper sizing of upstream and downstream piping systems, as well as the necessary controlelements. The model coincides well with the data, particularly with regards to proper predictionof surge events. The inaccuracies and limitations inherent in the current model are only problem-

atic if the entire rundown process needs to be described. The key variables-compressor characteristic-valve size and characteristic-discharge volumeare included into a simple, yet sufficiently accurate, modeling procedure.Furthermore, the related issue of recycle loop sizing and design considering the requirement tobring units on-line without overheating the system are discussed.

AcknowledgementsThe authors would like to acknowledge the contributions of Daryl D. Legrand and Roland Kai-ser, both with Solar Turbines Incorporated.

References

[1] Epstein, A.H., Ffowcs Williams, J.E., and Greitzer, E.M., Active Suppression of Compres-sor Instabilities, AIAA-86-1994, 1994.

[2] Blanchini, F., Giannatasio, P., Micheli, D., and Pinamonti, P., Experimental Evaluation of aHigh-Gain Control for Compressor Surge Suppression, ASME 2001-GT-0570, 2001.

[3] Arnulfi, G.L., Giannatasio, P., Micheli, D., and Pinamonti, P., An Innovative Control ofSurge in Industrial Compression Systems, ASME 2000-GT-352, 2000.


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[4] McKee, R.J., and Deffenbaugh, D., Factors that Affect Surge Precursors in CentrifugalCompressors, Proc. GMRC Gas Machinery Conference, Salt Lake City, Utah, 2003.

[5] Day, I.J., Axial Compressor Performance During Surge, Proc.10th

International Sympo-sium on Air Breathing Engines, Nottingham, UK,1991, pp. 927-934.

[6] Sentz, R.H., The Analysis of Surge, Texas A&M Turbomachinery Symposium, 1980.[7] ANSI/ISA S75.01, Flow Equations for Sizing Control Valves, 1995.[8] Kurz, R., White, R.C., 2004, Surge Avoidance in Gas Compression Systems, TransASME

JTurbo Vol.126, pp.501-506.[9] Bakken, L.E., Bjorge, T., Bradley, T.M., and Smith, N., Validation of Compressor Transient

Behav ior, ASME GT-2002-30279, 2002.

surge prevention cent r if compr sys

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