Centrifugal Compressors

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Engineering Design Guide: GBHE-EDG-MAC-1134

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Engineering Design Guide: Centrifugal Compressors CONTENTS SECTION SECTION ONE - ANTI-SURGE PROTECTION AND THROUGHPUT REGULATION 0 INTRODUCTION 1 SCOPE 1 2 MACHINE CHARACTERISTICS 2

2.1 Characteristics of a Single Compressor Stage

2.2 Characteristic of a Multiple Stage Having More Than One Impeller

2.3 Use of Compressor Characteristics in Throughput

Regulation Schemes 3 MECHANISM AND EFFECTS OF SURGE 3

3.1 Basic Flow Instabilities

3.2 Occurrence of Surge

3.3 Intensity of Surge

3.4 Effects of Surge

3.5 Avoidance of Surge

3.6 Recovery from Surge



4 CONTROL SCHEMES INCLUDING SURGE PROTECTION

4.1 Output Control

4.2 Surge Protection

4.3 Surge Detection and Recovery 5 DYNAMIC CONSIDERATIONS 5

5.1 Interaction

5.2 Speed of Response of Antisurge Control System 6 SYSTEM EQUIPMENT SPECIFICATIONS 6

6.1 The Antisurge Control Valve

6.2 Non-return Valve

6.3 Pressure and flow measurement

6.4 Signal transmission 6.5 Controllers

7 TESTING 7

7.1 Determination of the Surge Line

7.2 Records



8 INLET GUIDE VANE UNITS 8

8.1 Application

8.2 Effect on Power Consumption of the Compressor

8.3 Effect of Gas Conditions, Properties and Contaminants

8.4 Aerodynamic Considerations

8.5 Control System Linearity

8.6 Actuator Specification

8.7 Avoidance of Surge 8.8 Features of Link Mechanisms

8.9 Limit Stops and Shear Links



APPENDICES A LIST OF SYMBOLS AND PREFERRED UNITS B WORKED EXAMPLE 1 COMPRESSOR WITH VARIABLE INLET

PRESSURE AND VARIABLE GAS COMPOSITION C WORKED EXAMPLE 2 A CONSTANT SPEED ~ STAGE COMPRESSOR

WITH INTERCOOLING D WORKED EXAMPLE 3 DYNAMIC RESPONSE OF THE ANTISURGE

PROTECTION SYSTEM FOR A SERVICE AIR COMPRESSOR RUNNING AT CONSTANT SPEED

E EXAMPLE OF INLET GUIDE VANE REGULATION FIGURES 2.1 TYPICAL COMPRESSOR STAGE CHARACTERISTIC PLOTTED

WITH FLOW AT DISCHARGE CONDITIONS 2.2 TYPICAL COMPRESSOR STAGE CHARACTERISTIC PLOTTED WITH

FLOW AT INLET CONDITIONS 2.3 PERFORMANCE CHARACTERISTICS OF A COMPRESSOR STAGE

AT VARYING SPEEDS 2.4 SYSTEM WORKING POINT DEFINED BY INTERSECTION OF

PROCESS AND COMPRESSOR CHARACTERISTICS 2.5 DISCHARGE THROTTLE REGULATION 2.6 BYPASS REGULATION 2.7 INLET THROTTLE REGULATION 2.8 INLET GUIDE VANE REGULATION 2.9 VARLABLE SPEED REGULATION



3.1 GAS PULSATION LEVELS FOR A CENTRIFUGAL COMPRESSOR 3.2 REPRESENTATION OF CYCLIC FLOW DURING SURGE OF LONG

PERIOD 3.3 TYPICAL WAVEFORM OF DISCHARGE PRESSURE DURING SURGE 3.4 MULTIPLE SURGE LINE FOR A MULTISTAGE CENTRIFUGAL

COMPRESSOR 3.5 TYPICAL MULTIPLE SURGE LINES FOR SINGLE STAGE AXIAL-FLOW

COMPRESSOR 4.1 GENERAL SCHEMATIC FOR COMPRESSORS OPERATING IN

PARALLEL TO FEED MULTIPLE USER PLANTS 4.2 ILLUSTRATION OF SAFETY MARGIN BETWEEN SURGE POINT AND

SURGE PROTECTION POINT AT WHICH ANTISURGE SYSTEM IS ACTIVATED

4.3 ANTISURGE SYSTEM FOR COMPRESSOR WITH FLAT

PERFORMANCE CHARACTERISTIC 4.4 ANTISURGE SYSTEM FOR COMPRESSOR WITH STEEP

PERFORMANCE CHARACTERISTIC 4.5 SIMPLIFIED FILIPPINI ANTISURGE SYSTEM 4.6 ANTISURGE SYSTEM FOR COMPRESSOR WITH CONSTANT SPEED

BUT VARIABLE INLET CONDITIONS 4.7 CRITERLA FOR ANTISURGE CONTROL VALVE SIZE 4.8 ALTERNATIVE ARRANGEMENTS OF ANTISURGE VALVES FOR A

MULTISTAGE COMPRESSOR 4.9 TYPICAL REFRIGERATION SERVICE COMPRESSOR ILLUSTRATING

PASS-IN SIDESTREAM AND MULTIPLE BY-PASS LOOPS 5.1 SEPARATION OF ACTIONS OF OUTPUT CONTROL AND ANTISURGE

CONTROL LOOPS



5.2 ALTERNATIVE NON-RETURN VALVE ARRANGEMENTS FOR BOOSTERS

6.1 DIFFERENTLATING NETWORKS FOR PULSE DETECTORS 8.1 RELATION BETWEEN FLOW AND IGV ACTUATOR POSITION FIGURES APPENDICES: B1.1 SCHEMATIC COMPRESSOR SYSTEM ARRANGEMENT Bl.2 COMPRESSOR CHARACTERISTICS Bl.3 FILIPPINI APPROXIMATION TO POLYTROPIC HBAU Bl.4 CONTROL DIAGRAM FOR ANT 1 SURGE SYSTEM C2.1 SCHEMATIC OF COMPRESSOR C2.2 COMPRESSOR AND PROCESS CHARACTERISTICS C2.3 CONTROL DIAGRAM FOR ANTISURGE SYSTEM D3.1 SCHEMATIC OF COMPRESSOR ARRANGEMENT D3.2 COMPRESSOR CHARACTERISTIC D3.3 DYNAMICS OF SYSTEM E4.1 MACHINE & PROCESS CHARACTERISTICS E4.2 IGV ANGLE RELATION TO FLOII RATE E4.3 SCHEMATIC OF MECHANISM E4.4 IGV ANGLE Vs ACTUATOR STROKE GRAPH E4. 5 FLOWRATE V s ACTUATOR STROKE GRAPH E4.6 MOTION DIAGRAMS FOR LINK D



E4.7 MECHANICAL ADVANTAGE Vs IGV ANGLE E4.8 IGV MOVEMENT Vs INPUT SPINDLE ANGULAR POSITION E4 .9 ACTUATOR THRUST V s ACTUATOR STROKE E4.10 SHEAR PIN ARRANGEMENT E4.11 LIMIT STOPS BIBLIOGRAPHY DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE



SECTION ONE - ANTI-SURGE PROTECTION AND THROUGHPUT REGULATION 0 INTRODUCTION The function of compressor throughput regulation is to match the compressor performance to the process requirements. Further functions which require another control system are: (a) to prevent the machine transgressing limits which incur unsteady

operation with the possibility of damaging the process or the machine itself. Margins of safety in flow or in pressure differential should be incorporated.

