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Scientific Reports volume 14, Article number: 29538 (2024 ) Cite this article Environmental Water Jet Cleaning
High vibration poses significant safety and operational risks to compressor components. This study presents a thorough investigation of the root causes analysis and corrective actions to resolve vibration-induced failures faced on the suction system of a reciprocating compressor. Excessive pressure pulsations in the suction bottle and strong mechanical coupling between the cylinders and suction bottle due to high cylinder gas forces were identified as the root causes of vibration-induced failure of the suction system. The proposed solutions to eliminate this problem were implemented by splitting the suction bottle into two parts and making one bottle as an acoustic filter (Helmholtz resonator). The acoustic filter was optimized for the suppression of pressure pulsations above its Helmholtz frequency and also for the avoidance of acoustic and mechanical resonance. After applying the proposed solutions, the modified suction system achieved a 14.3-61.1% reduction in vibration and a 75.56% reduction in peak-to-peak pressure pulsation for the compressor operating at 1000 rpm. The processing capacity of the compressor has been increased by 6.82%. An acoustic filter is an effective low-cost technical measure for reduction of pressure pulsation and fluid-induced vibration, which greatly increases the reliability of the entire compressor system.
Underground gas storage (UGS) is principally used in response to demand variations and plays an important role in maintaining stability in gas networks to safely compensate for consumption peaks in winter. Storage compressors are used to inject gas into an underground gas storage field during the injection season, and gas is withdrawn from the storage field to the pipeline during the winter withdrawal period1. In general, large, double-acting, multi-stage reciprocating compressors are heavily used because of wide variation in storage conditions.
Owing to its inherent design, in a reciprocating compressor, pressure and volume fluctuations are unavoidable because the suction and discharge are not continuous. The resulting pressure pulsations contain many harmonics of the rotational speed2. Most of the vibration problems in reciprocating compressors result from high-pressure pulsations because pulsations always generate periodic dynamic forces to shake the compressor assembly, including the vessels, valves, and connected pipeline systems3,4. Serious vibrations in compressor systems can lead to a range of issues, such as loss of production efficiency, malfunction of instruments, cracks in pipelines, structural fatigue failure, and unscheduled system shutdowns, which can seriously affect the safety and reliability of reciprocating compressor systems.
Pulsation studies and vibration analysis focused on the following areas: vibration and pulsation theories, vibration diagnosis techniques, and vibration and pulsation control methods5,6,7. Vibration and pulsation theories include the construction of vibration and acoustic mathematical models; analytical and numerical solutions of the models; modal analysis to avoid mechanical resonance; pulsation and acoustic analysis via the transfer matrix method and finite element methods. Numerical simulations via finite element methods, such as ANSYS software, have become powerful and economical methods for analyzing pulsation characteristics and vibration behavior. Vibration diagnosis techniques focus mainly on measurement instruments and methods, the time domain, and spectrum analysis of test signals. Table 1 outlines various measures to reduce pulsation and vibration.
The abovementioned studies have provided a good basis and guide for vibration diagnosis and elimination. However, vibration control remains challenging for high-power reciprocating compressors, as compressors with high power tend to vibrate more. To the author’s knowledge, most reciprocating compressors for UGS have a power of 750 to 4500 kW. It is important to analyze and control the vibration of the reciprocating compressor system to ensure its safety. In this study we provided a methodology for troubleshooting excessive vibration in existing reciprocating compressor suction systems in UGS, identified the root causes of vibration and applied effective corrective actions to resolve these vibration problems.
A six-cylinder reciprocating compressor with a design speed of 1000 r/min and a rated power of 4000 kW was employed to pressurize natural gas at a UGS station21. As shown in Fig. 1, the reciprocating compressor has two stages; the suction and discharge bottles ensure stable working conditions in the compressor cylinders and reduce pulsations and vibrations.
Six-cylinder reciprocating compressor with suction and discharge bottles.
The compressor had been in operation for nearly 10 years at the time of this study. Severe vibrations were observed in the first-stage suction system under normal operating conditions. Notably, the first-stage suction system includes a suction bottle and related suction piping from three cylinders #2, #4 and #6 upstream of the inlet scrubber. The maximum vibration velocity of 40 mm/s was recorded for the first-stage suction bottle, which greatly exceeded the allowable limit of 18 mm/s specified in the ISO 20816-1 standard22. Moreover, the bolt connected to cylinder #4 failed multiple times, and a vibration-induced fatigue crack appeared on the second nozzle of the suction bottle near cylinder #4, as shown in Fig. 2.
