A vibration reduction method for a reciprocating compressor unit system

Abstract

The present invention discloses a method for reducing vibration of a reciprocating compressor unit system, belonging to the technical field of system vibration control. The present invention includes collecting the operating parameters of the reciprocating compressor unit system under operating conditions and the time domain vibration signals of each vibration measuring point in each test direction, and obtaining a corresponding frequency spectrum; establishing a mechanical vibration finite element analysis model of the system based on the design drawing information, the mass and stiffness of the equipment in the system, and obtaining the distribution of the structural natural frequency; identifying the vibration measuring points whose vibration amplitude exceeds the set limit through the frequency spectrum as vibration reduction positions; and designing and installing corresponding tuned mass dampers at each vibration reduction position according to the system natural frequency for vibration reduction. The present invention can analyze the vibration characteristics of a reciprocating compressor unit system that has been put into use during operation, propose targeted vibration reduction measures, and at the same time improve the life, safety and reliability of the reciprocating compressor unit system equipment.

Description

Technical Field

[0001] The invention belongs to the technical field of system vibration control, and in particular relates to a vibration reduction method for a reciprocating compressor unit system. Background Art

[0002] Compressor unit vibration is a major industrial challenge in the petroleum, chemical, and gas industries. Compressor unit vibration is primarily caused by equipment rotation and pulsation, with the highest levels of vibration occurring primarily at the cylinder ends, buffer tank inlets, buffer tank exhaust ports, and associated piping.

[0003] The existing solutions mainly include: modifying the pipe diameter, adding pipe supports, adding orifice plates, etc. These measures all require the modification of the compressor equipment, which has the following problems:

[0004] 1. Long transformation cycle, poor reliability, high cost, etc.

[0005] 2. If the transformation is not done properly, it will cause greater local stress in the compressor equipment, thereby aggravating the vibration.

[0006] 3. Poor repeatability.

[0007] On the other hand, while existing solutions focus on optimizing the design and improvement of equipment structure, in actual production applications, they are often unable to effectively reduce the vibration of compressor system equipment. Long-term operation of compressor system equipment under high vibration conditions will significantly reduce the equipment life, safety, and reliability. Summary of the Invention

[0008] The purpose of the present invention is to provide a vibration reduction method for a reciprocating compressor unit system, which can analyze the vibration characteristics of a reciprocating compressor unit system that has been put into use during operation, analyze the impact of vibration on the operation of the reciprocating compressor unit system, and based on this, propose targeted vibration reduction measures to prevent resonance from affecting the normal operation of the reciprocating compressor unit system, while at the same time improving the equipment life, safety and reliability of the reciprocating compressor unit system.

[0009] Specifically, the present invention provides a method for reducing vibration of a reciprocating compressor system, comprising:

[0010] Collect operating parameters under operating conditions: Collect the operating parameters of the reciprocating compressor system under operating conditions, collect the time domain vibration signals of each vibration measurement point in each test direction of the reciprocating compressor system, and obtain the corresponding spectrum diagram;

[0011] Obtaining the system's natural frequency: Based on the design drawing information, the mass and stiffness of the equipment in the system, a mechanical vibration finite element analysis model of the reciprocating compressor system in use is established to obtain the distribution of the structural natural frequency;

[0012] Determine the vibration reduction position: identify the vibration measurement point where the vibration amplitude exceeds the set limit through the spectrum diagram, and use it as the vibration reduction position P1n;

[0013] Install tuned mass dampers: At each vibration reduction position, corresponding tuned mass dampers are designed and installed according to the natural frequency of the system to reduce vibration.

[0014] Furthermore, the system natural frequency also includes the air column natural frequency.

[0015] Furthermore, before determining the vibration reduction position, the method further includes:

[0016] Establishing an airflow pulsation simulation analysis model for the reciprocating compressor unit system in operation; dividing each airflow pulsation system in the airflow pulsation simulation analysis model according to the position of the vibration measurement point, setting corresponding internal nodes, and setting boundary nodes at the boundaries of each airflow pulsation system;

[0017] The operating parameters of the reciprocating compressor system under operating conditions are used as inputs to the airflow pulsation analysis model to calculate the pulsation unbalanced forces of the internal nodes and boundary nodes; the pulsation unbalanced forces of the internal nodes and boundary nodes are used as excitation force inputs and applied to the mechanical vibration finite element analysis model to obtain the dynamic excitation force forced vibration response of the reciprocating compressor system, and the effective value of the vibration amplitude in the obtained dynamic excitation force forced vibration response is compared with a set limit value;

[0018] Determining the vibration reduction position further includes: using the vibration measurement point where the effective value of the vibration amplitude in the obtained dynamic excitation force forced vibration response exceeds the set limit as the vibration reduction position P2n; and using the union of the vibration reduction position P1n and the vibration reduction position P2n as all the vibration reduction positions that require vibration control.