(b) to limit the maximum attainable pressure. 1 SCOPE This Engineering Design Guide Section covers the design of a complete system for the throughput regulation and surge protection for a centrifugal compressor. It should always be applied to compressors in Groups 1 and 2 as defined in Engineering Procedure GBHE-EDP-MAC-3301. 2 MACHINE CHARACTERISTICS The construction and working principles of compressors are described in literature and in other sections of this Engineering Design Guide. 2.1 Characteristics of a Single Compressor Stage Historically, the performance of centrifugal gas compressor stages has been displayed by plotting characteristics in parameters derived from dimensional analysis. Such characteristics employed the pressure ratio r, against a mass flow function m, for a series of constant notional speeds of rotation U, where the customary parameters were defined as:



R = Pd ……… (2.1) Pi Pd = absolute total pressure at discharge bar abs Pi = absolute total pressure at inlet bar abs M = w x T i

½ ……… (2.2) Pi W = mass flow kg/s Ti = absolute total temperature at inlet k U = N ……… (2.3) T i

½ N = impeller rotational speed revs/s Other parameters are the Reynolds Number and the Mach Number. The Reynolds Number is usually neglected because the effect is small on commercially available machines where the discharge volume flow exceeds 0.25 m 3/s for centrifugal compressors or 3 m 3/s for axial flow compressors. The Mach Number is defined as V , where a a is the velocity of sound in the process gas, taken by convention at the

compressor stage inlet conditions. v is the velocity of a representative machine element, normally taken by

convention as the peripheral velocity of the impeller tip, but sometimes referred to as the peripheral velocity of the centrifugal impeller eye.



This set of parameters is much used in machine design. For system studies more useful characteristics are obtained by employing the concepts of polytrophic head and polytrophic efficiency. These are based on the impeller action being adiabatic but not isentropic so that changes follow the empirical relationship: P x V n = constant ……(2.4) where n is the polytrophic exponent determined by experiment for a particular compression stage when operated at a given flow and speed. Then the polytrophic head is defined as: HP = f x n x Z i Ro Ti r n-1/n - 1 ..…… (2.5) n-1 M where M is the molecular weight of the gas, Ro is the universal gas constant, 8.3143 kJ/kg/K, Zi is the compressibility at inlet conditions, defined as P iV I M R o T i where Vi is the specific volume of the gas in m3 /kg f is the polytrophic head factor which corrects for variations in n as the

pressure increases from Pi to Pd

For non-ideal gases it is taken to be the value calculated assuming n = δ. The polytrophic exponent n is to the isentropic exponent k by the polytrophic

efficiency ᶯP defined by:

ᶯP = n (k – 1) …….. (2.6) K(n – 1)



For an ideal gas, k = Ɣ = Cp /C v but for real gases k is defined only as the volume exponent. The exponent k is normally taken as the geometric mean of the values at inlet and discharge conditions. It is found empirically that the polytrophic efficiency is not affected by changes in k, Z or M but depends only on flow and rotational speed provided the Mach Number is low and the Reynolds Number is high. The polytrophic head can then be written as:

Caution is needed when interpreting manufacturers data because in some fields the sparseness of process gas data has led to a convention using a pseudo-polytrophic efficiency obtained by substituting Ɣ, for k and by assuming f = 1. The compressor characteristic is plotted as polytrophic head against volume flow for a given rotational speed. A single curve is obtained by using the actual volume flow at discharge conditions, see Fig 2.1. However, characteristics are conventionally plotted using the actual volume flow at inlet conditions as an approximation, see Fig 2.2. 2.2 Characteristic of a Multiple Stage having more than One Impeller In order to develop high head, compressors are built with a series of impellers so that the head rise is cumulative. The polytrophic head equation remains true, provided that there is no cooling and that there are no sidestreams. A compression stage is then conveniently defined as that section of the machine between such coolers or sidestreams. The performance of any particular stage is dependent upon the composition and temperature of the gas entering it and is thus a function of both intercooling and the performance of the earlier stages. Because cooling varies with atmospheric or cooling water temperature and transiently when load on the machine changes,



it follows that individual stage characteristics cannot be combined to give an invariant single overall compressor characteristic. To obtain the overall machine performance it is necessary to calculate the operating conditions for the first stage to derive the inlet conditions for the second and so on through the machine. 2.3 Use of Compressor Characteristics in Throughput Regulation

Schemes Fig 2.4 shows a typical process characteristic plotted as process inlet pressure against standard flow. The inlet pressure comprises a static head component, Ps, and a frictional component which increases with flow. Note that this is a steady-state representation and care should be taken to avoid misleading deductions about changes which are too fast for equilibrium to be maintained. The machine discharge pressure is equal to the process inlet pressure at the same flow. It is then convenient to re-plot the compressor characteristic by calculating from the polytrophic head curve the machine discharge pressures corresponding to a range of standard flows. The point of intersection of the compressor and the process characteristic curves, X, then indicates the mutually compatible conditions under which the compressor and the process will operate. Where two or more machines operate in parallel, their characteristics are combined to give the equivalent single characteristic. 2.3.1 Discharge Throttle Regulation The insertion of a series throttle valve upstream of the process will cause some increase in the resistance to flow, even when the valve is in the fully open position (from Xo to XI in Figure 2.5). Gradual closing of the valve will cause point X2 to move back along the compressor characteristic to reduce the flow. The vertical distance between X2 and Xo is the pressure dissipated across the valve and indicates the energy loss inherent in this form of control. This information provides the basis for determining the control valve size. A typical application is the distribution of flows to process units operating in parallel from the same machine.



2.3.2 Bypass Regulation For a constant speed compressor, a bypass valve may be used to turn down the plant throughput by recycling part of the flow. Bypass regulation can be used only when the compressor characteristic slope gives the required change in discharge pressure for an acceptable change in bypass flow. The construction of Fig 2.6 using the same ordinates as for discharge throttling, shows that the bypass gives an apparent decrease in process resistance. However, whilst the process flow falls to Q2 the total flow rises to Q1. This method avoids the problem of surging. Power losses are comparable to those incurred by discharge throttling. Bypass regulation invariably supplements other methods of regulation when it is necessary to extend the range of throughput control below the capacity limit set by approach to machine surge. Except on air compressors where simple blow-offs are invariably used, bypass regulation returns the process gas to the compressor stage inlet. Any condensation, together with the change in effective inlet gas temperature, alters the re-plotted compressor characteristic. It is then convenient to plot a family of curves for a range of machine inlet gas conditions. 2.3.3 Inlet Throttle Regulation This method reduces the pressure at the machine inlet, thus reducing the inlet gas density. The most convenient presentation is the family of curves obtained by plotting discharge pressure against standard flow for a range of machine inlet pressures (see Fig 2.7). On this diagram both the upstream and the downstream process characteristics can be plotted. The intersections (X) of the machine characteristic with the downstream process characteristic defines the machine working point. The corresponding intersections (Y) then define the working points for the inlet valve from which a new graph of pressure drop across the valve (P u – P i) against standard flow can be constructed. This curve enables the valve size to be selected.



This method of regulation is commonly used because the power loss is less than that for discharge throttling and because the machine operating point is also further from surge. 2.3.4 Inlet Guide Vane Regulation An IGV unit comprises a number of vanes, spaced round the circumference of the machine inlet, which can be moved in synchronism. The vanes impart an angular velocity to the gas, altering the head developed by the machine. For large movements of the vanes the angular deflection of the gas becomes less significant than the restriction caused by reduction of the passages between the vanes and the behavior reverts to that of inlet throttling. Customarily, the machine characteristics are presented as a family of curves (discharge pressure Pd against standard volume flow) each for a given angular setting of the guide vanes (see Fig 2.8), assuming that the speed and inlet conditions are both constant. Guide vanes change the shape of the surge line, giving a greater capacity range when the compressor discharges against a substantially constant pressure. Besides conferring greater flexibility, inlet guide vanes achieve the regulation more efficiently than inlet throttle regulation. The IGV unit is mechanically complicated and can only be provided as an integral part of the compressor. The construction is considered in more detail under Clause 8 of this Engineering Design Guide Section. 2.3.5 Speed Regulation Increasing the rotational speed increases the compressor performance. Given the Hp/Q characteristic (polytrophic head against volume flow) at speed N1 the characteristic for another speed N2 can be estimated by applying the relationships:



Note that these relationships are inaccurate when the compressor works at high Mach Number (see Fig 2.3) or low Reynolds Number. The compressor characteristics are plotted as a family of curves relating discharge pressure to standard volume flow, each curve representing a constant rotational speed (see Fig 2.9). This is used to obtain the relation between speed and flow, from which is obtained the specification for the speed governor on the driver. Speed variation provides the most economical method of throughput regulation when the process head is itself flow-dependent. The driver is customarily a steam turbine for variable speed duties: an important design consideration is that the compressor delivery is sometimes so sensitive to speed that conventional speed governors are taken to the limit of their performance. 3 MECHANISM AND EFFECTS OF SURGE The flow pattern through the machine changes with rate and usually becomes unstable when the flow is reduced. The point where this instability occurs is known as the surge limit and its locus over the whole range of throughput regulation is the surge line. Typically the surge limit occurs at 60 - 80% of design flow for centrifugal compressors and 85 - 95% of design flow for axial compressors (running at constant speed and constant inlet gas conditions). 3.1 Basic Flow Instabilities These are conveniently grouped as follows: (a) Intrinsic Perturbations

The steady-state characteristics for a machine are conventionally drawn as smooth continuous lines but at each point there is a spectrum of low intensity pressure perturbations covering a wide frequency band and giving rise to gas-borne noise levels ranging from 60 dB for a quiet ventilation fan to 140 dB for a large process gas compressor (see Fig 3.1).

