CONTROLLING - Gas Compression Magazine

EDITOR’S NOTEThis article is based on a paper pre-sented at the Gas Machinery Confer-ence held September 30 – October 3, 2018, in Kansas City, Missouri, USA.

INTRODUCTIONThere are two objectives for

compressor control: meeting the external process requirements and keeping the compressor within its operational boundaries. Typical control scenarios that have to be considered are process control, starting and stopping of units, and fast or emergency shutdowns.1-3

The control of centrifugal compressors must be considered from the perspective of the compressor and the process. Regarding the compressor, it is necessary to discuss the different control devices, such as variable speed, guide vanes, throttles, or recycle valves. It is also important whether a steady-state compressor map is still valid in the case of fast transients. Further, the different operating conditions of the compressor, such as surge, stall, and choke, have to be explained. The control system has to be addressed with regard to instrumentation and device requirements, as well as the control methods of the drivers.

For the process, one must under-

stand the relationship between the flow through the system and the pressures imposed on the compressor. These re-lationships are different depending on their rate of change; in other words, one must expect different system re-sponses for fast and slow changes as well as steady-state conditions.

Different upstream and midstream applications lead to different com-pression system characteristics and control requirements, which are the result of compressor requirements (such as high pressure ratio, wide operating range) and the process requirements. Multiple unit installa-tions, installations with multiple com-pressors per train, and installations in which the train has to serve multiple gas streams require specific process control considerations to match the

gascompressionmagazine.com | AUGUST 2019

CONTROLLING CENTRIFUGAL COMPRESSORS

BY RAINER KURZ AND KLAUS BRUN

ABSTRACTThe control requirements for centrifugal compressor packages are presented. Controlling machinery

must be understood within the interaction of the compressor and the compression system around it. The focus of this paper is on variable-speed, upstream, and midstream applications. The impact of the interaction between system characteristics and compressor characteristics, both under steady-state and transient conditions, is explained. Also considered are concepts to optimize and control the units. Special attention is given to the issue of surge avoidance. Additionally, the impact of the process and how the process dynamics interact with the compressor is analyzed, categorized, and explained.

Flow

High Flow

Low Flow

Best E�ciency Point

Isentropic Head

IdealHead

IncidenceLoss

Hea

d A

nd L

oss

FrictionLoss

Incidence Losses

(-)(+)

Figure 1. Head-Flow Characteristic Of A Compressor At Constant Speed. Operation At A Flow Higher Or Lower Than The Design Flow Cause Incidence Losses.

compressors with the process system behavior and the objectives of the sta-tion or system operator.

A discussion of control requires the goal of a control system to be defined. The requirements to protect the process as well as the equipment have priority. Within the range of allowable operating conditions, other goals need to be defined, too, particularly if the station involves multiple compression units, either in series or in parallel. Possible goals can be to minimize fuel consumption, to minimize maintenance cost, or to maximize throughput.

The framework described above also defines the structure of this study. The centrifugal compressor and its control will be addressed first, followed by a description of the process behavior under various scenarios. A prominent

role in these descriptions will be taken by surge avoid-ance considerations. The discussion will close with con-siderations on how to control multiple units.

The behavior of compressors during emergency shutdowns will not be discussed in this paper. The control system’s function is simply to initiate the shutdown and to open the recycle valve as fast as possible. Furthermore, the behavior during emergency shutdowns has been covered in great detail in a number of papers.4-8

COMPRESSOR BEHAVIOR AND COMPRESSOR CONTROLS

Centrifugal compressor behavior can be described by their head-flow-efficiency relationships. The basic rela-tionship for a compressor at constant speed is shown in Figure 1. The compressor shows a distinct relationship be-tween head and flow. In the case of machines with back-ward-bent impellers (the type generally used in upstream and midstream compression applications), the head of the compressor increases with reduced flow. Due to the in-crease in losses when the compressor is operated away

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from its design point, the curve eventually becomes horizontal and, subsequently, starts to drop again. The curve section with positive slope is usually not available for stable operation. When the flow is increased beyond the design flow, the losses also in-crease and increase the slope of the curve, sometimes to a ver-tical line. A more detailed treatment of this topic can be found elsewhere.2

Applying different control mechanisms such as speed varia-tion, adjustable inlet vanes, and adjustable diffuser vanes allow the compressor to operate on a family of curves.9 Additionally, compressors can be controlled by suction or discharge throttling and recycling (Figure 2).