[0019] Furthermore, the reciprocating compressor unit system is divided into various airflow pulsation systems in the system airflow pulsation simulation analysis model according to cylinders.

[0020] Furthermore, the step of installing the tuned mass damper includes:

[0021] In the mechanical vibration finite element analysis model, a corresponding tuned mass damper model is added at each vibration reduction position to obtain a global optimal solution of the mass, stiffness, and damping coefficient of the tuned mass damper model;

[0022] According to the global optimal solution of the mass, stiffness and damping coefficient of the tuned mass damper, the tuned mass damper at each vibration reduction position is designed and installed so that its natural frequency is equal to or close to the natural frequency of the system.

[0023] Furthermore, when obtaining the global optimal solution of the mass, stiffness, and damping coefficient of the tuned mass damper model, the parameter constraints of the tuned mass damper are defined as: mass ratio 1% to 5%, natural frequency ratio 0.95 to 1, and damping ratio 0.05 to 0.2.

[0024] Furthermore, the operating parameters include but are not limited to operating speed, gas flow, pipeline inlet pressure and exhaust pressure, pipeline inlet temperature and exhaust temperature.

[0025] Furthermore, the vibration amplitude is vibration displacement, vibration velocity and / or vibration acceleration.

[0026] Furthermore, the vibration reduction positions include the cylinder end, the buffer tank air inlet, the buffer tank exhaust port and the auxiliary pipes of the air inlet and exhaust ports.

[0027] Furthermore, the tuned mass damper includes a clamp, an elastic unit, a basic mass unit and an adjustment mass unit;

[0028] The clamp is tightly clamped on the outer circumference of the inlet and outlet gas buffer tank or the cylinder end of the compressor unit. The clamp is composed of several arc segments, and each arc segment is connected to form a ring. Several elastic units arranged in an array in a centrally symmetrical distribution are fixedly installed on the outer side of the circumferential surface of the arc segment of the clamp, and the basic mass unit is fixedly installed on the outer side of each elastic unit; one or several adjustment mass units are respectively fixed on each basic mass unit.

[0029] The beneficial effects of the reciprocating compressor system vibration reduction method of the present invention are as follows:

[0030] The reciprocating compressor unit system vibration reduction method of the present invention collects operating parameters of the reciprocating compressor unit system under operating conditions, collects time domain vibration signals of each vibration measuring point of the reciprocating compressor unit system in each test direction, and obtains corresponding frequency spectra. Vibration measuring points whose vibration amplitudes exceed set limits are identified through the frequency spectra as vibration reduction positions. Vibration reduction devices having operating frequencies corresponding to the natural frequencies of the system are installed at these vibration reduction positions to reduce resonance at these frequencies. The method can effectively reduce the vibration of the reciprocating compressor unit system without structural modification, has strong adjustability, and is applicable to all types of reciprocating compressor unit systems.

[0031] The vibration reduction method for a reciprocating compressor unit system of the present invention comprehensively considers the influence of the structural natural frequency and the air column natural frequency on the system vibration, and further analyzes the system vibration characteristics by performing airflow pulsation analysis on the reciprocating compressor unit system that has been put into use. It can find more locations where severe vibrations occur, and prevent the occurrence of severe vibration points at new locations after vibration reduction control is performed according to the operating parameters under the collected operating conditions, thereby achieving the optimal vibration reduction effect.

[0032] The vibration reduction method for a reciprocating compressor unit system of the present invention uses the collected operating parameters of the compressor unit system under operating conditions as input for simulation analysis. Compared with structural optimization through simulation means in the structural design stage, it can more accurately identify the vibration risk of the system, provide a reasonable basis for vibration reduction management, and solve vibration problems that are exposed after the equipment is manufactured and put into operation despite structural optimization.

[0033] By dividing the entire reciprocating compressor unit system into system airflow pulsation simulation analysis models according to cylinders and dividing them into airflow pulsation systems that do not interfere with each other, the analysis difficulty is reduced.

[0034] The reciprocating compressor unit system vibration reduction method of the present invention can effectively reduce the vibration of the compressor unit. The test data of a certain type of reciprocating compressor unit system at different vibration severe measurement points before and after vibration reduction are shown in Table 1.

[0035] BRIEF DESCRIPTION OF THE DRAWINGS

[0036] Figure 1 Flowchart of an embodiment of the present invention.

[0037] Figure 2 Schematic diagram of the arrangement of vibration measurement points for a reciprocating compressor system according to an embodiment of the present invention.

[0038] Figure 3 Schematic diagram of the arrangement of vibration measurement points for the first-stage intake and exhaust system of an embodiment of the present invention.

