Lecture Notes on Centrifugal Compressors

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Part 1: Overview of Compressors Pumps and compressors are machines used to transport fluids. Fluids are a general term for gases and liquids. What they have in common is that they have no fixed shape and depend on the container in which they are held. The difference is that liquids are incompressible fluids, while gases are compressible fluids. A pump is a machine for transporting liquids. A compressor is a machine that transports gas. A compressor is a machine that adds energy to gases. Used to transport gas or increase its pressure. Function of the compressor: 1. Compress gas to generate power. Such as: pneumatic tools, instrument automation control. 2 Used for refrigeration or gas separation. For example: refrigerators (ice makers), separation of petroleum cracking gases. 3 Used for synthesis and polymerization. For example: the synthesis of ammonia from nitrogen and hydrogen, high-pressure polyethylene. 4 Used for gas transportation. For example: long-distance transportation of natural gas. Classification of compressors: Based on their working principle, they are divided into two main categories – positive displacement type and dynamic type. 1 Volume-type types achieve gas compression and delivery through periodic changes in the volume of the working chamber. It is further divided into two categories: piston-type and vane-type. Reciprocating types include piston-type, while rotary types include rotary, sliding vane, diaphragm, rotor, and screw types. Speed-type machines rely on the high-speed rotation of blades to impart great kinetic energy to the gas, which is then converted into pressure energy. Based on the flow direction of gas discharge, they are classified as: centrifugal, where the gas is discharged radially along the blades. Velocity type, axial flow: The gas is discharged along the axis of the blades. Mixed-flow type: The gas is discharged axially at 45˚. Figure 1-1 Piston type Figure 1-2 Roots type Figure 1-3 Sliding vane type Figure 1-4 Screw type Figure 1-5 Centrifugal type Figure 1-6 Oblique-flow machine Figure 1-7 Axial-flow machine Figure 1-8 Transverse-flow type Part II: Centrifugal Compressors I. Overview 1. Classification of centrifugal fans. Based on the level of exhaust pressure, they can be classified as: ventilators, with an exhaust pressure below 0. 015 MPa (or less than 1500 mm of water column) blower, with an exhaust pressure of 0. 015—0. 35MPa compressor, with an exhaust pressure higher than 0.35MPa. Figure 1-9: Classification of jet and centrifugal compressors. Based on structure, they are divided into three types: horizontal split type, cylinder type, and isothermal type. Note: An isothermal type of compressor is one in which the gas compressed by each stage of impellers is cooled through an inter-stage cooler before being fed to the next stage. 3 The working principle of centrifugal compressors: Generally speaking, the main goal of increasing gas pressure is to increase the number of gas molecules per unit volume, that is, to reduce the distance between gas molecules. To achieve this goal, in addition to the positive displacement compression method that uses compression elements to compress the gas, there is also a gas dynamic approach. In this method, the working element of the machine – a high-speed rotating impeller – does work on the gas, causing the pressure of the gas to increase in the centrifugal field; at the same time, its kinetic energy increases significantly. As the gas then flows through the expanding ducts, this kinetic energy is converted back into static pressure energy, further increasing the gas pressure. This is the working principle, or principle of pressure increase, behind centrifugal compressors. For Model 4, there is currently **no unified standard numbering system; each manufacturer assigns its own numbers. Commonly used: DA □—□ □ Design sequence number ； Number of impeller stages ； Flow rate at the first stage of inhalation ； m³/min is the designation for centrifugal compressors. For example: DA220—72 refers to a centrifugal compressor with a gas flow rate of 220 m³/min in inhalation mode, 7 compression stages, and it is the second-generation design of this type of compressor. DA350—61: The gas flow rate under suction is 350 m³/min; it has 6 compressor stages and is the first centrifugal compressor of its kind to be designed. Also: □□-□ Number of impellers, Impeller diameter in cm, Shell structure type, Number of sections (1 is not indicated). Structure type code: MCL---Shell with horizontal split; BCL---Shell with vertical split. Example: 2MCL—456 means 2-stage compression, shell with horizontal split, impeller diameter of 45 cm, and 6 impellers. Others include: DH is biaxial grade-4 compression, and VK is biaxial grade-3 compression. 5 Advantages and disadvantages of centrifugal compressors: (compared to reciprocating compressors) Advantages: (1) Centrifugal compressors have a large gas handling capacity, a simple and compact structure, low weight, small unit size, and require less floor space. (2) It operates smoothly, is reliable in operation, has a high efficiency, and features few friction components; as a result, the demand for spare parts is low, and maintenance costs as well as the need for personnel are reduced. (3) In chemical processes, centrifugal compressors enable the compression of chemical media in a completely oil-free manner. (4) The centrifugal compressor is a type of rotating machine that is suitable for being directly driven by industrial steam turbines or gas-fired turbines. In general large-scale chemical plants, waste steam is commonly used to drive industrial steam turbines for power generation, thereby enabling the comprehensive utilization of thermal energy. Disadvantages (1): Centrifugal compressors are not currently suitable for applications with very low gas flow rates and excessively high pressure ratios. (2) The stable operating range of centrifugal compressors is narrow; although gas flow rate adjustment is convenient, their economic efficiency is poor. (3) Currently, the efficiency of centrifugal compressors is generally 5-10% lower than that of piston compressors due to high energy losses. Note: Compression ratio---the ratio of the absolute pressure at the outlet to that at the inlet. II. Overall Structure of Centrifugal Compressors A centrifugal compressor consists of two main parts: the rotor and the stator. The rotor includes a main shaft, as well as components such as an impeller fixed to the shaft, shaft sleeves, balance discs, thrust discs, and couplings. The stator consists of cylinders, various partitions positioned on the cylinder block, as well as components such as bearings. Sealing elements are also provided at the rotor shaft ends and wherever gas needs to be sealed between the rotor and the stator. In each section of the compressor, it is generally composed of several or one compressor stage. Each stage consists of an impeller and its corresponding fixed components. The fixed components include an inhalation chamber, diffuser, bend, return device, and volute, among others. Figure 8-1: Vertical cross-sectional structure of a centrifugal compressor. 1: Suction chamber; 2: Impeller; 3: Diffuser; 4: Bend; 5: Recirculator; 6: Vortex chamber; 7, 8: Seals; 9: Diaphragm seal; 10: Cover seal; 11: Balance disk; 12: Thrust disk; 13: Coupling; 14: Snap ring; 15: Main shaft; 16: Housing; 17: Bearing; 18: Thrust bearing; 19: Diaphragm; 20: Guide vanes. III. Basic equations of centrifugal compressors 1. Velocity triangle: Due to the high-speed rotation of the impeller, the gas moves in a circular path along with it (circular velocity u). Additionally, due to the effect of centrifugal force, the gas flows from the inlet to the outlet of the impeller along the blades, and the gas moves relative to the impeller (relative velocity w). For a stationary casing, the gas is in absolute motion (absolute velocity C). The absolute velocity C is equal to the vector sum of the circumferential velocity u and the relative velocity w. That is: C=u+w. The velocity triangle of the backward-blade impeller. IV. Power and efficiency of centrifugal compressors. The losses in centrifugal compressors can be divided into flow losses, leakage losses, wheel resistance losses, and mechanical losses; flow losses cause a decrease in pressure, leakage losses lead to a reduction in flow rate, while wheel resistance losses and external mechanical losses result in additional energy consumption. 