Figure 3 indicates effectiveness and efficiency of different control methods. Of particular importance in upstream and midstream applications is a compressor that can be operated at varying speeds, since this is the most effective and efficient control method (Figure 3). Using a throttle, recycling or adjustable vanes are very effective in reducing the volumetric flow, but they are not very efficient, because the power consumption is not reduced at the same rate as a speed-controlled machine.

It is a common perception that recycling gas in a centrifugal compressor system indicates inherent problems and is to be avoided. Certainly, recycling gas is not very efficient. However, a well-designed recycle system allows the compressor to stay on line even if the station flow is zero. This capability often allows for optimized designs and improved operation.

COMPRESSOR PROTECTIONWithin the control system, sub-systems protect the compressor.

In general, process control will be enabled as long as the compressor stays within acceptable, predefined boundaries. These boundaries may include the following:

• Maximum and minimum operating speed• Stability limit (aka surge line)• Choke or overload limit (on some machines)• Pressure, temperature, and torque limits

STATION LEVELFurther, on the station level, if multiple units are used, control

can be exercised by selecting the number of units in operation. Compressor configurations within a station can include single compressors, single compressors supplied from or delivering into multiple headers, multiple compressors operated in parallel, multiple compressors operated in series, multiple compressors, or compressors with multiple sections operated in a train. Varia-tions may include multiple compressors in a train, with control of intermediate pressures, or multiple compressor trains in parallel (Figure 4).

CONTROL CONCEPTS FOR MULTIPLE UNITS IN A STATION

With multiple units on a station, the question becomes how to control them to achieve certain objectives. These objectives may be minimizing the running hours of units, optimizing the capability to absorb load swings, or minimizing fuel consumption or emissions. The first objective requires that as few units run as possible, while the second may require all or most units to run at part load for most

Figure 3. Power Consumption For Different Control Methods

Figure 2. Recycle System


of the time. Minimizing fuel consumption (which also equates to minimizing carbon dioxide [CO2] production) will usually involve strategies to cover the load with as few units as possible running (i.e., running units as close to full load as possible or not running them at all). It then becomes a question of whether it is better to run all but one of the operating units at full load, with the remaining unit at part load, or to run all units at part load. It is important to note that in all scenarios the load (i.e., the power made available to the compressor) is the control variable. The compressor speed is a result.

For two identical units running, the answer is virtually always to run both of them at equal part load. This is accommodated by operating the compressors at equal turndown or at equal gas turbine load (i.e., equal gas producer speed).

For more than two identical units running, the difference in fuel usage between N units running at the same load, some running units at full load, and the remaining at part load, is usually very small. The optimum there is more often determined by the resulting operating points of the compressors.

For units that are not identical, it is usually better to load the more efficient unit and capture the load swings with the less-efficient unit. In some instances, these schemes are also dictated by the starting reliability of individual units; i.e., a low starting reliability may dictate operational schemes that are otherwise less fuel efficient.

CONTROL OF MACHINES IN SERIES AND PARALLELMany control schemes for multiple independent compres-

sors use turndown equalization — they seek to operate all compressors at the same turndown. Turndown equalization for compressors in series works backward from turndown equalization for compressors operating in parallel. In series operation, one has to increase the speed (and thus power) of the unit where turndown has to be reduced. For units in paral-lel, increasing the speed (and power) of the unit will increase its turndown. This is because of the different boundary condi-tions. In parallel operation, all units see the same suction and discharge pressure, but the flow will depend on the power that is fed into the compressor. In series operation, the flow has to be the same through all machines (unless recycle is used). Increasing the power for one of the machines will increase the amount of head that said machine will produce relative to the other machine, thus reducing its turndown.

