HG/T 2266-1992 Technical requirements for centrifugal compressors for oil refining and chemical industry
Some standard content:
Chemical Industry Standard of the People's Republic of China
HG/T2266-92
Centrifugal compressors for oil refining and chemical industry
Technical conditions
Approved on January 8, 1992
Implemented on July 1, 1992
Published by the Ministry of Chemical Industry of the People's Republic of China Standard exchange network www.bzaoeo.con Various standards industry materials are available for free download 1 Subject content and scope of application
2 Reference standards
3 Terminology
4 Technical requirements - Design and manufacture of compressors and auxiliary equipment 4.1 General
4.2 Casing, partitions and inlet guide vanes
4.3 Casing connections
4.4 External forces and torques
4.5 Rotating parts
4.6 Bearings and bearing housings
4.7 Shaft seals
4.8 Dynamics
4.9 Lubricating oil and sealing oil systems
4.10 Materials
4.11 Driver
4.12 Couplings and guards
4.13 Mounting base
4.14 Controls and instruments
4.15 Pipes and accessories
4.16 Special tools
5 Inspection and testing
5.1 General
5.2 Inspection
5.3 Testing
6 Marking, packaging, transportation, storage
7 Seller's information
7.1 Quotation:
7.2 Contract information
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Appendix A Typical data sheet of centrifugal compressors (supplement)Appendix B Seller's drawings and information requirements for centrifugal compressors (supplement)Appendix C Calculation of allowable external forces and torques for compressors (supplement) Appendix D Procedure for determining residual unbalance (supplement) Appendix E Specifications for common materials for major compressor components and comparison table of Chinese and American materials (reference) Appendix F Determination of allowable maximum tensile stress values for pressure-bearing components of compressors (reference) Appendix G Logic diagram for rotor dynamics analysis (reference) Appendix H Names of centrifugal compressor components (reference) Figure! Terminology diagram
Figure 2 Labyrinth shaft seal
Figure 3 Mechanical shaft seal·|| tt||Figure 4 Throttle ring shaft seal
Figure 5 Circular throttle sleeve floating ring shaft seal
Figure 6 Pump type floating ring shaft seal
Figure 7 Automatic gas seal·
Figure: Rotor response curve:
Bottom plate arrangement
Modified bottom plate support for leveling and grouting
Figure C1 Curve and calculation example of residual unbalance of external forces and external moments borne by the compressor
Figure G1 Logic diagram of rotor dynamics analysis
Centrifugal compressor parts name:
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Chemical Industry Standard of the People's Republic of China
Technical conditions for centrifugal compressors for oil refining and chemical industry Subject content and scope of application
HG/T2266-92
This standard specifies the design and manufacturing technical requirements, test methods and inspection rules, as well as packaging, transportation and other contents of centrifugal compressors and auxiliary equipment.
This standard applies to centrifugal compressors for conveying air or other gases. This standard does not apply to fans and blowers with gas boost values lower than 34kPa, nor to the integral gear speed increase assembled centrifugal compressors for general refinery instrument air as referred to in JB4113 standard. The seller can provide alternative design (see 7.1.1.13 for a list of differences). With the agreement of both parties, equivalent imperial sizes, fasteners and flanges may also be used. When this standard conflicts with the order contract, the conflicting part shall be subject to the provisions of the order contract. Note: For any clause with a number before the clause number of this standard, the relevant content shall be determined by the buyer. These determinations shall be directly filled in the data sheet prepared by the seller (see Appendix A). Otherwise, they shall be indicated and referenced in the quotation or order contract.
Steel pressure vessels
Basic dimensions of common threads
Common threads
Tolerances and fits
Methods for indicating grades of steel products
Methods for indicating grades of nonferrous metals and alloys
Technical conditions for high-quality carbon structural steel
Stainless steel bars
General pressure gauges
Ductile iron castings||tt ||Methods for measuring noise of fans and blowers Technical conditions for alloy structural steels
Radiography and quality classification of steel fusion welded joints 4216.14216.10 Grey cast iron flanges and gaskets 4457~4460 Mechanical drawing
Methods for indicating grades of cast iron
Methods for indicating grades of cast steel
Pipe threads sealed with threads
Technical conditions for turbine gear transmissions
9122~9131 (except GB9127) Steel pipe flanges and gaskets 9439||tt| |Gray iron castings
Magnetic particle inspection and quality rating method of steel castings 1992-01-08 approved by the Ministry of Chemical Industry of the People's Republic of China Exchange search network sm
Com Various standard industry information free download
1992-07--01 implementation
Technical drawing
Technical drawing
Title bar
HG/T2266-92
Folding method of copy drawing
12380~12384
Ductile iron pipe Flange
3 Terms
004 Chemical centrifugal compressors Terminology Chemical enterprise explosion and fire hazard environment power design specifications Boiler and steel pressure vessel butt weld ultrasonic testing Centrifugal and axial flow blowers and compressors Thermal performance test Pressure vessel welding process assessment
Steel pressure vessel magnetic particle testing
General refinery instrument air integral gear speed increaser assembled centrifugal compressors for fans Lubricating, sealing and regulating oil systems The terms used in this standard are defined in ZBJ72004 (see Figure 1). Several terms not included in ZBJ72004 are defined as follows: Alarm point: refers to the pre-set value of a parameter. When this value is reached, an alarm is triggered for a state that needs to be corrected as a warning. 3.1
3.2 Stop point: refers to the pre-set value of a parameter that requires an automatic or manual shutdown system. 3.3 Retention pressure: The pressure retained in the compressor system when the compressor is shut down. 3.4 Use of "design terms": This term (such as design power, design efficiency, design flow, design pressure, design temperature, design speed) should be avoided in the buyer's technical conditions. This term is only applicable to the designer and manufacturer of the compressor. The trip speed of the driver is specified in Table 1.