1 Flow channel loss: The loss incurred as gas flows through components such as the suction chamber, impeller, diffuser, bends, and recirculator. Including flow loss and shock loss. Flow losses, in turn, include frictional losses, boundary layer separation losses, local losses, and wake losses. (1) Friction loss—The energy lost due to friction between the gases as they flow, as well as between the gases and the walls of components such as the impeller, cover, and diffuser. (2) Boundary layer separation loss—In channels with decelerating pressure gradients, the near-wall boundary layer tends to thicken, and even separation vortices may form, resulting in separation loss. As shown in the figure: Schematic diagram of boundary layer separation (3) – Wake loss: The loss resulting from vortices that are generated when gas flows out of the channels between the blades, due to the fact that the blades themselves have a certain thickness. (4) Shock loss—When the gas flow rate is greater than or less than the compressor’s designed flow rate, the flow enters the impeller and diffuser at an angle different from the inlet angle of the blades; this results in shock between the flow and the impeller and diffuser, leading to boundary layer separation and thus generating loss, which is known as shock loss. Airflow separation in the impeller flow channel under different impact angles 2. Impeller friction loss: This is the energy loss that occurs as a result of friction between the impeller disk and the sides of the impeller casing while the impeller rotates at high speed in the gas. 3 Air leakage loss: the energy loss caused by air leaking between internal components or to the outside. 4 Power and efficiency A: The total energy increase of the gas with effective power Ne and effective volumetric flow rate Q after passing through the compressor. B Internal power Ni – the power actually consumed by the gas. The C-axis power Ns – the power input to the compressor is the shaft power Ns. The output power of engine D – the rated power of the engine is generally Ne≥1.3Nz. V. Performance curves of centrifugal compressors: The performance curve of stage 1 refers to the curves showing how the pressure ratio ε, efficiency η, and power N change as the inlet flow rate Qj to that stage varies. That is, the curves for ε-Qj, η-Qj, and N-Qj. These curves were obtained from experiments. 2 The performance curve of the centrifugal compressor—is similar to that of a stage. It refers to the curves showing how the overall pressure ratio ε, efficiency η, and power N vary with the inlet gas flow rate Q. It is also determined through experiments. 3 Characteristics of the compressor performance curve: The pressure ratio ε decreases as the flow rate increases. Power and efficiency increase as the flow rate increases; once a certain flow rate is reached, further increases in flow rate result in a decrease in power and efficiency. 4 Pipeline characteristic curve—When the conditions of the pipeline and the equipment are fixed, that is, when the pressures at both ends of the pipeline, the pipeline dimensions, length, the number and size of fittings, as well as the degree to which the valves are open are all constant, the Hi-Q curve represents the relationship between the work required per kilogram of gas as it flows through the pipeline and the volume of gas flowing through that pipeline per unit time Q (it can also be expressed in terms of the pressure at the pipeline ends and the gas flow rate, that is, the P-Q curve). 5 Operating point of the centrifugal compressor: By plotting the performance curve Pκ-Qj of the compressor alongside the pipeline characteristic curve Pe-Qj on the same coordinate system, with the horizontal axis representing Qj and the vertical axis representing pressure P, the intersection point M of the two curves represents the operating point of the compressor. When the performance curves of the compressor and the piping are fixed, the compressor can only operate at the focal point M of the two curves. Optimal operating point: The point on the curve where the efficiency is highest is usually referred to as the optimal operating point; it is generally the operating point that was taken into account during the design of the machine. At the operating points on either side of the optimal operating point, the efficiency decreases. 6 Surging: Surging, also known as \"fluttering,\" is a special phenomenon that occurs in centrifugal compressors. The compressor has a minimum design flow rate; when the actual operating flow rate is below this minimum value by a certain amount, the direction of the airflow relative to the blade inlet angle is not consistent, that is, the attack angle I > 0. In this case, gas separation (rotational separation) occurs on the non-working side of the blade. When the shock angle reaches a certain value, the rotating separation regions merge together and occupy the flow channel. When the compressor stops discharging gas, the gas in the pipeline flows back to compensate for the insufficient flow, and then exits again after being compressed by the impeller. After this burst of gas was released, the flow stopped again, and the gas flowed back in. By continuously changing the flow direction in this manner, \"low-frequency, high-amplitude\" pressure fluctuations are generated in the machines and pipelines, resulting in a noise similar to the roar of a bull. This is actually a strong airflow impact generated when the air flow alternates between reversing direction and being exhausted. This impact causes severe vibration in the machine, and if no measures are taken promptly, it will lead to serious damage to the compressor. This is \"surge\" in centrifugal compressors; surge is generally caused by the separation of the flow boundary layer in the blade diffuser, which then spreads throughout the entire flow channel. Characteristics of surge: pressure fluctuations that are \"low-frequency and high-amplitude,\" with a sound similar to an ox’s roar. 7 Three operating conditions of centrifugal compressors: (1) Surge condition – the condition at minimum flow rate. (2) Blockage condition (stall condition) — the condition at maximum flow rate. There are two possible reasons for this condition: one is that the airflow at a certain throat in the flow channel of the stage reaches a critical state; at this point, the volumetric flow rate of the gas is already at its maximum value, and no matter how much the back pressure of the compressor decreases, the flow rate cannot increase any further. This condition is known as a \"clogging\" condition. In another scenario, the flow channel does not reach a critical state; in other words, a \"blocking\" condition does not occur. However, at such high flow rates, the flow losses within the compressor are very large, and the exhaust pressure that can be generated is extremely low, almost approaching zero head. This pressure is merely sufficient to overcome the flow resistance in the exhaust pipe and maintain such a high flow rate, and this represents the compressor’s maximum flow rate condition. (3) Normal operating condition -- the operating range between the surge condition and the stagnation condition. 8 Relationship between the operating point of the centrifugal compressor and surge. The ε-Q curve of a centrifugal compressor is a curve that has a maximum at a gas flow rate other than zero. The area to the right of the peak is called the stable operating region, while the area to the left is called the unstable operating region (choking region). The gas volume corresponding to the highest point is the minimum flow rate Qjmin for compressor surge. 