CONTROL OF COMPLEX CONFIGURATIONSMore complex configurations may involve trains with three

compressor bodies with two side streams. The train is driven by a two-shaft gas turbine, allowing operation over a wide speed range (Figure 5). The speed of the train is the result of the equi-librium between power output from the gas turbine and power consumption of the compressor train. In other words, if the gas turbine power is increased, the train will increase its speed. Each compressor symbol (Compressor1, Compressor2, Compres-sor3) comprises the compressor and its aftercooler. Each com-pressor has its own recycle valve, and there are check valves be-tween the compressor sections. Therefore, we have four means of controlling the train: train speed and three recycle valves. We also have essentially two control systems: the process control system (depending on the case, one will control the flow or pres-

N N NStream 1

Stream 2

CompressorAnd Cooler 1



Stream 3

Figure 5. Three-Section Compressor Train With Two Side Streams

Figure 6. Compressors for series and parallel operation. Transfer from parallel to series arrangement (C1 is the LP compressor, C2 is the HP compressor, P1 and P2 are the suction and discharge pressure, respectively, SM is the surge margin for the compressors). The transfer is controlled by the control valve, with the check valves opening and closing automatically. The control valve is open for series operation and closed for parallel operation. The delivered flow is the same at the beginning and the end of the transfer.


Figure 4. Different Compressor Configurations: a) Single Section, Straight Through Compressor, b) Multi-Section Compressor, And c) Multi-Body Tandem.

(a) (b)

(c)

sures) and the anti-surge system for each of the compressors.

In general, it is advantageous to con-trol the largest (by mass flow) stream us-ing speed control, and the other streams will be controlled by either throttling or recycling. For these concepts to work, the flow-pressure behavior of each stream must be understood.

SERIES-PARALLEL CONSIDERATIONS For some applications — for example,

gas storage — it is advantageous to be able to switch two compressors from se-ries to parallel operation and vice-versa. In doing so, the operating range can be in-creased significantly. In series operation, the units can provide high head, while in parallel, the flow range is increased.

It is desirable to be able to switch from series to parallel operation and vice-ver-sa, while the compressors are running at or near full load. This is possible with an appropriate arrangement of valves. The most elegant solution involves a control valve and two check valves (Figure 6). With the control valve open, the machines op-erate in series, while with a closed valve they operate in parallel. The check valves will open and close automatically based on the pressure differential over these valves, and therefore do not have to be controlled. In this arrangement, the open-ing and closing speed of the control valve is not particularly important.

The transfer from parallel to series configuration is simulated in Figure 6. It is shown that neither machine will enter into surge. The low-pressure (LP) com-pressor will be designed with higher flow staging than the high-pressure (HP) com-pressor.

CONTROLSBasic Process Control With A Gas Turbine Driver

The control system for a gas turbine driver process control is set up to run the engine to maximum gas generator speed (i.e., full load), unless it runs into another limit first. Limits can be com-pressor suction pressure, compressor discharge pressure, or compressor flow. If, for example, suction pressure is con-trolled, the engine will run at full load unless the suction pressure drops below its set point. In that case, the gas pro-ducer speed is reduced. In the case of discharge pressure control or flow con-

trol, the engine will run at full load unless the discharge pressure or the compres-sor flow exceeds its set point.

Basic Process Control With Electric Motor Drives

For constant-speed electric motor drives, process control has to be ac-complished using one of the other pre-viously mentioned control methods. The complication lies in the fact that control methods such as recycle control and suction throttling have only limited capability to reduce the compressor ab-sorbed power (Figure 3). Therefore, the driver is required to make a certain level of oversizing to ensure there is enough power available to operate the com-pressor at some point on its constant speed line.

For electric motor drives that al-low variable compressor speed (i.e., variable speed gearboxes or variable frequency drives), speed is the usual control variable. Unlike the gas turbine drives described, the speed of the drive is adjusted to meet the process control objective, until the required power ex-ceeds the driver capability.

Surge AvoidanceSurge avoidance, while the compres-

sor is on line, is one of the process con-trols for the left boundary (low flow) of the compressor map. The intervention of surge control should be virtually un-noticeable. It should be as though the compressor has infinite turndown.