Bolt fracture and fatigue cracking in a suction bottle nozzle.
As reciprocating compressors suffered from safety and reliability issues, vibration and pressure tests were carried out without stopping the compressor unit to uncover the root cause of the vibration problems by analyzing the dynamic response and frequency characteristics of the vibration source. The dynamic forces in the compressor cylinders were investigated to evaluate the mechanical coupling between the cylinders and the suction bottle. The solutions to suppress vibrations and pressure pulsations were proposed by splitting the suction bottle into two parts. A novel acoustic filter with a baffle and an internal choke tube for pressure pulsation suppression was proposed, analyzed and applied.
Field measurements were used to investigate the excitation sources and response of the suction system. The measuring parameters included vibration and pressure, which were tested by means of the Leonova Infinity (see Fig. 3). Leonova Infinity is a multi-function, hand-held data logger, which can be operated via keypad and touchscreen23. The measurement sensors included accelerometers for vibration testing and pressure transmitters for pressure testing, the main specifications of which are given in Table 2. The signals from these sensors are sent through a signal conditioner to the analysis software via a USB cable for further processing, as shown in Fig. 3. Three vibration test points, V1-V3, and one pressure test point, the P1, are selected on the 1st stage suction bottle, as presented in Fig. 1. Each vibration test point is characterized in three directions: the horizontal direction is perpendicular to the compressor crankshaft, the vertical direction is coincident with gravity, and the axial direction is parallel to the compressor crankshaft.
On-site measurements for vibration diagnosis.
A vibration test is an effective way of identifying the dynamic characteristics of a system and diagnosing vibration faults without stopping the compressor unit. The vibration severity is sensitive to the running speed and load conditions of the compressor. Therefore, the reciprocating compressor was set to run at various speeds during the measurements, namely, 900 rpm, 950 rpm and 1000 rpm, and the vibration signals at points V1-V3 were recorded under two different kinds of operating conditions: (1) idle conditions and (2) load conditions. Table 3 lists the compressor operating parameters under three load conditions.
Figure 4 shows the comparison of the vibration velocity at points V1-V3 in the horizontal direction under idle and load conditions. The vibration velocity levels of points V1-V3 are above 18.0 mm/s at a speed of 1000 rpm under load conditions, which is considered an alarm limit per the ISO 20816-1 standard24. Therefore, the reciprocating compressor must be restricted to operation below 1000 rpm due to safety concerns. As shown in Fig. 4, the vibration velocity values at points V1-V3 are low and increase slightly with increasing compressor speed under idle conditions, indicating that the excitation frequencies generated by the compressor piston reciprocating action contribute little to vibration under idle conditions; however, the vibration velocity values at points V1-V3 are high and increase significantly with compressor speed under load conditions, indicating that the shaking forces due to pressure pulsations related to compressor speed have a significant effect on vibration under load conditions. As the vibration severity was weak under idle conditions but strong under load conditions at a certain compressor speed, the preliminary conclusion was that such vibration problems were caused by excessive shaking forces due to pressure pulsations acting on the suction bottle, rather than mechanical resonance.
Vibration velocity of points V1-V3 in the horizontal direction under idle and load conditions.
Vibration frequency spectrum analysis was carried out to detect the frequency and related peak values. The spectra of the vibration signals from points V1-V3 were analyzed to determine the dominant frequency components. Owing to the similarity of the vibration spectrum characteristics of points V1-V3, only the vibration velocity spectra of point V2 are plotted in Fig. 5. In vibration spectra, the amplitude indicates the severity, the frequency indicates the source, and the dominant and problematic frequencies that contribute most to vibration problems can be clearly investigated. Figure 5 shows obvious spikes, and two predominant vibration peaks are observed at frequencies of 29.583 Hz and 89.402 Hz at speeds of 900 rpm, 31.129 Hz and 94.078 Hz at 950 rpm, and 33.153 Hz and 99.286 Hz at 1000 rpm.
Vibration velocity spectra of test point V2 in the horizontal direction under load conditions.