9 Factors affecting and causing surge. When the performance curve of the compressor and/or the performance curve of the piping network change, the intersection point changes as well; in other words, the operating conditions of the compressor change, resulting in variable-condition operation. The characteristic curve (ε-Q) of a centrifugal compressor is related to the compressor’s speed, the properties of the medium, and the conditions of the inlet gas. The changes in the performance curve are shown in the figure. The variation in operating conditions of centrifugal compressors sometimes does not occur under direct, intentional control by humans (such as by adjusting valves), but rather as a result of indirect disturbances from the production system or even the drive mechanism. Centrifugal compressors in chemical plants often experience unexpected surging. Examples are as follows. a. A certain compressor originally had an inlet air temperature of 20°C, with its operating point at point A (see Figure a). Due to a failure in the cooler during operation, the temperature of the incoming air rose sharply to 60°C, and at this point the compressor suddenly experienced surge. The reason for this is that the increase in intake temperature causes the compressor’s performance curve to shift downward, from line 1 to 1’, while the pipeline network’s performance curve remains unchanged. As a result, the compressor’s operating point moves to point A’, and if this point falls on the surge line, surge will occur. b. A certain compressor was originally operating normally at point A as shown in Figure b, but later, for some reason, the intake pipe became blocked by foreign objects, resulting in surge. The reason for this is that the intake pipe is blocked, causing the compressor’s intake pressure to drop from Pj to Pj’, which results in the machine’s performance curve dropping to line 1’. The performance curve of the piping system remains unchanged; as a result, the operating point shifts to A’, falling on the surge line. c. A certain compressor originally operated normally at a speed of n1, with the operating condition at point A (see Figure C). Later, due to an insufficient supply of high-pressure steam during production, the speed of the steam turbine acting as the drive machine dropped to n2. At this point, the operating point A’ of the compressor moved into the surge zone, resulting in surge. In addition, there are also cases of surge caused by changes in the molecular weight of the gas. As the molecular weight decreases, the performance curve of the compressor shifts toward the lower left; if it enters the surge zone, surge will occur. In all the above cases, surge occurs due to a downward shift in the compressor performance curve, with no change in the pipeline network performance. Sometimes, surge is caused by changes in the performance curve of the piping network (such as the curve moving upward or becoming steeper). Diagram showing the situation of surge caused by changes in the pipeline network performance. The compressor was originally operating at point A’ (see Figure 6-13); later, due to instability in the production system, the pressure in the pipeline network increased significantly. As a result, the performance curve of the pipeline network shifted from line 2 to line 2’ (while the compressor’s performance curve remained unchanged), and this led to surge in the compressor. Another similar situation occurs when the exhaust valve is closed too tightly, causing the performance curve of the piping network to become steeper; once the compressor’s operating point enters the surge zone, surge occurs suddenly. When certain factors cause changes in the performance of both the compressor and the piping network, surge will occur suddenly as long as the intersection point of the two curves ends up within the surge region. For example, during the startup process of a centrifugal compressor (speed and pressure increase) and its shutdown process (speed and pressure decrease), both types of performance curves change gradually. Changing the speed alters the compressor’s performance curve, while changing the pressure in the system alters the performance curve of the piping network. During operation, it is essential to constantly ensure that the two change in coordination in order to keep the compressor operating within its stable operating range at all times. 10 Measures to Prevent Surge. The principle of preventing surge is to address the causes of surge, and immediately increase the compressor’s flow rate when surge is about to occur. Surge occurs because the flow rate of the compressor is too low, below its minimum flow rate, and the pressure in the piping system is higher than the discharge pressure provided by the compressor, resulting in gas backflow and significant airflow fluctuations. Common measures: (1) Vent a portion of the gas through an anti-surge valve. (2) Send some of the gas to the intake pipe via a bypass. (3) Disconnect the compressor from the air supply system. VI. Components of Centrifugal Compressors (I) Rotating Elements – Rotor: In a centrifugal compressor, the rotating body composed of elements such as the main shaft, impeller, balance disk, thrust disk, coupling, sleeve (or shaft sleeve), as well as retaining rings and fixing rings, is referred to as the rotor. Concept of critical speed: When the natural vibration frequency of the rotor coincides with its operating frequency (rotation speed), vibration intensifies and resonance occurs. The rotational speed at this point is called the rotor’s critical speed. A rotor whose operating speed is below the first-order critical speed is called a rigid rotor, while a rotor whose operating speed is above the first-order critical speed is called a flexible rotor. The actual rotational speed should be far away from the critical speed; otherwise, an accident will occur. To ensure the safe operation of the machine, it is required that the operating speed be kept away from the first and second critical speeds. The verification criteria are as follows: for rigid rotors, n ≤ 0.75nc1; for flexible rotors, 1.3nc1 ≤ n ≤ 0.7nc2. To prevent potential bearing oil film oscillations, the operating speed should be below twice the first critical speed, that is, n ≤ 2nc1. Since the rotor rotates at high speeds, it needs to be balanced. The method of balancing that allows the position of the rotor’s unbalanced mass to be determined while the rotor is at rest, and also enables the determination of the location and magnitude of the counterweights to be used, is known as static balancing. Static equilibrium is primarily used to balance the inertial forces of disk-shaped rotors. The method of balancing in which the position of the rotor’s unbalanced mass can be determined only while the rotor is in motion, as well as the location and size of the counterweights that need to be applied, is known as dynamic balancing. The dynamic balancing of a rigid rotor can be achieved using a balancing machine to counteract inertial forces and inertial couples, thereby eliminating vibrations of the rotor on its elastic supports. 1 The spindle serves to support rotating components and transmit torque. The spindle is generally designed to be stepped or flagellar in shape. The parts on the main shaft fit with the shaft, and generally the red-hot fitting method is used (heating, with a clearance of 0. 30-0. 50mm). When connected with keys, the keyways of the impellers at each stage should be offset by 180°, which is beneficial for strength and balance. 2 Impeller: The impeller is the most important component in a centrifugal compressor. It is through this rapidly rotating impeller that the mechanical power from the drive unit is applied to the gas, thereby imparting energy to it. It is the only component in the compressor that performs work; it is also known as the working wheel. An impeller generally consists of a shroud, a disk, and blades, forming a closed impeller; there are also semi-open impellers without a shroud. Classified by manufacturing method: riveting, casting, welding, electroerosion, etc. Classified by the bending pattern of the impeller: forward-bending, backward-bending, and radial. Forward-bending impellers produce the most work, backward-bending impellers produce the least, while radial impellers fall in between. However, the forward-curving impeller has low efficiency and is not used (only in fans). Compressors use backward-curving impellers, which are further divided into two types: normal curvature (β₂ɑ=30°-60°) (commonly used in compressors and referred to as the compressor type) and strong backward curvature (β₂ɑ=15°-30°) (widely used in water pumps; in compressors it is only used in the last few stages of medium and low-flow, high-pressure compressors, and is referred to as the water pump type). 