Air Compressor Surging and Pulsation Diagnosis for Centrifugal and Screw Units

The OEM's surge line will shift once it reaches the field, and how much it shifts depends on the discharge system differences

The surge line in the compressor's factory documentation was calibrated on the OEM's test stand. The discharge piping length, diameter, whether there's a buffer tank in between, and what test gas was used (most OEMs use air or nitrogen in closed-loop testing) are all different from your field installation. API 617's definition of the surge line is "the stable operating limit determined under specified test conditions." Pay attention to the words "specified test conditions." When conditions change, the line changes.

How much does it shift? There is no universal answer. Elliott and MAN Turbo machines in certain refinery wet gas applications have had field-calibrated surge lines shifted four to six percent to the right compared to the factory test report. The more the gas molecular weight deviates from the test gas and the more the discharge system volume and resistance characteristics deviate from the test stand conditions, the larger the shift.

This means the surge control line set in the anti-surge controller (most installations use CCC's S1 or S2 series, and some older installations still run their own anti-surge logic on a Triconex platform) has less margin than the engineer thinks. API 670's appendix gives vibration criteria recommendations for surge detection, but it cannot solve the problem of the surge line position itself being inaccurate.

Conducting a field surge test during commissioning is the only rigorous way to resolve this. The method is to activate the protection system and then gradually reduce flow or increase discharge pressure, pushing the operating point step by step toward the boundary, using high-speed data acquisition to capture the dynamic pressure signal of the first surge event, and then calibrate. CCC's S1 controller has a dedicated surge test mode to support this process. Not many installations have done it, because risk management approval is frequently not obtainable, especially on older installations that have been in operation for years where nobody wants to take this risk on an in-service machine.

Surge is a coupled oscillation between the discharge system and the compressor

Understanding surge as the compressor itself convulsing will mislead the diagnostic direction. The discharge piping plus buffer tank forms a Helmholtz resonance cavity, and the compressor is the energy input end of this resonance system. When the energy the compressor inputs can no longer sustain the downstream backpressure, the gas decelerates, reverses, discharge pressure drops, forward flow re-establishes, discharge pressure recovers, and the cycle repeats. The frequency is determined by the acoustic parameters of the discharge system and has nothing to do with compressor speed.

Industrial pressure gauge and discharge piping system — Discharge system geometry determines surge oscillation frequency

The same machine, connected to a compact system on the discharge side (short piping, small buffer tank or no buffer tank at all), can produce surge at three to four hertz, with each reversal making an audible bang that everyone can hear. Connected to a long pipe network plus a large receiver, the frequency drops to fractions of a hertz, showing up on the DCS discharge pressure trend as a slow oscillation, perhaps only a few percent of discharge pressure in amplitude, and the operator staring at the screen may not realize something is wrong.

Low-frequency surge has a longer duration per reversal event. Longer duration means the thrust bearing endures reversed axial loading for a longer time. Kingsbury or Waukesha tilting-pad thrust bearings respond to reversed loading very differently from steady-state forward loading. Under forward loading the oil film is established and load-carrying capacity is assured. When reversed loading is suddenly applied, the oil film must re-establish on the opposing thrust face, which takes time. During this transition, if the load already exceeds the transient load-carrying capacity of the oil film, the babbitt metal contacts the thrust collar directly.

During high-frequency surge this transition period is very short; the oil film has not been fully destroyed before the reversal ends and the axial force returns to the normal direction. During low-frequency surge the thrust bearing endures reversed loading for a longer time, and the damage mechanism leans toward heat accumulation and plastic deformation of the babbitt metal. During an overhaul inspection, if the babbitt surface shows discoloration turning purple with signs of peeling, this is characteristic of prolonged overheating, pointing to low-frequency surge or accumulated thermal damage from repeated medium-frequency surge events. If the surface shows scoring, embedded metal particles, and localized spalling, that looks more like oil film breakdown caused by high-frequency impact. API 670's monitoring recommendation for thrust bearing wear is to watch the axial displacement trend, which is correct, though axial displacement tells you "damage has already occurred," not "what mechanism is the damage occurring by."