[0039] Figure 4 Schematic diagram of the arrangement of vibration measurement points for the two-stage intake and exhaust system according to an embodiment of the present invention.

[0040] Figure 5 Schematic diagram of the arrangement of vibration measurement points for the final stage intake and exhaust system according to an embodiment of the present invention.

[0041] Figure 6 This is a schematic diagram of the arrangement of vibration measurement points for a vent pipeline according to an embodiment of the present invention.

[0042] Figure 7 FIG. 4 is a block diagram of a vibration testing system according to an embodiment of the present invention.

[0043] Figure 8 This is a vibration test curve diagram of the vibration measuring point 12-H direction according to an embodiment of the present invention.

[0044] Figure 9 Schematic diagram of the pulsation exceeding standard ratio of the second pulsation system according to an embodiment of the present invention.

[0045] Figure 10 Schematic diagram of the pulsating unbalanced force of the first-stage exhaust system according to an embodiment of the present invention.

[0046] Figure 11 Schematic diagram of the pulsating unbalanced force of the two-stage intake system according to an embodiment of the present invention.

[0047] Figure 12 The figure is a schematic diagram of the installation of a tuned mass damper for a horizontal compressor unit according to an embodiment of the present invention.

[0048] Figure 13 for Figure 12 Schematic diagram of the tuned mass damper structure.

[0049] Figure 14 yes Figure 13 Schematic diagram of the elastic unit.

[0050] Figure 15 yes Figure 13 Schematic diagram of the basic mass unit.

[0051] Figure 16 yes Figure 13 Schematic diagram of the adjusted mass unit.

[0052] Markings in the figure: 1-inlet and outlet air buffer tanks of the compressor unit, 2-first type of tuned mass damper, 21-clamp, 22-elastic unit, 221-elastic unit fixing interface, 222-elastic unit installation interface, 23-basic mass unit, 231-basic mass unit fixing interface, 232-basic mass unit installation interface, 24-adjustment mass unit, 241-adjustment mass unit fixing interface, 3-second type of tuned mass damper, 4-third type of tuned mass damper, 5-cylinder end, 6-compressor unit cylinder. DETAILED DESCRIPTION

[0053] The present invention will be described in further detail below with reference to the embodiments and accompanying drawings.

[0054] One embodiment of the present invention is a method for reducing vibration of a reciprocating compressor system. Figure 1 As shown, the vibration reduction method of the reciprocating compressor system of the present invention specifically includes the following steps:

[0055] 1. Collect operating parameters under operating conditions.

[0056] The operating parameters of the reciprocating compressor unit system under operating conditions are collected, the time domain vibration signals of each vibration measuring point of the reciprocating compressor unit system in each test direction are collected, and the corresponding spectrum diagram is obtained.

[0057] The operating conditions of the compressor unit system determine the vibration state of the system. When the compressor unit system is operating normally, the operating parameters include but are not limited to operating speed, gas flow, pipeline intake pressure and exhaust pressure, pipeline intake temperature and exhaust temperature. Usually, a device for obtaining corresponding operating parameters is installed in the operating compressor unit system. For example, the current operating speed can be directly obtained through the compressor's own feedback system, the gas flow can be read by the flow meter at the air inlet, the pipeline intake pressure and exhaust pressure can be read by the pressure gauge, and the pipeline intake temperature and exhaust temperature can be measured by the temperature sensor. For example, in this embodiment, the operating frequency of the compressor unit is collected to be 960rpm (16Hz), and the flow rate is 1.8×10 6 m 3 / day, the pipeline inlet pressure is 7MPa, the pipeline exhaust pressure is 14MPa, the pipeline inlet temperature is 20℃, and the pipeline exhaust temperature is 65℃.

[0058] The time domain vibration signal is the vibration state of the system during operation, and requires dedicated testing equipment for signal acquisition. The vibration measurement points of the reciprocating compressor system of this embodiment are arranged as follows: Figures 2 to 6 shown. Figure 2 CYL#1~CYL#6 are cylinder numbers. Figures 2 to 6 The circled numbers (1 to 75) are the location numbers where the time domain vibration signals were collected. Figures 3 to 6 The mid-coordinate represents the test direction: A (axial) represents the direction along the compressor crankshaft, H (horizontal) represents the direction along the compressor cylinder piston's motion, and V (vertical) represents the direction perpendicular to the ground. Three-axis accelerometers can be deployed at vibration measurement points to collect vibration signals (such as acceleration signals). These signals are then stored and processed using a dynamic signal analyzer.

[0059] The test system block diagram is as follows Figure 7 As shown, the vibration measurement point transmits the vibration signal to the analyzer through the three-axis acceleration sensor. The analyzer performs signal conditioning and data processing on the vibration signal, records the results and displays them.

