Centrifugal Compressors

GBH Enterprises, Ltd.

Engineering Design Guide:
GBHE-EDG-MAC-1134

Centrifugal Compressors
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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
2.2

Characteristic of a Multiple Stage Having More
Than One Impeller

2.3

3

Characteristics of a Single Compressor Stage

Use of Compressor Characteristics in Throughput
Regulation Schemes

MECHANISM AND EFFECTS OF SURGE
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

3

Recovery from Surge


4

CONTROL SCHEMES INCLUDING SURGE PROTECTION
4.1
4.2

Surge Protection

4.3

5

Output Control

Surge Detection and Recovery

DYNAMIC CONSIDERATIONS
5.1

6

Interaction

5.2

Speed of Response of Antisurge Control System

SYSTEM EQUIPMENT SPECIFICATIONS
6.1

Non-return Valve

6.3

Pressure and flow measurement

6.4

Signal transmission

6.5

6

The Antisurge Control Valve

6.2

7

5

Controllers

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
Pi

Pd

=

absolute total pressure at discharge

bar abs

Pi

=

absolute total pressure at inlet

bar abs

M =

……… (2.1)

w x Ti½
Pi

……… (2.2)

W=
Ti =

mass flow
absolute total temperature at inlet

U=

N

Ti

kg/s
k

……… (2.3)
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
a

, where

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
n-1
M

r n-1/n

-1

..…… (2.5)

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
RoTi
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

ᶯP

defined by:

= n (k – 1)
K(n – 1)

…….. (2.6)


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 pseudopolytrophic 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
4.1.1

Output Control
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
Where

Q s is the actual volume flow at surge,

4.1

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

Hp

is a polytrophic head

kJ/kg

TI

is inlet absolute temperature

K

M

is molecular weight

r

is the pressure ratio

A,B

Where

are constants

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


where
P

T

M

is the absolute upstream pressure at the
orifice plate

bar abs

is the absolute upstream temperature at
the orifice plate

K

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 rearranged 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 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 pseudoderivative 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 nonreturn (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 nonreturn 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)

6

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.

SYSTEM EQUIPMENT SPECIFICATIONS
6.1


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. Softseated 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 highcompression 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 equalpercentage 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 Singleseated 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 vibrationfree (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.


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