The intermittent suction/discharge flow in a reciprocating compressor generates intermittent periodic shaking forces that exhibit distinct amplitude and excitation frequency characteristics. The excitation frequencies corresponding to the compressor running speed are calculated via the following equation
where f is the excitation frequency related to compressor (Hz), i is the order of the harmonic, e.g. 1, 2, 3, and so on, N is rotational speed of the compressor (rpm, revolutions per minute), X dependents on piston type (single-acting or double-acting) and the crankshaft throw.
As the piston of the reciprocating compressor in the paper was operated as a double-acting type, the fundamental frequency occurred at a 2x compressor running speed (2nd harmonic). Moreover, the three cylinders #2, #4 and #6 were connected to the first-stage suction bottle (Fig. 1), and the dominant excitation frequency of the suction bottle was three times the cylinder fundamental frequency. Figure 5 shows that the frequencies of 29.583 Hz and 89.402 Hz correspond to 2x and 6x compressor speeds of 900 rpm (or 15 Hz), respectively; 31.129 Hz and 94.078 Hz are associated with 2x and 6x compressor speeds of 950 rpm (or 15.833 Hz); and 33.153 Hz and 99.286 Hz are related to 2x and 6x compressor speeds of 1000 rpm (or 16.667 Hz), respectively. In summary, all the dominant vibration peaks are located at two and six times the compressor speed (2nd and 6th harmonics).
Pressure pulsations in compressor systems can produce high periodic dynamic shaking forces on the system, which can cause excessive mechanical vibrations. Pressure measurements and analysis were performed at test point P1 (see Fig. 1) to determine the pressure pulsation characteristics. Figure 6 represents the pressure results of test point P1 at a speed of 1000 rpm. The peak-to-peak pressure amplitude is 0.5359 MPa, which represents approximately 6.78% of the average pressure of 7.91 MPa in Fig. 6 (a). Several distinct pulsation peaks are observed at the harmonics of the compressor speed of 1000 rpm (or 16.667 Hz), 0.152 MPa PK-PK is observed at 33.1667 Hz (2x compressor speed), and 0.125 MPa PK-PK is observed at 99.267 Hz (6x compressor speed) in Fig. 6 (b). Hence, the dominant frequencies were related to the 2x and 6x compressor running speeds.
Pressure results of test point P1 at a speed of 1000 rpm: (a) pressure‒time curve and (b) frequency spectrum.
The allowable pressure pulsation level for compressor parts, such as compressor valves, pulsation vessels, surge volumes, and piping systems, is specified in the API 618 standard, which can be calculated from the following equation25.
where Pall is the allowable pulsation level (%), a is the sound speed of gas in system (m/s), Pmean is the mean static pressure (bar), D is the inner diameter (mm).
As the gas pulsation waves travel at the speed of sound in a system, which speed can be obtained from Eq. (3).
where kv is the ratio of specific heats of gas, Z is the compressibility factor of gas, T is the absolute temperature (K), Rg is the gas constant [J/(kg·K)].
On the basis of the structural dimensions and physical properties of the natural gas within (see Table 4), the allowable pulsation limit for the 1st stage suction bottle was determined by Eqs. (2) and (3). As shown in Fig. 6 (b), the black line represents the allowable pulsation limit for the 1st stage suction bottle. All the PK-PK pressure pulsation spikes at the harmonics of the compressor speed are clearly well above the API 618 limit. Therefore, the vibration sources are the excessive shaking forces induced by high-pressure pulsations within the suction system due to the intermittent suction flow in the reciprocating compressor.
In addition to the pulsation shaking forces induced by pressure pulsations, other dynamic forces in reciprocating compressors can cause serious vibration problems. These dynamic forces result from the inertial forces of the piston and other reciprocating and rotating components and the gas forces induced by the gas compression loads inside the cylinder. Figure 7 shows a schematic of the double-acting compressor cylinder, it can compress gas on both the head end and crank end of the cylinder.
Schematic of the double-acting reciprocating compressor cylinder.
The inertial forces consist of reciprocating inertia forces and rotational inertia forces: the reciprocating inertia forces result from the acceleration of the reciprocating mass; the rotational inertia forces are caused mainly by the dynamic imbalance of the crankshaft, which can be offset by adding a counterweight. Therefore, the inertial forces can be calculated by Eq. (4)26.
where Fi is the inertial forces (N), mrecip is the mass of reciprocating components (kg), R is the crank radius (m), ω is the angular velocity of crankshaft (rad/s), θ is the rotating angle of crankshaft (°), L is the connecting rod length(m).