[0060] The vibration amplitude can be vibration displacement, vibration velocity and / or vibration acceleration. In this embodiment, taking vibration velocity as an example, for a vibration measuring point where the vibration velocity of the compressor unit exceeds a set limit (e.g., 18 mm / s), it indicates that vibration reduction control is required at that location. For example, the time domain signal vibration data of the vibration measuring point 12 in the H direction and the corresponding spectrum are as follows: Figure 8 As shown, the maximum vibration amplitude (i.e., vibration velocity) at this vibration measurement point in the H direction is 23.3 mm / s, exceeding the set limit (18 mm / s) and posing a significant safety hazard to the unit's safe operation. Other vibration measurement points are not detailed here. Table 1 lists the data for some vibration measurement points that exceeded the limit.

[0061] 2. Obtain the system's natural frequency.

[0062] According to the design drawing information, the mass and stiffness of the equipment in the system, a mechanical vibration finite element analysis model of the reciprocating compressor unit system put into use is established to obtain the distribution of the structural natural frequency.

[0063] For example, based on the design drawing information of the reciprocating compressor unit system, the mass and stiffness of the equipment in the system, a finite element analysis model of the mechanical vibration of the reciprocating compressor unit system in use is established using finite element analysis software (such as ABAQUS), and the distribution of the natural frequency of the reciprocating compressor unit system currently in use is obtained.

[0064] A reciprocating compressor system is a complex vibration system, typically with n natural frequencies (n is an integer greater than 1). When one or more of these natural frequencies coincide with the resonant region of the compressor's excitation frequency, the piping will resonate. This resonance is called structural resonance. By installing a vibration damping device whose operating frequency corresponds to the natural frequency of the structure, the structural resonance at that frequency can be reduced, effectively reducing the vibration of the reciprocating compressor system.

[0065] Table 2 shows the calculated results for the first 12 natural frequencies of the reciprocating compressor system in this embodiment. As can be seen, the lowest frequency of the compressor system is only 10 Hz, which is lower than the compressor operating frequency (16 Hz). Furthermore, the fifth and sixth mode frequencies also fall at twice the compressor operating frequency, making the system prone to resonance. Although the natural frequencies of the compressor system can be optimized using finite element simulation software during the initial design phase of the system structure, the complexity of the compressor system makes optimal optimization difficult, resulting in resonance during actual operation.

[0066]

[0067] (Optional) 3. Establish an airflow pulsation simulation analysis model for the reciprocating compressor system in use to further analyze the system vibration characteristics.

[0068] Gas flows through the pipes of a compressor system, and the shape of the gas column resembles the pipe's outer shape. When one or more orders of the gas column's natural frequency coincide with the resonant region of the compressor's excitation frequency, the pipe will also resonate. This resonance is called gas column resonance. Both gas column resonance and structural resonance result in severe system vibration. Therefore, to better eliminate the possibility of severe vibration, the system's natural frequency includes both the gas column natural frequency and the structural natural frequency.

[0069] Preferably, in another embodiment, an airflow pulsation simulation analysis model of the reciprocating compressor unit system put into use can be established through acoustic simulation software (such as Bentley Puls) based on the drawing information, equipment information, gas composition and overall layout of the reciprocating compressor unit system, and the natural frequency of the gas column can be obtained.

[0070] Preferably, in another embodiment, the air flow pulsation analysis of the reciprocating compressor system that has been put into use is performed to further analyze the vibration characteristics of the system, specifically including:

[0071] 3-1) Establishing an airflow pulsation simulation analysis model for the reciprocating compressor unit system in use; dividing each airflow pulsation system in the airflow pulsation simulation analysis model according to the position of the vibration measurement point, setting corresponding internal nodes to obtain the sound wave propagation characteristics at the internal nodes, and setting boundary nodes at the boundaries of each airflow pulsation system in the system airflow pulsation simulation analysis model.

[0072] An airflow pulsation simulation analysis model of the reciprocating compressor system in use may be established using acoustic simulation software (eg, Bentley Puls) based on the drawing information, equipment information, gas composition, and overall layout of the reciprocating compressor system.

[0073] Since the cylinder is the dividing point of the airflow pulsation system, each airflow pulsation system does not interfere with each other and can be analyzed independently. Therefore, it is preferred to divide the entire reciprocating compressor unit system into each airflow pulsation system in the system airflow pulsation simulation analysis model according to the cylinder, which reduces the difficulty of analysis.