Understanding the principles of surge avoidance initially will make understand-ing the remaining process controls easi-er. Some remarks about control dynam-ics and measurements follow. There are five essentials for successful surge avoid-ance.3

1. A Precise Surge Limit Model: It must predict the surge limit over the applicable range of gas condi-tions and characteristics.

2. An Appropriate Control Algorithm: It must ensure surge avoidance without unnecessarily upsetting the process.

3. The Right Instrumentation: Instru-ments must be selected to meet the requirements for speed, range, and accuracy.

4. Recycle Valve Correctly Selected For The Compressor: Valves must fit the


compressor map. They must be ca-pable of large and rapid, as well as small and slow, changes in capacity.

5. Recycle Valve Correctly Selected For The System Volumes: The valve must be fast enough and large enough to ensure the surge limit is not reached during a shutdown. The piping system is the dominant factor in the overall system response. It must be analyzed and understood. Large volumes will preclude the im-plementation of a single valve surge avoidance system.3-6

Since this paper does not cover the behavior of surge control systems during emergency shutdowns, all that needs to be mentioned here is that the volume of the pipes and vessels between the com-pressor discharge nozzle, the check valve, and the recycle valve should be kept as small as possible. If concerns about surge during emergency shutdown arise, a sep-arate hot recycle valve can be installed.

Figure 7. Behavior Of A Single-Stage Centrifugal Compressor At Positive And Negative Flow, Showing Isentropic Head Coefficient Versus Flow For Machine Mach Numbers From 0.57 To 0.85.11

Figure 8. System Characteristics And Compressor Map

Operation During Reverse FlowWhile the control system is supposed

to keep the compressor from operating in surge, the situation of the compres-sor is discussed when the system fails to prevent surge.

Test data exist describing the behavior of a compressor when it is subject to re-verse flow conditions. Kurz and White described the behavior of a compressor during emergency shutdown against a closed recycle valve, and recently, Be-lardini, et al., presented a detailed study on the compressor behavior in reverse flow.6,10 Aust11 published detailed mea-surements of a centrifugal compressor during surge cycles (Figure 7). Besides the normal, stable operation, the area of stable reverse flow and an unstable tran-sition area can be identified. Currently, the determination of the forces, as well as possible damage during surge cycles, is the subject of research. In general, control systems are designed to prevent the compressor from surging, but anec-dotal evidence suggests that most indus-trial compressors survive surge events without measurable damage.

SYSTEM BEHAVIOR The system within which the compres-

sor operates, that is, the piping, valves, and vessels, exhibits some relationship between the flow through the system and the pressure drop imposed by the system.

In the context of compressor appli-cations, it is important to understand this relationship, since it has a profound impact on the selection of the correct compressor. Further, these relationships tend to be different in steady-state op-eration versus transient operation.

The pipe system within which the

compressor operates will impose its characteristic on the compressor. There are three fundamental steady-state system characteristics that need to be considered (Figure 8):

(A) Strong head-flow relationship (B) Weak head-flow relationship(C) Integrative relationshipThe case of strong head-flow relation-

ship (A) is, for example, seen in gas pipe-lines. Under steady-state conditions, the pressure loss in the pipeline that impos-es the suction and discharge pressure on the compressor station increases significantly when the flow through the pipeline has to be increased. The pres-sure levels are thus dictated by friction losses, which depend on the gas velocity in the pipe.

In a weak head-flow relationship (B), the head requirement for the compres-sor head stays more or less constant with changes in flow. This behavior is found in refrigeration compressors, but also in situations in which the process dictates a constant suction pressure (e.g., separator pressure), while the dis-charge gas is fed via a short pipe into a larger flowing pipeline, so the compres-sor discharge pressure is more or less dictated by the pressure in the large pipeline. Friction losses therefore have a very small effect, and this results in very little changes in the pressure loss-es with flow.