How to separate rotating stall and surge on the frequency spectrum

When subsynchronous vibration and noise changes appear in the low-flow region of a centrifugal compressor, the first thing to do is determine whether it is rotating stall or surge. The subsequent course of action diverges completely.

Install a dynamic pressure transducer on the discharge piping (a piezoelectric sensor from PCB Piezotronics or Kistler works fine, with response frequency up to tens of kHz, more than enough for this application), vary the speed, and watch the pressure oscillation frequency response. If the frequency changes proportionally with speed, it is rotating stall. If it does not track with speed, it is surge.

A few more words on rotating stall. Among centrifugal compressors covered by API 617, designs with vaneless diffusers are more prone to rotating stall at low flow than those with vaned diffusers. In a vaneless diffuser, when the gas flow angle deviates from the design value, there are no vanes to correct the flow direction, and separation occurs more easily. Some machines just run with mild stall in a certain range below design flow, with vibration levels slightly elevated but within the API 670 alarm limits, losing a few points of efficiency. If this condition is misidentified as incipient surge, the operator or control logic may intervene unnecessarily, such as forcing increased recycle flow or reducing speed, which instead introduces new process disturbances.

Why centrifugal compressor surge happens

The operating point ended up on the left side of the surge line. That is the only direct cause. How the operating point got there needs to be broken down by situation.

Industrial compressor discharge and separator system — Operating conditions and system changes drive the compressor toward the surge boundary

Sudden reduction in downstream demand is the most common trigger scenario. An upstream process unit trips, a downstream user switches over, a batch ends, and flow drops. The anti-surge controller should open the recycle valve to maintain minimum compressor throughflow. There are many places within this "should" where things can go wrong, which will be discussed in detail later.

Gradual increase in downstream resistance causing discharge pressure to slowly climb. The time scale for this type of event is days or even weeks. Discharge-side heat exchanger fouling, separator liquid level control malfunction with the level running high and occupying gas-phase space, a manual valve being inadvertently left at half-open position. The diagnostic clue is to plot the historical trend of discharge pressure against flow and see if there is a slow upward drift. Many DCS historian databases only store single-variable trends and do not have the function to plot discharge pressure versus flow as an XY graph. If the engineer exports the historical data and draws the discharge-pressure-versus-flow trajectory overlaid on the compressor map, the picture often becomes clear at a glance.

Compressor performance degradation. Impeller fouling, labyrinth seal clearance increase, diffuser damage. These cause reduced head-generating capability at the same flow, equivalent to the surge line shifting rightward on the map. API 617 requires the OEM to provide predicted performance curves and a guarantee point, but does not require the OEM to provide a "predicted surge line shift under performance degradation." This information can only be obtained through periodic field performance testing (measuring actual head and flow and plotting them on the original curves for comparison).

Impeller fouling needs separate discussion. The deposit layer is unevenly distributed; the blade suction surface and the shroud inner wall are the worst-hit areas. This changes the blade exit angle and flow passage cross-section, and the impact varies in severity across different flow operating points. Sometimes the surge line does not shift uniformly but bulges outward in a specific speed range. A machine with comfortable surge margin at 85% speed might have no margin left at 78% speed. If the field performance test is only done at one point near rated speed, this localized sensitive zone will be missed.

Gas composition changes. Hydrogen purity fluctuations in refinery recycle hydrogen compressors, reaction gas composition changes with conversion rate in chemical plants, these alter the gas molecular weight, which in turn affects the relationship between polytropic head and volume flow, causing the operating point to drift on the map.

Among the above situations, the first one accounts for the majority of field surge events.

Anti-surge valve: failure modes beyond sticking

Anti-surge valve sticking (valve sticking) can be found through valve stroke testing. Both API 612 and the ISA-75 standard have requirements and methods for control valve stroke testing. This will not be elaborated on here.

There is a problem that cannot be found without dedicated testing: the valve position has arrived, but the flow has not.