[0074] like Figure 2 The compressor system shown is divided into 6 pulsation systems, namely the first pulsation system of the first-stage intake, the second pulsation system of the first-stage exhaust and the second-stage intake, the third pulsation system of the second-stage exhaust and the last-stage intake, the fourth pulsation system of the last-stage exhaust, the fifth pulsation system of the first-stage intake manifold, and the sixth pulsation system of the last-stage exhaust manifold.

[0075] 3-2) Using the operating parameters of the reciprocating compressor unit system under operating conditions as input to the airflow pulsation analysis model, calculating the pulsating unbalanced forces of the internal nodes and boundary nodes; using the pulsating unbalanced forces of the internal nodes and boundary nodes as excitation force input, applying them to the mechanical vibration finite element analysis model, obtaining the dynamic excitation force forced vibration response of the reciprocating compressor unit system, and comparing the effective value of the vibration amplitude in the obtained dynamic excitation force forced vibration response with the set limit.

[0076] The vibration amplitude may be vibration displacement, vibration velocity, and / or vibration acceleration. For example, in this embodiment, when the vibration velocity exceeds a set limit (18 mm / s), the vibration is considered severe and poses a safety hazard to the normal operation of the unit.

[0077] The operating parameters of the reciprocating compressor system under operational conditions, collected in Step 1, are used as input for the airflow pulsation analysis model. The acoustic simulation software automatically calculates the pulsation values ​​of the model nodes in the frequency domain and accurately calculates the pulsation imbalance force from these values. To evaluate the system's pulsation characteristics, the airflow pulsation analysis results can be compared with the requirements of the API 618 standard. API 618 is a standard for the design, manufacture, procurement, installation, operation, and maintenance of reciprocating compressors and their auxiliary equipment used in the petroleum, chemical, and natural gas industries, and is a key reference for reciprocating compressor-related activities. Figures 9 to 11 The calculation results obtained by using the compressor system operating parameters collected in Step 1 as input to the airflow pulsation analysis model show that the peak-to-peak pulsation value of the first-stage exhaust and second-stage intake systems (i.e., the second pulsation system) is 2.95 times the API 618 standard value, and the pulsation imbalance force also exceeds the corresponding standard requirements. This indicates that the reciprocating compressor system in operation is experiencing abnormal vibration caused by airflow pulsation. The other pulsation systems were analyzed using the same model and are not detailed here.

[0078] Using the mechanical vibration analysis model to analyze the system's dynamic excitation force forced vibration response can more objectively reflect the vibration conditions of the compressor unit system.

[0079] The exciting force is generally the unbalanced force caused by airflow pulsation and the gas force within the cylinder. The pulsating unbalanced force at each internal node and boundary node obtained from the above airflow pulsation analysis is used as the excitation force input and applied to the mechanical vibration finite element analysis model of the reciprocating compressor system in operation. Table 3 shows the effective value response of the system vibration velocity after the excitation force is applied to the reciprocating compressor system. The results show that the vibration amplitude of the first-stage air intake buffer tank and its associated piping, the first-stage exhaust buffer tank and its associated piping is large, and the vibration speed of the first-stage air intake buffer tank and the first-stage exhaust buffer tank exceeds the set limit (18mm / s), posing a safety hazard to the normal operation of the unit.

[0080]

[0081] This shows that while structural optimization through simulation during the reciprocating compressor system's design phase achieves vibration reduction based on theory, vibration issues that arise after the equipment is manufactured and put into operation cannot be identified before production begins. This is partly due to a discrepancy between the input information used during the structural design phase and actual operation. Therefore, by using the actual operating status of the compressor system, as input for simulation analysis in step one, we can more accurately identify the system's vibration risks and provide a sound basis for vibration reduction measures.

[0082] 4. Determine the vibration reduction position.

[0083] The vibration reduction position refers to the area of ​​the vibration measurement point in the compressor unit system where the vibration amplitude exceeds the set limit.

[0084] Using the spectrogram described in step 1, vibration measurement points whose vibration amplitudes exceed the set limit are identified and designated as vibration reduction points P1n. For example, in step 1, after processing the data from all vibration measurement points, the spectrogram can be used to identify vibration measurement points whose maximum vibration velocity exceeds the set limit (18 mm / s). These vibration reduction points are designated as vibration reduction points and recorded as P1n. Vibration reduction control must be implemented for these vibration reduction points P1n, as they can present problems in actual operation.

[0085] Preferably, in another embodiment, the vibration measurement points where the effective value of the vibration amplitude in the forced vibration response of the dynamic excitation force obtained through system airflow pulsation in step 3 exceeds the set limit are designated as vibration reduction locations P2n. For example, internal nodes or boundary nodes where the effective value of the vibration velocity exceeds the set limit are designated as vibration reduction locations P2n. The union of vibration reduction locations P1n and P2n is designated as all vibration reduction locations requiring vibration control. According to the vibration limit requirements for the compressor system, the vibration velocity of all controlled vibration reduction locations should not exceed the set limit (18 mm / s).