The high frequency perturbations merge into high level turbulence whilst the low frequencies are dominated by residual effects of flow separation, especially vane wakes. Stable vortex formations arising in this way give not only slow variations in the mean steady state values, but also



variations between static pressure measurements at neighbouring points in the pipework.

(b) Cyclic Flow

A single compressor stage usually has a head/flow characteristic exhibiting a maximum (Fig 3.2) so that at low flow the gradient of the characteristic is positive. When an attempt is made to operate on this positive gradient section of the characteristic, the flow perturbations mentioned above are amplified and a steady state of cyclic flow is quickly reached. Under these conditions the mean flow is still forward [4].

The point at which cyclic flow is first self-sustaining is not necessarily the maximum of the characteristic. It is very nearly so when the compressor has a well-matched vane less diffuser; the flow instability then normally first occurs first near the impeller tip.

Surge in a multistage machine will be initiated by one of the stages reaching its critical limit and this is unrelated to the shape of the overall machine characteristics.

(c) Stalling

A compressor stage may have a number of different internal flow patterns, each being quasi-stable, but which result in flow instability caused by periodic changes from one pattern to another.

Nearly all such patterns are produced by boundary-layer flow separation (i.e. stalling). This produces a number of cells of low-energy gas which then act as spurious boundaries for normal flow. Such cells usually propagate in the same direction as the impeller rotation but at about half its rotational speed. The phenomenon is termed 'rotating stall' and invariably occurs in axial flow compressors and in centrifugal compressors having vaned diffusers.



3.2 Occurrence of Surge 3.2.1 Non-steady Unidirectional Flow This is sometimes called incipient surge and is encountered in two situations; as 'rotating stall' and as cyclic flow in centrifugal fans. The critical stage may supply sufficient energy to excite acoustic resonance phenomena in associated piping/vessel systems having frequencies of 20 to 500 Hz and these often excite the natural frequencies of pressure gauges, giving warning of imminent surge. 3.2.2 Bidirectional Flow As the flow through the compressor is reduced, the eventual result is surging manifest as a relaxation oscillation of flow. The period of the oscillation depends on the impedance of the associated piping together with the volume of any attached vessels, but is commonly 0.5 to 5.0 seconds. For typical process gas compressors, the magnitude of the reverse flow 1s of the order of 30% of the design forward flow. 3.3 Intensity of Surge The intensity of surge increases as the mean forward flow is reduced. The potential intensity also increases rapidly with compression ratio. It is not uncommon to be able to run small centrifugal fans (with compression ratio less than 1.1) continuously in the nominal surge region without untoward effect. On the other hand, large multistage centrifugal compressors with compression ratios of 10:1 and vaned diffusers have been known to suffer impeller disintegration within 1000 hrs when running with externally imperceptible rotating stall. Intensity is correlated with the wave form of the pressure surge; the steeply falling pressure wave front gradient is of the order of the discharge pressure/second and may be as great as 100 bar/s (see Fig 3.3).



For machines with high aerodynamic loading, the surge initiation can be very fast and the passage of the first wave front back through the inlet system can sharply startle nearby personnel. 3.4 Effects of Surge 3.4.1 Effect on Processes and Process Control (a) When surge develops, the mean forward process flow drops sharply to an

unacceptably low level, which may stop the process even if there is no overt trip.

(b) The propagation of large cyclic pressure fluctuations through a system can

damage packing in gas scrubbers, catalyst in reactors and impose severe loads on piping anchors and vessel foundations.

(c) Reverse flow of gas through inlet gas filters normally dislodges the dirt

trapped by the filter media and cyclic flow encourages dust migration to the clean side. Long term surging can denature filter media.

Even small cyclic flows can markedly reduce the efficiency of liquid/gas separators by re-entrainment from knitted wire mesh elements.

(d) Flow measurement is notoriously susceptible to pulsations and surging

may thus cause incorrect control action.

Pressure measurement devices may also suffer from the repeated surges in pressure, necessitating recalibration.

3.4.2 Effects on Compressor and Driver (a) The axial gas pressure load balance on rotors is disturbed so that the

machine thrust bearings are subjected to overload and to load reversal. (b) The surging flow is reflected in fluctuations of the torque in the shaft, thus

exciting the natural torsional vibration frequencies.

There are two particular consequences of such torsional vibration: firstly the natural frequencies of free-standing blades in axial compressors and turbine drivers may be excited and the blade bending stresses increased by factors of 5 to 10; secondly, gear teeth may suffer impact loading, increasing stresses by a factor of 3 or more.



(c) Rotating stall directly excites the natural vibration frequencies of compressor impeller or diffuser vanes. Damage to a compressor subjected to occasional surges is cumulative in the sense that the stress cycles contribute to eventual material failure in fatigue.

The compressor life expectancy then depends principally on the machine type; typically from 500 surge cycles for a large multistage axial flow compressor to an indefinitely long life for a small robust centrifugal fan.

(d) When surging continues for more than a few cycles there may be an

additional risk of damage due to a rise in temperature.

This hazard is acute for uncooled multistage axial flow compressors where the inlet gas temperature may rise 3000C in 30 s; it is insignificant for cooled multistage centrifugal compressors of the isotherm type.

3.5 Avoidance of Surge It is possible to construct surge-free centrifugal compressors. The principal method is to use an impeller with markedly backswept vanes and to dissipate the vane wakes in a rotating diffuser which, in practice, forms an integral part of the impeller. The range of commercially available machines of this type is very limited. In general, it is necessary to avoid operation in the surging zone by introducing an auxiliary system which will maintain the flow through the machine when the process gas demand falls below some critical value. For multistage centrifugal compressors designed for high pressure ratios but operated at speeds much lower than the design speed, the small flow passages in the final stages present a high impedance to the preceding stages. In these circumstances the early stages may be driven into surge regardless of the operation of the external discharge bypass valve. It is then necessary to provide an auxiliary bypass or blow-off from an intermediate stage. This has the effect of reducing the surge limit over the range of low flows and speeds (see Fig 3.4).



3.6 Recovery from Surge Recovery from surge can be accomplished only by increasing the mean flow through the machine. If incipient surge occurs there is no difficulty; the process is fully reversible and a small increment in flow will cause the unstable flow pattern to disappear. True surge is a relaxation or limit-cycle oscillation subject to hysteresis. Stability can be restored only by increasing the mean flow to a value considerably above that at which the surge commenced (see Fig 3.2). As a guide, the flow should reach the full design value at the appropriate speed before allowing the antisurge system to reduce the newly established stable flow back to the minimum flow point. 4 CCNTROL SCHEMES INCLUDING SURGE PROTECTION The first task is to define the required steady-state conditions. The second is to identify the sources of disturbance, including deliberate changes in rate as well as emergency situations, and test the schemes to ensure that they will converge to the required steady state conditions from whatever disturbance is applied. 4.1 Output Control 4.1.1 Self-regulation In systems where the compressor runs at constant speed and where the upstream impedance is high, there is a degree of self regulation because any change in inlet pressure consequent upon a change in flow tends to alter the machine capacity to restore the flow. Where the compressor driver is a steam turbine and the speed of the set is consequently variable, the throughput may be constrained by a power limit. If the compressor inlet and discharge pressures remain nearly constant, operation at a power limit implies constant flow.