Table 1 Trip speed of the driver
Type of driver
When the turbine uses a Class A governor
When the turbine uses a Class B, C, or D governor
Gas turbine
Variable speed motor
Constant speed motor||tt| |Reciprocating engine
Note: The A, B, C, and D grades of the speed governor shall meet the speed control system grades specified in the following table: Speed control system
Speed range
(when specified)
4 Technical requirements - Design and manufacture of compressors and auxiliary equipment 4.1
Standard speed search network sm
Maximum speed
Trip speed (percentage of maximum continuous speed) 1t5
Maximum speed
Percentage of rated speed
Maximum speed
Specific operating point||tt ||2266-92
Minimum 1st critical speed
Minimum 2nd critical speed
A rigid shaft
Jumping speed-
A steam turbine
Jumping speed-
A gas turbine
Maximum continuous speed
100% speed
Normal speed
A certain operating speed
Minimum operating speed
Single-shaft gas turbine
Minimum operating speed
Steam turbine and dual-shaft gas turbine
Maximum 1st critical speed
126%=105%×1.2
115.5%=105%×1.1
110.3%=105%x1.05
Compressor rated pointwww.bzxz.net
105%=100%×1.05
Normal operating point
(98% assumption)
Specific operating point
(88%=98%×0.9)
4(78%=98%×0.8)
.66.3%=78%×0.85
Stability range
Adjustment range
Inlet flow
Figure 1 Terminology diagram
Note: (1) Except for special numerical relationships, the relevant values in this figure are only assumed values for illustration. (2) The 100% speed is determined by the energy head at point A and the flow required by the compressor design (to meet all specified working points, such as point C). (3) The energy head-flow curve at 100% speed should extend to 115% of the flow at point D, and the energy head-flow curve at other speeds should also be extended to the corresponding flow at each speed. For example, the energy head-flow curve at 105% speed should extend to a flow of 1.05×1.15 times the flow at point D; the energy head-flow curve at 90% speed should extend to a flow of 0.9×1.15 times the flow at point D, and so on. These points determine the flow limit line.
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HG/T2266-92
4.1.1 The equipment described in this standard (including auxiliary equipment) is designed and manufactured to have a service life of at least 20 years and an uninterrupted operation time of at least 3 years. This is a recognized design principle. 4.1.2 Unless otherwise specified, the normal energy head and flow of the compressor shall not be designed with negative deviations. Under the above conditions, its power shall not be greater than 104% of the normal value. Refer to the optional performance test provisions in 5.3.6.1. 4.1.3 The energy head-flow performance curve shall rise continuously from the rated point to the surge point. When the flow is at least 10% greater than the surge flow specified in the quotation, the compressor without bypass shall be able to work continuously. 4.1.4 Unless otherwise specified, the cooling water system shall be designed according to the following conditions: Flow velocity through the heat exchange surface
Maximum allowable pressure
Test pressure
Maximum pressure drop
Maximum inlet temperature
Maximum outlet temperature
Maximum temperature rise
Minimum temperature rise
Water side fouling coefficient
Shell corrosion allowance
1.5~2.5 m/s
>0.52 MPa
>0.79 MPa
0.35m2.K/kW
The cooling water system should have complete venting and drainage facilities. (This standard (G) indicates gauge pressure)
Note: If the minimum temperature rise is in conflict with the heat exchange surface velocity, the seller should notify the buyer. The heat exchange surface velocity is specified to minimize water side dirt; the minimum temperature rise is specified to minimize cooling water consumption. The buyer will review its final choice. 4.1.5 The equipment layout, including pipelines and auxiliary equipment, should be designed by the buyer and the seller. The equipment layout should leave enough space and safe passages for operation and maintenance. 4.1.6 The design of all equipment should take into account the maintenance and make it economical. The structure (providing bosses or positioning pins) and manufacturing of major components such as casings and bearing boxes should ensure accurate positioning during reassembly. 4.1.7 For vertically divided simple compressors, the inner casing should be designed to be easily withdrawn from the outer casing and easy to disassemble for inspection and replacement of parts.
4.1.8 The Buyer shall specify whether the equipment is installed indoors (heated or not) or outdoors (with or without a shed), as well as the climatic and environmental conditions of the equipment's work site (including minimum and maximum temperatures, abnormal humidity and dust, etc.). The compressor and its auxiliary equipment shall be suitable for operation under these specified conditions. In order to guide the user, the Seller shall list in the quotation the special protective measures required to be provided by the Buyer. If specified, the Seller shall provide antifreeze devices for the equipment.
4.1.9 The control of the noise sound pressure level of all equipment provided shall be completed by the Buyer and the Seller together. The equipment provided by the Seller shall not exceed the maximum allowable noise sound pressure level agreed upon by the Buyer and the Seller to comply with the applicable standards and regulations for local environmental noise limits. 4.1.10 The Buyer shall make suggestions to the Seller in the Inquiry for special requirements regarding the fluid. 4.1.11 The compressor shall be designed to operate without damage at the trip speed and the maximum allowable working pressure. 4.1.12 The performance of the compressor and the driver, as shown on the test bench and on the permanent foundation, shall meet the specified acceptance criteria. The performance of the entire unit after installation shall be the joint responsibility of the buyer and seller. 4.1.13 Many factors (such as piping loads, alignment under operating conditions, supporting structures, handling during shipment, and on-site handling and assembly) may have an adverse effect on on-site performance. In order to minimize the impact of these factors, the seller shall review and evaluate the buyer's piping and foundation drawings: If specified, the seller's representative shall: a. Observe the work of disassembling flanges to inspect the pipeline; b. Check alignment at operating temperature:
c. Be on-site for inspection at the initial alignment time.