[0086] Vibration damping positions P2n and P1n may intersect or differ from each other. The vibration damping positions in P2n that differ from P1n are denoted as P2x. Vibration damping control is also required for vibration damping position P2x. This is because, after vibration damping control is applied to vibration damping position P1n, the vibration occurring at position P2x may become prominent in real-world situations, forming a new vibration-damping point. Therefore, in the present invention, the union of P1n and P2n is used as the set of all vibration damping positions requiring vibration control, achieving optimal vibration damping.

[0087] In this embodiment, the identified vibration reduction positions include the cylinder end, the buffer tank air inlet, the buffer tank exhaust, and the auxiliary pipes of the air inlet and exhaust ports.

[0088] 5. Install the tuned mass damper.

[0089] At each vibration reduction position, a corresponding tuned mass damper is designed and installed according to the natural frequency of the system to reduce vibration.

[0090] Since the compressor system is already operational, shutting it down for structural optimization to address the vibration issue would not only affect production output but also incur higher costs and time. It should be noted that structural modification alone cannot completely address the system's vibration issues; improper modification can even lead to new vibration issues.

[0091] In this invention, a tuned mass damper is designed and installed at each vibration reduction location to achieve vibration reduction. As a vibration reduction device, the tuned mass damper does not require modification to the existing system structure, can be quickly installed and removed, and has an adjustable tuning frequency, making it a preferred choice for vibration control in compressor systems. Specifically, it includes:

[0092] 5-1) In the mechanical vibration finite element analysis model, a corresponding tuned mass damper model is added at each vibration reduction position to obtain a global optimal solution for the mass, stiffness, and damping coefficient of the tuned mass damper model.

[0093] Optimization software (such as HyperStudy) can be used for joint simulation, and the genetic optimization algorithm provided by the software can be used to achieve global optimization to find the global optimal solution for the mass, stiffness and damping coefficient of the tuned mass damper.

[0094] During global optimization, the parameters of the tuned mass damper (TMD) are constrained to meet the following constraints: a mass ratio of 1% to 5%, a natural frequency ratio of 0.95 to 1, and a damping ratio of 0.05 to 0.2. The mass ratio refers to the ratio of the TMD's mass to the mass of the damped system; the natural frequency ratio refers to the ratio of the TMD's natural frequency to the natural frequency of the damped system, with the stiffness-to-mass ratio equal to the square of the natural frequency; and the damping ratio refers to the ratio of the TMD's damping coefficient to the critical damping coefficient of the damped system. The goal of global optimization is to minimize the maximum vibration velocity at all damping locations, thereby obtaining the global optimal solution for the TMD's mass, stiffness, and damping coefficient.

[0095] 5-2) Design and install the tuned mass damper at each vibration reduction position based on the global optimal solution of the mass, stiffness and damping coefficient of the tuned mass damper.

[0096] According to the optimization results of the tuned mass damper parameters and comprehensive consideration of the structural characteristics of the vibration reduction position, the tuned mass damper structure is designed and manufactured and installed on the compressor unit system. Figure 12 This is a horizontal compressor unit, a specific implementation example of a reciprocating compressor unit. A first-type tuned mass damper 2 and a second-type tuned mass damper 3 are installed on the inlet and outlet gas buffer tanks 1 of the horizontal compressor unit for vibration reduction. A third-type tuned mass damper 4 is installed at the end 5 of the compressor unit's cylinder for vibration reduction.

[0097] like Figure 13 As shown, the first tuned mass damper 2 includes a clamp 21, an elastic unit 22, a basic mass unit 23 and an adjustment mass unit 24; and is used to control the inlet and outlet gas buffer tank 1 or the cylinder end 5 of the compressor unit.

[0098] The clamp 21 is tightly clamped on the outer circumference of the inlet and outlet gas buffer tank 1 or the cylinder end 5 of the compressor unit. The clamp 21 is composed of several arc segments, and each arc segment is connected to form a ring. Several elastic units 22 are fixedly installed on the outside of the clamp 21 in an array in a centrally symmetrical distribution manner. A basic mass unit 23 is fixedly installed on the outside of each elastic unit 22, and one or more adjustment mass units 24 are fixed on each basic mass unit 23.

[0099] like Figure 14 As shown, the elastic unit 22 is provided with an elastic unit mounting interface 222 , and the basic mass unit 23 is provided with a basic mass unit fixing interface 231 . The basic mass unit fixing interface 231 cooperates with the elastic unit mounting interface 22 to fix the basic mass unit 23 on the elastic unit 22 .