Industrial laboratory analysis and quality testing — Valve Cv deviation develops gradually in the most critical operating range

CCC's anti-surge controller (as well as other brands such as Honeywell and Yokogawa solutions) receives valve position feedback from a stem displacement sensor. The sensor tells the controller "the valve is at 17% open," the controller calculates from the valve Cv curve and the differential pressure across the valve that "17% open should provide this much recycle flow," and considers the margin sufficient.

The problem is that the Cv curve has changed and nobody knows. The operating conditions for anti-surge valves are harsh: frequent rapid small-stroke movements near the fully closed position, handling gas that may carry liquid droplets or trace solids. After two or three years in this service, Fisher or Masoneilan valve trims develop seal face wear and deposits in the small-opening range that cause the Cv to deviate significantly from the factory calibration curve. The 10% to 20% opening range is precisely the most frequently used working range in anti-surge scenarios, and also the range where Cv deviation manifests first.

The verification method is not complicated: at a stable operating condition, manually force the valve to a known opening, observe how much compressor flow increases, and compare with the theoretical value at that opening from the Cv curve. If the deviation exceeds 30%, the valve's performance in surge protection scenarios can no longer be trusted.

Resistance changes in the recycle piping are another blind spot. The recycle line has a silencer in series (for noise reduction, whose internal packing compacts or accumulates dust over time, increasing pressure drop), a recycle cooler (tube-side or shell-side fouling, increasing pressure drop), and a check valve (to prevent suction-side gas from flowing backward into the discharge side, whose disc may stick at a half-open position adding resistance). The pressure drop changes across these elements are gradual and have no dedicated monitoring points. The valve Cv has not changed, the piping resistance has changed, and the effective recycle flow has still decreased.

There is also a time-scale issue. Both API 521 and API 617 mention response speed requirements for anti-surge systems, and the general engineering consensus is that the total time from signal receipt to effective flow change at the recycle valve should be within a few hundred milliseconds. The full-stroke time measured during commissioning (for example, a Fisher Baumann 24000 valve body going from fully closed to fully open in 1.8 seconds) is not the same thing as small-signal step response. In an anti-surge scenario the valve needs to move quickly from, say, 3% opening to 18% opening. Within this small stroke range, stem packing friction, positioner dead band and gain settings, and the pneumatic actuator's volume and supply pressure all affect the actual response speed. Superimposing a 5% step on the controller output and observing the actual valve position response waveform using the CCC controller's built-in high-speed data recording or an externally connected high-speed acquisition system is a test that should be done, and many plant commissioning checklists do not include this item.

Dynamic closing characteristics of the discharge check valve

API 617 Section 2.7.2.3 mentions that the check valve should prevent reverse flow, but does not provide specific quantitative requirements for the check valve's dynamic closing speed. Piping specification sheets typically only consider flow capacity and pressure rating when specifying check valves; the column for closing speed is left blank.

When surge occurs, gas begins to reverse from the discharge side back toward the compressor. The discharge check valve should close to isolate the compressor from the downstream high-pressure pipe network. A conventional swing check valve relies on the dynamic pressure of the reverse flow to push the disc. When the reverse flow first begins, the velocity is low, the pushing force is insufficient, and the disc wobbles at a partially open position. By the time the reverse flow velocity is high enough, the disc slams shut, and the resulting water hammer pulse is superimposed on the pressure fluctuation from the surge itself. This water hammer pulse plus the kinetic energy of the surge reverse flow is all borne by the thrust bearing.

Spring-loaded non-slam check valves from Goodwin or Neles have a spring preload on the disc so that when the forward flow velocity drops to near zero, the disc has already begun its closing motion without needing reverse flow to push it. The closing process is smooth, and the water hammer pulse is much smaller. After installing a spring-assisted check valve, the peak axial force experienced by the thrust bearing during a surge event can be reduced by a considerable proportion.

This matter is rarely included in considerations during the compressor selection phase. The discharge check valve specification is typically determined by piping design, with the piping engineer selecting from the standard valve list in the piping specification, without having discussed with the rotating equipment engineer "what dynamic characteristics does the check valve at this location need."