4.1.2 Controlled Discharge Pressure This requirement occurs where the compressor feeds a header from which gas is distributed to multiple process plant units operating in parallel. The general situation is represented by Fig 4.1. A multi-machine installation introduces further complication because machines working in parallel should be stable and should share the load reasonably equitably. A danger arises when the machines have very flat characteristics so that flows are sensitive to small changes in back pressure. Under these circumstances one machine may take all the load whilst the other is driven into surge, or the load may swing from one machine to the other at regular intervals, commonly termed 'hunting'. Even though only one machine is normally in use, this problem may arise when the running machine has to be interchanged with the standby, because at some stage of the changeover both run on-line together. THE ONLY SATISFACTORY CURE FOR THIS INTERACTION IS TO INSERT SUFFICIENT SERIES RESISTANCE INTO EACH MACHINE'S DISCHARGE. Sufficiently close matching is obtained by applying the control signal in common to the chosen regulator on each machine. Caution is needed if the machines are obtained from the supplier as packages, each with its own discharge pressure controller. Integral action cannot be employed on each individual controller because this arrangement will not share the load. Only proportional-action is permitted. There are then two choices in setting up the control loops. One is to make the controllers sensitive (narrow proportional bands) and stagger the set points so that only one controller at a time is modulating its output. The other is to use a much lower sensitivity on each controller so that they can be used in parallel with the same set point. A single master pressure controller incorporating integral action should then be deployed in cascade to adjust the proportional-only controllers' set points.



4.1.3 Controlled Flow Flow control inherently increases discharge pressure in response to increasing restriction of delivery flow. Over-ride pressure control is required. A trip action which blocks the discharge would call for maximum pressure from the machine. The antisurge system should be actuated from the trip action itself, because the process is necessarily interrupted and no finesse is required. The sensitivity of flow to speed variation may make speed control impracticable. Multi-machine installations are seldom based on regulation by a common flow controller. Any such scheme requires a quantitative study using simulation procedures. 4.1.4 Controlled Inlet Pressure This is dealt with as Clause 4.1.2. 4.2 Surge Protection For systems well able to cope with the effects of surge pulses, protection can be provided solely by a detector (see Clause 4.3) linked to the control system or a machine trip. The standard arrangement prevents surging by incorporating a control system which detects the approach towards surge and takes action quickly to halt that approach. For a single stage, the parameter by which the approach to surge may be inferred is actual volume flow measured at the compressor stage discharge. If its value falls below a predetermined set point then a conventional controller opens a blow-off or bypass valve to prevent further reduction of flow through the compressor. It is important to note that this type of antisurge system is designed to prevent the compressor from ever getting into surge, and may not be effective for bringing the compressor out of the surge condition.



It is necessary to provide a margin between the surge line and the chosen protection line to cater for the following considerations: (a) If surge is approached rapidly by sudden interruption of the process flow,

the margin provides a short time delay between initiation of antisurge control action and arrival at the surge line, during which time the antisurge valve should open by the required amount.

Note that the surge may also be approached rapidly when there is a reduction in compressor speed. A particular case is the speed reduction consequent upon a voltage/frequency dip in the supply to an electric motor drive.

(b) The position of the surge line is not well known and may change with age

and fouling of the compressor surfaces and with the performance of the interstage coolers.

(c) The protection line does not follow accurately the shape of the surge line. For efficient and flexible operation of the plant the margin should be small; 10% of the surge flow is reasonable. 4.2.1 Single-stage Compressor with Constant Inlet Conditions (a) Fixed speed machine

The compressor characteristic is regarded as 'flat' when the fall in pressure ratio is less than 10% as the flow increases from the surge flow to the antisurge protection point.

The antisurge system for a compressor with a flat characteristic should be based on a flow measurement technique. A simple discharge meter will provide good antisurge measurement (Fig 4.3).

If the characteristic is steep, then a system based on pressure measurement should be used (Fig 4.4).



(b) Variable Speed Machine

For a single stage only, the surge line is of approximately parabolic form. This follows from the relationship given in Clause 2.3.5.

The surge line is given by the approximate relationship:

Q s 2 = k1 x H ps 4.1

Where Q s is the actual volume flow at surge,

measured at inlet conditions m3/ s H ps is the polytrophic head at surge kJ/kg k1 is a constant

The head is related to the pressure ratio by the equation, 2.7, which can be approximated by:

H p = T I (A x r + B) 4.2 M Where H p is a polytrophic head kJ/kg T I is inlet absolute temperature K M is molecular weight r is the pressure ratio

A,B are constants The differential pressure, h, across an orifice plate passing the actual volume flow, Q, is given by:



where

P is the absolute upstream pressure at the orifice plate bar abs

T is the absolute upstream temperature at

the orifice plate K

M is the molecular weight of the gas By combining equations 4.1 and 4.2 for the case when Hp = H ps then:

This is the equation relating surge flow to surge pressure ratio. Provided the volume throughput is greater than the surge flow by a suitable safety margin, then the compressor cannot surge, and:

where Qc is the actual compressor volume throughput at inlet conditions If an inlet flowmeter is used for antisurge control and measures the compressor throughput, then substituting for Q from equation 4.3, gives:



The pressure ratio, r, is Pd / Pi where Pd is the discharge pressure (bar abs), so that equation 4.6 becomes:

This relationship is the basis of the Filippini antisurge control system, [5]. The key features are that the parabolic orifice plate characteristic is matched to that of the compressor surge line, and that the temperature and molecular weight cancel out of the relationship. For a compressor with constant inlet pressure, the equation (after combining constants) is:-

The system is shown in Fig 4.5



4.2.1 Single-stage Compressor with Variable Inlet Conditions

(a) Fixed speed machine

The characteristic for a fixed speed compressor may be redrawn as a plot of discharge pressure against discharge volume flow for values of inlet pressure within the range over which the compressor will operate.

The procedure is to calculate the pressure ratio at a number of points (4 or 5) on the curve (including the surge point) using the equation 2.7 re-arranged as:-

Then convert the inlet flows, corresponding to the selected points on the curve, to discharge volume flows using the relation:-

It can be seen from equation 4.10 and also from Fig 2.1 that the actual volume discharge surge flow is constant and is independent of inlet or discharge pressure for any particular gas density. The appropriate antisurge controller is, therefore, an orifice meter installed in the discharge pipe and compensated to measure actual volume flow. If Qs is the actual discharge volume flow at which the antisurge system operates, then from equation 4.3:



If Td and M are constant, then h/Pd is also constant. Now h/Pd is the compensated signal to the antisurge controller and the compressor will not enter the surge region providing this signal exceeds the value given from equation 4.11. This antisurge system works for a range of gas densities as a change in M has a similar effect on both the compressor and orifice meter. Fig 4.6 shows an instrument configuration which generates a measured value signal proportional to h/Pd which is proportional to the square of actual discharge volume flow. It is desirable, whenever possible, to avoid the use of analogue devices performing division as they are a notorious source of inaccuracy. A simpler system, which achieves the same end but avoids the use of the dividing relay, is shown in Fig 4.7. This carries out the function:

(b) Variable speed machine When a compressor operates with both variable speed and variable inlet pressure, then the full Filippini system is most effective. The control algorithm is that shown in equation 4.7. A worked example of the design of such an antisurge control system is presented in Appendix B.

(c) Variable Molecular Weight Gases Some compressors handle a gas mixture which is subject to sudden changes in composition and hence molecular weight. Bypass regulation is not suitable for this case unless the process system has a high impedance.



Where the compressor works between large system capacities and the process impedance is low, then at the time when the effective molecular weight of the gas falls, the compressor cannot maintain the working compression ratio and the working point will move rapidly towards the surge line. THE CONVENTIONAL ANTISURGE SYSTEM IS INEFFECTIVE. A secondary control is required, initiated by some indication of an impending change of gas composition, and which immediately operates directly on the normal throughput regulation of the compressor.