3 Axial force and balance of the rotor. When a centrifugal compressor is in operation, the pressures on both sides of the impeller are not equal; that is, the pressure of the gas behind the impeller is higher than the pressure of the gas entering the impeller, which results in an axial force acting on the rotor in the direction of the low-pressure side. Axial force is harmful to the proper operation of compressors; it causes the rotor to shift towards one end, resulting in a loss of the correct relative position between the rotating parts and the fixed components. In severe cases, the rotor may collide with the fixed parts, leading to accidents. The magnitude of the axial force can be calculated based on the forces acting on the impeller of the centrifugal compressor. In centrifugal compressors, the axial forces on the rotor are generally balanced using thrust bearings. However, in order to reduce the load on these thrust bearings, a balance disk is often used to counteract most of the axial force, with the remaining portion being taken care of by the thrust bearings. A balance disk is a component that uses the pressure difference between the gases on its two sides to balance axial forces. It is mounted on the shaft at the high-pressure end of the compressor; one side of it is connected to the pressure at the high-pressure end outlet, while the other side leads to the atmosphere or the intake pipe. This helps to maintain a pressure difference between the two sides and also reduces leaks. A labyrinth seal is installed between the outer edge of the balance disk and the fixing element to prevent gas leakage. The balance of axial force can also be achieved by allowing air to enter on both sides of the impeller or by installing the impeller in reverse. It should be noted that the purpose of balancing the axial force on the rotor is to reduce this force and thereby lessen the load on the thrust bearings; however, a certain amount of axial force must still be applied to the thrust bearings, otherwise the rotor will move back and forth while in operation. 4 The thrust disc balances only a portion of the axial force; the remaining axial force is transmitted to the thrust blocks on the thrust bearings through the thrust disc, thereby achieving force balance. The contact surface between the thrust disc and the thrust blocks must be very smooth, and the gap between them should be filled with appropriate lubricating oil, so that the thrust blocks do not wear out under normal operation. When a centrifugal compressor starts up, the rotor moves towards the other end. To ensure that the rotor remains in its proper position, it needs to be held in place by thrust forces on both sides. This is because, at the time of startup, the gas pressure on either side of the balance disc has not yet been established; as long as there is gas flow, the rotor will move in a direction opposite to the normal axial force. Therefore, it is necessary to have thrust forces on both sides of the rotor to prevent accidents. 1- Thrust disc 2- Thrust block. The purpose of the shaft sleeve is to maintain a certain distance between the impellers on the shaft, preventing the impellers from shifting along the shaft. (II) Fixed components — The stator consists of components such as the casing, partition, seals, intake chamber, and volute. Fixed elements such as diffusers, bends, and recirculators are formed between the partitions. 1 Intake chamber: The function of the intake chamber is to introduce gas smoothly and evenly from the intake pipeline or intercooler into the impeller. Basic forms of the suction chamber: a) Suction chamber with axial air intake; b) Elbow-type suction chamber with radial air intake; c) Suction chamber with semi-spiral casing for radial air intake; d) Suction chamber with horizontal semi-spiral casing for air intake. 2) Diffuser: The function of the diffuser is to convert the kinetic energy of the high-speed airflow exiting the impeller into static pressure energy. When the gas flows out of the impeller, it still has a high flow velocity. For compressors, the focus is on increasing static pressure energy (pressure) rather than speed. The requirement regarding speed is merely to ensure that the desired volume of gas can be maintained in the gas pipeline with a certain cross-sectional area. To make full use of this portion of kinetic energy in order to increase the gas pressure, a diffuser with a gradually increasing flow area is installed behind the impeller. Diffusers generally come in various forms such as bladeless, bladed, and straight-wall diffusers. 3 Curves and return channels: In multi-stage centrifugal compressors, gas must change direction between stages; for this purpose, curves are used. These curves are annular spaces formed by the casing and partition plates. The passage connected behind the curve is the returner; its function is to direct the airflow evenly in the desired direction toward the next stage, and it consists of partitions and guide vanes. The guide vanes are usually arc-shaped; they can be cast as one piece with the cylinder or manufactured separately and then connected together with bolts. 4 Shell: The main purpose of the shell is to collect the gas that flows out after the diffuser or the impeller, and guide it out of the machine. The cross-sectional shapes of the volute include circular, plow-shaped, trapezoidal, and rectangular. (III) Centrifugal compressors with bearings are equipped with radial bearings and thrust bearings, both of which are sliding bearings. The function of a radial bearing is to support the rotor and maintain it in a specific radial position, enabling it to operate at high speeds properly. The thrust bearing bears the remaining axial force on the rotor, preventing it from moving axially toward the low-pressure side and maintaining the rotor’s axial position within the cylinder. It is generally installed on the low-pressure side of the rotor. 1 Introduction to sliding bearings: A sliding bearing consists of three parts: a bearing housing, a shaft sleeve or bearing shell, and a lubrication system. The journal rotates within the bearing, and lubricating oil forms an oil film between the journal and the bearing bush to reduce friction and wear. Sliding bearings can be classified into hydrodynamic bearings and hydrostatic bearings based on their friction properties and the principle by which oil films are formed ； Based on their structural design, they are divided into two main categories: integral type (bush type) and split type (bearing shell type). 2 Working principle of hydrodynamic bearings (1) Principle of formation of hydrodynamic lubrication. Hydrodynamic bearings utilize the viscosity of oil and the high-speed rotation of the shaft to carry the lubricating oil into the wedge-shaped space between the shaft and the bearing bore, thereby creating a pressure oil film that separates the shaft from the bearing. This allows the shaft to rotate within the bearing, resulting in hydrodynamic lubrication. (2) Three conditions for achieving dynamic pressure lubrication: A wedge-shaped gap must be formed between the shaft journal and the bearing working surface, and the relative sliding direction must ensure that the lubricating oil flows from the larger cross-section to the smaller one ； There should be sufficient relative sliding speed between the b-axis journal and the bearing ； c The lubricating oil must have a certain viscosity and must be supplied continuously. Note: There are five conditions stated on the skill assessment certificate: There should be a certain gap between the A-axis journal and the bearing working surface ； The B-axis journal has sufficient rotational speed ； C The amount of lubricating oil is sufficient and its viscosity is appropriate ； The D-axis journal and bearings should have precise geometric shapes and low surface roughness values ； E Multi-support bearings must maintain coaxiality. (3) When the rotor changes from rest to rotating at a certain angular velocity, as shown in the figure, the rotor journal is supported by the dynamic pressure of the lubricating oil flowing in the clearance between the journal and the bearing, thereby preventing dry friction and collision between the journal surface and the bearing surface. The dynamic pressure of the lubricating oil is generated because when the journal moves relative to the bearing, if the amount of oil entering through the larger gap above is greater than the amount of oil flowing out through the smaller gap below, the oil in the converging oil wedge is compressed, which immediately increases the dynamic pressure. This leads to a reduction in the speed at which oil flows in and an increase in the speed at which oil flows out, thereby maintaining equal flow rates for inflowing and outflowing oil, in accordance with the principle of continuity of flow. The converging oil film, also known as an oil wedge, is able to withstand certain