The gas in cylinder is compressed first on the head end and then on the crank end of the piston. The crank and head end pressures in the cylinder create alternating forces that act on the compressor cylinder assembly. The gas forces result from the differential pressures acting on the head end and crank end piston areas during each stroke, which can be obtained by Eq. (5). The gas forces can be a significant source of excitation and lead to vibration problems in the bottles and piping system close to the compressor cylinder.
where Fg is the gas forces in cylinder(N), PCE and PHE are the cylinder pressure on the crank end and the cylinder pressure on the head end (Pa) respectively, ACE and AHE are the internal cross-sectional area of cylinder on the crank end and the internal cross-sectional area of cylinder on the head end (m2) respectively.
In a multi-cylinder compressor, the gas forces differ in phase from cylinder to cylinder. Figure 8 shows the crankshaft throw arrangements of the six-cylinder reciprocating compressor. The six throws are arranged in three planes, and the phases between cylinders #2, #4 and #6 are 0, 120 and 240 degrees. As the three cylinders #2, #4 and #6 are connected to the 1st stage suction bottle (see Fig. 1), the three nozzles of the suction bottle can create strong mechanical coupling between the cylinders and the suction bottle. As a result, the 1st stage suction bottle suffers from differential motion along its length, which could cause high stresses in the bottle nozzles, leading to fatigue failure.
Crankshaft throw arrangement of the six-cylinder reciprocating compressor.
Using Eqs. (4) and (5), the inertial forces and gas forces acting on the 1st - and 2nd -stage cylinders were calculated for one crankshaft revolution on the basis of the compressor operating conditions and structural dimensions. Figure 9 shows the results of the gas forces and inertial forces in the cylinders versus crank angle. As the crankshaft rotates one revolution, both the gas forces and inertia forces vary from minimum to maximum values. The maximum gas force in the 1st -stage cylinders is 318.59 kN, which is greater than that in the 2nd -stage cylinders. Excessive gas forces are produced in cylinders #2, #4 and #6, and strong mechanical coupling is created between the cylinders and the suction bottle. It can be concluded that the underlying causes of the suction bottle fatigue crack and bolt fracture (see Fig. 2) are the high gas forces in the compressor cylinders and excessive shaking forces in the suction bottle due to pressure pulsations.
Compressor cylinder dynamic forces versus crank angle: (a) in the 1st -stage cylinders #2, #4, and #6; (b) in the 2nd -stage cylinders #1, #3, and #5.
As analyzed above, unacceptable pressure pulsations occurred in the 1st stage suction bottle because of the poor design of the suction bottle; moreover, the strong mechanical coupling between the cylinders and the suction bottle exacerbated vibration problems. It was difficult to balance the gas forces acting on the three nozzles of the suction bottle because the 1st stage suction bottle was shared by the three cylinders #2, #4 and #6. To eliminate vibration and pressure pulsations, the proposed solutions should focus on suppressing pressure pulsations in the suction bottle and isolating the mechanical coupling between the cylinders and the suction bottle. Therefore, the original 1st -stage suction bottle was split into two parts to minimize mechanical coupling due to the high gas forces in the cylinders, as shown in Fig. 10 (a). Figure 10 (b) shows the two pulsation bottles designed for vibration elimination: the small one mounted on cylinder #2 was designed as an ordinary pulsation bottle, the large one mounted on cylinders #4 and #6 was designed as an acoustic filter, and the suction piping leaving the scrubber was in line with the two suction bottles.
Layout of the 1st -stage suction system mounted on the reciprocating compressor: (a) original structure and (b) two suction bottles designed for vibration elimination.
Acoustic filter design was the key component in successful vibration control. Figure 11 shows the acoustic filter with a baffle and an internal choke tube. The baffle and choke tubes are used to create separate volume chambers for each cylinder. The acoustic filter has its own acoustic resonant frequency (or Helmholtz frequency), which can be calculated by Eq. (6)27.
where fH is the Helmholtz frequency of the acoustic filter (Hz), Ac is the internal cross-sectional area of choke tube (m2), Lc is the length of choke tube (m), Lc’ is the equivalent length of choke tube (m), Lc’=Lc+0.6dc, V1 and V2 are the volumes of chambers 1 and 2 (m3) respectively.