4.2.3 Features of High Compression Ratio Stages Strictly, the measurement systems described in the previous clauses are applicable only to single stage compressors which operate at a low Mach Number. High compression ratios are obtained from a stage operating at a high Mach Number or from a multistage machine. The surge lines of such compressors may have discontinuities (see Fig 2.3) or typically take an 'S' form. This means that the simple methods of generating the surge protection line which give a parabolic shape may not be satisfactory [6]. It is usual to insert a characterizing relay to modify the signal derived from inlet and discharge pressures. The modified signal is then fed to the set-point of the antisurge flow controller. The locus followed by the set-point then matches the surge line with an appropriate margin. Another implication of a high pressure ratio is that the main bypass will most generally be operating in the critical flow regime with choked flow. Provision of an intermediate bypass allows optimum sizing of the main valve but with more elaborate antisurge arrangements to ensure that the auxiliary valve is operated under the correct circumstances (see Fig 4.8).



4.2.4 Multistage Systems Such compressors may have ancillary control systems which regulate the process gas temperature at each stage inlet by adjusting the water flow through intercoolers, so that operation below dewpoint is avoided. It is then necessary to include the surge lines for the extreme conditions of intake gas temperature, pressure and humidity on the graph of the process/compressor characteristics. The envelope of these surge lines may itself be taken as a limit curve and used to define the desired surge protection line. Alternatively, the set point of the antisurge controller might be automatically reset by a signal from the dewpoint control system. Where the cooling duty is performed by unregulated intercoolers not integral with the machine casing, it is customary to provide an antisurge system for each stage. Such multiple antisurge systems may be arranged in a number of ways which exhibit different interaction effects (see Fig 4.9). It is not necessary to install a flowmeter for each stage; the volume flow for stages other than the first can be computed using interstage pressure and temperature measurements. Some compressors have casings with sidestream connections in addition to the inlet and discharge connections. The important point is that the incoming sidestream mixes with the discharge from the previous compression stage before admission to the next compression stage. The antisurge system then necessarily includes multiple bypass loops which interact.

4.3 Surge Detection and Recovery

Damage due to surging is cumulative in character; consequently the antisurge system should be activated quickly if surge occurs inadvertently. After the bypass valve has opened a sufficient amount to ensure recovery from surge, a controlled return to a satisfactory operating condition should be accomplished by the anti-surge system. If recovery proves impossible the compressor is tripped, usually after the pulse detector has counted a given number of surge pulses.



4.3.1 Pressure Pulse Detector A detector is normally installed which senses the discharge pressure and by differentiation generates a pulse output from the surge wave front (see Fig 3.3). Such a detector directly initiates opening of the bypass valve.

4.3.2 Inlet Temperature Detector During continued surge the inlet temperature increases. The rise could be used in principle to detect surging but the quick response and sensitivity required of the temperature measuring element make for a delicate construction unacceptable for normal plant use.

4.3.3 Torque Fluctuation Detection

Reverse flow in deep surge causes a sharp change in the torque transmission from driver to the compressor. In principle, this can indicate surge but adequate instrumentation for this purpose is not commercially available.

4.3.4 Detection of Incipient Surge Some compressors give warning of surge through an increase in gas pulsation level (see Fig 3.1). If this could be used as the basis for continuous self-adjustment of the antisurge system then the maximum possible range of throughput regulation would be usable. However, the problems of this approach have not been solved.



5 DYNAMIC CONSIDERATIONS

The response to changes which can be regarded as quasi-steady state represents one aspect of the control system performance. The other is its response to rapid changes, typically arising from trips. All physical systems contain elements which store material and energy so that their states cannot be changed instantaneously, giving rise to the concept of response time. This is the time which elapses between a change being impressed at one point and its effect being recognized at another. The behavior resulting from a disturbance which causes change within a time comparable to the combined response times in a system, will differ markedly from that obtained by considering a succession of intermediate steady states covering the same variation. Excess control loop gain will cause the system to become unstable and hunt. However, the rate of correcting a disturbance is almost directly proportional to the controller gain. All controller settings are, therefore, a compromise between damping the initial transient and rapid correction of the disturbance.

5.1 Interaction

So far, the treatment of throughput control and antisurge protection has implicitly assumed that the functions are separate; this is true of installations where the antisurge protection is an emergency measure called upon to act only on trip when the process is shut-down anyway. Mostly it is necessary to continue supplying the plant by using the antisurge control at the same time as the throughput regulation. Then, because they are both operating through a common impedance (ie the machine) any action by one control loop has a contradictory effect on the other. If the response times of the two loops are similar, their combined action is very likely to be unstable. This dilemma is difficult to resolve. The Simplest expedient is to detune by reducing the gain of the throughput controller or by altering the response times to give a ratio of at least 5:1. This degrades the throughput regulation performance. The margin between the antisurge system activation line and the true surge 11ne may also be increased. Carried to the limit, this implies using bypass or blow-off control alone (as mentioned in Clause 2.3.2).



A sophisticated approach is to cross-couple the signals in the control loops to compensate for the process interaction using multi-variable design techniques. The expense of this approach has not so far been justified for process plant. A practical expedient to avoid simultaneous operation of the two loops is illustrated in Fig 5.1 and operates as follows. The output P of the pressure controller is compared with the antisurge limit signal A in a selector relay S. This is a high selector and its output F forms the set-point of the inlet flow controller which modulates the speed, inlet guide vanes or inlet throttle of the machine. At the same time, this output is impressed, together with that of the pressure controller, on the summing relay R1. Its output, D, (see Note in Fig 4.5) is then given by:

Where B is the bias signal representing the fully shut position of blow-off (or bypass) valve. Because of the selector action, as long as P > A then:

F = P

And therefore D = B Then the pressure controller manipulates the set .point of the flow controller whilst the antisurge valve remains shut. If P falls below A, however:

The inlet flow then remains fixed at the anti-surge limit and the antisurge valve is used for regulation. Thus the pressure controller output is diverted between the two separate control actions depending on its value relative to the surge protection limit, giving a variable split-range facility.



The value of µ1 is adjusted to match the control sensitivities in the two modes. If further flexibility is required, a second relay, R 2 , can be added using the alternative connections shown. Adoption of this scheme for flow control requires the addition of a pressure override to ensure that the antisurge valve will open if the compressor discharge is blocked.

5.2 Speed of Response of Antisurge Control System

(a) The Antisurge Control Valve The slowest element in the antisurge control system is usually the control valve, when only the control valve dynamics need be related quantitatively to the process disturbances. The margin between the surge line and the antisurge activation line allows time for the antisurge valve to begin opening as the process disturbance develops. The fractional opening of the antisurge valve required to keep the compressor above surge is then calculated and the required valve speed obtained. The example in Appendix B illustrates how the permissible times may be estimated. This example is not a true dynamic calculation, as it makes no allowance for the effects of pipework volume which reduce the rates of change of pressure. Such an approach will result in a degree of over-design. A proper simulation model is needed to take full account of both valve stroke speeds and process dynamics. It is important to note that remote positioning of compressor discharge valves in long process lines so as to increase the volume between the valve and compressor may prove an economical way of avoiding the use of expensive fast-stroke antisurge valves.



(b) Controllers A surge protection margin of lOt implies that a proportional band less than 100% would be required for the valve to be opened fully by proportional action alone. Therefore a (P + I) controller with fast integral action is used. Derivative action is unsuitable because of the number of significant response times in the control loop and its amplification of measurement noise. A disadvantage of introducing integral action into a controller which is not in continuous use is integral "windup" or saturation, ie the integral term contribution to the controller output goes outside the normal working range, 4-20 mA, or 0.2 - 1.0 bar, because of the sustained difference between measured value and set point signals. The controller then takes a finite time to recover and to start to move the valve when the error signal reverses sign. This delay is unacceptable in an antis urge system which should therefore incorporate desaturation. This can take two forms, described as output limiting and integral limiting. Output limiting produces pseudo-derivative effects which cause the valve to start opening before the measured value crosses the set-point. This accelerated action is desirable but it can be troublesome in noisy systems. Fluctuations in the controller output signal then cause a valve plug to bounce on its seat, giving extra wear in addition to the danger of interfering with the throughput control. Integral limiting merely prevents integral "wind-up" and there is no pseudo-derivative effect. Whilst it is less rapid in action than output limiting, it is more satisfactory on noisy applications. Note that for disturbances which are slow enough for the balancing action of the output limiter to follow the changes. these systems are equivalent. (c) Non-return Valve The dynamic response of the system is significantly affected by the non-return (or check) valve which is located in the compressor discharge line downstream of the antisurge system bypass branch (see Fig 3.2).