Screw compressor pulsation

Screw compressors do not surge; this premise must be established first. API 619 is the standard for screw compressors, and the word surge does not appear in it. The working principle of positive displacement machinery means that as long as the rotors are turning, gas is forced forward.

Screw compressor pulsation originates from the discrete nature of the compression process. Each revolution of the male rotor produces several discharge events (equal to the number of male rotor lobes), and each discharge event is a pressure pulse. A machine with a 4-lobe male rotor at 3000 rpm has a fundamental frequency of 4×50=200 Hz, plus harmonics. Atlas Copco's ZR series, Kobelco's KNW series, and GHH Rand machines all operate on this basic principle; the differences lie in rotor profiles, Vi design, and noise reduction measures.

API 618 is the standard for pulsation and vibration control of reciprocating compressors and contains a complete simulation and analysis methodology. Screw compressors have no corresponding API standard that governs discharge pulsation analysis and control; API 619's coverage in this area is thin. Engineering practice typically follows the approach of API 618 for acoustic analysis, but 618 was written for the discrete pulse characteristics of reciprocating machines. Screw compressor pulse waveforms and frequency ranges are different, and direct application requires attention to applicability.

Built-in volume ratio mismatch

This topic has been chronically underappreciated in screw compressor operational practice. It is simultaneously the root cause of pulsation deterioration, efficiency loss, and noise increase.

Every screw compressor has a built-in volume ratio Vi determined by the rotor profile and discharge port geometry. Vi multiplied by the pressure at suction closure, then converted through an isentropic process (PV^k = constant, where k is the isentropic exponent), yields the pressure that the gas in the tooth space "should have" at the moment the discharge port opens. If this pressure equals the pressure in the discharge piping, the moment the discharge port opens the pressure inside and outside the tooth space is balanced, and the gas flows out smoothly.

Industrial compressor room with piping systems — Volume ratio mismatch silently wastes energy without triggering any alarm

Under-compression: the system discharge pressure is higher than the pressure inside the tooth space. The instant the discharge port opens, high-pressure gas from the piping rushes into the tooth space. This is a transient process that happens very fast, producing a steep pressure-rise pulse superimposed on the normal lobe-passing pulsation. On the frequency spectrum, the lobe passing frequency harmonic family increases in amplitude, and at the same time the noise floor between harmonics also rises, because the backflow process has an impulsive character with spectral content extending beyond just the harmonic frequencies. Efficiency loss comes from two directions: recompressing the backflow gas consumes power, and the turbulence induced by the backflow increases internal friction losses. Discharge temperature runs high.

Over-compression: the system discharge pressure is lower than the pressure inside the tooth space. The gas in the tooth space has been compressed beyond what is needed, and when the discharge port opens it expands outward. The impulsive character of this process is somewhat milder than under-compression. The efficiency loss direction is different: extra compression work was performed for nothing, and the excess pressure dissipates as turbulence and noise upon expansion at the discharge port. The discharge temperature response is not as pronounced as in the under-compression case, because the temperature rise from over-compression and the temperature drop from expansion partially cancel out.

Both types of mismatch carry an efficiency penalty, with under-compression carrying the heavier penalty. Atlas Copco provides in its technical manual typical curves showing the effect of Vi mismatch on specific power; at 15% deviation from the design pressure ratio, specific power degradation is roughly in the range of five to eight percentage points, with exact values depending on the machine model and rotor profile.

This energy waste will not trigger any alarm on any single machine. Specific power (kW per m³/min) is simply not a tracked parameter in the monitoring systems of the vast majority of plants. Operators watch discharge pressure, discharge temperature, and motor current, all within normal ranges, with no indication that the compressor is operating below its proper efficiency level. Taking actual numbers from a compressor station: six screw air compressors each rated at 160 kW, assuming an average efficiency loss of six percentage points (a conservative estimate for fixed-Vi machines running long-term at off-design pressure ratios), 7200 operating hours per year (three-shift operation with maintenance downtime allowance), local industrial electricity price 0.085 euros per kWh. One machine wastes an extra 160×0.06×7200×0.085 ≈ 5875 euros per year. Six machines comes to roughly thirty-five thousand. This number does not appear in the maintenance budget because nothing is broken, nor does it appear in fault records because there are no alarms. It just flows into the total figure on the electricity bill year after year.