In an integrative relationship (C), which exists, for example, in storage ap-plications,12 the compressor fills a large cavity. That means the compressor dis-charge pressure is increased as a func-tion of the cumulative flow into the cavi-ty, as a result of filling the cavity with gas. Line (C) is essentially showing a series of


For Given Load Setting (eg NGP), Compressor

Will Operate At Point Determined

By Available Power & Pipeline Characteristic

Figure 9. Available Power, Compressor Map, And Pipeline Characteristics

(e.g., NGP),

operating conditions at constant power. Similar conditions can be found in gas gathering applications where (on a much slower scale) the field pressure (and with it the compressor suction pressure) de-clines as a function of the cumulative flow out of the gas field. These fields additionally have a strong head-flow re-lationship; i.e., increasing the flow at any given time would lower the compressor suction pressure.

The interaction between compressor characteristic and system characteristic then becomes a basic ingredient for the control approach. Figure 9 shows how the power input provided by the driver can be used to control the compressor operating point within the constraint of the system behavior.

Compressor power (P) is a function of mass flow (W) and actual head (H), and thus related to the coordinates in the compressor map (Figure 9) of inlet density (ρ), inlet flow (Q), isentropic head (Hs), and efficiency (η):

(1) P = W x H = ρQ x

This defines the line of constant pow-er in Figure 9.

Further, the transient system behav-ior must be considered (Figure 10). For example, a pipeline can be operated in a transient condition by feeding more gas into the pipeline than what is taken off on the other end. This is usually re-ferred to as line packing. In general, pipe-lines are operated under slowly changing operating conditions. While a pipeline under steady-state conditions requires a unique station pressure ratio for a given flow (Figure 8), this is no longer true un-der transient conditions. If the pipeline operates under transient conditions, for example, during line pack after a fast in-crease in driver power, or, if one of the compressors has to be shut down, the steady-state relationships are no longer valid. Dynamic studies of pipeline behav-ior reveal a distinctly different reaction of a pipeline to changes in station operating conditions than a steady-state calcula-tion. In steady state (or, for slow chang-es), pipeline hydraulics dictate an increase in station pressure ratio with increased flow, due to the fact that the pipeline pressure losses increase with increased flow through the pipeline. However, if a centrifugal compressor receives more

Hsη

driver power, and increases its speed and throughput rapidly, the station pressure ratio will react very slowly to this change. This is because, initially, the additional flow has to pack the pipeline (with its considerable volume) until changes in pressure become apparent. Therefore, the dynamic change in operating conditions would lead (in the limit case of a very fast change in compressor power) to a change in flow without a change in head. If the power setting is maintained, the compressor oper-ating point would then start to approach the steady-state line again, albeit at a higher speed, pressure ratio, flow, and power.

Experimental data presented and analyzed by Blieske and Kurz indicates that the steady-state compressor map is still usable even in transient situations.8


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INTERACTION BETWEEN THE SYSTEM AND THE COMPRESSOR

For any situation, the process determines the suction and discharge pressure the compressor “sees.” Based on some control setting (available power, speed, guide vane setting) the compressor will react to the situation by providing a certain amount of flow to the system. Therefore, the flow into the system is a result of the compressor characteristic (its map) and some external control setting.

Different controls elicit different scenarios in these control situations. If we control the compressor by the level of power that is supplied, then the speed at which the compressor runs is an outcome of the interaction between compressor and process. If we control the speed of the compressor, the re-quired power is an outcome. The same is true for a constant speed machine (which in that sense is just a special case of a compressor that’s forced to operate at a set speed).

Recycle control and throttle control are essentially supple-mental ways to control the compressor in certain situations. Recycling gas still maintains the system suction and discharge pressure as long as the compressor stays on line, but it allows the compressor to provide more flow than desirable or avail-able from the system. Throttle control allows to reduce the system suction pressure, or the system discharge pressure, the compressor experiences.

There are obviously impossible outcomes. The compressor will not be able to operate at conditions where the speed is too high or too low, where the power demand is too high, or where the operation would cause an instability-like surge. If the compressor, due to constraints of power, speed, or flow range, is not capable of operating at the system-imposed suc-tion and discharge pressure, it will go into full recycle; i.e., it will operate within the constraints of a new system, that is, a throttle valve-controlled recycle loop.

