Operators of petroleum plants

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1 Validating anti-surge control systems At a gas processing plant, real time simulation was used to analyse the design of a surge protection scheme for a natural gas compressor, after questions arose over the system s effectiveness in the event of an emergency shutdown Nikhil Dukle and Krishnan Narayanan Compressor Controls Corporation Natural gas parameters in suction header (inlet boundary conditions): P=36.3 bar (a), T=35 C KT-8001 GT (not simulated) SDV-8000 VO5=0 m 3 Figure 1 Simulation model schematic M A-8001 VO= m 3 FV Cooler volume=5 m 3 VO4= m 3 VO3=5.8 m 3 ASV-6511 VO1=47.5 m 3 Natural gas parameters in discharge header (outlet boundary conditions): P=81.47 bar (a), T=108. C Operators of petroleum plants often contract compressor or turbine OEMs to supply an entire turbomachinery train. The scope of supply normally includes the design of the peripherals systems, such as the anti-surge control system for centrifugal and axial compressors. While there are advantages to having a single party responsible for the entire train, it can result in "finger pointing" and confusion when the various equipment OEMs disagree on the design of the peripheral systems. Engineers at a gas processing plant recently found themselves in this situation when adding a medium-pressure natural gas compressor as part of a major plant expansion. The gas turbine manufacturer, who was the responsible OEM, designed an anti-surge protection scheme for the process compressor, but the compressor manufacturer felt the design was inadequate. Specifically, the compressor manufacturer felt that in addition to the cold recycle loop provided in the initial design, a hot recycle loop was also needed to protect the compressor from surge, especially in the event of an emergency shutdown (ESD). Good anti-surge control system design is difficult, and it encompasses the proper sizing, selection, and location of all of the following elements: piping that comprises the recycle loops, the recycle valves, volumes of vessels in the recycle path, check valves, and last but no less important the anti-surge controller. It is very difficult, using steady-state analysis, to predict the effectiveness of a proposed anti-surge control system for various possible process-upset scenarios, due to the rapid onset of surge (approximately 300 milliseconds) and the interaction of the various components of the system. Because both manufacturers were basing their decisions on steadystate analysis, the plant engineers lacked confidence in either party s opinion. Validating anti-surge control system design is an area where advanced, highfidelity simulation can be applied successfully for scenario testing, and in an attempt to determine the true requirements for machine protection, plant personnel commissioned Compressor Controls Corporation (CCC) to perform a dynamic simulation of the two antisurge control system proposals. Using readily available data, the ability of the control systems to prevent compressor surge under the following scenarios was validated through simulation: ESD of the compressor train while operating in steady-state at design conditions Full closure of a process valve on the compressor discharge side with the compressor operating at rated condition in steady-state Full closure of a process valve on the compressor suction side with the compressor operating at rated condition in steady state. The results of these simulation scenarios assisted the plant engineers in analysing the competing solutions and choosing an appropriate design for the compressor anti-surge system. Simulation model A schematic of the dynamic simulation model used to describe the various elements of the compression system including both hot and cold recycle loops is presented in Figure 1. The entire system was broken down into the compressor, valves, and a number of volumes for associated pipes, knockout drums and the gas-cooler. The parameters of every element were determined from the data provided by the equipment manufacturers and the gas plant engineers. In the few cases where certain 1 PTQ SUMMER ptq.com

2 data were not available, values were assumed based on field experience and common-sense engineering practice. In addition, the simulation model was connected to real controllers in an in-the-loop setting. This enabled direct comparison of a generic antisurge controller employing standard PID-type algorithms and a purpose-built anti-surge controller employing advanced control methods and predictive algorithms. Model assumptions The following is a list of the major features of the model and associated simplifying assumptions: Natural gas parameters molecular weight, pressure and temperature in the suction and discharge headers are assumed to be constant and are set up as boundary conditions for the model. Three invariant coordinates compression ratio, reduced-speed and reduced-flow are used to approximate process compressor performance map. Compressor reduced-speed and compression ratio are used to calculate compressor inlet flow. Compression ratio and inlet flow are used to calculate shaft power. It is also assumed that compressor performance curves provided are valid in steady-state as well as transient conditions. The compressor map is defined only over the speed range provided by the manufacturer (that is, 70% to 105%). No attempt has been made to extrapolate the map for speed beyond this range. All pipes are modelled using a lumpedparameter approach to calculate pressures, temperatures and flows along the process. Pipes resistances are included into resistance of associated valves, and pipe volumes are split equally between the main upstream and downstream volumes connected by the pipe. Pressures and temperatures in process volumes are calculated using ordinary differential equations describing the accumulation of mass and enthalpy in the volumes. The volume associated with