[0100] like Figure 15 and Figure 16As shown, the basic mass unit 23 is provided with a basic mass unit installation interface 232; the adjustment mass unit 232 is provided with an adjustment mass unit fixing interface 241. The adjustment mass unit fixing interface 241 cooperates with the basic mass unit installation interface 232 to fix the multiple adjustment mass units 24 outside the multiple basic mass units 23.

[0101] When the first type of tuned mass damper 2 is in operation, the base mass unit 23 and the adjustable mass unit 24 vibrate relative to the clamp 21 along with the elastic unit 22, absorbing the vibration energy of the compressor unit's inlet and outlet gas buffer tank 1 and cylinder end 5. Each base mass unit 23 and adjustable mass unit 24 vibrates along with the first type of tuned mass damper 2, thereby accounting for multiple vibration directions and facilitating adjustment. This allows for more convenient and precise matching of optimal vibration parameters, resulting in the best vibration control effect.

[0102] The number of elastic units 22 can be adjusted. The installation direction and position of the elastic units 22 can be adjusted according to the direction of vibration control. For example, they can be distributed along the circumference of the clamp 21, or distributed radially along the inlet and outlet gas buffer tank 1 or cylinder end 5 of the compressor unit, or distributed axially along the inlet and outlet gas buffer tank 1 or cylinder end 5 of the compressor unit. By adjusting the number, material, shape, installation direction, and position of the elastic units, vibration-related properties such as elastic stiffness, damping, and mass can be adjusted, accurately matching the optimal vibration parameters, thereby achieving the best vibration control effect.

[0103] The first type of tuned mass damper 2 is a clamp that is tightly clamped to the outside of the inlet and outlet air buffer tank 1 or the cylinder end 5 of the compressor unit. It does not require any modification to the equipment and is easy to implement. The installation and implementation have little impact on the vibration control object and can effectively reduce the vibration magnitude of the inlet and outlet air buffer tank 1 or the cylinder end 5 of the compressor unit.

[0104] The first type of tuned mass damper (TMD) 2 features a modular design consisting of clamps, elastic units, base mass units, and adjustable mass units. This allows for flexible configuration based on equipment vibration conditions, quickly resolving issues with excessive vibration in multiple directions. By adjusting the number, material, shape, and orientation of the elastic units, base mass units, and adjustable mass units, vibration-related properties such as elasticity, damping, and mass can be adjusted, precisely matching optimal vibration parameters and achieving optimal vibration control results.

[0105] According to step 1, operating parameter data for a reciprocating compressor unit equipped with a tuned mass damper (TMD) was collected under operating conditions, and key vibration measurement points were compared before and after vibration reduction. The data comparison is shown in Table 1. The results demonstrate that the TMD's vibration reduction measures meet the system's vibration reduction requirements, and the vibration reduction method of the present invention ensures the stability, safety, and reliability of the unit's operation.

[0106] The reciprocating compressor unit system vibration reduction method of the present invention collects operating parameters of the reciprocating compressor unit system under operating conditions, collects time domain vibration signals of each vibration measuring point of the reciprocating compressor unit system in each test direction, and obtains corresponding frequency spectrograms. Vibration measuring points whose vibration amplitudes exceed set limits are identified through the frequency spectrograms as vibration reduction positions; and by installing vibration reduction devices at these vibration reduction positions with operating frequencies corresponding to the natural frequencies of the system, resonance at these frequencies is reduced, thereby effectively reducing the vibration of the reciprocating compressor unit system without modifying the structure.

[0107] The vibration reduction method for a reciprocating compressor unit system of the present invention comprehensively considers the influence of the structural natural frequency and the air column natural frequency on the system vibration, and further analyzes the system vibration characteristics by performing airflow pulsation analysis on the reciprocating compressor unit system that has been put into use. It can find more locations where severe vibrations occur, and prevent the occurrence of severe vibration points at new locations after vibration reduction control is performed according to the operating parameters under the collected operating conditions, thereby achieving the optimal vibration reduction effect.

[0108] The vibration reduction method for a reciprocating compressor unit system of the present invention uses the collected operating parameters of the compressor unit system under operating conditions as input for simulation analysis. Compared with structural optimization through simulation means in the structural design stage, it can more accurately identify the vibration risk of the system, provide a reasonable basis for vibration reduction management, and solve vibration problems that are exposed after the equipment is manufactured and put into operation despite structural optimization.

[0109] By dividing the entire reciprocating compressor unit system into system airflow pulsation simulation analysis models according to cylinders and dividing them into airflow pulsation systems that do not interfere with each other, the analysis difficulty is reduced.