The principles discussed also apply to transient situations, such as line pack in pipelines13 and even highly transient situ-ations, such as during an emergency shutdown.6,14 Again, the system (which is essentially the recycle loop as soon as check valves separate the recycle loop from the main system) impos-es a certain suction and discharge pressure on the compres-sor; the available power comes from the inertia of the drive train, and the compressor speed is a result of the interactions.

OPERATION IN A PIPELINE SYSTEMIn this example, the impact of different compressor charac-

teristics on the efficiency achieved under operating conditions imposed by the system is highlighted. A variable speed centrifu-gal compressor is compared with a reciprocating compressor. Unlike a centrifugal compressor, a reciprocating compressor will deliver a lower efficiency when the pressure ratio drops.15

The typical steady-state pipeline operation (Figure 8) will yield an efficiency behavior as outlined in Figure 11. This is the result of evaluating the compressor efficiency along a pipeline steady-state operating characteristic. Both compressors would be sized to achieve their best efficiency at 100% flow, while al-lowing for 10% flow above the design flow. Different mechanical efficiencies have not been considered for this comparison.

The reciprocating compressor efficiency is derived from valve efficiency measurements in Noall and Couch,15 with compression

Figure 11. Compressor Efficiency At Different Flow Rates Based On Operation Along A Steady-State Pipeline Characteristic (Pressure Ratio At 100% Flow = 1.4)

Figure 10. Typical operating points if transient conditions are considered — in this case, due to a fast engine acceleration from 50% to 100% load.13

efficiency and losses due to pulsation at-tenuation devices added. The efficiencies are achievable with low-speed compres-sors. High-speed reciprocating compres-sors may be lower in efficiency. The graph shows the impact of the increased valve losses at lower pressure ratio for recip-rocating machines, while the efficiency of the centrifugal compressor stays more or less constant.16

CONTROL OBJECTIVESIn a discussion of control, one has to

define the goal of a control system. The requirements to protect the process as well as the equipment have priority. But, other goals need to be defined, too, par-ticularly if the station involves multiple compression units, either in series or in parallel arrangements. Possible goals can be to minimize fuel consumption, mini-mize emissions, minimize maintenance cost, or maximize throughput.

Each of these goals has to be trans-lated into operations requirements for the compressors. For example, in a compressor station with three identical units, minimizing fuel consumption may be accomplished by running only the minimum number of units necessary for the duty. This generally also will minimize

maintenance costs (since the cumulative running hours are minimized), unless the gas turbine incurs additional main-tenance based on the number of starts. A simple but very effective scheme is outlined as follows:17 Assume we have a compressor station with one large unit (KC) and three smaller units (TC) all operating in parallel. For simplicity, the smaller unit produces half the power of the larger unit:

PKC

=2 PTC

and the compressors are aerodynamic scales, thus maintaining aerodynamic similarity. The parallel operation forc-es all units to operate at the same suc-tion and discharge pressure. Based on the above, the total available station power, P = 5 PTC.

The different load steps can be de-fined as follows:

Step 1: P = PTC

Step 2: P =PKC or P = 2 PTC

Step 3: P = PKC + PTC or P = 3 PTC

Step 4: P = PKC + 2PTC

Step 5: P = PKC + 3PTC

For all power demands that are high-er than step n, but lower than step n+1, the running units are equally loaded.

Further considerations have to be

Figure 12. Load Increase (Left); Pressure Control And Flow Control (Right)

Figure 13. Load Reduction (Left); Pressure Control And Flow Control (Right)

made to decide which options should be pursued for Steps 2 and 3. One consider-ation could be maintenance cost. It could be that the larger unit accrues lower main-tenance costs per fired hour than two of the smaller units. Similarly, if the larger unit is more efficient than the small units, one would opt for starting the larger unit in Steps 2 and 3.

Another option would be to analyze typical load cycle for the station. If the load typically rises beyond Step 3 rela-tively fast, it might be advantageous to start the large unit in Step 2. If, how-ever, the load often just stays between Steps 1, 2, and 3 (as is the case for sta-tions that have seasonally lower loads), then these steps may be better covered by the smaller units.