Structure of the acoustic filter.
The acoustic filter passes pulsations at frequencies below the Helmholtz frequency and filters pulsations at frequencies above the Helmholtz frequency. The typical frequency response of the acoustic filter to pulsation frequencies generated by the compressor is displayed in Fig. 12, which reveals the relationship between the Helmholtz frequency and the harmonics of the compressor speed. As the Helmholtz frequency is located between the first- and second-order compressor speeds, the pulsations at 2x and higher frequencies can be attenuated (see Fig. 12).
Typical frequency response to pulsation excitation of the acoustic filter.
The Helmholtz frequency can be tuned by altering the chamber volume, the diameter and length of choke tube, as given in Eq. (6). The influence of the tube diameter dc and the chamber volume V on the Helmholtz frequency fH was investigated by assuming that the acoustic filter was symmetrical (V1 = V2), the total volume V = V1 + V2, and the length of the choke tube was equal to the distance between the two nozzles of the acoustic filter. As illustrated in Fig. 13, the Helmholtz frequency fH of the acoustic filter increases with increasing tube diameter dc but decreases with increasing damper volume V.
The Helmholtz frequency of the acoustic filter should avoid coinciding with any harmonic of the compressor speed to avoid acoustic resonance. Moreover, for a double-acting compressor, the preferred Helmholtz frequency should have a 20% separation margin from the 1x and 2x compressor speeds to suppress high-frequency pulsations. As the reciprocating compressor operating from 900 to 1050 rpm generates harmonics of 30–35 Hz at its 2x running speed, the Helmholtz frequency of the acoustic filter should be below 24 Hz. The black dashed line in Fig. 13 represents the maximum Helmholtz frequency of 24 Hz for the acoustic filter. On the basis of the design point shown in Fig. 13, the acoustic filter is sized with a choke tube diameter of 0.15 m, a total volume of 0.384 m2, and a Helmholtz frequency of 23.67 Hz to eliminate pressure pulsations generated by a compressor operating at speeds of 900 rpm to 1050 rpm.
Helmholtz frequency of the acoustic filter at different choke tube diameters and chamber volumes.
Besides, excessive vibration could occur due to poor structural rigidity and mechanical resonance. The suction system in Fig. 10 (b) should be restrained by additional supports. Therefore, several supports were used to improve the stiffness of the suction system: pipe clamps were used to limit suction piping vibration, sway struts were used to restrain the movement of the two suction bottles, and both the clamps and braces were located where the vibration levels were expected to be high.
Mechanical resonance analysis was performed by the finite element method to obtain the various mechanical natural frequencies (MNFs) and mode shapes of the modified suction system to avoid mechanical resonance. Figure 10 (b) shows the boundary conditions after installation of the clamps and struts: the three flange ends are defined as fixed support; both the pipe clamps and sway struts are defined as elastic support with a certain stiffness; all structural components are modeled as isotropic, homogeneous and linear elastic material with a Poisson’s ratio v = 0.3, a Young’s modulus E = 210 GPa and a density ρ = 7850 kg/m3.
Figure 14 shows the effect of the stiffness of the elastic support on the first three MNFs of the modified suction system, which indicates that the first three MNFs increase slowly at the beginning, then increase rapidly, and finally level off as the stiffness increases. In general, only the first few MNFs of the system have a significant effect on vibration for practical applications; moreover, the first MNF of the system should be designed to be greater than 2.4 times the fundamental frequency (1x) corresponding to the compressor speed. Hence, the minimum value of the first MNF of the modified suction system should be 42 Hz, as the fundamental frequency range is 15–17.5 Hz for a compressor operating from 900 rpm to 1050 rpm, which indicates that the stiffness of the elastic supports should be greater than 1.6E + 4 N/mm (see Fig. 14).
Effect of the stiffness of elastic supports on the MNFs of the modified suction system.
Figure 15 represents the first three mode shapes of the modified suction system when the stiffness of the elastic support is set to 1.6E + 4 N/mm. The first vibration mode occurs at 43.95 Hz, the second mode occurs at 71.21 Hz, and the third mode occurs at 86.64 Hz. The areas highlighted in red indicate the weak areas of large relative displacement amplitude, which means that the suction piping and small pulsation bottle would reach high vibration levels if resonance occurred.