the gas cooler in the compressor discharge section is split equally between adjacent upstream and downstream volumes. Process valves are modelled using standard ISA sizing equations, and take into account the actual valve size and flow conditions. Suction and discharge block valves are assumed to have linear trim characteristics and a stroke-time of sec. Suction and discharge valve closure tests were performed at constant compressor speed The OEM anti-surge controller was modelled as a conventional PID algorithm with a 0 millisecond loop-execution rate. Anti-surge controllers provided by machinery OEMs typically employ standard PID algorithms or a minor variation thereof. The ESD scenario is simulated without any anti-surge controller in the control loop. On ESD, the anti-surge valves are opened and go full open based on their stroke times. The delay between the ESD signal and the opening of anti-surge valves is set to 0 milliseconds to simulate the time required for a solenoid valve to depressurise. Actuator stroke-time for full opening was assumed as.0sec for the hot recycle valve and.sec for the cold recycle valve. The stroke time for the cold recycle is set sec longer for the purpose of display in an effort to distinguish its trace from that of the hot recycle valve. This assumption does not have any significant impact on the results. Shaft deceleration rate on ESD is 4% per second. Validation Once the model was developed, a comparison was made between the design data provided by the OEM and the compression system model results under steady-state conditions with the compressor at the rated operating point. The values of speed, suction pressure, discharge pressure, discharge temperature, compressor power and inlet flow were compared, with a maximum error between design values and model-calculated values of %. The small error between the values validates the model under steady-state conditions, and provides good initial reconciliation for the analyses of transient process disturbances. In addition, gas flows through the cold recycle valve and hot recycle valve Figure Operating parameter trends on ESD, cold recycle only were compared. The model-generated flows exactly matched the design-calculated flows in steady-state conditions, which validated the recycle valve model with respect to valve size and trim. Simulation results Results for each test scenario are presented as a plot of time traces of pertinent compressor parameters. A proximity to surge variable, S S, describes the location of the compressor s operating point as compared to its surge limit, and is invariant to changes in suction conditions. If S S < 1, the compressor is operating in the safe zone of the available operating envelope; if S S = 1, the operating point is at the surge limit of the compressor; if S S >1, the operating point is in the non-stable zone to the left of the surge limit and the compressor is in a surge condition. Scenario 1 Compressor ESD Scenario 1 analysed the ability of the hot and cold recycle-loops, as proposed by the compressor manufacturer, to protect the compressor in the event of an ESD. This scenario was further broken down into three separate sub-scenarios (Cases 1 to 3) to analyse the effectiveness of the hot and cold recycle loops individually and together. Case 1: open cold recycle valve As seen in Figure, fully opening the cold recycle valve () immediately on ESD does not prevent surge. In the figure, the ESD occurs at time t = 0sec; the cold recycle valve begins to open 0millisec after ESD and opens fully in.sec. The hot recycle valve remains closed at all times. Note, however, that the operating point reaches the surge limit less than 1sec after the ESD. The reason for surge is the inability of the cold recycle valve to quickly and effectively de-pressurise the large volume in the cold recycle flow loop. This large volume is essentially made up of PTQ SUMMER 003

3 Figure 3 Operating parameter trends on ESD, hot recycle only the cooler and associated piping. The time constant of this volume when operating at design conditions is about 70sec, which is too long for effective and quick de-pressuring. Since the rate of speed reduction (and therefore compressor flow reduction) is greater than the rate of compressor pressure reduction, the compressor continues to surge as it coasts down. This is indicated by the proximity to surge indicator S S, which remains greater than 1 for the duration of coast-down. Case : open hot recycle valve As seen in Figure 3, fully opening the hot recycle valve () on ESD prevents surge, but the compressor operation reaches its surge limit (S S = 1). In the figure, the ESD occurs at time t = 0sec; the hot recycle valve begins to open 0millisec after ESD and opens fully in sec. The cold recycle valve remains closed at all times. Within 1.sec after ESD, the compressor operating point reaches its surge limit, but the compressor is protected, albeit with very little margin for error. Comparing this to Case 1, the hot recycle loop is obviously much more effective in de-pressuring the compres- sor discharge as the flow through the compressor reduces when speed begins to drop. The check valve between the hot and cold recycle loops is very effective in isolating the large cooler volume from the hot recycle loop as compressor discharge pressure decreases. In this case, the time constant of the volume between the compressor discharge and check valve upstream of the cooler is about sec; much smaller than the 70sec for Case 1. Because of the short time period between the disturbance and the approach-to-surge event, the influence of hot recycle valve opening on compressor suction temperature is ignored. Case 3: opening hot and cold recycle valves As seen in Figure 4, fully opening both the hot and cold recycle valves simultaneously on ESD is very effective at preventing surge. The compressor operation never reaches the surge limit, as compared to Case (maximum S S = 5 as compared to S