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4.1.14 Motors, electrical components and electrical devices shall be suitable for the hazardous environment area code specified by the buyer on the data sheet and shall comply with the requirements of HGJ21-89 and the specifications specified and provided by the buyer. 4.1.15 Spare parts for compressors and auxiliary equipment shall meet all the requirements of this standard. 4.1.16 If specified, compressors and compressor units shall be capable of being field tested with air. The performance parameters and precautions to be taken shall be agreed upon by the buyer and seller.
4.1.17 Appendix H (reference) lists the names of the main parts of centrifugal compressors 4.2 Casing, plate and inlet guide vanes
4.2.1 The casing thickness shall be suitable for the maximum allowable pressure and test pressure, and there shall be at least 3.2mm corrosion allowance. The circumferential stress value for casing design shall not exceed the maximum allowable tensile stress value at the highest operating temperature determined in Appendix F (reference). 4.2.2 The feet and adjustment bolts shall have sufficient rigidity to enable the machine to be moved with transverse and axial pre-threads. 4.2.3 The maximum allowable working pressure of the casing shall be at least equal to the set value of the safety shelf; if the set value of the safety shelf is not specified, the maximum allowable pressure shall be at least 1.25 times the specified line high exhaust pressure (gauge pressure). The protection of the system shall be provided by the buyer. 4.2.4 Unless the buyer explicitly requires, it is not allowed to design the casing to fit several allowable maximum pressure levels (pressure segmented casing). If the buyer agrees, the seller shall specify the mechanical limit value and maximum allowable pressure of each part of the casing. 4.2.5 The horizontal part of the casing shall have sufficient rigidity so that when disassembling and installing the upper casing, it will not affect the running space and bearing alignment between the rotor and the casing.
4.2.6 Under the following conditions, the casing shall be made of steel: a. Air or non-combustible gas when the maximum allowable pressure exceeds 2.76MPa (G), b. Air or non-combustible gas when the calculated maximum outlet temperature exceeds 260℃ (the maximum temperature is usually near the Ruizhen point) at any operating point within the allowable operating range at the maximum continuous speed. c. Combustible gas or released gas.
4.2.7 In operating conditions other than those specified in 4.2.6, the casing may be made of iron or other materials. 4.2.8 Unless otherwise specified, when the hydrogen partial pressure (at the maximum allowable pressure) exceeds 1.38MPa (G), the casing shall be a vertical section structure. The hydrogen partial pressure is equal to the specified maximum molar volume percentage of hydrogen multiplied by the maximum allowable working pressure. 4.2.9 The horizontally split casing shall adopt metal-to-metal joint surfaces, and appropriate sealants shall be added between the joint surfaces and fastened with appropriate bolts. Gaskets (including strip gaskets) shall not be used for the connection of the horizontal split surface. With the consent of the buyer, the connection of the horizontal split surface of the housing may be machined into an annular groove on the split surface and sealed with an O-ring. In the vertical split housing, when the connection between the end cover and the cylinder is made of gaskets, the position of the gasket shall be limited. The gasket material shall be suitable for all specified conditions of use. 4.2.10 In order to facilitate disassembly and reassembly, top screws, guide rods and housing locating pins shall be provided. When the joint surface is separated by screws, grooves shall be machined on the flange surface that receives the top screw to prevent leakage or poor fit of the joint surface. The guide rod shall be of sufficient length to prevent the studs on the internal parts and the housing from being damaged during loading and unloading. The lifting rings or eyebolts provided on the upper housing are only used for lifting the upper housing. The lifting method of the assembled whole machine shall be specified by the seller.
4.2.11 The use of screw holes for connection on pressure-bearing parts shall be avoided as much as possible. To prevent leakage from the pressure surface of the casing, the thickness of the metal around the smooth hole or screw hole and at the bottom of the hole should be at least equal to half the nominal diameter of the bolt, plus a corrosion allowance. 4.2.12 It is not allowed to use fillers to prevent leakage in the gap between the stud and the hole. 42.13 The surface roughness of the compressor mounting surface after fine machining should be Ra3.2~6.3um. The fastening hole or foundation bolt hole should be perpendicular to the mounting surface or other surface, and the hook flat seat diameter should be 3 times the drilled hole diameter. 4.2.14 When using an angle stud connection, the stud should be screwed in. The optimal depth of the stud hole should be 1.5 times. Both ends of the stud should be chamfered, and the threaded end should be removed by 1.5 mm. 4.2.15 Internal and external bolt connections should comply with the provisions of 4.2.15.14.2.15.4. 4.2.15.1 The detailed provisions of the butterfly pattern should comply with the provisions of GB196 and GB197. Standard report mm.bessb.ca Various standard industry information free download HG/T2266-92
4.2.15.2 For external connections, studs should be used instead of bolts. 4.2.15.3 For external connections, sufficient wrench space should be left around the studs to allow the use of socket wrenches or box wrenches. 4.2.15.4 Unless otherwise agreed by the Buyer, hexagon socket, slotted nut or butterfly nut type bolts shall not be used for external connections. 4.2.16 The stage partitions and inlet guide vanes shall be suitable for all specified operating conditions, as well as start-up, shutdown, tripping, return to stability and transient surge. When used for intermediate main processes, the Buyer shall specify the maximum and minimum pressures at each connection, and the Seller shall confirm that the provided partitions can meet the maximum pressure difference.
4.2.17 The design of internal connections shall minimize leakage and be easy to disassemble. 4.2.18 In order to minimize internal leakage, replaceable labyrinth seals shall be provided at all internal seals. Unless otherwise agreed by the Buyer, the labyrinth seal shall be designed on the static side and easy to replace. 4.2.19 If there is no special requirement from the Buyer, the partition shall be designed as a horizontal split structure. The partition shall be provided with eyebolt holes or other lifting measures shall be adopted for disassembly.