external loads because it generates hydrodynamic pressure that pushes the shaft journal to one side; under normal conditions, the shaft journal operates stably in this eccentric position. The length e of the line connecting the two centers, OO’, is called the eccentricity, while the angle between the line OO’ and the line of action of the external load W is called the deflection angle. Each eccentricity e corresponds to a specific shape of the oil wedge, and is linked to a particular load W and deflection angle; the greater the eccentricity e, the smaller the minimum oil film thickness hmin, and thus the greater the load-carrying capacity. Oil film oscillation of 3 bearings: Oil film oscillation is a self-excited oscillation of the rotating shaft caused by the oil film in sliding bearings, and it can generate amplitudes equal to or even greater than those that occur when the shaft reaches its critical speed. Oil film oscillation not only leads to failures in high-speed rotating machinery, but it can sometimes also cause damage to the bearings or the entire machine. When the rotor is subjected to some external disturbance, the forces acting on the journal are no longer in balance. In this case, the rotor rotates not only around the center of the journal but also around the center of the bearing, which leads to instability of the rotor and the formation of oil film oscillations. It is caused by the pressure of the oil film, and is also known as \"oil shock\" or \"slipping rotation\". Several properties of oil film oscillation: (1) Oil film oscillation occurs above twice the critical speed of the rotating shaft, and its swinging direction is the same as the rotation direction of the shaft ； (2) The whipping angular velocity of the oil film oscillation is independent of the rotational angular velocity of the shaft, and is approximately equal to the angular velocity at the critical speed of the shaft ； (3) Oil film oscillation is different from the vibration generated by the rotating shaft at its critical speed; once it occurs, an increase in speed will not stop it ； (4) Reducing the bearing width makes oil film oscillation less likely to occur ； (5) If the bearing support is designed to be self-aligning, allowing a slight degree of misalignment when installing the bearings at both ends of the shaft also helps to prevent oil film oscillation. Factors affecting oil film oscillation: (1) Shafting structure design – it affects the stiffness of the rotating shaft, and thus the critical speed ； It also affects the load distribution on the rotating shaft as well as the degree of shaft deflection ； During operation, the degree of eccentricity of the rotating shaft affects its critical speed, as well as the operating conditions of the bearings, that is, their performance. (2) Bearing load – A low load on the bearing can cause oil film oscillation or other abnormal vibrations. (3) Bearing oil inlet temperature: Oil temperature has a significant impact on oil film oscillations. When other conditions remain constant, a higher oil temperature results in lower oil viscosity, which in turn reduces the minimum oil film thickness. As a result, the operating point of the bearing, as well as the stiffness and damping coefficient of the oil film, all change. Under normal conditions, a higher oil temperature results in a smaller minimum oil film thickness and a larger eccentricity; as a result, oil film oscillations are less likely to occur in the bearings, which is why the stable operating speed increases. (4) Bearing shell clearance: The bearing shell clearance affects the stability of the bearing, mainly by influencing the minimum clearance required for its operation; this minimum clearance is an important factor for stable performance. The smaller the minimum gap, the more stable the bearing operates. (5) Other factors: Bearing tension, the stiffness of the support base and foundation, etc., have an impact on the stability of the shafting system. Qualitatively, increasing the support stiffness and damping improves stability, with increased damping having a particularly significant effect on enhancing stability. The oil film oscillation phenomenon has the following characteristics: (1) Oil film oscillation occurs at speeds above twice the first critical speed of the rotor; the direction of the oscillation is the same as the rotation direction of the shaft. Once oscillation starts, the amplitude increases sharply, and it does not decrease even if the speed is increased further, as shown in the figure. Severe vibration can sometimes lead to tile burning and the disruption of the oil film oscillation in the shafting system. (2) During oil film oscillation, the axis vortex frequency is usually the first-order natural frequency of the rotor, and the mode shape is the first-order mode shape. (3) During oil film oscillation, the direction of the axis vortex is the same as the rotation direction of the rotor, which is positive vortex motion. The self-excitation caused by dry friction is a reverse vortex. (4) Half-speed vortexing may occur when the rotational speed is below twice the first-order critical speed, with the vortex frequency being half of the rotational speed. The amplitude of semi-speed vortexing is small; if the speed is increased further, it will develop into oil film oscillation, as shown in Figure 5. Half-speed eddy currents usually occur in high-speed, light-load bearings. Figure 5 Half-speed vortex oil film oscillation (5). Oil film oscillation has an inertial effect; the speed at which oil film oscillation occurs when accelerating is different from the speed at which it disappears when decelerating. D Measures to prevent oil film oscillation: (1) Control the rotational speed, that is, keep it away from twice the rotor’s first critical speed. (2) Control the pressure, viscosity, and temperature of the lubricating oil. (3) Improve the accuracy of rotor balancing and the quality of bearing assembly. (4) Choose a bearing structure with strong seismic resistance. 4 Several commonly used vibration-damping bearings (1) Ordinary cylindrical bearings: These bearings operate with a relatively large degree of eccentricity at low speeds under heavy loads, which makes them stable. However, at high speeds under light loads they operate with a very small degree of eccentricity, resulting in instability; once oil film oscillations occur, it is difficult to suppress them. Therefore, cylindrical bearings are rarely used for high-speed, light-load rotors. (2) Elliptical bearings: These bearings are composed of upper and lower arc segments, as shown in Figure 1, and are widely used due to their ease of manufacturing. Its characteristic is that both the upper and lower arcs are significantly eccentric from the bearing center, resulting in two oil wedges. The oil film pressure of the oil wedge on it will act to suppress the instability of the aforementioned journal; due to geometric symmetry, this type of bearing allows the journal to rotate in both directions. Elliptical bearing (3) – Multi-blade bearing. This type of bearing is composed of several arc-shaped shaves, which can be symmetric or asymmetric. Its performance is similar to that of the elliptical bearing; each section has a greater degree of eccentricity, and there are more oil wedges. Since the shaft journal is subjected to the action of multiple oil wedges, its vibration suppression performance is better than that of the elliptical bearing. The asymmetric three-blade bearing (4), or multi-oil-foil bearing, is as shown in the figure. The vibration-damping performance of this type of bearing is similar to that of a multi-blade bearing; however, due to the asymmetry of the oil wedges, it allows the shaft journal to rotate in only one direction. The four-oil- wedge bearing (5), also known as a tilting-pad bearing, is composed of multiple movable pads that can deflect around a pivot point. This is currently considered the bearing with the best vibration suppression performance. It not only has a large number of oil wedges, but also, when external changes cause the center of the journal to momentarily deviate from its balanced position, the bearing shells can deflect around their pivot points, allowing them to automatically return to the balanced position; as a result, there are no factors that could sustain oscillation, giving it excellent stability. The four-wafer live bearing 4, being a thrust bearing, is also divided into upper and lower halves, just like radial