Mode shapes of the modified suction system with an elastic support stiffness of 1.6E + 4 N/mm.
The relationships between the MNFs of the modified suction system and the harmonics of the compressor speed are illustrated in Fig. 16. The six rectangles in green represent the first six harmonics for a compressor operating from 900 rpm to 1050 rpm, and the three blue dashed vertical lines from left to right represent the first to third mode MNFs of the modified suction system. The first mode and second mode MNFs have a separation margin from the first- to sixth-order harmonics; only the third mode MNF is within the fifth-order harmonic range, and all MNFs are separated from the significant excitation harmonics (the first- and second-order harmonics). In conclusion, there is no risk of mechanical resonance or acoustic resonance for the modified suction system.
Overlay of the MNFs of the modified suction system and the harmonics of speed from 900 to 1050 rpm.
As discussed above, the proposed solutions were effective in resolving the vibration problems encountered by the 1st -stage suction system. The final modifications implemented in the suction system included installing two new pulsation bottles to suppress pulsation, rearranging the suction piping layout, clamping the suction piping with two pipe clamps, and bracing the two new pulsation bottles with three sway struts. Figure 17 shows the 1st -stage suction system for the reciprocating compressor after the modifications.
Photograph of the 1st -stage suction system after the modifications.
After the modifications were implemented, new field vibration and pressure measurements were conducted to verify the vibration reduction performance under real operation conditions. The locations of the three new vibration test points V1-V3 are almost the same as before, as shown in Figs. 1 and 17. The operating parameters of the reciprocating compressor during the measurements after modifications are listed in Table 5. Notably, the reciprocating compressor was limited to operation below 1000 rpm because of safety concerns before the modifications.
Table 6 summarizes all the vibration velocity values before and after the modifications. When the compressor operates at 1000 rpm, the maximum vibration velocity obtained at point V3 in the horizontal direction decreases by 55.3%, from 33.38 mm/s to 14.91 mm/s, and the vibration levels are reduced by 14.3–61.1% after the modifications. Moreover, the maximum velocity achieved at point V3 during compressor operation at 1050 rpm is 16.61 mm/s, meeting the vibration levels specified in ISO 20816-1. As all vibration levels recorded at test points V1-V3 were within the standard limits, the operation and processing capacity of the reciprocating compressor was greatly improved after the modifications.
The velocity spectra of points V1-V3 are illustrated in Fig. 18, which shows that the vibration velocity spikes occur at every harmonic of the compressor speed; however, the dominant peaks are observed at the 1x, 2x and 4x compressor speeds (16.54 Hz, 33.06 Hz, and 66.05 Hz), and the vibration velocity spikes at other harmonics are weak.
Velocity spectra of points V1-V3 in the horizontal direction at a speed of 1000 rpm after the modifications.
The pressure signals were recorded at test point P1 (see Fig. 17) in the acoustic filter. The pressure‒time curve and frequency spectrum of point P1 at 1000 rpm after the modifications are plotted in Fig. 19, indicating that the peak-to-peak pressure amplitude is 0.131 MPa, which represents approximately 1.64% of the average pressure of 7.988 MPa. Comparing Fig. 6 (a) and Fig. 19 (a), the peak-to-peak pressure amplitude is reduced from 0.5359 MPa to 0.131 MPa, a reduction of approximately 75.56%. Figure 6 (b) and Fig. 19 (b) show that all PK-PK pressure amplitudes at the harmonics of the compressor speed sharply decrease and are within the API 618 limit after the modifications. In summary, the acoustic filter, with a Helmholtz frequency of 23.67 Hz, suppressed the pressure pulsations occurring at 2x and higher compressor speeds after the modifications.
Pressure test results for test point P1 at 1000 rpm after the modifications: (a) pressure‒time curve and (b) frequency spectrum.
Vibration problems in reciprocating compressors are inevitable due to pressure pulsations and dynamic loads. Mechanical coupling between the cylinders and all attached components due to high cylinder gas forces can exacerbate the vibration. This paper presents the measurement, diagnosis, analysis, and corrective actions of the vibration problems experienced by the suction system mounted on a reciprocating compressor. The conclusions that can be drawn from the above are as follows.