4.2.20 If the partition is required to be cooled, the upper and lower partitions of the horizontal split shall have independent cooling channels. The inlet and outlet of each coolant channel shall be connected to the manifold at the top and bottom of the casing respectively. 4.3 Casing connection
4.3.1 General
4.3.1.1 All process gas connections on the casing shall be suitable for the maximum allowable pressure of the casing specified in 4.2.3. 4.3.1.2 All connections of the buyer shall be easy to maintain without moving the machine. 4.3.1.3 Joints, pipelines, valves and accessories shall not adopt specifications such as nominal diameter DN32, 65, 90 and 125mm. 4.3.1.4 The material requirements (including impact value) of the pipes welded to the casing shall meet the material requirements of the casing, but shall not meet the material requirements of the connected pipes.
4.3.1.5 The welding of all pipes shall be completed before the hydraulic test (see 5.3.2). 4.3.2 Main process flow connection
4.3.2.1 The inlet and outlet connections shall be flanged, or connected with studs after hooking the plane, and the pipe opening orientation shall be in accordance with the provisions of the data sheet. The inlet and outlet pipes of vertically split compressors shall be arranged on the outer casing, not on the end cover. For vertically split cantilever compressors, the process gas inlet pipes can be arranged on the end cover.
4.3.2.2 Grey cast iron pipe flanges shall comply with GB4216.1~4216.10 standards; ductile iron pipe flanges shall comply with GB12380~GB12384 standards; steel pipe flanges shall comply with GB9112~9131 standards (excluding GB9127). Steel flanges with a nominal diameter greater than DN600mm shall comply with the large-diameter carbon steel flange standards agreed upon by both parties. 4.3.2.2.1 Thick flat flanges are allowed for flanges on non-cast iron casings. 4.3.2.2.2 Flanges with a thicker thickness and a larger outer diameter than the above applicable steel flange standards are allowed. 4.3.2.3 When connecting parts other than the above applicable steel flange standards are used, the buyer's consent is required. If specified, the seller shall provide all matching flanges and matching studs and nuts. 4.3.2.4 Cast iron flanges shall be flat flanges. For flanges with a nominal diameter equal to or less than DN200mm, the minimum thickness shall be PN5MPa flange of GB12381~GB12384.
4.3.2.5 The center circle of the bolt holes of all casing flanges shall have a certain degree of coaxiality with the inner hole of the casing flange so that the annular area of the machined sealing convex (concave) platform can accommodate a complete standard gasket without the gasket extending into the fluid. 4.3.2.6 The finished products of all flanges and pipes shall comply with the material and surface roughness requirements specified in GB4216.10, GB9125 or GB12348.
4.3.3 Auxiliary connections
4.3.3.1 Auxiliary connecting pipes shall include but not be limited to the following main pipes: ventilation, injection, drainage (see 4.3.3.2), cooling water, lubrication and sealing oil, flushing, buffer gas and balance disc cavity pipes. 6
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4.3.3.2For horizontal sub-shells, the seller shall provide all gas channel discharge interfaces: For vertical sub-shells, the discharge interface shall be set at the lowest point of each inlet section, the lowest point between the inner and outer shells and the lowest point of each exhaust section. If specified, each section including the balance plate cavity shall be provided with a separate discharge port
4.3.3.3The flange shall comply with the provisions of GB9112~9131 (except GB9127). 4.3.3.4The nominal diameter of the auxiliary pipe shall be at least DN19mm (see 4.3.1.3), and shall be connected by flanges of socket welding, or connected by studs after processing. For socket welding structures, a gap of 1.5mm shall be left between the pipe end and the bottom of the socket groove of the shell before welding. 4.3.3.5 When it is inconvenient to use socket welding flange connection or to connect the open structure with studs after processing, after negotiation with the buyer, screw interfaces with nominal diameter DN19~38mm can be used, which should be set according to the provisions of 4.3.3.5.1~4.3.3.5.3. 4.3.3.5.1 The screw holes and bosses of pipe threads should comply with the relevant provisions of GB7306 standard. 4.3.3.5.2 The pipe threads should comply with the RP/R thread standards of GB7306. 4.3.3.5.3 The threaded joints should be sealed by welding, but the joints of cast iron equipment, instrumentation and other joints that need to be frequently disassembled during maintenance are not allowed to be sealed by welding. The sealing welding should comply with the applicable standards agreed upon by the buyer and the seller. 4.3.3.6 The pipe joints should preferably not exceed 150mm in length and should be installed in the screw holes or socket welding openings. The minimum wall thickness of threaded pipe joints and socket welded pipe joints shall be as specified in Table 2. Each pipe joint shall adopt butt welding, socket welding or loose sleeve flange. 4.3.3.7 The screw holes that are not connected to the pipes shall be screwed with solid round head steel plugs that meet the requirements of Table 3. The pipe threads shall meet the Rp/R standard of GB-7306. These plugs shall at least meet the requirements of the casing material. The plugs that need to be removed shall be made of corrosion-resistant materials and the threads shall be lubricated. Sealing tape shall not be used for plugs used in oil circuits. Plastic plugs are not allowed. 4.4 External forces and external torques
4.4.1 The pipe openings of the compressor shall be designed to withstand the external forces and external torques calculated in Appendix C (Supplement). After considering the supporting position, supporting structure, length and reinforcement degree of the pipe opening, shape and thickness of the compressor, etc., the allowable external force and external moment shall be increased whenever possible. The seller shall provide the allowable external force and external moment on each pipe opening in the form of a table. Table 2
Nominal pipe diameter DN
Minimum wall thickness of threaded pipe joint
Minimum wall thickness of socket welding pipe joint
Tube burst size code
4.4.2 The casing and support shall be designed to have sufficient strength and rigidity to reduce the coupling caused by the application of allowable external force and external moment. The coaxiality error caused is limited to within 50 μm. 4.5 Rotating parts