bearings; there are positioning pins on the mid-surface, and these parts are connected together with bolts. A positioning sleeve is used between the spherical housing and the spherical seat to prevent relative rotation. Due to the spherical support, it is possible for adjustment to be made automatically based on the degree of shaft deflection. The thrust bearing works together with the thrust disc; the thrust disc mounted on the shaft rotates along with the shaft, and it transfers the thrust from the shaft to several stationary thrust blocks. A layer of babbitt alloy is also cast on the working surface of these thrust blocks, and the thickness tolerance of these thrust blocks is less than 0.01–0.02 mm. Mitchell thrust bearings and Kingsbury bearings are widely used in centrifugal compressors. During normal operation, the axial force in a centrifugal compressor always acts in the direction of the low-pressure side; the thrust block that bears this axial force is known as the main thrust block. When the compressor starts, the direction of the airflow force is toward the high-pressure side, and this force causes axial movement of that high-pressure side. To prevent such axial movement, another thrust block is installed; this block is located opposite to the main thrust block and is referred to as the auxiliary thrust block. A certain gap is left between the thrust disc and the thrust block to facilitate the formation of an oil film; this gap is generally within the range of 0.25 to 0.35 mm. Most importantly, the maximum value of this gap should be less than the minimum axial gap between the fixed component and the rotating component, so as to prevent the moving and stationary parts from coming into contact. The lubricating oil enters the spherical housing through the inlet at the bottom of the sphere, and then splits into two paths: one path goes to the radial bearings via the mid-surface, while the other path leads to the thrust bearings through two sets of inclined holes. Part of the oil that reaches the thrust bearings flows into the main thrust block, while the other part goes into the secondary thrust block. 5 Hydrostatic bearings: Hydrostatic bearings utilize a hydraulic system to supply pressurized oil between the shaft journal and the bearing, thereby separating the shaft journal from the bearing and ensuring that the bearing remains entirely in a state of fluid friction under various loads and rotation speeds. Hydrostatic bearings have advantages such as high load-bearing capacity, low frictional resistance, and long service life, but they require a complete hydraulic oil supply system. (IV) Sealing: To reduce the amount of air leakage through the gap between the rotor and the fixed components, sealing is often installed. Sealing is divided into internal sealing and external sealing. The function of internal seals is to prevent gas from flowing back between stages, such as the wheel cover seal at the wheel cover area, and the partition seal between the partitions and the rotor. Labyrinth seals are generally used. The external seal is designed to reduce and prevent gas from inside the machine from leaking out, as well as to stop outside air from entering the machine, such as the seals at the machine’s ends. At the shaft ends, labyrinth seals, pneumatic seals, water ring seals, oil film seals, floating ring seals, and snap-ring seals are generally used. 1 Labyrinth seal: A labyrinth seal consists of several annular sealing teeth arranged in sequence around the rotating shaft; gaps and expansion cavities are formed between these teeth. As the fluid to be sealed passes through these intricate gaps, a throttling effect occurs, thereby preventing leakage. Labyrinth seals are currently a commonly used sealing device in centrifugal compressors, employed for both external and internal sealing of the compressors. The gas flow in a labyrinth seal (see figure): as the gas flows through the gaps between the comb-shaped labyrinth seal elements, it undergoes an expansion process, with pressure dropping from P1 to P2 at the right end. This expansion occurs gradually; when the gas enters the sealing chamber through these gaps, the sudden increase in cross-sectional area causes the gas flow to form strong vortices, resulting in almost complete loss of velocity. There is a pressure difference between the gases on either side of the sealing surface. The pressure inside the sealing chamber is equal to the pressure in the gaps. According to the laws of gas expansion, as pressure decreases, velocity should increase and temperature should decrease. However, since the flow of gas through the narrow gaps is of a throttling nature, the kinetic energy obtained by the gas due to the pressure drop is completely lost in the sealing chamber and converted into useless heat energy. This heat energy then warms up the gas again, causing the temperature to rise back to its level before the pressure dropped. This process repeats itself each time the gas passes through another seal element and chamber, until the pressure reaches P2. It can be seen that a labyrinth seal utilizes the principle of throttling: each time gas passes through a tooth, its pressure decreases, and after passing through a certain number of teeth, a significant pressure drop occurs. In essence, a labyrinth seal creates a pressure difference resistance to the flow of gas, thereby reducing the amount of gas that can pass through. The commonly used types of labyrinth seals include the following. The smooth-type shaft serves as the optical axis; the seal unit is equipped with comb-like teeth or embedded tooth pieces, resulting in a simple structure. Smooth labyrinth seals vs. twisted labyrinth seals: To enhance the throttling and pressure-reduction effect of each tooth, twisted labyrinth seals were developed, offering a better sealing performance than those with a smooth design. The meander-type labyrinth seal and the stepped-type one both offer better sealing performance than the smooth-type; they are commonly used for sealing impeller covers. Stepped labyrinth seals typically have 3 to 5 sealing teeth. 2 Floating ring seal (oil film seal): The principle of the floating ring seal is based on the oil film formed between the floating ring and the shaft sleeve under high pressure; this film creates throttling and pressure reduction, thereby preventing gas from flowing from the high-pressure side to the low-pressure side. The floating ring seal enables an oil film to be formed in the gap between the ring and the shaft, while allowing the ring itself to move freely radially. The ring on the high-pressure side is called the high-pressure ring, while the ring on the low-pressure side is called the low-pressure ring. These rings can move freely in the radial direction but cannot rotate. The pressure of the sealing oil is usually about 0.5 Kg/cm2 higher than that of the process gas; this oil enters the sealing chamber and flows toward the high-pressure side through the gap between the high-pressure ring and the shaft, where it forms an oil film that seals off the high-pressure gas. The other stream of oil flows out through the gap between the low-pressure ring and the shaft, returning to the oil tank. Usually, there are several low-pressure rings in order to achieve a proper sealing effect. The floating ring seal is made of steel, with its end faces coated in tin bronze; the inner side of the ring is filled with babbitt alloy to prevent short-term contact between the shaft and the oil ring, as babbitt alloy serves as an anti-wear material. Floating ring seals can achieve complete leaklessness, and are widely used as shaft sealing devices for compressors. The clearance between the floating ring and the sleeve is very small, while the clearance of the inner ring is at (0. 5-1) ‰D (where D is the shaft diameter); the clearance of the outer ring is in the range of (1-1. 5) ‰D (where D is the shaft diameter); the clearance of the inner ring is smaller than that of the outer ring. Floating ring seals exhibit good adaptability to centrifugal compressors with high pressure differences and high rotational speeds. 7 Operation and Common Faults of Centrifugal Compressors (I) Performance Regulation of Centrifugal Compressors – The operation of changing the operating conditions of a compressor is referred to as the regulation of the compressor’s performance. Common adjustment methods include: throttling at the compressor outlet, throttling at the compressor inlet, using movable inlet vanes, changing the compressor’s speed, and using rotatable diffuser blades. 