S = 1). In the figure, the ESD occurs at time t = 0sec; the hot and cold recycle valves both begin to open 0millisec after ESD and open fully in.0 and.sec respectively. The Figure 4 Operating parameter trends on ESD, hot and cold recycle compressor operating point reaches the closest to its surge limit seconds after ESD. This is very similar to Case, except that the compressor operation does not reach the surge limit. Scenario Discharge valve closure Scenario 1 tests confirmed the compressor manufacturer was correct in regarding the absolute necessity of a hot recycle loop for protection during an ESD. Scenario then analysed the ability of the hot and cold recycle loops to protect the compressor in the event of a full closure of the compressor discharge valve, which was ramped from a fullyopen position to a fully-closed position over a period of 15sec. Initial tests revealed that the cold recycle valve, by itself, should be adequate to protect against slow disturbances, while both the hot and cold recycle valves may be required for protection against fast disturbances. This scenario was initially run using a conventional PID controller, representing the OEM controller, as described in Case 1. A purpose-built anti-surge controller with several advanced-control features was substituted for the conventional PID controller in Case. Case 1: discharge valve closure (DVC) with OEM controller As seen in Figure 5, the discharge valve begins to close at time t = 33sec, and reaches the fully-closed position in 15sec at time t = 48sec. The cold recycle valve comes into operation 17sec after the discharge valve begins to close. The time trace of S S, the proximity-to-surge variable, indicates the oscillations experienced by the compressor. Note that there is virtually no safety margin, that is, S S nearly reaches a value of 1 (surge). The compressor operating point is finally stabilised with the cold recycle valve at the 38% open position. Both hot and cold recycle controllers were configured with a surge-control margin of 5.5% of suction flow. Although the conventional controller was able to prevent surge, it had to be tuned very aggressively (K p = 8, K r = 4). This lends itself to potential instability in case of faster disturbances or when the compressor speed is changed to control a process parameter at the same time that the recycle valve is modulated. Both of these occurrences are very common, especially during start-up and partial-load operation. If the likelihood of oscillations is to be reduced, the controller will either have to be de-tuned, which will result in compressor surge, or the surge control margin will have to be increased, which reduces the efficient operating envelope of the compressor and will result in 3 PTQ SUMMER 003

4 , Figure 5 Operating parameter trends for discharge valve closure, PID controller, Figure 6 Operating parameter trends for discharge valve closure, purpose-built controller excessive wasteful recycling during startup and partial-load operation. Another detrimental aspect of conventional PID algorithms is that constant tuning parameters are employed for disturbances in both directions. Thus, there is always a possibility that the algorithm will be unable to provide appropriate control response for a new, untested large disturbance. This problem reveals itself later during the suction valve closure tests. Case : DVC with purpose-built antisurge As seen in Figure 6, the compressor discharge valve () begins to close at time t = 19sec. The valve reaches the fully-closed position in 15sec at time t = 34sec. The cold recycle valve comes into operation at time t = 34sec, approximately 15sec after the discharge valve begins to close, sec faster than in Case 1. The cold-recycle anti-surge controller effectively turns the operating point back toward the stable zone before it reaches the surge limit. S S reaches a maximum value of 4, and the compressor does not experience any surge cycles. This test clearly shows the effectiveness of the advanced, surge-prevention algorithms of the purpose-built controller when compared to a conventional PID controller. In order to understand why the purpose-built anti-surge controller is more effective than the OEM controller, it is necessary to look at the action of the control algorithms employed. The OEM (PID) controller acts on the error between the compressor s operating point and the surge control line, and no action is taken until the operating point moves to the left of the control line. The purpose-built controller uses a combination of closed-loop, open-loop, and anticipatory control responses, each with different tuning and control points. These algorithms allow the controller to begin moving the recycle valve even before the operating point crosses the surge control line. The response of the previously mentioned algorithms can be seen by looking at the detailed movements of the recycle valves. The anticipatory control algorithms begin to open the cold recycle valve at approximately t = 35sec, based on the rate-of-change of the compressor s operating point measured as a function of S S. Note that the opening begins when S S =, even though the compressor operation is to the right of the surge control line (S S = on the surge control line and S S = on the surge limit line). With larger and/or faster disturbances, it may be necessary to manipulate both the hot and cold recycle valves simultaneously. In such cases, the hot recycle valve will open for a very short time, arresting the movement of the operating point towards the surge limit. Thus, the control strategy employed is to use the cold recycle loop exclusively to protect against small disturbances and the hot recycle loop to protect against large disturbances. Once the disturbance is sufficiently arrested, the hot recycle valve will close to prevent overheating the gas at the compressor suction, while the recycle flow required for safe operation is slowly transferred to the cold