4.5.1 The shaft shall be made of heat-treated integral steel suitable for machining. The main shaft with a diameter greater than 200 mm after finishing shall be made of forged steel parts: The shaft with a diameter equal to or less than 200 mm after finishing is generally made of forged steel parts, but with the consent of the buyer, hot-rolled bars can also be used, but the quality and heat treatment of the bars shall comply with the standards for shaft forgings. 4.5.2 The shaft end that cooperates with the coupling shall comply with the applicable standards agreed upon by the buyer and the seller. 4.5.3 In addition to other protective measures for the shaft agreed by the buyer, replaceable sleeves shall be provided at the interstage seal gap, all carbon ring seals and the journal gas seal. Under the specified conditions of use, these sleeves shall be made of corrosion-resistant materials. The sleeve with the minimum clearance end seal shall be properly treated with anti-wear treatment to prevent leakage between the shaft and sleeve (see 4.10.1.7). 4.5.4 The design of the shaft-sleeve-impeller assembly shall ensure that the rotor will not produce temporary or permanent deformation. The impeller assembly method shall ensure sufficient coaxiality and balance accuracy under all specified operating conditions, including overspeed to trip speed. 4.5.5 The sensing surface of the probe used to detect the radial runout of the rotor shaft shall be coaxial with the bearing journal.All the sensing surfaces of the rotating shaft (the two sensing surfaces for measuring radial and axial runout) shall be free of scratches, marks or any other surface discontinuities (such as oil holes and keyways). These surfaces shall not be sprayed, coated or sleeved. The final surface roughness shall be Ra0.4~0.8um. This surface is best achieved by grinding or polishing with oilstone. These sensing surfaces shall be well demagnetized or otherwise treated so that the total electrical and mechanical runout does not exceed the specified values of a and b. a. The sensing surface for the radial runout probe is 25% of the maximum allowable bee-to-bee amplitude or 6.4um, whichever is greater. b. The sensing surface for the axial runout probe is 13um. Note: If the above specified values cannot be achieved after various methods are adopted, the buyer and seller shall negotiate to change their acceptance standards. 4.5.6 Each rotor shall have a unique identifiable mark. The mark shall be placed on the shaft end without the coupling. 4.5.7 A closed impeller consisting of a disk, blades and a wheel cover or a semi-open impeller consisting of a disk and blades may be used. The impeller may be welded, riveted, milled or cast. Other manufacturing methods such as electro-erosion and brazing are also permitted if the buyer agrees. 4.5.8 Both welded and riveted impellers may be made of forgings or castings. The welds in the flow channel shall be smooth and free of weld spatter. The impeller shall be heat treated after welding to eliminate internal stress. There shall be no sharp edges at the inlet and outlet of the blades. 4.5.9 Cast impellers shall be fully finished except for the flow channel. Only with the buyer's consent, repair welding is permitted. 4.5.10 Welding is not permitted to balance the impeller. 4.5.11 The design of pressure-bearing parts shall take into account the stress concentration factor caused by the geometric shape. The joints of the impeller, blades and disk, and the changes in the key and shaft section shall be designed with fillet structures that can reduce the stress concentration factor. 4.5.12 Integral thrust discs should be used first. If the shaft seal adopts floating ring type, mechanical contact type or gas shaft seal that needs to be frequently assembled and unassembled, replaceable thrust discs should be used: When an integral thrust disc is used, the thrust disc should have at least 3.2mm additional thickness so that it can be repaired when damaged. When replaceable thrust discs are used for easy assembly and maintenance, the thrust discs should be firmly fixed on the shaft to prevent micro-vibration wear.
4.5.13 The surface roughness of the two thrust surfaces of the thrust disc should not be greater than Ra0.4um, and the indicated value of the total axial runout of any thrust surface of the thrust disc should not exceed 12.7μm.
4.5.14 It is allowed to design the compressor into a structure without a balancing disc. 4.5.15 If necessary, a balancing disc, a balancing pipe and a balancing air outlet should be provided to reduce the axial load on the thrust bearing. One or more pressure gauge connections shall be provided to indicate the pressure in the balancing chamber, rather than the pressure in the balancing pipe. 4.5.16 The diameter of the balancing pipe shall be designed so that when the clearance of the labyrinth seal is twice the original design value, the balancing pipe can still convey the gas leakage of the balancing disc without causing the thrust bearing to exceed the rated load value (see 4.6.3.3). 4.5.17 In order to prevent static voltage on the shaft, the residual magnetism of the rotating element shall not exceed 0.5 mT (milli-Tesla). 4.6 Bearings and Bearing Boxes
4.6.1 General
4.6.1.1 The bearings shall be hydrodynamic radial bearings and thrust bearings. If other types of bearings are used, the formal consent of the buyer is required. 4.6.1.2 Unless otherwise specified, radial bearings and thrust bearings shall be equipped with bearing metal temperature sensors. The sensors shall comply with the applicable standards agreed upon by the buyer and the seller.
4.6.2 Radial bearings
HG/T2266--92
4.6.2.1 Bushing type or tilting pad radial bearings should be used. In order to facilitate assembly, a split structure should be adopted. The use of a non-split structure must be subject to the buyer's consent. The wheel bearings shall have precision-bored steel bearing bodies, replaceable bushings, tiles and housings with bearing alloy. The bearings shall be provided with anti-rotation pins and shall be axially positioned.
4.6.2.2 The bearing design shall suppress fluid dynamic instabilities and provide sufficient damping over the entire range of allowable bearing clearances so that the vibration of the rotor is limited to the specified maximum amplitude when the equipment is running unloaded or loaded at the specified operating speed (see 4.8.5.5).
4.6.2.3 The replaceable bushings, tiles and housings shall be installed in horizontally split bearing housings. When replacing these parts, it is not necessary to dismantle the upper casing of horizontally split machines or the end covers of vertically split machines. Unless otherwise agreed by the Purchaser, the bearings shall be designed so that the inner sleeve of the coupling can be replaced without disassembly of the inner sleeve of the coupling.
