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Vibration analysis for reciprocating compressors


This article originally was published in ORBIT magazine, Vol. 32, No. 2, April 2012.


Vibration analysis of reciprocating machines creates some unique challenges. This article explains the reasons and gives clarity on recommended monitoring and analysis practices and tools. Years of field experience have demonstrated that techniques which may be well understood for measuring and analyzing the vibration of purely rotating machinery can produce confusing results when applied to reciprocating machinery.

Vibration associated with rotational speed is the dominant motion for most industrial rotating machines. This “synchronous” (1X) behavior allows the direct application of traditional vibration analysis concepts towards addressing common machinery malfunctions – such as rotor unbalance. The typical frequencies observed with those common rotor-related malfunctions generally occur between a quarter of running speed and twice running speed and correlate excellently with machine mechanical conditions. Consequently, principles and diagnostic methodologies for these machines are broadly accepted and harmonized within the machinery diagnostic community.

This is not quite true for reciprocating compressors. Vibration analysis of these machines creates some unique challenges; many forcing functions produce a complex vibration signature that makes any attempt of using standard analysis techniques used for rotating equipment ineffective.

Figure 1
This drawing shows typical vibration monitoring locations for a reciprocating compressor. Sensors are installed at the crosshead guides (4 red hexagons) and on the frame (4 blue diamonds) [1]

Compressor frame vibration

Vibration measured at the frame results principally from the response of the mechanical system to the forces and movements that are occurring in the machine at the normal running conditions. These include the following factors:

Gas load forces: These forces act on the piston and stationary components at 1X and at integer multiples of running speed. They are generally significant up to about 10X and in the direction of the piston rod travel. For large slow speed compressors (up to roughly 500 rpm), gas forces are typically the largest contributor to piston rod and compressor frame load.

Inertial load forces: These forces are caused by the acceleration of the reciprocating components (piston, piston rod, and crosshead). These components represent a large amount of mass to be accelerated back and forth with each stroke. Inertial loads of 400,000 Newton (~90,000 pounds) of force or more are not uncommon with very large compressors.

Reciprocating & rotating masses unbalance forces: These forces are predominant at 1X and 2X compressor speed, and are caused by asymmetrical crankshaft design and imperfect manufacturing tolerances. They are usually much smaller than inertial and gas load forces.

Gas unbalance forces: These are caused by pressure in the pulsation bottles and pulsation at the cylinder nozzle area and on piping. Allowable pulsation levels are defined in API-618. Although these pulsating forces are usually much smaller than the forces listed above, they can be destructive to piping and piping support systems if they happen to correspond to resonant frequencies for the structures.

As a consequence of these factors, the extent of vibration is inherent with the reciprocating compressor design and its response to all the applied forces and movements. This causes these machines, even when in good condition, to vibrate much more than a comparable rotating machine. The examples in Figures 2 and 3 shows that many harmonics are produced by the complex shape of the frame velocity waveform.

Figure 2
Time waveform plot of the velocity signal from a frame-mounted vibration sensor. Observe that many different frequency components are present in the signal
Figure 3
Frequency domain (spectrum) plot of velocity signal shown in Figure 2. Fast Fourier Transform (FFT) processing allows us to see the various frequency components that are included in the complex waveform

Frame vibration frequencies typically include components below 10 Hz. For this reason, a velocity transducer (with extended low frequency response) is usually better suited than an accelerometer for detecting an increase of rotation-related forces (due to gas load or inertial loads, imbalance, foundation looseness, excessive rod load, etc.). The preferred location for the frame vibration transducer is on the side of the frame oriented in the direction of piston rod travel, on the centerline of the crankshaft and at a main bearing where dynamic load is transmitted (Figure 1). Magnitude for a filtered frame velocity signal is usually low (less than 7 mm/s); however, at low frequencies, even small amplitudes of measured velocity may correspond to large amounts of displacement.

On the other hand, measuring only frame vibration can be insufficient for effective condition monitoring, as the increase in frame velocity from incipient failures developing at the running gear or cylinder assembly will be small and typically covered by the larger signal that is produced by normal machine movement. Experience has shown that by the time the malfunction has been detected by the frame velocity transducer and the compressor shutdown, major secondary damage may have already occurred because of the malfunctions. These malfunctions include liquid or debris carryover, loose piston or piston nut, loose crosshead nut, or loose cylinder liner, and typically manifest themselves as impacts transmitted at the crosshead.

Monitoring vibration & impact

Vibration transducers monitoring rotating machinery generate “stationary” signals; this means they have constant frequency content over each revolution of the rotor (Figure 4).

In contrast, vibration measurements on reciprocating compressors present both stationary and non-stationary content. In particular, the signal generated by an accelerometer placed vertically on a crosshead guide is characterized by different frequencies with different amplitudes that occur at specific points in the revolution.

Figure 5 shows a typical waveform from a crosshead accelerometer. The signal shows high amplitude, short duration impulse peaks followed by a “ring down” that occur at certain parts of each crankshaft revolution. This signal is not filtered so the transducer is picking up the widest range of frequencies (typically from 10 Hz to 30 kHz).

Figure 4
Example of stationary vibration sample taken at an electric motor bearing. The higher frequency components are typical of the characteristic vibration produced by the interaction of the rolling elements with the bearing races
Figure 5
Timebase waveform of a crosshead acceleration signal

These acceleration peaks can be referred to as responses to impulse events occurring during compressor operation (valve opening and closing, gas flow turbulence, crosshead pin shifting at load reversal, etc.). Such impulses excite the structural resonances of the machine components - resulting in high frequency free vibration and the characteristic impact/ring-down profile.

As mentioned, the main source of vibration on the compressor frame is related to periodic forces. While the overall frame vibration increase is certainly a concern, the primary interest of crosshead vibration monitoring is detecting peaks associated with structure response to impulsive events. Conditions that increase the excitation of such resonances are generated by developing faults such as fractured or loose components or excess clearance.

Loose rod nuts, loose bolts, excessive crosshead slipper clearance, worn pins as well as liquid in the process can be detected at early stages of development using crosshead impact monitoring, thus allowing appropriate countermeasures and avoiding potential catastrophic consequences.

Of all vibration measurements that can be applied to reciprocating compressors, crosshead acceleration is probably the most effective protection measurement available, if appropriately employed.

While crosshead acceleration has proven itself to be a sound measurement for detecting mechanical failures, industry has little experience in applying and analyzing it, resulting in increased risks of false or missed alarms, and poor diagnostic value from diagnostic systems. The following paragraphs describe some basic requirements for a reliable monitoring system and diagnostic software.

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Article reprinted from Orbit magazine, Vol.32, No.2,  April 2012. This is the first installment in a mini-series of Recip Tip articles by Field Application Engineer (FAE), Gaia Rossi in Orbit magazine. Copyright 2012 General Electric Company. All rights reserved.