The valve duty arises principally from the results of a failure or trip of the compressor driver, when:- (1) It protects downstream plant against excessively rapid

depressurization.

(2) It protects upstream plant against over-pressure. It may not be possible to provide relief valves of sufficient size (at least 300% of rated forward flow). consequently a high integrity non-return valve is required.

(3) It prevents a high reverse speed of rotation due to the compressor

acting as a turbine when operating with reversed flow. For this purpose the valve closure should be fast (typically less than O.S s) and the valve must be located close to the compressor discharge to minimize the energy store represented by the amount of gas free to flow back through the compressor.

For multi-casing compressors having large intercoolers or other large capacity interstage vessels, auxiliary non-return valves are often provided between casings to limit the reverse speed.

For compressors with sidestreams. it is necessary to provide a non-return valve on each pass-out sidestream. On some major machines non-return valves have also been provided on each pass-in sidestream including the inlet, recognizing that this furnishes a degree of redundancy. For circulators or boosters, where the pressure rise across the machine is a small fraction of the inlet pressure there is no single model solution. The principal difficulty is that boosters will also function as turbines (albeit at very low efficiency) in forward rotation and with a large forward flow (see Fig 5.2). On some machines the design of bearings or seals may make reverse rotation unacceptable. It is important to note that provision of the conventional discharge non-return valve and antisurge system does not guarantee that reverse rotation will never occur. Reverse rotation locks or brakes should then be considered.



(4) It enables the antisurge system to act by isolating the compressor from the user plant when the plant pressure exceeds that developed by the compressor. This action is not restricted to a driver trip but may occur during a momentary speed dip or during surge recovery when the antisurge valve momentarily goes to the full-open position.

6 SYSTEM EQUIPMENT SPECIFICATIONS

6.1 The Antisurge Control Valve

The antisurge valve often remains in its closed position for long periods, but should open smoothly and quickly when required.

A small permanent flow through the bypass is often required (Clause 6.1.4) which implies that a tight shut-off is not essential. However, a tight shut-off is still desirable to avoid deterioration of the plug and seat. Soft-seated valves are not recommended; temperature limitations preclude their use in many installations and long periods in the closed position can cause distortion of the seat.

6.1.1 Valve Size and Characteristic

The minimum valve size is dictated by the necessity to pass the compressor rated flow at the stipulated discharge and inlet pressures.

It is equally important not to oversize the valve because the possibility of machine damage arises. In the case of motor-driven multistage high-compression ratio compressors, the design geometry of each stage is determined by the duty. Consequently, as the discharge pressure is reduced, the mass flow is eventually independent of the discharge pressure, resulting in a near-vertical compressor characteristic. Operation in this region disturbs gas pressure thrust balance arrangements, so that the machine bearing loads are large and variable. For economic reasons, valves are manufactured in a finite range of sizes. The most suitable valve could be over-sized. The maximum flow is best limited by using a restriction orifice-plate. Alternatively, the valve stroke could be restricted, either by a mechanical stop or by adjustment of the valve positioner, but these are open to misuse in service and are not recommended.



As illustrated in Fig 4.8, a single valve could have a capacity much greater than the design flow at normal discharge pressure. A supplementary bypass valve activated by auxiliary measurements should then be considered. The characteristic of an antisurge valve should be linear. This gives a greater fractional opening in the early part of its travel than an equal-percentage trim and also maintains a more constant loop gain over the range of controlled flow because of the generally low resistance of the bypass line. 6.1.2 Actuators The stroking time required of the antisurge valve is typically within the range 0.5.to 5 s. In some cases this is achieved by sing hydraulic actuation, but pneumatic operation is preferred even though it entails the use of a high-capacity positioner and booster relays. Pneumatic valve actuators equipped with positioners will saturate when held at either end of the stem travel. In this case, with the valve fully closed, there will be some delay in starting to open. For fast control loops where the process response is comparable with that of the control equipment, stability is degraded by the use of a positioner [11]. It is therefore possible that the performance of an antisurge control loop could benefit on both counts from the use of a boosted diaphragm actuator without a positioner. Nevertheless, it has been customary to specify positioners for this duty on the ground of high pressure drop across the valve. For the time being, it is considered advisable to continue this practice, with the option of bypassing the positioner on site if its use leads to intractable stability problems. This option does not apply to double-acting piston actuators. 6.1.3 Action on Instrument Air Supply Failure The purpose of the antisurge bypass valve is to protect the machine therefore it opens on air failure. If this action introduces other process hazards, then additional protective systems will be needed. A Single-seated unbalanced plug valve should be installed so that the machine discharge pressure will assist its opening.



6.1.4 Condensation in the Process Gas Absence of flow through the antisurge valve and its pipework provides an opportunity for vapor condensation to occur. The accumulation of liquid in the bypass line constitutes a major hazard to the compressor. Firstly, the inertia and viscosity of the liquid column seriously delays the establishment of an adequate antisurge gas flow; secondly, there is the possibility of mechanical damage arising from liquid sent into the compressor inlet. Accordingly, the bypass pipework both upstream and downstream of the valve should always be designed to drain naturally into the mainstream piping. Residual condensation should then be eliminated, either by steam or electrical heating or by arranging a small permanent flow of gas (about 1% of the process flow) through the bypass pipework. 6.2 Non-return Valve Normally the non-return valve is self-actuated to obtain the required speed of action without sensing problems. Typical constructions are the spring loaded multiple plate or poppet valve (similar to the valves used on reciprocating compressors), the single spear valve, and the hinged flap valve. The hinged flap valve is the least expensive construction but requires both a weight to balance the flap and a hydraulic damping device. If the damper is a simple restrictor in the line connecting the two ends of a hydraulic cylinder, it may make the operation of the valve too slow. More sophisticated dampers acting only at the extremes of the flap displacement may then be needed. Valves with dampers of this type are usually called 'nonslam' valves. Occasionally, for process reasons continued reverse leakage flow cannot be permitted. Self-actuated flap valves should then be supplemented (but not replaced) by tight shut-off isolation valves having power actuators. These externally actuated block valves should be located well away from the compressor because cases of spontaneous closing are known and if they are near the compressor discharge the surge resulting from such accidental closure is likely to have a high intensity and relatively high pulse rate.



The provision of a valve stem-travel trip is recommended, directly opening the antisurge bypass valve. For compressor service where deposits of dirt, polymers or corrosion products form, the self-actuated non-return valve may be unreliable. Consequently, either an externally actuated valve should be used or the system designed to accommodate the effects of omitting the valve.

6.3 Pressure and Flow Measurement

6.3. 1 Flow No instrument currently available measures the actual volume flow or true gas velocity. In practice, differential pressure type flowmeters are used with either an orifice plate or a Dall tube as the primary element and the velocity is inferred. In antisurge systems the criterion of flowmeter performance is repeatability, not absolute accuracy, but the installation requirements for achieving this are not well defined; designers therefore depend largely on experience gained from the operation of comparable systems. On simple compressors the favored position of the flowmeter is at machine discharge. However, compressors shed considerable vorticity and turbulence so that flowmeter types dependent on induced vorticity (Dall tube) are unsuitable for this position. For compressors with complex surge lines the best flowmeter position is at the inlet but here the lines have a comparatively large diameter so that problems of layout arise. Orifice-plate installation standards are given in BS 1042. It is usual to insert the orifice-plate in a length of pipe which is shorter than the ideal. The signal may then be subject to error but it will be repeatable provided the flow pattern at the orifice does not change with time. Vaned bends are satisfactory flow straighteners for this purpose. Dall tubes have performed satisfactorily between vaned bends 50 apart (made up of 3D upstream pipe length, 10 downstream and 10 for the Dall tube itself). Dall tubes are recommended because they have high pressure recovery. The Venturi meter is preferred for installations where deposition may be a problem.



Pitot tubes are not recommended because of their susceptibility to noise and flow maldistribution. Bend flowmeters consist of differential pressure impulses taken between the inner and outer radii of a pipe bend. Such flowmeters have no insertion loss but give a low differential pressure compared with an orifice meter and the impulse tappings are prone to blockage by solid particles. A major disadvantage is their insensitivity to flow direction. The inlet pipe connection to the casing of a gas filter can include a nozzle section across which the differential pressure is measured. There is no insertion loss but the position of the nozzle and the symmetry of the casing are crucial to repeatability.