4.6.2.4 Bearing housings of compressors equipped with sleeve-type radial bearings shall be capable of field installation of tilting pad bearings without re-machining.4.6.3 Thrust bearings
4.6.3.1 Hydrodynamic thrust bearings shall be of the multi-sector steel sleeve type of cast bearing alloy, designed to have equal thrust capacity in both directions, and arranged to provide continuous pressure lubrication on each side. Both sides of the bearing shall be tilting pad type with automatic load balancing to ensure that each pad bears an equal share of the load even if there are slight differences in pad thickness. 4.6.3.2 The thickness of each designed and manufactured gasket should have precise dimensions (thickness difference) to facilitate interchangeability or replacement of individual gaskets. 4.6.3.3 The thrust bearing size design should comply with the requirements for continuous operation under the most unfavorable specified working conditions. The thrust calculation should include but not be limited to the following factors:
a. Maximum design internal clearance of the seal and 2 times the maximum design internal clearance b. Step change in the diameter of the pressurized rotor;
c. Maximum pressure difference between stages;
d. Specified limit changes in inlet, interstage and energy head; e. External thrust transmitted by the coupling:
f. When the motor is directly driven, the maximum thrust from the motor sleeve bearing, 4.6.3.4 For gear coupling, the external thrust should be calculated according to the following formula: 0.25×9545P
Where: F--external thrust, kN;
P,--rated power, kW:
N,-rated speed, r/min;
D--coupling shaft hole diameter, mm,
Note: The coupling shaft hole diameter is approximately equal to the pitch circle radius, ND
4.6.3.5 The external thrust of the diaphragm coupling should be calculated based on the allowable maximum deflection specified by the coupling manufacturer. 4.6.3.6 If a thrust bearing is subjected to the thrust of two or more rotors (for example, in a gearbox), if the direction of these forces is such that their resultant forces are numerically added, then the resultant force should be used; otherwise, the maximum value of these thrusts should be taken. 4.6.3.7 The load borne by the thrust bearing shall not exceed 50% of the limit rated load specified by the thrust bearing manufacturer, and the thrust bearing shall be selected accordingly. The limit load rating refers to the load value under the following two conditions: the load when the minimum allowable oil film thickness is produced for continuous operation of the rotor without failure; the load when the local maximum temperature on the tile does not exceed the wax deformation or ductility strength of the bearing alloy. The smaller value of these two conditions should be taken. The specification of the thrust bearing shall be checked and confirmed by the buyer. When selecting the specification of the thrust bearing, the following factors shall be considered:
ee. Various standard industry materials are available for free download12 Integral thrust discs should be used first. If the shaft seal adopts floating ring type, mechanical contact type or gas shaft seal that needs to be frequently assembled and unassembled, replaceable thrust discs should be used: When an integral thrust disc is used, the thrust disc should have at least 3.2mm additional thickness so that it can be repaired when damaged. When replaceable thrust discs are used for easy assembly and maintenance, the thrust discs should be firmly fixed on the shaft to prevent micro-vibration wear.
4.5.13 The surface roughness of the two thrust surfaces of the thrust disc should not be greater than Ra0.4um, and the indicated value of the total axial runout of any thrust surface of the thrust disc should not exceed 12.7μm.
4.5.14 It is allowed to design the compressor into a structure without a balancing disc. 4.5.15 If necessary, a balancing disc, a balancing pipe and a balancing air outlet should be provided to reduce the axial load on the thrust bearing. One or more pressure gauge connections shall be provided to indicate the pressure in the balancing chamber, rather than the pressure in the balancing pipe. 4.5.16 The diameter of the balancing pipe shall be designed so that when the clearance of the labyrinth seal is twice the original design value, the balancing pipe can still convey the gas leakage of the balancing disc without causing the thrust bearing to exceed the rated load value (see 4.6.3.3). 4.5.17 In order to prevent static voltage on the shaft, the residual magnetism of the rotating element shall not exceed 0.5 mT (milli-Tesla). 4.6 Bearings and Bearing Boxes
4.6.1 General
4.6.1.1 The bearings shall be hydrodynamic radial bearings and thrust bearings. If other types of bearings are used, the formal consent of the buyer is required. 4.6.1.2 Unless otherwise specified, radial bearings and thrust bearings shall be equipped with bearing metal temperature sensors. The sensors shall comply with the applicable standards agreed upon by the buyer and the seller.
4.6.2 Radial bearings
HG/T2266--92
4.6.2.1 Bushing type or tilting pad radial bearings should be used. In order to facilitate assembly, a split structure should be adopted. The use of a non-split structure must be subject to the buyer's consent. The wheel bearings shall have precision-bored steel bearing bodies, replaceable bushings, tiles and housings with bearing alloy. The bearings shall be provided with anti-rotation pins and shall be axially positioned.
4.6.2.2 The bearing design shall suppress fluid dynamic instabilities and provide sufficient damping over the entire range of allowable bearing clearances so that the vibration of the rotor is limited to the specified maximum amplitude when the equipment is running unloaded or loaded at the specified operating speed (see 4.8.5.5).
4.6.2.3 The replaceable bushings, tiles and housings shall be installed in horizontally split bearing housings. When replacing these parts, it is not necessary to dismantle the upper casing of horizontally split machines or the end covers of vertically split machines. Unless otherwise agreed by the Purchaser, the bearings shall be designed so that the inner sleeve of the coupling can be replaced without removing the inner sleeve of the coupling.