Fixed-Vi machines have no way to self-adapt when system pressure fluctuates. Variable-Vi slide valve machines can theoretically track pressure changes and adjust Vi. Here is the complication: many machine models (such as certain SRM-profile twin-screw machines) share a single slide valve for both capacity regulation and Vi adjustment. The slide valve position simultaneously determines two things: the effective compression stroke length (controlling capacity) and the discharge port opening angle (controlling Vi). When downstream demand changes require a capacity adjustment, the slide valve moves to the position that satisfies the capacity requirement, and that position's corresponding Vi may not be the optimum value for the current pressure ratio. When the controller faces two conflicting objectives, discharge pressure control takes priority over Vi optimization (because discharge pressure directly affects supply pressure to the pipe network), and Vi gets sacrificed. Some manufacturers (such as GHH) use a dual slide valve design that decouples these two control degrees of freedom, allowing capacity and Vi to be adjusted independently. It costs more.

Acoustic resonance amplifying discharge pulsation

Discharge piping has its own acoustic natural frequencies, determined by the pipe geometry and the speed of sound in the gas inside. Speed of sound depends on temperature and gas composition (for an ideal gas, a = √(kRT/M), where k is the isentropic exponent, R is the universal gas constant, T is absolute temperature, and M is molecular weight).

If one of the pipe's natural frequencies coincides with the lobe passing frequency or one of its harmonics, that frequency component of the pulsation is amplified by resonance, and radiated noise increases dramatically.

Industrial equipment maintenance and inspection — Seasonal temperature changes shift acoustic resonance conditions

Speed of sound changes with temperature. As ambient temperature goes from 5°C in winter to 35°C in summer, the speed of sound in air goes from 334 m/s to 352 m/s, a change of over 5%. Pipe natural frequencies are proportional to the speed of sound and shift accordingly. A pipe natural frequency that in winter misses the lobe passing frequency by 8 Hz might land right on it in summer. This is why some screw compressor stations have noise complaints with a clear seasonal pattern.

If the problem is not approached from an acoustic resonance perspective, this phenomenon of "the same machine sometimes noisy sometimes quiet" is very difficult to explain. Maintenance personnel change gaskets, tighten flanges, check bearings, find nothing, and in two days the noise goes away on its own. The temperature dropped and the resonance condition was no longer satisfied.

The countermeasure: install a side branch resonator on the discharge piping (a length of closed-end branch pipe that provides quarter-wavelength resonant absorption at the target frequency). Burgess Manning and FläktGroup both make these products. Design requires precise target frequency and pipe acoustic impedance data. If the frequency calculation is off, it is equivalent to not having installed anything. And because the target frequency itself drifts with temperature, the resonator's effective operating range is limited. Some designs add an adjustable-length piston inside the resonator branch to accommodate frequency variation, adding mechanical complexity.

Rotor wear diagnostics

API 619 requires screw compressor performance testing to be conducted per ASME PTC 9 or ISO 1217. These standards specify how to measure input power and output flow under standardized conditions and calculate specific power. Performing periodic performance tests and comparing current specific power to the commissioning baseline is the most direct means of tracking rotor condition degradation.

Discharge temperature as a standalone indicator is unreliable. Discharge temperature is simultaneously influenced by suction temperature, suction pressure, discharge pressure, and the temperature and flow rate of cooling oil (for oil-flooded machines) or cooling water. A change in any one of these variables causes a discharge temperature change that has nothing to do with rotor condition.

The parameter with diagnostic value is discharge temperature deviation: at the current actual operating conditions (actual suction temperature, actual pressure ratio), calculate the theoretical discharge temperature using the isentropic compression formula, then add the known isentropic efficiency offset for that machine (obtained from commissioning data), yielding "what the discharge temperature should be at this operating condition with the machine in new condition." Measured value minus this theoretical value is the deviation. If the deviation is increasing month over month, and cooling system issues have been ruled out (decreased cooling oil flow, increased cooling water temperature, etc.), it points to increased internal leakage.