1 Outlet throttling of the compressor: A valve is installed in the compressor’s exhaust pipe, and the flow rate is adjusted by controlling the degree of opening of this valve. As the valve is closed, the resistance increases, and the pipeline characteristic curve becomes steeper as the valve is further closed, resulting in a corresponding decrease in flow rate. The smaller the valve is closed, the greater the resistance loss. For compressors with a steep performance curve, the losses are significant; therefore, they are generally used less often. a Diagram b Features Adjusting the opening degree of the throttle valve in the compressor outlet pipe is the simplest method of regulation. Its feature is that it does not alter the characteristic curve of the compressor; instead, it only changes the resistance characteristic curve of the piping network depending on the opening degree of the valve, thereby changing the operating point of the compressor, as shown in the figure ； Reducing the valve opening decreases the flow rate, and vice versa ； When the valve is closed, the resistance in the piping network increases; the pressure loss is primarily due to the additional local losses caused by the valve, which results in a decrease in the efficiency of the entire system. Moreover, the steeper the performance curve of the compressor, the greater the drop in efficiency ； This method is simple and easy to implement, with convenient operation. 2 Compression unit inlet throttling regulation: An inlet control valve is installed in the intake pipe of the compressor, and the flow rate and pressure are adjusted by changing the degree of opening of this control valve. Adjusting the throttle at the compressor inlet is a simple and power-saving method of regulation. As shown in the figure, changing the opening degree of the valve in the intake duct can alter the position of the compressor’s performance curve, thereby changing the flow rate or pressure of the airflow being delivered. The advantage is power savings. Moreover, the steeper the performance curve of the compressor, the more power can be saved. Another advantage of inlet throttling is that it shifts the compressor’s performance curve toward lower flow rates, allowing it to operate stably at reduced flow rates without experiencing surge. The disadvantage is that the throttling resistance causes a certain pressure loss and reduces the exhaust pressure. To ensure a uniform flow field at the compressor inlet, it is required that there be a sufficiently long straight pipe between the valve and the compressor inlet. Inlet airflow is a widely used control method. 3. Rotating inlet guide vane adjustment method (inlet pre-swirl adjustment). Inlet pre-swirl adjustment is more economical than inlet and outlet throttling adjustment, but the mechanism for rotating the guide vanes is relatively complex. Therefore, it is not widely used in centrifugal compressors, but is more commonly used in axial flow compressors. 4 Methods for adjusting the rotating diffuser blades: Centrifugal compressors equipped with blade diffusers have relatively steep performance curves, and when the flow rate decreases, severe separation occurs first in the blade diffuser, leading to surge. However, by changing the inlet angle of the diffuser blades to match the angle of airflow entry, the aforementioned disadvantages can be avoided. This results in a significant shift of the performance curve toward the low-flow region, thereby greatly reducing the surge flow rate and expanding the range of stable operating conditions, with only minor changes in pressure and efficiency. This regulation method can satisfactorily meet the requirements for flow control, but it has little effect on changing the outlet pressure. This type of adjustment mechanism is quite complex, which is why it is used less frequently. 5 Changing the compressor speed for adjustment: If the prime mover can have its speed adjusted, then by changing the speed it is possible to alter the position of the compressor’s performance curve; as the speed decreases, the performance curve shifts toward the lower left. Adjusting the rotational speed results in significant changes in both pressure and flow rate, which allows the stable operating range to be greatly expanded. This approach does not cause any additional losses nor requires any extra structural elements, making it an economical and simple solution. (II) Operation and Maintenance of Centrifugal Compressors 1 Key Points for Operation (1) Preparations before startup a. The drive motor and gear transmission should be tested individually and in combination, and must pass the inspections to be in good condition and ready for use. Install the coupling between the drive motor, the gear transmission, and the compressor, and recheck the alignment of the rotors to ensure it meets all requirements. b. The cleaning and adjustment of the unit’s oil system have been completed successfully; the oil quality meets the required standards, and the oil storage level is appropriate. Inspect the main oil tank, oil filter, and oil cooler; add oil if the oil level in the tank is low. Check the oil temperature; if it is below 24°C, a heater should be used to raise the oil temperature above 24°C. The oil cooler and oil filter should also be filled with oil and of any air removed, and the switching position between the oil cooler and the filter should be set to the side that needs to be in use. Check the main oil pump and the auxiliary oil pump to ensure they are operating properly and in the correct direction. Oil thermometers and pressure gauges should be complete, have appropriate ranges, and function properly. Dry nitrogen is pumped into the accumulator to keep the gas pressure inside it within the specified range. Adjust the oil pressure throughout the oil circuit system to meet the design requirements. Check that all interlock devices in the oil system are operating properly to ensure the safety of the unit. c. The inlet filters of the compressor should be clean and undamaged; the filter elements in the inlet filters have been replaced, and the filters are in good condition. d. The drain valves of the compressor cylinder and pipelines have been opened to drain the condensate; after that, they are closed partially, and finally closed once inflation is complete. e. Establish a cooling water circulation by introducing water to the intercoolers at various stages of the compressor, remove any air from the system, and bring it into operation. f. The process piping system must be in good condition; all blind flanges should have been removed and reinstalled. It is not allowed for the expansion, contraction, or vibrations of the pipes to exert stress on the cylinder body. g. Adjust the valves on the process gas pipelines to the appropriate positions as required for startup; generally, the inlet and outlet valves of the compressor should be closed, the reflux valve or vent valve used to prevent surge should be fully open, and the outlet valve of the process system should also be fully closed. The opening and closing of various types of valves should be smooth and accurate, without any sticking. h. Verify that safety valves and explosion-proof plates on the compressor pipes and associated equipment are properly installed; that the safety valves are calibrated to meet the required standards; and that the specifications of the explosion-proof plates are in compliance with requirements. i. The compressors and their associated machinery are equipped with complete instrumentation; the ranges, temperatures, pressures, and accuracy levels of these instruments meet the required standards, and important instruments should come with certification of calibration compliance. Check that the electrical wiring and instrument air system are in good condition. The control valves must be flexible and accurate, and the automatic control safety system must pass inspections to ensure accurate operation. j. All interlocks of the unit have been tested and adjusted, and all set values meet the requirements. The anti-surge protection control system has passed the calibration tests; all vent valves and anti-surge return valves should operate swiftly without any sticking. k. Based on the analysis, it has been confirmed that the gas composition within the process system before and after the compressor inlet and outlet valves meets the design requirements, or