Flowmeter impulse lines should be self-draining into the compressor piping. and of equal length so that pressure fluctuations in the pipe are applied simultaneously to each side of the transmitter diaphragm. The following meters are attractive in principle because they respond linearly to gas velocity; however, they are as yet untried.

(a) Vortex shedding meters count the frequency of vortices by a "bluff body" obstruction in the pipe. They are sensitive to random noise disturbances in the flow pattern and are likely to generate a false high-flow reading.

(b) Devices based on laser or ultrasonic beams are currently used for research purposes but are not yet suitable for plant applications.

6.3.2 Pressure

Gas density. and hence compressor and flowmeter performance, depends on absolute pressures. Atmospheric pressure varies by 0.1 bar. It is worth calculating the error produced if a gauge transmitter is used instead of an absolute transmitter in an antisurge control system. Measured pressures around the circumference of a large pipe close to the compressor discharge may differ by as much as 0.1 bar, consequently pressure transmitters should be installed in straight pipe runs well clear of the compressor discharge connection.



Impulse lines should be self-draining into the compressor pipework. 6.3.3 Surge Pulse Detectors

These function by differentiating a signal transmitted from the discharge pressure measurement. The pressure transmitter should not distort the surge waveform to an extent which would prejudice the generation of an unambiguous pulse output. Taking a factor of safety of 2, this means that. for a critical detection rate of R barfs. a pressure transmitter of span P bar should be capable of traversing its full output range in a time less than:

Where t = 5 s or less, avoid excessive volume loading on a pneumatic transmitter (a maximum of 1 litre is suggested).

With the span fixed, the time constant and the gain of the detector is adjusted to give the required output. Simple differentiators (Fig 6.1, a and b) have a time constant which attenuates the output amplitude. Increasing the time constant is therefore subject to diminishing returns; values less than 100 s are recommended. A sensitive differentiating unit is inherently susceptible to 'noise' and some form of protection is required. Noise which is High frequency compared to 6 x the surge cycle frequency should be filtered out. The differentiator should be biased to respond only to a fall in pressure. The rate of change in this direction is much higher than for the pressure recovery and this, together with suppression of the reverse pulse, leads to a more sharply defined output.



6.3.4 Gas Analysis Measurement Both the compressor surge line and the readings of differential pressure flowmeters are affected in roughly the same way by changes in the molecular weight of the gas, so that antisurge systems are partly self -compensating., Most gas analyzers require sampling systems which cause their total dynamic response to be too slow for re-setting an antisurge control loop. The oscillating spool type of meter (eg Joram Agar) might be suitable for non-condensing gases but has not yet been tried.

6.4 Signal Transmission

Because there are delays associated with the transducers for signal conversion and primary measurement, the use of conventional electronic equipment does not always produce the fastest system. As a rough guide, these delays are equivalent to about 20 m of pneumatic transmission line; therefore on compact installations pneumatic systems are acceptable. When an electronic system is used the I/P (current to pressure) converters should be sited close to the valve to minimize the length of pneumatic line, subject to the mounting being vibration-free (an important consideration at machine locations).

6.5 Controllers

Antisurge controllers are always embodied in fast response loops whose natural period is closely related to the stroking time of the antisurge valve. The integral action time will then be of the same order so that fast reset is needed, ie a minimum setting of less than 1 s. Integral desaturation is always required. The behavior of the whole loop when subjected to high amplitude fluctuations (> 10% FS, > 5 Hz), which may occur at the point of incipient surge, is unpredictable. A filter to attenuate the high frequency disturbance may be needed.



7 TESTING

The commissioning of the machine should be planned to include checks on safe operation. Consequently, the antisurge system is established first to ensure protection of the machine and the remaining control loops then tuned to avoid undue interaction.

7.1 Determination of the Surge Line

This is done experimentally because makers predictions for a new machine are not reliable. In the case of an existing installation normal wear and fouling will modify the surge characteristic and the original performance may not be fully restored after cleaning and re-assembly. The test points are plotted on the machine characteristic to obtain the true surge line. The surge protection line generated by the antisurge system is then compared with the true surge line and adjusted to give the required margin of safety. The margin required may be surprisingly large. In addition to the points raised in Clause 4.2, the optimistic effect of test conditions should be taken into account; slow changes allow the compressor to maintain temperature/pressure equilibrium, which may delay the onset of surge in contrast with the rapid disturbance occurring in service. Compressors with intercoolers have performance and surge characteristics markedly affected by relative changes between process gas and cooling water temperatures. Verification of surge lines should be carried out for each individual stage.

7 .2 Records

The actual values of the settings established during commissioning should be recorded in the machine manual. The record is important as a means of checking drifts in controller operation and also for ensuring that the performance can be reproduced if equipment has to be replaced.



There is a distinction between the documentation requirements of throughput and antisurge systems. The former are effectively normal process controls and their performance will be continually monitored. The latter are classed with alarms and trips. They should be included in the list for periodic inspection and testing.



8 INLET GUIDE VANE UNITS

8.1 Application

The IGV unit produces a family of compressor Q-H characteristics, each one corresponding to an angular setting of the vanes. As the angular setting increases the compressor characteristic becomes steeper. Thus guide vane regulation is well suited to compressor duties requiring roughly constant head: the compressor can run at constant speed and is able to operate over a substantial flow range. Bypass or discharge throttle regulation can be advantageously used to obtain a rapid response, with an IGV unit slowly adjusting the compressor characteristic for more efficient operation. Similarly a two-speed electric motor driver has been used in conjunction with an IGV unit in order to extend the operating range at high efficiency.

8.2 Effect on Power Consumption of the Compressor Over the range from -5o to +20o the IGV unit does not affect the compressor efficiency but the corresponding regulation range is also small, commonly 10%. of the rated flow. As the flow is reduced energy dissipation increases and the IGV unit approaches the performance of an inlet throttle valve. Tests on the process air compressor at Reysham Works AGD, which had both IGV and an inlet butterfly valve, showed that the IGV unit maintained a saving of 5% in power absorbed over the whole capacity range when the compressor discharge pressure was constant.

8.3 Effect of Gas Conditions, Properties and Contaminants

Stages where intake gas is clean and continuously above dewpoint but below 90oC are suitable for IGV units. Hotter gas is acceptable provided that any grease lubricated bearings are cooled by a separate cold gas supply, and that the connective linkages allow for differential thermal expansion.



Gas containing condensate droplets is NOT suitable for IGV units because such droplets are commonly nucleated by solid particles which are left by evaporation as a deposit on the vanes and their bearings and linkages. Current practice is to forbid the use of IGV units for stages following an intercooler which may operate intermittently below local dewpoint. IGV's for stages where the inlet pressure is high should be provided with a single seal to atmosphere, preferably by a rotary radial face seal rather than by a linear rod gland. Hydrocarbon gases may lay down polymer films on compressor surfaces. These films impose severe loads on vane actuation linkages; consequently current practice is to forbid the use of IGV's and specify regulation by variable speed or by inlet throttle valve. Ammonia/methanol synthesis gas is free from particulate contamination and IGV units have been used successfully provided that pipeline cleaning during construction and commissioning periods has been to a sufficiently high standard. Wet CO2 normally contains oxygen and corrodes steel surfaces. However, it is often economic to use carbon steel pipelines with a large corrosion allowance. IGV's are NOT recommended for such service.

8.4 Aerodynamic Considerations

8.4.1 Modes of Function

The theoretical head rise gH developed by an impeller is given by the Euler Equation:



Where U is the impeller tangential velocity, Vu is the swirl velocity, and subscripts 1 and 2 refer to inlet and outlet conditions, respectively. The IGV unit, when set within the range O o to +20o, alters the Vul term so that the head is reduced. No energy is dissipated consequently the compression efficiency remains unchanged. Above an angle of +20° losses become significant. The incoming gas is decelerated inefficiently in the turbulent region downstream from the vanes. At angular vane settings exceeding +50 o the flow passage area decreases and there is a large pressure drop across the vane. The regulation approaches that of inlet throttling. Provided that the machine compression ratio is sufficiently high the gas velocity in the passage between the vanes will reach sonic velocity with a vane angle of 70 - 75°. Current practice forbids continuous operation at or beyond this point because shockwave formation introduces unpredictable forces on the inlet guide vanes and interaction with impeller vanes may excite natural frequencies of vibration. Automatic control systems should limit the vane angle to 3° less than this critical setting. The maximum flow occurs at a negative angle of about 25°. Current practice is to limit the angle to -15° to avoid a significant fall in power efficiency.