4.6.2.4 Bearing housings of compressors equipped with sleeve-type radial bearings shall be capable of field installation of tilting pad bearings without re-machining. 4.6.3 Thrust bearings
4.6.3.1 Hydrodynamic thrust bearings shall be of the multi-sector steel sleeve type of cast bearing alloy, designed to have equal thrust capacity in both directions, and arranged to provide continuous pressure lubrication on each side. Both sides of the bearing shall be tilting pad type with automatic load balancing to ensure that each pad receives an equal share of the load even if there are slight differences in pad thickness. 4.6.3.2 The thickness of each designed and manufactured gasket should have precise dimensions (thickness difference) to facilitate interchangeability or replacement of individual gaskets. 4.6.3.3 The thrust bearing size design should comply with continuous operation under the most unfavorable specified working conditions. The thrust calculation should include but not be limited to the following factors:
a. Maximum design internal clearance of the seal and 2 times the maximum design internal clearance b. Step change in the diameter of the pressurized rotor;
c. Maximum pressure difference between stages;
d. Specified limit changes in inlet, interstage and energy head; e. External thrust transmitted by the coupling:
f. When the motor is directly driven, the maximum thrust from the motor sleeve bearing, 4.6.3.4 For gear coupling, the external thrust should be calculated according to the following formula: 0.25×9545P
Where: F--external thrust, kN;
P,--rated power, kW:
N,-rated speed, r/min;
D--coupling shaft hole diameter, mm,
Note: The coupling shaft hole diameter is approximately equal to the pitch circle radius, ND
4.6.3.5 The external thrust of the diaphragm coupling should be calculated based on the maximum allowable deflection specified by the coupling manufacturer. 4.6.3.6 If a thrust bearing is subjected to the thrust of two or more rotors (for example, in a gearbox), if the direction of these forces is such that their resultant forces are numerically added, then the resultant force should be used; otherwise, the maximum value of these thrusts should be taken. 4.6.3.7 The load borne by the thrust bearing shall not exceed 50% of the limit rated load specified by the thrust bearing manufacturer, and the thrust bearing shall be selected accordingly. The limit load rating refers to the load value under the following two conditions: the load when the rotor continuously runs with the minimum allowable oil film thickness without failure; the load when the local maximum temperature on the tile does not exceed the wax deformation or ductility strength of the bearing alloy. The smaller value of these two conditions should be taken. The specification of the thrust bearing shall be checked and confirmed by the buyer. When selecting the specification of the thrust bearing, the following factors shall be considered:
ee. Various standard industry materials are available for free download12 Integral thrust discs should be used first. If the shaft seal adopts floating ring type, mechanical contact type or gas shaft seal that needs to be frequently assembled and unassembled, replaceable thrust discs should be used: When an integral thrust disc is used, the thrust disc should have at least 3.2mm additional thickness so that it can be repaired when damaged. When replaceable thrust discs are used for easy assembly and maintenance, the thrust discs should be firmly fixed on the shaft to prevent micro-vibration wear.
4.5.13 The surface roughness of the two thrust surfaces of the thrust disc should not be greater than Ra0.4um, and the indicated value of the total axial runout of any thrust surface of the thrust disc should not exceed 12.7μm.
4.5.14 It is allowed to design the compressor into a structure without a balancing disc. 4.5.15 If necessary, a balancing disc, a balancing pipe and a balancing air outlet should be provided to reduce the axial load on the thrust bearing. One or more pressure gauge connections shall be provided to indicate the pressure in the balancing chamber, rather than the pressure in the balancing pipe. 4.5.16 The diameter of the balancing pipe shall be designed so that when the clearance of the labyrinth seal is twice the original design value, the balancing pipe can still convey the gas leakage of the balancing disc without causing the thrust bearing to exceed the rated load value (see 4.6.3.3). 4.5.17 In order to prevent static voltage on the shaft, the residual magnetism of the rotating element shall not exceed 0.5 mT (milli-Tesla). 4.6 Bearings and Bearing Boxes
4.6.1 General
4.6.1.1 The bearings shall be hydrodynamic radial bearings and thrust bearings. If other types of bearings are used, the formal consent of the buyer is required. 4.6.1.2 Unless otherwise specified, radial bearings and thrust bearings shall be equipped with bearing metal temperature sensors. The sensors shall comply with the applicable standards agreed upon by the buyer and the seller.
4.6.2 Radial bearings
HG/T2266--92
4.6.2.1 Bushing type or tilting pad radial bearings should be used. In order to facilitate assembly, a split structure should be adopted. The use of a non-split structure must be subject to the buyer's consent. The wheel bearings shall have precision-bored steel bearing bodies, replaceable bushings, tiles and housings with bearing alloy. The bearings shall be provided with anti-rotation pins and shall be axially positioned.
4.6.2.2 The bearing design shall suppress fluid dynamic instabilities and provide sufficient damping over the entire range of allowable bearing clearances so that the vibration of the rotor is limited to the specified maximum amplitude when the equipment is running unloaded or loaded at the specified operating speed (see 4.8.5.5).
4.6.2.3 The replaceable bushings, tiles and housings shall be installed in horizontally split bearing housings. When replacing these parts, it is not necessary to dismantle the upper casing of horizontally split machines or the end covers of vertically split machines. Unless otherwise agreed by the Purchaser, the bearings shall be designed so that the inner sleeve of the coupling can be replaced without removing the inner sleeve of the coupling.