that the system has been properly purged with nitrogen. l. Perform a barring test to check whether the unit rotor can rotate smoothly, with no friction or sticking. (2) Start the startup of oil system A. Start the auxiliary units first, then the main unit. The auxiliary oil pump can be started when the oil temperature is above 15°C, while the main engine can be started only when the temperature reaches above 24°C. B Gas displacement, feed process gas. The intake should be slow to allow the seal oil pressure to match the gas pressure. The centrifugal compressor unit of the C compressor must be fully prepared and pass inspection and acceptance before it can be started in accordance with the procedures specified in the regulations. For turbine-driven centrifugal compressors, the speed increases gradually from low to high after startup, and there is no issue of overload caused by too rapid acceleration as in motor-driven compressors. Generally, the inlet valve is opened fully, as are the backflow valve or vent valve used for anti-surge protection. During the acceleration process, it is not allowed to remain within a speed range of ±10% of the critical speed of either rotor. When passing through the critical speed, the acceleration should be rapid, and close attention must be paid to the vibration of the unit. Under load condition D, with the inlet valve fully open, the anti-surge return valve or vent valve is gradually closed to increase the outlet pressure step by step. During the process of closing the valve to increase pressure, close attention must be paid to surge; if signs of surge are detected, the valve should be opened promptly. After boosting the pressure, set the anti-surge automatic control valve to the “automatic” position. The normal shutdown sequence for the E compressor is the reverse of the startup sequence. Be sure to turn the valves on and off slowly. The lubricating oil pump and the seal oil pump should be shut down after the unit has been completely stopped and cooled down. Prevent reverse rotation after the compressor stops. (III) Common faults and troubleshooting. Abnormal vibration and noise of the compressor: Possible causes and corrective measures ① The alignment accuracy of the unit has been compromised; check the vibration of the unit. If the axial vibration amplitude is high and the vibration frequency matches the rotational speed, or sometimes is twice, three times... etc. of it, remove the coupling so that the prime mover can rotate alone. If there is no abnormal vibration in the prime mover, then misalignment is likely the cause, and re-alignment is necessary. ② Rotor imbalance: Check the vibration levels. If the radial vibration amplitude is high and the vibration frequency is n, then the amplitude is proportional to the amount of imbalance and n2 ； At this point, the rotor should be inspected to check for any dirt or damage; if necessary, the rotor should be rebalanced. ③ Check for friction and damage to the rotor blades – inspect them for signs of friction or damage, and repair or replace them as needed. ④ Check whether the main shaft is bent; if necessary, straighten it. ⑤ Inspect the coupling for any faults or imbalances – remove it, check its balance, and make any necessary repairs. ⑥ Check the bearings for abnormal conditions – examine the radial clearance of the bearings and make adjustments as required; also check the interference between the bearing housing and the bearing shells; if it is too small, increase it ； If the bearing alloy is damaged, the bearing sleeve must be replaced. ⑦ Poor sealing leads to friction between the sealing elements; the vibration patterns become irregular, and a metallic grinding sound can be heard during startup or shutdown. Repair or replace the sealing ring. ⑧ If there is poor gear meshing in the gear speed increaser, check its gear meshing condition. If the vibration is low but the frequency is high and it corresponds to a multiple of the number of teeth, with the noise changing rhythmically, then the misalignment between the meshing gears should be corrected. ⑨ If the foundation bolts are loose or the foundation is not solid, repair the foundation and tighten the foundation bolts. ⑩ If the oil pressure and oil temperature are abnormal, check the oil pressure, oil temperature, and operation status of each oil system, and make adjustments if any abnormalities are found ； If the oil temperature is low, heat the lubricating oil. ⑾ If there is dirt in the oil, which prevents it from being clean, this can cause wear on the bearings; check the quality of the oil, improve filtration, and change the oil regularly. Inspect the bearings and replace them if necessary. ⑿ Check for any debris that may have entered or accumulated inside the machine; inspect the airflow channels in the rotor and cylinders and remove any such debris. ⒀ Check for condensed water inside the machine; examine the interior of the compressor and remove any condensed water. ⒁ Check for compressor surge – verify that the compressor is operating away from the surge point, that there is sufficient surge margin, and adjust the operating conditions according to the specified performance curves by increasing the intake volume; also check whether the anti-surge device is functioning properly. ⒂ Gas pipelines can exert additional stress on the casing; these pipelines must be securely fixed to prevent excessive stress from being applied to the compressor cylinders ； The piping system should have sufficient elasticity to compensate for thermal expansion. ⒃ Machines operating near the compressor can cause its foundation and base to separate from each other, which increases the stress on the connecting pipes. ⒄ When the compressor load changes rapidly, adjust the opening of the throttle valve. ⒅ If any components are loose, tighten them and install additional anti-loosening mechanisms. Possible causes of compressor surge, along with corresponding solutions: ① The operating condition falls within the surge zone or is too close to it. Check the position of the compressor’s operating point on the characteristic curve; if it is too close to the surge boundary or within the surge zone, take action to eliminate surge. ② Insufficient surge margin. The surge margin set for various operating conditions should be between 1.03 and 1.50; it should not be too low. ③ Insufficient inlet flow. This can be caused by an insufficiently open inlet valve, a dirty or frozen filter element, blocked inlet passages, or a reduced or interrupted supply of gas. Identify the cause and take appropriate measures. ④ Excessive pressure in the compressor’s outlet gas system. This may occur if gas is not released or returned when the compressor slows down or stops, if the outlet check valve fails or is not properly sealed, allowing gas to flow back. Identify the cause and take corrective action. ⑤ The relief valve or return valve does not open in time when operating conditions change. When inlet flow decreases, the speed drops, or the speed rises rapidly, check the characteristic curve and open the relief or return valve promptly. ⑥ The surge prevention device is not set to automatic mode. It should operate in automatic mode during normal operation. ⑦ The surge prevention device malfunctions or does not function properly. Regularly check its performance; if it malfunctions, is inaccurate, or does not operate properly, repair or adjust it promptly. ⑧ Incorrect settings for the surge prevention parameters. Set these parameters accurately and test them regularly; correct any inaccuracies as soon as they are detected. ⑨ Too rapid acceleration or pressure increase. Changes in operating conditions should not occur too quickly; they should happen gradually and smoothly. ⑩ Pressure must be reduced before slowing down the compressor. Reducing pressure first before slowing down helps prevent surge. ⑾ Changes in gas properties or state. Before changes in gas properties or state occur, adjust the characteristic curve accordingly and set the surge prevention parameters based on the new curve. ⑿ Damage to compressor components. Damaged or missing seals between stages, balance disc seals, and “O” rings can cause surge. Regularly inspect these components to ensure they are in good condition. ⒀ Faulty check valves in the compressor’s outlet pipeline. Regularly check the check valves in the outlet pipeline to ensure they function properly, thereby preventing gas from flowing back when the compressor slows down or stops