For air the isentropic exponent is simply taken as 1.4. For process gases (such as natural gas, refrigerants, synthesis gas) the calculation requires the actual composition run through an equation of state.

Many plants' PLCs or DCS systems only have simple temperature monitoring without an embedded gas property calculation module to perform this correction. Honeywell's Uniformance or OSIsoft PI with appropriate calculation modules can accomplish this, but someone needs to configure it.

In the frequency domain there is another indicator: the amplitude distribution of the lobe passing frequency harmonic series. A new rotor produces a discharge pressure waveform close to a trapezoidal wave, and the FFT harmonics decay in a predictable pattern. After rotor wear, internal leakage disrupts the pressure buildup process within the tooth space, certain harmonics develop abnormally high relative amplitudes, and the "shape" of the harmonic distribution changes. The prerequisite is that baseline FFT data was collected during commissioning for comparison.

Liquid slugging

The impact signature when a screw compressor ingests liquid looks completely different from normal pulsation on a time-domain waveform. Normal pulsation is a periodic signal with good repeatability. Liquid slugging appears as randomly occurring isolated spikes, manifesting on the frequency spectrum as broadband energy elevation rather than discrete harmonics.

Warning signals are faint. Discharge temperature during a liquid slug event exhibits a rapid fluctuation pattern of first dropping (liquid evaporation absorbing heat) then rising (impact work generating heat), lasting a short time, requiring at least one-second sampling intervals to see the sawtooth pattern on a trend. Most plants' temperature acquisition cycles are ten or thirty seconds, too slow to capture it. Drive motor current also shows millisecond-duration spikes, and PLC scan cycles are likewise too slow.

Brüel & Kjær or Emerson online vibration monitoring systems, if configured with envelope analysis and peak detection capabilities, can theoretically extract the high-frequency impact components generated by liquid slugging from the casing vibration signal. This requires specifically setting parameters and alarm thresholds for liquid slug detection; factory default configurations most likely do not include this.

The reality is that most screw compressor liquid slugging events are only discovered after visible damage has occurred. The focus of prevention is on source control: whether the suction piping liquid separator design is adequate (API 619 Section 3.4 has requirements), whether drain traps are functioning properly, whether oil separator maintenance intervals are reasonable for oil-flooded machines, and whether operating conditions might produce condensation inside the suction piping (suction temperature below dew point).

Coupled surge between parallel centrifugal compressors through a common discharge header

The configuration of two or more centrifugal compressors operating in parallel sharing a common discharge header is very common in refineries, natural gas processing plants, and large air separation plants. API 617 Section 2.7 has some general requirements for parallel operation but does not provide detailed specifications for anti-surge control coordination.

Parallel compressor piping and instrumentation — Parallel compressor coupling through shared discharge headers creates system-level surge dynamics

Machine A approaches the surge boundary, and the CCC controller opens Machine A's recycle valve. Machine A's effective discharge flow drops, and it no longer contributes as much gas to the header. If total downstream demand has not changed, the entire shortfall is pushed onto Machine B. If Machine B is already running at high load, it cannot absorb the entire shortfall. Header pressure rises, Machine B's discharge pressure increases, and Machine B's operating point climbs along the constant-speed line toward the surge line. Machine B's controller also begins opening Machine B's recycle valve.

Now both machines have their recycle valves open, neither is contributing effective flow to the header, downstream pressure drops, both recycle valves close back, both machines reload, pressure rises again, both approach the boundary again. When the amplitude is large enough, one or both machines surge and trip.

The trip record reads "Machine A surge trip at 14:37, Machine B surge trip at 14:38." Root cause analysis is conducted separately for each. Neither machine shows a single-machine problem. The problem is in the coupling.

CCC's S2 controller supports "load sharing" and "load balancing" functions to coordinate parallel machines, and Compressor Controls Corporation has discussed this scenario in technical bulletins. Honeywell and Yokogawa solutions also have corresponding parallel coordination modules. The key is to treat parallel operation as a system when designing anti-surge protection, rather than protecting N independent machines separately. Many installations did not incorporate the dynamic coupling of parallel machines into the control philosophy document during the initial FEED phase. When this problem is discovered after startup, the cost of remediation is higher than if it had been designed in from the beginning.