recycle loop. The two controllers (the hot-recycle anti-surge controller and the cold-recycle anti-surge controller) collaborate to provide effective and efficient anti-surge protection. Scenario 3 Suction valve closure Scenario 3 analysed the effectiveness of the cold recycle loop in protecting the compressor in the event of a full closure of the compressor suction valve, which was ramped from a fully-open position to a fully-closed position over a period of 5sec. This is analogous to a loss-ofload condition. Scenario 3 was run using both a conventional PID controller model (described in Case 1) and a purpose-built anti-surge controller (described in Case ). The tuning constants for the two controllers are identical to those used for Scenario. Case 1: Suction valve closure (SVC) () with OEM controller As seen in Figure 7, the suction valve () begins to close at time t = 9sec. The valve reaches the fully-closed position in 5sec at time t = 54sec. The cold recycle valve comes into operation at time t = 51sec, approximately sec after the suction valve begins to close. The time trace of S S, the proximityto-surge variable, indicates the compressor s approach to its surge limit. The compressor enters the unstable surge region at time t = 53sec (S S >1) and oscillates back and forth while the 4 PTQ SUMMER 003

5 PID controller modulates the cold recycle valve in an effort to move the compressor back into the stable operating zone. The tuning constants were the same as those used in Scenario, Case 1. While these settings provided an adequate response for Scenario, they result in very aggressive valve movement for this scenario and cause unstable and unsafe operation. In fact, the compressor operation is initially pushed into the surge region because of the valve oscillation seen at time t = 50 to 53sec. The only alternative available to the operator is to manually position the cold recycle valve such that surging stops. If the algorithm is de-tuned to suit this scenario, then it becomes ineffective for Scenario. This is the dilemma presented by a constant-gain PID algorithm. While it can be effectively tuned for one scenario, it may prove to be totally ineffective for all others. Case : SVC () with purposebuilt anti-surge As seen in Figure 8, the compressor suction valve () begins to close at time t = 3sec. The valve reaches the Figure 7 Operating parameter trends for suction valve closure, OEM controller fully closed position in 5sec at time t = 48sec. The cold recycle valve comes into operation at time t = 43sec, approximately 0sec after the suction valve begins to close; the hot recycle valve does not open at all. The maximum value of S S is 3, indicating that the compressor operating point remained well within the stable operating zone. The advanced algorithms of the purpose-built controller are able to prevent compressor surge effectively, without any control-loop oscillations. In fact, the cold recycle valve begins to open through a combination of anticipatory and open-loop algorithms before the operating point even reaches the surge control line Figure 8 Operating parameter trends for suction valve closure, purpose-built controller Conclusion In the situation described in this article, where the end user was faced with trying to reconcile the differing recommendations of the two manufacturers, high-fidelity simulation was useful in evaluating the effectiveness of each competing solution. Clearly, the compressor manufacturer s recommendation to include a hot recycle loop was appropriate, a result that is not evident without simulation. Simulation was also indispensable in identifying a problem the end-user was unaware of that is, the inability of the OEM controller to adequately protect the compressor from surge in the case of all potential disturbances. Simulation also helped in identifying the many benefits of a purpose-built anti-surge controller: The anticipatory algorithms of the purpose-built controller were able to protect the compressor in the case of suction valve and discharge valve closures. In fact, purpose-built controllers can tailor their control response to the size and speed of the disturbance and can effectively protect against any disturbance. The OEM approach of opening the recycle valves only when the operating point reaches the surge control line was neither sufficient nor effective in preventing surge over a wide range of disturbances. In the case of disturbances larger than those tested, coordination between hot and cold recycle loops, as provided by the purposed-built anti-surge controllers, will be more effective both in surge prevention and in minimising heating of the compressor suction due to hot recycle. The use of simulation is obviously not always required or appropriate, but in cases where design questions exist, or the effectiveness of different controllers is being evaluated, high-fidelity simulation can be an effective and useful tool. Acknowledgements The authors wish to thank Gregory Lyulko and Al Sheldon, of Compressor Controls Corporation, for their assistance in the preparation of this article. Nikhil M Dukle is manager of simulation and sales support with Compressor Controls Corporation, Des Moines, Iowa, USA. He designs, tests, and commissions integrated control systems and turbomachinery simulation systems for internal and external use, and provides sales support. He has a BS in mechanical engineering from the University of Poona, India, and an MS in mechanical engineering from the University of Houston, Texas. ndukle@cccglobal.com Krishnan Narayanan is director of application technology development at Compressor Controls Corporation. He develops and implements new algorithms for turbo-machinery control systems. He has a BS in technology in mechanical engineering from the Indian Institute of Technology, Madras, an MS in mechanical engineering from the University of Minnesota, Minneapolis, and an MBA from Iowa State University in Ames, Iowa. 5 PTQ SUMMER 003