The presented case history discusses the analysis and investigation regarding the mechanical distress and performance deterioration of a critical refrigeration compressor installed in a world-class petrochemicals complex.
The compressor is a three-stage centrifugal machine driven by a 5,000-hp rated electric motor through a speed increasing gear. The compressed refrigerant gas exiting the compressor is condensed by an overhead air cooler. The condensed refrigerant enters a receiver and is then sent to a flash drum. Vapors from the flash drum are directed to the second-stage suction of the compressor through a side-load nozzle. The liquid refrigerant in the chiller (evaporator) exchanges heat with the circulating water.
The chiller is located within five feet of the compressor suction. Its vapor section is connected to the compressors first-stage inlet flange with a short, straight-run pipe. The inlet throttle valve is a butterfly type. Fig. 1 shows details of the simplified compressor loop schematic. Similarly, the arrangement of the compressor train is illustrated in Fig. 2.
| Fig. 1. Compressor-loop scheme. |
| Fig. 2. Compressor train arrangement. |
The cooling load for this compressor fluctuates with the petrochemical plants production rate. For continuous refrigeration capability, vapor in the condenser must be condensed at the same rate as liquid refrigerant is vaporized by the evaporator. Because the condensing temperature of the air cooler depends on the ambient temperatures, the refrigeration capacity of the condenser changes with increasing and/or decreasing condensing temperatures. When the cooling water flowrate through the evaporator increases in conjunction with hot summer days, the compressor operates at peak load.
The machine is critical equipment and without a spare; thus this refrigeration compressor train is included as part of the periodic vibration analysis program conducted by the plants condition monitoring and asset reliability unit. Since the first start-up, this machine has operated normally, as evident from steady-state vibrations, bearing temperatures, thrust positions and the other measured process parameters.
First signs of failure
However, the first signs of deviation from normal operating conditions became obvious when a step rise in the compressors vibrations was observed (changing from 21 microns peak-to-peak to 26 microns peak-to-peak) with dominant peaks occurring at 1x and 2x, together with a cluster of sub-synchronous frequencies of 0.3x, 0.5x, 0.8x, 1.5x and 2.5x at the discharge end of the radial bearing.
After analyzing the vibration spectrum, these frequencies were attributed to unbalance and mechanical looseness in the rotor. The alignment of the train had been checked during a recent outage of the chiller circuit, and it was within specified tolerance. As such, angular and/or parallel misalignment was not considered to be a contributing factor for increased vibration. After experiencing signs of rotordynamic malfunction, vibrations and thrust position of the compressor were closely monitored. During this period, the axial position of the compressor rotor had changed in the active direction only by a small amount0.002 in.
In parallel to the vibration analysis, compressor performance calculations were also done using data collected from the plants distributed control system (DCS). The results were compared with performance curves supplied by the manufacturer. It was soon revealed that the calculated head and efficiency were much higher than the values at the rated point. Secondly, the flowrate associated with calculated values was far to the right on the head (H)-flowrate (Q) curve. This circumstantial evidence indicated that the compressors impellers were operating in choke. In addition, there was a possibility that liquid was present in the suction gas because the gas exiting the evaporator had no superheat. In addition, the motor current was also recorded, which confirmed that the compressor train was operating in overload condition.
The unit was shut down after four days due to increasing amplitude, reaching 31 microns peak-to-peak. The steady-state vibration level of the compressor in the preceding months was 21 microns peak-to-peak, maximum. The compressor was decoupled, and alignment readings were recorded. The compressor was disassembled for inspection, followed by rotor removal. A thorough examination of the rotor showed that two vanes in the third-stage impeller were partially detached from the cover at the weld joint. This condition explained the presence of unbalance and mechanical looseness related to peaks in the vibration spectrum.
Rotor damage. Minor circumferential abrasion scratches were noticed on thrust bearing pads and both radial bearing surfaces, a condition often caused by dirt or debris passing through the oil film. To eliminate further abrasion damage, the duplex oil filters were replaced after flushing the entire oil system. Lubricating oil was reused, as it was determined that it had not degraded.
Evaporator. Inspection of the evaporator showed no damage to the internals. However the evaporators liquid-level controller was found to be defective following a calibration check. This controller was not able to maintain the set point, thus reporting incorrect information to the plant DCS. Result: Some liquid refrigerant remained entrained in the gas leaving the evaporator. As this liquid absorbed the heat of compression and evaporated, volume flow through the compressor increased, thus overloading the last stage (third stage) of the compressor. Sustained operating of the third stage at higher than the design pressure gradient contributed to material fatigue and initiated weld cracks.
After replacing the rotor and completing other maintenance activities, the compressor was assembled. Calibration checks of instruments in the compression loop were conducted, and the defective level controller was repaired. Start-up of the compressor was normal. The compressor now operates at normal vibration levels19 microns peak-to-peak maximum since the inspection/repair. Figs. 3 and 4, respectively, show the amplitude vs. frequency plots in X -Y directions for each radial bearing before shutdown and after repairs.
| Fig. 3. Amplitude vs. frequency plot before |
| Fig. 4. Amplitude vs. frequency plot |
In summary, the elevated vibration amplitude of the compressor was an after-effect of distress in the third-stage impeller. The level-controller malfunction and wrong information collected by the plant DCS resulted in a double-jeopardy situation.
In addition, the presence of liquids affected the density of the refrigerant gas, compression ratio, volume flow and motor load. Entrained liquid in the gas lowered the compressor discharge temperature and, therefore, the calculated polytrophic efficiency was higher (five points higher than that corresponding to the rated point). Note: The discharge temperature of a refrigeration compressor is not the same as the condensing temperature. The condensing temperature is governed by the temperature of the condensing medium.
Unbalance and looseness in the third-stage impeller showed the prominent effect on the vibration level on the closest radial bearing (discharge end bearing).
As mentioned earlier in this article, the rate of vaporization in the evaporator must match the rate of condensation to provide a continuous refrigeration effect. The higher the cooling load, the greater the flow velocity in the evaporator and the higher the potential for liquid to be entrained in the gas. It was recommended to plant operations that the vapors entering the compressor suction should have a minimum of 5° superheat. On start-up after the repairs, between 5° and 7° suction superheat was being maintained by the unit operators by adjusting the liquid refrigerant level in the chiller (evaporator) for a given cooling load. HP
||Neetin Ghaisas is a technical fellow and director of design engineering in Fluor Canadas Calgary office. He holds an MS degree in mechanical engineering and is a registered practicing professional engineer in the province of Alberta, Canada. Mr. Ghaisas has more than 30 years of significant experience in the specification, selection, application and troubleshooting of rotating equipment. He is a subject matter expert for Fluor Corp.s compressors and steam turbines. In Fluors Calgary office, he serves as a group leader of rotating equipment engineers. He is a member of API subcommittee on mechanical equipment and a member of the machinery function team in Process Industry Practices. |