4.6.2.4 Bearing housings of compressors equipped with sleeve-type radial bearings shall be capable of field installation of tilting pad bearings without re-machining. 4.6.3 Thrust bearings
4.6.3.1 Hydrodynamic thrust bearings shall be of the multi-sector steel sleeve type of cast bearing alloy, designed to have equal thrust capacity in both directions, and arranged to provide continuous pressure lubrication on each side. Both sides of the bearing shall be tilting pad type with automatic load balancing to ensure that each pad receives an equal share of the load even if there are slight differences in pad thickness. 4.6.3.2 The thickness of each designed and manufactured gasket should have precise dimensions (thickness difference) to facilitate interchangeability or replacement of individual gaskets. 4.6.3.3 The thrust bearing size design should comply with continuous operation under the most unfavorable specified working conditions. The thrust calculation should include but not be limited to the following factors:
a. Maximum design internal clearance of the seal and 2 times the maximum design internal clearance b. Step change in the diameter of the pressurized rotor;
c. Maximum pressure difference between stages;
d. Specified limit changes in inlet, interstage and energy head; e. External thrust transmitted by the coupling:
f. When the motor is directly driven, the maximum thrust from the motor sleeve bearing, 4.6.3.4 For gear coupling, the external thrust should be calculated according to the following formula: 0.25×9545P
Where: F--external thrust, kN;
P,--rated power, kW:
N,-rated speed, r/min;
D--coupling shaft hole diameter, mm,
Note: The coupling shaft hole diameter is approximately equal to the pitch circle radius, ND
4.6.3.5 The external thrust of the diaphragm coupling should be calculated based on the maximum allowable deflection specified by the coupling manufacturer. 4.6.3.6 If a thrust bearing is subjected to the thrust of two or more rotors (for example, in a gearbox), if the direction of these forces is such that their resultant forces are numerically added, then the resultant force should be used; otherwise, the maximum value of these thrusts should be taken. 4.6.3.7 The load borne by the thrust bearing shall not exceed 50% of the limit rated load specified by the thrust bearing manufacturer, and the thrust bearing shall be selected accordingly. The limit load rating refers to the load value under the following two conditions: the load when the rotor continuously runs with the minimum allowable oil film thickness without failure; the load when the local maximum temperature on the tile does not exceed the wax deformation or ductility strength of the bearing alloy. The smaller value of these two conditions should be taken. The specification of the thrust bearing shall be checked and confirmed by the buyer. When selecting the specification of the thrust bearing, the following factors shall be considered:
ee. Various standard industry materials are available for free download2 The thickness of each designed and manufactured gasket should have precise dimensions (thickness difference) to facilitate interchangeability or replacement of individual gaskets. 4.6.3.3 The thrust bearing size design should comply with continuous operation under the most unfavorable specified working conditions. The thrust calculation should include but not be limited to the following factors:
a. Maximum design internal clearance of the seal and 2 times the maximum design internal clearance b. Step change in the diameter of the pressurized rotor;
c. Maximum pressure difference between stages;
d. Specified limit changes in inlet, interstage and energy head; e. External thrust transmitted by the coupling:
f. When the motor is directly driven, the maximum thrust from the motor sleeve bearing, 4.6.3.4 For gear coupling, the external thrust should be calculated according to the following formula: 0.25×9545P
Where: F--external thrust, kN;
P,--rated power, kW:
N,-rated speed, r/min;
D--coupling shaft hole diameter, mm,
Note: The coupling shaft hole diameter is approximately equal to the pitch circle radius, ND
4.6.3.5 The external thrust of the diaphragm coupling should be calculated based on the maximum allowable deflection specified by the coupling manufacturer. 4.6.3.6 If a thrust bearing is subjected to the thrust of two or more rotors (for example, in a gearbox), if the direction of these forces is such that their resultant forces are numerically added, then the resultant force should be used; otherwise, the maximum value of these thrusts should be taken. 4.6.3.7 The load borne by the thrust bearing shall not exceed 50% of the limit rated load specified by the thrust bearing manufacturer, and the thrust bearing shall be selected accordingly. The limit load rating refers to the load value under the following two conditions: the load when the rotor continuously runs with the minimum allowable oil film thickness without failure; the load when the local maximum temperature on the tile does not exceed the wax deformation or ductility strength of the bearing alloy. The smaller value of these two conditions should be taken. The specification of the thrust bearing shall be checked and confirmed by the buyer. When selecting the specification of the thrust bearing, the following factors shall be considered:
ee. Various standard industry materials are available for free download2 The thickness of each designed and manufactured gasket should have precise dimensions (thickness difference) to facilitate interchangeability or replacement of individual gaskets. 4.6.3.3 The thrust bearing size design should comply with continuous operation under the most unfavorable specified working conditions. The thrust calculation should include but not be limited to the following factors:
a. Maximum design internal clearance of the seal and 2 times the maximum design internal clearance b. Step change in the diameter of the pressurized rotor;
c. Maximum pressure difference between stages;
d. Specified limit changes in inlet, interstage and energy head; e. External thrust transmitted by the coupling:
f. When the motor is directly driven, the maximum thrust from the motor sleeve bearing, 4.6.3.4 For gear coupling, the external thrust should be calculated according to the following formula: 0.25×9545P
Where: F--external thrust, kN;
P,--rated power, kW:
N,-rated speed, r/min;
D--coupling shaft hole diameter, mm,
Note: The coupling shaft hole diameter is approximately equal to the pitch circle radius, ND
4.6.3.5 The external thrust of the diaphragm coupling should be calculated based on the maximum allowable deflection specified by the coupling manufacturer. 4.6.3.6 If a thrust bearing is subjected to the thrust of two or more rotors (for example, in a gearbox), if the direction of these forces is such that their resultant forces are numerically added, then the resultant force should be used; otherwise, the maximum value of these thrusts should be taken. 4.6.3.7 The load borne by the thrust bearing shall not exceed 50% of the limit rated load specified by the thrust bearing manufacturer, and the thrust bearing shall be selected accordingly. The limit load rating refers to the load value under the following two conditions: the load when the rotor continuously runs with the minimum allowable oil film thickness without failure; the load when the local maximum temperature on the tile does not exceed the wax deformation or ductility strength of the bearing alloy. The smaller value of these two conditions should be taken. The specification of the thrust bearing shall be checked and confirmed by the buyer. When selecting the specification of the thrust bearing, the following factors shall be considered:
ee. Various standard industry materials are available for free download
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