Lubomirsky, et al.,18 pointed out in their study on pipeline availability and fuel optimization that in particular situ-ations that show significant variation in ambient conditions, this simple control schematic is rather powerful, both in terms of minimizing fuel consumption and minimizing the running hours, and thus the maintenance cost of the units.

SIMULATIONAll simulations in this section assume

compressor operating points at con-stant head. The flow therefore changes with the amount of power supplied to the compressor. The change in com-pressor efficiency and operating speed are accounted for in the simulation. Also, the gas turbine power output and heat rate are adjusted for power tur-bine speed and load. The available pow-er is based on realistic site conditions.

From a process standpoint, we can control pressure, power, speed, and flow. For compression applications, we want to control a process variable, like flow or suction or discharge pressure. Speed and power are irrelevant from a process standpoint, except for machinery protec-tion, or to maximize production (i.e., op-eration at full load or full speed). Among the process variables in a compressor station, pressures tend to change rela-tively slowly, while flow changes fast. For compression applications, pressure (ex-cept as a limiting factor) is also often not that relevant. It is often used, however, for control purposes in pipelines, pos-sibly because it is easier to monitor, es-pecially if reciprocating compressors are


involved. That leaves flow control as the most advantageous way of controlling a compressor application. Figure 12 shows how the compressor tends to react under different control scenarios. It follows that the fastest way to get to a new operating point is to accelerate the engine to full power.13

Both flow and pressure control essentially will make the en-gine operate at full load until the control objective is achieved. Pressure control will lead to a large flow increase into the pipeline, causing a faster pressure rise in the pipeline. In oth-er words, pressure control will usually bring the compressor faster to the new set point than flow control in systems with a strong head-flow relationship (Figure 8, Curve A).

The case in which the unit is supposed to run at reduced flow requires that the anti-surge system is considered (Fig-ure 13). Upon setting the control set point to a lower flow (or pressure), the gas turbine will reduce power (by reducing gas producer speed). This will lead the compressor operating point to approach the surge control line. Upon crossing the control line, the recycle valve will open to keep the compres-sor from surging. The combination of opened recycle valve and reduced power will bring the compressor to the new set point. Notably, there is no concern about interactions be-tween power control and surge control.

CONCLUSIONSThere are two objectives for compressor control: meeting

the external process requirements and keeping the compres-sor within its operational boundaries. Typical control sce-narios that have to be considered are process control, start-ing and stopping of units, and fast or emergency shutdowns. The paper discusses the interaction between a centrifugal compressor and the process, and as a result, the control re-quirements for centrifugal compressor packages. The focus

is on variable-speed, upstream, and midstream applications. The impact of the interaction between system characteris-tics and compressor characteristics, both under steady-state and transient conditions has been explained. Also considered were concepts to optimize and control the units. Special at-tention is given to the issue of surge avoidance.

ACKNOWLEDGMENTSThe authors acknowledge the valuable suggestions and ad-

vice Robert C. White, Solar Turbines Incorporated (retired), contributed to this paper.

ABOUT THE AUTHORSDr. Rainer Kurz is the manager, Systems Analysis, at Solar

Turbines. His organization is responsible for analyzing compres-sion requirements, predicting compressor and gas turbine per-formance, conducting application studies, and for field perfor-mance testing. Kurz attended the Universitaet der Bundeswehr in Hamburg, Germany, where he received the degree of a Dr.-Ing. He has written numerous publications about turbomachin-ery-related topics, is an ASME fellow, and a member of the Tur-bomachinery Symposium Advisory Committee.

Dr. Klaus Brun is the director of research and development at Elliott Group. His experience includes positions of product development, engineering, project management, and executive management at Southwest Research Institute, Solar Turbines, General Electric, and Alstom. He holds nine patents, has written more than 350 papers, and has published three text books on energy systems and turbomachinery. Brun is a Fellow of ASME; chair of the 2020 Super critical CO2 Power Cycles Symposium; and a member of the API 616 task force, the ASME PTC-10 task force, the Asia Turbomachinery Symposiums Committee, and Supercritical CO2 Symposium Advisory Committee.


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Behavior for Centrifugal Gas Compressors,” 3rd Middle East Turbomachinery Symp., held in Doha, Qatar (College Sta-tion, TX: Turbomachinery Lab, Texas A&M University, 2015).