Series multi-stage compression systems have similar inter-stage coupling. A first-stage anti-surge valve action changes the second stage's suction pressure and flow, and the second stage's operating point drifts accordingly. The dynamics of inter-stage coupling are more complex than parallel coupling because the disturbance propagates unidirectionally (from the upstream stage to the downstream stage). CCC uses so-called "decoupling" logic in its multi-stage anti-surge solutions, performing real-time calculation of inter-stage influence quantities within the controller and applying feedforward compensation.

The respective roles of dynamic pressure measurement and vibration measurement

API 670 is the core standard for rotating machinery condition monitoring, covering proximity vibration probes, accelerometers, axial displacement sensors, and temperature sensors. This system has strong diagnostic capability for mechanical faults: imbalance, misalignment, bearing damage, seal rubbing, rotor cracks all have signatures in vibration signals.

Surge is an aerodynamic phenomenon. Its primary signal is oscillation of pressure and flow, not vibration. Vibration is a secondary signal resulting from pressure oscillations acting on the rotor and casing through aerodynamic forces, having passed through a conversion from aerodynamic to structural mechanics, with information loss. A dynamic pressure transducer installed near the discharge flange (select from PCB, Kistler, or Kulite products based on range and temperature suitability), sampling at 2 kHz or above, yields a pressure waveform that can directly answer: is there surge or incipient surge? What is the frequency? What is the amplitude? Is it surge or rotating stall? Is there acoustic resonance amplifying a particular frequency?

API 670 does not list dynamic pressure measurement as a standard configuration. Most installations have centrifugal compressors equipped with a full set of vibration monitoring as required by 670, with no permanently installed dynamic pressure transducers on the discharge piping. When needed, one is temporarily installed for a diagnostic test.

The value of vibration data in surge matters lies after the fact rather than during the event. After a surge trip, check whether the axial displacement trend shows a step change (thrust bearing clearance may have suddenly increased), check whether the vibration spectrum shows new components (contact excitation from seal rubbing), and decide whether a shutdown and case-opening inspection is needed. Bently Nevada (now part of Baker Hughes) System 1 platform is quite good at trend analysis and spectrum comparison, though a "pre-surge versus post-surge data comparison" analysis template needs to be set up manually and is not included in the factory default configuration.

Conventional process pressure transmitters (such as Rosemount 3051) typically have a sampling rate of one per second or lower, whether analog output or HART/Foundation Fieldbus. They have absolutely no resolving power for the dynamic characteristics of surge and pulsation. These are process instruments, not dynamic diagnostic instruments.

Real-time operating point display on the compressor map in the DCS

There is one thing that involves no hardware modification whatsoever, purely DCS screen development and calculation configuration work. Project each centrifugal compressor's current operating point in real time onto the compressor map for the operator to see, mark its distance from the surge line, color-coded.

Compressor control and monitoring systems — Real-time map visualization closes the gap between individual parameter alarms and system awareness

What operators normally see is the individual numbers for discharge pressure, suction pressure, flow, and temperature, each with its own alarm limit. When all numbers are within limits, the screen is all green and everything appears normal. The operating point may be only 3% margin from the surge line, and no single parameter will reach its alarm because the alarms are set on absolute values, not on relative position on the compressor map.

Real-time operating point visualization closes this gap. What the operator sees is a point on a chart, its position and its direction of movement. When operating conditions push the point toward the surge line, the color changes from green to yellow to red, and the operator can see the trend and intervene manually before the anti-surge controller kicks in.

Implementing this function requires three things: digitized OEM performance curve data (request from the OEM or manually digitize from paper curves), a calculation module configured in the DCS to perform suction condition correction (converting actual suction temperature, pressure, and gas molecular weight to standard conditions, then locating the operating point position on the map), and DCS screen development. Honeywell Experion, Yokogawa CENTUM VP, and Emerson DeltaV all support this type of custom calculation and display. The workload is not large.