[0110] Although the present invention has been disclosed above with reference to preferred embodiments, these embodiments are not intended to limit the present invention. Any equivalent changes or modifications made without departing from the spirit and scope of the present invention are also within the scope of protection of the present invention. Therefore, the scope of protection of the present invention should be based on the content defined in the claims of this application.

Claims

1. A reciprocating compressor system vibration reduction method, characterized in that: include: Collect operating parameters under operating conditions: Collect the operating parameters of the reciprocating compressor system under operating conditions, collect the time domain vibration signals of each vibration measurement point in each test direction of the reciprocating compressor system, and obtain the corresponding spectrum diagram; Obtaining the system's natural frequency: Based on the design drawing information, the mass and stiffness of the equipment in the system, a mechanical vibration finite element analysis model of the reciprocating compressor system in use is established to obtain the distribution of the structural natural frequency; Establish an airflow pulsation simulation analysis model for the reciprocating compressor system in use; divide the airflow pulsation systems in the airflow pulsation simulation analysis model according to the positions of the vibration measuring points, set corresponding internal nodes, and set boundary nodes at the boundaries of each airflow pulsation system; use the operating parameters of the reciprocating compressor system in use under operating conditions as inputs to the airflow pulsation analysis model, and calculate the pulsation unbalanced forces of the internal nodes and boundary nodes; use the pulsation unbalanced forces of the internal nodes and boundary nodes as excitation force inputs, apply them to the mechanical vibration finite element analysis model, obtain the dynamic excitation force forced vibration response of the reciprocating compressor system, and compare the effective value of the vibration amplitude in the obtained dynamic excitation force forced vibration response with the set limit value; Determining the vibration reduction position: identifying, through the spectrum diagram, a vibration measurement point whose vibration amplitude exceeds a set limit as the vibration reduction position P1n; identifying a vibration measurement point whose vibration amplitude effective value exceeds the set limit in the obtained dynamic excitation force forced vibration response as the vibration reduction position P2n; and using the union of the vibration reduction positions P1n and P2n as all the vibration reduction positions requiring vibration control; Install tuned mass dampers: At each vibration reduction position, corresponding tuned mass dampers are designed and installed according to the natural frequency of the system to reduce vibration.

2. The reciprocating compressor system vibration reduction method according to claim 1, characterized in that: The system natural frequency also includes the natural frequency of the air column.

3. The reciprocating compressor system vibration reduction method according to claim 1, characterized in that: The reciprocating compressor unit system is divided into various airflow pulsation systems in the system airflow pulsation simulation analysis model according to the cylinders.

4. The reciprocating compressor system vibration reduction method according to any one of claims 1 to 3, characterized in that: The installation of the tuned mass damper comprises: In the mechanical vibration finite element analysis model, a corresponding tuned mass damper model is added at each vibration reduction position to obtain a global optimal solution of the mass, stiffness, and damping coefficient of the tuned mass damper model; According to the global optimal solution of the mass, stiffness and damping coefficient of the tuned mass damper, the tuned mass damper at each vibration reduction position is designed and installed so that its natural frequency is equal to or close to the natural frequency of the system.

5. The reciprocating compressor system vibration reduction method according to claim 4, characterized in that: When obtaining the global optimal solution of the mass, stiffness, and damping coefficient of the tuned mass damper model, the parameter constraints of the tuned mass damper are defined as: mass ratio 1% to 5%, natural frequency ratio 0.95 to 1, and damping ratio 0.05 to 0.

2.

6. The reciprocating compressor system vibration reduction method according to any one of claims 1 to 3, characterized in that: The operating parameters include but are not limited to operating speed, gas flow, pipeline inlet pressure and exhaust pressure, pipeline inlet temperature and exhaust temperature.

7. The reciprocating compressor system vibration reduction method according to any one of claims 1 to 3, characterized in that: The vibration amplitude is vibration displacement, vibration velocity and / or vibration acceleration.

8. The reciprocating compressor system vibration reduction method according to any one of claims 1 to 3, characterized in that: The vibration reduction positions include the cylinder end, the buffer tank air inlet, the buffer tank exhaust port and the auxiliary pipes of the air inlet and exhaust ports.

9. The reciprocating compressor system vibration reduction method according to any one of claims 1 to 3, characterized in that: The tuned mass damper comprises a clamp, an elastic unit, a basic mass unit and an adjustment mass unit; The clamp is tightly clamped on the outer circumference of the inlet and outlet gas buffer tank or the cylinder end of the compressor unit. The clamp is composed of several arc segments, and each arc segment is connected to form a ring. Several elastic units arranged in an array in a centrally symmetrical distribution are fixedly installed on the outer side of the circumferential surface of the arc segment of the clamp, and the basic mass unit is fixedly installed on the outer side of each elastic unit; one or several adjustment mass units are respectively fixed on each basic mass unit.

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