8.4.2 Vane Rotation Torque Generally IGV's have symmetrical profiles. At small deflections the location of the aerodynamic pressure centre is independent of the pitch angle and located approximately at the quarter chord point, ie 25% of the chord length from the leading edge. Most IGV units are designed with the pivot point behind the aerodynamic centre with the resultant moment tending to close the vanes.



8.4.3 Radial Flow Vane Arrangements Radial flow vane arrangements induce cylindrical whirl and depend upon the proper shaping of the inlet passage to maintain an even velocity distribution at impeller entry. The advantage of this configuration is that high values of whirl can be obtained (through conservation of angular momentum) with practical sizes of vane. It is the preferred arrangement for small volume flows ( < 2 m3/s). 8.4.4 Excitation of Compressor Impeller Vibrations Shockwave formation may occur near the closed vane position. Vane wake interaction may occur when an axial-flow IGV unit is positioned too close to the impeller entry, exciting 'flap' vibration modes of the vanes of an open impeller. Check the number of vanes in the IGV unit and the number of impeller vanes; they should not have a common factor. 8.4.5 Twisted Vanes Axial flow guide vanes have been twisted to provide a free vortex velocity distribution at the impeller entry for the vane angular setting corresponding to the most likely flow. No significant advantages have been reported for this configuration. 8.5 Control System Linearity It is essential that the control characteristic is reasonably linear otherwise the throughput controller cannot be tuned to give system stability. Consider Fig 8.1 showing a non-linear characteristic relating Flow to Actuator Position. If the system is tuned at low flow rate condition (point A on curve) the gain will have to be low. The corresponding integral action time should be reduced considerably otherwise the system will permit large deviations from the set point. On the other hand if the system is tuned at high flowrate condition (point B on curve), a high gain is desirable with a long integral action time.



These requirements are irreconcilable and current practice is to design for a reasonably linear characteristic over the desired operating range, keeping the gradient of the characteristic within a band of 2:1.

8.6 Actuator Specification

Recent practice is to specify actuators of the geared electric servomotor type. These have the advantages of: (a) irreversibility, with maintenance of the last position upon power

failure (b) constant velocity operation

(c) simple integration with electronic controllers

(d) press-button inching from the control panel

(e) effective limit stops

Past installations have successfully used pneumatic and hydraulic actuators. The latter are now considered to be a hazard when enclosed by the machine acoustic hood unless fire-resistant hydraulic fluids are used.

Note that pneumatic actuators cannot be self-locking, at best they drift to an end position.

The stroking speed is normally lilllited to avoid interaction with the antisurge system. There is also an aerodynamic limit. Near the open (0o) position the vanes can be regarded as an aerofoil cascade. The steady-state condition is reached upon an incidence change only after a time corresponding to the flow moving about 5 chords downstream. For typical machine dimensions this limits the angular velocity of the vanes to 0.5 rad/sec. Linkages designed to linearize the regulation have a mechanical advantage of 10 - 20. Thus the average stroking time should be based on a limit of 0.05 to 0.03 rad/sec giving minimum stroking time between 20 and SO seconds.

Some installations require only a manual hand wheel to allow occasional adjustment of the compressor. Specify such hand wheels to operate with at least 80 turns for full stroke to avoid excessive sensitivity in operation.



8.7 Starting Considerations

For large electric motor drives it is usually essential to reduce the power absorbed during starting. The IGV control is manually overridden to bring the unit nearly to the full close position (~ 80o). This setting is the first limit stop; it is based on obtaining the rated compression ratio at the rated volume flow, taken for the actual conditions at the compressor inlet and discharge.

It is ESSENTIAL to direct the attention of operators to the avoidance of running under manual control in the band between the IGV start position and the minimum flow position. In some cases a time-delayed alarm signal may be necessary. Because the motor rating is based on the ICV start position a microswitch senSor for this position should be provided, with the signal inserted into the permissive start relay sequence for the motor.

8.8 Features of Link Mechanism 8.8.1 Synchronism and Backlash

Systems where each vane is connected by its gear to a single gear driver have given considerable difficulties in controlling backlash. Systems using space links are preferred. Backlash due to clearances in bearings can be compensated by judiciously positioned spring loading. However, the preferred method is to specify preloaded needle roller bearings because these both eliminate backlash and reduce deadband due to friction. Manufacturers are reluctant to offer such bearings and usually have to be pressed to adopt them. Not every bearing can be of this type so that some spring loading will be necessary. Note that the vane moment reverses in the vicinity of the Oo position.

Pseudo-backlash has been encountered because of the deflection of links under load. Lang links in compression should be checked for buckling and may require special inspection during manufacture.



8.8.2 Linkage Characterization

The data needed are:

(a) Manufacturer's graphs of the compressor Q-H characteristics at various values of ICV angle. These graphs should be repeated for different inlet gas conditions.

(b) Process system characteristic. This may be a family of

characteristics.

The intersection paints of the machine and process characteristics give the relation between ICV angle and flow. This relationship will nat usually be unique but will appear as a band on the graph of angle against flaw. In developing the desired relation it is useful to remember that the regulation covers: (a) Deliberate changes in rate, during which the machine inlet

conditions remain constant.

(b) Maintenance of the given rate against external changes in machine conditions, e.g. changes in ambient conditions for air compressors.

Actuators are either the linear thrustor or the rotator type. The relation of this actuator characteristic to the ICV angle characteristic for linear change in flaw is the desired linkage inverse characteristic. Note that the actuator stroke should be sufficient to caver the IGV closed position for start-up. The desired linkage characteristic is then approximated by a synthesis of 4-bar chains. More complex mechanisms using sliders or gears are excluded on grounds of difficulty in meeting the backlash and deadband requirements.



8.8.3 Bearings and Lubrication Despite their very low operating speeds IGV mechanisms should not generate 'stick/slip' motion. This requirement leads to the following considerations:: (a) Rolling element bearings

The preferred type is the sealed preloaded needle roller unit. Other types can be used but with their ratings substantially reduced to avoid 'Brinelling'. Preloading is unnecessary when link backlash spring loading is used.

(b) Ball and Socket Joints

These are used for space linkages but preferably only one such linkage in the sequence should be included.

Specify Alvania EP2 grease for such joints because the additive dispersion greatly reduces the difference between static and dynamic friction forces and increases the load carrying capacity. Manufacturers tend to ignore specification of such special grease, consequently the inspecting engineer should be briefed to verify that it has been used.

(c) Dry Self-lubricated Bearings

Glacier DU bearings have been successfully used. Such bearings should be sized for the maximum actuator loading; such sizing must be included in the technical audit. Provision of link pre-loading springs is then essential.

8.9 Limit Stops and Shear Links

Limit stops are always required, either on the actuator, or on any hand wheel.

Should these fail or be inadvertently overridden, mechanical stops are required at some appropriate point in the linkage system.

As a final safety precaution, a shearlink should be provided as the weakest point in the whole system, positioned at an accessible point outside the machine casing.



DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE This Engineering Design Guide makes reference to the following documents: AMERICAN PETROLEUM INSTITUTE API 610 Centrifugal Pumps for General Refinery Services (referred to in Clauses 2.6 and 8.1). AMERICAN SOCIETY FOR TESTING AND MATERIALS ASTM D86 Distillation of Petroleum Products (referred to in Clause 2.4.1) ASTM D2892 Distillation of Crude Petroleum (referred to in text with Table 2A. Page 10). ENGINEERING DESIGN GUIDE GBHE-EDG-MAC-1014 Integration of Special Purpose Centrifugal Pumps

into a Process (referred to in Clause 4). GBHE-EDG-MAC-1117 Special Purpose Centrifugal Pumps (referred to in

Clause 8.1).

Centrifugal Compressors

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