Vibration tests were carried out on a reciprocating compressor operating at 900 rpm, 950 rpm and 1000 rpm under idle and load conditions. The vibration velocity values recorded at vibration test points V1-V3 revealed that the vibration severity of the suction bottle increased with increasing compressor load, and the vibration velocity far exceeded the allowable limit of 18 mm/s according to the ISO 20816-1 standard. The pressure obtained at point P1 indicated that all PK-PK pulsation amplitudes at the harmonics of the compressor speed were greater than the limit specified in API 618. Both the vibration velocity spectra of point V2 and the pressure spectrum of point P1 indicate that the dominant frequencies occur at 2x and 6x compressor running speeds. The 1st -stage cylinders #2, #4 and #6 produced gas forces of up to 318.59 kN, which caused high stresses in the suction bottle nozzles and led to failure. It can be concluded that the vibration-induced failures resulted from excessive periodic shaking forces acting on the suction bottle due to pressure pulsations and strong mechanical coupling between the cylinders and the suction bottle due to high gas forces in the cylinders rather than mechanical resonance.
To ensure that the pulsation and vibration levels were within acceptable limits, the original 1st -stage suction bottle was split into two parts. In particular, one pulsation bottle was designed as an acoustic filter with a Helmholtz frequency of 23.67 Hz to suppress pressure pulsations generated by a reciprocating compressor operating from 900 rpm to 1050 rpm. To increase the stiffness of the suction system and avoid mechanical resonance, several elastic supports with stiffnesses greater than 1.6E + 4 N/mm were used where the vibration levels were expected to be high. There was no mechanical resonance for the compressor running speed range, as the MNFs of the modified suction system had separation margins ranging from 1x to 2x the compressor speed.
Modifications to the 1st stage suction system were implemented: two new pulsation bottles were installed, two pipe clamps were installed on the suction piping, and three sway struts were installed on the pulsation bottles to restrict vibrational motion. With the application of the acoustic filter, the proposed solutions achieved 14.3-61.1% reduction in vibration and a 75.56% reduction in peak-to-peak pressure amplitude for the compressor operating at 1000 rpm. The PK-PK pulsation amplitudes at the harmonics of the compressor running speed were suppressed to an acceptable limit according to the API 618 standard. The maximum vibration velocity amplitudes satisfied the ISO 20816-1 requirement for compressors operating at 1050 rpm. The modified suction system was able to operate at 1050 rpm, meeting all the new process requirements.
In conclusion, the sources for vibration problems experienced with the suction system were eradicated by the proper acoustic filter design and suction piping configuration; the modified suction system eliminated the latent disaster of suction bottle fracture due to long-term abnormal vibrations and greatly increases the reliability of the entire compressor system. An acoustic filter characterized by a volume-choke-volume design is an effective low-cost technical measure for pressure pulsation reduction. The methodology presented for eliminating excessive vibration induced by pressure pulsations in reciprocating compressor can be applied to solve similar problems.
Data are available from the corresponding author on reasonable request.
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This work was supported by the Sichuan Industrial Development Fund for Industrial Technology R&D and Innovation Capacity Enhancement Project (2022JB001), and the open fund (OGE202101-02) of the Ministry of Education Key Laboratory for Oil and Gas Equipment.
School of Mechanical Engineering, Xihua University, Chengdu, 610039, Sichuan, China
Shuangshuang Li & Hui Yang
College of Mechanical and Vehicle Engineering, Chongqing University, Chongqing, 400044, China
Gas Storage Management Department, PetroChina Southwest Oil and Gasfield Company, Chongqing, 401147, China
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Shuangshuang Li conducted experiments and contributed to conceptualization, methodology, data analysis and drafting the manuscript; Hui Yang analyzed the data; Aijun Yin designed the study and performed the experimental tests; Guicheng Yu supervised the work. All authors reviewed the manuscript.
The authors declare no competing interests.
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Li, S., Yang, H., Yin, A. et al. Vibration and pressure pulsation elimination in a reciprocating compressor suction system using an acoustic filter. Sci Rep 14, 29538 (2024). https://doi.org/10.1038/s41598-024-81527-3
DOI: https://doi.org/10.1038/s41598-024-81527-3
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