3 R.C. White, R. Kurz, “Surge Avoidance for Compressor Systems,” Proc. 35th Turbomachinery Symp. (College Sta-tion, TX: Turbomachinery Lab, Texas A&M University, 2006).

4 K.K. Botros, S.T. Ganesan, “Dynamic Instabilities in In-dustrial Compression Systems with Centrifugal Compres-sors,” Proc. 37th Turbomachinery Symp. (College Station, TX: Turbomachinery Lab, Texas A&M University, 2008).

5 K.K. Botros, “Single vs. Dual Recycle System Requirements in the Design of High Pressure Ratio, Low Inertia Cen-trifugal Compressor Stations,” ASME GT2011-450022011 (West Conshohocken, PA: ASTM International, 2011).

6 R. Kurz, R.C. White, “Surge Avoidance in Gas Compres-sion Systems,” Trans. ASME J. Turbo. 126 (2004): p. 501-506.

7 M. Morini, M. Pinelli, M. Venturini, “Development of a One-Dimensional Modular Dynamic Model for the Simu-lation of Surge in Compression Systems,” ASME J. Turbo. 129 (2007): p. 437-447.

8 M. Blieske, R. Kurz, A. Garcia-Hernandez, K. Brun, “Cen-trifugal Compressors During Fast Transients,” Trans. ASME JEGTP 133 (2011): p. 072401.

9 P.C. Rasmussen, R. Kurz, “Centrifugal Compressor Ap-plications,” 38th Turbomachinery Symp. (College Station, TX: Turbomachinery Lab, Texas A&M University, 2009).

10 E. Belardini, D.T. Rubino, L. Tapinassi, M. Pelella, “Four Quadrant Centrifugal Compressor Performance,” Proc. 1st Asia Turbomachinery and Pump Symp., held in Singa-

Reprinted from the August 2019 issue of Gas Compression Magazine. www.gascompressionmagazine.com Copyright Third Coast Publishing Group.

pore (College Station, TX: Turbomachinery Lab, Texas A&M University, 2016).

11 N. Aust, “Ein Verfahren zur digitalen Simulation insta-tionaerer Vorgaenge in Verdichteranlagen” (A Method for the Digital Simulation of Unsteady Processes in Compres-sor Plants), Diss. UBwHH, Hamburg, Germany, 1988.

12 R. Kurz, K. Brun, “Assessment of Compressors in Gas Storage Applications,” Trans. ASME JEGTP 132 (2010).

13 R. Kurz, R.C. White, K. Brun, “Transient Operation in Pipeline Compressor Stations,” ASME. Turbo Expo: Power for Land, Sea, and Air, paper no. GT2014-25016; doi:10.1115/GT2014-25016.

14. J.J. Moore, A. Garcia-Hernandez, M. Blieske, R. Kurz, K. Brun, “Transient Surge Measurements of a Centrifugal Compressor Station During Emergency Shutdowns,” Proc. 38th Turbomachinery Symp. (College Station, TX: Tur-bomachinery Lab, Texas A&M University, 2009).15. M. Noall, W. Couch, “Performance and Endurance Tests of Six Mainline Compressor Valves in Natural Gas Com-pression Service” (Dallas, TX: GMRC Gas Machinery Con-ference, 2003).16. R. Kurz, B. Winkelmann, S. Mokhatab, “Efficiency and Operating Characteristics of Centrifugal and Reciprocating Compressors,” Pipeline and Gas Journal (2010). 17. R. Zamotorin, R. Kurz, D. Zhang, M. Lubomirsky, K. Brun,K., “Control Optimization for Multiple Gas Turbine Driven Compressors,” ASME. Turbo Expo: Power for Land, Sea, and Air, ASTM paper no. GT2018-75002; doi:10.1115/GT2018-75002. 18. M. Lubomirsky, R. Kurz, R. Zamotorin, “Calculation of Gas Pipeline Compressor Station Availability Factors Using the Monte Carlo Simulation Method,” Pipeline Simulation Interest Group Conf. (Houston, TX: PSIG, 2016).

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